/
A DETAILED
PERFORMANCE COMPARISON OF DISTILLATE
FUELS IN THE TEXACO STRATIFIED CHARGE ENGINE
by
Gordon D. Marsh
B.S.
United States Coast Guard Academy
(1971)
Submitted in Partial Fulfillment
of the Requirements for the
Degrees of Ocean Engineer
and
Master of Science in Mechanical Engineering
at the
Massachusetts Institute of Technology
May 1976
!
Signature of Author...
.
Certified by.....
- - I -- -
.....
./
Thesis Supervisor
Certified by..
Reader Ocean Eng. Dept.
X1
Pesls
R
i\
Accepted by.......
Chair
Department Committee on Graduate Students
JUN 21IES'
Ju
21 197
-2-
A DETAILED PERFORMANCE COMPARISON OF DISTILLATE
FUELS IN THE TEXACO STRATIFIED CHARGE ENGINE
by
Gordon D. Marsh
Submitted to the Department of Ocean Engineering on
9 May 1976 in partial fulfillment of the requirements for the
degrees of Ocean Engineer and Master of Science in Mechanical
Engineering.
Abstract
A stratified charge engine employing the Texaco Controlled
Combustion System has been operated over a large range of lodd
conditions on iso-octane, methanol, and a wide boiling point
(100-600F) residual fuel.
Basic performance, emissions and
combustion parameters were measured over a range of overall
equivalence ratios from
1500, 2000 and 2500 RPM.
4 = 0.1 - 1.0 at three engine speeds;
The basic performance and emissions
data were found to vary little between iso-octane and residual
fuels,
and to compare very well with similar data collected on
the same engine design at other research facilities.
The engine
operation on methanol was not entirely satisfactory due to an
improper match between the specific
fuel injection system used
for these experiments and the design requirements imposed by the
much lower heating value and higher stoichiometric fuel-air
-3-
mass ratio of methanol.
A direct, online data acquisition system, based on a
Digital PDP 11/10 computer was developed to obtain accurate
pressure-crankangle data for further combustion
dynamic studies.
and thermo-
The acquisition program also computes a mean
pressure - crankangle diagram and the statistics associated
with cycle to cycle variation.
The mean pressure - crankangle
data is then integrated to compute indicated mean effective
pressure.
The opportunity to analyze pressure - crankangle data
in this way substantially improves the accuracy and speed of
data collection.
A simple thermodynamic model based on homogeneous charge
engine combustion has been modified to compute the heat release
and fuel fraction burnt from the pressure - crankangle data.
The problems associated with calculation of these parameters in
diesel or stratified charge engines are discussed.
Recommendations
are made for further development of the online data acquisition
system and the thermodynamic model.
Thesis Supervisor:
Dr. Joe M. Rife
Title:
Lecturer, Department of Mechanical Engineering
-4-
ACKNOWLEDGEMENTS
The author wishes to express his thanks to Dr. Joe M. Rife
for his guidance and suggestions during the development of
this thesis.
The author also wishes to acknowledge the help and advice
of Professor J. B. Heywood ,
Dr. R.J. Tabaczynski and
Mr. Don Fitzgerald, provided throughout the course of this research.
The author's personal financial support was provided by the
United States Coast Guard.
This work was supported by the U.S. Army Tank-Automotive
Systems Development Center via subcontract to M.I.T. from
Texaco, Inc., under contract number DAAE07-74-C-0268.
I thank
the program supervisors, Mr. M. Alperstein of Texaco, Inc., and
Mr. Walter Bryzik of the U.S. Army Tank-Automotive Systems
Development Center for their advice and assistance.
-5-
TABLE OF CONTENTS
Title Page....... *..-.......................
....................
1
Abstract.........
............................
.. · ·.....
·........
2
Acknowledgments..
.......-..........
·....................
····
.··
·.
4
Table of Contents
............................
List of Figures..
............................
....................
··· ··
6
List of Tables...
.................
....................
99
*....--..
eeee..e..
I
Introduction
II
Test Engine Instrumentation
III
Performance and Emission Results........ ....................
IV
Pressure Ana.lysis Volume Data Analysis.. .................... 30
V
Conclusion a*nd Recommendations .......... ....................41
References ......
....................1 0
*q
-··
·
·
·
·
............. .................... 14
·
·
·
·
·
·
21
.................... 44
Appendix I
- Performance and Emission Data Reduction Program..46
Appendix II
- Online Digital Pressure Data Acquisition
Appendix III
- Computer Analysis Programs..............
Tables...
..........................................
.....
Figures.. .
....
.eeeeeeleeeeee.eeeeeeeeeeeeeeeeeeee*eeeeee.
.........
53
.........
58
........
132
........
142
-6-
LIST OF FIGURES
Figure Numbers
1.
Texaco Controlled Combustion System (Schematic)
2.
Roosa Master Injection Nozzle
3.
Schematic of Combustion Process
4.
Simplified PVY Plot
5.
Photograph of Test Installation
6.
Single Cylinder Engine Assembly
7.
Dynamometer Hydraulic Scale
8.
Cooling System
9.
Lubrication System
10.
Inlet and Exhaust System
11.
Fuel System
12.
Needle Lift Indicator
13.
Typical Oscilloscope Trace
14.
1500 RPM Motoring Log P vs Log V
15.
2000 RPM Motoring Log P vs Log V
16.
2500 RPM Motoring Log P vs Log v
17.
Schematic of Data Acquisition System
18.
Indicated Mean Effective Pressure vs. Equivalence Ratio,
for Iso-octane.
19.
Indicated Mean Effective Pressure vs. Equivalence Ratio,
for 100-600 Fuel.
-7-
Figure Numbers
20.
Indicated Specific Fuel Consumption vs. Equivalence Ratio,
for Iso-octane.
21.
Indicated Specific Fuel Consumption vs. Equivalence Ratio,
for 100-600 Fuel.
22.
Indicated Thermal and Volumetric Efficiency vs. Equivalence
Ratio for Iso-octane.
23.
Indicated Thermal and Volumetric Efficiency vs. Equivalence
Ratio for 100-600 Fuel.
24.
Exhaust Temperature vs. Equivalence Ratio for Iso-octane.
25.
Exhaust Temperature vs. Equivalence Ratio for 100-600 Fuel.
26.
Carbon Monoxide Emissions vs. Equivalence Ratio for
Iso-octane.
27.
Carbon Monoxide Emissions vs. Equivalence Ratio for
100-600 Fuel.
28.
Hydrocarbon Emissions vs. Equivalence Ratio for Iso-octane.
29.
Hydrocarbon Emissions vs. Equivalence Ratio for 100-600 Fuel.
30.
Nitric Oxide Emissions vs. Equivalence Ratio for Iso-octane.
31.
Nitric Oxide Emissions vs. Equivalence Ratio for 100-600 Fuel.
32.
Indicated Mean Effective Pressure vs. Equivalence Ratio
for Methanol.
33.
Indicated Specific Fuel Consumption for Methanol.
34.
Indicated Thermal and Volumetric Efficiency vs. Equivalence
Ratio for Methanol.
35.
Exhaust Temperature vs. Equivalence Ratio for Methanol.
36.
Carbon Monoxide Emissions vs. Equivalence Ratio for
Methanol.
37.
Hydrocarbon Emissions vs. Equivalence Ratio for Methanol.
38.
Nitric Oxide Emissions vs. Equivalence Ratio for Methanol.
-8-
Figure Numbers
39A.
Volumetricc Efficiency vs. Engine RPM.
39B.
Friction
lean Effective Pressure vs. Engine RPM.
40.
Log P vs.LI
Log V for Iso-octane,
= 0.76, 1500 RPM.
41.
Log P vs.LI
Log V for Iso-octane,
= 0.45, 1500 RPM.
42.
Log P vs. Log V for Iso-octane,~ = 0.28, 1500 RPM.
43.
Log P vs. Log V for Iso-octane,~ = 0.83, 2000 RPM.
44.
Log P vs. Log V for Iso-octane,P = 0.48, 2000 RPM
45.
Log P vs. Log V for Iso-octane,c
46.
Log P vs. Log V for Iso-octane,~ = 0.81, 2500 RPM.
47.
Log P vs. Log V for Iso-octane,4 = 0.52, 2500 RPM.
48.
Log P vs. Log V for Iso-octane,k = 0.25, 2500 RPM.
49.
Log P vs. Log V for 100-600,
= 0.77, 1500 RPM.
50.
Log P vs. Log V for 100-600,
= 0.43, 1500 RPM.
51.
Log P vs. Log V for 100-600,
52.
Log P vs. Log V for 100-600, ~ = 0.85, 2000 RPM.
53.
Log P vs. Log V for 100-600, c = 0.47, 2000 RPM
54.
Log P vs. Log V for 100-600,
55.
Log P vs. Log V for 100-600, ~ = 0.82, 2500 RPM.
56.
Log P vs. Log V for 100-600, c = 0.50, 2500 RPM.
57.
Log P vs. Log V for 100-600, c = 0.28, 2500 RPM.
58.
Fuel Mass Fraction Burnt vs. Crankangle for fixed
59.
Fuel Mass Fraction Burnt vs. Crankangle for variable ~b'
60.
PV
4
= 0.22, 2000 RPM.
= 0.27, 1500 RPM.
P = 0.26, 2000 RPM.
b'
vs. Crankangle with Burnt and Unburnt Gamma Values.
-9-
LIST OF TABLES
1.
Engine Specifications
2.
Injection System
3.
Summary of Instrumentation
4.
Summary of Fuel Properties
5.
Engine Operating Conditions for Emission Results
6.
Matrix of Averaged Pressure Crankangle Data
A-1.
Summary of Interactive Data Acquisition and reduction Programs
A-2.
Summary of Pressure Crankangle Data Tile Programs
A-3.
Summary of TCCS Combustion Analysis Programs
A-4.
Summary of General Thermodynamic Properties Subroutines
-10-
I
INTRODUCTION
A stratified charge engine is defined as a spark ignition,
internal combustion engine with a non-uniform fuel air mixture
in the combustion chamber,
Stratified Charge engines have been
recognized for good fuel economy potential, low emissions and
the ability to burn a wide range of fuel types.
In this
thesis, we analyse the performance of an engine based on the
Texaco Controlled Combustion System (TCCS),
Several features specific to this design make the engine
a strong candidate for small, low power applications.
Load
control can be achieved by changing the amount of fuel injected and inlet air throttling is unnecessary in most applications.
As a result, the engine can be run at very lean
overall equivalence ratios giving excellent fuel economy and
low emissions.
Fuel is injected late in the compression stroke
just before the combustion process and is ignited with a spark
discharge.
The residence time at elevated temperatures is
therefore shorter than the time for compression ignition and
hence the engine does not display either octane or cetane requirements.
As illustrated schematically in Figure 1, the Texaco
Controlled Combustion System uses an open combustion chamber with
high air swirl, direct fuel injection and electronic ignition.
Air swirl is generated by
the inlet air flow, and amplified
*Numbers in parenthesis refer to the bibliography at the end of
this paper.
-11-
during the compression stroke by constraining the vortex in
the combustion chamber.
The combustion chamber is essentially
a cup with a cylindrical upper section and a toroidal bottom
formed in the head of the piston.
Fuel is injected with a
Roosa Master Pencil Nozzle with a flat seat and a single hole
orifice as shown in Figure 2.
The positive ignition system
uses a high energy multiple spark unit with controlled duration.(2)
The spark plug electrodes are carefully aligned to promote the
formation of a stable flame front.
High pressure injection of fuel into the swirling air
begins near the end of the compression stroke.
Air swirl in
the combustion chamber promotes mixing and controls the penetration and trajectory of the fuel spray in the cup.
The com-
bustible fuel-air mixture formed by turbulent mixing and air
entrainment in the fuel jet is then ignited by the long duration
spark discharge and burns downstream of the spark plug as shown
in Figure 3.
Following B.C. Jain ( 3 ) we divide the combustion process
into three stages; a rapid combustion phase controlled by the injection rate and a slower "burn up" phase which is controlled by
the rate on air entrainment and mixing of burning products and
air downstream of the spark plug, and a heat transfer dominated phase
which follows after mixing is complete." This sequence is illustrated
on the diagram shown in Figure 4.
During an isentropic compres-
sion or expansion of an ideal gas the value of PVY remains
-12-
constant.
After
a delay covering the jet transit time for the
injector to the flame front, the value of PVY rises rapidly,
This rapid combustion phase appears to be controlled by the injection rate.
After the last fuel injected passes the spark
plug, the rate of change of PV¥ is substantially slower and the
rise is controlled by the rate of mixing of the plume of rich
products with the surrounding air and residual gas.
This
phase ends when all mixing is complete or the exhaust valve
opens.
Any fall in the PVY curve prior to the exhaust valve
opening can be attributed to heat losses.
This research project is a part of a larger program which
includes work on a jet mixing model, a performance model and
photographic studies with a rapid compression machine. ( 4)
The following areas of research are covered in this thesis;
(1)
The completion of the engine test setup and the
development of all necessary
instrumentation to
record the variables of interest.
(2)
TCCS engine performance and emissions are presented for the three test fuels over a wide
range of operating conditions.
Differences in
fuel characteristics are also presented.
(3)
The problems associated with heat release calculations in stratified charge and diesel engines
I7
-13-
reviewed, with the further development of an
existing thermodynamic model to predict heat
release and mass fraction of fuel burnt outlined.
(4)
Log P vs Log V diagrams and plots summarizing the
output of the thermodynamic model are presented
for a matrix of comparable test data,
(5)
The development of computer programs to process
raw data and accomplish online pressure data
acquisition, along with listings of all computer
programs are included as appendixes.
-14-
II
TEST ENGINE AND INSTRUMENTATION
Single Cylinder Test Engine
The engine used in these experiments is located in the
Sloan Automotive Laboratory at MIT, and is arranged as shown
in Figure
5
CFR - 48
crankcase that has been modified to accept a cylinder
.
The single cylinder test engine is based on a
sleeve assembly, head, piston, crankshaft and overhead cam and
valve train assembly for the 3 7/8 inch bore by 3 7/8 inch
stroke LIS - 183 TCCS geometry, as shown in Figure
6
Engine specifications and dimensions are shown in Table 1.
engine is coupled to a dynamatic
The
eddy current dynamometer
equipped with a hydraulic scale, as shown in Figure
7
.
The
basic instrumentation is listed in Table 3 and whenever possible
redundant measurements have been introduced to provide alternate
data sources and to qualify experimental results.
The basic engine support facilities were constructed by
Lazarewicz
5)and follow standard practices as shown in Figures 8
through 11 .
The engine cooling system is arranged to enable
heat rejection measurements.
A
rotameter and throttling valve
on the return line from the engine allows both flow regulation
and measurement while holding system pressure above 5 psig, as
shown in Figure 8 .
Maximum water temperature at the engine cy-
linder outlet was held at 190 ±5 0 F.
The lubrication
-15-
system shown in Figure 9 consists of two separate loops; a low
pressure circulating loop for temperature control and a high
pressure bearing feed and filter loop.
The operating oil temp-
erature and pressure were held at 165°F and 42 psig respectively.
A pressure alarm incorporated in the filter system activates a
siren and cuts off the fuel supply and ignition system when
pressure falls below 20 psig.
Inlet air flow is measured with
an ASME square edge orifice with flange taps and water manometers
as shown in Figure
10.
Following recommendations made by
Lazarewicz, the inlet and exhaust systems were rebuilt and the
inlet settling tank and air heater were closely coupled to the
engine with a short inlet pipe.
Injection and Ignition Systems:
Fuel system specifications appear in Table 3 and fuel flow
is measured gravimetrically as shown in Figure
pump pressure is held at 25 psig.
Fuel
11 .
Transfer
injector leakoff and
the injector pump bleed are returned to the fuel reservoir mounted
on a standard laboratory scale.
Ignition
.
and injection timing, intensity and duration
are monitored with a 565 Techtronix Oscilloscope.
Ignition timing
can be precisely set by a means of a vernier scale and adjustment
arm.
Injection timing is varied through the use of an American
Bosch TMB - 12 Manual Timing mechanism.
A pressure transducer
-16-
is mounted just ahead of the fuel injector in the high pressure
supply line and the Needle Lift Indicator shown in Figure 12
has been developed to replace the standard fuel injector needle
lift and cracking pressure adjustment assembly
.
Both
outputs are displayed on the oscilloscope cathode ray tube
along with the cylinder pressure crankangle markers and BDC
reference pulse as indicated in Figure
13.
As presently de-
signed, resolution of fuel injector timing is limited to 2 CA°
by the time base necessary to include one entire revolution on
the CRT display.
The first peak in the pressure trace cor-
responds to the point of initial needle lift, however; as indicated by the needle lift trace, the injector does not open appreciably for about 2 CA0 .
Determination of injector cutoff and
any secondary injection is only possible with the Needle Lift
Indicator.
Substantialsignal processing problems occur with the
linear displacement amplifier
and a new design is
-
recommended.
Emission Sampling System:
Engine emissions are measured using the Sloan Laboratory
Exhaust Gas Analysis Cart (Table 3).
The exhaust sample is re-
moved from an engine exhaust tank consisting of several cylinder
volumes and pulled through a heated teflon line and filter to the
Gas Analysis Cart.
Stainless steel pipe is used between the
-17-
engine and exhaust surge tank to reduce the potential for
hydrocarbon reactions induced by "rust" and the elevated
temperatures of the exhaust flow.
An additional spun glass
particle filter was installed in the sampling line when it was
discovered that at high loads and equivalence ratios near
= 1.0,
the gas analysis equipment was severely contaminated with carbon.
This technique
appears to have solved this immediate problem.
The development of a computer program
to calculate basic
engine performance and emissions from the recorded data is
described in Appendix I with the complete program listing in
Appendix III.
Pressure Volume Measurement:
Accurate combustion pressure and volume measurements are
absolutely required for mathematical engine simulation, calculation of heat release rates, engine pumping losses and the
compiliaticnof statistics associated with cyclic variations and
peak pressures.
Acceptable pressure volume records can only be
obtained when the entire monitoring system receives careful
attention.
First, high resolution signal recording equipment is re-
quired if the effort expended to obtain accurate pressure and volume
measurements is to be worthwhile.
Oscilloscopes were used in these
experiments for qualitative analysis but they lack the accuracy
required for quantitative resolution.
The real time digital
computer with analog to digital converters can provide the
-18-
resolution necessary; and for these experiments, an online
digital data acquisition was developed for Sloan Laboratory
PDP 11 Computer Facility.
The analog to digital converter
provides resolution to within 0.48 psi over a range of 0-1000 psia,
well in excess of current engine pressure transducer performance
as described in following
were 5 CA .
paragraphs.
The sampling intervals
Appendix II provides details of the system hard-
ware and the development of the software used for online
computer sampling.
Appendix III contains a listing of the
pressure-crankangle data management program and the assembly
language program that actually performs the data acquisition.
In these experiments the cylinder head geometry prevented
the use of a large diaphragm, water cooled pressure transducer
and a Kistler 609A piezoelectric pressure transducer was chosen.
Considerable experimental art is required to obtain acceptable
performance from available pressure measuring equipment, including this specific unit.
For example, quartz piezoelectric
transducers are high impedance devices and contamination of the
electrical connectors can significantly degrade performance.
All
connections must be thoroughly cleaned with a freon base solvent
and sealed with heat shrink tubing.
Transducers are also subject
to the thermal cycling that is fundamental to the combustion
process in engines.
In a chopped flame test by Jain and
-19-
Lazarewicz
(5)
, the 609A transducer showed an apparent 6 psi
response when directly exposed to an acetylene flame at typical
engine frequencies.
This response was reduced with a coated
diaphragm and a thin coating of silicone rubber was applied as
recommended in the literature ( 6
)
The transducer was in-
stalled in a recessed adapter inserted through the water jacket
of the cylinder head.
The adapter
cavity is designed to
minimize attenuation and protect the transducer from engine
temperature fluctuations.
The engine coolant also serves to cool
the transducer.
Careful preparation of the transducer is wasted if the cylinder
volume is not known with similar accuracy.
The cylinder volume
is computed from engine dimensions and recorded crankangle data.
Cylinder clearance volumes were determined by careful measurement of pertinent engine dimensions.
The piston top dead center
was located and the flywheel position pointer adjusted with a
depth micrometer as recommended by Lancaster
(7)
.
A crankshaft
driven rotary pulse generator, supplying 720 pulses plus a marker
every revolution was then aligned with the flywheel with an
accuracy of approximately 1/4 degree. The alignment was tested
by analyzing pressure crankangle diagrams and Log P
- Log V plots
of motoring runs as shown in Figures
.
14, 15
and
16
Lancaster
provides a detailed explanation and interpretation of the plot
orientations(
).
-20-
A clamped disk balanced pressure indicator was also
mounted through the engine water jacket to provide a technique
for dynamic calibration for the cylinder pressure transducer.
The balance pressure indicator is a pressure activated switch
with a reference pressure applied to one side of a thin membrane
disk and the cylinder pressure to the other.
The balanced
pressure indicator is connected to the oscilloscope display as
shown in Figure 17.
When the cylinder pressure rises above the
reference pressure plus the disk contact pressure, the disk is
deflected to ground the center electrode.
The change in potential
is converted by the cathode ray tube grid modulator to momentary
changes in signal intensity on the oscilloscope display as shown
in Figure
13.
These pulses are then used to dynamically
calibrate the signal from the piezoelectric transducer.
-21-
III
BASIC PERFORMANCE AND EMISSIONS
The multifuel capability of the TCCS engine was investigated using methanol, iso-octane and a wide boiling
point (100-600 F) fuel.
summarized in Table 4.
Properties of these test fuels are
Performance and emissions data in
Figures 18 through 39 represents engine operation over the range
of fuel-air ratios and engine loads summarized in Table 5.
The
engine was naturally aspirated and exhausted to ambinet pressure throughout the test series.
All performance data was
measured with injection timing set for maximum brake torque
and ignition system timing set to commence 2 CA0 prior to the
start of injection.
The injection duration exceeded the ig-
nition duration of 20 CA0 except at light load.
The overall
equivalence ratio was used as the absissa in Figures 18 through
38.
a)
The equivalence ratio was determined using two techniques;
the measured fuel and air flow and b)
calculated from the
exhaust gas composition using the method of Stivender
(8)
.
Data
was considered reliable when the difference between these two
values was less than 0.025.
A deviation greater than this was
always traced to operator error, air leaks or faulty equipment.
The engine was not operated at fuel air ratios above stoichiometric.
Previous experience with this engine has shown that any
performance gain above
= 1.0 are achieved with substantially
-22-
increased hydrocarbons, CO and degraded fuel economy( 5 )
Engine Performance with Iso-Octane and Wide Boiling
Point Fuel:
The indicated mean effective pressure (IMEP) versus
equivalence ratio for iso-octane and the 100-600, wide boiling
point fuel are shown in Figures 18 and 19.
The maximum IMEP
is not developed, with either fuel, in the range of equivalence
ratios shown; rather, the effective upper limits for engine
operation are determined by a "smoke limit" near
= 1.0.
In
addition, there is a distinct flattening of the power curve near
stoichiometric conditions.
The secondary dependence on RPM
exhibited by the IMEP curves can be traced to several factors.
The volumetric efficiency increases with speed in the range
tested as shown in Figure 39.
PV
plots also show a slight de-
crease in heat transfer with increasing speed during the heat
transfer dominated phase of combustion.
As will be discusses in
the following chapter, the burning angle appears to decrease
slightly with increasing RPM, and this would also serve to increase the IMEP.
The data for iso-octane and 100-600 split
differently with speed.
(9)
Texaco (9 )
Similar results have been observed by
This difference cannot be satisfactorily explained
with the basic data comparison presented in this thesis and additional study to resolve this potential conflict is recommended.
The indicated specific fuel consumption (ISFC) is shown for
-23-
iso-octane in Figure 20 and for 100-600 in Figure 21.
The
data sets are of similar character with a clear minimum at an
equivalence ratio near
equivalence ratios.
= 0.3 and a steep rise at leaner
This rise in ISFC at lean fuel air ratios
is accompanied by increased cyclic variations and incomplete
combustion; and appears to be a characteristic of the fuel
injection system used with this engine.
The indicated thermal efficiency (i)
provides the best
comparison of the actual combustion process since it properly
accounts
for the different heating values of the three fuels
used in these experiments. The indicated thermal efficiency is
equal to the reciprocal of the ISFC x lower heating value and
the indicated thermal efficiencyis equivalent for engine
operation on both fuels, with a miximum value of approximately
50% reached near c = 0.3 as shown in Figures 22 and 23.
The volumetric efficiency (.)
changes with load to re-
flect changes in the quantity, composition and state of the
residual gas in the combustion chamber as shown in Figures 22
and 23.
The effect of engine speed is shown in Figures 22, 23
and 39.
The effect of load is shown in brackets on Figure 39.
The exhaust temperature data follows the same trends as
described for the IMEP data as shown in Figures 24 and 25.
The friction mean effective pressure (FMEP) versus RPM
-24-
for the single cylinder test engine is shown in Figure 39.
This
data is representative and the actual FMEP showed little variation
throughout the test series.
Emissions with Iso-Octane and Wide Boiling Point Fuel:
The TCCS concept is designed to burn the fuel immediately
after injection in a fuel rich, mixing controlled plume in
order to achieve multifuel capability and low emissions.
In the
course of these experiments, it was consistently observed that
emissions, particularly hydrocarbon and carbon monoxide , are
more sensitive to small variations in engine operating conditions
than the basic performance data.
Careful system timing is re-
quired to obtain satisfactory emissions levels and the convention
adopted places the start of ignition immediately ahead of injection.
However, if timing is adjusted so that fuel injection
preceeds ignition higher IMEP levels are achieved with a corresponding significant increase in hydrocarbon emissions at
equivalence ratios of
> 0.6.
This system sensitivity is thought
to be the source for data scatter at high load conditions and all
lines were drawn using a least square regression analysis
technique.
The increased cyclic variations and degraded emissions
observed at very low load test conditions are thought to be
caused by an injection - ignition system phenomena.
Two
-25-
potential mechanisms for cyclic variations are advanced.
High
speed movies of TCCS combustion in a Rapid Compression Machine
show that the ignition arc
discharge time is short compared
_............
to
the
arc cycle time
(10)
.
Z
_ .B.
..
_
....
. _ ._
__
.........
............................
This ignition characteristic
suggests that the initial fuel jet may pass the spark plug in
front
the interval between ignition pulses and a stable flame
may not be formed in the leading edge of the fuel jet.
This
mechanism introduces the possibility of cycle variation in
the initial stages of combustion.
At light load conditions
the injection duration is less than the ignition duration, and
with small injection quantities, the unburnt fuel vapor that
passes the electrodes before a stable flame front is formed can
rapidly mix to a fuel air ratio below the limit of combustion.
This mechanism may partially account for the degradation of
hydrocarbon emissions at low load.
A second possible mechanism is traced to fuel injector
characteristics.
As indicated in Figure 13, needle lift is not
always crisp and there is often a 2 CA° period at the start of
injection when the needle opens only a small amount.
This
starting transcient may have substantially influenced the
initial stages of fuel jet formation.
As shown by Jain, a low
momentum jet would be swept outside of the electrode radius by
air swirl.
The percentage of fuel vapor missing the plug
electrodes would increase for light load conditions with the
-26-
smaller fuel quantities required.
As in the previous case, this
phenomenon introduces a mechanism for cycle variation and
hydrocarbon and CO formation in the combustion chamber.
Carbon monoxide emissions are shown in Figure 26 for isooctane and in Figure 27 for 100-600.
obtained at an equivalence ratio near
The lowest CO levels are
= 0.45.
The CO levels
observed with iso-octane are approximately 50% lower than observed
with 100-600 at a given equivalence ratio.
The mean minimum
value observed with iso-octane 4 gr/ihp-hr or 3.5 gr/ihp-hr
less than that observed with 100-600.
The hydrocarbon emissions show a sharp increase at equivalence ratios less than
28 and 29.
= 0.3 for both fuels, as seen in Figures
This sharp rise is accompanied by increased cyclic
variations which are attributed to the injection and ignition
effects discussed previously.
dynamic
A small rise in HC emissions is also
observed near stoichiometric equivalence ratios.
Nitric oxide emissions are shown in Figures 30 and 31.
The
trends indicated by the data have the same characteristic shape
as demonstrated in homogeneous charge spark ignition engines;
however, the peak occurs near an overall equivalence ratio of
= 0.6 whereas in a homogeneous charge engine the measured levels
are generally higher and the maximum level occurs near stoichiometric fuel-air ratios.
-27-
Methanol Performance and Emissions:
Methanol was chosen as the third test fuel because it
The
represented a severe test of TCCS multifuel capability.
fuel properties of methanol are significantly different from
iso-octane or the wide boiling point fuel.
Methanol has a
much lower specific heating value and a higher stoichiometric
air-fuel ratio and thus requires injection quantities nearly
A
twice as large to achieve the same equivalence ratio.
Bosch injection system was used in these tests and the pump and
nozzle geometry were selected for the reference fuels.
No
attempt was made to modify the pump or nozzle for the methanol
experiments.
As a result, the fuel system was operated off
design.
The performance and emissions data for methanol are presented in Figures 32 through 38.
Since the fuel system was not
optimized for this fuel, wide scatter in emissions data was
observed.
The basic performance plots in Figures 32 through 35
show the trends to be expected with a properly matched fuel injection system.
Figure 32 showing IMEP data exhibits no
sensitivity to RPM; however, "misfire" increased with higher
RPM
i
h,h
1riA
rOnAt4liniOc
Thc
icnr nA4troA
cr',-f4fc
Fl-F1 aln-r
sumption, volumetric efficiency, and thermal efficiency data, as
shown in Figures 33 and 34 has trends very similar to the
-28-
corresponding trends observed with oso-octane and 100-600 fuel.
The indicated thermal efficency (7i)
is almost identical to
Figures 22 and 23; the maximum value of approximately 50% occurs
at an equivalence ratio near ~ = 0.3.
the Texaco
This demonstrates that
controlled combustion process is compatible with
alcohol fuels; however, fuel system modifications are necessary
to properly match the engine with this fuel type.
Comparison of M.I.T. and Texaco Data
The single cylinder engine emissions and performance data
compare favorably with the TCCS engine data observed at the
Texaco Engine Development Laboratory
(9)
.
The maximum IMEP
observed at MIT with 100-600 fuel exceeded the values observed
at Texaco by nearly 5% at a given equivalence ratio.
The two
engines exhibited identical values of volumetric efficiency at
each given operating point.
Comparative plots by Lazarewicz
showed good agreementwith all emissions data except hydrocarbon
(5)
emissions( 5 )
In these experiments, the level of hydrocarbon
emissions observed was considerably reduced when compared to
values obtained by Lazarewics; however,the level is still higher
than observed at Texaco.
This difference in HC levels can be
explained by the different sampling techniques used at the two
facilities.
The Texaco data was acquired in bag samples before
analysis, while at the Sloan Laboratory a heated teflon line is
used to sample directly from the exhaust tank.
The heated
-29-
sampling line used at MIT eliminates the potential for condensation of hydrocarbons, and in general, higher hydro carbon
levels are measured in experiments with a heated sampling line.
-30-
IV
ANALYSIS OF PRESSURE VOLUME DATA
Accurate records of cylinder pressure-volume data are an
important tool for evaluating performance.
The indicated mean
effective pressure and pumping mean effective pressure can be
computed by integrating the average pressure-volume diagram.
In addition, logarithmetic plots of P, V data provide estimates
of the combustion delay time, the duration of effective heat
release, and the ratio of specific heats, y, during the
isentropic compression and expansion phases.
Finally, pressure-
volume data is required as an input for the thermodynamic models
used to compute heat release and fuel burning rates.
Logarithmetic Pressure Volume Diagrams:
A matrix of comparable test data for engine operation on
iso-octane and 100-600 is presented in Table 6 and logarithmetic
P-V diagrams are shown in Figures 40 through 57.
When pressure-
volume data is plotted on a logarithmetic diagram, the isentropic portions of the compression and expansion process appear
as straight lines.
The slope of the linear segments is equal to
(-1) .y, where y is the ratio of specific heats.
The beginning
and end of combustion are marked by a departure from and return
to the straight isentropic compression and expansion lines;
since the effect of combustion is equivalent to heat being added
-31-
with a consequent change in y.
Table 6.
These points are indicated in
The apparent burning time for the iso-octane and
100-600 fuels is nearly the same and decreases with increasing
RPM.
The apparent average burning time is 5.2 ms at 1500 RPM,
3.5 ms at 2000 RPM and 2.7 ms at 2500 RPM.
Decreases in
engine load only slightly decrease the time required for
burning.
The combustion delay time can be determined if the start
of injection is known.
The injection is tabulated in Table 6
and indicated in each figure.
In previous work, this delay has
been attributed to the required jet transit time of the fuel
from the injector to the stable flame front established at the
(3)
spark plug electrodes(
However, as shown in Table 6 the fuel
used also influences the delay time.
This indicates droplet
evaporation rates may also influence the combustion process.
The small dip in the log P log V diagrams during the injection
period also indicates the effects of fuel vaporization and this
dip is more pronounced when the engine is operated on iso-octane
due to its higher latent heat of vaporization.
Heat Release Calculations:
The rate of fuel burning is a basic parameter in most
engine models and techniques for calculating fuel burning rates
from engine pressure data have received considerable attention.
Our understanding of combustion in homogeneous fuel air mixtures
-32-
is well developed;
however, the increased complexity of
heterogeneous charge engine combustion precludes a strictly
thermodynamic solution which does not account for mixing in the
combustion chamber.
This statement applies to both stratified
charge and diesel engines and most of the previous
work has been done with diesel combustion
systems.
Until
recently, efforts to predict heat release rates in a diesel
engine were highly empirical.
Lyn has developed relationships
to predict heat release rates in an open chamber diesel based
on the fuel injection rates
(11)
.
Borman and Kreigher have
developed a thermodynamic model to predict burning rates from
diesel engine pressure data
(12)
.
In the Borman model, the fuel
was assumed to be homogeneously mixed at each time step.
This
assumption implies very lean equivalence ratios at the start
of injection which increase to the overall equivalence ratio.
This is physically inconsistent since mixing considerations of
the fuel jet imply initially rich combustion followed by progressive mixing down to the final lean overall equivalence ratio.
Still other investigations have assumed micro mixtures of
burning droplets in which all combustion takes place at stoichiometric considerations
(13)
.
The TCCS performance model developed
by Jain assumed that burning took place at equivalence ratios
determined by jet mixing and air entrainment and used a specified
constant equivalence ratio for the burnt products during its
-33-
rapid combustion phase (3 )
These assumptions are critical and
control the shape of the burning rate diagram computed from
pressure-volume data.
The thermodynamic model used in these experiments to predict
the cumulative mass of fuel burnt from pressure-volume data is
an extension of the two zone homogeneous charge combustion model
as outlined below.
The closed system is defined as all air,
residual gas and fuel vapor in the cylinder prior to ingition.
The mass conservation equation can be written as:
v=
4.1
v/M = xvb + (l-x)v u
and the first law as:
e = (Eo-W-Q)/M = xeb + (l-x)eu
where
x = charge mass burnt/total charge mass
= total energy of the charge at time t
E
M = the total charge mass, air + fuel + residual gas
Q = the cumulative heat since t
o
v = the combustion chamber volume
W = the work done since t
0
e = the average specific energy
4.2
-34-
v = the average specific internal energies
e ,e = the appropriate average specific volumes
Vb'V
= the appropriate average specific internal energies
Subscripts:
b refer to the burnt zone
u refer to the unburnt zone
Further more at a given equivalence ratio ~;
v
u
= v
u
u
(P,T
vb
e
u
b=
4.4
b)
b
(P,T)
4.5
eb (P,Tb)
4.6
= e
eb =
4.3
(P.T)
u
u
where
P = the cylinder pressure
T ,Tb = the appropriate average zone temperatures.
ub
Assuming the unburnt zone undergoes adiabatic quasistatic
compression and expansion, T
is calculated from;
U
(ap
s= [ Vu(P1T)-
ah
(u
)T
]
4.7
-35-
Equations 4.1 and 4.2 can be combined to eliminate x, Tb is
then found by iterative technique.
Once Tb is known, x,
the mass fraction burnt can then be calculated from either 4.1
or 4.2.
This model has been expanded by Martin for use in a
heterogeneous charge engine.
The burnt zone is considered as
burnt products + residual gas uniformly mixed at an externally
designated equivalence ratio
b at each time step.
The unburnt
zone is divided into two components; unburnt air + residual gas,
and the injected but unburnt fuel vapor.
Heterogeneous com-
bustion models require a relationship between the fuel fraction
burnt and the charge mass burnt since combustion takes place at
other than the overall equivalence ratios.
The following ratios
can be defined:
y = unburnt fuel mass/charge mass
z = fuel mass burnt/total fuel mass
The equation for specific volume and specific energy of the unburnt zones are then written as:
v
= (uf + (l-x-y))v ua)/(l-x)
4.8
= (Yeuf + (1-x-y) eu )/(l-x)
4.9
-36-
Equations for y and z are then expressed in dimensionless ratios.
FP (-R)
y +Fl
FPb(l-R)
(-R
1F+
1+ RFT + (1-R)F%6
6b (1+FT)
T
(1+RF
x
4.10
4.11
x
+ (I-R)F~b)
where
b
=
the burnt zone equivalence ratio at time t
= the average overall equivalence ratio at time t
F = the stoichiometric fuel-air ratio
R
=
the residual fraction
The computational procedures used to solve for x is the same
as in the homogeneous charge model; and z can then be found with
Eauation 4.11; however, particular attention must be paid to the
change in
and
b with time.
The overall equivalence ratio only
varies during the injection process, and its
change is directly
related to the fuel injected in each time step.
But it will be
shown that proper selection of the burnt zone equivalence ratio
is not straightforward.
As originally developed by Martin, the modified model assumed a constant unburnt equivalence ratio equal the final overall equivalence ratio and that fuel burned immediately upon in(4)
jection (4).
Early efforts to use the model in this form with
held constant predicted a slow start of combustion prior to
b
-37-
the actual rapid combustion phase, and a final burnt fuel
fraction greater than one.
The mddel has been refined by in-
cluding the variations in unburnt equivalence ratio during
injection, and by assuming that the injected but unburnt fuel can be
treated as fuel vapor.
Procedures for allowing the burnt zone
equivalence ratio to change with time have also been included
in the computer program.
Sensitivity of the thermodynamic model to
Figure 58.
b is shown in
These computations assume no mixing, and the burnt
gas equivalence ratio
b is held constant.
The computations are
Ii
I
I.
based
on pressure-volume data and on overall equivalence
ratio of
= 0.45 at 1500 RPM.
Four values of
b are shown, the
overall average equivalence ratio, stoichiometric and two rich
mixtures.
When burning is assumed to take place at the overall
equivalence ratio
= 0.45, the fuel fraction burnt does not
reach 1.0 and the burning time is much larger than predicted
I '_-
by the log P - log V diagram.
This homogeneous premixed case is
clearly one limiting example.
With either rich or stoichiometric
burnt gas equivalence ratio, the rapid combustion phase does not
significantly vary, however, the curves diverge widely and exceed
a value of unity during the mixing controlled combustion.
During this period, the mixing rates and burning rates are
comparable, by definition,and a mixing model is clearly needed
-38-
to account for the change in
b due to air entrainment by the
burning plume if the mass burned is to be correctly related to
the physical processes in the engine.
The fuel-air mixtures in the combustion chamber is heterogeneous, and combustion is not limited, a priori to any single
equivalence ratio at a given time, the burnt and unburned gases
may continually mix throughout the combustion process.
already noted the choice of
As
b for each time step during the
mixing controlled combustion phase is important and an accurate
entrainment model is required.
The air entrainment model pro-
posed by Blizard and Keck and developed for the TCCS engine by
Jain (3 ) is used to calculate the mass burned for the same PV
data used in the sensitivity study as shown in Figure 59.
En-
trainment rates predicted by this model were high and the overall equivalence ratio was reached in 15 CA , introducing an
unrealistic dip in the mass fraction burnt curve.
As indicated
by the two previous examples, the calculation of burning rates
in
a stratified charhe engine requires additional development
to include mixing before plausible results will be obtained.
We postulate that a model developed to predict NO may serve as
a tool to aid in untangling the mixing phenomena.
The mechanisms
of NO formation in homogeneous charge engines are relatively
well understood(l4); and the extension of a homogeneous model to
-39-
stratified charge engines appears to be plausible.
PV¥ Results
An examination of the value of PVY just before, during and
after combustion gives a very good idea of the net heat input
or output to the working fluid due to heat release as a result
of chemical reaction and/or heat loss((3).
of PVY
Figure 60 is a plot
for the same pressure time data as analyzed in Figures
58 and 59.
The cylinder pressure and volume at the start of
injection is used as the reference (PoVo).
The values of the
specific heat ratio before and after combustion were determined
from Figure 41.
The solid line represents heat release at
constant yu while the dashed line represents heat release at yb.
When the end boundary conditions are applied, namely that the
process must start as unburnt and end with its maximum coincident with the burnt maximum, the lines define a very narrow
region within which the true heat release curve must fall.
The
dotted line represents the results of the thermodynamic program
discussed in the last section with the burnt products equivalence
ratio held at stoichiometric.
The rise of this line above the
peak value evident in the burnt curve is explained by the rise
in the value of fuel mass fraction burnt in Figure
58 to greater
than 1.0.
Several conclusions can be drawn from plots of this nature.
-40-
The dip in the unburnt curve during the injection period provides further evidence of fuel vaporization, a fact also discussed in the section covering logarithmetic data plots.
If the
above curves are normalized using the maximum difference in
burnt and unburnt curves, a line starting as unburnt and changing
to the burnt curve near TDC closely approximates the actual
cumulative fuel fraction burnt curve.
PV' plots with y de-
termined from log P, log V plots can be used to qualify results
from a more complete thermodynamic analysis.
normalized PV
Note that when
curves are compared to the results of the
<
Figure 58, only the fuel fraction burnt curves in which 0.95 <
during the rapid combustion phase fall within the defined
boundary region.
The method discussed above can, with careful normalization,
provide a very good estimate of cumulative heat release.
These
estimates can, by comparison with data from detailed thermodynamic, be used to qualify assumptions made in the necessary
mixing models.
b
1.1
-41-
V
1.
CONCLUSIONS AND RECOMMENDATIONS
The multifuel capability of the TCCS concept has been
demonstrated by tests with iso-octane, a wide boiling
point fuel and methanol.
The thermal efficiency of the
engine is independent of the fuel used.
However, proper
matching of the fuel injection system is required to
achieve satisfactory exhaust emissions levels.
2.
The limits on engine operating range are determined
by hydrocarbon and CO emissions.
The equivalence ratio
upper limit is determined by a "smoke limit" near
metric and the lower limit near
stgoichio-
= 0.3 is determined by
cyclic variations and high hydrocarbon emissions.
3.
Detailed emissions data has been acquired for the
three test fuels.
Emissions with iso-octane and the wide
boiling point fuel exhibit similar trends and compares
favorably with previous available data.
4.
Further research is required to explain the effects of
RPM and different fuel types on indicated mean effective
pressure at high load conditions with the TCCS system.
-42-
5.
Techniques for obtaining online digital data with a
small laboratory computer have been demonstrated.
The
pressure-volume data obtained has been shown to be sufficiently accurate for use in performance models.
6.
It has been shown that accurate pressure-volume data
can be used to provide a good estimate of cumulative heat
release through the use of logarithmetic and PV ¥ plots.
7.
The problems involved in predicting heat release rates
have been discussed and the importance of mixing in diesel
and stratified charge combustion clearly demonstrated.
A
detailed thermodynamic analysis of burning rates will
require better modelling, both for the mixing of the fuel jet
with air before it is entrained in the flame front and for
the entrainment of air by the burning plume.
8.
It is recommended that a NO
developed for this engine.
prediction model be
This model will aid in under-
standing the role of mixing in stratified charge combustion
and can be used to qualify mixing models developed for heat
release calculations.
-43-
9.
Parametric studies involving off design operation are
required to explain the sensitivity of hydrocarbon emissions
to small variations in injection and ignition phasing.
-44-
REFERENCES
1.
Alperstein, M. et al., "Texaco's Stratified Charge Engine Multifuel, Efficient, Clean and Practical,"
2.
Canup, R. ED, "The Texaco Ignition System - A New Concept
for Automotive Engines,"
3.
SAE Paper 750347.
Jain, B.C., "Performance Modelling of the Texaco Controlled
Combustion Stratified Charge Engine,"
4.
Lazarewicz, M.L.,
Engine on Gasoline,"
6.
S.M. Thesis, MIT, 1975.
"Performance of Texaco Stratified Charge
S.M. Thesis, MIT, 1975.
Brown, W.L., "Methods for Evaluating Requirements and Errors
in Cylinder Pressure Measurement,"
7.
PhD. Thesis, MIT, 1975.
Martin, M.K., "Photographic Study of Stratified Combustion
Using a Rapid Compression Machine,"
5.
Lancaster,
SAE Paper 67008.
D.R., Krieger, R.B., and Liensch, J.H., "Measure-
ment and Analysis of Engine Pressure Data','
8.
SAE Paper 70026.
Stevender, D.L., "Development of A Fuel Based Mass Emission
Measurement Procedure,"
9.
SAE paper 740563.
SAE Paper 710604.
Personnel Communication M. Alperstein, Ltrs. dated 12 and 24
February, 1976, providing experimental data sheets.
10.
Wong, V., "A Photographic and Performance Study of Stratified
Combustion Using a Rapid Compression Machine,"
May, 1976
S.M. Thesis,MIT
-45
11.
Lyn, W.T., "Study of Burning Rate and Nature of Combustion
in Diesel Engines,"
IX Internation Symposium on Combustion,
1962.
12.
Kreiger, R.B. and Borman, G.L., "The Computation of Apparent
Heat Relsease for Internal Combustion Engines,"
ASME Paper
66-WA/DGP-4.
13.
Chiu, W.S. et al., "A Transient Spray Mixing Model for
Diesel Combustion,"
14.
SAE Paper 760128.
Komiyama, K., "Predicting NO
Emissions and Effects of
Exhaust Gas Recirculation in Spark-Ignition Engines,"
SAE Paper 730475.
15.
Leary, W.A, "Metering of Gases by Means of the ASME Square
Edged Orifice with Flange Tops,"
MIT Sloan Laboratory Notes,
1951.
16.
Leary, W.A., "Hydraulic Scale Data,"
MIT Sloan Laboratory
Notes, 1946.
17.
Spindt, R.S.,"Air-Fuel Ratios from Exhaust Gas Analysis,"
SAE Paper 650507.
18.
"LP511 Laboratory Peripheral System User's Guide,"
Digital Equipment Corp. - HLPGA-C-D.
-46-
APPENDIX I
PERFORMANCE AND EMISSIONS DATA REDUCTION PROGRAM
An interactive data reduction program was written for the
Sloan Laboratory PDP 11/10 data analysis facility.
The program
consists primarily of I/O and is designed specifically for the
TCCS engine; however, modifications for other single cylinder
engines are possible.
All inputs are requested in the same units
that are used on the experimental data sheet.
Equations used to
compute air flow rate in grams/second, coolant flow rate, volumetric efficiency, brake horsepower, and engine emissions require
clarification,
Air Flow:
The equation for the flow rate of air through the ASME square
edged orifice meter with flange taps is given by the following
equation (5).
W = 51.94 D2 KY
2
T1
Gy AP
w = mass flow rate grams/second
P2 = orifice diameter inches
K = flow coefficient
Y = expansion factor
P1 = static pressure before orifice in in. Hg
T1 = temperature before orifice O R
-47-
G = specific gravity of gas
y = super compressibility factor
p = pressure drop across orifice in in. H20
The computer form of this equation will work with two orifice
diameters D2 = 0.515 and D2 = 0.71.
The program assumes the
following variable values:
Y = 1 - 7.3 x 10
-G
-G
wet = Gdry
-4
Ap
l+w
1+1.608w
Y= 1
The maximum and minimum Reynolds number can be written as
functions of Y and D2 with N
= 0.85, RPM Max = 3000 RPM
Min = 1000, and inlet air temperature = 900F
Re
min
=
17174.72
D2
ReMa x =-
Max
68701.3
D2
and appropriate values of K as a function of D 2 determined
0.0366
K = 0.6152 D 2
Coolant Flow Rates:
The equation for coolant flow rates was determined by linear
regression analysis of calibration data points.
46 F water was
used and a density correction was applied to obtain the least
squared curve fit given below, good at 175°F.
-48-
m = 0.0479 (h) - 0.0081
r = 0.994
m = lbm/sec
h = rotameter height
r = goodness of fit
Volumetric Efficiency:
The engine volumetric efficiency is strictly a function of
engine dimensions and operating conditions as shown below:
2
m(
V
)
PV
PiV
( TIR )
m
= air flow rate in lbm/min
N = revolutions/min
PI = inlet air pressure
V = cylinder volume
R = specific gas constant
TI = inlet air temperature
after the inclusion of engine geometry and unit conversions
3.TDTia (TI+460)
N[.0193
Patm
0.361
I]
-49-
Engine Power Output:
The mean effective pressure and horsepower are determined
from the following equations (16).
hp = mep LAN
66000
hp =
NAh
K
After including specific engine geometry and an overall
dynamometer constant of K = 6000
mep = 2.88847Ah
NAh
6000
hp = NAh
where
A = area of engine piston, in.
L = stroke, ft.
N = engine RPM
mep = mean effective pressure, psi-.
Ah = dynamometer scale height, in Hg.
Specific Emissions:
The average exhaust composition is a function of the
equivalence ratio.
A model developed by Stivender (8)
is used
to determine the exhaust based equivalence ratio as a check of
the equivalence ratio measured from inlet flow rates; and to
compute the engine emissions in grams of pollutant/indicated
-50-
horsepower hour.
Required inputs are the indicated specific
fuel consumption (ISFC), emissions concentrations on a volume
basis and the fuel carbon:hydrogen ratio.
The model as pre-
sented does not apply to alcohol fuels and a method presented by
Spindt
7)
was used for the methanol experiments.
The model fits one undetermined equilibrium constant for the
water-gas reduction to direct measurements:
[H2 0]
K
[o
2
[CO2
]
[CO]
[H 2
[H2
= 3.8
]
The combustion reaction for a typical hydrocarbon fuel with air
may be expressed in the following form:
CH
+ (n) 02 + (3.76n)N2 -
(a) C
2
+ (1-a-6c) CO +
+(b) H2 0 + (c) C6H 14 + (d) NO -+ (y/2-b-Tc)H 2 +
(n-a-(l-a-6c)/2 - b/2 -d/2) 02 + (3.76n-d/2)N 2
The molecular weight of fuel as it appears in the above
equation can be written as
Mf = 12.01 + 1.008 y
The molecular weight of air is assumed as
M
a
= 28,96
-51-
The air fuel ratio can then be written by a carbon and
oxygen balance as
[CO]
Me
[C02] + [02]
A= 4.76
E
F
M
[
a
[H20] + [NO]
2
[HC] + [CO] + [CO 2 ]
The HC concentration is measured wet.
All other pollutant
concentrations are dry and must be corrected by the following
relationship
[ ]wet =
[
]dry (1-[H201)
An empirical correlation used to determine the exhaust water
concentration where the concentrations were wet and K = 3.8, as
shown below
5 y( [CO2
[H 2 0] = .5y([C02]
- [Co]
- CO]
3.8[C02]
Specific pollutant emissions can be written as
M
x
IS "X" (gr/ihp-hr) = M
[
where "X" is the species of interest.
[X] wet
[HC] + [CO] + [CO 2 ]
ISF
When the above equation is
used to indicate C6H14 emissions the ["X"] term is [HC]/6 since
[HC] is determined by a count of single carbons.
-52-
For output consistency fuel consumption is based on the
observed air flow and the equivalence ratio calculated from the
exhaust products.
Appendix III.
The computer program listing is included in
-53-
APPENDIX II
ONLINE DIGITAL PRESSURE DATA ACQUISITION
Accurate pressure volume data was required and an online
digital acquisition system was developed for these experiments.
The advantages of direct data acquisition include improved accuracy
and speed.
With these routines, a large number of data records can
be collected and statistically analyzed.
Consequently, failure
of the experimental techniques can be detected by immediate review
and preliminary analysis of the digital data.
The online data
acquisition system and program described in this appendix is based
on a Digital Equipment Company (DEC) 11/10 computer with 16 K of
core memory, a DEC RK -05 random access disk, two teletype
terminals and the DEC Laboratory Peripheral System (LPS).
The
DEC LPS consists of an 8 channel multiplexed analog to digital
converter with variable gain preamplifiers, and internal timing
clock, and two Schmidt triggers.
This system has the capability
to sample each channel along with its multiplexed pair simultaneously and then perform sequential converstion on the two
signals.
The sample window length is 5
signal conversion takes 25 pS.
nano-seconds and each
Maximum sampling rate on a single
channel is 45 Hz, with 12 bit conversion.
A schematic of the data acquisition system is shown in
Figure 17.
The reference marker pulse at the start of the compression
-54-
stroke is input to channel 0 with the cylinder pressure input to
channel 10.
Marker pulses every 5 CA ° are used as an input to the
Schmidt trigger.
It appears that further system refinements will
permit 1 CA0 sampling increments for a single pair of inputs.
Accurate pressure records must be matched with accurate
crankangle records to be of further use and variations in crankshaft angular velocity precludes accurate determination of the
cylinder volume when the pressure signal is only logged against
time.
Consequently, combustion pressure and crankangle position
signals make up a data pair and are sychronized with the aid of
the Schmidt trigger.
Each sample interval consists of 144 data
pairs comprising two engine revolutions.
The actual crankangle
position is not known at the start of a sampling interval
and
the sampled data is reordered at the end of each data sample interval using a reference signal at 185
before top dead center.
Two techniques can be proposed for obtaining average engine
performance.
The first involves a complete heat release
analysis of individual combustion records followed statistical
averaging of these results.
The extended interval between
sampled data sets and the computational expense required by this
method precludes its use.
Consequently, a second technique in-
volving the computation of mean cylinder pressure records from
consecutive data sets was used.
These mean records are then
analyzed to obtain engine performance.
If a Gaussian distribution
-55-
is assumed, the probability density function can be expressed as:
p(X)
=
cy
-(X -X )
exp [
2
2a
The following equations are used to calculate the statistical
properties of data variations from N records of data:
N
Mean
( )
=
1N
1
E
X.j
1j
j =i
X. represents the average value of the i
the element of the data
vector X.
N
Variance:
(2)
S.2
i
1
(N-)
(X. -X .)
1J
1
1
j =1
Rearranging these equations, the standard deviation can be
computed with a single pass of the data:
N
1i
IN-l =
1
j
-
)22
ij
N
j = i
This routine permits calculation without an unmanageable number
of element arrays.
However, the standard deviation as calculated
is not normalized and can only be expressed as a percent of the
-56-
observed element mean.
Assuming that we are sampling from a normal population
it is possible to construct exact confidence intervals for p,
the true mean, even when
distribution.
i
is unknown, by use of the Student - t
A 1-a confidence interval for pi is expressed as:
1
1/2
S.
- t )(_!
<<
S.
i
+
( -(
i X + ta/2 (
1-
))
For a large sample size the distribution of Si can be closely
approximated as normal and a 1- a confidence interval is:
S.
S.
<
. <
/2
1+
1
1
Za/ 2
where
S. Z.
.
1
(C/F)
In addition to the statistical information outlined above,
the outline data acquisition program integrates the mean pressure
volume diagram to calculate the indicated mean effective pressure
and the pumping mean effective pressure by using Simpson's Rule
-57-
for non evenly spaced ordinates as shown below:
N
W =
[(Vj+ 2 (0)
-()
(5P(0) + 8Pj+(0) -Pj+2()]
j = 1,3,5..
Note that cylinder volume is a geometric function of crankangle
position.
Log P and Log V vectors are also displayed for cycle evaluation and the program listing containing in stream documentation
is
included in Appendix
III.
for the sampling subprogram
The assembly language commands
are explained in Referencec8).
-58-
APPENDIX III
COMPUTER ANALYSIS PROGRAMS
Listings for all computer programs and subroutines used
in these experiments are included in this appendix.
The programs
and subroutines can be grouped in four areas as indicated by
Tables A-1 through A-4.
All programs contain in-stream
documentation of major equations and computational schemes
and each subroutine contains a brief section describing its
purpose, calling sequence, and the definition and dimensions of
arguments.
All programs with the exception of subroutine
SAMPLE are written in ANS Fortran IV.
written in DEC Assembly Language.
Subroutine SAMPLE is
-59-
Table A-1
Summary of Interactive Data Acquisition and Reduction Programs
Program
Purpose
REZLTS
Calculates basic performance
and emissionsfrom experimental data
ONLINE
Performs online pressure data acquisition and calculates mean pressure
crankangle statistics
Subroutine
SAMPLE
Provides assembly
language commands
used to control analog to digital
conversion
-60-
Table A-2
Summary of Pressure Crankangle Data
File Preparation and Control Programs
Program
ANALIZ
Purpose
Prepares pressure crankangle data
files and basic performance information for additional thermodynamic analysis.
Subroutine
RECALL
Returns pressure-crankangle data
from ONLINE for further calculations.
XPRNT1,
Provides printout of internal
variables in XCLC2 for each time
increment.
XPRNT2
STORE
Provides file storage options
for output from ANALIZ.
Table A-3
Summary of TCCS Combustion Analysis Program
Purpose
Subroutine
XCLC2
Calculates fuel fraction burnt, and
PVY versus crankangle from pressurecrankangle data
GASVEL
Calculates the average gas velocities
at the periphery of the piston cup*
HEAT2
Calculates heat transfer rate in
TCCS engine during combustion
PLUME
Calculates air entrainment rates
for burning gas plume**
* Appendix B, Reference ( 4).
**Appendix II, Reference ( 3).
-62-
Table A-4
Summary of General Thermodynamic Property Subroutines*
Subroutine
Purpose
AFTEMP
Calculates adiabatic flame
temperature from given initial
state
CLDPRD
Calculates burnt gas properties
at low (<1000OK) temperatures
DERIVS
Calculates derivatives of
properties; for HPROD
HPROD
Calculates burnt gas properties
at temperatures >11000K
TEMP
Calculates T(h,P) for burnt gas
TSUBU2
Calcualtes T(P) from given initial
state of unburnt gas following an
isentropic process(assumes no
fuel vapor)
UPROPZ
Calculates properties of unburnt gas
(assumes no fuel vapor)
*Appendixes C and D; Reference ( 4)
-63-
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-132-
Table 1
Engine Specifications
DIMENSIONS*
bore
3.875 in.
stroke
3.875 in.
connecting rod
6.625 in.
clearance volume
4.570 in.
VALVE TIMINGS
CLOSES
OPENS
inlet valve
10 BTDC (1)
0
exhaust valve
(2)
55 BBDC (1)
45
(2)
(1) at 0.006 in, valve lift
defined in Figure ( )
55 ABDC (2)
45
(1)
10 ATDC (2)
o
(1)
(2) valve face flush with head
and Appendix B of (
)
-133-
Table 2
Injection System
PUMP
pump:
cam:
plunger:
reaction valve:
APE - B Bosch
6/1
7mm.
20 mm
NOZZLE
nozzle: Roosa-Master XNM 1029
orifice
diameter
orifice
L/D
nozzle cracking pressure
needle
lift
.023 in.
1.0
2000
.010 in.
-134-
Table 3
Summary of Instrumentation
Temperatures
Air Orifice Inlet
Water Outlet
Bearing Oil
Air Inlet
Exhaust
Fuel Inlet
Water Inlet
Crankcase Oil
Fuel Returns
Instrument - All Points
Chromel - Alumel Thermocouple
Omega DS - 500
Digital Readout
Resolution 1F
Pressures
Method
Resolution
Inlet Air
Water Manometer
.1 in.
Crankcase Vacuum
Water Manometer
.1 in.
Oil Pressure
Panel Gage
Exhaust
Mercury Manometer
.1 in.
Dynamometer Load
Mercury Manometer
.1 in.
Injection Line
Kistler 601 Piezoelectric
transducer, Kistler 504E
2 PSI
30 PSI*
Charge Amplifier
Combustion Chamber
Kistler 609A Piezoelectric
Transducer, Kistler 503D
Charge Amplifier
.2 PSI**
-135-
Table 3 (cont)
Flow Rates
Method
Air Inlet
Resolution
ASME Square Edged Orifice
with water manometers
Fuel
Laboratory Scale and Timer
Cooling Water
Rotameter
0.05 g/sec
0.01 g/sec
.2 lbm/sec
Position
Crankangle
Trump Ross Rotary Pulse Generator
720 pulses per revolution
.2 CA0
plus Marker
Fuel Injector
Needle Lift
AVL NH1 - 100 B LDT
1
m
Gas Analysis Cart
Exhaust
Hydrocarbons
Scott Model 215 FID HC Analyzer
Nitric Oxides
TELCO Model 10 A Chemilumenscent NO Analyzer
Carbon Dioxide
Beckman Model 315A NDIR CO2 Analyzer
Carbon Monoxide
Beckman Model 315A NDIR CO Analyzer
Oxygen
Scott Model 150 Paramagnetic 02 Analyzer
* Accuracy is effectively limited by the system transfer
function and the 565 oscilloscope.
** See Text
-136-
Table 4
Summary of Fuel Properties
Methanol
CH3OH
Molecular Wt.
32
H:C Ratio
4 0
Specific Gravity
Boiling Point
Iso-Octane
C8H18
Lower Heating Value (Btu/lbm)
2,25
211
8580
19080
0.155
1,828
.692
149
Stoichiometric F/A Ratio
Distillate
4125
114
.796
F
Cross-cut
0.0665
.80
(106-648)
18038
0.0692
FIA - %
29.5
Aromatics
3.0
Olefins
67.5
Saturates
Octane No. RON
Cetane No.
106
100
76.6
28,3
-137-
Table 5
Engine Operating Condiltions for Emission Results
-
0
RPM
1500
2000
2500
0
IMEP
ISFC
S
s
0.759
-24
104.7
0.384
0.625
-23
93.2
0,358
0.446
-23
81.9
0.290
0.278
-23
56.8
0.284
0.261
-23
53.6
0.263
0.172
-23
39.1
0.249
0.143
-20
24.7
0,331
0.113
-23
17.5
0,379
0.828
-26
122.2
0.353
0.650
-26
108.9
0.320
0.475
-26
91.3
0,289
0.348
-24
73.1
0.269
0.218
-23
44.2
0.285
0.161
-24
25.9
0.362
0.811
-29
130.3
0.349
0.639
-28
119.3
0,349
0.524
-28
105.4
0,286
0.371
-28
82.0
0.262
0.249
-28
55.5
0.265
0.193
-28
38.1
0.306
0.124
-28
21.7
0,344
= Start of Injection
Ignition Always Preceeded by
Ignition Always Preceeded by 2 CA°
-138-
Table 5 (con't)
Cross Cut Distillate
RPM
1500
2000
2500
0
IMEP
ISFC
0.924
-22
112.1
0.432
0.762
-20
105.4
0.392
0.600
-20
93.0
0.358
0.442
-19
79.7
0.299
0.426
-20
76.3
0.317
0.282
-19
55,5
0.284
0.211
-19
41.3
0,290
0.129
-18
22.8
0.327
0.927
-25
112.9
0.416
0.859
-23
112.4
0.390
0.692
-23
98.5
0.372
0.472
-23
88.4
0.296
0.334
-23
63.6
0.303
0.258
-23
51.4
0.290
0.180
-23
34.9
0.304
0.184
-20
32.4
0,338
1.000
-26
122.5
0.439
.850
-26
120.7
0.391
.685
-25
108.9
0.353
.546
-26
95.6
0.324
.367
-25
69.3
0.303
.291
-26
56.6
0.300
.212
-23
39.7
0.314
.218
-20
19.4
0.387
-139-
Table 5 (cont)
Methanol
RPM
0
s
1500
2000
2500
IMEP
ISFC
0.767
-25
106.9
0.878
0.636
-28
101.7
0.755
0.486
-26
89.0
0.676
0.394
-26
77.7
0.637
0.324
-26
66.4
0.616
0.261
-24
50.8
0.655
0.219
-24
40.7
0,697
0.811
-28
108.9
0.900
0.741
-28
106.9
0.838
0.654
-28
104.6
0.766
0.591
-28
100.5
0.731
0.504
-28
92.7
0.682
0.414
-28
84.6
0.632
0.364
-25
72,5
0.643
0.291
-25
58.6
0.640
0.221
-25
26.9
1.077
0.696
-35
106.3
0.824
0.641
-35
102.3
0.795
-140-
Table 6
Matrix of Averaged Pressure Crankangle
Iso-Octane
RPM
1500
2000
2500
0.
is
0.
ie
9bs
0
106.6
-24
9
-12
35
81.9
82.3
-23
0
-13
35
0.28
56.8
60.0
-23
-3
-12
24
0.83
0.83
122.2
118.3
-25
13
-12
28
0.49
0.48
91.3
88.5
-24
3
-9
24
0.22
0.22
44.2
44.3
-23
-5
-7
23
0.78
0.81
130.3
122.4
-29
15
-8
30
0.52
0.52
105.4
99.6
-28
7
-9
30
0.25
0.25
55.4
47.4
-28
-4
-9
29
IMEP
be
~e
0.77
0.76
104.7
0.46
0.45
0.29
IMEP.
m
1
Subscript Nomenclature
o = Observed from fuel air flow
e = Measured from exhaust gas composition
M = Dynamometer measurement
i = Integrated average pressure data
is = Injection start
ie = Injection end
bs = Start of heat release from Log P Log V plots
be = End of effective heat release from Log P Log plots
be
-141-
Table 6 (cont)
Cross Cut Distillate
RPM
1500
2000
2500
RPo0
e
IMEP
IEM
IMEP.
I1i
0.
is
0.
ie
0
bs
be
0.76
0.78
105.4
99.2
-20
10
-11
35
0.43
0.43
76.3
73.3
-20
4
-13
36
0.27
0.27
51.4
47.2
-20
-5
-11
35
0.87
0.85
121.0
114.6
-24
15
-14
29
0.49
0.47
88.4
84.9
-23
-2
-14
26
0.26
0.26
51.4
46.5
-23
-12
-15
31
0.83
0.82
121.3
124.8
-26
15
-14
31
0.52
0.50
95.0
100.1
-26
0
-14
24
0.29
0.28
56.1
59.7
-26
-9
-13
26
-142-
FI-. 1 -TEXACO CONTROLLED COMBUSTION
SYSTEM (SCHEMATIC)
__II
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-143-
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-144-
SPARK PLUG
\
AXIS
PLU ME
\L STAGE
ILETE
JS
AIR
SWIRL
AXIS
SCHEMATIC
OF THE COMBUSTION
FIG 3
___
·___
__1__1
PROCESS
-145-
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FIGURE 5
PHOTOGRAPH OF TEST ISTALLATION
_lbl
-147-
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(D
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m
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-148-
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atz
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art
apart
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-150-
ENGINE
RESERVOIR
ANGER
FIG 8
COOLING SYSTEM
-151-
FNCGINF
TO HEAD
STEAM
WATER
-AT EXCHANGER
TO CRANKCASE
IN
VALVE
PUMP
FIG, 9
LUBRICATION SYSTEM
--'''=1e7'. fK2975$9P}m.NAft.f
-152-
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-155-
1.
IGNITION DURATION
2.
FUEL INJECTOR NEEDLE LIFT
3.
4.
5.
6.
7.
FUEL INJECTION LINE PRESSURE
CYLINDER PRESSURE
CRANKANGLE INDICATOR, 1/5 CA0
185 CA0 BTDC REFERENCE SIGNAL
BALANCE PRESSURE INDICATOR
FIG.13-TYPICAL OSC I LOSCOPE TRACE
-136-
I
A
I.".
I
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I
I
I
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MOTORING 1500
I I
I
RPM
1.2
1.0 _
-
8
w
cr
D
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cl:
-
(D
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0 _
III
-.2
-
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-.9
I
1
I
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-.7
-. 5
-.
3
LOG (VOLUME / VOLUME MAX)
FIG,14 1500 RPM MOTORI NG LOG P Vsv LOG V
- --
---~~----~--
I
-.1
I
.1
-157-
II.'4
I
I
I
I
I
I
I I
I
i
I
I
I
I
I
I
I
1.2
1.0
.8
n,."
-W.
LLJ
)
I::L
CL
oC.4
0
_1
.2
0
-. 2
-i1.i
-. 9
I
I
-. 7
-. 5
-. 3
LOG (VOLUME/VOLUME MAX)
FIG. 15 2000 RPM1 MOTORING LOG P vs LOG
-----
.-
.-·9lllls
-ICI1--
-·-----
-
--
I
- I
I
I
.I
-158-
1.4
I
I
I
.....
_
_~~--
_
I
I
~~
I
I
...
I
I
I
I
MOTORING 2500 RPM
I.2
1.0
H
8
-w
cL
:D
6
(n
(n
~o
w
cc
o_
0
-J
4 _
2 _
0
_
-. 2
_
-
.I
I
I
-. 9
I
I
I
.
I
-. 5
-. 3
-. 7
LOG(VOLUME/VOLUME MAX)
I
-.1
FIG 16 2500 RPM 1MOTORING LOG P vs LOG V
_______________II__III*Cl_____lsCIC*__
I
.1
-159-
CD,
Z
On
C)
o
c:
a
)_4
0
CD
H
U-
(,_)
H
LU
r~
0
Pl--
I
-J
-I-,
U
0
I:_
CD
I
LJ
U
._J
<
Cl:
--------
-160-
I
I
I
I
I
I
I
I
I
ISO-OCTANE
140
130
120
Un
a_
I
_
II0
W
100
n
Q_
U
U
_
I&
90 _
-
80 _
W
I_
LL
u
uL
70 _
zUZ
60 _
C
U
p-
50
_
40
_
Zi
z
I
30 _
I
20
I0 I
I
0
0
I
.2
FIG. 18
I
I
I
.4
EQUIVALENCE
I
I
.6
RATIO
-------
lilPoBI---
I
INDICATED MEAN EFFECTIVE PRESSURE VS
EQUIVALENCE RATIO
--------
I
.8
-- -----
FOR
F
ISO-OCTANE
1.
-161-
I
I
-- _~
140
130
I 20
I
I
I
I
I
I
I
100-600 D
~
m
m
n
a-
110 --
D
100
U)
U)l
a.
LU
90
LLI
80
0
LU.
II
>
IL
u(D
z
,
r'0
2t
70
w·
60
_
50
_
Z
40
30 _
20 _
A
1500 RPM
*
*
2000 RPM
2500 RPM
10 _
0
I
0
I
I
I
I
.4
.2
INDICATED
I
.6
EQUIVALENCE
FIG.19
I
I
I
.8
RATIO
!EAN EFFECTIVE PRESSURE VS
EQUIVALENCE RATIO, FOR 100-600 FUEL
1.
-162-
.5
I
I
I
I
I
I
ISO-OCTANE
sty
r
I
Il
4
A
z
0
T-o
J
z0..
Li
A
I
m
3
--
:D
LL
A
A
u
I
.
a:
0
Li
Q_
w
u
C--
0
2
2 _
A 1500 RPM
* 2000 RPM
* 2500 RPM
lI
0
0
I
.2
I
1
I
.4
.6
EQU IVALE NCE RATIO
I
I
I
.8
FIG. 20 INDICATED SPECIFIC FUEL CONSUMPTION VS
EQUIVALENCE RATIOj FOR ISO-OCTANE
1.
-163-
.5
I i
I
I
I
I
I
I
I
100- 600 D
cr
A
I
I
',
.4
_-
A/w
w
.
-
2
A
I
0
A
a
u
3
0
_
0
A
LL
E:
0
0
2 _-
z
1500 RPM
* 2000 RPM
* 2500 RPM
A
0
_
L
I
C)
I
.2
FIG. 21
I
I
I
I
I
.6
.4
EQUIVALENCE- RATIO
I
I
.8
INDICATED SPECIFIC FUEL CONSUMPTION VS
EQUIVALENCE RATIO, FOR 100-600 FUEL
1.
-164-
100
I
!r~L-.11,,~
-0
> 80
Z
z
I I i
~
I
I
ISO-OCTANE
250 0
-,-.-
15oo
-
U
LL
UL
60
C)
0:::
:D
-0
40
Cd
>v
* 1500 RPM
* 2000 RPM
I 2500 RPM
r20
I
H
I
0
0
I
.2
I
I
I
.4
EQUIVALENCE
I
.6
RATIO
I
I
.8
-
I
1.
FIG. 22 INDICATED THERMAL AND VOLEMETRIC EFFICIENCY
VS EQUIVALENCE RATIO FOR ISO-OCTANE
ljm
I
-165-
i.A
A
A
r] .
.
IUU
-
.
_
_
_
I
I
_
.
-a_
_
_
.
I
._
I
100-600 D
.
I
_
_
.
.
I
I
0-2
> 80
0
z
LL
--
Li 0
0
A
1500-
LiJ6
0
u
LI
:D
I-'J0 4.0
A
A
~2O
o
2
A 1500 RPM
· 2000
~* RPM * 2500 RPM
-~~-
Li,,
0
I
0
I
.2
I
I
I
.4
EQUIVALENCE
I
I
.6
RATIO
I
.8
I
1.
FIG. 23 INDICATED THERMAL AND VOLUMETRIC EFFICIENCY
VS EQUIVALENCE RATIO FOR 100-600 FUEL
~s
l
r~
)--_
I--
-166-
_
I
__
I
I
I
I
__
I
I
ISO-OCTANE
14
12 _
IL
0
10 _-
U
a:
0~
bli
8 _
:D
x
ii
6 _
4 _
IPM
!PM
RPM
2 _
I
0
0
I
.2
I
I
.4
I
EQUIVALENCE
I
.6
I
I
.8
I
RATIO
FIG. 24 EXHAUST TEMPERATURE VS EQUIVALENCE RATIO
FOR ISO-OCTANE
1.
-167-
_
I
__
I
_·
I
14
I
·
I
I
I
·
I
I
100- 600 D
12
(lJ
0
x
10
LL
0
uI
r
8
nD
6
0:
w
u3
x
X
4
A 1500 RPM
2 _
I
0
0
.I
.2
I
I I
.4
EQUIVALENCE
*
2000 RPM
a
2500 RPM
I
.6
RATIO
I I
I
.8
FIG 25 EXHAUST TEMPERATURE VS EQUIVALENCE RATIO FOR
100-600
FUEL
1.
-168-
I
I
I
I
35
I
I
I
I
ISO-OCTANE
30
cr 25 _
Ir
A 1500 RPM
0r
* 2000 RPM
* 2500 RPM
(9
20
0
(I)
15
w
0
0
10
a
A \
A
.
A
0
U
5 _
.
.
I
0
0
I
.2
'I
I
I
.4
EQUIVALENCE
I
.6
I
.8
I
1.
RATIO
FIG 26 CARBON MONOXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR ISO-OCTANE
-169-
I
I
I
r
i
r
I- I
i
100- 600 D
35 _
30 _
'r
I
25 _
0-
"r"
Ir
n
(9
20 _
zr
0
A
i
0V)
U
15 _
0
A
0
u,
*/
10 _
.
,
I
-
A
5 _
1500 RPM
* 2000 RPM
a 2500 RPM
A
II
0
A
0
I/
.2
II
I·
II
.4
EQUIVALENCE
FIG, 27
II_
.6
I·
II
.8
I(
RATIO
CARBON M1ONOXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR 100-600 FUEL
1.
-170-
18
16
14
xI
12
'
'r10
I
(D
n
o0
z
0o
8
W
or)
L
6
()
Ir"
4
2
0
0
.2
.4
EQUIVALENCE
.6
.8
RATIO
FIG. 28 HYDROCARBON EMISSIONS VS EQUIVALENCE
RATIO FOR ISO-OCTANE
__
111
1.
-171-
I
I
I
I
I
I
I
I
1.8
100-600 D
1.6
I .4
x
I
1.2
a9
1.0
(I)
(n
8
A
(I)
U
.
6
uI
A 1500 RPM
.
* 2000 RPM
a 2500 RPM
4
A
A
0
2
A
I
A
I
A
I
0
0
I
.2
1
1
I
.4
EQUIVALENCE
I
.6
RATIO
I 1
.8
1
.9
FIG,29 HYDROCARBON EMISSIONS VS EQUIVALENCE
RATIO FOR I00-600 FUEL
__r__ls____F__________sl__·__
____I__
I.
-172-
-
7
-
I-
Irt
-1
v
I
-I
I
1~~
ISO-OCTANE
-·
6.
Ii
-r
I
5.
U
o
0o
11
N
0
2
.
4.
0
A
(o)
z
A
A
0
or)
uj
3. _
a
A
0
A
z
1500 RPM
2000 RPM
2500 RPM
A
2. _
*
*
.
0
1._
a
I
0
0
_ I
.2
I
I
I
.4
EQUIVALENCE
FIG,30
--
I
I -
.6
I
.8
RATIO
NITRIC OXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR ISO-OCTANE
-- ---I--
1.
-173-
_·_
__
I
I
I
7.
·
·
_·
_·
·
I1
I
I
I
I
100-600 D
6.
a
.
:r
I
.
a.
I
A
5.
A
(5
!
A
AA
4.
0
.
0
U')
2
0z
bn
3.
0
zU)
x
z
.
2. _
* 2000
* 2500
a
I.
A
I
_
I
0
I
I
I
I
.4
EQUIVALENCE
.2
0
FIG,31
I
I
.6
RATIO
i·--CII_1X_
I
.8
I
NITRIC OXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR 100-600 FUEL
_____llq
m
A 1500
.I-
1.
-174-
140
130
120
0
oO
a-
110
u
0l.
LI
0n
100
OO
Au
90
LLJ
Er
Q-
LL
2lu
u
0
LLuI
80
70
60
ZiJ
5: 50
z
<i
0
U
40
C3
30
20
I0
0
0
.2
.4
EQUIVALENCE
.6
RATIO
.8
FIG. 32 INDICATED MEAN EFFECTIVE PRESSURE VS
EQUIVALENCE RATIO FOR METHANOL
---
-175-
I.
I
I
I
I
TI
I
I
I
METHANOL
I'l
.8
2
-J
Z:
0
_0
nI
S
A
6
LL
I.L,
a:
00C
u-
LI)
<:
n
Or
C]
u
.4
A 1500 RPM
0 2000 RPM
-
.,...,
7
0
L
I
0
FIG. 33
I
.2
I
I
I
.4
.6
EQUIVALENCE RATIO
.8
I
I
I
I
INDICATED SPECIFIC FUEL CONSUMPTION FOR
METHANOL
1.
-176-
1001I(
............................
I
I
__
________
_ _ _ ___
I
I
I
I
I
I
L...........
___
I
METHANOL
z
-
o
0
80
2000
1500
-
-
-
Ld
G
60 _-
05
H
Dli
40 _
A 1500 RPM
C 2000 RPM
20 _
-
I
H
II
0
II
(I
II
.4
.2
0
·I
EQUIVALENCE
FIG
.6
RATIO
·1
·I
.8
34 INDICATED THERMAL AND VOLUMETRIC
EFFICIENCY VS EQUIVALENCE RATIO FOR
METHANOL
_
II
__eclllra-·-·-·-----------
·I
1.
-17 7-
I
I
I
I
I
I
I
I
I
METHANOL
14
12
O&
0,1
0
10
uLL
0
r-
8
A
I0
A
'r'
LIJ
6
A
0
n
I-
I
A
LLI
4
2
_-
A
A
1500 RPM
*
2000 RPM
_
I
0
I
0
.2
I
I
I
I
.4
EQUIVALENCE
.6
RATIO
I
I
I
.8
FIG. 35 EXHAUST TEMPERATURE VS EQUIVALENCE
RATIO FOR METHANOL
____1__1_
__I
_I___
I.
-178-
I
I
I
I
I
I
I
I1
METHANOL
35
30
I
Q-
25
::
(9
20
_
(I)
z
0
(I)
ii
15 _
0
U)
0
I0 _
A
·
A
A
A 1500 RPM
* 2000 RPM
A
0
5 _
A
I·
II
0
.2
·I
A
·I
_·I
.4
EQUIVALENCE
·
I 1I
.6
RATIO
·I
·I
.8
FIG. 36 CARBON MONOXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR METHANOL
1
-----
---- ^--·I1IIIIF-
--- I-----
-
1.
-179-
I
1
1
/I 1
-
I
I
18
METHANOL
I6
I4
-
12
(9
I0
A 1500 RPM
0 2000 RPM
w
C-)
(I)
8
0
A
U
C-)
I
A
A
6
0
0
4
0A
0
2
A A
--·
*
A
I
0
I
.
0
.
.2
FIG. 37
-----
CIII-
.---1--___-___---____-·^---·-··119··
I
I
I
.
.4
EQUIVALENCE
I
E
A
I
___
.6
RATIO
I
___
I
.8
HYDROCARBON EMISSIONS VS EQUIVALENCE
RATIO FOR METHANOL
--·--·-----
1.
-180-
II
I
7.
~~
~~
I
_____
~~~~~~I
I
I
I
I
METHANOL
6.
I
0r
I
'I-,
5.
.
0
(.9
I
(Ni
0
4.
z
U)
z
0
o)
~n
(n
*
0
3. _-
*
0
A
uJ
0
Z
A
2.
_
A
A
1500 RPM
0
2000 RPM
I. _
0
I
0
I
.2
I
I
i
.4
EQUIVALENCE
I
I
.6
RATIO
I
.8
I
FIG, 38 NITRIC OXIDE EMISSIONS VS EQUIVALENCE
RATIO FOR METHANOL
I--------
^-I-
-_---._------m----
--
1.
-181-
100
LU
I
I
I
1500
2000
RPM
2500
90
IL
LUI
LU
80 I-
0
70
VOLUMETRIC EFFICIENCY VS ENGINE RPM
FIG.
50
I
40 _
O)
30
_
I
L
LI
20 _
I0 _
0
1500
2000
RPM
FIG. 39
__ IIC__
FRICTION MEAN EFFECTIVE PRESSURE
VS ENGINE RPM1
2500
-182-
1.8
_·
_ __
_ _
1.6_
1.4
-\
1.2 _
H
INJ
wC 1.0 _
-J
(n
vn)
8
_
.6 _
.4
2_
1500 RPM
n _
~ J
-I. I
ISO-OCTANE
0 0.76
I
I
I
-. 9
-1.0
-. 8
-. 7
-. 6
LOG ( VOLUME / VOLUME MAX)
FIG. 40 LOG P vs LOG V FOR ISO-OCTANE,
= 076, 2500 RP1
--P·1SIU···s*IID·I··_-·I·----.·__
_._ _
.5
-183-
1.8
-I
.
·
·
_
i,6 _
1.4 -\1_
i.2
_
H
INJ
Li
1.0 _
D
(I,
b-i
8_
0
-i
.6 _
4 _
2_
1500 RPM
0
1.1
-1.0
ISO-OCTANE
I
-. 9
-. 8
-,7
-. 6
LOG ( VOLUME / VOLUME MAX)
FIG 41
LOG P vs LOG V FOR ISO-OCTANE,
= 0,45, 1500 RPM
-·
--
`--
,/0.45
---
·iil
.5
-184-
1.8
I
I
I
I
I
1.4 A\
i. 2
m
W
w7U
1.0 _
(n
in
Li
0.
.8
(9
0
J
6_
4_
2_
1500 RPM
0-,
_
L
1I
I
-1.0
ISO-OCTANE
I
-
~
~
_
~
~
~
~ ~
---
.
--
-. 9
-.7
-. 6
-.8
LOG(VOLUME / VOLUME MAX)
FIG 42 LOG P vs LOG
.
0.28
_
028i
FOR ISO-OCTANE,
1500 RPM
.
.
-185-
1.8
T
·
--i
·
I
I
1.6_
1.4
1.2
INJ
w 1.0
:
Un
(n
w
O~ .8 _
o
0
_J
6 _
.4
.2 i2000 RPM
I
n' _
~ J
_
-1.I
-1.0
-,6
-. 8
-. 7
LOG ( VOLUME / VOLUME MAX)
-. 9
FIG. 43 LOG P VS LOG
FOR ISO-OCTANE,
= 003, 2000 RPM
-
----I
0.83
ISO-OCTANE
-----~-------
--
-·------------
-186-
1.8
--
1,
.6
*\
1.4
1.2
M
m
1.0
i--
INJ
U
U-,
8
0
-J
.6 _
_
4
.2 _
2000 RPM
n
-
_
I I
- 1.0
ISO-OCTANE
1
I
-. 7
-. 9
-. 8
-. 6
LOG(VOLUME/VOLUME MAX)
FIG.44
LOG P. VS LOG V FOR ISO-OCTANE,
' = 0,.48, 2000 RPfl
-s-
lI·I(C---P(W
-
-·--
, 0.48
_5
-187-
l.b
!.6
!.4
1.2
,, 1.0
w
n
(n
Q.
0wL
J
.V
.6
.4
.2
0)
-1.1
-1.0
-. 9
-. 8
-. 7
-. 6
LOG(VOLUME / VOLUME MAX)
FIG. 45 LOG P VS LOG V FOR ISO-OCTANE,
= 0.22, 2000 RPM
--13*91--
·---
-s
-. 5
-188-
1.8
I
I
I
I
\-
1.6
1.4
1.2
m
H
w
cr
INJ
1.0
m
U)
(n
w
oo
8_
(9
0
-J
6_
4_
.2
_
2500 RPM
0
-
1.1
--
-1.0
ISO-OCTANE
--
0o. 8
--
-. 9
-. 8
-. 7
-. 6
LOG(VOLUME / VOLUME MAX)
FIG.46 LOG P VS LOG V FOR ISO-OCTANE,
= 0.81, 2500 RPM
-
-
.5
-189-
1.8
'-
1.4 __
1.2 _
A
INJ
u
1.0 _
~~
INJ
a-
.8
0D
On
o)
nr
0_J
6
_
.4 _
2
_
2500 RPM
OL
I,i
-
1
-1.0
_ ·
ISO-OCTANE
_I
0 0.52
_· I
_·I
-. 9
-. 7
.6
LOG ( VOLUME / VOLUME MAX)
FIG,47 LOG P vs LOG V FOR ISO-OCTANE,
= 052, RPM
"PC---
---
-- "----------r----r--
.5
-190-
1.8
I
I.
6 _
1.4_
1.2_
w
a:ioo
INJ
1.0 _
(n
.
8_
0
-J
.6
4_
2_
2500 RPM
0
-
I
ISO-OCTANE
0 0.25
1
.
-1.0
I
-. 9
-. 7
-. 8
-. 6
LOG ( VOLUME / VOLUME MAX)
FIG, 48 LOG P VS LOG V FOR ISO-OCTANE,
= 0.25, 2500 RPM1
~~'--I·"PBU"*IC--r~~~---~---------~~
-------·-^----·I~-
·---
-191-
-
1.8
1.6
·
I
,
I
I
1
w
\I
1.4
1.2
W
wcl:
FlLi
1.0
(n
0)
t.LI
v
cr
8
(9
0
-J
.6
4 _
.2
1500 RPM
·" ·
v
-
_
1I
-1.0
100-600 D
0 0.27
I
-. 9
-. 7
-. 8
-. 6
LOG ( VOLUME / VOLUME MAX)
FIG.49 LOG P VS L.OG V FOR 100-600,
= 0,77, 1500 RPMI
-s_.__...._
__
.5
-192-
1.8 _
1.6
\\
m
,\1
1.4
1.2
__
m
u
1.0
m
wLI
n
lLi
or
8 _
00
-i
6_
4 _
.2 _
1500 RPM
I
0
-
I.I
-1.0
100-600 D
I
1I
, 0.43
I
FIG, 50 LOG P vs LOG V FOR 100-600,
= 043, 1500 RPM
--
S
--·-----
I
-. 9
-. 8
-. 7
-. 6
LOG ( VOLUME / VOLUME MAX)
-__
-5
-193-
1.8
'N1
,6
1.
I\
1.4-'-
I-
1.2
H
W
Li
1.0
oo
~J
cr
8 _
a.
(.
(9
0
-J
6 _
.4
2_
1500 RPM
0
-v
.L
I
i
I.1
-1.0
100-600 D o0.77
I
I
FIG, 51
LOG P vs LOG V FOR 100-600,
= 0,27) 1500 RPM
_~-
.
-
~
-
I
i
-. 9
-. 8
-,7
-. 6
LOG ( VOLUME / VOLUME MAX)
-----`"1~~1^1--~~~"~~I----"I
1_1-
-194-
1.8
I
I
I
I
I
!.6
1.2_
m
A:
1.0_·
Ln
a. 8 _
0
-o
.J
6_
.4
.2 _2000 RPM
0
-
.I
I
-1.0
100-600 D
: 0.85
-. 9
-. 8
-. 7
-. 6
LOG ( VOLUME / VOLUME MAX)
FIG 52 LOG P vs LOG V FOR 100-600,
= 0.85, 2000 RPM
.1-1
..
-"----------I
.5
-195-
I
^
.0
I
I
I
I
I
1,6
1.4
11-1 1-1
1.2
-
w 1.0
D
(n
(I)
w
: .8
a.
If
r
0
-J
.6
.4 __
.2
2000 RPM
Q
-1.1
v
I
-1.0
I
100-600 D
I
0.47
I
FIG. 53 LOG P vs LOG V FOR 100-600,
= 0 47, RPM
-
-
__·__lls
I
-. 9
-. 8
-. 7
-. 6
LOG(VOLUME / VOLUME MAX)
___-·I··--·-··-··I----__.
-196-
1.8
I
I
I
I
I
1.6
1.4
\
1.2
H
INJ
1.0
(1)
tr
a. 8
(9
a
-J
6 _
.4
.2 _
2000 R PM
1
-- 1.0
n)
iJ_
-
I I
100-600 D
+ 0.26
i
-. 9
-. 7
-. 6
-. 8
LOG (VOLUME /VOLUME MAX)
FIG. 54 LOG P vs LoG V FOR 100-600,
, = 0.26, 2000 RPM
~~-··lls~~~·PllsP(--·*
Ii~~~~~~~··ll~~~·---~~-~~~
- -
------
I
.5
-197-
1.8
_
I
1.6_
_
1.4
1.2
H
INJ
U
W
1.0 _
U)
0K
o3
8_
0
-
6_
.4 _-
2
_
2500 RPM
0
- I
I
I
i
I
II
-. 9
-. 8
-. 7
-. 6
LOG (VOLUME / VOLUME MAX)
-1.0
.1
# 0.82
100-600 D
FIG. 55 LOG P vs LOG V FOR 100-600,
0.82, 2500 RP1
....
1~1c
_
S
------..
~
11---~
-
_5
-198-
I1. ?%
C5
1_
I
I
I
-I--
I
11-
1.6
1.4
1.2
F._2
I-L
c-
1.0
.8
(fD
.J
.6
.4
.2
~_
2500 RPM
I
-I.0
n
-1.1
100-600 D
1
-. 9
-. 8
0.50
I
-. 7
LOG(VOLUME / VOLUME MAX)
F ,.G.56 LOG P VS LOG V FOR 100-600,
= 0 50, 2500 RPf1
·__
. __I_
_
_-__
I
-. 6
-199-
1.8
_
_·
·
·
I
i
1.6
1.4
1.2_
INJ
w
1.0_
_J
(n
n
:D
0ir
L-
8_
vD
0o
6
_
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2
_
2500 RPM
Coi
-
,
I
-1.0
-
I.1
L
-r_
-. 9
100-600D
I
_I
-. 8
0 0.28
I
---- L
-
-. 7
·
-. 6
LOG( VOLUME/VOLUME MAX)
FIG. 57 LOG P vs LOG V FOR 100-600,
, = 028, 2500 RPM
I----··-----------
-200-
0
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