EPA-460/3-76-008-a
March 1976
STUDY
ON OXIDES OF NITROGEN
AND CARBON FORMATION
IN DIESEL ENGINES -
FINAL REPORT
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Offiee of Mobile Source Air Pollution Control
Emission Control Technology Division
Ann Arbor, Michigan 48105
-------
EPA-460/3-76-008-a
STUDY
ON OXIDES OF NITROGEN
AND CARBON FORMATION
IN DIESEL ENGINES -
FINAL REPORT
C.J. Kan, T.J. T)son, and M.P. Heap
Htras) stems, Inc.
2400 Michelson Drive
Irvine, (California 92664
Contract No. 68-01-0436
EPA Project Officers: J.I,. Bascunana and G.D. Killredge
Prepared for
COORDINATING RESEARCH COUNCIL, INC.
30 Rockefeller Pla/a
(New York, N. Y. 10020
A PR AC CAPE 20-71
and
ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Emission Control Technology Division
Ann Arbor, Michigan 48105
Marth 1976
-------
This report is issued by the Environmental Protection Agency to report technical
data of interest to a limited number of readers. Copies are available free of
charge to Federal employees, current contractors and grantees, and nonprofit
organizations - as supplies permit - from the Air Pollution Technical Information
Center, Environmental Protection Agency. Research Triangle Park, North
Carolina 27711; or, for a fee, from the National Technical Information Service,
5285 Port Royal Road, Springfield, Virginia 22161.
This report was furnished to the Environmental Protection Agency by Ultrasystems,
Inc., Irvine, California, in fulfillment of Contract No. 68-01-0436. The
contents of this report are reproduced herein as received from Ultrasystems,
Inc. The opinions, findings, and conclusions expressed are those of the
author and not necessarily those of the Environmental Protection Agency.
Mention of company or product names is not to be considered as an endorsement
by the Environmental Protection Agency.
Publication No. EPA-460/3-76-008-a
11
-------
ACKNOWLEDGEMENTS
This report presents the outcome of almost four year's work and many
individuals have contributed to the final product. The authors would like to
express their thanks to several present and former Ultrasystems staff members,
G. P. Carver, C. McComis, K. Maloney, L. J. Muzio, C. H. Waldman and
R. P. Wilson, Jr., for their efforts in the early stages of the model's develop-
ment and in the conduct of the experimental investigations .
The CAPE 20-71 Steering Committee of the Coordinating Research Council
and the EPA originally defined the project and guided the work to it's final
outcome. The ultimate value of this project is largely due to the Committee,
without whose help and criticism it would have been difficult to complete. The
authors sincerely acknowledge their gratitude to the following individuals who
have served on the Steering Committee during the lifetime of the project:
F. J. Hills, Committee Chairman Mobil Research & Development
Corporation
J. L. Bascunana Environmental Protection Agency
N. G. Beck International Harvester
T. C. Belian Coordinating Research Council
J. E. Bennethum General Motors Research Laboratory
G. L. Borman University of Wisconsin
W. L. Brown Caterpillar Tractor
J.C. Hoelzer International Harvester
G. D. Kittredge ' Environmental Protection Agency
G. A. Lavoie Ford Motor Company
D. F. Merrion General Motors-Detroit Diesel Allison
P. C. Meurer International Harvester
J. M. Perez Caterpillar Tractor
3. M. Shahed Cummins Engine
A. D. Tuteja General Motors-Detroit Diesel Allison
A. V. Wilson Cummins Engine
A. E. Zengel Coordinating Research Council
-------
TABLE OF CONTENTS
Page No,
1.0 Summary 1-1
2.0 Introduction 2-1
3.0 The Methodology of Diesel Engine Modeling 3-1
3.1 The Objectives of Modeling 3-1
3.2 Review of Existing Models 3-2
3.3 A Qualitative Description of the Ultrasystems 3-4
Model
4.0 Detailed Description of a Model to Predict Pollutant 4-1
Formation in Diesel Engines
4.1 Cycle Thermodynamics and Heat Transfer 4-1
4.2 Fuel Injection and Spray Characterization 4-16
4.3 Air Spray Entrainment and Droplet Vaporization 4-17
4.4 Ignition Delay 4-21
4.5 Air Product Mixing 4-21
4.6 Nitric Oxide Formation 4-33
4.7 Soot Formation and Oxidation 4-38
4.8 Nitric Oxide and Soot Conservation Equations 4-39
4.9 Equilibrium Compositions 4-40
4.10 Fuel Representation 4-41
4.11 Method of Solution 4-42
4.12 Indicated Engine Performance and Engine Emission 4-43
5.0 The Empirical Nature of the Model 5-1
5.1 The Nature of the Empirical Constants 5-1
5.2 Experimental Engine and Exploratory Computations 5-6
5.3 Sensitivity Analysis 5-13
5.4 The Role of Droplet in NO Formation 5-17
6.0 Comparisons Between Engine Emission Levels and Model 6-1
Predictions
Effect of Engine Operational Parameters 6-1
Effect of Charge Intake Parameters 6-7
Effect of Fuel Injection Parameters 6-11
7.0 Limitations of the Model and Recommendations for 7-1
Future Work
8.0 Nomenclature 8-1
9.0 References 9-1
Appendix A - Simplified Representation of Fuel Spray A-l
Appendix B - Coefficients for Equations (4-25 and (4-26) B-l
Appendix C - Characterization of Diesel Combustion by C-l
Direct In-Cylinder Sampling
Appendix D - Sampling Valve Data Test Series 3 D-l
ii
-------
1.0 SUMMARY
A modular model describing the processes of energy release and
pollutant formation in direct injection diesel engines has been developed and
tested. The model consists of separate modules describing the following
processes:
• Cylinder thermodynamics and heat transfer;
• Fuel injection and spray characterization;
• Air spray entrainment and droplet vaporization;
• Ignition delay;
• Air product mixing;
• Heat release;
• Nitric oxide formation;
• Soot formation and oxidation.
It was recognized at an early stage of development that a model of this type must
be able to evolve with time. Consequently, the total package was constructed to
ensure that any of the modules listed above could be readily updated as more
definitive information becomes available. In essence, the present model pro-
vides a framework which can be readily used by other workers, after simple
modification, for their own specific purposes.
The model contains a considerable number of empirical constants
which have been divided into two groups:
• Those constants whose values are perhaps not universal, but
which remain unaffected by gross changes in engine design, or
do not strongly influence predictions of pollutant formation.
• Those constants which require adjustment to allow accurate pre-
dictions and which will require a certain degree of readjustment if
the model is used outside its original frame of reference.
A sensitivity study has been carried out to ascertain the influence coefficients
of these adjustable constants on nitrogen oxide and soot formation.
1-1
-------
Experience gained during the development of the model and during
subsequent application has highlighted two areas of major importance:
• Macromixing. The results of the sensitivity study indicated that
the most sensitive parts of the model are those parameters related
to fuel/air mixing.
• Spherical droplet diffusion flames. The model suggests that non-
premixed heat release regions can not be adequately described by
single droplet diffusion flames.
Chemistry inherently sets the lower limit on pollutant formation in diesel engines,
however, it is not the governing mechanism. Macromixing, fuel and air con-
tacting on the macroscale is the limiting mechanism. Consequently, it is not
surprising that those parameters included in the model which influence fuel/
air mixing have the strongest influence on pollutant formation.
Two regions are identified in the model which appear to be of major
importance from the aspect of NO formation:
• Those regions where combustion products are well mixed on a
scale which is small compared to the total volume of that region,
and therefore, local gradients can be neglected. The NO forma-
tion rate is calculated by a non-equilibrium homogeneous model
based upon extended Zeldovich kinetics.
• Those regions in the vicinity of fuel droplets which allow the
formation of spherical diffusion flames. These regions are typified
by steep gradients, and molecular diffusion plays a strong role.
Calculations indicate that the total exhaust NO level is dominated by NO pro-
duced in the former regions. In the model as it now stands, droplet diffusion
flames do not contribute a significant fraction of the total NO production under
normal diesel operating conditions. This can be attributed to:
• A large fraction of the droplets are contained within rich zones
which cannot support spherical diffusion flames;
1-2
-------
• The droplet lifetime is restricted to the time period close to top
dead center where the available volume is small and, therefore,
the volume available for the single flame is small, thus minimizing
the total flame surface area;
• The droplet environment contains NO which will reduce the net
rate of NO formation.
Model predictions have been compared with experimental data
generated in a single cylinder diesel engine for the following operation and
engine parameters:
* Load (or overall air/fuel ratio)
• Fuel injection timing
• Engine speed
• Compression ratio
• Exhaust gas recirculation
• Air swirl
• Fuel nozzle diameter
• Turbocharge.
Based upon comparison with single cylinder test data, it is concluded
that the overall predictive capability of the model with respect to pressure history
and exhaust NO level is adequate. However, its ability to predict exhaust soot
emissions can only be considered fair. The model includes an arbitrary
empirical relationship describing carbon formation in fuel rich regions since a
more fundamental understanding of the chemistry of carbon formation is not
presently available.
Experience with the model indicates that several minor modifications
are possible which will improve and extend the model; these include:
modifications to the equilibrium species tables to allow fuels of
different hydrogen/carbon ratios to be easily treated and to
extend these tables to higher pressures and richer mixtures;
limits can be imposed on the equilibrium assumptions to enable
carbon monoxide and hydrocarbon emissions to be included;
1-3
-------
the ignition model can be extended to more adequately reflect
the fuel history; and
mixing between burning zones can be included.
Four areas have been identified in which the model is deficient and which
require more extensive work to include in the model:
fuel injection schedules can rarely be approximated by a constant
injection rate. The present model is limited to a constant fuel
injection rate;
the model considers that the jet behaves throughout its lifetime
as a steady state jet, the influence of the transient nature of
the jet is neglected;
fuel/air mixing dictate pollutant emission levels and a more
realistic mixing model which reflects the three dimensionality
of the process and which includes fuel vapor segregation should
be included. An effort must be made to ascertain the degree to
which the well-ordered jet structure is maintained throughout the
cycle; and
as other components in the model are improved, the deficiencies
in the equilibrium heat release model will become more apparent
and some kinetic overlay will be called for.
In conclusion, it should be recognized that the model has only been compared
against one set of single cylinder engine data. Before further improvements are
made to the model, more extensive tests will be required to fully uncover both
the weaknesses and the strong points in the model. Modifications should also
be made to the basic model to allow prechamber engines to be accommodated.
1-4
-------
2.0 INTRODUCTION
This report presents the culmination of three years effort to provide
diesel engine designers with an analytical tool to aid in the development of
clean, efficient diesel engines. The initial phase of this program was concerned
with the construction and operation of a single cylinder engine to provide data
on pollutant formation and to give some insight into the mechanism of pollutant
formation. Experimental information was generated with which to test the
analytical model to be developed in a later phase. The successful completion
of this work has been adequately described by Wilson et ar . The second
phase of the work, namely the construction, modification and exercise of a
mathematical model to predict pollutant formation in a diesel engine is des-
cribed in this report.
The general objective of the total study was to provide industry with
a working computerized tool to aid engine development. This tool was to be
developed in an iterative manner in conjunction with diagnostic experiments
planned to provide information on the phenomena occurring within the cylinder
during combustion. The model development included in this report is confined
to direct injection engines, although the experimental investigations described by
Wilson et ar include data on prechamber configurations. At the initiation of
the project is was hoped that the model development would include both con-
figurations. However, the realities of modeling such a complex process as
intermittent combustion in a compression ignition engine dictated otherwise.
The progress of combustion and, therefore, energy release and pollutant for-
mation is dominated by the fuel/air mixing process. Since the mixing processes
are entirely different in the two configurations, this would necessitate two
separate models. Constraints imposed by time and budget precluded the develop-
ment of two models and it was decided to concentrate the modeling effort on
direct injection engines. The constraints stated above also prevented applica-
tion of the model to other engine geometries and multi-cylinder engines.
2-1
-------
3.0 THE METHODOLOGY OF DIESEL ENGINE MODELING
Diesel engines are a fact, they operate satisfactorily in all manner
of equipment. A significant reason for attempting to develop an analytical model
describing the processes occurring within a diesel engine can be that such a
model lies on the critical path to the construction of a more efficient, cleaner
engine. As a design tool the analytical model is at present in its infancy. The
complex interaction of physical and chemical processes occurring in a diesel
engine involve phenomena which are not completely understood at present.
However, each small advance in the state-of-the-art of modeling helps to
build what can be a powerful tool in the hands of engine designers.
The mathematical simulation along will not enable improved diesel
engines to be built. Development and prototype testing, rejection and redesign
are mandatory. However, the availability of an analytical model which will
guide this iterative process could greatly reduce development costs. This is
particularly applicable to attempts to reduce pollutant emissions through engine
redesign without incurring the penalty of increased fuel consumption. Pollutant
formation in diesel engines is controlled by several interrelated parameters
which mades complete redesign based upon intuition difficult. An accurate
analytical model which has been tested and the range of applicability established,
will help to guide this intuition, thus saving valuable experimental test resources.
Before presenting the detailed mathematical formulation, a qualita-
tive description of the model developed during this program will be given and
contrasted with other models.
3 .1 Review of Existing Models
Wilson et al reviewed, in considerable detail, the characteristics
of the various models which predate the present model. However, a concise
description of their salient features will be presented to allow an adequate
comparison to be made with the new Ultrasystems' model.
3-1
-------
• The NR£C (Northern Research and Engineering Corporation) Model
[Bastress, Chang and Dix ( 2 )]
This model includes a prescribed heat release schedule which has
two stages:
an initial premixed triangular distribution;
a secondary Gaussian distribution.
In both of these stages the fuel-air ratios over which heat release takes place
are specified. Random mixing is assumed to account for subsequent dilution
of the combustion products and Zeldovich kinetics are used to describe NO
production. The model requires the arbitrary specification of six parameters:
a fuel vaporization rate, fuel burning rate, dilution rate, an overall heat transfer
coefficient, a mean air-fuel ratio for heat release, and an air-fuel ratio excursion
parameter for the premixed heat release schedule. One major disadvantage of
the NREC approach is that there is no link between the value of these parameters
and the actual physical processes. Such a link will be necessary if the model
is to be used to assess the effect of design changes which influence the details
of the combustion process .
/ 3 \
• The CAV Model [Khan, Greeves , and Probert , and Khan and
Greeves^ 4 ' 1
The CAV Model represents an advance from the previous model in
that it includes a fuel spray model. The spray is modeled by a conical free jet
which impinges upon a plane wall, thus simulating the impact of the fuel jet on
the piston bowl. Experimentally determined ignition delays are input as is a
triangular premixed heat release distribution. The pre-exponential constant In
the modified Zeldovich NO formation mechanism was increased by a factor of
5.37 over the literature value in order to match experimental exhaust data. This
model also includes an Arrhenius type expression to allow soot formation to be
predicted.
• The Cummins Model [Shahed, Chiu and Yumlu ^ ', and Shahed
Chiu and Lyn (6} ]
A preliminary version of Cummins Model (Shahed et ai ' ) cor-
rectly represented the cycle thermodynamics, but ignored mixing and spray
3-2
-------
combustion. An empirical burning law based on the work of Lyn and an
(8 9)
empirical ignition delay correlation based on data of Lyn and Valdmanis '
were used and it was assumed that heat release took place under stoichiometric
conditions and the subsequent mixing between stoichiometric packages of com-
bustion products and excess air was not included. The NO model is also based
upon the Zeldovich mechanism with adjusted rate constants to correctly predict
experimental data.
The latest version of the Cummins model contains considerable improve-
ments over the preliminary model, particularly in the areas of mixing and fuel
spray combustion. The formulae describing unsteady jet mixing were obtained
via curve fitting experimental data obtained for various engine operating con-
ditions and geometry (e.g., various chamber pressures, temperatures, swirl
levels, injection pressures and hole sizes) . The stoichiometric burning and
non-dilution mixing assumptions were totally discarded. Thus, a more precise
mathematical representation of the NO formation processes, which is capable
of reflecting the important effects of local oxygen concentration history on NO
formation has been included.
3 . 2 A Qualitative Description of the Ultra s_vstems ' Model
The Ultra systems' model of pollutant formation in diesel engines has
( ] )
been developed along the guidelines set down by Wilson et al . However,
there are important differences in the details of the various model components.
The model was developed recognizing that the complexity of the process will
not allow the total elimination of all empirical constants from the analytical
expressions. The initial concept of eliminating all arbitrarily assigned con-
stants has not been achieved. Certain adjustable constants have been included
and these are discussed in detail in a later section. The aim of this section
is to describe the model in qualitative terms before the mathematical formula-
tion is presented in Section 4.
Figure 3-1 presents a general view of the various components of the
model which has been constructed in such a way that any of these components
can be modified as additional information becomes available. The model
simulates conditions occurring in the cylinder of a four stroke direct injection
-
-------
CO
I
,u
Engine Description
Operational Conditions
Engine
Performance
Pollutant
Emission
Fuel Spray Dispersion
Therm odynamtc
Cycle Analysis & Heat
Transfer
Pollutant Formation
NO
Homo- \ / Hetero-
geneous I I geneous
Soot
Heat Release
\Dlffusion/
Fig. 3-1 Elements of the Ultrasystems Model
-------
diesel engine from the crank angle degree of intake valve closure to the angle
at which the exhaust valve opens. The model is built around the thermodynamic
cycle and heat transfer analysis components which both receives and supplies
information to and from the other components of the model. The interaction of
the separate modules can be seen in Figure 3-1.
Fuel Soray Dispersion
In the fuel spray dispersion module the fuel jet is modeled as though
it were a steady axisymmetric unconfined wall jet. Figure 3-2 shows a pictorial
representation of the initial stages of the jet before impact. The fuel injection
schedule is considered to be uniform and the fuel forms droplets of a specified
size distribution immediately upon injection. In order to ascertain the subse-
quent history of all the fuel elements in 1° crank angle intervals it is considered
to be distributed equally between four shells. Thus, each element of fuel can be
denoted, both by its age and by its initial location in the jet. The fuel jet
entrains air in the manner prescribed by well-established theories of turbulent
jet mixing. Thus, the total mass in each shell will increase and those shells
on the periphery of jet will have fuel air ratios lower than those shells on the
jet axis.
Fuel Injection Schedule
m
Locii of
9inj=X°
Fuel droplets
Fuel vapor
Air
Locii of
9 = (x+l)°
Swirl \ Direction
equal fuel mass
Jet Boundary
and 9 = 1 intervals
Figure 3-2. Pictorial Representation of the Fuel Jet
3-5
-------
During the entralnment stage, the fuel droplets are treated as a single
component hydrocarbon compound which will begin to vaporize when a stable
wet bulb temperature is obtained. The rate of evaporation is influenced by the
swirl velocity of the air sweeping through the jet, thus as the fuel element
passes along the jet, its vaporization rate is increased, both by virtue of the
increase in temperature of its environment, and also because of the increased
angular velocity. The fuel that is vaporized remains in the parent shell. Thus,
immediately before ignition the fuel jet consists of parcels of fuel, all mixing
with air, which contain droplets in various stages of vaporization. The multi-
orifice fuel injector is handled by dividing the total available air equally between
the separate jets and assuming all jets are exactly similar.
Quite obviously, this is a very simple picture of a complex process
which Ignores all three dimensional effects, non-steady phenomena and influ-
ence of confinement. Several omissions should be enumerated:
• Cross flow — the swirl introduced with the air charge increases
entrainment, but the jet remains axisymmetric. With sufficiently
high swirl velocities the cross flow effects will deflect the jet
axis and distort the circular section into a kidney shape (see
Figure 3-3). Air entrainment will be different on the upwind and
downwind sides. There is the possibility that vapor and small
droplets will migrate in the direction of the cross flow.
• Non-steady jet —no attempt is made to include the transient
effects of an intermittent jet, i.e. , the nature of the leading and
the trailing edges.
• Three dimensional effects - the Jets exist in a finite and continu-
ously changing space. The jet impact upon the piston bowl is
not normal to a plane surface.
• Jet interaction —at no time are the fuel jets considered to
interact.
3-6
-------
Real Injection Schedule
m
Swirl \ Direction
Migration of vapor
and small droplets
Jet Cross Section
Figure 3-3. Schematic Showing Simplified Cross Flow Effects
Heat Release
The complex chemistry of ignition is not included in the model. A
shell containing air and fuel vapor will ignite when an empirical ignition
criterion has been satisfied. This criterion depends upon the properties of the
fuel and the average pressure and temperature over the period of time before
ignition takes place. Once ignited the vapor reacts with the air producing
equilibrium combustion products. Initial heat release rates were too slow
when a turbulent flame speed model was used to ignite other shells. An
improved simulation has been achieved by allowing successive ignition in
other shells when they also satisfy the ignition criteria .
3-7
-------
Once ignition has occurred, the location of that shell is no longer
tracked in space, but only in time. Mass transfer between air and zones
containing combustion products and vaporizing fuel droplets is no longer
considered to be controlled by the turbulent jet mixing model. All the air
is assumed to be equally available to all the parcels of fuel and the mixing
rate is dependent upon the mass of air available and the mass of each package
of burned products. Air is vitiated with combustion products, but no provision
is made for mixing to take place between packages of products. This random
mixing process is related to the turbulence properties of the air charge, but
it still includes an empirical constant.
The mechanism of heat release depends upon the state of the fuel
and oxygen availability. The chemistry of hydrocarbon combustion is not
included. Ignited fuel vapor, or fuel vaporizing in a rich zone is immediately
transformed into equilibrium products. Droplets vaporizing in oxygen rich
atmospheres can support spherical diffusion flames. No criterion is included
for flame blow-off or the production of wake flames.
Pollutant Formation
The model does not consider that the pollutant species, nitric oxide
and soot are in chemical equilibrium. The rates of formation and destruction
of these species are calculated from simple kinetic considerations assuming
that the species involved in their production and oxidation are in equilibrium.
Extensive reviews of the kinetic mechanism of NO formation in combustion
processes are available in the literature . The model includes the following
reactions:
N
which is usually referred to as the Zeldovich mechanism. This approach does
not include any coupling of the hydrocarbon chemistry with the nitric oxide
formation, either in terms of superequilibrium atom concentrations, or Fenimore
type reactions , e.g.,
HCN + N
3-8
-------
Once NO has been formed in lean zones, it could be reduced in rich zones by
(12)
such reactions as
NO + CH * HCO + N
which would allow the formation of nitrogen by
N + NO * N2 + O
Although NO9 has been observed both in flat flames and turbulent diffusion
flames, no attempt has been made to allow for the formation of NO~ within the
cylinder. This might occur by the reaction of NO with HO2 and its presence
in the exhaust would depend upon the quench rate since the high temperatures
would favor its conversion to NO.
The model allows for the formation of NO both in the homogeneous
burned gas zones and during droplet combustion. The spherical droplet diffusion
flames were originally included because it was believed that diffusion zones
of this type were characteristic of the regions of pollutant formation in diesel
engines. Also, mathematical analyses were available for single droplets con-
tained in an infinite environment. Initial calculations indicated that these
analyses gave the "improbable" situation that the flame volume was greater
than the avilable cylinder volume. Consequently, the diffusion flame analysis
was modified to include a bounded environment "filling up" with combustion
products.
The model for soot formation is similar to that used earlier by other
workers. The only difference is in the definition of zonal equivalence ratio
to include both reacted and unreacted oxygen. Oxidation of soot is also
included.
3-9
-------
4.0 DETAILED DESCRIPTION OF A MODEL TO PREDICT POLLUTANT
FORMATION IN DIESEL ENGINES
The physical and chemical processes controlling heat release and
pollutant formation in diesel engines are extremely complex. The model, des-
cribed in detail in this section, has been developed upon the premise that
many of these processes can be decoupled from each other and, therefor-e, the
total process can be subdivided into several tractable components. Although
each of these components contains certain simplifying assumptions to make
them mathematically amenable, every effort has been made to ensure that they
retain their physical significance. Thus, a modular model has been developed
which reflects the present state-of-the-art but which will allow any of its
components to be updated independently as an improved understanding of the
controlling phenomena is achieved. The separate components of the model
are:
• The cycle thermodynamics and heat transfer
• Fuel injection and spray characterization
• Air spray entrainment and droplet vaporization
• Ignition delay
• Air product mixing
• Combustion
• Nitric oxide formation
• Soot formation and oxidation
A complete description of symbols, superscripts and subscripts included in these
various components is given in Section 8, Nomenclature.
Cycle Thermodynamics and Heat Transfer
The cycle thermodynamic analysis serves as the skeleton upon which
the separate components of the whole model are assembled and it describes the
thermodynamic state of the cylinder contents during the compression and expan-
sion processes for that period of the cycle when both the intake and the exhaust
valves are closed.
At any instant the cylinder contents, which may be either gaseous or
liquid, are considered to exist in separate zones:
Air zone: the mixture of fresh air, recirculated exhaust gas (if any)
and residual gas. Subsequent to ignition the "air" zone is vitiated
and mixed with the burning gas zone. The properties related to this
zone are identified by subscript "a".
4-1
-------
Vaporizing liquid fuel zone: for thermodynamic purposes, all the
pre-ignition liquid fuel is considered to exist in one single zone.
This zone is identified by subscript "fl".
Burning liquid fuel zone: all the post-ignition liquid fuel, which
is either engulfed by combustion products or diffusion flame
envelopes is considered to exist in one single zone which is
identified by the subscript "fib".
Fuel vapor zone: thermodynamically, all the vaporized fuel
(either mixed or unmixed with air) is considered to exist in a
single zone prior to ignition, designated by the subscript "fv".
This approach might well introduce errors into the calculation
of the thermodynamic properties during the ignition delay. How-
ever, it is considered that these errors will have little influence
on overall pollutant predictions.
Burning zones: the zones which contain the combustion products
are identified by subscripts which indicate both the time of
ignition and the proximity to the fuel spray axis at the time of
ignition (see Figure 3-2). Each burning zone is subsequently fed
either by the combustion products generated by diffusion flames or
by fuel vapor (when conditions do not allow diffusion flames to be
established) from the fuel droplets engulfed by the zone itself.
Although mass transfer between burning zones and the air zone,
i.e., dilution and vitiation, is taken into consideration, the mass
transfer between burning zones is not considered. Burning zones
are signified by double subscript "bm,I".
The basic assumptions contained in the cycle thermodynamics model are:
pressure is uniform throughout the cylinder;
each zone is considered to be thermodynamically
homogeneous;
each zone radiates independently to the cylinder wall (i.e. ,
radiative exchange between zones is not considered);
the kinetic energy of the fluid within the cylinder is
neglected except when it is required to characterize the
velocity of the cylinder contents;
the sole source of heat necessary for fuel vaporization is
provided by the "air" zone;
the specific heat of both the vapor and liquid fuel is constant;
a "mixed burned" assumption is used to account for the com-
bustion of the premixed air and vaporized fuel and equili-
brium combustion products are formed immediately followina
ignition;
4-2
-------
the thermodynamic properties and species concentration
(except those of nitric oxides and soot) of the "air" and
burning zones are assumed to be in their equilibrium state
based upon the zonal temperature, equivalence ratio and
the pressure.
It should be stressed that the zones above are conceptual and
although their total volume at any instant is that of the cylinder volume, their
physical shape is irrelevant. The mass transfer between each of the zones
is calculated by other relevant model components. Thus, the rate of liquid
fuel vaporization is controlled by the fuel vaporization model, which relates
to the physical process. However, for thermodynamic purposes, once
vaporized all of that fuel is assumed to exist in one hypothetical zone. Fig-
ure 4-1 presents a pictorial representation of the cylinder contents showing
the mass transfer between the various zones. The basic mathematical formula
describing the thermodynamic state of these zones can be expressed by the
following set of conservation equations.
Basic Energy Conservation Equations
Following the first law of thermodynamics, the basic energy con-
servation equations can be written as:
"Air"
mu = -h m,-. — [h m , ] + <_j h, ,m T - P 3 V
a a a u a apo om,i v,i a
- (\ QC + ^ra + <*el) + (hfj - V S (4'1}
The last term represents the pumping work and kinetic energy dissipation
associated with the injection of the fuel into the air which is generally negli-
ble and is included only for self-consistency.
4-3
-------
m
fj
Figure 4-1 Conceptual Zones of Cylinder Contents
4-4
-------
Liquid Fuel (Vaporizing)
Vfl
Liquid Fuel (Burning)
"fib "fib
(4-3)
Fuel Vapor
"Vfg
Burning Systems
- [h
fg
V" - ' V>c * « rfg'
^ ru, r - h m^ T + h,,m,JU T^[hm ,+hrnj (4-5)
bm,I bm,I - a D,I fd fdb,I a apb,I fg fpD,r
Here dot represents the time rate of change in crank angles (i.e. , d/d=) .
The relative zonal volumes of gaseous systems to the total cylinder
volume under uniform pressure, are defined by
4-5
-------
m. R. T.
B
fg PV (4-7)
_ bm,l Dm, I bm,I (4-8)
bm,I pv
The instantaneous total volume of the engine cylinder is computed by:
RL-VRL-RW sin2* , (4-9)
v (e) = v + —-3— D R (i - cos 0
c 4 e w V R
W
The fraction factors for zonal convective heat transfer rate are evaluated by:
(T - T ) B 2/3
= ua w ; Ma
B (4-10)
.2/3
= ( Tfg " V % (4-lD
~
x =
~
(4-12)
* i / ~]
y _ bm, I w/ bm,.
bm, I Q
where normalizations were made by B, which is
3 = (T - T ) £ 2/3 - (T - T ) [3 2/3 + T(T - T ) # (4-13)
ua w; pa ufg Lw> pfg Ubmj V bm,I
and the ~ean cylinder wail temperature T is used in above equations.
-------
Mass Conservation Equation
The law of conservation of mass in each zone is described by
the following mass rate balance equations:
'Air"
Fuel Vapor
Burning System
Liquid Fuel (Vaporizing)
. . • . (4-15)
m,, - m - m - m
fl f] fv fid
Liauid Fuel (Burning)
(4-16)
mflb
m = m. - ni ^ (4-17)
fg fv fpb
mfpb,l + %b,I + m apb,I ~ mv,I (4-18)
4-7
-------
Equivalence Ratio
In order to determine the composition and theromodynamic properties
of each zone, instantaneous equivalence ratios must be calculated. This can be
achieved by maintaining separate accounts for the mass of pure air and the mass
of fuel in each zone. The equivalence ratio of the. "air" zone is governed by
= (m___m, _- m, m )/f
• 2
^_ \»*m »••» ***^ *** I i * rn
a p^a f,a f,a p^a/ os
(4_lg)
where the total flux of fuel and pure air due to the .various rate processes, i.e
vitiation, dilution and combustion can be expressed by:
(4-20)
almv,I -«a(mD+mapb (4-21)
Similarly, the equivalence ratio of the burning zone is governed by:
(4-22)
where
m.
(4-23)
4-3
-------
and T! and « are defined by:
f
(4-24)
(1 + f)
1
oc =
(1 + f)
Caloric Equations of State
The combustion products are assumed to be in thermodynamic and
chemical equilibrium and the properties of the "air" and burning zones can be
calculated based upon the pressure, the local temperature and the local
equivalence ratio. Consequently, it is possible to express the internal energy
and the gas constant of the combustion products and the "air" by the following
general forms, i.e.,
U (P,T, 0) =
1 +f 0
OS
a. T
1=
-1
cal
gm
(4-25)
where
9 =
1000
and f is the stoichiometric air-fuel ratio based on the fuel used, and
OS
4-9
-------
(FI + r20) + exp [(r3 + r^"1 + rg In P)0
(r.
,-1
In P
(4-26)
The polynomial coefficients a , b , d and r. can be obtained, for a
i i / i
particularly specified fuel composition by nonlinear least-square curve fitting
of the calculated results from the NASA equilibrium program (Gordon and
McBride ). The fuel consists of 50 percent n-heptane and 50 percent
toluene and has a H/C ratio of approximately 1.7 and the coefficients which
have been obtained are tabulated in Tables B-l and B-2 in Appendix B for the
range:
300
-------
Over-all Conservation Equations
The overall energy conservation equation is obtained by summing up
equations (4-1) through (4-5), i.e.,
m u m,,u,. m,,, u ... m, u, *r? m, _ u, .
a a fl fl fib fib fg fg I bn,I bm,I
-PV H- hf,mf. -
where Q is the overall radiation loss from all zones, i.e. ,
where
i = a , fg , (bm , I)
and J is the mechanical equivalent of heat.
(4-27)
(4-28)
The over-all mass conservation equation is obtained by summing
equations from (4-14) through (4-18), thus:
m + m-- +m + m. + E m = m (4-29)
a fl fib fg j bm , I fj
The equation of state can be written as
(4-30)
4-11
-------
Working Energy Conservation Equations
Internal energy and mass conservation relationships are expressed
by polynomials, thus the energy conservation equations for the "air" and
burning zones can be solved for the following temperate rate equations:
T = A P + Ba
a a a
(4-3 la)
where
(m
R-T- 8",
a a
a P
- T
a dP
(hf. - ufl) mf.
/C.
m (
a v 6T a a
a
and
Tbm,I Abm,I
(4-3 Ib)
where
/ om,
m,I I f
Ibm,I +
'bm,I
au
bm,I _ T
a p bm
4-12
-------
X
bm,I 1C
bm,I
bm,I\ bm,I
3(Wi . g"bm.r
bm,I
+ (h - h T)
bm,I a bm,F
+ m
.
apb,r
and
(h
fd
mfdb,I + (hfg
C. _ — m,
bm,I bm,I
bm, I
bm,I
bm,I
bm,I
ar
bm,I
bm, I
Similarly for fuel vapor we have
Tr = A, P + B,
fg rg ig
(4-31c)
wnere
Rr T.
fg fg
^
fg
P c <•
vig
r
fg
(hfv - V "fv
mfg ( °vfg
R
fg
If Eqns (4-27) &(4-29) are introduced into the differentiated over-all equation of
•
state, p can be written by
P =
P
V + ]
?"'
«
T, m. + / m
T i 9 + y
1 : a^- i T
>i (Ri
+ T.
1 ^ B
a±v i
V
T
^ +T, -rr1- / i
i
(4-31d)
4-13
-------
where i = a, fg, (bm,I) and the volume rate is expressed by:
V =
"4.
cos 9
\
\
1
Equations (4-3 la) through 4-3 Ic) are the working form of energy conversation
equations.
Convective Heat Transfer
The gross convective heat transfer to the cylinder walls is expressed
(14)
in a similar method to that used by Annand
Qc- KIT
(4-32)
where
* is the mean thermal conductivity of cylinder gas
R is the Reynolds number based on the mean piston velocity and
cylinder bore
T is the average gas temperature
TW is the average cylinder wall temperature
A is the surface area of the cylinder chamber
D is the bore
It should be noted that two empirical constants, A and B , have been intro-
c n ^ 16)
duced whose accepted values range from 0.17 to 0.49 for A , and B =0.7V .
Radiative Heat Transfer
The rate of radiative heat transfer from each zone to the cylinder wall
is expressed by
/ -KLT\
'rbm,!
AI
a 1 - e
. T
w/J d9
(4-33)
4-14
-------
K is a mean attenuation coefficient which will depend upon the concentration
of radiating species (gases and particles) , the number of particles, the particle
cross section and the optical properties of those particles. McAuley gives
2 *?
a value ofKas5.7xlO cm /gm x the zonal fuel concentration. L, and AT
are the characteristic length and surface area of the zone respectively.
Similar expressions are also used for evaluation of the rate of radiative
heat transfer from fuel vapor and "air" zones to the cylinder wall. However,
they are negligible because of the very low zonal temperatures or the very low
fuel concentration.
4 .2 Fuel Injection and Spray Characterization
For the purposes of this work, the fuel jet is treated as an
incompressible jet with a constant flow rate and the injection velocity is
derived by:
Vfjo = V^nozzle^ (4'34)
where m,. the modified fuel flow rate is defined as
mass of fuel injected per power stroke ,. _->
" "nozzle
4-15
-------
It is assumed that the fuel spray breaks into droplets immediately upon
injection and that the droplets have a specified size distribution. The drop-
size distribution is described by an upper-limit distribution function, the
fraction of the total fuel mass, having a dropsize within the interval d.
being given by
Am,
m
tot
erf
5 In
d - d.
^ max i
- erf
(4-36)
The maximum droplet size, d _ , is directly proportional to the
max
nozzle diameter, and is inversely proportional to the product of the Weber and
Reynolds numbers to the 1/4 power, and to the pressure of the cylinder air
charge at the time of injection to the 1/5 power, i.e.,
,1/2
max ~ max nozzle al
1/4 A-3/8 -1/8 1/4 -1/5
'1
'1
(4-37)
where
n
Ap is the pressure difference across the injector orifice in (dyne/cm )
2
a. is the liquid surface tension in (gm/sec )
.
is the fuel viscosity in (gm/cm-sec)
d , 'is the nozzle diameter in (cm)
The two empirical parameters in the above expressions 6 and C
have values of 0.85 and 37.5, respectively, which represent the best fit
to experimental data by T««\17^
max'
4-16
-------
4. 3 Air Spray Entrainment and Droolet Vaporization
The analytical representation of the phenomena associated with the
fuel spray dispersion is perhaps the single most important component of any
model of diesel engine combustion. The complexity of the actual situation,
which involves a heterogeneous, confined unsteady jet subjected to cross
flow defies rigorous mathematical description at present. Consequently,
simplifying assumptions and empirical relationships must be included to allow
the rate of fuel-air mixing and droplet vaporization to be quantified. The
analysis developed for inclusion in the overall model has the following
objectives:
• To describe the gross fuel-air distribution within the fuel jet;
• To track the trajectory of fuel elements of known "age";
• To define the state of a fuel-air parcel at the onset of ignition,
i.e. , the fuel-air vapor ratio, the liquid vapor fuel ratio and
the dropsize distribution.
The gross oversimplification is made in the analysis of the fuel
spray dispersion that the fuel jet can be considered as a steady unconfined
wall jet of air and vaporizing fuel droplets, shown schematically in Figure 4-2 .
The similar transverse velocity profiles are based upon constant injection
conditions throughout the injection period. The steady state properties of the
jet are determined in Eulerian coordinates. Each element of fuel is labeled
upon injection, both with respect to its age (i.e. , crank angle at injection) and
its radial location within the jet. Both the location and the conditions of each
fuel-air parcel are determined in Lagrangian coordinates by integrating the
Lagrangian Eulerian (time space) relationship. The rate of droplet vaporization
is determined by the properties of the surrounding air which will vary, both
in space and time. This approach, which may be termed a "pseudo-steady"
spray model,neglects:
the effect of piston motion upon the spray geometry;
relative motion between the droplets and the air;
4-17
-------
Note: I is equivalent to [(i-1) X (No. of shells) + ]]
Figure 4-2 Schematic of Fuel Spray
transient
region
4-18
-------
the transition zone from a free to a wall jet;
the fuel volume;
the curvature of the piston bowl;
preferential fuel vapor transfer, all vapor remains associated
with the parent fuel parcel;
the influence of confinement both by the piston and cylinder
walls and the adjacent fuel jets.
However, the following phenomena are included in the analysis:
droplet heat-up and vaporization;
jet entrainment due to impingement upon the wall;
the influence of swirl on the jet entrainment rate.
The relationships governing the behavior of an unconfined jet impinging upon
a plane wall in a quiescent atmosphere can be drived from fundamental
(18)
turbulent jet theory (see Abramovich ) and the detailed formulations have
been included in Appendix A. The influence of air swirl induced by the
masked inlet valve on the fuel-air mixing rate is approximated by
m
7—V = 1.0 + C , (v /v-. ) (4-42)
(m ) sef s fjo
3 O
where C , is an adjustable constant obtained by matching experimental data.
(19)
Wise and Agoston proposed a correction to the normal vaporiza-
tion rate to account for liquid droplets in crossflow. A similar expression
has been used to account for the influence of the swirling air motion on
4-19
-------
droplet vaporization. The ratio of the vaporization rate, rh , to the
vaporization rate without swirl (m ) is given by:
1/21/2
RePr (4-43)
where the Reynolds number is based upon both the droplet diameter and the
swirl velocity and it is assumed that Pr = 1 . The empirical constants
Cevapissetat0.276.
The vaporization rate without swirl is given by
Cp (I. - T,)
Under high pressure high temperature conditions the heat of vaporization is a
function of droplet temperature T,:
_, - T, \ 0.38
/ _£rtt
298 IT- 298
where L and T are the latent heat at 298 K and the critical temperature
298 cnt
of the fuel respectively. The droplet temperature is assumed to have a uniform
value equal to the stable- wet-bulb temperature based upon the conditions of
the surroundings at the beginning of the fuel injection. Methods for evaluating
the stable wet-bulb temperature of a vaporizing droplet are discussed in the
recent work of Rosner and Chang ^ . A mean heat-up time T, is used to
account for the time taken for the droplet to attain the stable wet bulb
temperature.
4-20
-------
4.4 Ignition Delay
The objective of the ignition delay model is to provide a relationship
whereby the time of ignition of each fuel/air parcel can be evaluated. The
(2 1)
ignition delay time is calculated using the expression from work of Shipinski
Tig= Aig(Nce)Blg ®°ig exP(E.g/RT) (4.45)
where T. , the time of ignition in seconds, and the pressure and temperature
are the average values calculated over the delay period, in atmospheres and
°K respectively. The ignition delay model includes four empirical constants
A. = 1,224 x 10"4
ig
B. = -0.69
ig
C. = -0.386
E. = 9290 cal/mole
ig
representing the best fit to the experimental data .
4. 5 Air Product Mixing
High speed photographic records were made during Phase I of the
present program (Wilson et ar ') which tended to suggest that after the initial
vapor phase heat release, the contents of the cylinder approach well-stirred
conditions on the macroscale. To accommodate these observations the model
allows mass transfer to take place between "air" and "burning" zones in a
random manner. After ignition of the vapor/air mixture in a fuel/air parcel the
combustion product-droplet-air mixture continues to mix with the air zones, but
not with other fuel/air parcels. Thus, the air zone is being continuously vitiated
and vitiated air is being mixed with the vaporizing droplets . In the light of
recent in-cylinder sampling investigations (see Appendix C) it appears that the
concept of random mixing may not be representative of actual conditions. The
mixing process used in the model does not retain any memory of the jet structure
once ignition has occurred. However, although only preliminary results are
available, sampling valve data suggest that mixing process may be dependent
upon jet properties some time after ignition has taken place.
The air product mixing is modeled as a completely random process,
i.e. , the well-ordered jet is considered to have broken down and the mass
4-21
-------
rate of dilution and vitiation is derived with the following provisions:
only the "air" and burning zones are involved;
turbulence is unsteady but homogeneous throughout the engine cylinder;
the overall dilution rate decreases as the mass of available air
decreases;
overall dilution rate vanishes with the total mass of combustion
products;
the zonal dilution rate depends upon the zonal surface area in contact
with "air" zone. This area is assumed to be proportional to the
zonal volume
Thus, a rate of dilution can be expressed by
- JL maambm
rmix <*+™ (4'46>
where m is the mass of the "air" available for mixing, and is equal to the mass
of unburnt "air" less the mass of "air" mixed with unburnt fuel vapor. An
( 2 )
expression similar to Eqn (4-46) has been used by Bastress et al where
r was input as a arbitrary adjustable constant. Physically,
r in Eqn (4-46) is the characteristic mixing time (measured in crank angles) and
is related to the characteristic turbulent length scale / and the turbulent diffusivity
D by:
the turbulent diffusivity is in turn related both to the scale of turbulence,
i , and the scale of turbulent velocity v , by:
turb
D « i - v
turb
4-22
-------
Finally, a and v . are scaled by the volume surface area ratio and mean flow
velocity.
V cc U = V ,
turb char
The characteristic charge velocity is defined by:
1/2
v
char
K(
/ H" V
P s
»a° +
) + mf. v
m
where v is instantaneous piston velocity and v is the swirl velocity based
upon swirl angular velocity and half bowl radius assuming solid body rotation,
vfio is t*le ^ue-'- injecting velocity.
_L
'mix """ 77
or
mix
RPM x L
C . U
mix
(4-47)
Here empirical constant C is introduced. The zonal dilution rate for each
mix
burning zone is estimated by
dm
D.I =
3bm,I
dm
D
(4-48)
bm,I
The vitiation rate from each burning zone is assumed to be
to the dilution rate and therefore can be evaluated by
directly proportional
dm.
= G
dm.
•vidi
(4-49)
4-23
-------
4.6 Combustion
Combustion may take place within the cylinder in several modes.
Fuel can exist in either the vapor state, or liquid droplets, the air can be
mixed with the fuel prior to reaction or a flame can separate the reactants and
diffusion of both reactants will, therefore, dictate the rate of heat release.
The rate and timing of the heat release will control both engine performance
and pollutant formation. In reality this complex process will involve droplets
burning as clouds, combustion close to solid surfaces, fuel impingement on the
piston bowl vaporization and subsequent combustion. However, the simplified
approach included in this model only takes account of:
• Homogeneous premixed, fuel vapor mixed with the air prior to
ignition
• Heterogeneous, fuel burns as single droplets surrounded by a
diffusion flame in a fuel lean zone.
• Homogeneous diffusion, diffusion flames cannot be supported around
droplets existing in fuel rich zones; immediately upon vaporization
this fuel reacts with its surroundings giving equilibrium products for
the composition and conditions of that zone.
4.6.1 Homogeneous (Premixed) Combustion
Equation 4-45 determines the point of ignition for the premixed
fuel vapor-air produced in each parcel during the preparation to burn period.
The temperature of the mixture after combustion of the fuel vapor is assumed to
be the equilibrium temperature at the equivalence ratio, pressure and specific
internal energy pertaining to that parcel. This assumption will only be valid
when the residence time is greater than 10 seconds since certain radicals
will not have equilibrated before that time. This time scale is slightly less
than one degree C.A. for an engine operating at 1500 rpm.
4-24
-------
4.6.2 Heterogeneous (Droplet) Combustion
When fuel droplets exist in an environment containing oxygen, reactions
can occur in a narrow zone surrounding the droplet-the diffusion flame envelope.
In recent years numerous theoretical and experimental studies have been carried
out to establish the mechanism of combustion of single droplets in an infinite
(22 ,23)
atmosphere (Williams ' ). Due to the complexity of the process, none
of this work can be directly applied to heterogeneous combustion in diesel engines .
The droplets do not exist in an infinite environment and interaction will take plsce
between droplets. The atmosphere is not quiescent thus allowing the formation
of wake flames or producing complete flame blow-off. Calculations indicate
that droplet combustion does not take place as though single droplets existed
in an unbounded environment since the volume necessary to accommodate such
flames was found to be many times the engine displacement. The ultimate
solution will involve some form of unsteady combustion coupled with effects
due to relative motion between the droplet and its surroundings . As a very small
first step, the classical droplet theories have been modified to provide flame
confinement and to allow continuous vitiation of the atmosphere surrounding the
droplets .
The basic assumption involved in the analytical solution for con-
fined pseudo-steady state droplet combustion which has been included in the
model are:
combustion takes place under quasi-steady state conditions;
the rate of chemical reaction involved in heat release are infinity
fast
the pressure throughout the system is uniform
the system is spherically symmetric
average values are used for the specific heat and thermal conductivity
of the gas and liquid phases
Lewis number is unity (i.e., PD = - = constant)
liquid temperatures are constant, the droplet temperature is given
by the stable wet-bulb temperature of vaporization thus temperature
transients in the liquid phase are neglected
4-25
-------
radiant heat transfer is negligible
single component fuel composition
body forces and internal motion within the droplet are negligible
the droplets exist in a large but finite space which is continuously
vitiated during the entire droplet lifetime.
convective effects are neglected, the diffusion flame limit depends
solely on ambient oxygen concentration.
The governing differential equations of mass, species and energy conservation
are written as (Williams ^ 23 ^ )
Continuity (4-50)
7? ~o7"~ °"2 PV) *°
Species
pv<*i- 1 d (PDr2^ +w (4"S1)
dr r ~:Tr~
Energy
The equation of state is
PT = Constant (4-53)
The boundary conditions at the droplet surface r , the flame surface rf, and the
assumed radius of influence r are presented below.
4-26
-------
(i) at the droplet surface (r = r .)
jL
T = T4 (4-54)
YQ2 = 0 (4-55)
dY
- (4~56)
Condition (4-56) is derived from species continuity at the liquid-gas interface.
Condition (4-57) is deduced from an energy balance assuming a constant
liquid phase temperature.
(ii) at the flame surface (r = rj
Y02 = ° (4"58)
Yf = 0 " (4-59)
= ( P D
Condition (5-60) states that the deliveries of fuel vapor and oxidant to the flame
surface are in stoichiometric proportion. Condition (5-61) prescribes that the
heat of combustion is equal to the sum of the heat transferred inward to the
droplet and the heat transfer outward to its surroundings.
4-27
-------
(iii) at a large distance from droplet (r = rj
T - T (4-63)
90
Y - Y (4-64)
02 XO2,«
Yp - 0 (4-65)
Y -Y (4-66)
N2
Y -Y (4-67)
NO *NO,«
Analytical Solutions
Analytical integration of the governing differential will be described in
this section.
If we let
and
C T
Zm = - E
T (v^ - yfT) Wf q
where q is the heat of combustion (in cal/gm) and v's are the stoichiometric
coefficients for a single stem chemical reaction
Sv.
Equations (4-51) and (4-52) become
dZ;
_i
r
pv'dr - 2 dr
4-28
-------
where w is defined by w. = w(v," - \j.')W. . If we further specify that
or
3 =- Z. H- ZT
we have
M. dg _ 1 d / 2 da
2 dr 2 dr lr ~3F
where
M_ _ jv
_2 ~ 3D
Integrating the above equation once, with respect to r, we have
ds = °1 -M/r
dr "T" e
Further integration gives the general solution as
, = £1 e-lvl/r + c
d M e U2 (4-68)
here C, and C2 are the constants of integration, to be determined by imposing
appropriate conditions in accordance with the assigned definition of 3. The
important relations deduced from the above general solution are explained
briefly in the following.
Mass Burning Rate
The expression for mass burning rate, which is obtained by defining
3 s ZT + Z^y and utilizing the boundary conditions at the droplet surface and
the far field,can be written as
r
mf = 4?r pDr^ ( -f~- rj In ( 1 + BZ) (4-59)
4-29
-------
Flame Radius
The location of flame surface can be identified by defining 3
and making use of boundary conditions at r^ and r^, thus,
r ~ r1
OQ 1
OC -
(4-70)
where
B4 " 1 - Y
— PC ( T - T )
L L p ^ « V
fl
02,<
f-1
In. (1 +B2)
L n (1 + B4)
] = Stoichiometric Oxygen-fuel Mass Ratio
q = Heat of Combustion
Flame Temperature
In a similar manner, the temperature at the flame surface is obtained
from
fl
'fi
(4-71)
1 +
•02," /J
Additional information, such as major species and temperature distributions
around a fuel droplet, are also required when it becomes necessary to evaluate
the production of nitric oxide in these diffusion flames. These relations.
4-30
-------
which can be obtained from methods similar to those described above, are
listed as follows without further explanation.
Temperature Distribution
(T, - T J (e - - - - e
f T = Tf +
(e
(T - T
v £ 05
- e
e
r < r
r > r.
(4-72)
(4-73)
Fuel Distribution
Y
Yf =
fl
and
'1 + Y
Y
02 ,
f,l
Oxygen Distribution
Y
, -M/r
O2, » e - e
•02
r > r.
(4-74)
(4-75)
Nitrogen Distribution
Y
-M/r -M/r,
- e
N2
, -M/r.
(e 1 - e
(4-76)
r, < r < r
i CO
where nondimensional burning rate is defined by
"
(4-77)
4-31
-------
fraction at the droplet surface can be calculated by
and the nitrogen mass
i . v (4-78)
1 *
YN2,1
* /X
where
44
X= --
it is assumed that the only species present at the droplet surface are fuel
vapor and the products of complete combustion, w represents the number of
carbon atoms in a fuel molecule and x is the number of hydrogen atoms.
Determination of r^ in the Finite-Space Formulation
The solution for the bounded single droplet burning model requires the
determination of a distance r.y, for which the boundary conditions specified
in equation 4-63 though 4-67 are applicable. This distance is evaluated by
invoking the zonal oxygen concentration relation
r P v *
max oo O2,« J O2
rf
where rfflax is a function of equivalent droplet center-to-center spacing. It is
assumed that the volume of each zone is apportioned equally between all the
droplets and rmav is evaluated by
/ V \ 1/3
rmax' °'23873 -5H
where N is the total number of droplets in "I" th zone or more precisely:
r = 0.62 D,
max d
where D , is droplet center-to-center spacing
4-32
-------
Combining Eqn (4-69) and (4-71) and assuming an ideal gas Eqn (4-76)
can be rewritten as:
-M/rr -M/r 2
/ -ivi/ r.. -ivi/ r c. .
, r ( e f - e ) r dr
•j /^ max
r
J (e-M/rf _ e-^/r) + i_ (e^A - e'
Tco (4-80)
In the above equation, Tf is readily determined from Eqn (4-70) and r, is an
implicit function of r^, thus Eqn (4-80) must be solved for r^ by graphic or trial
and error methods.
A schematic of a droplet diffusion flame in a confined space is shown in
Figure 4-3.
4-33
-------
a--3 Schematic of droplet diffusion Flame in a Confined
4-34
-------
4.7 Nitric Oxide Formation
The rate of the chemical reactions involved in the formation and
destruction of nitric oxide are extremely temperature sensitive. Under con-
ditions typical of normal diesel engine operation, characteristic residence
times are small compared to the times necessary to achieve equilibrium.
Consequently, the amount of nitric oxide formed will depend upon the time-
temperature-pressure history of the reactants. If heat release took place in
one homogeneous zone, temperatures would be too low to allow the formation
of significant quantities of NO because of the extremely lean overall opera-
ting conditions. Thus, it becomes apparent that nitric oxide is formed in
zones whose characteristic size is much less than the characteristic cylinder
size and that local homogenity will exist on various scales within those
zones.
Two zones of NO formation are identified in the model depending
upon the scale of local homogenity:
• Reactants can be well mixed on a scale which is small compared
to the total reacting volume and local gradients can be neglected.
These well mixed zones will have a wide distribution of equiva-
lence ratio. The rate of NO formation in regions of this type are
calculated by the homogeneous NO equations.
• Reactants are separate and the reaction zone is characterized by
steep gradients and molecular diffusion plays a strong role in
NO formation in these regions. Reaction zones of this type can
typically exist around a single droplet, a cluster of droplets or
at the boundary of a fuel and air zone. In this model NO produced
in diffusion flames is assumed to form around single fuel droplets
and is calculated by the heterogeneous NO equations.
4-35
-------
4.7.1
Homogeneous NO Formation
The calculation of the rate of homogeneous NO formation is based upon
the Zeldovich mechanism^ ' and assumes complete combustion of the reactants
before NO formation begins and also that oxygen atoms are equilibrated with
oxygen molecules. The governing kinetic equations are:
(1)
O + N,
NO + N
(4-82)
(2)
(3)
N
ro]/ roj 1/2 = K
NO + O
(4-83)
(4-84)
eq3
The literature values of the forward and backward rate constants for these reactions
are tabulated in Table 4-1 . The instantaneous NO production rate (gm/crn - sec)
in any zone can be determined by
30
(*>
iHom
0111,1
K
bm,
L/2 1/2
X02
1-K2(XNOXNO/XN2X02)1
1 + K3 (X-NO/02}
(Cnor)
(4-8S)
Where M is the averaged zonal molecular weight and
bm, I
4-36
-------
K. is equal to 0.22078 KrlK ,
-1 tl eq o
K2 is equal to
is equal to K /Kf
Table 4-1 Literature Values of NO Rate Constants
Rate
Constants
Kfl
Si
Kf2
K
b2
Keq 3
*
Values
1.36 x 1014 exp (-75,400/RT)
3.12 x 1013 exp (-400/RT)
1.33 x 10 10 T exp (-7,080/RT)
3.2 x 109 T exp (-39,100/RT)
5.0 exp (-58,300/RT)
Units
3 ,
cm /mole-sec
cm /mole-sec
cm /mole-sec
3 .
cm /mole-sec
/mole \
Reference
Baulch'25'
Baulch125'
Wray & Teare
(26)
Wray & Teare
(26)
Caretto(27)
*R = 1.986 cal/mole- K
T = Temperature in °K
The mass involved in the NO formation process in diffusion flames
surrounding the burning droplets is evaluated by mass integration over flame zone:
m
df
-E4-T/
2
P r dr
(4-86
4-37
-------
which has been discounted from the homogeneous formation process shown in
Equation (4-85). The summation in above equation is made over all droplets
involved in this zone.
The nitrogen bound in the fuel which will be responsible for a very
small amount of NO has not been considered in this analysis.
4.7.2 Heterogeneous NO Formation
The nature of the heat release process in a diesel engine dictates
that the majority of fuel droplets will burn in vitiated air which also contains
nitric oxide. At high NO concentrations the net rate of NO production is
strongly dependent upon the concentration. Consequently, it is important
that the concentration profile surrounding the burning droplet is known. The
analytical solution described in Section 4.5.2 is used to define both the
temperature and concentration field around the droplet and the NO kinetics
discussed earlier are assumed to be valid, thus a local rate of NO formation
can be defined. Consequently, only the following NO mass conservation
equations must be solved, provided that it can be assumed the formation of
NO has no influence on the flow field surrounding the droplet
dY \
N0\
— + w 4-87
dr / NO
The local NO production rate can be conveniently expressed in terms of
species mass fraction by:
1 1 - 0.9953
J
/^i K*^
/ ol bZ
K K
\Kfl Kf2 /
1+ 1.142857 ( -^
1 \%2/
v 2
\ N0 1
/ Y Y
' N2 "02
YNC
Y02 J
(4-88)
4-38
-------
and the two-point boundary conditions are specified by
(4-89)
YNO
In the above equations, Y and Y and T are already known explicitly from the
relationships described in Section 4.5.2. The solution to the above two point
boundary value problem can be achieved by following a standard "trial and error"
procedure
guessing the initial condition, i.e. , Y.JQ i at the droplet surface
integrating Equation (4-87) numerically throughout the entire flow
field from r, to r^ based on known Yf, Y-^ and T distributions.
matching (Y )(calculated) to (YNO^ ) (specified)
repeating the procedure until agreement is reached
The contribution of the NO formed in the droplet diffusion flame to the overall zonal
•
production rate is calculated by integrating the NO production rate, w ,
over the diffusion flame zone, i.e.
r \2
= 47T r " I w.._ |-
NO/HET
4-39
-------
4.8 Soot Formation and Oxidation
Little is known concerning the detailed mechanism of soot formation
and the model of soot formation is based upon empirical relationships derived
from experimental observations. It is believed that the following processes
are involved in the formation of soot:
dehydrogeneration: breakdown of hydrocarbon molecules to form
molecules with a higher C/H ratio
polymerization: formation of higher molecular weight molecules
agglomeration: formation of solid particules
The most favorable conditions for soot production in diesel engines
are those high temperature zones deficient in oxygen. These conditions are
produced during the ignition delay period or during the early stages of combustion.
An Arrhenias type equation for the rate of soot formation has been suggested
(4)
by Khan and Greeves which is based upon the local unburnt equivalence
ratio, partial pressure, temperature and includes several empirically determined
coefficients. The Khan and Greeves relationship is not directly applicable
to this study due to the singularities in defining the zonal unbumt equivalence
ratio (when it is based solely upon unreacted oxygen) for a fuel rich zone.
A modified rate equation based upon local equivalence ratio of the
burned products their pressure and temperature has been used in this study is
(4-92)
The activation energy Egf is taken to be 40 kcal/mole. P is pressure in atm.
T is averaged temperature of cylinder contents and the pre-constant A is fitted
sf
to exhaust soot data.
The mass rate of oxidation used is the expression proposed by Lee
etal.( 28 }:
- (Eso/RT)
4-40
-------
Where P is in atm, and A has a value of 6.51 x 10 . The averaged soot density
3
ps, and particukfte diameter, dg, are assumed to be 2.0 gm/cm and In respec-
tively. The activation energy Ee is set equal to 39.3 Kcal/mole (also see
(7 Q\ ^
Field eta!U3;).
4.9 Nitric Oxide and Soot Formation Equations
The total zonal NO production rate is the algebric sum of the homogeneous and
heterogeneous contribution:
HET
(4-94)
The summation in above equation is made over all droplets.
The species conservation equations for NO in the "air" and burning zones are
written as:
"Air"
1
m
} -(Y ) 1 m
NO;bm,I UNO;a V, I
NOa
(4-95)
Burning Zone
bin, I
(Y ) - (Y ) (m_ . + m )
v 1"'V NO J D,I apo,I
a bm,I
(4-96)
4-41
-------
The net rate of soot formation is
w = w , - w
s sf so
(4-97)
The species conservation equations for soot in the "air" and the burning zone
m
EI K'bm.I
m
(4-98)
Burning Zones
"Am.:
m
bm,I
(Ys>a
)
4.10
(4-99)
u T £JU T - U T U
'bm,I fdb,I ipb,r s bm,I
Equilibrium Compositions
The equilibrium species distributions (in mole fractions) of combustion
products are computed by means of a computer program developed by the NASA
Lewis Research Laboratory. For fuel-lean mixtures, eleven species including
CO2/ CO, H2O, H2, NO, OH, N2, O2/ H, N and O are considered and CH4 is
added for fuel-rich mixtures. The Argon content of the air is neglected.
For a given fuel, the species distribution are calculated as functions
of three parameters: pressure, temperature and equivalence ratio. For com-
putational efficiency, a set of species distribution tables, covering the pressure,
temperature and equivalence ratio ranges of normal engine operation, are cal-
culated and stored in magnetic tape. Consequently, a 3-D table look-up is the
only operation needed whenever equilibrium composition, for specified pressure,
temperature and equivalence ratio, is required.
4-42
-------
4.11 Fuel Representation
The laboratory report on the diesel fuel used in this study gave the
following properties:
Table 4-2
H/C (by mole, approx)
Gravity, °API
Total sulfur, wt %
Flash point, F
Pour point, F
Cloud point, F
Cetane number
Viscosity, CS @ 100°F
Aroma tics, %
Distillation
Initial boiling (°F)
50% point (°F)
End point (°F)
Residue
Total nitrogen, wt %
1.7
34.5
0.3
170
20
24
46.7
2.5
37.6
386
514
658
1%
< 0.05
The model representation of this diesel fuel is a mixture of 50% (by volume)
n-heptane (C_H, J and 50% toluene (C_H0) . The sulfur and nitrogen content of
7 lo / o
the fuel are neglected. This fuel model is used in the parametric calculation of
the equilibrium thermodynamic properties and species distribution of the combustion
products.
The extended properties of the fuel mixture based upon the properties
of the fuel model constituents are listed in the table.
4-43
-------
Table 4-3
Molecular wt. 96.
Critical pressure (atm) 34.3
Critical temperature (°K) 567.0
Density (at 298°K) 0.851
Boiling point (1 atm) 377.6
Heat of Combustion (cal/gm) 10249.
4.12
The Method of Solution
The temperature of each gaseous zone is determined 'by integrating
the energy equations (4-31a) through (4-31c). Since the liquid fuel temperature
is frozen at the initial wet-bulb temperature, the energy equations (4-2) and
(4-3) can be used to determine the energy transfer, for phase transition during
both the fuel vaporization and the combustion periods. The internal energy is
determined from the temperature, pressure and equivalence ratio. The initial
temperature of a burning zone is calculated using the Newton-Ralphson method,
based on the known internal energies of fuel and "air" before ignition, instan-
taneous heat losses and pressure.
The rate of heat transfer is described by Eqns (4-32) and (4-33). The
mass transfer between zones is described by Eqns. (4-35). (4-42), (4-43), (4-48) and
(4-49). The total mass of the individual zones is calculated by integrating Eqns
(4-14) through (4-18). The instantaneous cylinder pressure is then determined
from over-all equations of state. The over-all energy balance equation (4-27) and
mass balance equation (4-29) are used to check the accuracy of the computational
procedures at the end of each numerical integration step.
The pollutant zonal mass fractions are obtained by integrating the species
conser/ation Eqns. (4-95) and (4-96) for NO concentration and by integrating
4-44
-------
Eqns (4-98) and (4-99) for soot.
The governing differential equations in the analytical model of
diesel engine combustion are all first order differential equations and can be
represented symbolically as:
(4-100)
The symbolic variables Y. are functions of crank angle 6. Numerically
this first order system could be integrated by various finite difference techniques.
For accuracy and computational efficiency, the modified iterative, second order
Runge-Kutta method has been employed. The recurrence formula for the finite
difference equivalence of eqn (4-100) can be written as:
Y. n + * (0+40) = Yi (0) +
(4-101)
+ f IX n (9 + 40),0 +
where n is the index of integration and the initial (n=0) approximations are
evaluated by
= Y. (0) +A0 • f(Y ,0) (4-102)
At each integration step the iteration is continued until a stable pressure and
mass averaged temperature is obtained.
4.13 Indicated Engine Performance and Engine Emission
The engine performance parameters, including indicated mean effective
pressure, indicated horsepower and indicated specific fuel consumption are
calculated by the following relationship:
4-45
-------
Indicated Mean Effective Pressure (IMEP)
IMEP =
JTD 2 R
T
-180
5vc
ISO
(4-103|
where Q and
are the intake valve closing angle and exhaust valve opening
angle respectively. The sum of the first term and last term on right hand side
of Eqn (4-103) is small and can be ignored when comparing engines of similar
manifold pressures and valve timings.
Indicated Horsepower (IHP)
The total horsepower developed in the engine is calculated by:
(IMEP) (V ) (RPM)
396000 xN
(4-104)
Indicated Specific Fuel Consumption (ISFC)
The indicated specific fuel consumption given by:
ISFC
(60) (mf1) (RPM)
N (IHP)
(gm/hp-hr)
(4-105)
is a comparative parameter that is indicative of the efficiency of an engine
in converting chemical energy into work and N = 2 for four stroke engine, N
= 1 for two stroke engine.
4-46
-------
5 . 0 THE EMPIRICAL NATURE OF THE MODEL
One of the major objectives of this program was to develop a model
describing the formation of pollutants in a compression ignition engine which
contained the minimum number of arbitrarily assigned empirical constants.
The previous section indicated that limitations in the fundamental knowledge of
many of the physical processes described in the model necessitated that several
empirical constants be included. Although the physical significance of the
empirical constants included in the Ultra systems' model has been discussed
in the previous section, it is necessary to summarize the empirical constants
used in the model and to ascertain the sensitivity of predicted engine per-
formance to the value of those constants.
5 .1 The Nature of the Empirical Constants
The empirical constants included in the model have been divided into
two groups and the basis for this division is:
the level of uncertainty associated with their numerical value;
their importance with respect to the prediction of engine performance;
the availability of experimental data to allow re-evaluation.
The first group of constants are considered to have nonadjustable values for
the purpose of this study and are listed in Table 5-1. Although it is recog-
nized that several of the constants listed in Table 5-1 cannot be considered
as well established (e.g., Es activation energy for soot formation), some
bounds must be placed on the number of "adjustable" constants. Consequently,
those constants included in Table 5-1 were selected because:
they had values which were generally recognized and had been
established from acceptable experimental evidence (e.g. , C
Cb'Cmax' 6'Eig'Big'CigandCevap);
they had been used previously and no new data was available
to necessitate re-evaluation (e.g. , pD, AgQ, E&, Cfj and GW.);
they were considered to have a minor influence on the calcula-
tion of engine performance (e.g. , Cf.).
5-1
-------
Table 5-1 Nonadjustable Empirical Constants
I Constant
1
1
1
Ol 1
1 1
1
C
a
Cb
po
C
evap
C
max
t
C..
fjo
c
fj
Cw,
E
ig
ig
c
ig
E
8
A
so
Description
I're-constant In Annand type of convectlve
heat loss expression Eqn (4-32)
Tower for Reynolds number dependent
In Annand type of convectlve heat
less expression, Eqn (4- 32)
Product of density and mass diffusion
coefficient used as a constant In
droplet diffusion flame model
Convectlve con reel Ion on droplet
vaporization, Eqn (4-43 )
See Eqn (4-3?)
See Eqn (4-36)
Initial free Jet spread
coefficient, Eqn (A-l )
Free Jet main region, spread
coefficient, Cqn (A-4 )
Wall jet spread coefficient
Cqn (A- 14 )
Activation energy for Ignition
reaction, Cqn (4-45 )
Centane number dependence for
Ignition delay, Eqn (4-45 )
Pressure dependence for Ignition
delay, Eqn (4-45)
Activation energy for soot
formation reaction, Eqn (4-92)
Pre-constant for soot oxidation
reaction, Eqn (4-93)
Literature
Value
0.17
0.7
3. Ox 10"4
g -cm/sec
0.276
37.5
O.B5
0.135
•
0.175
0.204
9290 . cal/
g-mole
-0.69
-0.386
40 kcal/
g-mole
6.5x 104
References
McAulay et all"'
•
S,tke,<3°>
McAulay et al11*'
Dracco^3 ''
Wlse(l9)
Ranz et al* '
Frossllng^33'
Lee
Lee<17>
(34)
Newman
Newman^34*
Gauntner et al' '
(21)
. Shlplnskl et al
Shlplnskl et al' '
(21)
Shlplnskl et al* '
Khan & Greeves
Lee et al<2fl>
Field et al'29'
Remark
.Iterature value
sed
Literature value
used
.Iterature value
used
Literature value
used
Value obtained via
curve fit to Lee's data
Value obtained via
curve fit to Lee's data
Literature value used
Literature value used
Derived from Gauntner' s
results
Literature value used
Literature value used
Literature value used
Literature voluo used
Literature value used
-------
Those empirical constants which have been termed adjustable are those which:
• have a high level of uncertainty associated with them, and are
considered to have a significant effect upon the prediction of
pollutant formation and engine performance.
• are peculiar to the present model.
The value of these constants will depend, to some degree, upon the characteris-
tics of the engine being simulated. However, their value can be ascertained
within limits from single cylinder experiments and exploratory model calcula-
tions. Seven constants have been placed in the second group which includes:
C . C r (Equation 4-42) allows the rate of air entrainment by the jet
to be modified to take account of cross flow effects. This parameter has a
significant influence on the early stages of combustion, which is mainly
premixed, and is therefore, responsible for the rapid pressure rise immedia-
tely following ignition.
C . The parameter C . accounts for the "random mixing" process that
m L2C mix
is assumed to occur between burned zones, which will contain varying amounts
of liquid and gaseous fuel, and the "air" zone. Thus, C , will have a strong
m IX
influence on the processes occurring during the later stages of heat release.
Variations in C . primarily affect the predicted cylinder pressure after ignition,
mix
and the value of C . is chosen primarily to give a fit to the measured pressure
mix >
curve and secondarily, to match observed exhaust NO levels.
A. A. allows the ignition delay to be predicted correctly. Values for
ig ig
A. can be found in the literature, but these cannot be considered universal
ig
since the ignition delay depends upon several operational parameters such as
nozzle type, chamber configuration, etc. Therefore, it is considered desirable
to tune A. for a particular engine configuration.
A ,- The model does not attempt to include the detailed kinetics of soot
si
formation, and it is considered that there is a large degree of uncertainty
attached to the parameter A , which is obtained by matching the predicted and
measured soot concentration levels.
5-3
-------
C , The parameter C ... has been included to provide some vitiation
vidi vidi
of the air prior to mixing with the burning zones during the air/product mixing
processes. The value of C .,. influences both the droplet burning rate and
the early stages of NO formation. The suggested value is somewhat arbitrary.
However, its influence on the prediction of performance and pollutant formation
is concluded to be minimal.
rh The droplet heat up time is a function of droplet size, convective
transfer and the surrounding temperature. Detailed modeling of transient heat
transfer effects within each individual droplet is considered to be outside the
scope of this study and must be justified before consideration is given to its
inclusion. A uniform value, which is approximately equal to the physical
ignition delay (0.2 ms), has been chosen for all droplet sizes. It has been
assumed that vaporization of liquid fuel does not occur during the heat up time.
Cnor It has been well established that it is not possible to decouple the
kinetics of NO formation from those of hydrocarbon combustion. Super-
equilibrium radical concentrations are known to occur in reacting regions
and elementary reactions other than those described by Equations 4-83
through 4-84 take part in NO formation. Consequently, it is reasonable to
expect that any NO prediction which is based solely upon the Zeldovich model
would be an underprediction in those zones where combustion is taking place.
A value of CnQr = 5.37 has been used previously by Kahn and Greeves^ 4 ' to
allow exhaust NO levels to be calculated using the Zeldovich mechanism.
Table 5-2 lists the seven empirical constants which are considered to be
adjustable, together with their suggested values and the method used to
determine this value.
5-4
-------
Table 5-2. Adjustable Empirical Constants
Constants
Csef
Cmix
nor
Aig
Asf
Cvidi
Thu
Equations
4-42
4-47
4-85
4-88
4-45
4-92
4-49
—
Suggested
Value
4.3
7.8E-3
5.37
1.47E-4
3.52
0.1
0.2 ms
Remarks
Correlate to pressure-crank angle data
in initial stage (premixed) combustion
Correlate to pressure -crank angle
data in the final stage of combustion
Match the measured exhaust NO value.
Same value has been used by Khan
and Greeves(4 )
Match the measured ignition delay
data
Match the measured exhaust soot
concentration
—
Droplet heat up time, assumed to be
equal to typical physical delay time
5.2
Exploratory Computations
Engine performance data and exhaust pollution levels obtained from
a series of single cylinder experiments were used to evaluate the values of the
constants listed in Table 5-2. The direct injection single cylinder engine
had the following characteristics:
2340 cm3
13 .97 cm
15.24 cm
Displacement
Bore
Stroke
Connecting rod
Piston bowl/bore
Fuel nozzle type
Nominal orifice size
Cone angle
30.48 cm
0.59
Roosa pencil injector
0.0254 cm
160°
5-5
-------
The values of the adjustable constants used in later calculations
were determined by fitting the predictions to measurements obtained when the
engine was operated under the following conditions (baseline case):
Speed
Load
Compression ratio
Swirl
Fuel injection timing
Injection duration
Air pressure
Mean Wall temperature
1500 rpm
equivalence ratio 0.6
17:1
Medium
20° BTDC
17° CA
30 in. Hg. absolute
500° K
Initially the model was tuned using the baseline case to give an
acceptable fit of the measured and calculated cylinder pressure as a function
of crank angle. The computed pressure curve for these engine conditions is
presented in Figure 5-1, together with the measured pressure curve. In order
to establish confidence in both the model and the values chosen for the
empirical constants, further computations were made with injection times of
-25° and -12°. These computations were made without subsequent adjustment
of any of the constants. The resulting calculations can be compared with mea-
surement in Figure 5-2 and it can be seen that there is excellent agreement
instantaneous cylinder averaged NO concentrations for all three injection
timings. Included on Figure 5-3 are the measured exhaust NO values for
A
these conditions. The calculated averaged instantaneous soot concentration
(in gm/gm-of-fuel) for all three cases are shown in Figure 5-4 and the calculated
engine performance parameters are listed in Table 5-3.
Table 5-3
Calculated Engine Performance Parameters
8ln)
Indicated Mean Effective Pressure (psi)
Indicated Horsepower (hp)
Indicated Specific Fuel Consumption (Ib/hp-hr)
-12
83.3
22.5
0.441
-20
82.2
22.1
0.446
-28
74.25
20.05
0.494
5-6
-------
en
I
c
*j
90
60
70
60
50
40
30
20
10
CalcuL
Measur sd
ted
-60 -50 -40 -30 -20 -10 TDC 10 20 30 40 50
60
Figure 5-1 Calculated and Measured Pressure-Crank Angle History for Baseline Case
= 0.6 Speed - 1500 rpm 0 = 20 and Medium Swirl)
-------
'JO
CM
I
uo
6
4-J
JO
a.
70
60
50
30
20
10
Calculated
Measured
50 -40 -30 -20 -10 0
-60
Figure 5-2 Calculated and Measured Pressure Crank Angle Histories of Cases
(ein, = -28° and -12°. 0 = O.6. speed = 15OO ppm)
-------
HUUU
[NOjppm
tt\e\n
ouuu
2000
1000
/
i
/
1
,
i
/
i
i
/
i
I
t
1
i
i
1
/ /
s S
X
/
/
/ Experime
/
/
i
i
/
•
/
/
/
f
fl
1 /
I i
1 1
I i
1 1
j f
I .
/ '
/ '1
/ /
' 1
y \
Cor
(]
&[nj = -28°CA
^ |
ntal Data ;
Phase
—
Spread Bet
I and Pha
dlnj = -20°CA
i
__^*~*~^~
in] — —
s**~~ " ~~~~
12°CA
- — — — —
Ixhaust
centra tion
•ha sell)
3S22
veen
se II
2387
2050
1293
-30 -20 -10 TDC 10 20 30 40
e
Figure 5-3 Comparison of Calculated Instantaneous NO with
Measured Exhaust Levels for Various Injection Timing
5-9
-------
10
-2
10
-3
4
E
01
o
0
CO
10
10
-5
-30
5-4 Calculated Instantaneous Value of Soot (in gm/gm of
total fuel injection)
5-10
-------
The engine performance is, of course, strongly linked to the rate
of fuel-air mixing and the state of the fuel, i.e. , liquid or vapor. Figure 5-5
presents the distribution of fuel between these states from the onset of injec-
tion (-20°) to almost complete combustion (+60°). Fuel can exist as:
unbumed liquid fuel, i.e., that liquid fuel which is heating up
but not vaporizing or which is vaporizing;
unburned fuel in the vapor state which is composed mainly of fuel
vaporized before the onset of ignition;
burning liquid fuel, i.e., those droplets contained in burning zones.
Figure 5-5 shows that initially the mass of unburned liquid fuel is equal to the
mass of fuel injected. During this period the droplets are being heated, but
are not vaporizing. Before ignition the mass of fuel vapor increases, but it
is rapidly burned out once ignition is initiated (at 9 = -13°). From this time
on the mass of unburned liquid fuel also decreases. The plateau in the liquid
fuel curve represents the constant droplet heat-up time. After all the premixed
vapor has burned, the conditions are such that fuel can only exist as a droplet
in a burning zones or as unburned liquid fuel.
Elements of fuel are tagged by two indexes upon injection, reflecting
their crank angle at birth, and their location with respect to the fuel spray axis.
After ignition only a single index is used, reflecting age since the location is
irrelevant in the random mixing process. The history of two such zones is given
in Figure 5-6. The calculation of equivalence ratio 0 does not include fuel in
the liquid state, only vapor and products are considered. Thus, as fuel either
evaporates or bums, the equivalence ratio increases and then decreases by
dilution. The temperature of the lean zone increases from the onset of ignition
because fuel is able to bum; however, the temperature of the rich zone hardly
varies because of the heat requirements associated with vaporization. In the
extremely rich zone virtually no NO is produced because temperatures are too
low by the time oxygen becomes available, and the NO that is found in this
zone is due mainly to dilution. This highlights one of the weaknesses of the
model which does not allow for NO formation by Fenimore type mechanisms .
5-11
-------
en
I
1.0
0.8
c
o
W
M
a)
0.4
0.2
-20
r
(m,./m?.) Overall Fuel Injection
(m,./m?.) Unturned Liquid Fuel
/m?.) Burning Liquid Fuel
(m, /m?.) Unburned Fuel Vapor
Overall Fuel Burned
-10 TDC 10 20 30 40 50
Figure 5-5. Distribution of Fuel in Different States (Baseline Operation)
60
-------
10,000
8,000
6,000
E
a
a.
4,000
Fuel Rich Zone (No. 21)
Fuel Lean Zone (No. 24) -
2,000
-10
NO
2.4
2.2
2.0
1.8 -
1.6. -I
1.4 -9-
1.2
1.0
0.8
0.6
3,000
2,500
2,000 |-
1,500
1,000
0.4
10'
20
30
40
50
60
Figure 5-6. Instantaneous Zonal Temperatures, Equilvalence Ratios
and NO Mole Fraction for a Typical Fuel Rich and a
Typical Fuel Lean Zone
5-13
-------
However, it can be seen that there is a very rapid rate of production of NO in
the fuel lean zone. Figure 5-7 presents similar comparisons for the instantane-
ous zonal mass of liquid fuel and soot for the same two zones. In contrast to
the NO production, soot is hardly formed in the lean zone, emphasizing that
which is well known, i.e., conditions which reduce NO formation tend to
enhance soot formation.
5.3
Sensitivity Analysis
A sensitivity analysis has been carried out to assess the influence
of the adjustable constants on the calculated emission levels. The range
covered in this analysis for each constant is given in Table 5-4 and the results
of the computations are presented graphically in Figures 5-8 and 5-9 for nitric
oxide and soot, respectively. Both the calculated emissions and the empirical
constants have been normalized by the baseline values.
Table 5-4. Values of Empirical Constants for Sensitivity Analysis
Case
1
2
3
4
5
Variable
Csef
mix
C
nor
Cvidi
Thu
Baseline
4.3
7.8E-3
5.37
0.1
2.0 x 10'4
1
2.15
3.9E-3
1.0
0.0
1.0 x 10'4
2
8.6
1.56E-2
8.0
0.2
3.0 x 10"4
5-14
-------
I x 10
-1 _
1 x 10
-2
ta
in
a
2
1 x 10
-3
1 x 10
-4
— — Fuel Lean Zone (No. 24)
Fuel Rich Zone (No. 21)
10
10
-4
Q
u
2
en
ta
*
O
O
CO
co
10
-5
-10
10
20
30
40
SO
10-
-o
60
Figure 5-7. Instantaneous Zonal Masses, Masses of Burning Fuel
and Soot Concentrations for Typical Fuel Rich and
Fuel Lean Zones
5-15
-------
2.0
1.8
1.6
1.4
1.2
XNO
o
1.0
0.8
0.6
0.4
0.2
0.0
Variables
Csef
C
mix
nor
Cvidi
THU
Co or To
4.3
7.8 E-3
5.37
0.1
2.0 E-3
Line
2077 ppm
0.0
1-0
C/C0 or r/r
3.0
"igure 5-8 Effect of Empirical Constants on I-'C Prediction
5-16-
-------
2.0
1.8
1.6
1.4
1.2
o
o
n
\ 1.0
4^
§
0.8
0.6
0.4
0.2
Variables C or r Line
o o
- Csef 4.3
Cmix 7'8E-3
— Cvldl
2.0 E-3
(soot)Q =139 mg/m3
0 o
Figure 5-9 Effect of Empirical Constants on Soot Prediction
5-17
-------
It was apparent from the calculations concerning NO formation in
lean versus rich zones that the major "valve" on calculated NO is the equiva-
lence ratio of each zone. Thus, it is not surprising that the constants which
influence mixing, strongly affect the NO calculations. It is, however, sur-
prising that the strongest influence is provided by Cmix which is associated
with the mass transfer processes which occur after ignition, rather than Csef
which controls entrainment by the fuel jet. This can be attributed to the fact
that the calculated peak temperatures for lean mixtures occur close to TDC and
the mass transfer rates between the air zone and burning zones are higher,
thus producing more lean burning zones.
T, has a stronger influence on soot emissions than on NO. Also,
the relative influence of C - and C . on soot production is different from that
set mix
on NO. Increased mixing after ignition tends to decrease soot emissions.
Table 5-5 lists predicted peak pressures and the crank angle at
which it occurs for the range of constant variables once more illustrating the
significance of the mixing parameters.
Table 5-5. Effect of Empirical Constants on the Prediction of Peak
Pressure (Baseline is P
max
78.83 atm at 3° CA)
Variables
Csef
Cmix
Cvidi
Thu
Percent Changes
in Value
+100
- 50
+100
- 50
+100
-100
+ 50
- 50
Predicted
^max
82.42
76.49
88.49
72.55
79.44
78.22
80.64
76.75
Peak Pressure
Timing °CA
+3 TDC
+3 TDC
+4 TDC
+ 1 TDC
+3 TDC
+3 TDC
+4 TDC
+3 TDC
5-18
-------
5.4 Droplet Diffusion Flames as a Source of Nitric Oxide
The model of pollutant formation in diesel engines grew from a belief
that spherical droplet diffusion flames characterized the total class of high
temperature diffusion flames which were responsible for pollutant formation.
No claim was made that such a belief would improve heat release predictions
nor that it actually described the burning process. However, it did dictate
the development of the model. Since a spherical diffusion flame was assumed
the influence of droplet interaction, and the relative velocity between the droplets
and their surroundings was ignored. In the early stages, droplets are confined
by the jet and even without violent mixing it is difficult to believe that single
droplets existing in an infinite environment will actually simulate the process.
It is more real to consider that the droplets exist as a cloud and act as a source
of fuel which is transported to the boundary of the cloud. More serious is the
tacit assumption of a quiescent atmosphere, droplet velocity lag would produce
wake flames which are dissimilar in structure to spherical diffusion flames or
even produce blow-off. Fuel vapor can be stripped from the parent droplet, it
may partially mix with oxident before combustion, and is therefore, unlikely
to remain associated with the parent droplets.
Calculations have been made in an attempt to ascertain the contri-
bution to the total NO production of the two modes of NO formation included
in the model. This was done by separately suppressing the homogeneous and
heterogeneous NO formation mechanisms. The calculated instantaneous NO
concentration history for the baseline case for these computations and the
normal predictions are shown in Figure 5-10. The NO formation from the droplet
diffusion flame appears to be only significant during the early stages of NO
formation. It can be seen that contributions of NO from the homogeneous and
heterogeneous phases are not additive and NO formed in homogeneous zones
appears to suppress NO formation in the heterogeneous phase.
5-19
-------
3000T
2000
=
a
1000
j
— Homogene
Only
- Heterogen
to NC Cor
/
A
//-'
//
//
X
JO Concentration
ous Contribution to NC C
ecus (Diffusion Flame) C
icentration Only
_^ *
n
1
*
X
X
^
;oncentratioj:
oncentration
"*
3000
2000
1000
-20
-10
TDC
10
20
30
Figure 5-10 Instantaneous Diffusion Flame and Homogeneous
Contributions to NO Concentration (Baseline)
5-20
-------
The relatively minor influence of the heterogeneous phase on the
total production of NO as shown by the model can be explained by:
• The majority of the droplets are burned in fuel rich zones where
diffusion flames cannot be established because of the lack of
oxygen. Less than 20 percent of the droplets have sufficient
oxygen (assumed to be 2 percent in these calculations) to
allow the formation of diffusion flames.
• Most of the droplets react in atmospheres containing NO which
may actually result in the destruction of NO, if not, it will certainly
reduce the rate of production.
Figures 5-11 and 5-12 present the results of calculation demonstrating
the influence of the droplet environment on the production of NO. The calcula-
tions were carried out with a droplet radius r, of 15M. and a constant ambient
pressure and temperature of 50 atm and 2000°K respectively, and the major
ambient species were assumed to be in equilibrium. The effect of ambient NO
concentration (for ^ = 0.4) is shown in Figure 5-11, the NO formation rate
decreases linearly with increasing ambient NO concentration, and when the
environment contains NO concentrations in excess of the equilibrium value
the rate becomes negative. Figure 5-12 shows the influence of ambient
oxygen concentration for both zero ambient NO and equilibrium ambient NO.
The NO formation rate increases rapidly with increasing ambient oxygen
concentration.
5-21
-------
01
I
ro
to
1U
5.
V
10
1 °
u>
o
r-l
X
0 _5
•e
-10
0
0
x^^
x
\
X
\
2000°K equil
1 1
X^
X^^
X.
2500°K eq
1
P = £
iO
0 = 0.4
T°° = 2000
oo
X
Uil
2750°K
equil
.0 4.0 8.0 12. 16. 20. 4. 8
Y v i n3
y^'KT/"V A i\J
NO ,<»
Figure 5-11 Effect of Ambient NO Concentration Formation
Around a Burning Droplet
-------
10
™ 1
m
10
-2
10
-3
0.0
0.2
0.4
0. b
0.8
1 0
.10
-4
Figure 5-12
Effect of. Ambient Oxygen Concentration on Rate
of NO Formation Around a Burning Droplet
5-23
-------
6.0 COMPARISONS BETWEEN'ENGINE EMISSION
LEVELS AND MODEL PREDICTIONS
The tuned model predictions have been compared to the engine
performance data reported by Wilson et ar . The characteristics of the
single cylinder experimental diesel engine were described in Section 5.2 as
were the engine conditions used for the baseline data. The response of the
model to the following parameters has been investigated:
• Engine operational parameters - timing, load and speed and
compression ratio;
• Charge intake parameters - air swirl, turbocharging and exhaust gas
recycle;
• Fuel injection parameters - diameter and number of orifices.
Calculations of both soot and NO formation are carried only as far as 30° after
top dead center since initial calculations indicated formation and distribution
rates were negligible beyond this point because of the low zonal temperatures.
Calculated NO levels are given in ppm and soot in mg/m^.
6 .1 Effect of Engine Operating Parameters-
Effect of Timing
The effect of fuel injection timing on exhaust NO level is well known.
Timing retardation significantly reduces the emission of NO due to:
reduction of ignition delay;
heat release in the expansion stroke where bulk temperatures
and pressure are declining
reduction of the total residence time.
These effects are more pronounced at high load, that is, when the duration of
fuel injection is longer. The comparisons of calculated and measured exhaust
NO concentration are shown in Figure 6-1. For 0=0.6, a delay 8° CA injection
reduces the NO by 35 percent, conversely a 70 percent increase in NO is
observed by 8° CA timing advancement over baseline case. The effect of
timing on the level of soot in the exhaust is shown in Figure 6-1. Higher
6-1
-------
2.0
1.5
1.0
[NO
0.5
0.0
(SOOT)
Calculated 2077 (ppm) 139 gm/m
Measured 2387 (ppm) 10%
(Phase II)
opacity
1.5
(Soot)
(Soot)
l.O
0.5
Figure 6-1 Effect of Timing
6-2
-------
exhaust smoke is detected when the timing is retarded. Approximately 12 percent
reduction in peak pressure is observed for every 8° CA injection timing delay.
Effect of Load
The effect of overall equivalence ratio, 9, on NO and soot emission
is shown in Figure 6-2. It has generally been observed that both NO and soot
emissions decrease with load. The calculations were made with a constant
injection rate and the load was varied by reducing the injection duration. The
explanation for the increased soot emission at the lowest load (9=0.3) is
uncertain, and this is contrary to all the other experimental observations, but
it could reflect a decrease in the rate of soot oxidation because of the lower
temperatures.
Effect of Speed
Calculations were made for two engine speeds while the overall
equivalence ratios and injection rates were maintained constant, thus the
duration of fuel injection was increased from 17° to 23° corresponding to an
engine speed increase of 1500 rpm to 2100 rpm. The comparison between the
calculated and measured data is shown in Figure 6-3. The effect of residence
time in a high temperature and pressure environment which is inversely propor-
tional to the engine speed, is the dominating factor. The other effects of higher
speed in the real engine such as higher injection pressure, higher valve overlap
are not reflected by the calculations, however, compared to the residence time
effect these are of minimum significance.
Effect of Compression Ratio
Decreasing the compression ratio will cause a decrease in both the
cylinder temperature and pressure, both of which will tend to reduce the rate
of NO formation. However, there are other important influential factors such
as the effects of heat transfer and ignition delay.
The overall effects of the compression ratio on emission level are
shown in Figure 6-4. The model predicts the trends; however, the comparisons
of absolute magnitudes are very poor. The exaggerated influence of compression
6-3
-------
1.5
1.0
NO]
0.5
0.0
(soot)
Calculated 2077 ppm 139 mg/m3
Line
Measured 2050 ppm 10% opacity —-o— —
0.3
2.0
1.5
(Soot)
(Soot),
1.0
0.5
0.0
0.45
0.6
Figure 6-2
Effects of Load
6-4
-------
l.Si
1.0
0.5
0,0
[N0]o (Soot)Q Line
Calculated 2077 ppm 139 mg/m3
Measured 2050 ppm 10% opacity — — — —
1.5
1.25
1.0
.75
O.S
Figure 6-3 Effect of Engine Speed
6-5
-------
1.0
(NO)
iNOT
0.6
0.2
(NO)
(Soct)
Q o
Calculated 2077 ppm 139 gm/m3
Measured 2050 ppm 10% opacity
14 15 16 17
CR
Figure 6-4 Effect of Compression Ratio
1.0
(Soot)
(Soot).
0.8
0.6
0.4
is
5-6
-------
ratio can be attributed, in part, to Cnor which had been fitted for a higher
compression ratio.
6.2 Effect of Charge Intake Parameters
Effect of Exhaust Gas Recirculation (EGR)
The effect of EGR was computed assuming a constant mass flow of
intake charge (i.e., mass of fresh air intake plus mass of recirculated exhaust
gas) and maintaining both the overall fuel/air ratio (i.e., proportionally reducing
the fresh air intake and the mass of fuel) and the fuel injection rate constant.
Figure 6-5 shows that excellent agreement is obtained between predicted and
calculated NO emission from the measured values. However, there is complete
disagreement with the measured soot values which probably reflects the model's
lack of real understanding of the density of soot formation and does not take
adequate account of the influence of temperature.
Effect of Air Swirl
The influence of the level of air swirl on NO formation can be seen
in Figure 6-6, the NO concentration increases with increasing swirl. This
effect is caused by the improved fuel-air mixing which leads to earlier heat
release, thus the higher local temperatures and the increased oxygen availability
promote NO formation. The experimental data from Phase I and that obtained
during experiments with the sampling valve are scattered. Data reported
by CAV is also plotted for comparison and to give an indication of the normal
trend.
The comparison of measured and calculated exhaust soot level is
poor, particularly-at higher swirl levels. This error is probably caused by an
over-estimate of the influence of temperature on soot formation which is used
in the present study. Thus, the higher dilution due to the higher swirl does not
significantly lower soot emission in the calculated results.
Effect of Turfaocharaing
The effects of turbo charging on NO are plotted in Figure 6-7 and are
compared with the experimental data obtained during Phase I. Poor agreement
is shown between the predictions and measurements at low levels of super-
charging. Computations at higher manifold pressure and higher loads were
6-7
-------
1.6
1.2
NO]
0.8
0.4
0.0
(Soot) Line
Calculated 2077 ppm .
Measured 2050 ppm 10% opacity — -O
0.0
10 20
(meg/mtotal)%
30
4.0
3.0
(Soot)
(Soot)c
2.0
1.0
0.0
Figure 6-5 Effect of Exhaust Gas Recirculation
6-8
-------
1.6
1.4
1,2
,'!.<
0,8
O'.i
— — -O- —
Calculation [NO] = 2077 ppm
pviase I data [NO] = 2050 ppm
~~ ~"* ~~ P^ase II sample valve data
[NO]o = 2387 ppm
— -Q- — CAV data (Khan & Greeves, 1973)
/
•Calculated (? = 0.6) (soot) = 139 mg/m3
I. 6
— -O- — P^ase I data $ = 0.45 (soot) =6.2%. opacity
\
\
V
1.2
(Soot)
(Soot)^
0.3
0.4
1.0
2.0 3.0
Relative Swirl Magnitude
0.0
Figure 6-6 Effect of Swirl
6-9
-------
1.4
1.2
NO
NO
1.0
0.8
0.6
NO.
(Soot)
o ' o
Calculated 2077 ppm 139 mg/m3
Measured 2050 ppm 10% opacity
Line
1.4
1.2
1.0
\
\
\
0.8
\
0.6
1.0
1.2
1.4
P (atm)
1.6
Figure 6-7 Effect of Turbocharge
1.8
2.0
5-10
-------
limited by equilibrium species tables and thermodynamic property expressions;
however, in future work this could easily be remedied.
6.3 Effect of Fuel Injection Parameters
Injector Orifice Diameter
The diameter of the fuel injection orifice affects bdth the fuel dis-
persion rate and the drop size distribution. Calculated exhaust pollutants for
various fuel orifice diameters are plotted in Figure 6-8, along with the experi-
mental data. The computations are based on constant total orifice area and a
constant total fuel mass. The emission of an engine with larger and therefore
fewer orifices showed substantial reductions in NO. The phenomena is mainly
caused by a smaller total surface-to-volume ratio of vaporizing and burning
fuel droplets in association with an increased mean droplet diameter which, in
turn, reduces rates of vaporization, mixing and thus energy release rates.
6-11
-------
1.4
1.0
[NO
(N
0.8
0.6
0.4
Data [NQ]0 (soot)Q Line
Calculated ($=0.6) 2077(pprn) 139 mg/m3
Measured ($=0.45) 1781(ppm) 6.2%. opacity—-O--
Measured ($=0.76) 1934(ppm) 11. 6% opacity --Q--
ci
8.0
5.0
(Soot)
(Soot).
4.0
2.0
0.0254
0.030S
d nozzle {cm)
0.0356 0
1.0
Figure 6-8 Effect of Orifice Diameter
5-12
-------
7.0 LIMITATIONS OF THE MODEL AND RECOMMENDATIONS
FOR FUTURE WORK'
While it is unlikely that a diesel engine will be designed solely on
the basis of a mathematical model, such a model can provide insight which
will allow more rapid development of engines. The preceding sections have
described the development of an analytical model which simulates the mechanisms
of heat release and pollutant formation occurring within a direct injection diesel
engine. Several simplifying assumptions have been made during the development
process to allow the complex phenomena, which occur in practice, to be treated
mathematically. Thus, the ultimate goal, that of building a model based entirely
upon physical and chemical laws and free from empiricism has not been achieved.
What has been accomplished is the construction of a modular model which will
predict, after tuning, most of the pollutant emission characteristics of a single
cylinder diesel engine. Its modular nature will enable the rapid update of any
of the model's components as improved fundamental and experimental informa-
tion becomes available.
Recognizing that gaps in fundamental knowledge make the inclusion
of several empirical constants mandatory, the success of the model must be
judged from predictions after adjustment of these constants for a particular
class of engine. To date, the model has only been compared to single cylinder
engine data. On the basis of these comparisons it can be concluded that the
overall predictive capability of the model with respect to pressure history and
exhaust NO levels is good while its ability to predict exhaust soot emissions
can only be considered as fair. The model's predictive performance has been
summarized qualitatively in Table 7-1.
7-1
-------
Table 7-1. Summary of Model Predictive Capability
Variables
Fuel Injection Timing
Load
Engine Speed
Compression Ratio
Exhaust Gas Rec irculation
Swirl
Turbocharge
Nozzle No. and Diameter
Exhaust NO
good
good
good
poor
good
good
poor
good
Exhaust Soot
good
fair
good
fair
poor
poor
fair
fair
Initial concepts concerning the formation of pollutants in compression
ignition engines directed the model towards the consideration of droplet com-
bustion. It was believed that spherical diffusion flames formed around droplets
were typical of the "diffusion flame" nature of the combustion process. Experi-
ence with the model tends to refute this hypothesis. Undoubtedly, diffusion
flames exist within the heat release zone, however their scale is apparently
very different from those modeled by a spherical diffusion flame surrounding
a single droplet. In order to include this type of heat release zone, a pseudo
steady state confined droplet model has been formulated. Model calculations
indicate that droplet contribution to the total production of NO is only signifi-
cant in the early stages of heat release, and has a negligible influence on the
final exhaust value. The results of these calculations can be attributed to:
• A large fraction (approximately 80 percent) of the droplets are
contained within rich zones which cannot support spherical
diffusion flames.
• The droplet lifetime is restricted to the time period close to
top dead center where the available volume is small and
therefore, the volume available for each separate spherical flame
is small, thus minimizing the total flame surface area.
• The droplet environment contains NO produced during premixed
combustion which will reduce the net rate of NO formation.
7-2
-------
Exercise of the model has highlighted several deficiencies and
omissions which can be easily modified in the near term; these include:
• Modifications to extend the equilibrium species tables to higher
pressures and richer mixtures.
• Limits can be imposed upon the equilibrium assumption to enable
carbon monoxide and hydrocarbon levels to be predicted.
• The model contains a very simple ignition criteria and it should
be modified to take account of the equivalence ratio and
temperature history of the fuel parcel.
• Mixing between burning zones should be included.
Four general areas of ignorance have been identified which require
considerable effort to allow improvements to be made to the model; these include
a transient fuel spray model, non-uniform fuel injection rate, an improved fuel
spray dispersion model, and an improved energy release model. Both physical
and numerical experiments will be necessary to allow improvements to be made
in these components.
Transient Fuel Spray
The present model considers that the fuel spray behaves as though it
were a steady state jet. No attempt has been made either to model the leading
or the trailing edge of the jet. More experimental information is required before
a realistic approach can be included in the model which will describe these
transient phenomena.
Fuel Air Mixing
In view of the conclusions concerning fuel droplets, it is suggested
that numerical and physical experiments be conducted to establish the importance
of droplets. The phenomena of ignition may well be strongly linked to droplet
properties. However, it appears that the major portion of the jet could be
treated as a gaseous fuel. This would simplify certain aspects of the model
and allow a more rigorous treatment of the fuel/air mixing process to take
account of the following.
7-3
-------
air swirl;
jets in cross flow; and
jet interaction and three dimensional effects
An essential prerequisite for an improvement in the fuel/air mixing
model is more detail of the mixing processes occurring in the cylinder under
fired conditions. This is particularly true of conditions after ignition. A major
unanswered question concerns the lifetime of the well-ordered jet structure.
Non-Uniform Fuel Injection Schedule
The fuel injection schedule is assumed to be constant. As a first
simple step towards taking account of a non-uniform injection schedule the
model could be modified to calculate the entrainment into a series of steady
jets with different injection rates.
Heat Release Model
Improvement in the heat release model will be necessary if a more
real approach to-pollutant formation is required. This could occur if engines
with low pollutant potential were being considered or fuels containing signifi-
cant quantities of bound nitrogen were to be utilized. Under these conditions
it will be necessary to treat the hydrocarbon oxidation process in more detail
than the present equilibrium assumption.
In view of the fact that the model has only been tested against one
set of experimental data, a necessary first set in deciding where the model
needs the most refinement is it's application to other engine geometries and
multicylinder engines.
7-4
-------
8.0 NOMENCLATURE
A Area (cm2) or pre-exponential factors.
b Jet half width (cm)
C Constants
C Constant pressure specific heat (cal/gm-°K)
C Constant volume specific heat (cal/gm-°K)
D Diffusion coefficient (cm2/sec)
D Engine bore (cm)
d Droplet diameter (cm)
dm,^ Maximum droplet diameter (cm)
max
E Activation engine in Arrhenius rate expression
erf Error function
f,F/A Fuel/air ratio by mass
G Mass fraction of a drop size group
h Enthalpy (cal/gm)
I Integral function
I Heat equivalent of work
j Stoichiometric oxygen - fuel mass ratio
K Reaction rate constants, or absorption coefficient in measure of density
of radiating particle per unit path length
L Latent heat or characteristic length
m Mass (gm)
m Mass flow rate (gm/CA)
M Molecular weight
M Nondimensional droplet burning rate
N Cetane Number
ce
P Pressure (atm)
P Pressure derivative (atm/CA)
Q Rate of convective heat transfer (Cal/CA)
c
0D Overall ratio of radiative heat transfer (Cal/CA)
. K
q Zonal rate of heat transfer (Cal/CA)
q Heat of Combustion (Cal/gm)
R Gas constant
Connecting rod length (cm)
8-1
-------
NOMENCLATURE (Cont'd)
RW Crank radius or half stroke length (cm)
R Reynolds number
6
r Radial coordinate (cm)
T Temperature f K)
t Time (sec)
u Internal energy (cal/gm)
V,V Volume and clearance volume (cm3)
w
v Velocity (cm/sec)
W Production rate (gm/cm -sec)
(W) Production rate(gm/sec)
w Number of carbon atoms in a fuel molecule
X, [ 1 Mole fraction
x Axial or downstream coordinate (cm), or number of hydrogen atoms in a
fuel molecule
Y Mass fraction
y Radial or cross-stream coordinate (cm)
Greek Symbols
ff Stephen Baltzman constant
or Absorptance or a - l/(l+f)
ft Volume ratio
y Fraction factor of convecture heat transfer
9 Crank angle
* Equivalence ratio defined as § = (F/A)/(F/A)stoicj,
p Density (gm/cm3)
£ Nondimensional radial coordinate, y/b
x Thermal conductivity (cal/cm-sec-°K)
^ Absolute viscosity (gm/cm-sec)
<5 Parameter for drop size distribution
r Characteristic time
17 Defined by r\ = f/(l+f)
8-2
-------
NOMENCLATURE (Cont'd)
Subscripts and Superscripts
a Air (instantaneous)
a,a Air available for air/product mixing process
apb Air in premixed burning process
b Burning process, or backward reaction
bm Combustion products (Instantaneous)
c Convecture heat transfer or clearance
char Characteristic properties
crit Critical conditions
CO2 Carbon Dioxide
D Dilution process (cumulated)
el,e2 Vaporization process
eq Equilibrium
fv Fuel vaporized (cumulated)
fg Fuel vapor (Instantaneous)
fid Liquid fuel ignited (cumulated)
fib Liquid fuel in burning process (instantaneous)
fdb Liquid fuel burned in diffusion flame process (cumulated)
fpb Fuel vapor burned in premixed process (cumulated)
fl Liquid fuel in vaporizing process (Instantaneous)
fj Fuel injection process (cumulated), or free jet
f Fuel, or forward reaction
H2O Water
HET Heterogeneous process
HOM Homogeneous process
hu Heat-up process
i Zonal identification, i = a, fg, (bm,I)
ig Ignition
inj Fuel injection
I I'th burning zone
j Shell index
1 Liquid phase or droplet surface
max Maximum
mix Mixing process
8-3
-------
NOMENCLATURE (Cont'd)
N2 Nitrogen
NO Nitric Oxide
O2 Oxygen
o Initial condition
os Stoichiometry
r Radiation
s Soot; Swirl
sf Soot formation process
so Soot oxidation process
turb Turbulence
v Vitiation process (cumulated)
w Wall
wj Wall Jet
§ Center-line
(• ) Rate in crank angle (i.e. d/d9)
(~) Averaged value
" Ambient conditions
Note: Where other symbols are used, they are defined in the text.
8-4
-------
9.0 REFERENCES
1 Wilson, R. P. Jr., Waldman, C.H. and Muzio, L.J., "Foundation for
Modeling NOX and Smoke Formation in Diesel Flames", Final Report
for Phase I, EPA-460/3-74-002a, January, 1974
2 Bastress, E.K. et al, "Model of Combustion and Nitrogen Oxide Formation
in Direct and Indirect Injection Com press ion-Ignition Engines" SAE
No. 719053, 1971
3 Khan, I.M., Greeves, G. and Probert, D.M., "Prediction of Soot and
Nitric Oxide Concentrations in Diesel Engine Exhaust," Inst. Mech.
Engr., Paper C142/71, 1971
4 Khan, I.M. and Greeves, G., "A Method of Calculating Emissions of
Soot and Nitric Oxide from Diesel Engines," SAE Paper No. 730169,
January, 1973
5 Shahed, S.M., Chiu, W.S. and Yumla, V.S., "A Preliminary Model for
the Formation of Nitric Oxide in Direct Injection Diesel Engines and Its
Application in Parametric Studies," SAE Paper 730083, 1973
6 Shahed, S.M., Chiu, W.S. and Lyn, W.T. "A Mathematical Model of
Diesel Combustion", Institution of Mechanical Engineers (1975).
7 Lyn, W. T. , "Study of Burning Rate and Nature of Combustion in Diesel
Engines," IX Symposium (International) on Combustion, 1963
8 Lyn, W. J. , and Valdmanis, E. , "The Effects of Physical Factors on
Igntion Delay" SAE Paper 680102, 1968
9 Lyn, W. T. and Valdmanis, E., "The Effect of Physical Factors on Ignition
Delay," Proceedings of Inst. of Mech. Engrs., Vol. 181, 2A, No. 1, 1966
10 Bowman, C. T. , The Fourteenth Symposium (International) on Combustion,
p. 729, The Combustion Institute, 1973
11 Fenimore, C. P., "Formation of Nitric Oxide in Premixed Hydrocarbon
Flames", Thirteenth Symposium on Combustion, The Combustion
Institute, p. 373, 1971
12 Myerson, A. L. , Fifteenth Symposium (International) on Combustion,
p. 1085, The Combustion Institute, 1975
13 Gordon, S. and McBride, B.J., "Computer Program for Calculation of
Complex Chemical Equilibrium Compositions, Rocket Performance,
Incident and Reflected Shocks, and Chapman-Jouguet Detonation",
NASA, SP-273, NASA Lewis Research Center, 1971
14 Annand, W.J.D. , "Heat Transfer in Cylinders of Reciprocating Internal
Combustion Engines", Institution of Mech. Engrs. 1962
9-1
-------
15 McAulay, KJ., et al, "Development and Evaluation of Simulation of
Compression-Ignition Engine", SAE Paper No. 650451, 1965
16 Borman, G., "Mathematical Simulation of Internal Combustion Engine,"
Ph,D Thesis, University of Wisconsin, 1964
17 Lee, D. W., "The Effect of Nozzle Design and Operating Conditions
on the Atomization and Distribution of Fuel Sprays", NASA Kept.,.
425,1932
18 ' Abramovich, G.N., "The Theory of Turbulent Jets", M.I.T. Press, 1963
19 Wise, H. andAgoston, G.A., "Burning of Liquid Droplet", Literature
of Combustion of Petroleum, American Chemical Society, Washington,
D.C., p. 116-135, 1958
20 Rosner, D.E. and Chang, W.S., "Transient Evaporation and Combustion
of Fuel Droplet Near Its Critical Temperature". Combustion Science and
Technology, Vol. 7, p. 145-158, 1973
21 Shipinski, J., Myers, P.S. ndUyehara, D.A. , "A Spray-Droplet Model
for Diesel Combustion", Proc. Instn. Mech. Engrs., Vol. 184, PE 3J,
p. 28-35, 1969
22 Williams, A., "Combustion of Droplets of Liquid Fuels: A Review",
Combustion and Flame, Vol. 21, p..1-31, 1973
23 Williams, F.A., "Combustion Theory", Addsion-Wesley Publishing Co.,
1965
24 Zeldovich, J., "The Oxidation of Nitrogen in Combustion and Explosion",
Acta Physicochimica USSR 21, 557, 1946
25 Baulch, D.L., Drysdale, D.D., and Lloyd, A.C., "High-Temperature
Reaction Rate Data No. 1", Dept. of Physical Chemistry, The University;
Leeds, England, May 1968
26 Wray, K.L. and Teare, J.D., "Shock Tube Study of Kinetics of Nitric
Oxide at High Temperatures", I. Chemical Phys. 36, 2582, 1962
27 Caretto, L.S., Sawyer, R.F. and Starkman, E.S., "The Formation of
Nitric Oxide in Combustion Processes", University of California Rpt.
TS-68-1, 1968
28 Lee, K.B., Thring, M.W., and Beer, J.M., "On the Rate of Combustion
of Soot in a Laminar Soot Flame," Combustion and Flame 6 p. 137-145,
1962
29 Field, M.A. et al, "Combustion of Pulverized Coal" BCURA, Cheney
& Son Ltd., Banbury, England, 1967
9-2
-------
30 Sitkei, G. and Ramanaiah, G. V., "A Rational Approach for Calculations
of Heat Transfer in Diesel Engines", SAE Paper No. 720027, 1972
31 Bracco, F.V. , "Nitric Oxide Formation in Droplet Diffusion Flames",
14th International Symposium on Combustion, p. 831-842, 1972
32 Ranz, W.D., Marshall, W.R., Chem. Eng. Prog. Vol. 48, p. 141-173,
1952
33 Frossling, N., Gerlands Beitr. Geophysics, Vol. 52, p. 170, 1938
34 • Newman, J.A. and Brzustowski, T.A., "Behavior of a Liquid Jer Near
the Thermodynamic Critical Region", AIAA, Vol. 9, No. 8, 1971
35 Gauntner, J.W., Livingood, J.N.B. and Hrycak, "Survey of Literature
on Flow Characteristics of Single Turbulent Jet Impinging on a Flat
Plate", NASA TN D-5652, Lewis Research Center, 1970
9-3
-------
APPENDIX A
Simplified Representation of Fuel Spray
The basic fuel spray model assumes a two-phase round jet in a quiescent
ambient atmosphere which impinges on a flat wall in a finite distance. A schematic
representation of the jet is shown in Figure 4.2 . The jet is characterized by four
distinct flow regions:
• The initial region: extends from the orifice exit plane to the apex of
constant velocity core x,
• The main region: includes that part of the jet from the end
of the potential core to the wall. This region is characterized by the
decay of centerline velocity and self-pre serving transverse profiles
• The transient region: the region where the jet is deflected and trasnformed
from a free jet to a wall jet
• The wall region^ includes that portion of the jet where it spreads
three-dimensionally and axisymmatically with respect to axis
of injection.
(i) Initial Region
The growth law in this region is calculated by
, pa , (A-l)
db = Q.135 1 + —
dx V f
(34)
where constant C( - 0.135 was used by Newman and Brzustowskr '.
(ii) Main Region
(18)
Self-preserving velocity and fuel concentration profiles (Abramovich ),
are assumed.
A-l
-------
where £ = -I— and the constant growth law is assumed:
b
db
dx
(A-3)
(A-4)
constant
where Cf. = 0.175 corresponds approximately to 10 degree half spray angle.
Utilizing the law of conservation of momentum, the decay of centerline velocity
in main region can be derived as:
1 1
2
-VfS
fo \ _
^
(A-S)
Similarility from the law of conservation of fuel mass, the decay of fuel concentration
in the main region can be obtained as:
I
f
(A-6)
The values of the integrations
(A-7)
are listed in Table (A-l) for various n.
J^ and Im appearing in the
flux of momentum, respectively.
and Im appearing in the above equations are the flux fuel mass and the
A-2
-------
The locus of the mass element in "j" shell is expressed by
= 2?rb
/
(A-8)
(iii) Transient Region
The initial conditions of the wall Jet are obtained from conversation of mass
and mechanical energy during the transition. The initial width and velocity of
3-D wall jet are expressed as:
b v
v
f2
v
•few
(A-9 )
(A-IO)
The initial fuel density of the wall jet at the location nearest to the wall is
obtained from conservation of fuel mass:
0 b v II
w £w f2 3
(A-1L)
Where subscript "o" represents the values at r = o and the numerical values
of the integrations
1.5 ,n , „
I = 27T/
n J
A-3
-------
are evaluated and listed in Table (A-l). It should be noted that the energy
dissipation and entrainment during the transition from the jet to the wall jet are
not considered.
(iv) Wall Region
Self-preserving velocity and fuel concentration profiles are assumed i.e.
- (i - r )
The boundary layer in the vicinity of the wall is not considered here. The
growth law of the wall Jet is:
$3L = c = constant (A-14)
dr wj
where C , = 0.204 is taken from the work by Gauntner, et all35'.
The fuel concentration nearest to the wall, derived from mass conservation relationship
is
/
/ Pf\ Tf
o p v' (r + bw b'I_
a / a € 3
* (A-15)
A-4
-------
The velocity decay, derived from conservation of momentum flux, is
2
- _5 Tf + L! y \ + 4 (r + b ) b p I, J
V b/ "a'4
where initial jet momentum flux is
' bwV (v
Similar to Eqn (A-9) the Trajectories of fuel elements in the wall region are
computed from
_1_ A5'5 1 ,4+-i £2'5 * = "° "
5.5 j "
" 4 *J
Table A-l
n
1
2
3
4
5
6
7
2.5 *J
Numerical
I
n
3.7699
2.8274
2.3122
1.9854
1.7487
1.5738
1.4730
J 2?rP' v' (r + b ;b'
I* €, W^
Values of I & L
n in
I.
fn
1.3446
0.8105
0.5592
0.4209
0.3311
0.2709
0.2276
A-5
-------
APPENDIX B
Coefficients for Equations (4-25) and (4-26)
These polynomial coefficients are obtained from least-square curve fitting
of computed results from NASA Equilibrium Program' ', particularly for H/C
equals 1.7. For fuel having different H/C, these coefficients have to be
reevaluated.
Table B-l. 0 s 1.0
1
2
3
4
5
6
7
8
9
10
11
-5.5492E+01
7.8374E-02
1.6734E-04
-1.1376E-07
4.0737E-11
-5.4383E-15
7.3465E+02
3.7817E-02
-1.0774E-04
6.1112E-08
-1.4435E-11
9.9489E-16
1.3700E+01
-1.1-17E+00
-5.715SE-01
-2.6780E+01
7.2448E+00
1.3090E+00
5.0189E-01
-5.2147E-01
5.3444E-02
-2.6396E+00
1.6069E+00
6.8789E-02
4.3293E-03
-9.8361E-01
6.6435E+00
1.1172E-01
2.3055E+00
-2.5753E+01
-4.0113E-01
i
1
2
3
4
5
6
7
8
9
10
11
-8.5550E+02
2.5318E-01
•1.6176E-05
•1.0564E-07
7.2508E-11
1.2100E-14
Table B-2. 3.5 >
b.
-1.1796E+02
6.5986E-01
-1.1693E-03
7.1490E-07
-1.8777E-10
1.7905E-14
1.0
d.
9.7714E+00
4.3547E+00
-7.2560E-02
-9.6828E+00
-1.5763E+01
2.3124E-01
-6.4446E+00
7.2392E+00
5.2083E-03
2.2071E+01
-2.5775E+01
r.
4.3246E-02
1.2509E-02
2.3144E-01
-3.8400E-01
-3.3850E-03
-4.0000E+00
2.9256E-01
-1.7973E-02
B-l
-------
APPENDIX C
CHARACTERIZATION OF DIESEL COMBUSTION
BY DIRECT IN-CYLINDER SAMPLING
C-l. INTRODUCTION
The model of diesel engine combustion described in the body of this
report includes several arbitrarily adjusted constants and is based upon a
healthy dose of conjecture. In an attempt to provide experimental evidence to
support the model development, an effort has been made to characterize the
cylinder contents during the period of heat release. Wilson et aP reported
earlier unsuccessful efforts to utilize ultraviolet emission and absorption
spectroscopy for this purpose. A sampling valve,which was designed by the
Ricardo Company and provided to Ultra systems by the General Motors Research
Laboratories, has been used in an attempt to provide data to aid further model
development.
In addition to the overall objectives stated above, one important
feature of the project was to establish the validity of the results from
in-cylinder direct sampling. The project could not be concluded in the pre-
sent study, but answers to the following questions vrere sought:
• What kind of spatial and temporal resolutions are possible?
• On what scale is it possible to characterize gradients and
unmixedness in the system?
• To what degree is the sample analyzed, representative of
the system being measured?
C-2. EXPERIMENTAL EQUIPMENT
The characteristics of the single cylinder diesel engine used in
these experiments have been given elsewhere . The sampling valve was
designed around a powerful electromagnet and armature originally intended
to operate a Societe Francaise d'Etudes et de Development de 1, Injection
(S.O.F.R.E.D.I.) diesel fuel injector. A detailed description of the development,
C-l
-------
(2)
construction and operation has been given by Nightingale . Details of the
sampling valve tip used in this investigation are given in Figure C-l. Valve
lift and open times are dependent upon spring strength, which acts to close
the valve, and the magnitude and duration of the current pulse passing through
the electromagnet. Since cylinder pressure acts to open the valve, adjustments
must be made to provide constant sampling conditions for different loads and
crank angles . The actual valve opening time represents a compromise between
that necessary to provide sufficient sample for analysis and a desire to maintain
a high temporal resolution. Figure C-2 provides a schematic of the sampling
system. In the final series of tests the triggering delay electronics for acti-
vating the electromagnet which allowed an accurate adjustment of the triggering
crank angle were supplied by AiResearch.
Experience with the sample valve indicated that the titanium valve
tip and seat could not maintain satisfactory operation under conditions of high
load and insertion. In the initial series of tests an insertion of 1/4 in. was
only possible at part load. Leakage occurred after a short period of time due
to the forging action of the valve stem. This problem was overcome in the
final test series by using Waspolloy (0.50 Nc; 0.19 CN, 0.14 CO, 0.04 MO
and 0.03 Tc) for the valve seat and titanium carbide for the stem tip. These
design modifications allowed operation under high engine load condition-s with
full insertion without leakage. Included in Figure C-l is an alternate valve
tip design which utilizes a sphere as the tip, rather than a conical section.
The sphere is a press fit into the end of the stem and the seat is coined rather
than machined thus allowing rapid inexpensive renewal of the valve component
not subject to failure-
The sampling probe could be located in three different ports in the
cylinder head. Rotation of the fuel injector then allowed samples to be
obtained on these three radii at any distance from the fuel jet axis. Figure
C-3 shows the radial location of the three sampling ports .
C-2
-------
o
I
CO
RICARDO SAMPLING VALVE
SEAL
•9/64 DIA
COOLING WATER
CYLINDER
HEAD
/2%— TITANIUM
.096 O.D. x
.068 I.D.
STAINLESS
STAINLESS
TIP
90?
3/16 DIA,
ALTERNATE TIP DESIGN
.0625 OIA
TUNGSTEN CARBIDE
SPHERE
v\\\\\\\\\
\
\
\
STAINLESS
VALVE STEM
I
COINED
SEAT
0.28
Figure C-l. Details of Sampling Valve Tip
-------
DC Power
Supply
Sampling
Valve
Electronics
Capacitor Charging
Voltage (Input)
o
I
Spring Tension
Adjustment
(Input)
Sample
Capacitor
Discharge
Trigger Delay
.6 Sample
Valve Open
(Input)
Sample Valve
Roto meter
H2O Trap Filter
O2 Analyzer
1
NO, NO
Analyzer
x
(Paramagnetic)
(Chemllumlnescent)
Crank
Angle
Signal
Figure C-2. Schedule of Sampling System
-------
R/R
Bowl
.0.95
0.65
.
CYLINDER HEAD
Rb»1.62
Figure C-3. Location of Sample Valve Ports
C-3. RESULTS
Four test series, all of which were not completed, have been carried
out using the sampling valve:
Series 1 — All exhaust values were higher than the local values
at 40° after TDC .
Series 2 — Problems with the sampling train were corrected, but
engine failure occurred before the test matrix was
completed.
Series 3 — Complete test matrix, but discrepancies apparent in
results.
Series 4 — Partial test series with new valve tip and seat after
the engine had been rebuilt.
C-5
-------
An impression of the differences in the results obtained from the various test
series can be gained from the curves of NO versus sampling valve opening
^t
time presented in Figure C-4. In Series 1 all the sample valve NO values
J\
dropped below the exhaust value. Although local values must not necessarily
agree with the exhaust level at 40 ATDC it was thought that the sample valve
values should show some distribution about the final level. Examination of
the sample train indicated that NO was being converted to NO2 in the sample
line with subsequent loss of NO2 in the condensate. Sample line lengths were
reduced, a chemical drier was eliminated from the train and a water trap fitted
close to the outlet of the sample valve. After these modifications, results
typified by those shown in Figure C-4 for Series 2 were obtained. Unfortun-
ately, before the second test series could be completed, an engine failure
occured and after repair the pollutant emissions for the same nominal engine
conditions were different.
In the third series of tests a complete test matrix was carried out
and the curves presented in Figure C-4 give an impression of their characteris-
tics in comparison with earlier results for the same engine condition. Table
C-l lists the test matrix which was carried out in the third series of tests, and
a complete listing of experimental data is included in Appendix D.
Three problems are apparent upon examination of this data:
• Variation in exhaust levels for what are nominally the
same engine conditions.
• Differences in the influence of swirl on exhaust emissions from
test to test.
• Errors in the measurement of NO concentrations.
J^
An impression of the reproducibility of the exhaust NO levels can be gained
from the results presented in Table C-2. Even discounting the obvious rogue
values, the results exhibit a considerable degree of scatter. This scatter may
be associated with slight changes in engine parameters such as ignition timing.
Whatever their cause, they indicate that the in-cylinder sampling results could
have a greater level of uncertainty.
C-6
-------
4000
3000
a.
Q.
2000
Q
X
O
z
111
O
O
OS
1000
Speed: (500 rpm
Load: £ = .5
Timing
-------
Table C-l. Experimental Condition for In-Cyllnder Investigations
Series
A
B
C
D
E
F
G
H
I
R.P.M.
1500
1500
1500
1500
2100
1500
1500
1500
1500
Load
0.6
0.6
0.6
0.6
0.6
0.35
0.35
0.35
0.3
No. of Orifices
in Injector
6
4
6
6
6
6
6
6
6
Timing
-20
-20
-28
-12
-20
-20
-28
-12
-20
Swirl
Low
Medium
High
Medium
Medium
Medium
Medium
Low
Medium
Medium
Medium
Low
Medium
No. of
Valve Locations
8
7
6
6
8
8
4
2
2
2
2
6
8
Remarks
*
*
*
*
o
I
QO
Series with sample valve inserted 0.25 in.
-------
Table C-2. NO Exhaust Emission Levels Measured For
^t
Different Sample Valve Locations (Series 3)
-20°
2431
2333
2435
2382
2360
2669
2840
2300
2460
-28°
3210
3100
3245
3540
3600
3823
4440
3640
-12°
1254
940
1412
1395
1085
1230
1209
1300
1300
Swirl Medium
Load <£• 0.6
RPM 1500
As discussed earlier, between the second and third test series there
was an unexplained jump in the baseline emission levels. What is perhaps
more disquieting, is that there was also a change in the behavior of the engine.
Data presented in Section 6 shows that the measured influence of swirl level
on NO emissions was not reproduced from Phase I to Phase II. The emission
jC
levels tabulated in Table C-3 indicate that the "influence of swirl" also varied
between the second and --third sampling valve test series.
09
-------
Table C-3. Comparison of the Influence of Exhaust
Emission Levels for Series 2 and 3
1500 rpm 9 = 0.6
Low Swirl
NO
-20ATC
Series 3
Medium Swirl
8.8
8.8
8.8
8.8
8.6
8.3
9.0
NO
x_
2431
2435
2383
2396
2300
2460
2411
o
2_
8.7
8.6
8.7
8.8
8.8
8.7
8.3
High Swirl
N0x
2534
2200
2200
2190
2100
2450
8.4
8.6
8.6
9.0
8.6
8.9
Series 2
2230
2290
2470
8.8
7.5
8.1
1340
1430
1200
1460
,0
,8
7.6
8.0
2800
2830
2670
2880
8.4
7.8
8.8
8.9
C-10
-------
The dead volume in the sampling system was reduced as much as
possible. However, the sample analyzed always contained a significant
fraction of NO2, and it is hot known whether all of this is due to oxidation in
the sample line or whether NO2 is actually formed in the cylinder. Conse-
quently, the concentration of interest is that of the total nitrogen oxides since
the NO level will not be sufficient to characterize the pollutant content of the
sample. NO levels were determined by passing the sample through a conver-
Jt
ter which normally converts the nitrogen dioxide to nitric oxide, thus allowing
the total nitrogen oxides to be measured by the chemiluminescent analyzer.
However, when the sample contains fuel gases such as hydrocarbons and carbon
monoxide (the sample would probably also include hydrogen, but this was not
measured) NO levels lower than NO levels can be recorded. Typical results
JC
are presented in Table C-4 where it can be seen that for samples withdrawn
around TDC the NO levels are obviously in error. Thus, the results will mis-
J^
represent actual NO levels when fuel gases are present tending to give lower
JC
rates of NO increase since these are the samples normally containing fuel
Jt
gases.
The modified valve design allowed measurements to be made during
the fourth test series with much deeper valve insertion distances. Several of
the results obtained with this design are presented in Figure C-l. Time restric-
ted the extent of the test matrix which was limitted to one sample valve posi-
tion (No. 2) with the fuel nozzle rotated such that the sample valve was located
30° downstream from the initial jet orientation. Probe insertion was varied
between the flush condition and 11/16 inch penetration. Swirl and injection
times were varied, while load and speed were held constant at 0 = 0.5 and
1500 rpm. Considerable care was taken to ensure that exhaust levels were
reproducible and that actual sample valve data was repeatable.
In the fourth series of tests' the sample valve was operated over a
wider cycle range since it was expected that at sample times greater than +100°
the sample valve concentration would approach the exhaust values. As shown
in Figures C-4, 5, 6 and 7, this was not the case for this one sample valve
C-ll
-------
Table C-4. Examples of Erroneous NOX
Concentrations Due to Converter Errors
Sample
Time
-32
-22
-12
-7
-2
3
-22
-12
-7
-2
3
Measured
NOx ppm
291
201
126
1880
2722
3580
152
255
245
1085
2130
Measured
NO ppm
166
113
322
1600
2300
2811
140
350
410
960
1820
Calculated
NO2 ppm
125
88
-196
280
422
769
12
-95
-165
125
310
Measured
°2 %
19.5
19.0
14.2
11.0
10.5
10.7
19.8
9.8
8.5
12.0
13.2
C-12
-------
4000
3000
5
a.
a.
I
Ul
§ 2000
O
z
tu
8
a
1000
Speed:
Load:
Timing:
Swirf Le- el
A High
o Medium
a Low
450 -100 -50 0 +50
Crank Angle, Degrees
+f.OO +(.50
Figure C-5 . Dependence of Sample Valve Measurement
on Swirl Level - 11/16" Valve Insertion
C-13
-------
4000
3000
2
a
a
ox
S 2000
Ui .
O
X
O
Z
UJ
ee.
2
1000.
Speed
Load
Timing
Swtrt-Level
A High
o Medium
a Low
•100 -50 0 50 100
Crank Angle, Degrees
ISO
Figure C-6. Dependence of Sample Valve Measurements
on Swirl Level - Flush Valve Position
C-14
-------
4300
3000
2
a
a.
O
u
o
X
O
z
UJ
2000
1000
Speed: 1500 rpm
Load: <£= .5
Swirl: Medium
Valve
Insertion: Flush
S
S
X
UJ
•ISO -100 -50 0 50 100 ISO
Crank Angle, Degrees
Figure C-7. Dependence of Sample Valve Measurement
on Injection Timing
C-1S
-------
location the final sample yielded NOX levels lower than the exhaust value.
Some idea of the unmixedness existing in the cylinder can be gained from the
results showing the influence of valve insertion depth shown in Figure C-4.
Peak NO levels are considerably greater than the exhaust value. Also, the
peak value does not increase monotonously with increasing insertion depth.
Increasing swirl increases exhaust NO levels which is reflected
in the peak NO levels recorded by the sample for two insertion depths (see
Figures C-5 and 6). It should be noted that peak levels tend to occur later
in the cycle for lower swirl levels. The dip in the NOX level close to TDC
with the lowest swirl level when the valve was in the flush position appear
to be real as the results could be repeated.
Variation of the injection timing produces a radical change in the
character of the NO versus crank angle curves when the valve is flush with
X
the cylinder head. The rise in NO is delayed in agreement with the delay in
fuel injection.
C-4. DISCUSSION
Two issues must be addressed relating to the correspondence be-
tween the measured concentrations and those actually existing in the cylinder
before discussing the implications of the results, these are:
1. Is the sample ingested into the valve representative of the con-
centration at the probe position and sampling time in the absence
of the probe ?
2. Does the sample undergo chemical transformation in the probe?
Complete answers to these questions are not available although it is possible
to define possible effects and construct plausable arguments as to their
magnitude. Regarding the first question posed above,the following effects
can perturb the sample:
- influence of the inserted probe tip on local gasdynamics and com-
bustion;
C-16
-------
- ingestion of the quench layer adjacent to the probe tip;
- influence on the local gasdynamics of the sink strength during the
sample period;
- size and duration of the sample as they affect spatial and temporal
resolutions
Simple fluid mechanical arguments suggest that the quench layer
and injector trapped mass do not represent a significant effect at the sample
rates under consideration. This conclusion has been borne out by varying
the sample rate and the duration of the valve opening time. The presence of
the probe itself undoubtedly outweighs the sink and sample size effects. For-
tunately the flow is predominantly across the-probe and hence the largest
perturbation to the flow lies in the wake, downstream of the probe. All of
these items taken together probably mean that the sample should be considered
as representative of a zone whose scale is that of several probe diameters.
The long valve open time (** 1 msec) is probably the single most severe effect
with regard to both temporal and spatial resolutions due to high swirl velo-
cities which can convect material from 30° upwind during this period.
With regard to chemical transformation within the probe, the fol-
lowing effects may be of concern:
- lack of adequate quenching within the probe;
- catalytic wall reactions;
- presence of carbon in the probe;
- presence of condensate in the sample line
Analysis of the first point requires the recognition that the probe's internal
gasdynamics is an unsteady filling and emptying process. The stem fills
over a period of 0.5-1.0 msec and empties over a period of 80 msec. During
this process the sample at first expands very rapidly and then undergoes a re-
heating through shock compression and boundary layer stagnation. Although
the required quench rates are severe, an analysis of this process indicates
that the favorably lower pressures in the probe combined with the cool walls
is adequate to terminate the nitrogen oxide production. This conclusion seems
to be validated by the independence of measured NO with sample rates for
^S,
rates greater than 300 SCCM. However, conditions may not be adequate to
insure complete quenching of all reactions concerning carbon monoxide burnout.
C-17
-------
The possible presence of carbon in the prove and its effect on NO
reduction remains an open question. Sample lines must be kept short to reduce
NO to NO2 conversion; this effect must be significant with typically 10 percent
conversion even with short sample lines.
In the results presented in Figures C-4 to C-7 the probe was located
well out near the bowl edge and somewhat downstream of the jet impingement
point. It seems reasonable to hypothesize that samples collected at this point
will be lean and will, therefore, be representative of regions of high rates of
NO formation. As time proceeds it can be assumed that gases in the vicinity of
the sample valve will eventually show a decrease in NO concentration for several
possible reasons:
- as time proceeds work extracted by the piston will cool the zones
and reduce the NO production rate.
- as the temperature drops further mixing with available air will serve
to dilute the NO concentration faster than production can compensate
- as time proceeds the sampling point will begin to experience zones
containing "younger" fuel; that is, fuel which entered the system
after ignition. Such zones are likely to have been initially richer
in composition before reaching the sampling point because their
fuel was not consumed in the intial stages of heat release. Hence
such zones would never experience high NO levels even though at
the time of sampling these may have reached a lean (but cool) con-
dition .
Indeed all of the test conditions exhibited this rapid rise to high NO values
followed by decay to quite low values.
In all cases the NO concentration at the time of exhaust valve
opening was below the exhaust emission measurements by typically 1000 ppm.
An obvious and easy explanation for this would be that there is a significant
loss of NO in the sampling system by some unspecified mechanism. While
not ruling this out completely it appears not to be the explanation. Although
the sampling system has never been calibrated under complete simulation of
diesel conditions it has been satisfactorily checked against high pressure cal-
ibration gases. However, the possibility of carbon contamination of the valve
remains an open question.
C-18
-------
Without obtaining further data at different probe locations it is
impossible to be definitive in providing an explanation for the observed' be-
havior. However, it is possible to construct a variety of plausible explana-
tions which could represent the actual situation. For example, as suggested
above it may be that late in the cycle the probe is sampling relatively "young"
zones where the fuel came into the system late in the injection period. As
a consequence such a zone would experience rather rich combustion early
in its history with attendant low NO levels . Subsequently such a zone ex-
periences dilution by air entrainment to lean conditions but the temperatures
are too low at that time to produce significant NO. A necessary condition
for this or any other mechanistic explanation to be valid is that there exists
considerable unmixdeness late into the expansion stroke. Otherwise the
late sample measurements would have to reflect the entire cylinder content
and hence the exhaust value. Now the measured fuel/air ratios (deduced from
the excess oxygen) confirm that this unmixedness does indeed exist. Sample
equivalence ratios at the exhaust valve opening point are presented in
Figures C-4, 5, 6, and C-7. The mass average of these values must be
0.5 (the exhaust value).
The distribution of equivalence ratio with crank angle is presented
in Figures C-8 and C-9 for several swirl levels and at two valve insertions.
An interesting peculiarity is the dip in equivalence ratio after the initial spike
followed by a period of growing richer and then a final dilution period. A
possible explanation again rests on the concept of "younger" and previously
richer zones appearing in the neighborhood of the valve later in the cycle.
Another explanation might be that this behavior reflects a jet interaction
phenomenon where at approximately 25 ca the sample position begins to
experience the presence of a neighboring jet. Figure C-7 provides additional
evidence that unmixedness persists late into the cycle. Data is presented
there for different valve insertion depths. Substantial gradients exist at
9= 150° in both NO and equivalence ratios .
C-19
-------
O
i
1.0
^
O .75
Of.
UJ
o
w, .50
ID
o
UJ
LJ .25
_j
Q.
UJ
0
.50
Speed- 1500 rpm
Load: 0 = . 5
Timing: 0j, = - 20°
Valve Insertion: 11/16"
0 50
Crank Angle , Degrees
100
Figure C-8. Dependence of Sample Equivalence Ratio on Swirl
- 11/16" Valve Insertion
-------
1.0 r
Speed: 1500 rpm
Load: <£ = . 5
Timing: 0|n.= -20°
Valve Insertion: Flush
Swfrt Level
A High
o Medium
o Low
.75
.50
o
UJ
5
S .25
lu
-50
+50 +JOO
Crank Angle, Degrees
+150
Figure C-9-
Dependence of Sample Equivalence Ratio
on Swirl - Flush Valve Position
C-21
-------
OS. CONCLUSIONS
Experience gained from in cylinder sampling suggests, that if
properly applied, this technique could provide valuable information which
would aid in the development of improved mathematical models. However,
certain questions regarding sample integrity remain to be answered. The
tentative conclusions suggested by the data are:
• Significant unmixedness persists throughout the cycle
• It does appear that the mixing process cannot be described by a
random homogeneous turbulence field with variations of species
concentration occurring in a random fashion in space and time.
The mixing process appears well ordered and repeatable, that has
its origins in the well-ordered Jet mixing patterns occurring only
in the cycle.
• There is considerable need to generate much more data, particularly
with the sample valve inserted to cover the total free volume. Data
obtained with the valve flush with the cylinder head is of limited
value because of the significant vertical gradients.
C-22
-------
REFERENCES
1. Wilson, R.P. Jr., Waldman, C.H., and Muzio, L.J., "Foundation for
Modelling NOX and Smoke Formation in Diesel Flames," Ultrasystems
Final Report, APRAC Project CAPE 20-17, (1974).
2. Nightingale, D.R., "A Fundamental Investigation into the Problem of NO
Formation in Diesel Engines , " Paper presented at the SAE Off. Highway
Vehicle Meeting Milwaukee, Sept. 8-11 (1975), SAE 750848.
C-23
-------
APPENDIX D
SAMPLING VALVE DATA TEST SERIES 3
Sampling value data test series 3 are presented in this appendix under the
following order.
No. of
points
8
7
7
8
7
8
4
1
1
4
Speed
(rpm)
1500
1500
1500
1500
1500
1500
1500
1500
2100
1500
Load
0
0.6
0.6
0.6
0.6
0.6
0.35
0.35
0.75
0.6
0.6
Timing
(ATDC)
-20
-20
-20
-12
-28
-20
-20
-20
-20
-20
Swirl
Level
Medium
High
Low
Medium
Medium
Medium
Low
Medium
Medium
Medium
Fuel
Injector
6
6
6
6
6
6
6
6
6
4
D-l
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-20
£• U «
Medium
6
NO
X
CO
UHC
co2
0,
Exhaust Concentrations
2460 ppm
0.005 percent
0
8 . 6 percent
8 . 7 percent
Sample Valve Location
R = 2.7 cms
9 = 0 from jet axis
Z = 0 depth (cms)
Sampling
Time
\)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
137
90
600
1560
2146
2688
2937
3650
4240
4070
CO
%
-003
.080
.430
.500
.450
.900
.400
.180
.150
.360
UHC
ppm
«*»*V^HM
730
950
1250
900
420
350
0
0
0
0
co2
%
0.50
0.50
2.95
3.90
3.70
7.20
7.50
8.10
8.00
9.00
°2
%
19.3
19.0
15.8
12.0
8.5
10.0
8.5
8.0
8.1
6.8
D-2
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm
0.6
-20-
Medium
6
NO
X
CO
UHC
CO2
2411 ppm
0.006 percent
0
7 .8 percent
8.3 percent
Sample Valve Location
R = 2.7 cms
9 = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
139
123
660
820
1489
2090
2188
3150
3800
4120
CO
%
0.001
0.003
5.10
5.50
4.50
0.70
0.65
0.66
0.60
0.50
UHC
Ppm
360
5400
9500
12500
2500
950
850
650
0
0
co2
%
0.30
0.58
6.70
6.40
6.10
7.00
8.00
7.70
7.20
7.10
0
0/2
/a
19.2
14.5
4.1
2.9
7.5
9.3
8.3
8.5
6.0
5.5
D-3
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°1
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
140
143
1370
3542
3280
3980
3950
4035
4039
3300
Exhaust Concentrations
1500 rpm
0.6
-20.0
Medium
6
Sample Valve
R = 3.9 cms
9 = 0° from
Z = 0 depth
CO
%
0.005
0.004
0.4
2.7
2.5
1.9
1.6
0.85
0.94
0.6
N0x
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
500
1250
650
500
0
0
0
0
0
0
2431 ppm
0.007 percent
0
9 . 1 percent
8 . 7 percent
co2
%
0.7
0.4
4.9
8.4
8.9
9.3
9.1
9.4
9.6
9.4
°2
%
19.5
20.0
14.0
7.0
6.3
5.8
6.5
7.1
6.1
6.5
D-4
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-20
£* \J •
Medium
6
Sample Valve
Exhaust
NO
X
CO
UHC
co2
°2
Location
Concentrations
2435 ppm
0.01 percent
0
8 .5 percent
8 . 6 percent
R = 3.9 cms
9 = 15° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
77
84
350
1480
2190
2730
2880
3170
3840
3890
CO
%
0.005
0.65
0.65
1.6
1.7
1.55
1.5
1.15
0.86
0.8
UHC
ppm
1500
6500
8000
4000
1700
1200
1200
1050
0
0
co2
%
0.47
0.8
1.0
2.1
4.6
6.0
7.4
8.1
8.9
9.3
°2
%
20.0
16.5
13.5
10.3
7.8
7.1
6.9
6.5
6.2
5.6
D-5
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
pom
315
327
680
2546
3440
2360
3570
2940
2750
2027
Exhaust Concentrations
1500 rpm
0.6
-20.
Medium
6
Sample Valve
R = 3.9 cms
9 = 30° from
Z = 0 depth
CO
%
0.12
1.1
2.6
4.5
4.2
6.6
2.0
1.4
0.6
0.9 '
N0x
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
130
400
1600
1000
220
200
160
130
110
90
2383 ppm
0 . 2 percent
0
8 . 1 percent
8 . 7 percent
co2
%
0.9
3.6
4.2
7.0
9.0
8.1
9.1
9.5
9.0
8.1
°2
%
18.9
15.0
11.8
7.0
5.5
3.5
5.5
5.3
6.6
9.5
D-6
-------
Speed
Equivalence Ra
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
-22
-12
- 7
- 2
3
8
13
23
33
43
tions Exhaust Concentrations
(nominal)
N0x
ppm
133
123
890
1112
2100
2950
3300
3800
4018
3700
1500 rpm NO
°-5 coX
-20
Medium ,H^C
6 C02
°2
Sample Valve Location
R = 3.9 cms
9 = 30 from jet axis
Z = 0 depth (cms)
CO UHC
— 2L_
0.2
1.9
10.0
6.6
5.4
2.2
1.4
0.8
0.8
0.8
ppm
600
1050
5750
3000
2000
1050
900
700
0
0
2426 ppm
0 . 1 percent
0
8 . 5 percent
8 . 8 percent
CO2
S. —
0.61
1.55
6.50
7.00
7.00
8.60
8.80
9.10
9.50
9.20
°2
_A_
19.5
16.5
2.7
4.9
5.5
6.4
6.8
6.9
6.0
6.0
D-7
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
280
300
2220
2980
3015
3007
3800
4118
3700
3590
1500 rpm
0.6
-20.
Medium
6
Sample Valve
R = 3.9 cms
9 = 45° from
Z = 0 depth
CO
/y
0.15
0.28
2.35
3.80
3.40
2.00
1.35
0.80
0.45
0.60
Exhaust
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
820
920
660
600
0
0
0
0
0
0
Concentrations
2396 ppm
0.006 percent
0
8.5 percent
8.8 percent
co2
•V^H^^B^
1.35
1.50
9;10
9.40
9.00
9.00
9.00
8.00
9.50
9.00
18.0
17.5
5.3
3.9
4.5
5.7
6.3
6.1
6.2
6.2
D-8
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 . 6
Timing (A.TDC) -20.
Swirl Level Medium
Fuel Injector 6
Exhaust Concentrations
2300 ppm
0.15 percent
0
8.8 percent
O7 8.8 percent
NO
x
CO
UHC
Sample Valve Location
R = 5.6 cms
9 = 30 from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
92
75
821
370
1160
2140
2170
2733
2800
3050
CO
0.005
0.35
10.00
10.00
10.00
9.00
5.00
1.80
1.30
1.20
UHC
PPm
100
100
100
120
0
0
0
0
0
0
co2
0.67
0-65
6.50
4.00
6.40
7.40
8.00
8.40
9.60
9.90
°2
19.4
19.1
1.3
0.4
1.2
1.9
4.0
6.1
5.0
4.7
D-9
-------
Input Conditions
Speed • 1500 rpm
Equivalence Ratio (nominal) 0.6
Timing (ATDC) -20.
Swirl Level High
Fuel Injector 6
Exhaust Concentrations
2100 ppm
0.005 percent
0
8 .5 percent
8.6 percent
NO
x
CO
UHC
CO.
Sample Valve Location
R = 2.7 cms
9 = 0° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
115
143
1560
3400
3419
2280
1312
1581
2233
2355
CO
%
0.002
0.45
0.75
2.20
8.80
10.00
10.00
7.00
3.80
2.00
UHC
ppm
320
1200
1500
400
0
0
0
0
0
0
co2
%
0.40
1.85
5.50
8.60
8.40
8.40
8.50
10.10
11.60
11.80
°2
%
19.8
17.0
12.3
6.0
1.0
0.1
0.0
0.0
0.3
1.3
D-10
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-20
t* \j •
High
6
N0x
CO
UHC
co2
°2
Exhaust Concentrations
2450 ppm
0 .2 percent
0
8 . 6 percent
8 . 9 percent
Sample Valve Location
R = 2.7 cms
9 = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°1
-22
-17
-12
- 7
- 2
3
8
13
28:
38
NOX
ppm
112
40
180
860
2010
3000
3220
2210
2250
950
CO
%
0.05
0.40
5.40
9.00
10.00
4.40
6.50
8.00
4.00
2.40
UHC
ppm
1600
4500
9200
10200
14000
2600
1100
1000
1000
770
co2
%
0.6
0.8
5.0
7.8
6.6
9.9
10.0
9.9
12.0
12.0
°2
%.._
20.5
19.0
9.2
2.6
1.2
2.7
0.8
0.2
0.6
1.2
D-ll
-------
Input Conditions
Speed
Equivalence Ratio (nominal
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-20.0
High
6
Sample Valve
Exhaust
NOX
CO
UHC
co2
°2
Location
Concentrations
2534 ppm
0.25 percent
0
8 . 0 percent
8 .4 percent
R = 3.9 cms
9 = 0° from jet axis
Z = 0 depth (cms)
Sampling
Time
( )
-22
-12
- 7
- 2
3
3
13
23
33
43
NOX
ppm
264
330
1518
1354
2360
4300
3234
3340
2900
2248
CO
%
0.08
0.17
0.46
3.8
4.3
1.7
1.8
1.4
1.0
1.0
UHC
ppm
160
180
180
150
150
130
130
130
-0
60
co2
%
0.8
1.5
5.1
6.1
6.9
9.0
9.3
10.5
11.0
11.0
°2
%
18.6
17.6
12.4
8.4
7.0
6.4
4.6
3.7
3.5
3.6
D-12
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 . 6
Timing (ATDC) -20.
Swirl Level High
Fuel Injector 6
N0x
CO
UHC
o.
Exhaust Concentrations
2200 ppm
0.25 percent
0
8.6 percent
8.6 percent
Sample Value Location
R = 3.9 cms
9 = 15 from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
NOX
ppm
265
227
1968
1852
1936
3637
3419
2324
1800
1385
CO
%
0.14
0.36
2.2
1.2
8.6
2.0
2.3
1.3
1.0
0.8
UHC
ppm
190
450
1700
500
400
380
280
230
170
150
C02
%
0.8
1.0
5.6
6.1
7.4
8.9
7.9
8.1
7.0
6.0
°2
%
19.4
18.9
11.5
10.9
2.1
6.1
6.6
7.5
9.1
9.6
D-13
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
ft
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
pom
152
350
410
1085
2130
1745
1530
1490
1470
1140
Exhaust Concentrations
1500 rpm
0-6
-20.
High
6
Sample Valve
R
9
Z
= 3.9 cms
= 45° from
= 0 depth
CO
0.05
0.60
1.00
5.20
1.20
2.20
1.40
1.20
0.55
0.45
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
58
78
120
240
95
86
80
90
80
60
2200 ppm
0 . 3 percent
0
8 . 4 percent
8 . 6 percent
co2 i
1
3
3
5
5
7
6
7
7
6
.3 1
.4 1
.2 1
.2
.8 1
.0
.8
.2
.2
.8 1
19.5
15.2
15.2
8.8
11.2
8.6
9.4
9.6
9.5
10.4
D-14
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0.6
Timing (ATDC) -20.
Swirl Level High
Fuel Injector 6
Exhaust Concentrations
NO 2190 ppm
x
CO 0.3 percent
UHC 0
CO 8.7 percent
^
O9 9.0 percent
Sample Valve Location
R = 5.6 cms
9 = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
170
256
344
660
1750
1705
1650
1414
1036
650
CO
%
0.20
0.50
0.70
7.00
5.40
2.60
1.50
0.73
0.45
0.15
UHC
ppm
900 .
1250
1600
1250
1000
800
740
470
0
0
C02
%
0.69
2.10
2.60
4.60
7.70
6.50
5.90
6.00
4.90
2.30
°2
%
19.5
17.5
16.9
8.2
6.0
9.5
11.2
12.5
13.2
14.0
D-1S
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm
0.6
-20.
Medium
4
NO
X
CO
UHC
C00
2190 ppm
0.05 percent
0
8.9 percent
9.0 percent
Sample Valve Location
R = 5.6 cms
9 = 0 from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-22
-12
- 7
- 2
3
3
13
23
33
43
N0x
ppm
67
72
190
470
1206
2025
2140
1902
2045
1925
CO
%
0.04
1.50
5.00
1.10
3.00
6.00
3.40
1.50
0.65
0.60
UHC
ppm
120
150
7500
3000
600
400
390
300
260
200
co2
%
0.3
0.8
3.2
2.3
6.0
8.5
9.0
9.5
9.6
9.4
°2
%
20.5
18.3
12.5
17.2
10.0
4.0
4.4
6.0
7.2
7.0
D-16
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
x
ppm
135
70
188
432
718
770
916
917
1038
980
1500 rpm
0.6
-20-
Low
6
Samole Valve
R = 2.7 cms
9 = 0° from
Z = 0 depth
CO
0.05
0.80
1.80
1.70
0.60
0.23
0.15
0.10
0.08
0.09
Exhaust
NOX
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
300
2500
4000
4000
800
500
0
0
0
0
Concentrations
1685 ppm
0 . 1 percent
0
8 . 2 percent
8.3
C°2
0.50 1
1.02 1
2.80 1
3.30 1
4.30 1
3.60 1
3.70 1
3.80 1
3.90 1
4.20 1
19.3
17.9
14.1
13.5
13.4
14.8
15.0
14.7
14.5
14.2
D-17
-------
Input Conditions Exhaust Concentrations
Speed 1500 rpm NO 1830 ppm
Equivalence Ratio (nominal) 0.6 rnx n , nprrfant
Timing (ATDC) -20. ~° ° ' 1 Percent
Swirl Level Low ^ "
Fuel Injector 6 2 '
O 9.0
Sample Vavle Location
R = 2.7 cms
9 = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
28
38
N0x
ppm
91
360
800
856
1010
1113
1200
1223
1190
CO
%
0.05
1.05
1.00
0.85
0.30
0.10
0.11
0.11
0.10
UHC
ppm
540
1400
900
900
700
440
450
350
-0
C02
%
0.45
3.20
4.30
4.40
4.40
4.00
3.90
4.00
3.90
°2
%
20.0
16.0
13.5
13.5
14.0
15.5
15.7
15.9
15.5
D-18
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 .6
Timing (ATDC) -20.0
Swirl Level Low
Fuel Injector 6
Exhaust Concentrations
NO 1805 ppm
>\
CO 0.15 percent
UHC 0
COn
o
8.5 percent
8.8 percent
Sample Valve Location
R = 3.9 cms
0 = 0° from j
Z = 0 depth (cms)
0 = 0° from jet axis
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
130
172
663
767
1460
1800
1817
1830
1590
1370
CO
%
0.1
3.4
6.0
4.4
2.6
1.5
0.95
0.50
0.30
0.25
UHC
ppm
210
2000
2500
1500
400
300
250
200
200
200
co2
%
.56
2 = 3
5.2
5.0
6.0
6.3
6.7
6.2
5.9
4.3
°2
JL
19.5
14.3
8.0
8.9
9.5
10.1
10.5
11.0
11.8
14.0
D-19
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 . 6
Timing (ATDC) -20.
Swirl Level Low
Fuel Injector 6
Exhaust Concentrations
NO 1780 ppm
JC
CO 0 .15 percent
UHC 0
CO2 8.6 percent
O9 8.8 percent
Sample Valve Location
R = 3.9 cms
6 = 15° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
PPm
135
224
760
1245
1295
1935
1880
1790
1750
1684
CO
%
0.13
8.5
10.0
7.0
3.6
1.1
0.8
0.4
0.36
0.4
UHC
pprn
350
4000
8000
3000
1200
750
650
450
400
210
CO2
%
0.43
4.6
6.5
7.2
7.5
6.8
6.0
6.0
6.0
6.0
°2
%
20.1
7.0
2.0
4.5
7.2
11.0
11.9
12.5
12.5
13.0
D-20
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
x
ppm
173
198
1681
2005
2390
2160
2142
1867
1493
1380
1500 rpm
0.6
-20.
Low
6
Sample Valve
R = 3 .9 cms
9 = 30° from
Z = 0 depth
CO
%
0.05
0.25
3.4
3.8
2.4
1.4
0.7
0.2
0.14
o'.13
Exhaust
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
200
260
400
350
240
200
200
170
110
100
Concentrations
1555 ppm
0 . 1 percent
0
8 . 6 percent
8 . 8 percent
co2
%
0.75
1.25
7.4
7.2
7.6
6.8
5.8
5.1
5.0
4.3
19.8
18.5
8.4
7.4
9.5
10.9
11.8
13.0
13.3
14.5
D-21
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
x
ppm
ISO
185
350
290
1136
1410
1220
1250
1040
755
Exhaust Concentrations
1500 rpm
0.6 .
-20.
Low
6
Sample Valve
R = 3.9 cms
9 = 45° from
Z = 0 depth
CO
/O
0.08
0.26
1.2
3.6
4.0
2.0
1.5
0.4
0.2
0.2
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
240
400
550
460
360
310
160
110
48
40
1717 ppm
0. 15 percent
0
8 . 6 percent
8 .8 percent
co2
%
0.66
0.9
3.2
5.5
7.2
7.5
6.8
4.3
4.0
3.6
0,
i
Ji
19
19
15
8
6
7
8
5
6
6
,5
.5
,5
,5
,5
2
D-22
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0.6
Timing (ATDC) -20.
Swirl Level Low
Fuel Injector 6
NO
x
CO
UHC
CO0
Exhaust Concentrations
1670 ppm
0.1 percent
0
8.5 percent
8.6 percent
Sample Valve Location
R = 5.6 cms
Q = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
120
109
200
471
1011
1790
1690
1580
1100
1100
CO
%
0.20
0.17
0.46
4.50
7.20
3.80
2.50
1.60
1.10
0.50
UHC
ppm
5000
7300
9500
8100
6700
5700
4700
4100
3100
2600
co2
%
0.61
0.52
1.50
2.60
5.40
6.60
6.20
6.20
5-60
4.80
°2
%
19.5
19.8
18.5
13.2
6.7
7.5
9.3
10.5
12.0
12.8
D-23
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0. 6
Timing (ATDC) -12.
Swirl Level Medium
Fuel Injector 6
Exhaust Concentrations
NO 1300 ppm
X
CO 0.06 percent
UHC 0
CO2 8.0 percent
O9 8.4 percent
Sample Valve Location
R = 2.7 cms
9=0° from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
80
90
280
800
1195
1330
1760
2074
1948
CO
%
0.04
0.06
0.84
0.80
0.83
1.00
0.46
0.47
0.30
UHC
ppm
300
500
1300
2000
600
350
0
0
0
co2
%
0.48
0.40
2.30
4.10
4.60
4.90
6.20
6.80
7.00
°2
%
19.1
19.1
16.2
13.3
12.5
12.0
10.9
9.5
8.9
D-24
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 . 6
Timing (ATDC) -12
Swirl Level Medium
Fuel Injector 6
Exhaust Concentrations
NO 1300 ppm
x
CO 0.05 percent
UHC 0
CO.
8.4 percent
8 .6 percent
Sample Valve Location
R = 2.7 cms
9 = 30 from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-17
-12
- 7
- 2
3
8
13
28
38
NO
X
Dom
56
94
155
280
820
890
990
850
1160
1250
CO
%
0.05
0.25
1.50
2.00
5.00
4.20
6.50
2.50
0.35
0.20
UHC
ppm
420
480
500
1800
2300
3000
7100
1200
580
400
C02
%
0.7
1.2
2.5
3.4
8.0
7.2
7.0
6.0
5.4
5.3
°2
%
20.0
19.5
18.2
14.2
4.8
6.0
6.0
9.5
13.0
13.5
D-25
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-12.
Medium
6
NO
X
CO
UHC
\J 11\_/
co2
Exhaust Concentrations
1254 ppm
0 . 1 percent
n
u
7 . 9 percent
Sample Valve Location
R = 3.9 cms
9 = 0 from jet axis
Z = 0 depth (cms)
8.6 percent
Sampling
Time
(o)
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
102
58
70
238
984
1094
1394
1350
1290
CO
%
.06
.09
.25
.35
.60
.30
.20
.15
.23
UHC
ppm
175
250
700
400
150
150
85
-0
-0
co2
%
0.35
0.36
0.60
1.20
3.40
3.60
4.50
6.00
6.40
°2
%
19.5
19.0
18.2
17.5
14.7
15.1
13.2
12.0
11.0
D-26
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
i 9
- 1£ .
Medium
6
Sample Valve
Exhaust
NO
X
CO
UHC
co2
°2
Location
Concentrations
1412 ppm
0.15 percent
0
8 .0 percent
8 .4 percent
R = 3.9 cms
9 = 15 from jet axis
Z = 0 depth (cms)
Sampling
Time
(O)
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
200
108
236
280
370
1014
923
718
737
CO
%
0.10
4.00
7.60
5.20
3.70
1.20
0.40
0.28
0.19
UHC
ppm
300
4500
8000
7000
3900
1100
500
320
210
co2
%
0.4
1.6
3.8
3.5
3.8
4.2
3.2
3.4
3.3
°2
%
19.5
14.0
7.0
9.5
11.9
12.5
15.1
15.0
-15.2
D-27
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-12
- 7
- 2
3
8
13
23
33
43
NO
x
pom
197
201
852
2160
1880
1470
1150
976
912
1500 rpm
0.6
-12.
Medium
6
Sample Valve
R = 3.9 cms
9 = 30° from
Z = 0 depth
CO
%
0.05
0.15
4.60
2.80
2.50
1.70
0.28
0.15
0.15
Exhaust
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
200
220
300
250
200
190
190
160
110
Concentrations
1395 ppm
0 . 1 percent
0
8 . 0 percent
8 . 8 percent
co2 c
% J
0.54 19
0.60 19
7.40 5
8.50 5
7.70 8
7.00 9
5.00 12
4.40 13
4.40 13
5.2
5.5
8.0
9.5
D-28
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (A.TDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm NO
R = 3.9 cms
9 = 45° from jet axis
Z = 0 depth (cms)
1085 ppm
-12.
Medium
6
Sample Valve
CO
co°2
°2
Location
0
0
8
8
.01 percen
.5 percent
. 6 percent
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
83
56
56
72
1055
1730
1328
1015
785
560
CO
%
0.05
0.05
0.05
0.05
0.50
0.40
0.40
0.20
0.10
0
UHC
ppm
94
100
120
130
110
100
100
88
55
52
CO2
%
0.7
0.7
0.8
1.2
4.8
5.4
5.0
5.5
4.4
0
°2
%
19.6
19.8
19.6
19.5
13.5
13.0
13.5
12.5
14.5
0
D-29
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing &TDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-12
- 7
- 2
3
8
13
23
33
43
ppm
88
77
231
430
980
1440
2080
1900
1830
1500 rpm
0.6
-12.
Medium
6
Exhaust
NO
X
CO
UHC
co2
°2
Concentrations
1230 ppm
0.008 percent
0
8 . 2 percent
9 .2 percent
Sample Valve Location
R = 5 .6 cms
8 = 0° from
Z = 0 depth
CO
"*
0.005
0.15
4.80
4.30
3.40
7.80
3.60
1.90
0.70
jet axis
(cms)
UHC
ppm
30
100
4000
850
150
60
0
0
0
co2
^MWMM^^H
0.3 :
0.3
4.2
5.2
7.7
8.0
9.5
8.5
8.0
20.5
19.2
9.5
8.5
6.1
1.8
3.8
7.2
7.9
D-30
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(Q)
-12
- 2
3
8
13
23
33
43
NO
x
opm
72
70
230
199
735
1330
1130
957
730
1500 rpm
0.6
-12.
Medium
6
Sample Valve
Exhaust
N0x
CO
UHC
co2
°2
Location
Concentrations
1209 ppm
0 . 1 percent
0
8 .2 percent
9.2 percent
R = 5 . 6 cms
9 = 30° from jet axis
Z = 0 depth
CO
%
0.09
0.09
0.65
10.00
6.80
3.80
1.15
0.50
0.40
(cms)
UHC
ppm
560
540
560
460
900
800
420
370
240
co2
%
0.35
0.30
1.80
3.80
5.40
6.10
5.80
4.60
4.00
°2
%
19.8
19.9
17.8
5.8
6.9
8.7
10.9
13.7
16.3
D-31
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(Q)
-32
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
291
201
322
1880
2722
3580
4400
4785
5400
5664
5500
1500 rpm
0.6
-28.
Medium
6
Sample Valve
R = 2.7 cms
9 = 0° from
Z = 0 depth
CO
%
0.04
0.10
1.70
1.15
0.55
0.26
0.16
0.35
0.24
0.15
0.16
Exhaust
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
200
600
2500
600
300
300
250
225
-0
-0
-0
Concentrations
4440 ppm
0.06 percent
0
7 . 9 percent
8 . 1 percent
co2
%
0.4
0.5
3.2
5.9
6.5
6.6
7.0
8.0
8.8
8.9
8.7
°2
%
19.5
19.0
14.2
11.0
10.5
10.7
10.0
8.2
7.5
7.3
7.0
D-32
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm NO
n c 2
CO
x
-28.
Medium
6 ^2
O,
3640 ppm
0 .05 percent
0
8.3 percent
8.9 percent
Sample Valve Location
R = 2.7 cms
9 = 30° from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-22
-17
-12
- 7
- 2
3
8 '
13
28
38
NO
X
ppm
240
250
580
905
1750
2460
3800
4000
4150
4200
CO
%
0.15
1.50
6.20
7.00
1.00
0.35
0.40
0.40
0.25
0.35
UHC
ppm
500
580
1400
4000
700
640
520
440
280
240
co2
%
0.6
2.2
6.2
6.2
5.2
5.8
6.5
7.4
8.0
8.1
°2
%
20.0
16.2
6.2
5.0
11.7
12.0
11.2
9.8
8.8
8.8
D-33
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0.6
Timing (ATDC) -28.
Swirl Level Medium
Fuel Injector 6
CO
UHC
CO
2
O*
Exhaust Concentrations
3710 ppm
0.15 percent
0
8.0 percent
8.6 percent
Sample Valve Location
R = 3.9 cms
9 = 0° from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-27
-17
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
440
460
2940
2490
2670
3400
3988
4720
5485
5120
CO
%
0.06
0.30
0.60
0.45
0-60
0.45
0.25
0.23
0.18
0.12
UHC
Ppm
96
400
170
160
170
160
-0
-0
-0
-0
co2
%
0.8
1.3
5.5
5.8
6.0
6.0
6.7
8.0
8.4
7.9
°2
%
18.6
17.5
11.1
12.9
11.9
11.9
11.1
9.0
8.5
9.0
D-34
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (A.TDC)
Swirl Level
Fuel Injector
Sampling
Time
(Q)
-27
-17
-12
- 7
- 2
3
8
13
23
33
43
NO
x
ppm
325
221
644
1322
1681
2317
2731
3521
3571
2490
2360
1500 rpm
0.6
-28.
Medium
6
Sample Valve
R = 3.9 cms
Exhaust
NO
X
CO
UHC
CO£
°2
Location
Concentrations
3245 ppm
0 . 1 percent
0
8 .2 percent
8 .8 percent
9 = 15° from jet axis
Z = 0 depth
CO
%
0.23
3.50
5.20
4.60
5.50
3.20
1.30
1.20
0.60
0.45
0.40
(cms)
UHC
ppm
190
270
3000
1000
570
460
420
360
250
200
180
co2
%
0.7
2.6
4.8
5.5
6.5
6.0
6.6
7.0
6.4
5.8
5.0
°2
%
19.2
12.3
8.5
9.6
6.7
8.2
10.0
9.0
10.6
11.3
13.0
D-35
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0. 6
Timing (ATDC) -28.
Swirl Level Medium
Fuel Injector 6
Exhaust Concentrations
3540 ppm
CO 0.1 percent
UHC
NO
x
CO,
0
8.0 percent
8.6 percent
Sample Valve Location
R = 3.9 cms
9 = 30 from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-32
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
560
1830
2370
2760
2440
3070
3840
3645
3300
2820
CO
%
0.15
3.20
3.40
2.00
0.76
0.70
0.46
0.30
0.34
0.26
UHC
ppm
100
300
350
240
200
160
140
130
90
80
co2
%
1.0
7.4
7.8
7.3
6.0
6.6
7.0
6.7
7.4
6.4
°2
%
18.3
7.0
6.6
7.9
10.0
10.6
10.2
9.8
10.0
10.8
D-36
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.6
-28.
Medium
6
NO
X
CO
UHC
co2
°2
Exhaust Concentrations
3600 ppm
0 . 1 percent
0
8 .7 percent
8 .5 percent
Sample Valve Location
R = 3.9 ppm
9 = 45° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-27
-17
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
215
322
2336
2340
2360
2188
2525
2804
2560
2020
CO
%
0.06
1.10
1.15
1.00
0.40
0.25
0.40
0.25
0.20
0.15
UHC
ppm
70
100
100
90
90
80
70
60
60
60
CO2
%
0.66
0.71
5.80
5.50
5.20
4.80
5.60
5.70
5.90
5.00
°2
%
19.5
19.4
11.5
12.0
12.7
13.5
12.0
11.2
11.9
13.1
D-37
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°)
-27
-17
-12
- 7
- 2
3
8
13
23
33
43
NO
x
opm
180
260
1635
1600
1555
2500
3507
3499
2590
1711
1430
Exhaust Concentrations
1500 rpm
0.6
-28.
Medium
6
Sample Valve
R = 5.6 cms
9 = 30° from
Z = 0 depth
CO
^^^^^
0.21
0.25
5.20
9.00
6.20
2.30
0.90
0.60
0.50
0.40
0.25
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
240
360
450
540
570
490
440
370
240
180
180
3823 ppm
0 . 1 percent
0
8.5 percent
8 . 8 percent
co2
fa-
0.65
0.65
6.80
6.00
6.60
6.70
6.70
6.40
5.60
4.70
2.40
19.7
20.2
6.8
4.8
6.6
8.9
11.1
10.8
12.5
14.0
14.6
D-38
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
0.35
-20
u \J •
Medium
6
Sample Valve
Engine
NO
X
CO
UHC
co2
°2
Location
Concentration
1880 ppm
0.05 percent
0
5 .4 percent
13.2 percent
R = 2.7 cms
9 = 0° from jet axis
Z = 0 depth (cms)
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
28
38
NO
X
ppm
72
75
500
1190
1285
1070
1370
2280
2420
CO
%
0.01
0.05
1.70
1.15
0.65
0.20
a. 07
0.03
0.03
UHC
ppm
250
2000
2600
900
550
500
400
250
-0
co2
%
0.05
0.15
3.80
6.00
5.30
4.40
5.00
6.00
5.80
°2
%
20.5
20.2
14oO
11.2
12.0
15.5
14.5
13.2
13.1
D-39
-------
Speed
Equivalence Ra
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-17
-12
- 7
- 2
3
8
13
18
23
28
38
Conditions Exhaust Concentrations
o (nominal)
1500 rpm NO
•35 ^ *
on n CO
Medium
NO
X
ppm
105
80
95
580
715
905
1350
1472
1810
2100
2030
2420
6
Sample
R =
9 =
Z =
CO
0.05
0.10
1.50
8.00
2.00
0.35
0.20
0.15
0.10
0.15
0.15
0.15
co2
°2
Valve Location
2.7 cms
30° from jet axis
0 depth (cms)
UHC
ppm
280
700
2800
6000
1100
600
500
450
0
0
280
250
2000 ppm
0.05 percent
850 ppm
5 . 2 percent
13.1 percent
C02
0.30
0.40
1.80
6.20
5.00
3.60
4.50
4.50
4.50
4.60
5.80
5.20
°2
19.
19.
17.
5.
13.
15.
14.
14.
14.
13.
12.
13.
D-40
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm
.35
-20.0
Medium
6
NO
X
CO
UHC
C02
o,
1840 ppm
0. 10 percent
0
5 .5 percent
12 .3 percent
Sample Valve Location
R = 3.9 cms
9=0 from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
-7
-2
3
8
13
23
33
43
NO
x
ppm
154
163
234
1715
1730
1830
1570
1630
2012
2250
CO
%
0.06
0.15
0.50
0.65
1.00
0.60
0.17
0.10
0.10
0.60
UHC
PPm
1400
1500
1300
840
900
780
760
600
400
300
co2
%
0.03
0.40
1.40
4.60
5.40
5.40
4.50
4.20
4.70
5.10
°2
%
19.5
19.0
17.8
13.1
12.6
12.5
14.0
14.5
13.5
13.2
D-41
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
Exhaust Concentrations
1500 rpm NO
:fo.o c°x
Medium UHC
6
°2
Sample
R =
9 =
Z =
CO
_JL
0.05
5.40
7.60
2.00
1.00
0.50
0.30
0.15
0.16
0.15
Valve Location
3.9 cms
15 from jet axis
0 depth (cms)
UHC
ppm
170
240
6800
1100
570
450
370
120
100
100
1718 ppm
0. 10 percent
0
5.2 percent
12.7 percent
co2 c
'° ___
0.25 19
2.30 11
4.60 7
3.60 13
4.00 15
3.60 14
3.40 15
3.00 16
2.40 17
3.20 16
>2
%_
.6
.2
.0
.0
.1
.7
.7
.0
.5
.5
D-42
-------
Speed
Equivalence Ra
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
te)
-22
-12
- 7
- 2
3
8
13
23
33
43
itions Exhaust Concentrations
D (nominal)
1500 rpm
.35
-20.0
Medium
6
NO
X
CO
UHC
°2
1765 ppm
0 .06 percent
0
5 . 1 percent
13.2 percent
Sample Valve Location
NO
X
ppm
150
158
1700
2020
1373
1017
970
1140
1045
1000
R = 3.9 cms
9=30° from
Z = 0 depth
CO
_2S_
0.06
1.10
4.20
4.80
1.40
0.53
0.24
0.18
0.18
0.16
jet axis
(cms)
UHC
ppm
170
240
500
600
300
280
240
210
150
140
co2
_^_
0.40
2.40
8.00
7.80
7.10
4.40
3.80
3.40
3.70
3.40
°2
JL_
19.8
16.0
5.5
5.0
8.8
13.5
14.5
15.2
15.2
15.6
D-43
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
35
• *J w
-20.0
Medium
6
NO
X
CO
UHC
CO,,
Exhaust Concentrations
1824 ppm
0.07 percent
0
5 .0 percent
13.2 percent
Sample Valve Location
R = 3.9 cms
o
9 = 45 from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
147
148
825
2145
2268
1600
1300
1430
1400
1275
CO
%
0.05
0.70
0.60
1.00
2.00
1.00
0.30
0.10
0.08
0.07
UHC
ppm
100
120
125
105
98
95
80
70
70
70
co2
%
0.30
0.33
3.80
7.00
7.40
5.60
4.00
4.10
4.00
3.70
°2
%
19.8
19.8
14.5
9.7
8.3
11.5
14.0
14.5
15.0
15.2
D-44
-------
Speed
Equivalence Ra
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
>nditions Exhaust Concentrations
3 (nominal)
NO
X
ppm
211
270
183
620
940
2150
2647
2976
2704
2530
1500 rpm NO
35 X
^0.0 C0
Medium UHC
6
°2
Sample
R =
9 =
Z =
CO
— /0
0.20
1.00
2.00
1.20
2.70
4.00
2.20
0.55
0.30
0.25
Valve Location
5 .6 cms
0 from jet axis
0 depth (cms)
CO
UHC 2
PPm °/°
50 0.65
125 1.40
550 2.60
200 2.00
150 4.80
95 7.50
80 7.70
0 8.00
0 7.00
0 6.50
2156 ppm
0.15 percent
0
5 . 3 percent
12.0 percent
°2
£
19.5
17.8
15.8
17.1
11.2
5.5
7.0
8.2
9.0
9.4
D-45
-------
Input Conditions
Exhaust Concentrations
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-12
- 7
-2
3
8
13
23
33
43
NO
x
ppm
160
132
887
570
1200
1694
1860
1540
880
928
1500 rpm
.35
-20.0
Medium
6
Sample Valve
NO
X
CO
UHC
co2
00
2
Location
2200 pp
0.09 pe
0
6.0 pen
12.2 pen
R = 5 . 6 cms
9 = 30° from jet axis
Z = 0 depth
CO
%
0.11
0.15
4.00
10.00
5.20
1.05
0.65
0.30
0.20
0.20
(cms)
UHC
ppm
250
400
360
1000
310
280
240
160
70
90
C°2
%
0.32
0.32
6.60
4.60
6.60
6.50
5.80
4.60
3.40
4.40
20.0
20.1
8.1
4.1
6.5
11.4
12.5
13.8
17.2
17.0
D-46
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
Sampling
Time
(o)
-22
-17
-12
- 7
- 2
3
8
13
18
28
38
NO
1500 rpm
35
• w *J
-20.0
Low
6
N0v
X
CO
UHC
C°2
°2
1570 ppm
0.10 percent
800 ppm
5 .4 percent
13.0 percent
Sample Valve Location
R = 2 .7 cms
9 = 30° from
Z = 0 depth
CO
%
0.00
0.00
0.60 1
jet axis
(cms)
UHC
ppm
310
480
700
1.70 2500
1.00
0.30
0.15
0.15
0.15
0.10
0.15
900
640
500
460
380
300
240
CO,,
2
%
0.60
0.30
1.00
4.40
4.60
4.00
3.80
3.30
3.20
3.00
3.20
0,
2
%
20.3
20.0
18.5
13.5
13.5
15.2
16.0
16.2
16.2
16.5
16.7
D-47
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
1500 rpm
.35
-20.0
Low
6
NO
X
CO
UHC
co2
o.
1539 ppm
0 .10 percent
0
5.6 percent
12.8 percent
Sample Valve Location
R =3.9 cms
9=0 from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
-7
-2
3
a
13
23
33
43
NO
X
oom
76
155
622
930
1489
1618
1537
1623
1540
1255
CO
%
0.06
3.00
7.00
4.40
1.80
0.75
0.30
0.18
0.10
0.12
UHC
ppm
500
5000
5500
1700
770
660
610
500
0
0
co2
%
0.35
2.10
6.00
6.70
6.10
5.40
5.10
5.00
4.40
4.00
°2
%
19.9
15.0
6.0
8.5
11.1
13.1
13.8
14.2
14.2
14.8
D-48
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
1500 rpm
.35
-20.
Low
6
NO
X
CO
UHC
CO,,
Exhaust Concentrations
1463 ppm
0.01 percent
0
5.1 percent
13.4 percent
Sample Valve Location
R = 3.9 cms
9 = 15 from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
163
138
462
1120
1203
1153
1178
950
960
1110
CO
%
0.06
5.40
10.00
5.20
1.30
0.70
0.28
0.20
0.18
0.20
UHC
ppm
200
240
7600
1800
540
420
320
250
150
110
C°2
%
0.31
2.80
5.00
6.20
5.20
4.40
4.20
4.00
4.00
3.80
°2
%
20.1
11.5
3.8
7.3
11.5
13.5
14.7
15.2
15.1
15.5
D-49
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
-Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
Exhaust Concentrations
1500 rpm
.35
-20.0
Low
6
Sample Valve
R = 3.9 cms
NOx
CO
UHC
co2
°2
Location
1620 ppm
0.06 percent
0
5 . 1 percent
13.2 percent
9 = 30° from jet axis
Z = 0 depth
CO
0.04
0.20
1.40
1.60
1.00
0.27
0.10
0.07
0.08
0.08
(cms)
UHC
ppm
200
250
310
330
280
240
200
180
140
130
co2
0.45 19
0.66 19
6.60 10
7.20 9
6.80 9
5.00 ' 12
4.60 14
4.20 14
3.60 15
3.60 15
°2
.9
.1
.5
.5
.8
.9
.0
.5
.2
.6
D-50
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Exhaust Concentrations
Sampling
Time
(o)
-22
-12
-7
-2
3
8
13
23
33
43
ppm
300
260
711
1080
1570
1870
1950
2640
2884
2920
1500 rpm NO
1) .75 X
-20.0 CO
Medium UHC
6 C02
°2
Samole Valve Location
R = 5.6 cms
9=0 from jet axis
Z = 0 depth (cms)
CO UHC
'°
0.46
7.10
10.0
10.0
10.0
7.0
4.7
2.5
2.2
2.6
ppm
35
1250
1000
800
250
140
0
0
0
0
2420 PPm
0.35 Perce
0
10.8 Peres
4.5 Peres
co2 o2
— —
1.2
4.5
6.2
6.6
7.0
8.5
9.3
10.5
11.4
11.6
— '2. —
19.5
7.2
1.2
0.1
0.3
2.5
4.0
3.2
2.3
1.3
D-51
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(o)
-22
-12
- 7
-2
3
8
13
23
33
43
NOX
ppm
133
680
1457
1366
1740
1370
1560
3080
3500
4020
Exhaust Concentrations
2100 rpm
.60
-20.0
Medium
6
Sample Valve
NO
X
CO
UHC
co2
°2
Location
2440 ppm
0.35 percent
0
8.5 percent
8 . 3 percent
R = 5 . 6 cms
9=0 from
Z = 0 depth
CO
%
0.006
0.200
2.300
5.700
10.000
5.000
0.800
1.070
0.760
0.500
jet axis
(cms)
UHC
ppm
500
500
1700
550
450
300
0
0
0
0
CO0
2
%
0.40
1.20
7.90
8.70
7.50
8.50
9.00
11.00
11.50
11.3
on
2
%
20.2
19.0
7.4
3.9
0.4
5.0
7.0
4.2
3.8
4.2
D-52
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0 . 6
Timing (ATDC) -20.
Swirl Level Medium
Fuel Injector 4
CO
UHC
co2
O.
Exhaust Concentrations
2253 ppm
0 .05 percent
0
8.9 percent
8.9 percent
Sample Valve Location
R = 2.7 cms
9 = 0° from jet axis
Z = 0 depth (cms)
Sampling
Time
(0)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
153
130
300
1080
1640
2100
2340
2411
2600
2722
CO
%
.06
.40
.65
.10
.20
.93
.85
.25
.04
.20
UHC
ppm
250
600
1000
4000
1700
150
150
120
100
95
co2
%
1.0
0.8
1.3
3.5
4.3
6.0
6.4
5.7
5.9
7.2
°2
%
19.5
19.9
19.1
16.8
15.0
11.5
11.0
12.5
12.2
11.5
D-53
-------
Input Conditions
Speed 1500 rpm
Equivalence Ratio (nominal) 0.6
Timing (ATDC) -20.
Swirl Level Medium
Fuel Injector 4
N0x
CO
UHC
CO,,
Exhaust Concentrations
2450 ppm
0 .05 percent
0
8.5 percent
8 .4 percent
Sample Valve Location
R = 2.7 cms
9 = 54° from jet axis
Z = 0 depth (cms)
Sampling
Time
(o)
-22
-12
- 7
- 2
3
8
13
23
33
43
NO
X
ppm
131
402
1500
1600
1000
1566
1840
2692
3500
3250
CO
%
0.05
0.80
2.50
5.00
4.00
0.60
0.20
0.20
0.30
0.20
UHC
ppm
125
150
300
900
1000
200
150
120
100
75
co2
%
0.6
4.0
7.0
7.0
4.8
3.0
4.9
7.1
8.1
8.1
°2
%
19.9
15.0
9.0
6.0
8.8
15.0
13.2
11.0
9.5
8.9
D-54
-------
Input Conditions
Speed
Equivalence Ratio
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(0)
-22
-12
- 7
- 2
3
8
13
23
33
43
(nominal)
NO
X
com
106
91
124
1333
2240
2258
2440
2140
1519
1360
1500 rpm
0.6
-20.
Medium
4
Sample Valve
R = 3 .9 cms
9 = 0° from
Z = 0 depth
CO
%
0.04
0.38
0.55
0.15
1.20
0.75
0.60
0.55
0.30
0.20
Exhaust
NO
X
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
ppm
200
750
850
300
280
250
210
210
160
140
Concentrations
2100 ppm
0 .07 percent
0
8 . 9 percent
8 . 6 percent
co2
%
0.4
0.5
0.7
3.5
5.4
6.6
6.0
6.5
5.4
5.0
°2
%
20.3
20.0
19.5
15.9
12.0
11.0
11.9
11.0
12.5
13.5
D-55
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(°)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
80
156
435
350
290
1008
1540
2000
1730
1670
1500 rpm
0.6
-20.
Medium
4
Sample Valve
R = 3 . 9 cms
9 = 33° from
Z = 0 depth
CO
%
0.1
3.0
8.0
6.0
3.8
0.6
0.5
0.6
0.7
"0.7
Exhaust
N0x
CO
UHC
co2
°2
Location
jet axis
(cms)
UHC
PPm
260
900
2200
2800
700
380
240
200
120
120
Concentrations
2120 ppm
0.05 percent
0
8.7 percent
8 . 8 percent
co2
%
1.0 2
3.2 ]
5.5
4.3
4.4 ]
4.4 ]
4.5 ]
7.5
9.2
9.5
?2
20.0
14.0
7.5
9.2
11.0
14.0
13.8
9.5
6.8
6.5
D-56
-------
Input Conditions
Speed
Equivalence Ratio (nominal)
Timing (ATDC)
Swirl Level
Fuel Injector
Sampling
Time
(Q)
-22
-12
- 7
- 2
3
8
13
23
33
43
N0x
ppm
197
357
362
477
1040
1450
1640
2620
2734
2380
Exhaust Concentrations
1500 rpm NO
-it.
Medium ^C
4 C°2
°2
Sample Valve Location
R = 5 . 6 cms
0 = 45 from jet axis
Z = 0 depth (cms)
CO UHC
%
0.9
3.0
10.0
10.0
10.0
4.8
1.3
1.4
1.0
1.1
ppm
110
160
380
1600
5000
1000
360
300
120
100
2364 ppm
0 . 1 percent
0
8 . 6 percent
8 . 5 percent
co2
%
0.9
3.8
4.4
3.8
5.2
6.8
5.8
7.5
8.4
8.5
%
18.8
13.5
1.8
1.2
2.1
7.1
11.2
8.2
8.1
8.5
D-57
-------
TECHNICAL REPORT DATA
(Please read las&ucnons on che reverse before completing)
1. REPORT NO. 2
EPA-460/3.-76-008-a
4. TITLE AND SUBTITLE
Study of the Oxides of Nitrogen and Carbon Formation
in Diesel Engines
7. AUTHOR(S)
C. J. Kau, T. J. Tyson, M. P. Heap
9. PERFORMING ORG \NIZATION NAME AND ADDRESS
Ultra systems, Inc.
2400 Michelson Drive
Irvine, California 92715
12. SPONSORING AGENCY NAME AND ADDRESS
U.S. Environmental Protection Agency
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Ann Arbor, Michigan 48105
3. RECIPIENT'S ACCESSIOf*NO.
5. REPORT DATE
Issued May 1976
6. PERFORMING ORGANIZATION CODE
8. PERFORMING ORGANIZATION REPORT
NO.
10. PROGRAM ELEMENT NO.
11. CONTRACT/GRANT NO.
68-01-0436
13. TYPE OF REPORT AND PERIOD COVERED
Final Report June '73-June '76
14. SPONSORING AGENCY CODE
15. SUPPLEMENTARY NOTES
Co-sponsor: Coordinating Research Council, 30 Rockefeller Plaza, New York,
New York, 10020
16. ABSTRACT
A mathematical model describing heat release and pollutant formation in direct injec- j
tion diesel engines has been developed and tested. The model includes several empiri-j
cal constants which can be tuned to fit the requirements of a particular engine. Sen-;
sitivity studies indicate that the model is most responsive to those constants which <
control fuel/air mixing. Numerical experiments strongly suggest that diffusion flames;
modelled by spherical droplet flames are unsuitable for this type of system. The j
model has been tested against results obtained with a single cylinder diesel engine.
Reasonable predictions of the influence of engine design and operation parameters on j
NO emissions were obtained. However, predictions of smoke emissions were not satis- '
factory.
KEY WORDS AND DOCUMENT ANALYSIS
DESCRIPTORS
Mathematical Modeling
Soot
Computer Program
Fuel Sprays
Internal Combustion En<
Fuel Consumption
b.IDENTIFIERS/OPEN ENDED TERMS
Mobile Sources
Exhaust Gas Recirculation
Divided Chamber Engine
:. COSA
21-07
n Field/Group
(Recipro-
Diffusion Flames
Air Pollution
Diesel Engines
Combustion
Emission
Nitric Oxide (NO)
Nitrogen Oxides
eating Engines
21-02 (Combus-
tion)
3. DISTRIBUTION STATEMENl
Release Unlimited
3ECI
I This Repon)
21. NO. OF PAGES
206
20. SECURITY CLASS (This page)
22. PRICE
EPA Form 2220-1 (9-73)
------- |