EPA-460/3-74-011
July 1974
A STUDY
OF THE DIESEL
AS A LIGHT-DUTY
POWER PLANT
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Emission Control Technology Division
Ann Arbor, Michigan 48105
-------
EPA-460/3-74-011
A STUDY OF THE DIESEL
AS A LIGHT-DUTY POWER PLANT
by
M. L. Monaghan, C. C. J. French, and R. G. Freese
Ricardo and Company Engineers (1927) Ltd.
Bridge Works
Shoreham-by-Sea, Sussex BN 45FG
Contract No. 68-03-0375
EPA Project Officers:
T. C. Austin, J. J. McFadden, and K. H. Hellman
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Emission Control Technology Division
Ann Arbor, Michigan 48105
July 1974
-------
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 22151.
This report was furnished to the Environmental Protection Agency by
Ricardo and Company Engineers (1927) Ltd. , in fulfillment of Contract No.
68-03-0375. The contents of this report are reproduced herein as received from
Ricardo and Company Engineers (1927) Ltd. 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-460/3-74-011
11
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1-1
CONTENTS
SECTION 1
EXECUTIVE SUMMARY
SECTION 2
INDEX AND LIST OF FIGURES
SECTION 3
GENERAL SUMMARY AND INTRODUCTION
SECTION 4
LITERATURE SURVEY, ASSESSMENT OF
PROBLEM AREAS AND TRADE OFFS
SECTION 5
ENGINE CONFIGURATION STUDY
SECTION 6
POWER PLANT RATING
SECTION 7
PROGRAMME PLANS
SECTION 8
APPENDIX 1 - LIST OF KEYWORDS
SECTION 9
APPENDIX 2 - LIST OF REFERENCES
SECTION 10
APPENDIX 3 - GLOSSARY OF TERMS
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1-2
THE DIESEL ENGINE AS A LIGHT DUTY POWER PLANT
SECTION 1
EXECUTIVE SUMMARY
Introduction
In America today the gasoline engine is virtually the only engine
used for light duty purposes due to its low initial cost and
refined performance. However, current and proposed legislation
and the prospect of increased fuel costs may render the low
emissions gasoline engine less attractive than some of the other
available power plants.
In other parts of the world the diesel engine, which is the most
efficient practical engine produced today and which has good
emissions characteristics, has long been used as a light duty power
plant in applications where fuel economy and durability are of great
importance. The lightweight, high speed diesel has been developed
specifically for European and Japanese conditions while low fuel costs
have until now prevented any major effort being expended on a light
duty engine for America.
The purpose of this study was to see if the diesel engine could be used
as a viable power plant for an American passenger car. It was felt
necessary to carry out the study because European effort has tended
towards the lighter weight European vehicles while American effort
has been towards meeting legislative limits with the gasoline engine
while retaining traditional performance standards.
Proposed legislation indicates that a light duty vehicle for the near
future should be able to meet emissions targets of : -
HC 0.41 g/mile
CO 3.4 g/mile
NOX 1.5 g/mile
when tested by the CVS-CH procedure.
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1-3
A light duty vehicle for longer term application should be able to meet
emissions targets of :-
HC 0.41 g/mile
CO 3.4 g/mile
NOX 0.4 g/mile
For the purposes of this study the vehicle considered was a 4/5 seat
sedan with a loaded weight of about 1600 kg (3500 Ib), capable of
0-9(7 km/h (0-60 mph) in 13.5 s and 32-113" km/h (20-70 mph)
in 15 s, i.e. a compact size sedan but with "standard1 performance
capabilities.
The first phase, a literature survey, involved a study of all existing
light duty and relevant heavy duty diesel literature from published
sources in Europe, America, Japan and also from Ricardo in-house
technical reports. Visits to diesel manufacturers and users were also
made. At the end of this phase, broad conclusions on the feasibility and
likely problem areas of an American light duty diesel could be made.
The next phase required that brief design studies be made to cover
all the potentially viable diesel power plants. This involved the
calculation of performance and outline design of any type of diesel
which could power the target vehicle in the above emissions climate.
For the third phase a rating methodology was derived which allowed
a numerical comparison of all the potentially viable power plants to
be made.
The fourth and final phase of the study involved a consideration of the
results of the previous sections in order to make recommendations
for further action and effort to achieve the most desirable light duty
vehicle. Although these recommendations were to be aimed mainly
at light duty use, other areas uncovered in the course of the study
were not specifically excluded.
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1-4
Conclusions
On completion; of the study the following broad conclusions were drawn : -
1. .A diesel pow.eredpassenger car could be built using
present technology.
It would have equal acceleration and general performance
to typical gasoline powered passenger cars.
The noise level would be perfectly acceptable and in fact
would be less than 1 dBA greater than a gasoline powered
vehicle in an S.A.E. drive-by test.
The engine would be hardly any larger than the equivalent
V-8 gasoline engine and would weigh only 68 kg (150 Ib)
more.
There would be very little if any visible smoke under
normal driving conditions.
The diesel powered vehicle would meet the primary
emission targets without the use of catalysts or other
special equipment.
The diesel powered vehicle will deliver up to 50% greater
fuel economy than the equivalent gasoline powered vehicle,
depending on the driving cycle.
2. The study indicates that a diesel powered vehicle could not
meet the secondary emissions target of 0.4 g/mile NOx
if the target vehicle and performance standards are adhered
to and if current technology is assumed.
Virtually no work has been carried out to determine the
emissions capabilities of the light duty diesel for these
secondary targets.. More basic research is required to
determine the ultimate emissions potential of the light
duty diesel.
It would seem that a diesel powered vehicle could meet
a 1.0 g/mile NOX standard with the use of modulated exhaust
gas recirculation but lower standards than 1.0 g/mile would
require a reduction in vehicle performance and/or weight.
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1-5
Recommendations
As-a: result of the study the following recommendations were made :-
1... A V-8 97 kW(!30 bhp) naturally aspirated indirect
injection diesel engine should be constructed and
demonstrated in a 4/5 seat sedan.
2-, A 6 cylinder indirect injection diesel engine should be
built and work carried out with the engine to investigate
the application of turbo-chargers and "Comprex" pressure
exchangers to automotive engines.
3,. Advanced fuel injection systems could help to bring about
ftirther reductions in pollutants and make it possible to
use the direct injection diesel: as a possible automotive
power plant, while novel forms of construction may allow
some reduction in the cost of fuel injection equipment.
Thus work should be pursued on advanced and novel forms
of fuel injection equipment.
4*. A. thoro ugh investigation should be made of the potential
of the low compression ratio , ignition assisted, indirect
injection diesel engine in view of its possible fuel economy
advantages.
5- A fundamental investigation should be carried out into the
formation of unburnt hydrocarbons in diesel engines. The
development of a catalyst for control of hydrocarbons is also
desirable.
6.. Work should be initiated to develop control systems for
load and speed modulation of exhaust gas recirculation.
7. Work on the development of improved starting aids such as
'instant1 glow plugs and programmed starting sequences
should be carried out.
8. The clear superiority of the diesel for such specialised
applications as taxi cabs and light delivery vehicles indicates
that a programme to demonstrate and encourage the conversiot
of these vehicles to diesel power should be instituted.
9% An investigation into particulate formation in the diesel engine
should be carried out together with the development of smoke
and particulate traps.
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Literature Study
Current light duty diesel experience is mainly concerned with European
type vehicles and thus the literature study was carried out to indicate
the feasibility of the diesel engine for American light duty use and the
likely problem areas and possible trade-offs if it were adopted.
The survey covered the period 1919 to the present day although the
great bulk of the literature studied was published in the period 1950
to the present. A key-word indexing system was devised so that
pertinent points could be extracted easily from the literature and so
that data relevant to a particular point could be assembled quickly
and thoroughly. Sources of information in this section of the study
were proceedings of learned societies , technical journals , Ricardo
published information, Ricardo unpublished information and discussions
with manufacturers and users of light duty diesel engines.
The overall conclusion of the literature study was that the diesel engine
is a potentially viable power plant for the target vehicle and that it could
meet the primary emissions targets without the use of a catalyst.
Specific conclusions on various performance aspects were as follows :-
Smoke - The target vehicle should emit very little visible smoke and
under normal driving conditions there should be virtually none.
Odour - Simple timing controls should minimise the slight odour of the
high speed diesel under light load conditions while the use of an indirect
injection combustion chamber and the proposed restricted power rating
should minimise the full load odour.
Gaseous Emissions - A naturally aspirated indirect injection engine
capable of powering the prototype vehicle should be able to meet the
primary emissions targets of 0.41 g/mile HC, 3«4 g/mile CO and
1.5 g/mile NOX without the use of catalysts, although some small amount
of exhaust gas recirculation may be required to ensure sufficient margin
for production compliance. The secondary target of 0.4 g/mile NOX could
not be achieved with current technology if the target vehicle size and power
to weight ratios are assumed. The use of conventional direct injection
systems would make it difficult to meet both NOX and CO standards while
the injection timing retard required would cause high HC levels.
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1 - 7
Particulates - The diesel, like all engines employing heterogeneous
combustion, emits more exhaust particulates than the gasoline engine,
but neither the effect of particulates on health nor the importance of
the nature of particulates is known at present. Filter systems could
probably remove diesel particulates if necessary.
Noise - Attention to engine and vehicle design would make the diesel
passenger car perfectly acceptable. Drive-by noise levels would be
only slightly higher (1 or 2 dBA) than the gasoline powered vehicles.
Idling noise can be annoying with current diesel engined vehicles but
it is unlikely to be a significant cause for complaint with the target
vehicle.
Volume - The volume of the diesel engine will be greater than that
of the equivalent highly rated gasoline engine, but there should be
no installation problems in the target vehicle.
Weight - A 97 kW (130 bhp) diesel engine may be 136 kg (300 Ib)
heavier than a highly rated gasoline engine of the same power output
but only about 68 kg (150 Ib) heavier than a more normal and lower
rated gasoline engine.
Fuel Economy - The diesel engine is undoubtedly the most efficient
current prime mover for light duty use and in city conditions the
fuel consumption advantage may be as much as 50% (twice the miles
per gallon).
Fuel - The increasing cost of fuel makes the high economy diesel
more attractive and there should be no difficulty in increasing the
quantity of automotive diesel fuel.
First Cost - The diesel engine will cost between 1.5 and 2 times as
much as the gasoline engine in America (i.e. % 300 against % 200
production cost ) and about half of the cost increase is due to the
fuel injection equipment.
Maintenance - Longer periods between overhauls and the similar
nature of the maintenance requirements to those of the gasoline
engine will mean that overall maintenance costs will be less for
the diesel powered vehicle.
Starting - Starting is inferior to the gasoline engine but the maximum
delay with heater plugs will be 30 seconds.
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1-8
Hot Driveability - The hot driveability of the diesel engine will be
very similar to that of the gasoline engine and no major transmission
changes will be required.
Cold Driveability - The cold driveability of the diesel engine will be
as good as its hot driveability and very much superior to the gasoline
engine.
Torque Rise - The torque rise of the diesel engine will be similar
to that of the gasoline engine.
Durability - The diesel engine will have greater durability than its
gasoline counterpart.
Coolant. Heat Loss - The diesel has reduced heat loss at low load and
idle so that ' slow speed traffic overheating ' will not be a problem
although winter morning de-icing may be more difficult. The
increased heat losses at full load may dictate the use of a larger
radiator, although this may only be required for trailer-towing
vehicles.
Vibration and Torque Recoil - The unthrottled, high compression ratio
diesel undoubtedly vibrates more than the gasoline engine especially at
idle. This tends to give an impression of harshness with current
vehicles but this impression disappears once the vehicle is in motion.
The problem can be minimised by careful attention to engine and
transmission mountings.
Manufacture - The similarities between diesel and gasoline engines
allow them to be made on the same production lines if necessary although
the diesel engine does require closer control of tolerances.
Ancillaries - Ancillaries are similar and often identical to those of
the gasoline engine.
Production techniques for the high volume production of conventional
fuel injection equipment have already been developed but there may be
scope for cost reductions with greater production numbers.
Lubrication - The diesel engine requires oils with greater dispersant
and anti-corrosive qualities than does the gasoline engine.
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1-9
Engine Configuration Study
The literature study and Ricardo in-house knowledge indicated that
a diesel engine could provide a viable power plant for a passenger
car and thus the study required that all potentially viable diesel
variants be designed in sufficient detail to allow their likely performance
to be assessed realistically. The vehicle to be powered by any of the
candidate power plants was a 4/5 seat sedan weighing less than 1600 kg
(3500 Ib) , and capable of meeting the EPA standard car performance
specifications, i.e. 0- 97 km/h (0-60 mph) in less than 13.5 s. ,
32-113 km/h (20-70 mph) in less than 15 s and capable of overtaking
a 80 km/h (50 mph) truck in less than 15 s. The emission targets were
Primary (or short term)
HC 0.41 g/mile
CO 3.4 g/mile
NOX 1.5 g/mile
Secondary (or long term)
HC 0.41 g/mile
CO 3.4 g/mile
NOX Oo4 g/mile
and were to be obtained when the vehicle was tested according to the
CVS-CH procedure. Computer calculations indicated that a 3-speed
automatic gear box would require an installed power of approximately
97 kW (130 bhp) for engines with normal torque characteristics, and
thus all the diesel power plants were designed with this in mind.
While no specific gasoline configuration exercise was required, two
gasoline engines were studied to provide a basis for comparison
with the diesel engine types.
The diesel engine types considered covered all of the major combustion
systems, all likely engine configurations and operating cycles (e.g.
4 stroke, rotary, turbo-compound) and different possible installations
(e<»g. air-cooled, water-cooled), but the only ones schemed in detail
for this study were those engine configurations and types which were
considered at all suitable for passenger car application.
The following paragraphs summarise the results of the configuration
study while Table 1 shows some of the more important engine
parameters for the diesel engines.
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o mm
Bore
in
c. , rnm
Stroke
in
kW
Power ,
hp
Swept litre
Volume in3
Weight ^
Box m3
Volume ft3
Specific power- kW/1
o
swept volume hp/in
Power/unit kW/mm
piston area hp/in^
Specific power- kW/m3
box volume hp/ft3
Specific weight kg/1
- swept volume lb/in3
Specific weight kg/kW
- power Ib/hp
V8
Gaso-
line
97
3.82
76
3.00
96
128
4.5
275
250
550
_
21.3
0.47
.
-
55.6
2.00
2.60
4.30
IL6
Gaso-
line
88
3 .46
82
3.22
96
128
2.99
183
186
410
_
32.11
0.70
-
-
60.2
2.19
1.87
3.12
V8
NA
IDI
88
3.46
98
3.86
96
128
4.78
292
320
700
0.32
11.2
20.2
0.44
0.0020
1.70
302
11.4
66.9
2.41
3.31
5.47
V6
TC
IDI
90
3.54
100
3.94
96
128
3.84
234
309
680
0.33
11.6
24.4
0.55
0.0025
2.15
291
11.05
80.9
2.91
3.22
5.31
In Line
6 TC
IDI
90
3.54
100
3.94
96
128
3.84
234
327
720
0.36
12.7
24.4
0.55
0.0025
2.15
265
10.1
85.6
3.08
3.41
5.62
V6
TC
DI
93
3.66
94
3.70
96
128
3.84
234
300
660
0.31
11.0
24.4
0.55
0.0023
2.03
306
11.6
78.5
2.83
3.12
5.16
In Line
6 TC
DI
93
3.66
94
3.70
96
128
3.84
234
310
680
0.34
12.0
24.4
0.55
0.0023
2.03
280
10.7
80.7
2.91
3.22
5.31
V6
2 Stroke
Loop IDI
99
3.89
114
4.50
96
128
5.25
320
340
760
0.36
12.7
18.3
0.40
0.0021
1.82
265
10.1
65.7
2.37
3.59
5.93
In Line 6
2 Stroke
Uniflow DI
83
3.28
114
4.50
96
128
3.74
228
365
800
0.47
16.8
26.0
0.57
0.0029
2.52
201
8.0
98.4
3.55
3.78
6.25
4 Cyl.
Compound
DI
93
3.66
93
3.66
96
128
2.52
153
305
670
0.32
11.3
38.0
0.83
0.0035
3.04
299
11.3
120.6
4.34
3.17
5.23
2 Rotor
2 Stage
Rotary
-
-
96
128
^
227
500
0.26
9.2
-
-
368
13.9
_
2.36
3.91
I
*-»
o
SUMMARY TABLE OF MAJOR CHARACTERISTICS OF
POWERPLANTS CONSIDERED IN ENGINE CONFIGURATION STUDY
CO
£
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1-11
1. V-8 Gasoline Engine
8 cylinders 97 mm x 76 mm (3.82" x 3.00")
4.5 litre (275 CID)
96 k'w (128 bhp) at 66.7 rev/s
290 N.m (210 Ib.ft) at 41.6 rev/s
250 kg (550 Ib) estimated weight
For the primary emissions target (1.5 g/mile NOX) the engine would
require sophisticated close tolerance carburettors, modulated exhaust
gas recirculation, air injection into the exhaust manifold and an oxidation
catalyst in the exhaust system.
For the secondary emissions target (0.4 g/mile NOX> a reducing
catalyst and a further catalytic device 'getter box1 or some similar
system would be needed.
In primary emissions build, fuel consumption would be approximately
18 1/100 km (13 mpg) over the CVS-CH cycle while the need to run
closer to stoichiometric in the secondary emissions build would reduce
this to 19 1/100 km (12.5 mpg).
Drive-by noise level would be 74 dBA under US Federal Test conditions.
2. In-Line Gasoline Engine
6 cylinders 88 mm x 82 mm (3.46" x 3.i'.3")
3 litre (183 CID)
96 kW (128 bhp) at 83.3 rev/s
236 N.m (171 Ib.ft) at 50 rev/s
186 kg (410 Ib) estimated weight
For the primary emissions targets the engine would be equipped with
gasoline injection, modulated exhaust gas recirculation, air injection
into the exhaust manifold and an oxidation catalyst.
For the secondary emissions targets the addition of a reducing
catalyst and protective 'getter1 box will be required.
Fuel consumption over the CVS-CH" cycle should be 15.5 1/100 km (15 mpg)
Drive-by noise level will be 77 dBA.
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1 - 12
3. Naturally Aspirated 4-stroke
V-8 Indirect Injection piesel Engine
8 cylinders 88 mm x 98 mm (3.46" x 3.86")
4.78 litre (292 CID)
96 kW (128 bhp) at 66.7 rev/s
290 N.m (210 Ib.ft) at 33.3 rev/s
320 kg (700 Ib) estimated weight)
For the primary emissions targets the engine should only require
retardation of the injection timing to reduce NOX levels to 1.2 - 1.5
g/mile although some slight exhaust gas recirculation might be
necessary to ensure production compliance.
It is not thought that this type of engine could meet more stringent
NOX limits than 1.0 g/mile with conventional injection equipment
and current combustion systems.
Fuel consumption with this engine should be 11.5-10.5 1/100 km
(20-22 mpg) on the CVS-CH cycle.
Drive-by noise should be 76 dBA.
4. Boosted 4-stroke 6 cylinder
Indirect Injection Diesel Engine
6 cylinders - 90 mm x 100 mm (3.54" x 3.94")
3.84 litre (234 CID)
96 kW (128 bhp) at 60 rev/s
290 N.m (210 Ib.ft) at 33,3. rev/s
327 kg (72CHb) (in-line-6) estimated weight
309 kg (680-lb) (V-6) estimated weight
This engine would be lightly boosted by a turbocharger to attain the
above performances. In-line and V configurations were schemed.
For the primary emissions targets, only injection retard should be
required although some slight exhaust gas recirculation might be
needed to ensure production compliance.
As for the naturally aspirated engine, no secondary emissions
configuration was schemed.
The target vehicle, equipped with this engine,, should achieve 12-11
1/100 r.,n (19-21 mpg) during the CVS-CH cycle, the poorer part load
consumption of the turbocharged engine causing this penalty when
compared with the naturally aspirated engine.
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1 -r 13
Drive-by noise level should be 75 dBA.
5. Comprex on 4-stroke 6-cylinder
Indirect Injection Diesel Engine
6 cylinders - 90 mm x 100 mm (3.5-i" x 3.94")
3.84 litre (234 CID)
96 kW (128 bhp) at 60 rev/s
290 N.m (210 Ib.ft) at 33.3 rev/s
327 kg (720 Ib)estimated weight
This engine, only schemed in in-line form due to manifolding problems
with the pressure exchangers, would employ the 'Comprex' to provide
a slight improvement in torque over more of the speed range than with
a turbocharger and to give a more driveable engine. It is considered
that suitable sound insulation couli be fitted to the 'Comprex'.
The emissions of the engine will be the same as for the turbo-charged
version and fuel consumption and drive-by noise should be similar at
12 - 11 1/100 km (19 - 21 mpg) and 75 dBA.
6. Boosted 4-stroke 6-cylinder Direct
Injection Diesel Engine
6 cylinders - 93 mm x 94 mm (3.66" x 3.70")
3.84 litre (234 CID)
96 kW (128 bhp) at 60 rev/s
290 N.m (210 Ib.ft) at 33.3 rev/s
310 kg (680 Ib) (In-line 6) estimated weight
300 kg (660 Ib) (V-6) estimated weight
It was considered that only a boosted direct injection engine would
have adequate power for a reasonable weight and volume and would
allow sufficient injection timing retard for emissions control.
The engine schemed out would only be capable of meeting NOX limits
of 2.5 g/mile and would be retarded by 10° crank and have 10%
exhaust gas recirculation even to achieve this. HC levels at these
retarded timings would be very high and an efficient oxidising catalyst
has been assumed to allow the 0.41 g/mile target to be met. (Such a
catalyst does not exist at present.)
Predicted fuel consumption is 11 - 10 1/100 km (21 - 23 mpg) on the
CVS-CH cycle and predicted drive-by noise level is 82 dBA.
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1-14
8. Loop Scavenged 2-stroke 6-cylinder
Indirect Injection Diesel Engine
6 cylinders - 99 mm x 114 mm (3.89" x 4.50")
5.27 litre (321 CID)
96 kV (128 bhp) at 45 rev/s
400 N.m (290 Ib.ft) at 33.3 rev/s
340 kg (760 Ib) estimated weight
For this engine a 90 V-6 configuration was chosen.
It was considered that the primary emissions NOX level of 1.5 g/mile
could be met, but there is little doubt that the HC target of 0.41 g/mile
woqld be unattainable with present combustion systems and fuel
injection equipment while the levels are likely to be so high that
no current catalyst is known to be capable of ensuring compliance.
Predicted fuel consumption is 13 - 11.5 1/100 km (18 - 20 mpg) arid
the predicted drive-by noise level is 75 dBA.
9. Uniflow 2-stroke 6-cylinder
Direct Injection Diesel Engine
6 cylinders - 83 mm x 114 mm (3.28" x 4. 50" )
3.74 litre (228 CID)
96 kW (128 bhp) at 45 rev/s
414 N.m (300 Ib.ft) at 33.3 rev/s
365 kg (800 Ib) (In-line 6) estimated weight
354 kg (780 Ib) (V-6) estimated weight
Two configurations of this engine were schemed out, an in-line 6
cylinder and a 90 V-6.
With this form of combustion system it is unlikely that the NOX targets
can be met and there is some doubt as to whether the HC levels will
be below the target without the use of a catalyst. These predictions
assume current technology.
This engine should allow the target vehicle to achieve 12.5 - 11.5 1/100 km
(19-21 mpg) over the CVS-CH cycle, and drive-by noise level should be
75 dBA under U.S. Federal Test conditions.
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1-15
10. Compound 4-stroke 4 cylinder Direct
Injection Diesel Engine
4 cylinders - 93 mm x 93 mm (3. 66" x 3. 66")
2.52 litre (154 CID)
Turbine and compressor directly geared to crankshaft
96 kW (128 bhp) at 50 rev/s
318 N.m (230 Ib.ft) at 50 rev/s
305 kg (670 Ib) estimated weight
For light duty application the directly geared turbo-compound
configuration was chosen since it gives good part load fuel
consumption and is the most suitable construction for mass-
produced passenger car application. The torque curve is such
that a new transmission system is required since the torque is
at maximum at full speed full load and falls at all conditions
below this. The high rating will prevent the primary NOx target
of 1.5 g/mile being achieved while the need for low compression
ratios and the very wide range of fuel quantities to be burnt in
the combustion system will almost certainly give a very high
hydrocarbon level. If the engine were produced and if a suitable
transmission could be used, fuel consumption might be 11.2 -
10.2 1/100 km (21 - 23 mpg) on the CVS-CH cycle.
Drive-by noise level could be as low as 75 dBA.
11. 2-staqe 2 bank Rotary Diesel Engine
2-stage, 2 rotors - 5.35 litre (326 CID) L.P. rotor
96 kW (128 bhp) at 66.7 rev/s (output shaft)
290 N.m (210 lb,ft) at 25 rev/s
227 kg (500 Ib) estimated weight
This engine would consist of two banks of a two-stage rotary engine.
The first L.P. stage would be 5.35 litre (326 CID) per lobe (2:3
geometry) while the second stage would be 1.343 litre (82 CID) per
lobe. The combustion system would be of swirl chamber type.
The output shaft would be linked to the rotors to provide a 2:3 speed
reduction.
This engine would probably be able to meet the NOX limit of 1.5 g/mile.
However, the difficulties of achieving good combustion with this type
of engine imply that both HC and CO levels will be extremely high and
the target levels certainly would not be attainable without some major
improvement in catalyst technology.
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1-16
Fuel consumption would be little better than with the gasoline engine over
the CVS-CH cycle and is predicted to be 16.4 - 14.7 1/100 km (14-16 mpg).
There is some evidence that the engine would be quiet and a drive-by
noise level of 75 dBA is predicted.
Power Plant Rating
One of the major aims of the study was that a methodology should
be derived which would allow a quantitative assessment of the
relative merits of various power plants for light duty vehicle use.
Although the study was concerned only with a comparison of gasoline
and diesel configurations, the methodology was developed so that it
could be applied to any liquid hydrocarbon power plant and thus should
be of value in other studies.
The advantages of such methodology are that its application will allow
a direct quantitative rating of various power plants and that it should
also be possible to identify those factors and aspects which render a
particular power plant suitable for a particular duty. The second
advantage allows an assessment of changes in a particular area as well
as high-lighting areas worthy of effort to make a particular configuration
more suitable for use in a given environment.
The fitness of a power plant for light duty use can be assessed under
the following broad headings :-
a) Emissions
b) Package (Size, weight, etc.)
c) Costs
d) Nature (Driveability)
e) Others (Convenience and minor safety aspects)
These headings were too broad for a detailed measurement of different
power plants and a more detailed list of performance aspects was drawn
up such that the various aspects would cover all facets of the operation
of light duty vehicle power plants fuelled by liquid hydrocarbons.
The individual performance aspects were as follows :-
1. Smoke
2. Particulates
3. Odour
4. NOX
5. HC
6. CO
7. S02
8. HC reactivity
9. Evaporative emissions
-------
1 - 17
10. Misc. emissions
11. Drive-by noise
12. Package volume
13. Package weight
14. Fuel economy
15. Fuel
16. Vehicle first cost
17. Maintenance cost
18. Startability
19. Hot driveability
20. Cold driveability
21. Torque rise
22. Durability
23. Coolant heat loss
24. Fire risk
25. Idle noise
26. Vibration and torque recoil
The performance aspects were generally those studied in the literature
survey although certain others were added so that the rating methodology
would be complete.
The final rating involved the assignment of weighting factors to each
of the performance factors, followed by the rating of each candidate
engine under the same heading.
The weighting factors were assigned by a group of eighteen Ricardo
engineers, each with experience of the emissions field and of the
American automotive situation, who were each instructed to act in
isolation in assessing the factors. The weighting factors, which are
listed in Table 2, are the arithmetic means of the values assigned
by the members of the group.
The rating of the individual engines involved the devising of a scale
which would cover both quantitative and qualitative assessments.
As expected, some difficulty was experienced in relating purely
subjective impressions to a linear quantitative scale, but after some
consideration the following system was adopted as giving several easily
relatable, subjective key points to the numerical scale, the numbers
without definition being an interpolation of the surrounding merit
definitions.
-------
1 - 18
Merit Rating Scale
0 Totally unacceptable
1
2 Bad
3
4 Poor
5 Acceptable
6
7 Good
8
9 Best practical
10 Perfect
-------
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
1 - 19
TABLE 2
Final Weightings Used in Study
Aspect Weighting
Smoke 4.48
Particulates 2.14
Odour 4.48
NOX 3.92
HC 3.99
CO 3.61
SO2 . 3.48
HC reactivity 1.83
Evaporative Emissions 1.60
Miscellaneous Emissions 0.98
Noise (Drive-by) 6.32
Package volume 2.61
Package weight 2.59
Fuel economy 12.20
Fuel cost 5.40
Vehicle first cost 4. 65
Maintenance cost 4.35
Startability 4.85
Hot driveability 4.48
Cold driveability 3.52
Torque rise 1.98
Durability 4.80
Heat loss 2.18
Fire risk 3.55
Idling noise 3.83
Vibration and torque recoil 2.18
-------
1-20
The rating system evolved allowed an immediate quantitative
assessment of the overall merit of the power plant which was
accomplished by multiplying each aspect "rating1 by its
appropriate 'weighting1 and summing all products. With a
total weighting of 100 and a merit scale of 0 - 10 as above,
the maximum possible score is 1000.
A committee was used to assign ratings to the various power
plants. The committee consisted of five experienced members
of the Ricardo staff and great care was taken to ensure that the
committee had no bias to either diesel or gasoline power plants.
The power plants considered were those described in the "engine
configuration" section of the report with the addition of the two
gasoline engines described briefly in the same section.
Each of the power plants was to be considered for the primary
and secondary emission levels of the study, i.e.
Primary Targets
HC 0=41 g/mile
CO 3.4 g/mile
NOX 1.5 g/mile
Secondary Targets
HC 0.41 g/mile
CO 3.4 g/mile
NOX 0.4 g/mile
but since it was considered that none of the diesel power plants
could possibly meet the secondary targets (in particular the 0.4
g/mile NOX figure) the rating system was only applied to an
environment embracing the primary targets <,
The ratings for each performance aspect are given in Table 3.
The designation (O) indicates that the particular engine is completely
unable to meet the particular target, which might be considered to
be a disqualification for light duty use.
-------
1-21
TABLE 3
Final Ratings for Each Aspect
Aspect
1 Smoke
2 Particulates
3 Odour
4 NOX
5 HC
6 CO
7 S02
8 HC reactivity
9 Evap. emissions
10 Misc. emissions
11 Noise (drive by)
12 Package volume
13 Package weight
14 Fuel economy
15 Fuel cost
16 Vehicle first cost
17 Maintenance cost
18 Startability
19 Hot driveability
20 Cold driveability
21 Torque rise
22 Durability
23 Heat loss
24 Fire risk
25 Idling noise
26 Torque recoil
o>
c
•iH
"o
w
CO
0
oo
9
7
7
5
5
5
7
5
5
5
7
7
6
5
5
6
5
6
7
5
7
6
5
5
8
8
_c
I-H
o
w
CO
t
1 — 1
CJ
8
7
7
5
5
5
7
5
5
5
6.5
8
7
5.5
5
7
6
6.5
7
6
7
5 = 5
5
5
8
8
5
I-H
oo
6
2
4
5
5
6
4
7
7
5
6
6
5
7.5
5
' 5
6
5
7
8
7
7
5
6
5
6
0
1— 1
Q
o
1— 1
o
4,5
2
4
5
5
5
4
7
7
5
6
5
5
7
5
5
6
5
6
8
7
6
5
6
5
5
mprex
0
O
Q
0
O
vO
5
2
4
5
5
5
4
7
7
5
5
4
5
7
5
5
6
5
i
8
7
6
5
6
5
5
^
3
i — i
o
:4.5
2
3
(0)
(0)
5
4
7
7
5
4
5
5
8
5
5
5
5
6
8
7
6
5
6
3
5
X
CD
!H
a
0
u
3
i-H
O
5
2
3
(0)
(0)
5
4
7
7
5
3
4
5
8
5
5
5
5
7
8
7
6
5
6
3
5
O
(H
Crt
iv
3
CM .
i
2 |
2
2
5
(0)
(0)
4 I
7
7
c
5
6
8
7
5
5
4
2
2 i
7 i
8
7 s
2 |
4 !
6
5
7
-------
1-22
The product of the weighting and rating for each performance aspect
was summed up for each power plant and the results of this operation
are shown in Table 4 : -
TABLE 4
Power Plant Final Score
(Rounded to nearest
whole number)
1
2
3
4
5
6
7
8
9
10
11
V-8 gasoline
6 cylinder gasoline
V-8 IDI
6 cylinder IDI T/C
6 cylinder IDI Comprex
6 cylinder DI T/C
6 cylinder DI Comprex
2-stroke loop scavenge
2-stroke uniflow
Compound
2 stage rotary
608
620
587
556
554
(500)
(497)
(516)
(515)
(465)
(434)
In order to establish the validity of the "committee method1 of
rating two re-runs were done for the gasoline engines and the
IDI diesels . The final runs for these power plants were found
to be within - 2-J% of the above figures and the relative order of
scores was not found to change. Because of this the final scores
from the first complete run shown in Table 4 were taken as
representative of the relative merit of the various power plants.
The rating methodology has the disadvantage that a rating of 0 in
any aspect can be hidden by good ratings for other aspects, and
the bracketed figures in Table 4 are for those power plants which
scored a 0 (i.e. totally unacceptable) in one or more performance
aspects.
The results indicate that the gasoline power plants are superior to
the diesel power plants and that only the indirect injection 4-stroke
diesels are viable for the duty considered. All the other diesel
power plants are unacceptable due to the inability to score better
than zero on one or more performance aspects (generally emissions)
-------
1-23
With the weightings adopted, the gasoline engine's superiority for
passenger car use in America is demonstrated but the closeness
of the final scores indicates that only quite minor changes in the
weightings would bring the scores equal. It is apparent that in an
emissions and fuel conscious environment there are many applications
which would change the individual weighting systems sufficiently to
make the automotive diesel an attractive alternative to the gasoline
engine.
Without a breakthrough in the reduction of exhaust emissions of
diesel engines, no diesel power plant can meet the secondary
emissions target however.
-------
N&:*%&
GASOLINE ENGINE
K:;:!:;:;:;:;::::^ DIESEL ENGINE
OUTLINES OF A 96 kW V-8 NATURALLY ASPIRATED I.D.I.
DIESEL ENGINE AND A CURRENT 51 (3OO CIDJ
AMERICAN V8 GASOLINE ENGINE
IV* (/I
•s
22
-------
2 - 1
SECTION 2
INDEX AND LIST OF FIGURES Page
Contents 1-1
1. Executive Summary 1-2
Conclusions 1-4
Recommendations 1-5
2. . Index and List of Figures
Index 2-1
List of figures 2-2
3. General Summary & Introduction
Summary 3-1
Introduction 3-4
4. Literature Survey
Summary 4-1
Introduction 4-2
Scope of survey 4-3
Performance aspects 4-3
Conclusions 4-50
Areas requiring future work 4-56
5. Engine Configuration Study
Summary . . 5-1
Configuration study 5-2
C.VS-CH consumption $ emissions
estimation 5-49
6. Power Plant Rating
Summary 6-1
Performance aspects 6-2
Weighting factors 6-3
Rating scale 6-5
Results of rating 6-7
Final rating 6-33
Results 6-34
Conclusions 6-35
7. Programme Plans 7-1
8. Appendix 1 - Keyword System 8-1
9. Appendix 2 - List of References 9-1
10. Appendix 3 - Glossary of Terms 10-1
-------
2-2
LIST OF FIGURES
SECTION 1
Fig. 1-1 Outlines of a 96 kW V-8 Naturally Aspirated I.D.I.
Diesel Engine and a Current 51 (300 CID)
American V-8 Gasoline Engine.
SECTION 4
Fig. 4-1 Bar Chart of Institution and Conference Papers Surveyed
4-2 Bar Chart of Journals Surveyed
4-3 Study of European Diesel Vehicle Operators Contacted to
Obtain User Experience Data
4-4 2 litre D.I. Conversion
Performance Comparison - D.I. and Comet V Builds
4-5 Comparison of Full Load Characteristics for Three
Combustion Chambers on 2-litre 4-cylinder Engine
4-6 NO Emission and Performance Characteristics of
D.I. and Swirl Chambers
4-7 Effect of Exhaust Gas Recycle on Performance and
Smoke of a Six Cylinder Swirl Chamber Engine over its
Load Range at 28 rev/s (1700 rev/min)
4-8 The Response of Naturally Aspirated Engine
NOX Emissions to Load
4-9 The Response of Turbocharged Engine
NOX Emissions to Load
4-10 Light Duty Diesel Vehicle Emission Levels - 1975 FTP (CVS-CH)
4-11 Data on Current Small 4-Cylinder High Speed Diesel
Engines for Automotive Applications
4-12 Comparison of Diesel and Gasoline Engine
Weight/Swept Volume
4-13 Diesel versus Gasoline Weight Analysis
for 4 Cylinder Engines
4-14 Power/Litre of Some Current Gasoline and Diesel Engines
-------
4-51
Odour at full load can be minimised by combustion chamber development.
The proposed reductions in smoke levels should alleviate this problem.
The identification of several odorous components has been achieved
but quantitative assessment has yet to be perfected. The A.D. Little
Odormeter may advance technology in this area by a significant step.
Gaseous Emissions
In general, turbocharging increases NOx further by increasing the
charge temperature but it allows further retard for the same smoke
limit.
Exhaust gas recirculation is effective in reducing NOx levels (particularly
over the CVS-CH cycle) but durability has yet to be proved and it does
tend to increase smoke emissions.
Although water injection has the benefit of reducing NOx without
significantly affecting engine performance, the logistics of the installation
and the problems of engine durability make this measure unattractive.
Timing retard is undoubtedly the most effective single parameter for
the reduction of NOx and the fact that the smoke limited performance
of the DDI engine tends to deteriorate less with retard than the DI
gives it a major advantage in this field.
The limited data available indicates that emission levels from 2-stroke
engines should be of the same order as from 4-stroke engines of
similar performance.
Heavy duty experience leads to the conclusion that the use of a
conventional direct injection chamber will increase both NOx and CO
levels while HC levels might rise rapidly with retarded timings. It
seems almost certain that a high speed (67 rev/s) (4000 rpm)
conventional, naturally aspirated direct injection engine would not
achieve the primary emission levels due to its low smoke limited
performance at retarded timings.
For a naturally aspirated 4-stroke indirect injection engine it can be
predicted that 3.4 g/mile CO can be achieved; 0.41 g/mile HC could
be attained on prototype vehicles although this figure may not be
held in production ; and 1.5 g/mile NOx could just be obtained from
a prototype current generation engine although some exhaust gas
recirculation may be necessary to allow a margin for production
compliance.
-------
4-52
Although 0.4 g/mile NOx has been achieved with a highly modified
prototype engine in a European type vehicle, it is extremely unlikely
that this figure would be achieved with a heavier vehicle and with a
higher power to weight ratio.
Any diesel powered vehicle would have less difficulty in achieving
the target objectives if both weight and power to weight ratio were
reduced. A lighter, lower powered vehicle would also have improved
fuel economy.
Particulates
The early suggested Californian requirement would present a major
problem for all engines with a heterogeneous combustion system and
this is undoubtedly a problem area for the diesel engine. However,
the true effect of particulates on health is unknown at the moment
and the problem could be solved by the addition of filter systems
although this move would carry a high cost penalty.
It is considered that work should be initiated into the distribution of
particulates from different types of engines and their true health
hazard determined before any legislation is finalised. An understanding
of their formation within the engine might also be a useful tool for their
control and thus a fundamental investigation using experimental and
analytical techniques should also be started.
Noise
The drive-by noise levels of diesel powered vehicles are slightly higher
than those of gasoline powered vehicles but there is no reason why
light duty vehicles should not meet proposed noise legislation.
The idle noise is annoying to the by-stander as well as the driver with
present European vehicles.
There is no reason why diesel powered light duty vehicles should be
unacceptable to either the driver or by-stander if sufficient attention
is paid to details of the engine and vehicle construction.
Volume
For the same power output the diesel engine is likely to be greater in
volume than a highly rated gasoline engine, but the increase in volume
is unlikely to pose any major problems.
-------
2-3
Fig. 4-15 Vehicle Fuel Consumption versus Inertia Weight
during LA4 1975 (CVS-CH) Test Cycle
4-16 Tolerances Controlling Piston to Cylinder Head Clearance
4-17 Effect of Inlet Valve Timing on Low Speed Torque
4-18 Torque Comparison on a 300.HP D.I. Diesel Engine
4-19 Heat Rejected to Coolant for Similar Diesel and
Gasoline Engines over the Load and Speed Range
SECTION 5
5-1 Estimated Torque Curve for 0 97 x 76 mm V-8
Gasoline Engine in Low Emissions Build
5-2 Estimated Fuel Consumption Curves for a097x76mm V-8
Gasoline Engine in Low Emissions (1.5 g/mile NOX) build
5-3 Estimated Torque Curve for a 088 x 82 mm 6 Cylinder
"European Type" Gasoline Engine in Low Emissions Build
5-4 Estimated Load Range Consumption Curves for a 6 Cylinder
088 x 82 mm "European Type" Gasoline Engine in 1.5 g/mile
NOX build
5-5 Estimated Performance Curve for Naturally Aspirated
088 x 98 mm V-8 Comet V Engine
5-6 Load Range Fuel Consumption Curves for Naturally Aspirated
088 x 98 mm V-8 Comet V Engine
5-7 N/A - V-8 - I.D.I. Diesel Engine Installation Drawing
5-8 N/A - V-8 - I.D.I.'Diesel Engine
Comparison of Crossflow/Unisided Cylinder Heads
5-9 N/A - V-8 - I.D.I. Diesel Engine Cross-Sectional Arrangement
5-10 Estimated Performance Curve for Boosted Six Cylinder
090 x 100 mm I.D.I. Engine in Low Emissions Build
(1.5 g/mile NOX in a 3500 Ib Passenger Car
5-11 Estimated Load Range Fuel Consumption Curves for a
Boosted Six Cylinder 090 x 100 mm I.D.I. Engine in
Low Emissions Build (1.5 g/mile NOx) in a 3500 Ib
Passenger Car
-------
2-4
Fig. 5-12 96 kW 6 Cylinder Turbocharged I.D.I. Diesel Engine
Installation Drawing showing Comparison between
In Line 6 / V-6.
5-13 6 Cylinder Turbocharged I.D.I. Diesel Engine
Cross and Longitudinal Arrangement Drawing
5-14 Preliminary Layout of V-6 60 Bank Angle Turbocharged
Diesel Engine with 6 Throw Crankshaft
5-15 Preliminary Layout of V-6 Turbocharged Diesel Engine
with 3 Throw Crankshaft
5-16 V-6 Turbocharged I.D.I. Diesel Engine
Cross and Longitudinal Arrangement Drawings
5-17 Estimated Performance Curves for Boosted 093 x 94 mm
Six Cylinder D.I. in Minimum Emissions Build
5-18 Estimated Load Range Fuel Consumption Curves for a
Boosted 093x94mm Six Cylinder D.I. Engine
in Minimum Emissions Build
5-19 96 kW Six Cylinder Turbocharged D.I. Diesel Engine
Installation Drawing showing Comparison between In-Line6/V-6
5-20 6 Cylinder Turbocharged D.I. Diesel Engine
Cross-Sectional Arrangement Drawing
5-21 Estimated Performance Curve for 2 Stroke Loop Scavenge
I.D.I. Diesel Engine
5-22 96 kW Loop Scavenge - I.D.I. 2 Stroke Diesel Engine
Installation Drawing
5-23 V-6 - I.D.I. - 2 Stroke Loop Scavenge Diesel Engine
Cross-Sectional Arrangement Drawing
5-24 Estimated Performance Curve for 2 Stroke Through
Scavenge D.I. Diesel Engine
5-25 96 kW 2 Stroke Through Scavenge Diesel Engine
Installation Drawing
5-26 6 Cylinder 2 Stroke Through Scavenged D*!. Diesel Engine
Cross and Longitudinal Arrangement Drawing
-------
2-5
Fig. 5-27 V-6 2 Stroke Through Scavenged D.I. Diesel Engine
Cross and Longitudinal Arrangement Drawing
5-28 Estimated Performance Curve for Four Stroke
093 x 93 Compound D.I. Engine
5-29 96 kW Compound In-Line 4 Cylinder D.I. Diesel Engine
Installation Drawing
5-30 Compound In-Line 4 Cylinder D.I. Diesel Engine
Cross and Longitudinal Arrangement Drawing
5-31 Estimated Performance Curve for Two Stage
Rotary Diesel Engine
5-32 96 kW 2 Stage/2 Bank I.D.I. Rotary Diesel Engine
Installation Drawing
5-33 2 Stage/2 Bank I.D.I. Rotary Diesel Engine
Preliminary Cross Sectional Arrangement Drawing
5-34 2 Stage/2 Bank I.D.I. Rotary Diesel Engine
Preliminary Longitudinal Arrangement Drawing
5-35 Naturally Aspirated 96 kW V-8 I.D.I. Engine
Predicted HC ppm Levels in Low Emissions Build
5-36 Naturally Aspirated 96 kW V-8 I.D.I. Engine
Predicted NOX ppm Levels in Low Emission Build
5-37 Naturally Aspirated 96 kW V-8 I.D.I. Engine
Predicted HC ppm Levels using Current Fuel
Injection Equipment
5-38 Naturally Aspirated 96 kW V-8 I.D.I. Engine
Predicted NOX ppm Levels in Optimum Performance Build
5-39 Typical HC Emissions from a Lightly Boosted Conventional
D.I. Engine in Optimum Performance Build
5-40 Typical NOX Emissions from a Lightly Boosted Conventional
D.I. Engine in Optimum Performance Build
5-41 Typical HC Emissions from a Lightly Boosted Comet V
Engine in Optimum Performance Build
5-42 Typical NOX Emissions from a Lightly Boosted Comet V
Engine in Optimum Performance Build
SECTION 6
Fig. 6-1 Light Duty Vehicle Power Plant Survey Results
-------
3-1
SECTION 3
GENERAL SUMMARY AND INTRODUCTION
This report summarises the results of a study which was conducted
on behalf of the Office of Mobile Source Air Pollution Control of the
Environmental Protection Agency and which was aimed at identifying
the problem areas and estimating the effect of the wide use of the
diesel engine as a light duty power plant in the United States.
The diesel engine is widely used in Europe for light duty vehicles and
for taxi cab services, roles in which its well known advantages in
respect of fuel economy and emissions are of great importance, and
in which the off-setting problems of weight, bulk, noise, cost and
possibly odour are accepted. The prospect of the high costs, greater
weight, increased bulk, and lower specific power of an emissions
controlled gasoline engine may reduce the impact of some of these
problems while the undoubted fuel economy advantage of the diesel
engine may be accentuated in an emissions controlled United States
environment.
Since there is little United States experience of light duty diesel
engines, the first stage of the study entailed the collection of related
data and a preliminary account of the feasibility of the diesel as a
light duty power plant. This first stage involved a study of published
and unpublished literature from technical journals and learned societies
throughout the world. Visits were also made to existing diesel users
and manufacturers throughout Europe so that operating and service
factors could be estimated realistically. A careful assessment of the
literature and visits indicated that a diesel engined passenger car should
be able to meet the proposed emission targets of 0.41 g/mile HC,
1.5 g/mile NOx and 3.4 g/mile CO although the secondary target of
0.4 g/mile NOX did not appear likely. It appeared that the main problems
of the diesel engine might be odour, particulates, noise, bulk, cost and
starting. In the areas of reliability, driveability and economy there seemed
to be considerable advantages to the diesel engine.
The next stage of the study involved the preliminary design of potentially
viable diesel power plants. These power plants had to be capable of propellint
a 1600 kg (3500 Ib) car with an acceleration potential equal to that of a
'standard' American car. (Essentially they were 97 kW (130 bhp) engines).
-------
3-2
The concepts studied included reciprocating and rotary diesel type engines
with and without turbocharging, compounding and low heat loss components.
Two gasoline power plants were also outlined for comparison purposes.
The configuration study covered the following contenders for the power
plant :-
1. 8-cylinder 'American' gasoline
2. 6-cylinder 'European1 gasoline
3. 8-cylinder IDI diesel 4-stroke
4. 6-cylinder IDI diesel 4-stroke turbocharged
5. 6-cylinder IDI diesel 4-stroke with 'Comprex1
6» 6-cylinder DI diesel 4-stroke turbocharged
7. 6-cylinder DI diesel 4-stroke with 'Comprex1
8. 6-cylinder IDI diesel 2-stroke loop scavenged
9. 6-cylinder DI diesel 2-stroke uniflow
10. 4-cylinder compound diesel 4-stroke
11. 2-stage, 2-bank rotary diesel
Drawings were prepared for all the diesel power plants and performance
curves and emissions levels were calculated.
The third stage of the study was an assessment of the relative merits
and viability of the power plants for both the primary and secondary
targets.
In order to compare the various power plants a quantitative rating
method was devised. This involved an estimate of the relative
importance of 26 performance aspects and assigning to them
numerical 'weightings'. Assessment of the merit with which each power
plant met an aspect allowed a numerical 'rating' to be assigned.
Summation of the weighted ratings allowed the power plants to be
compared numerically.
Use of this rating system on the power plants indicated that the gasoline
engines were slightly superior to the diesel engines for the primary
emissions targets and that only the 4-stroke IDI diesels were really
viable for the duty considered.
None of the diesel engines could be considered as being practical
power plants for the secondary emission targets.
The fourth and final stage of the study was an assessment of the
implications of the above results so that proposals for further
work and research could be made if it seemed desirable.
-------
3-3
It was concluded that :-
a) In view of the large extrapolations required
from current European practice an 8-cylinder
IDI diesel 4-stroke engine should be built and
demonstrated.
b) A 6-cylinder IDI diesel should be built and
work on turbochargers and 'Comprex1
pressure exchangers should be carried out.
c) Work should be pursued on advanced and
potentially cheaper fuel injection systems.
d) The low compression ratio, ignition assisted
IDI engine should be investigated thoroughly.
e) A fundamental investigation into the formation
of hydrocarbons in diesel engines should be
carried out together with an attempt to produce
catalysts with a lower light-off temperature.
f) A control system for the modulation of exhaust
gas recirculation should be developed.
g) A quick warm up starting aid should be developed.
h) Diesel engines should be demonstrated in taxi cab
and light delivery truck service to encourage a
rapid switch to diesel power for these duties.
i) A study into particulate formation in diesel exhaust
together with the development of soot and particulate
filters is desirable.
-------
3-4
Introduction
There is little doubt that the low emissions gasoline engine offers
the prospect of higher cost, lower specific power and high specific
bulk and weight while the increasing cost of fuel renders higher
efficient alternative power plants more attractive.
The diesel engine is demonstrably the most efficient practical
prime-mover produced today and it is known to have favourable
emissions characteristics. The most efficient diesel engines
are not used in light duty vehicles because of their bulk and
limited speed range while those diesel engines currently used
in light duty vehicles are smaller than would be required for an
American passenger car and have not yet been developed to meet
very stringent emissions targets*
In Europe and Japan small high speed diesel engines are used widely
for light duty vehicles and taxi cab service while in some countries
there is also a minority of diesel powered passenger cars. Thus
there is already a great mass of experience of light duty diesel
vehicles in most roles.
The present American environment, however, cannot be related
directly to European and Japanese experience. The diesel engined
passenger cars used in Europe and Japan have been of smaller size
than typical American cars and have also had lower power to weight
ratios. Current and proposed American emission standards are
different from those obtaining in the rest of the world so that there
has not, as yet, been any attempt to run existing light duty diesel
engines in service in low emissions build.
Consideration of the diesel engine for an American passenger car
immediately reveals that it will be considerably larger and will
spend most of its life running under different conditions than those
currently found elsewhere. It was these considerations which led to
the objectives of this study, i.e. an examination of the feasibility
of using the diesel engine as a power plant for an American passenger
car, an examination of potential problem areas and a study of the
trade-offs possible within these areas. The determination of the
possibilities of finding solutions to any outstanding technical problems,
and finally to put forward recommendations for the future increase in
light duty diesel engines (if any be thought desirable) and to indicate
the best routes to achieve this increase.
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4-1
SECTION 4
LITERATURE SURVEY, ASSESSMENT OF
PROBLEM AREAS AND TRADE OFFS
This section contains a digest of the articles considered in a
survey of light, duty (and appropriate heavy duty) diesel
literature and the conclusions obtained from visits to manufact-
urers and users of light duty diesel engines in Europe. The
published literature covered the period from 1919 to the present
although most of the information came from the period from
1950 to the present. Where possible and pertinent the published
information was augmented by technical data from Ricardo in-
house reports»
Study of the literature, which was conducted under the headings
of 'Performance Aspects', allowed problem areas and potential
trade-offs to be identified as well as indicating some of the
possible gains from the use of light duty diesel engines in America.
The major conclusions were that there were no major problem
areas in the use of light duty diesel engines in America. Smoke,
odour, participates, noise, volume, weight, cost, starting and
refinement would be inferior to existing gasoline engines but there
would be advantages in respect of emissions, fuel economy,
maintenance, driveability, durability and overall economics.
Emissions predictions indicated that the diesel engined American
passenger car might just achieve the primary emissions targets
of HC - Oo41 g/miie. CO » 3o4 g/mile, NOx - 1.5 g/mile with-
out catalysts, but that the secondary emissions targets of 0.4
g/mile NOx would be practically impossible to achieve.
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4-2
Introduction
Although diesel engined light duty vehicles and taxi cars are
widely used in Europe and Japan, all the current service
experience undoubtedly relates to engines which are smaller and
which run under different load conditions from those which would
be applicable to present and proposed American light duty con-
ditions. Current service experience moreover is with a gener-
ation of engines which were not emissions biased although much
work has been carried out recently to determine the emissions
characteristics of these engines.
It was these conditions which indicated that a study of published
literature, Ricardo in-house reports and data, and discussion with
current light duty diesel manufacturers and users would yield
much useful information on the likely feasibility of the diesel
engine for American light duty use and which would also indicate
the problem areas and potential trade-offs if diesel engines were
adopted for American use.
Although defined in the configuration study, the goals of the com-
plete study had to be kept in mind throughout this literature study.
The vehicle and application considered was a 1600 kg (3500 Ib)
4/5 seat sedan capable of achieving 0-97 km/h (0-60 mph)
in 13.5 s, 32 - 113 km/h (20 - 70 mph) in 15 s, and the D.O.T.
high speed pass manoeuvre in 15 s so that an installed power of
about 96 kW (128 bhp) was required.
The emissions targets were that the vehicle should achieve the
following figures measured by the CVS-CH test procedure.
a) Primary Targets
HC 0.41 g/mile
CO 3.4 g/mile
NOx 1.5 g/mile
b) Secondary Targets
as above but
NOx 0.4 g/mile
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4-3
Scope of the Survey
Since its first application, much has been written about the diesel
engine and a considerable proportion of the published work has
dealt with light duty engines. For the survey the period 1919 to
the present day was selected as appropriate although the majority
of the information was uncovered for the period 1950 to the pre-
sent. For this period all published and unpublished reports and
translations were checked and Figs. 4-1 and 4-2 show the scope
of the survey. For convenience in analysis the light duty diesel
engine was considered under various performance aspect para-
meters, as shown in Table I below:-
TABLE I
Performance Aspects
1. Smoke
2. Odour
3. Gaseous Emissions
4. Particulates
5. Noise
6. Volume
7. Weight
8. Fuel Economy
9. Fuel
10. First Cost
11. Maintenance
12. Starting
13. Hot Driveability
14. Cold Driveability
15. Torque Rise
16. Durability
17. Coolant Heat Losses
18. Fire Risk
19. Vibration & Torque Recoil
20. Idle Noise
21. Reliability
22. Economics
23. Manufacture
24. Performance
25. Design
26. Ancillaries
27. Lubrication
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4-4
In order to assist in the identification of major conclusions and
problem areas, a keyword system was used to classify the items
of information uncovered in the literature search and this keyword
system, a three level one, is described in Appendix 1 (section 8).
The literature search covered some 700 items and although it is
not possible to refer to them all in detail, a complete list may be
found in Appendix 2 (section 9).
In addition to the literature study, several European companies
involved with the day to day operation of diesel powered light duty
vehicles were contacted to obtain a unbiased opinion of the true
operational characteristics of diesel passenger cars in current
European conditions. The companies contacted are listed in Fig.
4-3.
Exhaust Smoke
Any suggestion to make wider use of the diesel engine may come
under public attack because of the diesel's anti-social reputation in
this respect. It is important to realise however that black smoke
is not an inherent characteristic of a diesel engine but only of one
that is either ill-maintained or is overloaded.
Apart from Thermal or Mechanical Loading limits, which are a
function of engine design, it is the onset of excessive black smoke
in the exhaust which limits diesel engine output. Smoke limitations
are therefore output limitations.
Black Smoke
Exhaust smoke may be of one of two forms, either black or blue/
white. Black smoke occurs when there is insufficient oxygen for
complete combustion of the fuel and since in the diesel engine com-
bustion is heterogeneous, it can occur even though there is an
overall excess of oxygen in the charge.
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4-5
When a fuel drop is heated in a combustion chamber, complete
oxidation of the carbon and hydrogen occurs provided there is a
plentiful supply of oxygen and provided further that there is time
for combustion to be completed before the reactions are quenched
by expansion of the gases. If there is a shortage of oxygen however,
the hydrogen preferentially takes it quota, leaving too little oxygen
for complete combustion and under this condition carbon particles
are formed which appear as black smoke in the exhaust (19).
The exact chemical reactions leading to the formation of smoke are
however unknown although they may involve €2 or €3 fragments of
hydrocarbon decomposition as a basic building brick leading to a
final polybenzenoid structure for the final smoke particle. It may
even be that there are a number of alternative paths leading to such
formation.
Soot production takes place during the early part of the combustion
process of all diesel engines but under conditions in which the
exhaust smoke emission is low, the soot is consumed during the
later parts (9).
The formation of soot particles is a characteristic of diffusion flames
such as are typical of combustion in a diesel engine and photographs
of such combustion exhibit intense white or yellow radiation of
incandescent particles under all engine conditions. There is a
dramatic difference in this respect from premixed combustion in a
gasoline engine, which exhibits a much lower degree of luminosity.
At high fuel inputs, with a consequent smoky exhaust, the radiation
persists until cooling takes place during expansion which leads to
quenching of the combustion products and the emission of black
exhaust smoke (38).
Blue/White Smoke
The smoke emission from diesel engines does not consist uniquely of
black smoke but at times may include white or blue/white smoke
resulting from misfire, the colour of the smoke depending on the
size of the droplets.
This smoke is produced by fuel droplets in abundant oxygen which do
not reach ignition temperature and which in a partially vaporised form
-------
4-6
pass out of the exhaust as a cloud. The problem can occur during
the winter when the engine is first started after a long soak at low
temperature and on some engines the smoke can persist for a
considerable time after start-up. The most general solution is an
increase in compression ratio but at very low temperatures even
this may not be a complete solution (19). Other solutions are the
use of a higher Cetane Number fuel or by decreasing the mid boiling
point of the fuel (38) .
Blue smoke can also result from excessive lubricating oil consumption
(573) and while this may be the result of wear due to high mileage,
problems may arise early in the engine life due to unsatisfactory
running in of the piston rings.
Combustion Chamber Type and Black Smoke
Diesel engine combustion chambers may be classified into one of three
types. The first of these, and the one most extensively used for truck
and larger engines, employs direct injection, where the fuel is
injected directly into the space above the piston. Prechamber engines
employ injection into a flask shaped chamber into which part of the
air charge has been compressed, with a restriction between the pre-
chamber and the main chamber. Swirl chamber engines are a special
form of prechamber engine in which the entrance to the prechamber is
so arranged as to give an intense swirling motion to the air charge in
the prechamber; a larger portion of the air charge is compressed into
the swirl chamber than would be the case with a prechamber, 50% as
against 25% or so, and the restriction between the swirl and main
chamber is less pronounced than would be the case in a prechamber
engine.
The three types of combustion chamber have different smoke character-
istics. Direct injection engines have a relatively slow onset of smoke
as load is increased and furthermore have a lower effective air
utilisation. As a result, with low smoke limits, the direct injection
engine must be more severely derated than a swirl chamber engine
(Fig. 4-4). Prechamber engines on the other hand, while they have
similar smoke levels to swirl chamber engines at full load, (Fig.4-5)
(20,695) can give plateau' smoke as the load is reduced, i.e. the
smoke does not fall as low as it does on a swirl chamber or with
direct injection (16, 18, 20).
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4-7
Fuel/air mixing is a vital factor affecting smoke production. On direct
injection engines, both increased rate of injection and advance in
injection timing reduce the amount of exhaust smoke (110) . Air swirl
rate is also important and it is essential to optimise swirl and rate of
injection (Fig. 4-6).
Indirect injection engines are less sensitive to injection rate but have
a reverse characteristic in so far as injection timing is concerned,
retardation of timing giving reduced smoke output. The insensitivity
of injection rate is presumably due to the mixing being controlled by
high air velocities rather than by jet mixing.
Fuel Type and Black Smoke
While prechamber engines tend to be less exacting in their fuel require-
ments than direct injection engines, increasing fuel volatility in all
engines helps to reduce black (and indeed white) smoke. While, as
has already been mentioned, increasing the Cetane Number helps
reduce white smoke, it should not be increased too far or increased
black smoke results (294).
Improvements in exhaust smoke can be obtained by fumigating, all (543)
or part (568) of the fuel with the induction air. Ricardo results- siagjgest
however that this can lead to blue smoke in the exhaust at part load
unless the fumigated fuel be cut off under these condition®.
One minor but interesting smoke problem was reported! during: tfee
Vehicle User Survey by the British Petroleum Company, who mentioned
that vehicles with fuel pump maximum fuel stop settings set to- eomply
with smoke requirements on continental European fuel with a relative
density of 0.825 - 0.835, had a smoke problem in the United Kingdom
where the average relative density of fuel is 0.84. This could well
cause difficulties in other locations.
Dual fuel engines with gaseous main fuel's, ignited by injecting a small
charge of diesel fuel, can also give a cleaner exhaust (2l) (61) r but
emission tests on such dual fuel engines show high levels of hydrocarbons
in the exhaust.
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4-8
Anti-Smoke Fuel Additives and Hang-on Devices
There are numerous references to tests with alkaline (usually barium)
based fuel additives. All agree that large reductions in smoke levels
are obtained but most long term tests show that combustion chamber
deposits are increased by the use of these additives (7, 9, 76, 109,
292, 293, 294, 80), since the resulting fuels have a higher ash content
than untreated fuel (19) . There is also concern as to possible health
hazards with metallic additives in the fuel.
It is clear that anti-smoke additives do affect the combustion process
and do not merely act as soot dispersants since there is a reduction
of mass soot emission as well as a reduction in invisible smoke (9).
Tests with exhaust catalysts have resulted in slight changes in exhaust
conditions, there being a small improvement in the smoke levels at
the full load/speed condition in one test engine (107) . Very little
work with such devices has however been reported in the literature.
Smoke in Turbocharged Engines
With the levels of turbocharging likely to be applied to a passenger car
engine, the turbocharger will be matched with the engine at the peak
torque condition. Hence with any acceptable torque back-up curve
the exhaust will be clean under any steady state condition at this and
higher speeds. At lower speeds however the torque will be limited
by the onset of exhaust smoke and in addition, with turbocharging'
there could be a transient smoke problem over most of the speed range
unless a boost sensor be fitted to the maximum fuel stop.
In this area, the Comprex may have decided advantages over
conventional turbochargers but more experience is necessary before it
is possible to guarantee that smoke would be eliminated under starting
and transient conditions.
Smoke Measurement
Smoke measurement is not the simple problem which it appears to be
at first sight as may be seen from the large number of attempts that
have been made to design and develop smoke meters. Only a few have
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4-9
survived as practical instruments and these may be classified as
follows: -
a. Sampling light-obscuration meters (Volvo, Hartridge,
DDR.RD M.4)
b. Sampling filter soiling meters (Bosch, Von Brand)
c. Full flow obscuration meters (UTAC, USPHS, Beckman,
Celesco)
d. Sampling, ligh-dispersant meters (Bosch, Tyndall)
As has been shown, smoke emission from diesels does not consist
uniquely of black carbon particles but at times includes white smoke.
Light obscuration meters indicate such white smoke as increasing
the apparent density whereas filter soiling meters, where the soiling
is measured by light reflection, indicate reduced density. Thus the
direct correlation betweeen these meters, which should in theory be
possible with black carbon particles, breaks down.
Smoke Legislation
Current United States smoke legislation for truck diesel engines
limits the average visible smoke from the vehicle to 20% opacity
during acceleration and 15% during the lug mode of the Federal smoke
cycle with a 50% limit to the peak cycle. While these levels may be
acceptable for a limited number of trucks, many of which spend the
greater part of their time on highways and not in cities, Ricardo think
that such levels from diesel engined light duty vehicles would be
unacceptable and would suggest a level of approximately 5-7% opacity
at the rated speed and 8-10% at peak torque speed.
These proposed levels are in fact in line with those laid down in Britain
for this size of engine, and it is Ricardo's belief that current European
diesel passenger cars could comply with these limits.
Typical smoke legislation for countries other than the United States is
as follows: -
In Britain control of diesel engine smoke in road vehicles is covered
by British Standard AU 141a:1971. This requires a 100 hour type test
-------
4-10
for each basic engine type for certification. The AU 141a smoke
limit is in the form of a curve which defines the smoke limit at
different levels according to the engine output; the higher the
output the lower the smoke limit. The limits apply to speeds at
full load from 100% down to 45% of the maximum speed, or 16.7 rev/s
(1000 rev/min) whichever is the higher, smoke density being mea-
sured by a light obscuration method.
The European Economic Community (EEC) have produced a smoke
test code for diesel engined road vehicles (EEC test code 72/306/EEC).
This includes steady state testing over the same road/speed range as
the British Standard and with limits similar to but slightly higher than
those of BS AU 141a. In addition a free acceleration test procedure
is included with a smoke limit set slightly above the limit of the steady
state tests, so that the transient response can be checked and also
road-side tests carried out on vehicles. In this test code, smoke
measurement is by opacity smoke meter.
Czechoslovakia
The test is a filter soiling type using meter type NC 112, allowing 40%
discoloration of the filter in new vehicles and 50% discoloration in other
vehicles.
Finland
A full load acceleration test, in gear, is specified using a filter soiling
smoke meter.
E. Germany
Steady state and free acceleration tests are under investigation using
RD M-4 smokemeters.
Japan
A steady state, full load test at 40, 60 and 100% rated speed with smoke
measurement by filter soiling smoke meter and a limit of 50% discolor-
ation at each test condition.
Jugoslavia
Exhaust density must not exceed No. 2 Ringelmann.
Mexi co
Smoke emission not to exceed No. 2 Ringelmann for more than 10 seconds.
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4-11
South Africa
A free acceleration test with Hartridge meter is specified, with a
smoke limit of 70 HSU (Hartridge Smoke Units).
Spain
Free acceleration limit of 65 HSU operates from 1970 in Madrid
only.
Sweden
The test involves a loaded vehicle in gear, limiting the maximum
speed to 45 - 50 km/h. Full load acceleration is tested using either
Bosch or Hartridge meters. Limits for passenger vehicles for more
than 30 persons are 30 HSU or 2.5 Bosch. The limit for other vehicles
is 45 HSU or 3.5 Bosch.
Switzerland
Full load steady state using Bosch meter and free acceleration test
using Bosch integrating smokemeter. For steady state test, limits
range from 6 Bosch for up to 3 litre engines to 4.5 Bosch above 8 litres.
This applies up to 600 m altitude, 0.5 Bosch being allowed additionally
for each 400 m above this.
Odour
Exhaust odour, like smoke, is a social problem of the diesel engine
but is much more difficult to quantify. Nevertheless, a considerable
amount of work has been carried out in this area, mainly using
subjective testing with 'sniff panels' but also including chromatography
in association with subjective testing.
Early work indicated that exhaust odour was related to aldehyde concen-
tration but that it was dependent on the sulphur4 content of the fuel (6).
Later work by A.D. Little has suggested however that only two groups
of components are important in so far as diesel odour is concerned.
One group is composed of aromatics such as alkyl benzenes, indenes,
indans, tetralins and naphthalenes. The other group of components
is composed of partial oxidation products of components found in the
aromatic fraction of the diesel fuel. This work also suggests that the
sulphur content of the fuel is not important (14).
Peak odour intensities have been found to occur at no load and full load
while engine speed is not an important parameter (652). Comparisons
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4-12
between two stroke and four stroke diesel engines gave much higher
odour levels from two stroke engines. Later model two stroke engines
with modified low sac volume fuel injectors had however similar
odour characteristics to equivalent four stroke engines (653) .
The high odour levels at no load, which can be accentuated at high
speeds by misfire, can be minimised by advancing the injection timing
at light loads to prevent misfire conditions. Steps taken to reduce
smoke at full load by limiting the engine load will similarly reduce
full load odour.
Major changes in fuel constitution, e.g. the use of a fuel with very
low aromatic content, have been reported to give changes in odour
characteristics. The use of such a fuel also gives a lower intensity
of odour as compared with a normal No. 1 or No. 2 diesel fuel (655) .
It has been stated that the levels of exhaust odour found in a Mercedes
diesel car are not low enough to prevent a significant amount of
objectionability (654), but so far as Ricardo are aware,
exhaust odour from light duty engines has not been a serious cause
of complaint in Europe even in taxi service where the density of such
vehicles can be quite high in congested urban areas. While it is very
difficult to estimate the likely effect of a large increase in the
density of such diesel vehicles, it would seem reasonable to assume
that densities in American cities up to those found in Europe should
not cause offence.
Experimental work has been carried out with catalytic exhaust mufflers
(653). These have been shown to reduce odour under idle and acceleration
conditions but have no effect on other loads. Such mufflers are believed
to have been used on some bus fleets.
As diesel fuel does not readily evaporate, fuel odours can be detected in
diesel engined vehicles as a result of bad housekeeping. It is essential
to make sure that there are no leaks from the fuel system and there
is no spillage when refilling the fuel tank. With these precautions such
fuel odours should not be detectable.
Gaseous Emissions
The three major gaseous pollutants are NOx, CO and HC. NOx is largely
formed as NO but also contains small amounts of NO2- Only the NO is
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4-13
measured if the exhaust is metred by means of a non-dispersive
infra-red analyser „ HC will contain a very wide spectrum of
hydrocarbons of all types, some of which will be of similar form
to those found in the raw fuel while others will have partially
oxidised or otherwise changed in the engine. They arise from
incomplete combustion of the fuel.
The oxides of nitrogen are formed from the constituents in the air
although it has been suggested that dissolved nitrogen in the fuel
can play an important part. The oxides of nitrogen are formed
under high temperature and pressure conditions during the
combustion part of the cycle. The rate of formation is highly
temperature dependent and anything that can be done to reduce the
peak cycle temperatures will reduce NOx levels. Unfortunately
it is difficult to do this without loss of efficiency since this is also
a function of peak cycle temperature.
The carbon monoxide levels produced by the diesel are very low, at
least by gasoline engine standards, since the flame conditions do not
favour this product of partial combustion under normal operating
conditions. It is only when excessive fuel is injected and the
controlled pattern of mixing and combustion breaks down that carbon
monoxide emissions rise to significant levels. This can occur, although
normally at fuels air ratios above the limits set by legislative smoke
requirements ° Thus within these smoke limits, control of CO is not
a problem and for example CO levels from a well developed pre-
chamber engine are less than 5% of those emitted by non-emission
controlled gasoline engines.
As has been explained, the hydrocarbons range from pure fuel through
to complex aldehydes, oxygenated hydrocarbons etc., and at very high
fuel:air ratios can include 3.4 benzpyrene, recently suspected as
being carcinogenic. HC emissions can show apparently random
variations across the engine operating range, mainly due to the
extreme complexity of the reactions involved in chemical-kinetic terms.
Sources are possibly the same locally over-rich flame zones responsible
for soot formation and in some cases weak mixture zones in which the
flame is unstable. Late admission of fuel is known to increase the
levels significantly since the falling chamber temperature effectively
freezes all reactions before combustion is completed.
Whereas NOx formation is the result of efficient, high temperature
combustion (and is therefore difficult to reduce without seriously
affecting engine efficiency) , both CO and HC are the results of marginal
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4-14
combustion conditions either throughout or within local pockets
(sometimes on a molecular scale) of the chamber.
The diesel engine emits much larger quantities of one other gaseous
pollutant than does the gasoline engine. This is SC>2 which arises
from the sulphur in the fuel. Current diesel fuels contain up to
0.5% of sulphur while gasoline contains only about 0.05%. It is not
possible to reduce the SC>2 levels by engine modifications, and while
it would be technically possible to remove the sulphur during the
refining process, the cost is likely to be high.
Effect of Combustion Chamber Type on Emissions
It has been recognised for many years that, when optimised for
performance, the direct injection engine emits more HC, CO and
NOx than does the indirect injection engine. As a number of authors
have pointed out however, and as may be seen from Figure 4-6, the
direct injection engine has a much steeper response of NOx reduction
with retard of ignition timing, and at very retarded timing there is
little to chose between the two types in so far as NOx levels are
concerned. These degrees of retard of timing introduce excessive
hydrocarbon levels however, and with naturally aspirated direct
injection engines, the high smoke levels would involve excessive
derating to maintain an acceptable exhaust smoke level»
Pischinger I 21) selects the direct injection engine as a potential low
emissions power plant because of the steep NOx/injection timing
response and claims to prevent the onset of smoke by induction port
and combustion chamber design modifications. As mentioned above
however, in Ricardo's view the high hydrocarbon levels likely to
result from the required degree of retardation will present a problem.
EPA tests on vehicles with prechamber (Mercedes 220D) and swirl
chamber (Peugeot 504, Opel 2100D, Nissan Datsun 220c) engines have
indicated no significant differences in the emission characteristics
between these engine types although any engine may have high HC
levels due to injection irregularities.
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4-15
Effect of Engine Load on Emissions
As the load on most engines is reduced at constant speed, the specific
hydrocarbon emissions increase. Similarly on all indirect injection
engines and on many direct injection engines, the specific NOx levels
increase with reduction of load. As a result, for a given power output,
a larger, lower rated engine will have higher emission levels than
will a smaller higher rated one. With a given vehicle weight therefore
and driving over a fixed driving cycle, provided that all vehicles can
drive the cycle, the one with the lowest power/weight ratio, i.e. the
smallest engine, will have the lowest HC and NOx emissions. In
general terms however there will be little to choose between their
CO levels.
Turbocharging
In general, turbocharging without aftercooling gives somewhat higher
NOx levels than those obtained with naturally aspirated engines. It will
however clean up the smoke which otherwise would arise from retarded
injection timing. If the compression ratio is reduced to limit maximum
cylinder pressures, high HC levels may result at low load where the
turbocharger is giving no effective boost.
With turbocharging and aftercooling, as has been shown by Walder (16),
a reduction of NOx of approximately 2Q% is possible together with a
smaller improvement in CO levels. The HC levels are unaffected.
Gaseous Emissions from Two Cycle Engines
As the majority of engines used in truck service are four cycle, the
greater part of the data on emissions from diesel engines is from
four cycle engines. The very successful Detroit Diesel 71 and 53
series are however through scavenged two cycle engines and available
data indicates that these engines have similar levels of NOx and HC
emissions to four cycle engines. It may be inferred therefore that
both through scavenged two cycle and four cycle engines have similar
emissions levels.
So far as Ricardo know, there is no data available for the exhaust
emission levels from loop scavenged two cycle engines. With exhaust
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4-16
ports, oil control and hence HC levels may give problems and, with
less effective scavenging, smoke and hence CO levels may tend to be
higher than with through scavenging, unless severe derating be
employed. On the other hand, the internal exhaust gas recirculation
inherent in poor scavenging should give lower NOx levels although
there will be of course no cooling of the recycled gas.
NOx Reduction
The most effective way of reducing the quantity of NOx in the exhaust
is to retard the injection timing. On a direct injection engine however
retarding the timing leads to an increase in smoke levels and hence
for a constant smoke limit it is necessary to reduce the power output
of the engine. Some improvement is possible by retarding the timing
and increasing the rate of injection. This can lead to reduced levels
of oxides of nitrogen with less derating of the engine but normally
there is still some loss of efficiency.
On indirect injection engines, at least over the range of injection
timings of interest, smoke levels reduce as the timing is retarded,
and hence there is no need to derate the engine. There is still however
a loss of efficiency due to the retardation of timing.
Apart from injection timing retard, there are two other effective
methods of NOx reduction. One is by recirculation exhaust,
preferably after cooling it. and the other is by water injection either
with the induction air or as an emulsion with the fuel. Both methods
reduce the temperatures within the charge and EGR also reduces the
oxygen available to react with the nitrogen in the air.
Exhaust Gas Recirculation
The effectiveness of exhaust gas recirculation in reducing NOx has
already been proven beyond doubt by many authorities with both light
duty (647) and heavy duty(6, 16, 12) cycles but durability, particularly
that of the lubricating oil, has yet to be established. From data
available to Ricardo., there are indications that at high exhaust re-
circulation levels the wear rate of engine components may be higher
than without such recirculation. There is insufficient data however
to predict whether or not a problem will arise from such wear although
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4-17
oil thickening is likely to be a problem with a consequent need for
frequent oil changes.
By reducing the oxygen available for combustion, exhaust gas re-
circulation results in a need for derating to avoid excessive smoke.
Fig. 4 -7 demonstrates the smoke response of an EDI engine over the
load range with varying amounts of EGR. It will be noted that increasing
recycle rates significantly reduce the smoke limited performance of
the engine while the fuel consumption is unaffected. Assuming therefore
that the engine is rated near to its smoke limit without EGR, no
appreciable improvement in full load emission levels can be achieved
using recycled exhaust unless the engine is derated. As load is
reduced, increasing amounts of recycle gas may be introduced.
The scope for such modulated exhaust gas recirculation will vary from
engine type to engine type. With indirect injection, the NOx emissions
actually rise with reducing load for some way down the load range
and hence modulated EGR can be very effective. With direct injection
however, where the NOx levels fall quite rapidly with load, the effect
of modulation is less pronounced (Fig.4-8).
Turbocharged engines, with both direct and indirect injection, have
rapidly falling NOx concentrations (Fig. 4-9) with reduction in load but
normally have a considerable excess of air for combustion at full power
conditions. Hence there is little need to restrict exhaust gas recirculation
at full power. There may however be smoke problems at the maximum
torque speed as derating here will make the torque curve from the
engine less attractive.
It must be remarked however that when driving a light duty vehicle with
a high power/weight ratio over the CVS-CH test procedure, the engine
will not run at full power and modulation of exhaust gas recirculation
to reduce NOx at these conditions may not therefore be necessary.
Water Injection
Water injection has been found to be a powerful agent for reducing NOx
(6,16), but there are major problems in practice in applying this
method. The major advantage is that significant reductions in NOx can
be obtained with no loss in engine performance.
Problem areas are:-
-------
4-18
A water tank approaching the size of the fuel tank is
required, the actual size depending on the proportions
of water/fuel employed.
Antifreeze is necessary in winter and this may lead to
HC problems. When Ricardo experimented with alcohol
as an antifreeze the exhaust hydrocarbons were excessive.
To guarantee low emission levels it is necessary to
incorporate a mechanism to shut the engine down in case
of failure of the water injection system.
Cylinder liner corrosion and wear and/or lubricating oil
contamination with water may be a problem.
Some experimental work is being carried out with injection of water/
fuel oil emulsions. This may overcome some of the problems but
it will probably be very difficult to obtain acceptable stability of the
emulsion. Either partial or complete breaking of the emulsion would
be disastrous from a fuel handling and engine operating point of view.
Reduction of NOx by Reduced Engine Compression Ratio
It has been suggested that one way to reduce NOx levels would be to run
the diesel engine at a low compression ratio. In Ricardo's experience
however, the lowest overall levels of HC and NOx are obtained at high
compression ratios,due to the onset of misfire at light loads and
starting when using a low compression ratio. This results in high HC
emission levels. The use of more effective starting and ignition aids
has been proposed to overcome this but to the best of our knowledge
no-one has yet successfully demonstrated such a device.
Emissions under Light Duty Cycle (CVS-CH) Conditions
While, as has been pointed out earlier, the power/weight ratio of
current light duty diesel engined vehicles is much lower than is necessary
to satisfy the performance requirement for this survey, it is of greatest
interest to look at results from vehicles operating under the CVS-CH
cycle test procedure in order to predict emission levels for a diesel
engined passenger car.
-------
4-19
The literature contains a large number of references relating to the
effect of engine variables on diesel exhaust emissions when
operating over the 13 mode steady state test procedure. There are
unfortunately however only a very small number relating to light
duty CVS-CH cycle testing. Furthermore, in Ricardo's experience
there is no constant relationship between 13 mode and CVS-CH test
results so that it is not possible to read directly across from one
to the other. Despite this, there is in general terms a relationship
between the two types of testing in that parameters which tend to
reduce emissions in one test are very likely to do so in the other.
Although Ricardo have some data arising from tests which they have
carried out for their clients, very little additional CVS-CH data
has been found apart from EPA tests on standard European diesel
cars and one Mercedes 220D with modified injection equipment.
All these vehicles have indirect injection engines. The results of these
tests are summarised in Figure 4-10.
One other report on light duty emissions tests on diesel vehicles
covers comprehensive tests by Daimler Benz on an experimental
vehicle. After considerable development and using high rates of
exhaust gas recirculation the secondary target levels were just
achieved. Significant reductions in HC and CO levels were obtained
by the introduction of a catalyst into the exhaust system and HC
levels were reduced by modifying the fuel injection equipment (647).
All of the light duty vehicles tested are somewhat lighter than the
specified 3500 Ib car for this study. In addition, power levels are
much lower and in fact the vehicles are only normally just able to
drive the cycle.
To a first order of accuracy, the emissions in g/mile when driven
over a fixed cycle and with the same power/weight ratio for each
vehicle, will vary with the weight of the vehicle. Furthermore,
since specific NOx and HC levels on indirect injection and many
direct injection engines rise as load is reduced, an increase in
power/weight ratio will increase the emissions from the vehicle
when driven over a fixed cycle.
The best emission levels which have been obtained from current
generation diesel engined light duty vehicles are:-
HC 0.2 g/mile
CO 1.5 g/mile
NOx 1.5 g/mile
-------
4-20
HC levels of 0.21 g/mile have been achieved by EPA on a standard
Nissan Datsun 220c and Ricardo tests on other vehicles have
indicated that substantial reductions in HC levels can be achieved
by minor modifications to fuel injection equipment without altering
engine performance characteristics. Thus it is felt that the HC
emission levels of 0.3 - 0.4 g/mile achieved in the Mercedes and
Opel vehicles could be reduced to circa 0.2 g/mile.
As a result of applying a correction factor for vehicle weight and
increased engine size, predicted emission levels from a 130 bhp
3500 Ib inertia weight vehicle over CVS-CH cycle are:-
HC ' 0.4 g/mile
CO 3.0 g/mile
NOx 2.0 g/mile
Timing retard will reduce NOx levels to the order of 1.2 g/mile
without seriously affecting CO levels. Combustion noise levels in
this retard mode should be lower and vehicle driveability will be
unaffected. The fuel consumption will probably increase by about
10%, water jacket heat losses will remain constant but blue smoke
and misfire and hence HC levels, particularly when cold, may be
a problem. Some form of temperature sensor controlling injection
advance at light load may be necessary to control this.
Some further reduction in HC and CO levels may be possible by
optimisation of the gear ratios within the transmission. Some
preliminary tests by Ricardo on a diesel powered vehicle in which
gear ratios were altered showed substantial benefits could be
obtained by changing from a four speed to a three speed gear box.
The tests were in fact run with the standard four speed box but in the
second run first gear was not used, but all gear change points were
selected as if the gearbox were of standard three speed form. While
the NOx emissions were unaffected, the HC and CO levels were
reduced by approximately 50%. While as large a change as this might
affect the ability of the vehicle to give the target performance,
transmission ratio modifications must be regarded as a possible
method of reducing HC and CO levels from diesel vehicles.
The secondary project objectives, i.e. to obtain 0.4 g/mile NOx, are
very much more difficult to achieve. Daimler Benz have already
demonstrated a passenger car (647) which just attains this level in a
highly modified build, but with no room for production variations or
deterioration throughout the life of the vehicle. The level of 0.4 g/mile
was only achieved with 'increased1 EGR flow rates (exact value not known)
-------
4-21
and although the value of EGR in terms of effectiveness in reducing
NOx is not questioned, engine durability with high recycle rates
has yet to be proved., The practicality of achieving this level of
NOx with a production solution in a European type car is very much
doubted and the chances of achieving 0.4 g/mile NOx with even a
prototype version of an American type high-powered vehicle are
thought to be very remote.
In summary, a 97 kW (130 bhp), conventional naturally aspirated
swirl chamber diesel engine in a practical 3500 Ib vehicle should be
able to achieve the project target emission levels of CO 3.4 g/mile,
HC 0.41 g/mile and NOx 1.5 g/mile with timing retard and relatively
minor modifications to the fuel injection equipment. Some EGR may
be necessary to ensure a sufficient margin for production compliance.
A less powerful engine in the same vehicle would result in the
additional benefits of lower HC and NOx but an inability to meet the
performance requirements.
Particulates
With its complex heterogeneous combustion, local rich areas leading
to particulate emissions must exist in diesel engines. It has been
postulated by Wheeler in an unpublished Ricardo note that it is in these
rich areas immediately surrounding the fuel droplets that pyrolysis
occurs forming graphite sub particles of great reactivity which
subsequently oxidise rapidly upon meeting freely available oxygen or
coagulate into clumps like strings of beads. If the latter happens
slightly quicker than the former then soot escapes from the chamber
still actively coagulating. Even the highest temperature black soot
is still a hydrocarbon containing about 3% hydrogen by weight (CsHs).
The most important property of this soot is its high absorptive
capability for all active molecules. Thus as it cools and its attraction
increases it absorbs first high boiling hydrocarbons, then lower
boiling, then aldehydes and other oxygenates, then SO2 and finally
when down to ambient temperatures, fill-in of the remaining 'holes' is
completed by water molecules.
As mentioned in the smoke section, the droplet aerosols of blue and
white smoke are also products of very incomplete combustion (in
this case mainly due to marginal temperature conditions within the
cylinder).
-------
4-22
Thus there are two completely separate particulate problems in the
diesel engine, blue/white aerosol emissions at low load which can be
relatively easily removed by engine modifications, and the gas
carbon (soot) formed by pyrolysis in rich areas which increases
as stoichiometry is approached and is regulated by legislation
limiting its visual appearance.
Californian suggested legislative levels were set at 0.1 g/mile.
This level was originally intended for enforcement in 1975 (383).
Typical diesel powered, European type passenger cars appear to
emit particulates in the order of several times this level (0.5 - 3
g/mile) and it is difficult to envisage practical methods of reducing
particulate levels at source. A few authorities, notably Berliet,
(649, 650) have attempted to develop soot filters. Their early system
used Kaowool as a filter element, this being impregnated with an
oxidation catalyst. More recently, it has been shown that the catalyst
is not necessary for burning off the carbon particles, although it is
of some secondary use in that it can oxidise the CO resulting from the
carbon burn off to CC>2 (as well as oxidising the in cylinder generated
CO and HC).
The pure filter (i.e. non catalytic) system works by reason of the
fact that the residence time of the soot particles is increased
sufficiently to allow natural burn off to occur. With burn off occurring
at 450 - 500 C ( i.e. at a high load factor), the major problem behind
the development of a practical system is the incorporation of a
sufficiently large storage system in the exhaust to trap and store (hold)
soot particles emitted at part loads without causing excessive exhaust
back pressure increases in typical duty. It is in this area that much
research would be necessary into vehicle operating modes and engine
smoke characteristics before a system could be developed for
production.
To sum up therefore, at present the only controls on the particulate
emissions from diesel engines are in the form of smoke legislation
and most modern engines have been developed to comply with the most
severe of these (U.K. smoke regulation based on recommendation
BS AU 141a) . Californian suggested particulate controls of 0.1 g/mile
over the CVS-CH cycle are much more severe than this, current diesel
engines emitting 0.5-3 g/mile, and in order to comply with 0.1 g/mile
levels it would be necessary to develop a soot filter system. Much
development would be necessary to ensure a sufficient storage volume
for particulates emitted at light load and at below burn off temperature.
-------
4-23
Noise
To the passer-by the noise of the engine is probably the first obvious
indication that a particular vehicle is fitted with a diesel engine,
and for this reason it is an obvious cause of consumer resistance.
The larger the vehicle the greater is the likelihood that the engine
will be of direct injection form and the more obtrusive the noise to
be. However, the smaller high speed diesel engines fitted to
passenger vehicles and delivery vans will probably be of indirect
injection form and consequently less noisy (34). Although little
direct comparison is available, the indications are that for a similar
size, the IDl engine may be up to 5 dBA less noisy than the DI under
standard drive-by test conditions (39) and a comparison with a
gasoline powered vehicle shows an increase of 2.5 dBA for the diesel
after direct conversion from gasoline to a Ricardo Comet type of
diesel engine (142) (Ref. 112 shows a predicted 2 dBA increase,
Ref. 479 shows a predicted 4 dBA increase).
Noise at idle conditions is a prime consideration as, although the
absolute levels may be low, the characteristic of the noise is subject-
ively annoying and of much greater impact than for gasoline engines ,
especially for passenger vehicles (18). Comparisons have shown the
diesel to be considerably more noisy than the gasoline engine at idle
(2) but these tests were carried out on a four cylinder engine where
low frequency vibration of the diesel due to torque reaction under
high compression ratio conditions, since the engine is not throttled
when idling, gave subjectively a harsh idle characteristic. This
problem would be largely eliminated in a 6 or 8 cylinder engine and
the idle noise would be less objectionable but would still be very
pronounced as compared with a gasoline engine.
It has been shown that the majority of noise in the diesel engine is
combustion generated and it follows therefore that modifications to
the combustion and/or fuel injection system which gave a smoother
pressure diagram, would result in a quieter engine (2, 3, 631).
These measures include the use of pilot injection (2, 33, 89, 200) ,
injection timing changes (2, 201, 221, 324), nozzle changes (454)
and injection rate changes (3, 200). For instance, refs. 2, 33, 89,
and 200 quote reductions of 2 - 5 dBA using pilot injection, and refs.
201 and 324 quote reductions of l£ to 4 dBA with up to 6 injection
timing retardation. Although pilot injection is known to reduce the
noise by modifying the pressure diagram its application, in practice,
has not been achieved over the whole load and speed range of an
engine due to the variable fuelling characteristics of current fuel
injection equipment and the difficulty of obtaining an accurate balance
-------
4-24
between cylinders. The effect of injection equipment in reducing
the characteristic diesel clatter (164, 2) is important at idle
conditions, and one manufacturer (Peugeot) has introduced a
device which effectively reduces the injection rate at. idling.
Attention to the compression ratio is important in keeping combustion
noise to a low level. Refs. 153, 201, 253 and 531 report significant
reductions in noise with increases in ratio. (201 reports 1 dBA
reduction with CR increased from 18 to 20 in a Ricardo Comet
combustion chamber).
Fuel quality is also important in keeping diesel knock to a minimum.
Low Cetane number increases the noise level due to poorer ignition
quality and therefore longer delay between injection and combustion,
resulting in a more rapid pressure rise (2, 131, 243). However,
no advantage is to be gained above about 55 Cetane.
Noise of the engine is caused by vibration of the outer surface
resulting from forces within the engine being transmitted by the
engine structure to the outer surfaces. It follows therefore that
reduction of noise can be achieved by modification to these forces
and to the transmission paths, and by structural changes, bringing
about a reduction in the vibration of the engine outer surfaces. Much
research has been, and is being, carried out on noise reduction by
attention to engine construction and many papers have been published
including those by Priede (83, 301 etc.), Jenkins (4), List (48) and
Scott (2). In addition researchers have investigated the effects of
changes to piston pin offset (3, 378, 380), valve mechanisms (23, 377)
and timing gear drives (2), but the main conclusions are that more
noise reduction can be obtained by attention to the engine structure
than by other means.
By isolation and damping of the unstressed cover, a noise reduction of
5-8 dBA has been achieved (53, 111), others record 3-5 dBA
reduction (23, 84). By using a complete 'structured1 engine (174, 4)
up to 10 dBA reduction in noise wag obtained. However, this engine
was of experimental form, having a stiff frame containing the crankshaft
and cylinders, with the areas in between covered with non-vibrating
material, and it is not a practical' manufacturing proposition at the
present time. Future engines, however, may show a trend towards
'unconventional1 construction techniques if noise limiting legislation
becomes more severe.
-------
4-25
In order to reduce the noise of existing engines, in addition to
isolation and damping of the unstressed parts of the engine as above,
resort may be made to shielding and enclosureso Much research
has already been carried out on these techniques and it has been
shown.that 5-8 dBA reduction may be obtained by shields to the
engine (23, 267, 174, 84, 272, 371, 373), and by total enclosure
up to 20 dBA reduction has been obtained (23, 83, 84, 375). These
experiments were mostly carried out on the test bed where
conditions are ideal but, when attempting to shield or enclose an
engine in a vehicle, a more difficult situation'arises. It has been
shown that 5-6 dBA reduction can be obtained under these conditions
however (275), although_such a system of shielding or enclosing
increases the problems associated with maintenance.
Priede (83) has shown that engine noise can be calculated using an
empirical prediction formula, i.e. the noise (dBA) of a diesel engine
at 1m = 30 log-jQ N + 50 logjQB - 48.5, where N is the engine speed in
rev/s and B is the cylinder bore diameter in mm. This applies to a
4-stroke normally aspirated engine. Variants of the formula can be
used to predict noise values for gasoline engines, 2-stroke engines
and boosted engines „ These for mulae show that the noise of the
engine is a function of the rotational speed and cylinder bore diameter,
and ref „ 83 .concludes that for a lower noise 'engine a configuration
of. small bore, more cylinders, lower rotational speed and turbo-
charging would be desirable. Experience has shown that drive-by and
test bed noise can be related and that the noise of a vehicle at 7.5m
is some 15 -_17»5 dBA less than an engine at 1m „ This relationship
is dependent to a certain extent on the installation in the individual
vehicle due to variabilities in shielding by bodywork etc.
To keep vehicle noise to a minimum, careful attention should be paid
to the design of exhaust and intake silencers. It should be noted that
the intake on a diesel is unthrottled all the time (367)„ Optimisation
of silencing systems can more rapidly be carried out by use of a
computer aided silencer design program,
It is likely .that radiator fan noise could be a problem in both gasoline
and diesel powered vehicles (175, 373), and the use of thermostatic
control devices would aid the solution. With air cooled engines fan
noise can be excessive but by careful design this can to a large extent
be overcome (43, 257). Air cooling may have other advantages, for
instance the ducting system could.be used as a noise shield if careful
attention was paid to design.
-------
4-26
Interior noise can be a considerable problem in diesel engined
passenger cars unless care is taken. All the measures mentioned
above would reduce the amount of noise transmitted to the interior
and particular attention should be paid to engine mountings to
prevent vibration - particularly at idling conditions. Mos"t of the
published data on interior noise is concerned with commercial
vehicles but a small amount of passenger car information is
available (142), and shows an increase of 2 - 3 dB over the gasoline
version of the vehicle at 48 km/h (30 mph) and 3^- 4^ dB at 80 km/h
(50 mph). During acceleration the comparisons were ps follows:
48 km/h (30 mph) 3 dBA louder, 80 km/h (50 mph) 2^ dBA louder
and 113 km/h (70 mph) 1^ dBA louder. 'The latter figure suggests
that at speed the diesel car is less noticeably noisier than the
gasoline version and with more efficient sound insulation would be
no noisier o Provided the small diesel vehicle is constructed in such
a way that the engine noise is not too obtrusive, there is little reason,
on grounds of noise level, why such a vehicle should not be acceptable.
Volume
The close physical resemblance between the diesel and gasoline engine
results in very similar specific box volume when the comparison is
made on the basis of swept volume. Due to the slightly increased block
height (to ensure adequate piston compression height) and the greater
need for water cores between cylinders, the diesel is however slightly
bulkier than its gasoline counterpart. The cylinder head of the diesel
engine may also be slightly deeper due to its increased complexity.
When however a comparison is made on the basis of engine power
output, the diesel suffers a large penalty. This, as emphasised in the
next section, is due to the substantially higher specific output of the
gasoline engine. As a result, the gasoline engine may have to be
compared with a naturally aspirated diesel engine of some 50% higher
swept volume.
Increasing the specific output of the diesel by boosting would reduce
this disadvantage although it is likely that there -wduldstill be a height
penalty against the diesel (depending on engine configuration). The
improvement would be achieved at the cost of the increased technological
requirements of the engine (higher cylinder pressure demanding more
sophisticated gasket materials and stud patterns as well as higher
thermal loadings leading to the necessity of oil cooled pistons, oil
-------
4-27
coolers etc.). It is felt that reduced box volume alone is insufficient
to warrant this degree of complexity although the attendant reduction
in engine weight may be.
Forthcoming U.S. Vehicle Safety regulations are likely to have an
effect on engine length with a need to allow sufficient crush space
and it may be in this area that the diesel is at its greatest disadvantage
with regard to physical size, although the use of the vee configurations
in order to minimise length is equally applicable to the diesel as it is
to the gasoline engine.
In conclusion, the diesel in its naturally aspirated form has a significantly
larger box volume than a gasoline engine of the same power output.
Boosting the diesel would reduce this penalty but it is debatable whether
the increased engine complexity is warranted by reduction in box volume
alone. Due to the normal excess hood space in cars of this class, this
is not considered a major problem area although the introduction of
severe vehicle safety regulations may demand extra hood length for the
diesel powered passenger car, thus incurring a first cost and fuel economy
penalty.
The increase in volume over the gasoline engine will be of the order of
50% which, assuming similar bore/stroke ratios, will give an increase
of linear dimensions of 12%.
Weight
Experience with European powerplants and engine conversions reveals
that for reasons explained in the last section, there is only a small
difference (12 - 18%) in the specific weight of the two when the
comparison is made on the basis of swept volume (648 and Figs. 4-11
and 4-12)= Important reasons for the increased weight of diesel engined
vehicles are the weight of most diesel engine accessories (starter
motor, generator, fuel injection equipment, battery etc.) which are
substantially greater than their gasoline engine counterparts -
approximately 30% of the total engine weight penalty in Fig. 4-13 (648)
being due to the extremely large, heavy duty battery fitted to the diesel
powered vehicle»
For a given application, the dominant factor governing the relative weights
of diesel and gasoline engines however is that of specific power output.
-------
4-28
Because of its heterogeneous combustion mode, the high speed
indirect injection engine can use at most 90% of the air available
within the cylinder without exceeding socially acceptable levels
of exhaust smoke, and engines with other combustion systems
would use substantially less. The gasoline engine is capable of
consuming and efficiently using 100% of its air charge; this,
combined with its lower friction losses (Curve No. 5, ref 648)
and higher useful operating speed accounts for the much higher
specific power output of the gasoline engine in terms of bhp/litre.
Fig. 4-14 (Fig. 4 from 648) demonstrates the difference in specific
power output between the two engine types. At best the high speed
diesel engine delivers 24 kW/litre (0.5 bhp/in3) whilst its
European gasoline counterpart delivers a minimum of 30 kW/litre
(0.65 bhp/in3). Many American cars do in fact have a somewhat
lower rated gasoline engine, of similar specific output to the diesel
engine, but current American compact cars are powered more
closely following 'European philosophy1 although they are normally
somewhat heavier.
A greater use of aluminium components would go some way to
reducing the weight penalty at the expense of cost but it would of course
be open to the gasoline engine to move in the same direction.
Alternatively, boosting of the diesel engine can be employed although
any disadvantages of this must then be accepted.
For the 1600 kg (3500 Ib) vehicle under consideration, it is estimated
that an engine power output of 96 kW (128 bhp) is required to enable
performance targets to be achieved. Current American gasoline
engines of this output would have a swept volume of about 4.5 litres
(275 in ) and would weigh about 250 kg (550 Ib). It would however
be possible to produce a more compact gasoline engine of about 3 litres
(180 in3) which would produce 96 kW (130 bhp) at 80 - 100 rev/s
and would weigh about 186 kg . (410 Ib).
Fuel Economy
The fuel consumption of U.S. cars has been increasing over the years
due to increases in vehicle weight and engine size as a result of
increases in vehicle size and the fitting of automatic transmissions,
air conditioning and emission and safety devices (461, 464). The
standard size Ford sedan weight increased for example from 1610 kg
-------
4-29
(3550 Ib) in 1965 to 1940 kg (4275 Ib) in 1973, and it was necessary
to increase the engine size from 5 litre to 5.75 litre (300 to 350 in3)
Over this period the fuel economy deteriorated by 20% (463) .
The demand for gasoline in the U.S. is predicted to increase by 50 -
70% between 1973 and 1985. Three possible ways to moderate this
growth in fuel demand and still protect air quality criteria are:-
1. Wider use of mass transportation and car pools
2. Use of smaller cars
3. Use of diesel cars
Of various methods considered to conserve transportation energy, the
greatest saving can be achieved by conversion to small cars and
economy vehicles (467).
The calculated peak thermal efficiency for ideal engines shows the
gasoline to be 45% and the diesel 55%; the diesel advantage increases
however under low load conditions (20) . The fuel economy gain by the
diesel diminishes as the engine speed is increased due to the higher
friction and pumping losses (18).
While there have been a limited number of models of light duty diesel
engined vehicles in service, there are a considerable number of
references to the comparative fuel economies of diesel and gasoline
engined vehicles . Typical data are given in the following extracts from
the literature :-
The potential saving in fuel for the diesel as compared to the
gasoline vehicle is 5% for highway driving, 30% for mixed
duty and 50% for taxi cab service (479) .
Early London taxi company fuel consumption data suggested diesel
taxi 11.1 1/100 km (21.2 mpg) against 21.4 l/100km (ll mpg)
for gasoline, the diesel figure being taken over 8.8 x 10 km
(5.5 x 10 miles) but gasoline figures are likely to have im-
proved since that time (15).
Fuel consumption for Paris vehicles , 11.3 1/100 km (20.8 mpg)
for diesel and 15.7 1/100 km (15 mpg) for gasoline (15).
In London taxi service the gain in mpg is almost 2 : 1 for the
diesel over gasoline. In general use the diesel engined car
-------
is 30 - 50% better than gasoline (112).
By fitting diesel engines in place of gasoline engines in
U.S. Post Office ^, 1 and 5 ton vehicles, fuel economy
increases ranging from 18/23%, 10/28% and 150/180%
respectively were obtained, in controlled tests at
Aberdeen Proving Ground. From vehicles in actual service,
the £ and 1 ton vehicle economy was 50% greater than the
gasoline (31 ) .
From customer and service records, 770 kg (1700 lb)
vans fitted with Perkins 4-99 engines used for local
delivery, municipal work and distance haulage showed
fuel economy between 71.5 and 129. 8% better than gasoline
engined vehicles (363).
EPA tests carried out on diesel engined cars - Mercedes
220D, Opel Rekord 2100D and Peugeot 504. Diesels give
70% better fuel economy than gasoline, by the same
test procedures (l9l).
Ricardo measured road fuel economies of three different model diesel
engined cars were as follows : -
Litre/100 km
7.0
5.9
6.0
mpg (country route)
33.8
39.8
39.4
litre/100 km
8.9
7.2
8.2
mpg (town)
26.3
32.8
28.8
Reference 490 gives constant speed consumption.
Peuqeot 504
Mercedes 220D
Opel Rekord
litre/lOOkm
60
80
100
120
140
km/h
km/h
km/h
km/h'
km/h
5
6
8
10
12
.0
.1
.0
.3
.2
mpg
46.6
38.7
29.2
22.8
19.2
litre/ 1 00km
4.
5.
7.
8.
11
9
8
2
7
.4
mpg
47
40
32
27
20
.9
.5
.5
.1
.6
litre/ 100km
5.
6.
7.
9.
11
3
3
8
9'
.6
mpg
44.6
37.5
30.0
23.8
20.2
Austin A40 - measured fuel economy 45% better with diesel than with
gasoline engine (198).
-------
4-31
Comparing direct injection and indirect injection diesels of the same
cylinder size, the DI has no significant fuel advantage over the EDI
for automobile use. The DI gives about 10% better fuel economy on
the test bed but no difference on the road. This is due it is believed
to the ability to hold on longer in a high gear with the high torque
of the indirect injection engine.
To sum up therefore, the features of good fuel economy combined
with relatively low emissions are the major arguments for the case
of diesel powered passenger cars. In practical terms the fuel
economy advantage of the diesel is at its greatest at the light load
end of the engine operating range. European experience has shown
significant economic gains in the use of diesel powered vehicles
particularly for light load, local delivery and in city use. Rising
fuel prices as well as increasing service costs are steadily
increasing this advantage.
Actual fuel economy gains of the diesel vehicle over its gasoline
powered cousin are difficult to quantify because of the dominating
effect of vehicle operating mode. European in-city use has revealed
fuel economy gains of 50% to 100% in terms of miles per gallon. In
general use this improvement reduced to 30 - 50%. No figures are
available from vehicles with a significant highway driving mode but it
is felt that such vehicles would return fuel economy advantages of
less than 25%0 EPA data comparing diesel with gasoline powered
vehicles on a simulated .inertia weight basis showed that economies
of approximately 70% (again mpg) were realised (Fig. 4-15).
In the prototype vehicle, the high power/weight ratios must cause
the engine to be driven at light specific load factors and will also
allow higher gear ratios to be used, resulting in the use of low
average engine rotational speeds. Both these factors should enhance
the fuel economy advantage of the diesel (as opposed to comparing
two equal but low powered diesel and gasoline vehicles where the high
load factor and rotational speeds would tend to minimise the advantage
of the diesel)»
Although current European diesel powered vehicles comply with the
project emission targets while in production build, the high powered
heavier prototype vehicle will need engine modifications to ensure
compliance with the primary project targets. Data obtained from
the European vehicle suggests that if similar technology engines are
used, large amounts of timing retard will be needed to better 2 g/mile
NOx and a small but significant fuel economy penalty will be incurred.
-------
4-32
Assuming approximately 10 injection retard is used over the lower
half of the useful torque band, the resultant fuel consumption
penalty at steady state conditions would be in the order of 8 - 10%.
In real life driving conditions the penalty in fuel economy would be
less, probably 5 - 8%.
Fuel
It must be remembered that some 16% or so of the increase in miles
per gallon between the diesel and gasoline engined vehicles arises
from the higher density and corresponding higher calorific value of
diesel fuels as compared with gasoline. Whether this is a real gain
which can be fairly attributed to the diesel engine can only be
decided from a study of the energy economies of the production of
distillate fuels from crudes.
The world consumption of mid-distillate fuels is increasing rapidly
in both the under-developed and in the highly industrialised countries.
In those countries in the early stages of industrial development,
the directing of the greater part of their resources into construction
channels is reflected in measures taken to encourage the building of
all types of vehicles which use diesel engines and the discouragement,
by discriminating taxes or other means, of the manufacture of motor
cars for private sale and pleasure use. In highly industrialised
countries which have a moderately or very cold winter, space heating
of all new and many old buildings is increasing, with a consequent
increase in demand for light distillates for the heating installations (277)
The use of residual fuels is becoming less popular due to their high
sulphur content and there is a trend towards the use of lower viscosity
fuels, i°e. the mid-distillates (192).
In the U.S. , crude oil consumption is increasing at a rate of 100 million
barrels/year, and the estimated increased demand for mid-distillate
fuels is some 44 million barrels/year. By changes in refinery
techniques it would be possible to double the current production of
mid-distillates at the expense of gasoline and residual fuel production.
By means of these changes in refinery techniques and based on the
projected 1978 U.S. consumption, it is estimated that the production of
automotive diesel fuel could be increased from 130 to 1610 million
barrels/year, i.e. a 12-fold increase (194).
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4-33
Widening the cut of mid-distillates would give a fuel having a lower
Cetane number and flash point and an increased aromatic content.
The reduction of Cetane number would adversely affect noise and
starting (173). The increased aromatic content would result in
some increased engine wear (640).
For the high speed diesel engine the most important fuel qualities
are:-
1. High ignition quality to give short ignition delay
and good cold starting.
2. A boiling range which will give the highest possible
air utilisation and ease of starting combined with a
minimum of deposits.
3. Flash point low enough to enable use of the
volatile fractions.
4. Vapour pressure low enough to avoid vapour
lock,,
5. Sulphur content generally below 1%.
(555)
(Ricardo Note:- The level of this last item will probably be driven
down by future legislation) .
Amyl Nitrate is an ignition accelerator for diesel fuels - its use would
effectively increase the Cetane rating of the fuel.
Barium based additives are very effective in reducing diesel exhaust
smoke although this effectiveness varies depending on the engine
design. (109)° However, these anti-smoke additives are not attractive
because of pollution problems concerning particulates (173) .
A series of tests carried out at South West Research Institute on a rail-
road engine operating on propane as fuel was suspended before satis-
factory operation had been achieved (537).
On transit vehicles it is claimed that the use of propane results in
reduced fuel costs compared with diesel or gasoline vehicles. It is
also claimed that with propane as a fuel, engine wear and maintenance
is less and there are no problems with smoke andodbur (541).
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4-34
(Ricardo Note: - This last item refers to spark ignition propane
engines - the use of propane in diesel engines does not appear
attractive) .
Fuel at Low Temperatures
When fuel is cooled between cloud point and pour point, some
paraffinic constituents preciptate as wax crystals and can accumulate
in fuel systems causing fuel starvation. Flow improving additives
alter the crystal growth to give better cold flow (643).
Flow improvers give crystals which are smaller, less cohesive and
do not form a gel structure. These improvers do not alter cloud point
or the quantity of wax which separates. (642)
The flow of automotive diesel fuel at low temperature can be improved
by fuel dilution with kerosene. (314)
First Cost
The manufacturing cost differential between gasoline and diesel
engines is difficult to establish, in part because manufacturers who
make both types will seldom provide the necessary breakdown of costs.
The selling price of diesel vehicles may be no guide to the manufacturing
cost, being more an indication of what the market will stand or a means
of controlling the extent of the market. The relative production
quantity affects the cost of the engine and also that of the fuel injection
equipment which forms a major item in the increased cost. The data
available indicates that in Europe the light duty diesel engine costs 50%
more to produce than the gasoline engine. About half of this increase
is the cost of the fuel injection equipment. The remainder is attributed
to the greater complexity of the engine design in the combustion
chamber region, superior materials, larger starter motor etc. (479).
Three reasons account for the present difference in cost between the
diesel and gasoline engines: greater weight and complexity, closer
manufacturing tolerances on some parts, and more limited scale of
production (277).
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4-35
First cost is inevitably influenced by the number of engines built and
the degree to which common tooling can be used between diesel and
gasoline versions of a basically common engine. The diesel is
likely always to be more costly than the gasoline engine for the
following reasons:-
1. Unless there is a major breakthrough to some simpler
form of injection equipment, the cost of this item will
be higher than for gasoline engine ignition system and
carburettor.
2. Because of the higher cylinder pressures in the diesel,
some of the basic engine parts need to be stronger and
due to the combustion chamber form the cylinder head
will be more complex.
3. Because of the heavier starting load, the starter motor
and battery will need to be larger. (15)
For the diesel engined vehicle the difference in cost will be of the
order of 8 - 15% depending on the scale of production and on whether
or not a basically common design with common tooling is used for
the engine ( 15).
Ricardo believe that comparing British engines of equal capacity, i.e.
1.6 litres (98 CID) in four cylinders developing 37-54kW(50-72
bhp) respectively, the difference in cost between the diesel and
gasoline engine is approximately % 120. Of this difference over half
is accounted for by the increased cost of the fuel injection equipment
and starting aid compared with the cost of the ignition system and
carburettor for the gasoline engine.
Increasing the power output of the diesel engine to equal that of the
gasoline engine, either by increasing the engine capacity or by turbo-
charging, will increase the cost of the diesel engine by % 120.
This high first cost is one of the few economic disadvantages of the
diesel engine. Basically it is this initial penalty that the improved
fuel economy of the diesel must overcome before the economic
advantages of this engine type can be realised, although the longevity,
reliability and fuel economy advantages of the diesel should reduce
the depreciation rate of the vehicle compared with a gasoline powered
car unless the limit is set by body deterioration.
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4-36
Investigation of available data therefore reveals the major contributory
causes to the current first cost penalty of the diesel to be:-
1. Lower specific output
2. Complex fuel injection equipment
3. Relative production levels
4. Greater complexity
From the limited data available to Ricardo, it is estimated that the
true production costs of European automotive diesel engines are
approximately 2-2^ times that of a comparable (non-emission
controlled) gasoline engine. These figures are representative of true
current production costs, therefore production volume could have a
large influence. Absolute cost figures are not known with authority
but it is felt that the true production costs of current high volume
European diesel engines are probably in the order of % 3.25/kW
(# 2.4/hp). The cost of fuel injection equipment for the diesel engine
is obviously far higher than for ignition/carburation systems, and
although the general technology and in particular production tolerances
on injection equipment are far more severe than in any other section
of the engine, the basic design for most components (for example
pumping elements, control valves and injector nozzles) lend them-
selves to high volume production. Thus increased production levels
could reduce not only bare engine production costs, but also those of
ancillary equipment and in particular fuel injection equipment.
Further, the cost of fuel injection equipment does not increase pro rata
with engine power output, whether this is achieved by larger or a
greater number of cylinders and the relative percentage cost penalty
of a 97 kW(l30 bhp) diesel engine compared with a gasoline engine of
the same output is likely to be smaller than from the two comparable
engines of lower output.
Other factors which contribute towards the higher cost of the diesel
include more complex and accurate machining requirements around the
combustion chamber and the need for sophisticated materials in pre-
chamberso
In these high compression ratio engines, the total combustion volume
is very small and tolerances controlling piston/head bumping height
must be minimised in order to remain within the band of correct
chamber proportions for optimum performance. This means that all
components which have some control over this must be manufactured
to more exacting requirements than their gasoline equivalents (see
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4-37
Fig. 4-16), and a slight additional cost penalty is thus incurred.
Although the diesel has been shown to be considerably more expensive
than the gasoline in terms of first cost, it must not be forgotten that
in a typical passenger car the cost of the gasoline engine accounts
for generally less than 10% of the total retail price of the vehicle. This
proportion would increase with a diesel engine to perhaps 15 - 20%,
still only a modest proportion and one which might easily be justified
in a fuel economy conscious climate.
Maintenance
User experience indicates that major maintenance intervals for light
duty diesels may be three times as long as for gasoline engines (15).
The use of pintle type nozzles in IDI engines gives good reliability
and long service, 290,000 km (180,000 miles) is being quoted with
routine cleaning (100). Other sources quote 160,000 - 240,000 km
(100,000 - 150,000 miles) without removing injectors, even for
cleaning (112) and vehicle operation of 320,000 km (200,000 miles)
without attention was experienced by one taxi fleet operator.
Routine maintenance costs are difficult to define. Some operators find
the cost less for diesel than for gasoline engines due to the lack of
spark plugs, HT components, etc. (479), but others suggest that
routine servicing is slightly more costly due to the higher frequency
at which filter elements should be changed. Ref. (15) has a table
showing service costs of % 98 for gasoline and % 112 for diesel. In
general, the opinion of operators suggests that major engine maintenance
is much less frequent for diesel than gasoline engines but is probably
slightly more expensive per service. The net result however is reduced
maintenance cost and less time off the road due to breakdowns and
servicing. Extra cost of diesel vehicle recovered in first 40,000 km
(25,000 miles) was quoted by one operator (479).
From the literature survey it is apparent that the routine maintenance
costs of the diesel are very similar to those of the gasoline engine
and that if anything the shorter diesel lubricant life may make-the
diesel's servicing marginally more expensive.
Service life of pintle nozzles of 160,000 km plus (100,000 miles plus)
with the rest of the fuel injection equipment having similar maintenance
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4-38
requirements, will virtually eliminate major overhauls in a
passenger car with an average life of 160,000 km (100,000 miles).
Starting
While the diesel engine is totally immune to the problems of damp
ignition systems which affect gasoline engines, there are difficulties
in obtaining a quick start under cold ambient conditions.
The limiting temperatures for a start without aids depend on the
combustion chamber type as follows :-
Pre-chambers - It is necessary to use a starting aid (heater
plug) to obtain a quick and smoke-free start from below
10 - 15°C (50 - 59°F) with this chamber.
Swirl chambers - Some form of starting aid needed for the
first start of the day when ambient temperature is below
15 C (59 F) , subsequent starts are satisfactory if Pintaux
nozzles are used.
DI - Easy cold starting without aid down to -10 to -15 C
(14 - 5 F) is normal.
'M1 System - The cold starting of the 'M' system is inferior
to that of the DI because of the heat losses arising from the
high air swirl employed and a starting aid is required (34) .
The cranking speed of the engine is important for effective starting
and a minimum of 2 rev/s(l20 rpm) should be maintained. It is
important that the starter should not throw out of engagement when
the engine first fires (277).
The cranking speed can be increased under low temperature conditions
by changing to low viscosity lubricating oil. By changing from 10 to
5 SAE lubricating oil an increase of 1/3 in cranking speed can be
obtained.
A sufficiently high compression ratio is vital for good starting and
indeed the ratio is normally chosen from consideration of good starting
and low noise and is set at a higher value than would be desirable from
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4-39
considerations of best fuel economy.
A number of aids to assist starting are available in addition to the
provision of excess fuel which is universally employed.
1. Heaters installed in water jacket and/or oil sump.
These are usually household current operated and
provided the engine is isolated from the atmosphere
so that a general area of warmth is generated around
the engine, they are effective.
2. Heater plugs installed in the combustion chamber.
The term heater is a misnomer as they really ignite
the fuel rather than heat the chamber. There are
three types of heater plug: heavy gauge, shielded, and
thin wire. Tests showed that the ignition delay period
was least with the heavy gauge and longest with the
thin wire (54).
For satisfactory starting the element should reach at
least 1000°C (1832°F). (423)
With swirl chamber engines the gain in starting with a
heater plug can be as much as 30 C (54 F) and the
lowest limit for starting with heater plugs is around
-25°C (-13°F). (423)
A wait of 15 - 30 seconds is necessary for the heater
plug to warm up. (112)
3. Heater installed in the inlet manifold.
Manifold heating can be achieved by an electric heater
or by combustion of fuel within the manifold. The
probable gain in starting with an electric heater for
a DI is of the order of 10°C( 18°F) . (423)
Combustion heaters such as the CAV Thermostart or
Kygas burn fuel in the manifold and give gains in the
order of 10 - 15°C (18 - 25°F) on both DI and EDI.
(423)
With combustion heaters it is important to control the
fuel quantity or excessive flows will starve engine
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4-40
combustion of oxygen. (423)
In IDI engines manifold heaters are less effective than
heater plugs in the cylinder but may be easier and
cheaper to install and can operate with smaller delay
time. (277)
4. Aids based on fuels having low ignition temperatures.
Starting fluids are usually based on ether with addition
of lubricating oil and are very effective and can start
DI or IDI to -40°C. (423)
5. Devices applied to injection equipment.
Simple retard of injection timing for starting and the use
of Pintaux nozzles is employed with swirl chamber
o / o
engines which can improve starting by 10 C (18 F).
(423)
6. Other Methods.
Adding a small quantity of oil or gas oil to the inlet mani-
fold directly behind the inlet valve assists starting, by
improving the seal of the piston rings and increasing
compression ratio. (423)
Closing the inlet valve earlier gives an increase in
effective compression ratio but will give a penalty in
engine performance at full speed.
While Cetane number is an important fuel quality for
starting, at lower temperatures where a starting aid of
some form is necessary, the Cetane number is less
important. (564)
White/blue smoke on starting is reduced by using higher
Cetane fuel (38, 19)
White/blue smoke on starting is reduced by using more
volatile fuel. (38)
As mentioned earlier, some diesel fuels give trouble in vehicles in
cold weather due to clogging of fuel filters by wax crystals. Improved
design of the fuel system can help as can the use of flow improvers
in the fuel or kerosene dilution of the fuel.
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4-41
Starting of Passenger Cars in the USA
For the first cold start of the day, no insurmountable problems are
foreseen for the diesel powered passenger car provided the correct
starting procedure is followed. If heater plugs are allowed to
achieve working temperatures of 10OO C (1832 F)(entailing a delay
of 2O - 30 seconds with current devices) a satisfactory start in
0,0 o (
ambient conditions down to -20 to -25 C (-4 to -13 F) can be
achieved; the additional use of simple ether-based aids extends this
range down to -40 C. Care must be taken when using aspirated
special fuels with a very high ignition quality (e.g. ether) as
excessive amounts can cause damage due to extremely high cylinder
pressures and it is preferred that if this form of aid is to be used
regularly it should be administered automatically.
White/blue smoke can be a problem for some engines immediately
after a cold start, particularly at light load when fuel atomisation
and mixing is often poor. Any fuel introduced into the chamber late
in the combustion process is thus a potential cause of blue smoke,
and it is mainly in the field of development of fuel injection equipment
and reduction of production tolerances on injection timing settings
that improvements can be made- On the combustion chamber side,
any feature which decreases warm up time will minimise the time
period during which blue smoke is likely to occur, and both Daimler-
Benz with their pre-chamber and Ricardo with the Comet Mk.V
incorporate insulated chamber members in order to reduce this
problem of warm up time. Other means of cleaning up blue smoke
include increasing the Cetane number of the fuel or increasing the
compression ratio of the engine (already raised for starting consider-
ations far above the optimum performance requirement), or
increasing the exhaust back pressure, thus raising the mean load
factor and in-cylinder temperatures.
One subjectively annoying aspect of starting small high speed diesel
engines is the delay while waiting for the heater plugs to warm up.
Prototype plugs with virtually instant warm up times have been tested
but these have always encountered durability problems and it is
unlikely that an instant warm up plug could be developed to a production
stage within the next two or three years. A fast warm up plug (say
1O seconds) would be extremely desirable, or alternatively current
production plugs could be made more attractive if a programmed
start were adopted in which the heater plugs were activated manually
or automatically via some remote sensor/switch before the driver
gets into the car so that the subjective delay time is minimised.
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4-42
Even in warm ambient conditions, e.g. 15 C (59 F),it is desirable
to use some degree of pre-heat of the heater plugs in order to avoid
excessive cranking times (with fuel being injected and passing straight
through the engine) . Thus some form of programmer which
determines the minimum required pre-heat time to ensure an
immediate start would be advantageous. In order to predict required
pre-heat times both ambient and cylinder head temperatures would
need monitoring.
Thus the starting problem of the light duty diesel falls into the
following categories :-
1. Subjective annoyance during delay while heater plugs warm
up. This can be a problem even in mild ambient conditions.
A possible cure is a simple programmed start device or the
development of a fast warm up heater plug. 'Instant1 warm
up heater plugs would eliminate this problem altogether but
experience has shown durability and allied problems.
2. Excessive cranking with a warm engine due to the driver over-
riding the heater plugs leading to unburnt fuel being passed
through the engine. This problem could be eliminated by
incorporating an automatic programming system which senses
cylinder head and possibly ambient temperatures and
automatically selects optimum preheat time before the starter
can be engaged. The ultimate solution is a completely
automated start.
3. White/blue smoke - can be minimised by careful development
of fuel injection equipment for low speed operation. If this is
not successful, other techniques such as raising the exhaust
back pressure thus increasing engine working load are known
to be effective, but in this case it is accompanied by a fuel
economy penalty.
4. Starting during extreme winter conditions - in this respect the
starting of the diesel is inferior to that of the gasoline engine
in that a specific starting sequence must be followed to ensure
a start - as in (2) automatic programming devices may help
to reduce the subjective annoyance but the delay is still there.
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4-43
Hot Driveability
With the same torque curve for the diesel as the gasoline engine,
driveability will be similar to that of the gasoline engine. The diesel
engine normally has an all speed governor, the throttle control being
an engine speed control. If desirable however it would be possible
to fit a load control together with a governor giving idling and over-
speed control.
Current flywheels are small enough to give easy and rapid gear
changes, particularly on six cylinder engines. Transmission system
requirements are similar to those of the gasoline engine and no
special provisions need be made.
• * : t
Diesel engined vehicles in production, even with four cylinder engines,
can be driven down to about 25 km/h(l5 mph) in top gear. Engine flexibility is
very good and hot driveability should not therefore present any problems.
(Fig. 16, Ref. Nos. 489, 648).
Driveability is unaffected by emission control measures, such as retard
of injection timing. (112)
Cold Driveability
Once the diesel engine is running it will deliver power unfailingly
without the danger of stalling or hesitation which is a common fault with
gasoline cars if the accelerator is depressed suddenly when the engine
is cold. Driveability is also unaffected by emission control measures
such as retard of injection timing. (112)
Torque Rise
In the gasoline engine the torque curve shape reflects primarily the
breathing capacity of the engine modified somewhat over the speed range
by the change of mechanical and indicated thermal efficiencies. The
diesel engine torque curve is influenced by the same factors but can
be radically altered by a change in injection equipment specification.
Additionally, changes in combustion efficiencies will be greater in the
diesel. At low speeds the maximum torque must be reduced, in part
because of the higher proportionate heat losses and in part because of
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4-44
mis-matching of the air and fuel movements resulting in an early
onset of exhaust smoke. At high speeds the torque will hold up
within the limits of the breathing of the engine r
The ideal torque curve shape depends on the engine application. By
changing valve timings, injection timings or combustion chamber
characteristics the torque curve shape can be varied. (287)
(Fig. 4-17)
A torque curve back up of rather more than 10% would seem reason-
able for automotive diesel engines whether normally aspirated or
turbocharged. The speed range over which the torque back up will be
achieved will be less with the turbocharged than with the normally
aspirated engine, in general 65% and 55% of maximum speed,
respectively. (39l).
With Comprex supercharging the rated output can be increased by 40%
and at half speed, the torque can be increased to 70% above the
normally aspirated values. By suitable matching, the peak torque can
be set at whatever speed is required between 50 and 75% of full speed.
(l76)(Fig. 4-18).
A naturally aspirated swirl chamber engine producing 96 kW (128 bhp)
from 4.8 litre (292 CID) at 66.7 rev/s (4000 rpm) 6 Bar (88 lb/in2
bmep) would have, say, 10 - 15%torque back up down to a speed of
33.3 rev/s (2000 rpm).
The torque to be transmitted will depend on the engine type chosen. If
a high speed gasoline engine is replaced by a low speed diesel engine,
the increased torques may require a somewhat heavier transmission
but if the two engines run at similar speeds the same transmission
may be employed.
Durability
It is difficult to put definitive values on the durability of the high speed
diesel engine. Most user experience is concerned with commercial
vehicle operation; trucks and buses, etc., but operators of passenger
car vehicles which have been surveyed indicate that their experience
may be similar to that of the larger vehicle operators.
durability of the diesel engine is tied to wear rates, since excessive
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4-45
wear is the normal reason for component failure. This in turn may
be attributed to the duty for which the particular vehicle is subjected
and to the sulphur content of the fuel. Research has shown that
diesel wear on start up is much less than for the gasoline engine but,
due to higher loadings, steady state wear may be higher (456) .
Overall durability, however, should be greater than for the gasoline
engine if only because of the higher quality materials and more
rugged construction required for the diesel engine (15). Wear rates
of O.lmm/1000 hours at top ring level and 0.0016 mm/1000 km for
liner wear are quoted (100) if the liner temperature is kept above
100 C (212 F). The effect of fuel sulphur content was shown to
produce three times the wear rate when the sulphur content is
increased from .06% (desulphurised) to 1.4% (564). The sulphur
content was also shown to have an effect on carbon deposits which in
turn affect durability. Carbon deposits were also shown to be
affected by fuel Cetane number; reducing the number from 55 to 35
doubled the amount of carbon (564).
Experience with taxis suggests 320,000 km (200,000 miles) on one vehicle
without attention. Comparison of the durability of fuel injection
equipment with spark ignition equipment shows the diesel equipment
to require less maintenance and to have a greater life (15). The life
of injection equipment is variable but is quoted as 288,000 km
(180,000 miles) (100), and 176,000 km (110,000 miles) (639). In
this period routine cleaning of the nozzles may be required but some
users do not consider this necessary, particularly for the self
cleaning pintle type of nozzle used in indirect injection engines.
The factor which determines the time at which the diesel engine should
undergo a major overhaul is cylinder bore wear. When run on a low
sulphur fuel (i.e. less than 0.4%), current European engines achieve
approximately 160,000 km (100,000 miles) between overhauls and
during this time it may be necessary to inspect and clean the fuel
injectors once or perhaps twice.
Thus the durability of the complete engine package is known to be at
least as good as that of the gasoline equivalent. It is likely in fact
that the diesel engine will have a durability that considerably exceeds
that of the bodywork of the vehicle.
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4-46
Coolant Heat Losses
Heat transfer from the combustion gases is dependent on the temperature,
pressure and velocity of the gases, and in the diesel engine, on radiation
from carbon particles in the flame. At high loads, radiation plays an
important part, and while the mean gas temperatures of the diesel are
somewhat below those of the gasoline engine, the pressure and gas
velocities will be greater and as a result the heat losses from an
indirect injection dieseL at high load are greater than from a gasoline
engine of the same power output.
At part load, radiation plays a much smaller part and the mean gas
temperatures in the diesel fall rapidly and as a result, although the
pressures and gas velocities are not reduced appreciably, the diesel
engine has a lower heat rejection at low loads than does the gasoline
engine.
The differences are clearly brought out in Figure 4-19 from which it
can be seen that the heat loss to the coolant at full load can be 25%
higher for the diesel engine. In view of this higher heat loss, it is
normally necessary to fit a larger radiator than for the equivalent
gasoline engine where the size is often set by the need to prevent
boiling of the coolant under idling conditions with a high ambient
temperature. A radiator size of some 15% larger is necessary. (150).
The lower heat loss to the coolant from the diesel at light loads could
be a minor disadvantage in winter conditions in that it may marginally
affect cab warm up time with a cold engine. Under light load
conditions it may be desirable to fit shutters to blank off the radiator
to ensure a sufficiently high coolant temperature to give effective
defrosting of the windshield and to provide cab comfort. (31)
It is of course possible to use air cooling for the engine. This over-
comes the problem of leakage and coolant freezing. (257)
There are disadvantages with air cooling in that:-
1. With higher temperatures and the difficulty of getting effective
cooling of hot spots, the air cooled engine is generally rated
at a lower power.
2. The hotter cylinder walls result in a lower volumetric efficiency.
3. With no water jackets to absorb sound, the air cooled engine is
noisier.
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4-47
4. It is not so easy to provide vehicle heating.
The first of these is the most difficult, and the problems are accentuated
by the very high local heat flows which occur in portions of the cylinder
heads of indirect injection engines. There are in fact no air cooled
diesel engines employed in light duty service today.
Vibration and Torque Recoil
On approaching idling conditions from a higher speed on a diesel engine,
the exciting torque and the transmission of the engine mounting both
increase so that a rapid increase in the magnitude of the vibratory
forces applied to the vehicle frame is to be expected. With the
gasoline engine the exciting force is substantially smaller and is
further reduced by throttling of the air intake (367) .
After installing diesel instead of gasoline engines in £, 1 and 5 tonne
vehicles, driver reaction was that diesel vehicles were noisier and
suffered more from vibration (31).
If one makes the engine mounting soft enough to accommodate the
displacement involved at idling speeds, it is almost impossible to
avoid the engine passing through the natural frequency of the mounting
on starting up or stopping (15).
For these reasons very low natural frequencies of the engine on its
rubber mounting are chosen; i.e. below the lowest operational speeds
of the engine (367).
The use of automatic transmissions in vehicles has a beneficial effect
on the problem of torque recoil vibration by virtue of the fluid coupling
involved (15).
In general, very close attention to detail in the development of the
engine/body package is necessary to ensure an acceptably 'smooth'
vehicle, especially at low engine speeds and at idle, but Ricardo
believe that it is possible although it is never likely to be as good as
with a throttled gasoline engine.
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4-48
Manufacture
In overall design the automotive diesel engine is conventional and
could be made on the same production line as a gasoline engine.
Many gasoline engine parts could in fact be incorporated. The -special
feature which the diesel engine demands above all others is the
reduction to a minimum of the dead space outside the combustion
chamber. Clearance between the piston and cylinder head must
therefore be kept to a minimum and it is chiefly for this reason
that a simple shape, i.e. a flat cylinder head with vertical valves,
is maintained. The necessity for close tolerances on a large number
of parts in order to hold the piston to head clearance down to a
reasonable value will add to the cost of the engine as compared with
the gasoline engine. Other features of the diesel, apart from the
fuel injection pump drive, which differ from gasoline engine practice
result from the higher cylinder pressures in the diesel cycle. Thus
there is a main bearing between each cylinder, the piston and
connecting rod are extremely sturdy and the cylinder head is held
on by a thicket of studs (626).
The compression ratio of a typical 2^ litre, 4 cylinder Comet engine
can vary considerably due to production dimensional tolerances. By
far the greatest contribution to the variation arises from the
dimensions controlling the piston head clearances (122).
With regard to ancillaries, all these are very similar to those
currently used on American gasoline engines (although some, such
as the starter motor and alternator,may need to be slightly larger)
apart from the fuel injection equipment and a small air pump (vacuum
or positive pressure). The latter is needed for the power braking
system as the diesel, which does not throttle its intake air, does not
have a convenient, usable low grade pressure system as does the
gasoline engine.
Fuel injection equipment, although manufactured to very close tolerances,
is designed specifically for high volume production and is currently
manufactured in very large quantities on current technology equipment.
Lubrication
The internal combustion engine imposes a severe duty on its lubricant.
The lubricant must perform several independent functions including
the control of wear and corrosion yserve as a piston seal and coolant as
-------
4-49
well as flush away carbonaceous, partly burnt fuel residues from
critical areas. It must also function satisfactorily over a wide
temperature range with both temperature extremes introducing
their own individual problems.
Diesel engine lubricants are formulated to control corrosive wear of
piston rings, liners and bearings and to prevent the fouling of various
engine parts by the build up of deposits (570).
All modern lubricants contain additives to improve their performance.
The following properties are normally bestowed on the lubricant by
the addition of additives: -
1. Detergent/dispersant qualities so that insolubles such as
soot or oxidation products remain in suspension in the oil
and do not deposit on engine surfaces. This is much more
important in a diesel engine than in a gasoline engine.
2. Anti-oxidant qualities to discourage lacquer formation in
piston ring grooves and consequent ring sticking and to
reduce the rate of viscosity increase caused by oxidation products.
3. Anti-corrosive qualities to prevent the corrosion of bearing
liner materials. Also adequate alkalinity to neutralise any
sulphuric acid products that condense on the liner and hence
to suppress corrosive wear of liner and rings arising from
the higher sulphur content of diesel fuels.
4. Extreme pressure qualities to reduce valve train scuffing,
particularly tappets and cams.
5. Viscosity index improvements to give the oil multi-grade
characteristics.
6. Pour point depressants to prevent oil thickening at low
temperatures.
7. Anti-foaming and anti-rust properties.
It appears that prechamber engines are more severe from the additive
depletion aspect in their demands on heavy duty lubricating oil than
are their direct injection counterparts. (570)
-------
4-50
Oil change periods are normally dictated by additive depletion although
other factors can influence the change period, e.g. contamination of
lubricant with fuel, coolant or abrasives (570).
A reasonable oil consumption for an automotive diesel engine is 0.5%
of its gross fuel consumption.
Analysis of oil samples after short engine runs both with and without
exhaust gas recirculation showed that the oil used during the test with
exhaust gas recirculation had a significantly higher insolubles content
than the oil from the non-EGR test, suggesting that some engine wear
increase may occur as a result of EGR and that more frequent oil
changes will be required with EGR.
CONCLUSIONS
Smoke
Black Smoke - This should not be an aesthetic problem if the engine
complies with the Federal Smoke Regulations for heavy duty vehicles.
(This should be obtainable by attention to local mixing and the overall
air:fuel ratio at rated conditions. Ratings can be controlled to take
account of altitude effects).
The high power;: weight ratio of the American car should mean that
visible smoke conditions will only be obtained for extremely short
periods during hard accelerations.
Turbocharged engines may have a low speed transient problem but
the 'Comprex' pressure exchanger might be a solution to this.
Blue/White Smoke - This can be unpleasant from the sidewalk
particularly as it is formed under idle conditions, but the problem
can be minimised by careful attention to combustion chamber design
and fuel injection characteristics.
Odour
The small high speed diesel can have an odour problem, particularly
at light load conditions if misfire is approached. The problem can be
minimised by the addition of a light load advance mechanism.
-------
4-51
Odour at full load can be minimised by combustion chamber development.
The proposed reductions in smoke levels should alleviate this problem.
The identification of several odorous components has been achieved
but quantitative assessment has yet to be perfected. The A.D. Little
Odormeter may advance technology in this area by a significant step.
Gaseous Emissions
In general, turbocharging increases NOx further by increasing the
charge temperature but it allows further retard for the same smoke
limit.
Exhaust gas recirculation is effective in reducing NOx levels (particularly
over the CVS-CH cycle) but durability has yet to be proved and it does
tend to increase smoke emissions.
Although water injection has the benefit of reducing NOx without
significantly affecting engine performance, the logistics of the installation
and the problems of engine durability make this measure unattractive.
Timing retard is undoubtedly the most effective single parameter for
the reduction of NOx and the fact that the smoke limited performance
of the IDI engine tends to deteriorate less with retard than the DI
gives it a major advantage in this field.
The limited data available indicates that emission levels from 2-stroke
engines should be of the same order as from 4-stroke engines of
similar performance.
Heavy duty experience leads to the conclusion that the use of a
conventional direct injection chamber will increase both NOx and CO
levels while HC levels might rise rapidly with retarded timings. It
seems almost certain that a high speed (67 rev/s) (4000 rpm)
conventional, naturally aspirated direct injection engine would not
achieve the primary emission levels due to its low smoke limited
performance at retarded timings.
For a naturally aspirated 4-stroke indirect injection engine it can be
predicted that 3.4 g/mile CO can be achieved; 0.41 g/mile HC could
be attained on prototype vehicles although this figure may not be
held in production ; and 1.5 g/mile NOx could just be obtained from
a prototype current generation engine although some exhaust gas
recirculation may be necessary to allow a margin for production
compliance.
-------
4-52
Although 0.4 g/mile NOx has been achieved with a highly modified
prototype engine in a European type vehicle, it is extremely unlikely
that this figure would be achieved with a heavier vehicle and with a
higher power to weight ratio.
Any diesel powered vehicle would have less difficulty in achieving
the target objectives if both weight and power to weight ratio were
reduced. A lighter, lower powered vehicle would also have improved
fuel economy.
Particulates
The early suggested Californian requirement would present a major
problem for all engines with a heterogeneous combustion system and
this is undoubtedly a problem area for the diesel engine. However,
the true effect of particulates on health is unknown at the moment
and the problem could be solved by the addition of filter systems
although this move would carry a high cost penalty.
It is considered that work should be initiated into the distribution of
particulates from different types of engines and their true health
hazard determined before any legislation is finalised. An understanding
of their formation within the engine might also be a useful tool for their
control and thus a fundamental investigation using experimental and
analytical techniques should also be started.
Noise
The drive-by noise levels of diesel powered vehicles are slightly higher
than those of gasoline powered vehicles but there is no reason why
light duty vehicles should not meet proposed noise legislation.
The idle noise is annoying to the by-stander as well as the driver with
present European vehicles.
There is no reason why diesel powered light duty vehicles should be
unacceptable to either the driver or by-stander if sufficient attention
is paid to details of the engine and vehicle construction.
Volume
For the same power output the diesel engine is likely to be greater in
volume than a highly rated gasoline engine, but the increase in volume
is unlikely to pose any major problems.
-------
5-35
The main variables are listed below :-
Engine type - 4 stroke
2 stroke (valve in head, loop
scavenged-opposed piston)
Engine compression ratio
Boost pressure
Cylinder pressure limits
Air fuel ratio
Valve timing
Turbine efficiency
Compressor efficiency
Use of charge cooler - charge cooler efficiency
Engine /turbo components drive arrangement
Thermal loading limitations
Bore/stroke ratio
Although there is a fairly extensive literature study of the subject,
the little practical experience that exists is concerned with large
engines for which this configuration is most suited. The performance
predictions are also concerned with large engines and are mostly
confined to maximum output conditions without consideration for
part load operation.
In order to arrive at fairly broad estimates for the performance of
a small engine, the available information has been used and modified
where possible to take account of the scaling down of the components.
Engine Configuration
The simplest form of compound engine is that in which the output is
taken from the diesel engine with the turbo-compressor unit geared
to the crankshaft. The gearing has to cope only with the difference
in power between the compressor and the turbine, and there is no
requirement for power equilibrium between these components
such as exist in normal turbocharging.
-------
5-36
With a gas generator system all the power output is taken from the
turbine through reduction gearing. For the relatively low power
requirements of the engine under study the small turbine would
have a very high rotational speed. Although this arrangement has
the theoretical ad vantage of a high-low speed torque compared with
the compound engine, it is not considered to be feasible for the
present application in view of these high rotational speeds and
the unknown performance of small turbines at high expansion
ratios.
Engine Type
From the available data a 2-stroke engine is seen to offer the best
thermodynamic performance and power/weight ratio with an opposed
piston type being superior to the valve in head form. It also permits
greater control over the division of power between the diesel engine and
the turbine compared with a 4-stroke engine. However, although the
2-stroke engine is basically simpler it is considered that the 4-stroke
might be preferred since it has much lower thermal stresses and
piston heat flow and has greater volume of experience behind its
development.
Estimated Performance
The values tabulated below have been derived from data predicted
for large engines with the following assumptions :-
Boost ratio 4:1- Inlet pressure 4.12 Bar ( 60 lb/in2 abs)
Charge cooler effectiveness 80%
Charge cooling water temperature 85 C
Compressor efficiency 74%
Turbine efficiency 82%
Cylinder pressure limit 207 Bar (3000 lb/in )
The final column shows the engine swept volume for 96 kW (128 bhp)
at an engine speed of 50 rev/s.
-------
4-53
Weight
When compared with a highly rated gasoline engine of 97 kW (130 bhp)
a diesel engine of the same power output may be as much as 136 kg
(300 Ib) heavier, but the weight penalty may be as little as 68 kg
(150 Ib) if a more normal low rated gasoline engine is chosen for
comparison.
Fuel Economy
The diesel engine is the most efficient current and practical power
plant for light duty use.
Its fuel economy advantage over the gasoline engine will increase
as the load is reduced and may be as much as 50% (twice the miles
per gallon) in city use.
Fuel
As fuel costs increase the high economy diesel engine becomes more
and more attractive.
Automotive diesel fuel production could be increased by up to twelve
times the current levels by 1980.
If it became necessary to introduce wider cut fuels noise could be
increased and cold starting worsened, emission levels could also
increase.
First Cost
The first cost of the diesel engine will be significantly higher than that
of the gasoline engine for the same power output and emission levels.
Such evidence as is available indicates that the diesel engine in
America will cost between 1.5 and 2 times as much as the gasoline
engine. For a 97 kW (130 bhp) engine this means some # 300 against
£ 200 production cost and about half of this extra cost is due to the
extra cost of fuel injection equipment.
Maintenance
Minor maintenance requirements of the diesel engine (oil change,
filter elements, etc.) are equivalent to the gasoline engine although
the cost of these items may be slightly higher.
-------
4-54
The periods between major overhauls for the diesel are greater so that
over the life of the vehicle overall maintenance costs will be less than
for a gasoline engine-
Starting
Starting is inferior to the gasoline engine particularly in mild ambient
conditions although a programmed start would reduce the annoyance
of the 30 second (maximum) delay for heater plugs.
The light duty diesel engine will start down to -25 C (-1.3 F) with heater
plugs, but without external aids, and down to -40 C (-40 F) with
simple external aids. Below this temperature more complex external
aids are required.
Hot Driveability
For engines of equivalent power the low speed torque will be greater
with the diesel than with the gasoline engine.
Although the operational speed range of the diesel is less than that
of the gasoline engine its low speed smoothness is frequently better.
Cold Driveability
The cold driveability of the diesel engine is as good as its hot drive-
ability and is considerably superior to that of the carburetted gasoline
engine.
Torque Rise
The gasoline and diesel engines will have similar torque characteristics.
If the diesel engine runs at an appreciably slower speed than the
gasoline however, the high torques may require a slightly heavier
transmission.
Durability
The durability of the diesel engine will be greater than that of its
gasoline counterpart. The diesel may in fact be too durable and out-
live the vehicle, but it is difficult to reduce engine life without
seriously impairing durability and reliability.
-------
4-55
The durability of the diesel fuel injection system is greater than that
of the gasoline engine's ignition and carbitration system.
Coolant Heat Loss
The reduced heat loss at idle and low load will reduce the problems
of * traffic- jam -boil-over1 but may make winter morning de-icing
and warm-up slightly more difficult.
The increased heat loss to coolant at full load may dictate the use of
a larger radiator, but this will possibly only be required for vehicles
towing trailers.
Vibration and Torque Recoil
The torque recoil from the unthrottled, high compression ratio
diesel is undoubtedly greater than from the gasoline engine, particularly
at idle. This tends to give the impression of harshness which disappears
once the vehicle is in motion.
The problem can be minimised by attention to the engine mountings.
Manufacture
The diesel engine is sufficiently similar to the gasoline engine to allow
it to be made on the same production line if necessary, although
greater control of tolerances is required.
Ancillaries are all similar to those of the gasoline engine and in many
cases are identical.
The production techniques for the high volume production of fuel
injection equipment have already been developed in Europe although
there may be some scope for cost reductions with even greater
production quantities.
Lubrication
The diesel engine demands lubricants with higher dispersant and anti-
corrosive properties than the gasoline engine. With present day oils
this implies that oil changes might be slightly more frequent.
-------
4-56
Areas likely to Benefit from further Work
1. Emissions
A fundamental attraction of the diesel engine is that its emission
characteristics can avoid the use of catalysts. A careful study of
the origins of the controlling parameters in the formation of unburnt
hydrocarbons in light duty diesels would yield much valuable
information and perhaps extend the use of non-catalyst controlled
power plants.
2. Particulates
Current ignorance on the spectrum of particulates from a given
engine, the relative dangers of particulates of various sizes and
compositions, and the effects of blanket legislation would indicate
that work should be in:' ated into these unknown areas before any
legislation is passed.
3. Starting
Further work on 'instant1 glow plugs and programmed starts may
increase the acceptability of the diesel engine.
4. Fuel Injection Equipment
About 50% of the cost differential between diesel and gasoline engines
is due to fuel injection equipment while both emissions and noise
might benefit from more versatile injection systems. Because of
this research into cheaper and more versatile fuel injection equipment
should be considered.
-------
EPA PROJECT RFP WA 73-R176
BAR CHART OF INSTITUTION AND CONFERENCE PAPERS SURVEYED
PUBLICATION
SAE TRANSACTIONS
INDIVIDUAL SAE PAPERS
I.MECH.E. PROCEEDINGS
I.MECH.E. A.D. PROC.
J.S.M.E. BULLETIN
J.S.A.E. BULLETIN
CIMAC CONFERENCE PAPERS
FISITA CONFERENCE PAPERS
MIRA JNCLASS. REPORTS
MIRA V.A. REPORTS
MIRA E.A. REPORTS
ASME PAPERS
ASME TRANSACTIONS
ASME JOURNAL
YEAR
50
51
^_
52
53
5*
55
!
56
57
58
59
60
,
f
61
62
63
6*4
65
66
67
68
69
70
71
72
73
-------
trM
r\rr- WH
BAR CHART OF JOURNALS SURVEYED
JOURNAL
MTZ
ATZ
DIESEL 6 G.T. PROG.
DIESEL 6 G.T. PROG. W.W. ED.
DIESEL & G.T. PROG. U.S. ED.
SAE JOURNAL
C.M.E. (I.MECH.E. JOURNAL)
I.MECH.E. A.D. JOURNAL
AUT. DESIGN ENGINEERING
SI A JOURNAL
AUTOMOBILE ENGINEER
GAS & OIL POWER
AUTOMOTIVE INDUSTRIES
MECHANICAL ENGINEERING
(ASME JOURNAL)
G.M. ENGINEERING JOURNAL
COMBUSTION ENGINE PROGRESS
YEAR
50
51
52
53
5*4
55
56
57
E.
58
53
60
61
62
63
6/4
65
66
67
68
69
70
,
71
72
73
i
-------
4-59
FIGURE 4-3
A STUDY OF THE DIESEL AS A LIGHT DUTY POWER PLANT
LIST OF EUROPEAN DIESEL VEHICLE OPERATORS
CONTACTED TO OBTAIN USER EXPERIENCE DATA
OPEL, RUSSELSHEIM, ¥. GERMANY.
PEUGEOT, PARIS, FRANCE.
DEUTZ, PORZ, W. GERMANY.
B.L.M.C., LONGBRIDGE, ENGLAND.
B.P., LONDON, ENGLAND.
C.A.V., LONDON, ENGLAND.
LONDON GENERAL CAB CO., LONDON, ENGLAND,
VERBAND FUER DAS PERSONENWERKEHRSGEWERKE,
HAMBURG, W. GERMANY.
SLOTA TAXIS, PARIS, FRANCE.
AUSTIN HIRE & TAXI SERVICE LTD., WORTHING,
ENGLAND.
-------
4-60
FIG. No. 4-4
Drg. No 0 2-GtOCtB
Date £8-7-74
2 LITRE D.I. CONVERSION
PERFORMANCE COMPARISON - DJ. AND COMET V BUILDS
EXHAUST SMOKE MEASUREMENTS AT OPTIMUM CONDITIONS.
FIXED STATIC PUMP TIMING
OPT. SWIRL 55 rev/s FOR D.I.
JbARO; D.!. 7£4 *r»m Hg.
COMET 77O mm
-------
COMPARISON
t, < FIG. No. 4-5
4-61 DRG.No. D26203
DATE*- JULY 74
OF FULL LOAD CHARACTERISTICS FOR THREE
COMBUSTION CHAMBERS ON 2-LITRE 4-CYLINDER ENGINE
i
(FROM PAPER BY DR. EISELE - DAIMLER BENZ)
D.I. - DIRECT INJECTION
P.C - DAIMLER BENZ PRECHAMBER
S.C. s SWIRL CHAMBER COMET V
BMEP
I- 5
8*
o
CD
I
D.l.y
PC.
S.C,
EXHAUST SMOKE
HO i
IOO-
9O -
BO-
70J
-7-5
- 7
-6-5
o
.0
h 6 I
55
- 5
FUEL CONSUMPTION
I OOP
i
2O
rcv/min
2OOO 3OOO
1
4OOO
3O 4O 5O 6O
ENGINE SPEED-r«v/s
•=0-5-
O-4-
0-3-J
-35O
-300
O
UJ
-25O
U
ui
•200ft
7O
-------
4-62
FIG. No. .4-6
DRG No. D as71 3 A
DATE:- £8-7-74
NO EMISSION AND PERFORMANCE
CHARACTERISTICS OF D.I. AND SWIRL CHAMBERS.
AVERAGE DATA FROM NATURALLY ASPIRATED ENGINES AT
2.1:1 AIR : FUEL. RATIO & PISTON SPEED CIRCA G-Sm/SEC.
E
q
d
O
asoo
aooo
I5OO
1000
50O
oL
350
O
U.
£50
aoo
4 SPRAY D.I.
SWIRL CHAMBER
NO
SPECIFIC CONSUMPTION
'EARLY
LATE
0)
I
-------
FIG. No. 4-7
I^b£.2!i.f£5£9Sl^^ Drs NO. D a&c
& SMOKE OF A, SIX CYLINDER SWIRL CHAMBER Date £8-7-74
EFFECT OF EXHAUS
F3 C~ .* i. j- #•' i -~ *«*-, ,*> A. ft /* f»i ,»*i ; «" f*«, "«• /-^, r" .^> £*,.» ; j .« -• (5 P« '
K&L.."< w-ufctJ L**X£» L-OC-u.fc.lJ i
-------
4-64
THE RESPONSE OF NATURALLY ASPIRATED
ENGINE NOX EMISSIONS TO LOAD.
FIG No 4-6
Drg. No. OZ6O&9
Date 38-7-74
GCYL.
L COMET V ENGINE
G CYL. 5-40 L Dl ENGINE.
1 FULL SPEED,
j RETARDED £°
TIMING
-------
4-65
THE RESPONSE OF TURBOCHARGED
FIG. No 4-9
Drg. No. DZ&O70
Date £8-7-74
ENGINE NOx EMISSIONS TO LOAD.
GCYL. G-81 COMET V ENGINE
G CYL. 5-41 L D.I. ENGINE
}
FULL SPEED, TIMING
RETARDED 6°.
-------
VEHICLE
MERCEDES 220D
* MERCEDES 220D
OPEL REKORD
PEUGEOT 504
DATSUN 220C
DATSUN 330C
INERTIA
WT
3500
3500
3000
3000
3500
4500
SWEPT
VOL. IN3
134
134
127
129
132
198
COMB.
CH.
PRECH
PRECH
COMET
COMET
COMET
COMET
TEST
LAB
EPA
EPA
EPA
EPA
EPA
EPA
G'BOX
4 AUT
4 AUT
3 AUT
4 MAN
4 MAN
4 MAN
NO. OF
TESTS
5
5
4
5
2
3
HC
g/mile
0.34
0.28
0.40
3.11
0.21
1.70
CO
g/mile
1.42
1.08
1.16
3.42
1.69
3.81
NOX
g/mile
1.43
1.48
1.34
1.07
1.72
1.71
MPG
(US gall)
23.6
24.6
23.8
25.2
24.0
21.4
1/100 km
10.0
9.5
9.9
9.3
9.3
11.0
E
o
x
H
d
c
H
w
f
O
£
M
2
h— I
W
CO
HH
§
s
w
* MODIFIED F.I.E. - ALL OTHER ENGINES IN PRODUCTION BUILD (NOTE DATSUN 330C
WAS A NON-STANDARD MARRIAGE WITH A 1973 FORD F250 PICK-UP)
-J
en
3
T3
f)
O
2C
2
S'
0)
i
h-k
To
-------
Engine Make
and Model
Peugeot XLD
Prototype ®
BLMC 1.5
Perkins 4.108
Prototype
Peugeot XDP4.88
Isuzu C190
Opel 2100
Prototype
Peugeot XDP4.90
Nissan SD 22
Citroen CRD90
Mercedes 220D
Land Rover 2\
Ford York
Mercedes 240D
BLMC 2.5
Mitsubishi 4DR50
Bore x
Stroke
mm
75 x 71
78 x 71
76 x 76.
73 x 89
79 x 89
80 x 89
88 x 80
86 x 84
88 x 85
86 x 90
90 x 83
83 x 100
90 x 85.5
87 x 92.4
90 x 89
94 x 86
91 x 92.4
89 x 101.6
92 x 100
Swept
Volume
1
1.25
1.36
1.39
1.49
1.76
1.80
1.95
1.95
2.07
2.09
2.11
2.16
2.17
2.20
2.29
2.36
2.40
2.52
2.66
Power
Output
kW
(DIN)
30
33.5
27
29.5
36.5
38
45
46 gross
45
49 gross
48.5 .
48.5
46
45
50 gross
46
48
48
56.5
Rated
Speed
rev/s
83.3
83.3
66.7
66.7
66.7
66.7
75.0
73.3
73.3
70.0
75.0
66.7
75.0
70.0
66.7
60.0
70.0
58.3
58.3
Weight
kg
152K
216
185
208
174
170
205
199
208
185
186
206
184
243
222
268
261
kg/1
122*
155
125
118
97.5
87.4
105
96.5
99.5
87.5
85.5
95
83.9
106.5
94.5
106
98.2
kg/kW
5.10*
7.80
6.29
5.70
4.58
3.80
4.44+
4.45
4. 65"*"
3.82
3.84
4.49
4.11
5.39+
4.80
5.62
4.60
kW/1
23.7
24.8
19.8
19.7
20.7
20.4
23.0
21. 4+
21.6
21. 6+
23.0
22.4
21.0
20.4
19. 7+
19.6
20.1
18.9
21.3
® Included for reference, circa 1956
Weight includes integral gear box
Assuming DIN power 10% less than
d
g
M
O
o
w
to
d
H
O
2
§
-------
4-68
COMPARISON OT DICTSEL AND GASOLINE ENGINE
WEIGHT/SWCPT VOLUME
GASOLINE CUBVg TAKEN FROM I. MgCM-g-PAPEB C34S/73
FIG No. 4-12
Drg No 0246J2A
Date 0.1.74
-------
4-69
DIESEL V. GASOLINE WEIGHT ANALYSIS
FOR 4 CYLINDER ENGINES
Figure 4-13
Component
Combustion &
Ignition
Equipment
Cylinder Head
(Cast Iron)
Diesel Gasoline
Item
4 Cyl. DPA +
Delivery Valves
4 Injectors
Complete
4 Heater Plugs
Statistical
Weight
kg
7.13
1.85
.34
9.32
25
Item
Distributor
Lucas
Plugs + HT Lead
Coil Lucas LA12
Carburettor
Statistical
Cylinder block difference due to increased height etc.
Pistons & Con.
Rods
Crankshaft
Flywheel
Starter Motor
Vacuum Pump .
Battery
Sound absorption
material
Vehicle
4 Units
Insufficient
statistical
data - estimated
only
1 Lucas M45G
1 off Pierburg
1 Lucas BT11A
72 amph
Estimated
8.1
16.5
10.5
1
29.5
2
4 Units
Insufficient
statistical
data - estimated
only
1 Lucas 2M100
None
1 Lucas BT7A
43 amph
Weight
kg
1.13
.79
1.47
3.84
19
5.9
12.3
7.9
-
19.5
Diesel
Penalty
kg
5.5
6
2
2.2
4.2
2.6
17
1
10
2
13
ESTI MATED TOTAL PENALTY 35.5
Mercedes 220D
Peugeot 504D
Peugeot 204D XLD
1375
1279
955
Mercedes 220
Peugeot 504
Peugeot 204
(1130 cc)
1335
1230
935
40
49
20
-------
4-70
POWER/LITRE OF SOME CURRENT
GASOLINE' & DIESEL ENGINES.
FIG. No. 4-14
Drg.No. DZ6O7I
D.te £8-7-74
O GASOLINE ENGINES
* DIESEL ENGINES
ENGINE CAPACITY - Litr«
-------
4-71
FUEL ECONOMY - m.p.g. (U.S.) FUEL CONSUMPTION - L/IOO Km. ._
FIG. No. 4-15
DRG. No. D26O9
VEHICLE FUEL CONSUMPTION Vs INERTIA _ _
VEJGHT DURING LA4 1975 (CVS-CH) TEST CYCLE.
I
ao
15
10
5
£5
10
15
10
5
1C
+ DIESL P
JRES FOR 1957- &7 GASOLINE VEHICLES.
JRES FOR 1973 GASOLINE VEHICLES.
OWERED VEHICLES- EX EPA DATA.
500 1000
VEHICLE INERTIA W
xvs^
X
.^'"
«
: t
isoo aooo
EIGHT -kg.
• t^m
^^^^rr-
3
74
i
>oo aooo 3000 4000 sooo
VEHICLE INERTIA WEIGHT- Ib.
"\
-------
•*- I £.
FIG. No. 4-16
DRG. No. S5802
DAU:- 31-7-74
0.25
BLOCK
GASOLINE
0.10
THROW
1
/
DIESEL
0.050
CON ROD
CENTRES
0.15
CON ROD
CENTRES
TOLERANCES CONTROLLING PISTON_TO
CYLINDER HEAD CLEARANCE MM.' "
FOR ENGINES OF b5 m/m TO IIOm/m STROKE
0.100
BLOCK
-------
4-73
FIG. No A-H
Drg. No. D2GO7£
Date Z8 - 7 - 74
EFFECT OF INLET VALVE TIMING ON
LOW SPEED TORQUE.
NLET CLOSING
- rev/s
ENGINE SF
-------
4-74
TORQUE - COMPARISON!
ON A. BOO HP Dl-DIESEL ENJQNE
FIG. No 4-18
DRG. No. 025 7 22
DATE.-
IOO
o
8-
01
3
a
(*
o
M^XIDYNE TC.y* \
(MTZ 1912) \
NATURAL ASPIRATED
-------
4-75
HEAT REJECTED TO COOLANT FOR SIMILAR
DIESEL & GASOLINE ENGINES OVER THE LOAD AND
SPEED RANGE
FIG. NO 4-19
Drg. No D 26O73
Date £8-7-74
DIESEL.
GASOLINE.
SWEPT VOLUME, BOTH ENGINES = I • 49 L
ES OF CONSTANT*
HEAT REJECTED
BRAKE OUTPUT
,.
, i , i I . . . ,X4 . . . - , ..... -t
'• • - • I ! ' ' t • I : • * • < i i t • ; -i
• • ' • I • t • I < . • ^ • • , » ..... • 4
IG-7rev/s J
ENGINE BRAKE OUTPUT
-------
5 - 1
SECTION 5
ENGINE CONFIGURATION STUDY
This section contains details of the diesel engine configurations which
were schemed as potentially viable light duty diesel power plants.
Information on two comparable gasoline power plants is also included.
The diesel power plants were all designed to propel the target vehicle
for the study, a 4-5 seat sedan with a loaded weight of about 1600 kg
(3500 Ib) capable of 0-96 km/h (0-60 mph) in 13.5 s and 32-112 km/h
(20-70 mph) in 15 s. Computer calculations indicated that a bare
engine power of about 97 kW (130 bhp) was required if a conventional
3-speed automatic transmission, was assumed.
The engines included in the configuration study were as follows :
U V-8 Gasoline 4^stroke
2. IL-6 Gasoline 4-stroke
3, V-8 Indirect Injection Diesel 4-stroke
4. 6 Cylinder Indirect Injection Turbocharged Diesel 4-stroke
5. 6 Cylinder Indirect.Injection 'Comprexed1 Diesel 4-stroke
6. 6 Cylinder Direct Injection Turbocharged Diesel 4-stroke
7. 6 Cylinder Direct Injection 'Comprexed1 Diesel 4-stroke
8. 6 Cylinder Indirect Injection Loop-scavenged Diesel 2-stroke
9. 6 Cylinder Direct Injection Uniflow Diesel 2-stroke
10. 4 Cylinder Direct Injection Compound Diesel 4-stroke
11. 2-stage 2-bank Rotary Diesel ' 4-stroke'
Drawings and performance curves were prepared for all the diesel power
plants which were schemed for the primary emissions environment,
i.e. HC - 0.41 g/mile, CO - 3.4 g/mile, NOX - 1.5 g/mile.
No diesel configurations were prepared for the secondary emissions
target (HC - 0,41 g/mile, CO - 3.4 g/mile, NOx - 0 .4 g/mile)
since it was considered that no current technology diesel engine could
meet this NOy target.
-------
5-2
Introduction
The literature survey and Ricardo in-house knowledge both indicate
that a diesel engine would provide a viable power plant for a
passenger car. Given this general conclusion the study required
that the most promising diesel variants be schemed in sufficient
detail to allow a reliable assessment of their potential.
Requirements of Vehicle
For the purposes of this study two target emission levels were
envisaged. These were :-
Primary (or short term)
HC Oo41g/mile
CO 3,4 g/mile
NOx 1.5 g/mile
Secondary (or long term)
HC Oo41 g/mile
CO 3<,4 g/mile
°-4
The vehicle for these environments was to be a passenger car
(typically a 4/5 seat sedan) weighing less than 1600 kg (3500 Ib)
test weight and capable of meeting the EPA standard car performance
specifications, i.e. 0-97 km/h (0-60 mph) in less than 13.5 s,
32 - 112 km/h. (20-70 mph) in less than 15 s , capable of over-
taking a 80 km/h (50 mph) truck in less than 15 s.
To assist in the definition of the power plant a computer program was
written to allow the maximum power output of the engine , the shape
of its torque curve and the transmission system necessary to achieve
these acceleration capabilities to be calculated.
Use of this program showed that with a three-speed automatic gearbox
and the final drive ratios selected to give 136 km/h (85 mph) top speed
a bare engine power of 97 kW (130 bhp) was required at the rated speed
to satisfy the acceleration target. This implies that the vehicle has
-------
5-3
acceleration "in hand1 at its designed top speed and this is indeed true
of the diesel power plants which will have a governor cut-off at their rated
speeds in order to protect the engine. Various gear ratios and amounts
of torque back up were tried and the minimum power plant requirements
were finally selected as :-
Rated Speed 96 kW 128 bhp
.75 Rated Speed 76 kW 102 bhp
.5 Rated Speed 53 kW 71 bhp
.25 Rated Speed 17 kW 22.7 bhp
i.e. 96 kW ( 128 bhp) engine with 10% torque back up at 50% speed and
a 4:1 speed range.
Feasibility of achieving Target Emissions
While each and every type and layout of diesel engine will have problem
areas when used as a passenger car engine, the only area in which
performance targets have been set is that of exhaust emissions.
As has been explained in the report on Task 1, the great majority of
diesels emission data available refers to tests carried out on engine
test beds under the 13 mode test procedure and in the experience of
Ricardo and others it is not possible to correlate this data directly
with results which would be obtained with CVS-CH testing. It would
in fact be possible to estimate the results of cycle testing from ' maps'
of data collected under steady state conditions but such maps are not
available for many of the engines of interest.
It is estimated however that a 128 bhp indirect injection engine
fitted in a 1600 kg (3500 Ib) car would require an NOX reduction
of some 30-40% from current levels in order to achieve the
1.2 to 1.4 g/mile necessary to guarantee 1 = 5 g/milejin a production
car. Such reductions should be possible, with some penalty in
fuel economy, by retarding the injection timing but it might be
necessary in some cases to employ modulated exhaust gas
recirculation.
Ricardo do not believe however that it is possible with current
knowledge to achieve the secondary target levels of 0=4 g/mile NOX
with a diesel engine fitted to the specified vehicle. No attempt has
therefore been made in this report to rate any of the diesel engine
designs at any tighter emissions target than 1.5 g/mile NOX.
-------
5-4
GASOLINE ENGINE
While no engine configuration studies have been carried out on
gasoline engines, it was necessary to have estimates of performance
and package size for comparison with the various diesel engines.
Two gasoline engines were studied, the first being a V-8 following
current American practice both in design and performance, and the
second a 6 cylinder engine more typical of European practice.
V-8 Gasoline Engine
Specific Performance and Emissions
Using typical specific performance levels for this type of engine, i.e.
21.0 -22,0 kW/1 (0.46 - 0.48 bhp/CID) , the exercise target
power output demands a swept volume of 4.5 1 (275 CID). This
capacity in conjunction with a rated speed of 6607 rev/s (typical of
most standard American engines) produces 96 kW (128 bhp) at a
brake mean effective pressure of 6.26 Bar (92 Ib/in ) and a peak
torque of 285 N. m (210 Ib.ft) at 41. 6 rev/s gives 25% torque back
up resulting in satisfactory driving characteristics with a simple
three or four speed transmission. The torque curve for this engine
can be seen in Figure 5-1.
The precise engine specification considered is as follows :
0 97 mm (3.82") x 76 mm (3oOO") x 90°, V-8
4.5 1 (275 CID)
96 kW (128 bhp) at 66.7 rev/s (6026 Bar, 92 Ib/in2 bmep)
285 N.m (210 Ib.ft) at 41.6 rev/s (7.83 Bar, 115 Ib/in2 bmep)
Estimated weight 250 kg (550 lb)»
The above performance levels can be achieved with the engine in low
emissions build but external hang on emissions control devices are
necessary, their complexity depending on the severity of the emissions
target. The primary project targets of 1.5 g/mile NOX, 0°4 g/mile HC
and 3.4 g/mile CO in a 1600 kg (3500 Ib) inertia weight car can be
achieved by using close tolerance sophisticated carburettors in
conjunction with modulat.ad exhaust gas recycle, air injection into the
exhaust and an oxidation catalyst. All these devices can be considered
current technology and in this case the trade-off for low emissions
is well known (increased power plant weight, first cost, maintenance
-------
5-5
cost, a demand for lead free fuel, and a depreciation in vehicle fuel
economy). The load range fuel consumption curves shown in Figure 5—2
are estimated for this engine in 1.5 g/mile NOX build.
The secondary project emissions objectives of 0»4 g/mile NOX while
maintaining HC and CO levels demand the additional use of a
reducing catalyst and possible further catalytic devices (getter box)
to protect the reducing catalyst from oxygen spikes. Durability of
such systems has yet to be proven although recent developments
appear promising (Gould 3 System). Due to the need to run with
low oxygen concentration in the exhaust (less than 0.5%) it is
necessary to run with a mixture strength approaching stoichiometric
conditions with a resultant fuel economy penalty. Depending on the
exact baseline emission levels of the engine, it may be possible to
reduce EGR levels compared with the 1.5 g/mile NOX build with
resultant increases in thermal efficiency, particularly at the light load
end of the operating range. The combined effect of these two changes
is likely to be a small loss in vehicle fuel economy compared with
current vehicles.
From EPA data, a fuel consumption level of 18 1/100 km (13 mpg) has
been assumed for the gasoline powered vehicle when in 1.5 g/mile NOX
build. Reducing NOx emissions to 0.4 g/mile will probably increase this
to a level in the region of 19 1/100 km (12.5 mpg)»
From data on current American V-8 engines, engine weight will be
250 kg (550 Ib).
Using established noise prediction formulae, bare engine noise levels
on a test bed would be 95 dBA at rated load and speed» This correlates
with a drive-by noise level of 74 dBA at 15.23 m under standard
American test conditions,,
6 Cylinder Gasoline Engine
The first and most obvious difference between American and European
engines of this power range is that of specific performance. Whereas
the American V-8 gasoline engine runs to around 67 rev/s and develops
21 kW/1 (0.47 bhp/CID), the European engine runs to at least 83 rev/s
and develops a minimum of 32 kW/1 (0.7 bhp/CID) normally nearer
36 kW/1 (0.8 bhp/CID). The major reason for these vastly differing
-------
5-6
philosophies is one of economics; mainly because of taxation
European fuel prices have always been high in terms of real spending
power with a resultant continuing trend towards small capacity, high
economy, engines and cars. On the other hand, the price of fuel in
America has always been low and fuel consumption has never been a
major consideration; engine size has therefore increased over the
years to improve driveability and to some extent as a sales feature.
A total power requirement of 96 kW (128 bhp) and a torque back up of
20% to enable the prototype vehicle to achieve its target performance
could be achieved easily with a 3 litre, six cylinder engine running to
83.3 rev/s. A typical oversquare cylinder configuration of 88 mm bore
x 82 mm stroke (3.46" x 3.22") has been selected. The use of petrol
injection will ensure good distribution between cylinders and should
allow the engine to run at generally leaner mixture strengths than if
carburettors were used. Estimated performance in the form of torque
curves and load range fuel consumption loops can be seen in Figures
5-3 and 5-4. The load range consumption curves have been based on
levels which are currently being achieved by current production engines.
Note that fuel consumption levels at light load conditions (on a bmep
basis) are similar to those of the V-8 engine. At high load factors
most current V-8 engines still suffer a throttling effect (often the
intake valves or ports are restrictive in order to further improve
low speed torque) with a resultant fuel consumption penalty whereas
the more high tuned six cylinder engine with its need for improved
breathing characteristics should return low fuel consumption at these
conditions.
These performance figures and the project primary emissions targets
could be achieved using the same external hang-on control devices as
the V-8 gasoline engine, i.e. modulated exhaust gas recycle, air
injection and an oxidation catalyst.
Again, as in the case of the V-8 engine, the more severe NOX levels
of Oo4 g/mile should be achieved by the addition of a reducing catalyst
system to the above package with the same order of magnitude increase
in vehicle fuel consumption levels compared with the 1.5 g/mile NOx
build.
Overall fuel consumption levels during CVS-CH testing have been predicted
at 15.5 1/100 km (15 mpg). This results in a total fuel consumption saving
of over 10% compared with the V-8 gasoline engine.
-------
5-7
From European data, the predicted engine weight is 186 kg ( 410 lb),
a reduction of 64 kg ( 140 lb) compared with the V-8 engine.
Estimated drive-by noise levels of this engine in the prototype vehicle
are 77 dBA at 15 m (50 ft). The reason for theSdBA increase in
drive-by levels compared with the V-8 is the higher rotational speed
of the 3 litre engine.
The completed specification for this engine is :-
0 88 mm (3.46") x 82 mm (3.22") x 6 (in-line)
2.991 (183 CID)
96 kW (128 bhp) at 83.3 rev/s
232 N.m (171 Ib.ft) at 50 rev/s
Estimated weight 186 kg (410 lb)
-------
5-8
NATURALLY ASPIRATED IDI ENGINE
Specific Performance and Emissions
Of all the diesel candidates, this configuration is the closest to
current technology and as such its emissions and performance
predictions are those in which Ricardo have the greatest degree
of confidence.
It would be possible to use either a swirl chamber or a prechamber
of the form used by Daimler-Benz. The swirl chamber has a
higher air utilisation at high speeds however and its smoke level
drops more markedly as load is reduced, avoiding the 'plateau
smoke1 at low load which can be a problem with prechambers It
was decided therefore to use a swirl chamber for the design study.
The effects of vehicle inertia and engine swept volume have already
been calculated with the conclusions that this candidate in a 1600 kg
(3500 Ib) passenger car would give approximately 2 g/mile NOX
during CVS-CH testing, with the engine in optimum performance
build. Retarding the injection timing by 6 - 8 (crank) should lower
NOX levels to 1.2 - 1.5 g/mile, and it is at this retarded timing
condition that engine performance characteristics have been
calculated. No other devices should be necessary to achieve the
1.5 g/mile NOX target although a modest degree of exhaust gas
recycle might be necessary if large production compliance type
safety margins were demanded. The CO target of 3.4 g/mile should
be achieved without any modification to the engine but some fuel
pump/injector development might be necessary before the HC target
could be attained. The modifications required to reduce HC levels
are unlikely to affect engine performance adversely.
Thus, when considering the primary emissions targets, the only low
emissions device likely to affect engine performance is the timing
retard necessary for low NOX levels. This will adversely affect both
the specific power output and fuel consumption of the engine. In
order to maintain the necessarily low smoke levels for passenger car
operation, the specific power output of the engine has been taken as
20 kW/1 (0.44 bhp/CID) - normally 22.0 - 22.5 kW/1 (0.48 -
0.49 bhp/CID) for this type of engine in optimum performance build.
A fuel consumption penalty of approximately 5% compared with
optimum performance has been assumed over the load and speed
range.
-------
5-9
In the inters! of fuel economy, engine speed is normally limited
by friction and breathing considerations to a minimum piston
speed of 13.0 m/s (2600 ft/min). Thus with a rated speed of
66.7 rev/s (in order to accept the same drive line as the
gasoline engine) and a known swept volume, engine bore and
stroke are automatically defined as follows :-
Bore 0 88 mm (3.46")
Stroke 98 mm (3.86")
Figures 5-5, 5-6 illustrate predicted test bed performance levels.
As stated earlier, fuel economy and power outputs are all based
on current practice, allowing for some derate and fuel economy
penalty due to the retarded injection timing.
From the test bed fuel consumption levels a fuel economy level of 11.5-
10.5 1/100 km (20-22 mpg) has been predicted for the 1600 kg
(3500 Ib) inertia weight prototype while driving CVS^CH tests;
current lower powered European diesel passenger cars of the same
inertia weight average 10-9 1/100 km (24-26 mpg).
With current technology engines, empirical formulae have been devised
which enable overall noise levels of engines to be predicted at the
design stage with a remarkable degree of accuracy (normally - 1 dBA).
The predicted levels of this engine on a test bed is 97 dBA which
should result in Californian drive-by noise levels of 76 dBA when it
is installed in a typical American vehicle.
DESIGN NOTES
V-8 Specification Figures :- 5-7, 5-8 and 5-9
Bore 88 mm - 3.46"
Stroke 98 mm - 3.86"
Displacement 4.781 - 292 in
CR 20:1
bhp 96 kW (128 bhp) at 66.7 rev/s
bmep 7.5 Bar (109 Ib/in ) at 33.4 rev/s
power/unit piston area 0.00196 kW/mm
Piston speed 13 m/s - 2560 ft/min
-------
5-10
Crankcase Controlling Features
Crankshaft
Crank shaft pins and journals based on maximum firing pressure of
86 Bar (1250 lb/in2).
2
Crankpins 420 Bar (6100 lb/in )
Journals 283 Bar (4100 lb/in2) aS y
Connecting Rod and Piston
Small end eye 690 Bar (10,000 lb/in2)
Piston Boss 530 Bar ( 7,600 lb/in2) aS °n y
Straight cut rod - 170 mm (6.7") CRS( — = 3.48 )
Low piston height of 55 mm (2.16")» 63% bore, was chosen as
no piston cooling oil was required.
Cylinder
Press fit dry liners for engine compactness and low cost.
Cylinder centres at 1.31 x cylinder bore controlled by bearing layout.
Camshaft
Central camshaft, operating valves through push rods.
Single camshaft chosen for minimum engine width.
Water circuit
Water pump on front face of block, discharging coolant through both
banks of crankcase and cylinder heads through transfer pipes.
Cylinder Head
Figure 5-8 shows two possible port arrangements, cross flow
and uni-sided.
Due to the complexity of inboard push rod,Comet chamber and a
regular bolting pattern, a uni-sided head with up-swept inlet port
was chosen.
This gives more freedom for the cylinder head casting and reduces
the overall length of the engine.
-------
5-11
Fuel Pump
Centrally mounted pump, gear driven from camshaft. Inboard
injectors give short fuel pipe lengths.
Auxiliary Drives
Internal gear train chosen for long life and versatility. 'V belt
drives to alternator, vacuum pump and water pump. Gear driven
hydraulic pump.
Overall Engine Package
Length 758 mm - 29.84"
Height 604 mm - 23.75"
Width 692 mm - 27.25"
Box Volume 0.32 m3- 11.18 ft3
Estimated weight 320 kg - 700 Ib
Boosted - Indirect Injection
Whilst the naturally aspirated engine should make an attractive
power plant, the use of boosting will give an improvement in
specific power output and hence a reduction in bulk and weight.
Boosting with both turbocharging and with the use of Comprex have
been considered. Turbocharging does of course introduce other
difficulties. A waste gate will be essential with the turbocharger
matched to the engine at peak torque speeds. A boost responsive
maximum fuel stop, or the equivalent, is necessary to prevent
puffs of black smoke when accelerating and the time lag on
acceleration due to turbocharger inertia may require a change in
driving patterns.
The use of the Comprex should go a long way to eliminate this lag
and to give a lift to the torque curve at low engine speeds. It is
also quite easy to introduce modulated exhaust gas recirculation
without the need of additional valves and controls. On the other hand,
the device is somewhat bulky, is unproven in service, and at this
time it is difficult to estimate its first cost. There can also be
difficulties with smoke under starting conditions.
The indirect injection engine has a more severe thermal loading than
does the direct injection engine, and this is accentuated by boosting.
Oil cooled pistons will be essential for the boosted engine.
-------
5-12
The required power output is conveniently given by a six cylinder engine
of 90 mm (3.54 inch) bore and 100 mm (3.94 inch) stroke, and
comparative arrangement drawings on both in-line and Vee versions
are given in Figure 5-12. Once again, the Vee engine shows a sub-
stantial gain in length but the difference in bulk and weight is less
significant.
It is predicted that a lightly boosted six cylinder IDI will have very
similar emissions characteristics to the naturally aspirated engine
although NOX levels at the same combustion timings are likely to be
marginally higher because of the increase in cylinder temperatures
and pressures encountered in the boosted engine. These increased
NOX levels can be compensated for by means of an aftercooler to
reduce intake charge temperatures, although this approach may be
impractical in passenger car applications because of first cost,
weight and bulk penalties, or by further retarding the injection'
timing. The extra amount of retard required to arrive at a compar-
able NOx level to the naturally aspirated V-8 is difficult to quantify
but Ricardo have assumed 2 which would result in a fuel consump-
tion penalty of 2 - 3%. Therefore the total fuel consumption penalty
of a 1.5 g/mile NOX boosted IDI engine is likely to be up to 8%
compared with the same engine in optimum performance build:
a penalty of 6 - 8% has been assumed when constructing the load
range consumption curves.
In the interest of improved vehicle economy, the maximum rotational
speed of this engine has been limited to 60 rev/s.
Overall Configuration :-
Bore 0 90 mm (3.54")
Stroke 100 mm (3.94")
Rated Speed 60 rev/s
Figures 5-10 and 5-11 show detailed predicted performance levels
of this engine. It is assumed that identical performances could be
obtained from both in-line and Vee configurations.
The boosted engines will suffer a small but definite fuel economy
penalty at light load conditions (only just becoming apparent at the
extreme lower end of the load range curves) because of the power
required to drive the boosting device, and this in conjunction with
a need for a greater timing retard will result in a small vehicle
fuel economy penalty compared with the naturally aspirated engine.
-------
5-13
A level of 12-11 1/100 km (19-21 mpg) during the Federal test procedure
has been predicted. Bare engine predicted noise levels are 96 dBA at 1 m
which should result in Californian drive-by noise levels of 75 dBA when
installed in a typical American passenger car.
DESIGN NOTES
In-Line 6 Specification
Bore
Stroke
Total Cylinder Displacement
CR
bhp
bmep
power/unit piston area
Piston speed
Crankcase Controlling Features
Crankshaft
Figures 5-12, 5-13
90 mm - 3.54"
100 mm-3.94"
3.84 1 - 234 CID
19: !
96 kW (128 bhp) at 60 rev/s
9.3 Bar (1351b/in2 at 33.4 rev/s
0.00247 kW/mm2 (2.15 hp/in2)
12.0 m/sec - 2350 ft/min
Crank shaft pin and journal loadings based on maximum firing pressure,
Pmax of H3 Bar (1650 lb/in2)
Crankpins
Journals
450 Bar (6550 lb/in ) _
345 Bar (5000 lb/in2) GaS °nly
Connecting Rod and Piston
Small end eye
Piston bosses
640 Bar (9300 lb/in2) „
565 Bar (8250 lb/in2) GaS °nly
Conventional straight cut connecting rod with centres of 175 mm (6.9")
3.5 )
Piston compression height controlled by the requirement for oil cooling
with fixed crankcase located spray nozzles.
-------
5-14
Cylinders
Dry sleeve cylinder liners adopted for overall engine compactness
and low cost.
Cylinder centres of 1.23 x cylinder bore controlled by coolant
jacket requirement between adjacent barrels.
Tappets removed through side of crankcase.
Oil Galleries
Individual lubricating oil and piston cooling oil galleries. (Piston
cooling oil unfiltered, hence smaller filter requirements).
Coolant Pump Position
Coolant pump located on the front face of the block discharging
directly into the block jacket.
Cylinder Head Features
Uni-sided port lay out adopted for reasons of casting, simplicity
and manifold compactness (turbocharger and manifold on same
side of engine).
Pushrods, combustion chamber and injector positioned together
on opposite sides of ports.
Fuel Pump
High mounted distributor type fuel pump, gear driven off the main
camshaft timing train.
Mounted on same side of engine as the combustion chamber, hence
compact fuel.pipe runs.
Auxiliary Drives
Two sets of twin ' V ' belt drives. Alternator and water pump together
with individual drive to the vacuum pump.
-------
5-15
Overall Engine Package
Length
Height
Width
Box Volume
Estimated Weight
976 mm (38.41")
678 mm (26.7" )
543 mm (21.4" )
0.36 m3 (12.7 ft3)
327 kg (720 Ib)
(incl SAE.4
bell housing)
V-6 Specification
Bore
Stroke
Total Cylinder Displacement
CR
bhp
bmep
power/unit piston area
Piston speed
Figures 5-12, 5-14, 5-15, 5-16
90 mm (3.54")
100mm (3.94")
3.84 1 (234 CID)
19:1
96 kW (128 bhp) at 60 rev/s
9.3 Bar (135 lb/in2) at 33.4 rev/s
0.00247 kW/mm2 (2.15 hp/in2)
12.0 m/sec 2350 ft/min
Crankcase Controlling Features
Crankshaft and Cylinder Bank Displacement
Table of Cylinder Bank Displacement and Crankshaft
Configurations for a V-6 Engine
Cylinder Bank Displacement Firing Primary and
and Crank Configuration Intervals Secondary Balance
120 simple 3 throw crank Even
120 - 6 throw crank
pins displaced 60
Uneven
Primary balance shaft
required. Two secondary
balance shafts required
for complete balance.
Primaries in balance.
Two secondary shafts
required for complete
balance.
90 simple 3 throw crank Uneven
Primaries in balance.
Two secondary shafts
required for complete
balance.
-------
5-16
90-6 throw crank Even One primary balance
pins displaced 30 shaft required. Two
secondary shafts required
for complete balance.
o
60-6 throw crank Even Primaries in balance.
pins displaced 60 Two secondary balance
shafts required for
complete balance.
o
120 bank displacement rejected due to excessive overall engine width.
o
90 bank angle considered with a simple 3 throw crankshaft as
opposed to the 6 throw arrangement. Although uneven firing
intervals result,the primary forces are in balance. Crankcase
length of 425 mm (16.75") can be realised with angled split
connecting rods.
o o
60 bank angle considered with a 6 throw crankshaft, pins offset 60 .
This arrangement has both even firing intervals and primary balance
as shown in the above tables. Crankcase length of 507 mm (19.96")
can be achieved with angled split connecting rod.
o
90 bank angle and simple 3 throw crank selected on its merits of
minimum length and inherent primary balance.
Crankshaft pin and journal loadings based on maximum firing pressure
of 113 Bar (1650 lb/in2).
max
Crankpins 455Bar (6600 lb/in )
Journals 345 Bar (5050 lb/in2) °n y
Connecting Rod and Piston
2
Small end eye 640 Bar (9 300 lb/in )
Piston bosses 565 Bar (8250 lb/in2) S y
Angle-split connecting rod with centres of 175 mm (6.9")
( 7- = 3.5)
Piston compression height controlled by requirement for oil cooling
by fixed crankcase located spray nozzles.
-------
5-17
Cylinders
Dry sleeve cylinder liners adopted from cost considerations.
Cylinder centres of 1.45 x cylinder bore controlled by the
crankshaft requirements.
Camshaft
Single camshaft located within the centre of the Vee with pushrod
operated valves adopted for reasons of minimum cost and overall
height profile.
Tappets removed through side of crankcase.
Oil Galleries
Individual piston cooling galleries for each bank located with the
crankcase side walls. Separate lubricating oil gallery, located
adjacent to the camshaft, thus minimising filter requirement.
(Piston cooling oil unfiltered).
Coolant Pump Location
Single water pump mounted on frcnt face of block, transfers across
to the opposite bank via external pipe at rear.
Cylinder Head Features
Cross-flow porting arrangement adopted for simplicity of turbo-
charger and manifold installation* Inlets on the inside of the Vee,
exhausts on the outside.
Pushrods, combustion chamber and injector located together on
inside of Vee. (Due to wide.cylinder centres, reasonably uncongested
head of layout should result).
Fuel Pump
Distributor type fuel pump mounted within the Vee directly above the
camshaft, and gear driven off the main timing gear train.
Fuel pump adjacent to> injectors, hence compact fuel pipe runs.
Auxiliary Drives
Two sets of twin ' V ' belt drives. Alternator and water pump together,
with individual drive to vacuum pump.
-------
5-18
Overall Engine Package
Length 731 mm (28.8")
Height 693 mm (27.3")
Width 650 rr.m (25-6")
Box Volume 0. 33 m3 (11. 6 ft3)
Estimated Weight 309 kg (680 Ib)
In-Line 6 Cylinder with Comprex Blower
Figure 5-12
Identical concept and layout to the T/C version but for the
substitution of the exhaust driven turbocharger for a twin 'V
belt driven blower.
Overall engine package box volume will be reduced marginally.
Naturally Aspirated - Direct Injection
Due to its high thermal efficiency and long life, the direct injection
diesel engine is used almost universally in the heavy duty,
commercial vehicle. While the operating speed range of such an
engine is adequate for heavy vehicles, there are however difficulties
in obtaining the wide speed range which is desirable for an engine
for a light duty vehicle.
The main reason for this difficulty is to be found in the fuel/air
mixing requirements which are necessary for good combustion.
In the engine bore size under consideration, it is necessary to
employ considerable air swirl in the cylinder. This swirl is
normally generated by the appropriate design of the inlet air port but
it is found that if the fuel and air flows are matched for the high
engine speeds then the air swirl is insufficient at low speeds and the
engine power has to be reduced at low engine speeds to avoid smoke.
Conversely, if the swirl is matched at low speeds, air swirl is
excessive at high speeds and smoke will result at the high engine
speeds.
-------
5-19
It may also be necessary to employ very high injection pressures at
high engine speeds to avoid excessively long injection periods with
the nozzle hole sizes which are required to give good combustion at
low speeds.
Another difficulty with high speed direct injection engines is the
high noise level which arises from the long combustion delay period
as a result of which much of the fuel is in the cylinder before
combustion is initiated.
For these reasons direct injection engines, despite their higher
thermal efficiencies, are not employed in any high speed application,
i.e. with speeds much in excess of 50 rev/s. Such a restriction in
speed involves a considerable increase in the engine swept volume
and hence in the bulk and weight of the engine for a given power output.
As a matter of interest, several years ago, Ricardo converted a small
high speed indirect injection engine (21 swept volume) to direct
injection and carried out a limited amount of test bed and vehicle
running.
(Emission measurements were only carried out over the 7-mode
cycle but the direct injection engine had twice the NOX and CO levels
of the comparable indirect injection engine and much higher HC levels. )
The following point emerged from the result :-
On the test bed, the direct injection engine gave about 10% better
fuel economy but with 8% less smoke limited power output. The
improved fuel economy was not borne out by road tests with city
driving, however, when no difference in fuel consumption could be
measured. This was believed to be due to the higher torque of the
indirect injection engine enabling it to be driven with high gears
engaged for a longer period of operating time. A similar finding
had been reported by one of Ricardo's clients with a pair of somewhat
larger engines.
The direct injection engine was excessively noisy.
As would be expected, the direct injection engine was more
sensitive to injection timing and required more advance over the
speed range. Provided that one is not aiming for very low emission
levels it is indeed not necessary to provide any speed advance for
an indirect injection engine.
-------
5-20
The major problem with direct injection is however set by the
emissions targets. It is unlikely that an NOX level of 1.5 g/mile
could be achieved on production engines, even with a considerable
amount of exhaust gas recirculation, and the retardation of exhaust
timing which will also be necessary will give excessive hydrocarbon
levels and the fuel consumption will be little if any different from an
indirect injection engine of the same emission levels.
While it might be possible to use an exhaust catalyst to reduce the
hydrocarbon levels, there are likely to be difficulties in developing
a system with adequate catalyst activity especially under light load
running when the hydrocarbon emissions will take the form of blue/
white smoke due to misfire.
Due to these problems and the severe derating which would be necessary
under retard timing conditions to avoid excessive smoke, no proposals
have been put forward for a naturally aspirated direct injection engine.
Boosted - Direct Injection
The use of boost will of course give a higher specific output and hence
will reduce engine bulk and weight.
It will also enable the injection timing to be retarded without a heavy
reduction in power since the exhaust smoke levels will be much
reduced.
The turbocharged engine will have similar problems of turbocharger
lag and transient smoke as the turbocharged indirect injection engine,
but the latter may be treated again by means of a boost controlled
maximum fuel stop. The Comprex again offers possibilities.
While it seems unlikely that a boosted direct injection engine could
meet the emission targets, it is clearly desirable to include such
an engine in the overall comparison of types.
The specific power output of the boosted DI has been assumed to be
exactly the same as that of the boosted IDI. This has allowed the
same engine capacity to be used but in the case of the DI, virtually
'square1 (0 93 mm x 94 mm) cylinder dimensions have been selected.
This choice was not affected by combustion chamber considerations
but was selected simply to show the other extreme to the under-square
IDI configuration.
-------
5-21
The resultant weight reduction by adoption of a square configuration
was estimated to be in the order of 18. kg (40 lb) for in-line variants
and 9 kg (20 lb) for the vee engines. A noise penalty of approximately
1 dBA will be incurred with this stroke:bore ratio, when compared with
a DI of the same proportions (0 90 mm x 100 mm) as the IDI engine.
Overall predicted noise level of the bare engine is 104 dBA with a
resultant predicted drive-by noise level of 82 dBA.
Estimated performance levels of this candidate are shown in Figures
5-17 and 5-18. Identical performances should be obtained from both
in-line and vee configurations.
The load range consumption curves in Figure 5-18 have been constructed
assuming the injection timing has been retarded 10 in an attempt to
achieve low NOX levels with a resultant fuel economy penalty of 10%.
From heavy duty data it is estimated that this timing retard in conjunction
with 10% EGR should result in NOx levels in the order of 2.5 g/mile.
This amount of EGR should not affect fuel economy levels.
HC levels at these retarded timings are likely to be very high and it
is assumed that an efficient oxidising catalyst would be necessary in
order to achieve 0.1 g/mile. Such a system is unlikely to affect
fuel economy levels.
From the predicted fuel consumption curves it is estimated that vehicle
fuel economies of 11-101/100 km (21-23 mpg) are likely to be returned
during the Federal test cycle.
DESIGN NOTES
In-Line 6 Cylinder Specification Figures :- 5-19
Bore 93 mm (3.66")
Stroke 94 mm (3.70")
Displacement 3.841 (234 CID)
CR 17.5:1
bhp 96 kW (128 bhp) at 60 rev/s
bmep 9.3 Bar (135 lb/in2) at 33.3 rev/s
power/unit piston area 0.00233 kW/mm2 (2.03 hp/in2)
piston speed 11.3 m/s 2230 ft/min
-------
5-22
Crankcase Controlling Features
Crankshaft
Pins and journals based on a maximum firing pressure of 124 Bar
(1800 lb/in2).
2
Crankpins 450 Bar (6500 lb/in )
Journals 290 Bar (4200 lb/in2) 3S °n Y
Connecting Rod and Piston
Small end eye 450 Bar (6500 lb/in2) Gas Qnl
Piston bosses 580 Bar (8400 " "
Angled split connecting rod with centres of 169mm (6. 65")
(- = 3.6)
Piston height controlled by piston oil cooling requirements.
Cylinder
Press fit dry liners for engine compactness and low cost.
Cylinder centres controlled by liner and water jacket clearances
and chosen at 1.25 x cylinder bore.
Camshaft
Single block mounted camshaft, with pushrod operated valves,
adopted for minimum engine cost and overall height profile.
Oil Galleries
Individual lubricating and piston oil cooling galleries.
Piston cooling oil is unfiltered.
Water Circuit
Water pump discharging directly into front face of cylinder block.
Cylinder Head Features
Uni-sided port arrangement chosen for simplicity in the cylinder head
casting, and ease of directing manifold pipes to and from the turbocharger.
-------
5-23
Fuel Pump
High mounted distributor pump gear driven from camshaft giving short
fuel pipe lengths.
Auxiliary Drives
One set of twin 'V belt drives to alternator and water pump. One set
of 'V belts to vacuum pump.
Hydraulic pump mounted on front gear cover and gear driven.
Overall Engine Package
Length 939 mm - 36.95"
Height 667 mm - 26.25"
Width 543 mm - 21.375"
Box Volume 0.34 m3 - 12 ft3
Estimated weight 310 kg - 680 Ib.
COMPREX BLOWER SYSTEM
The comprex blower was mounted on the side of the block and twin belt
driven. The manifolding is simplified and gives a slimmer engine and
a reduction in box volume.
The auxiliary drives were repositioned, and the position for the vacuum
pump was changed to block mounted and driven from the camshaft.
V-6 Specification Figures :- 5-19 and 5-20
Bore 93 mm (3.66")
Stroke 94 mm (3.70")
Displacement 3.841 (234 CID)
CR 17.5:1
bhp 96 kW (128 bhp) at 60 rev/s
bmep 9.3 Bar (135 ib/in2) at 33.3 rev/s
Power/unit piston area 0.00233 kW/mm (2.03 hp/in2)
Piston speed 11.3 m/s (2230 ft/min)
-------
5-24
Crankshaft Controlling Features
Crankshaft
Pins and journals on a maximum firing pressure of 124 Bar
(1800 lb/in2).
2
Crankpins 460 Bar (6700 Ib/in0) r
,2* *jas onlv
Journals 325 Bar (4700 lb/in ) y
Balancing requirements as stated for V-6 Comet engine
Connecting Rod and Piston
2
Small end eye 700 Bar (10,200 lb/in )
Piston bosses 580 Bar ( 8,400 lb/in ) y
Angle cut split connecting rod with centres of 174 mm (6.86")
-------
5-25
Fuel Pump
Centrally mounted punap, gear driven from camshaft. Inboard injectors
give small fuel pipe lengths.
Auxiliary Drives
Internal gear drive used for long life and versatility. ' V' belt drives to
alternator, vacuum pump, and water pump. Gear driven hydraulic pump.
Overall Engine Package
Length
Height
Width
Box Volume
Estimated weight
.708 mm - 27.87"
654 mm - 25.75"
674 mm - 26.50"
0.31 m3T 11 ftS-
300 kg - 660 Ib.
Comprex blowers were not deemed a viable proposition due to
complexity of manifold pipes to and from the blower.
-------
5-26
Loop Scavenged, Two Cycle, Indirect Injection
While it is difficult to obtain a reasonable performance and fuel
economy with a two cycle indirect injection engine, the lower NOX
emissions with indirect injection make it desirable to assess the
possibilities.
An engine of this type was manufactured for a short period by the
Turner Manufacturing Company in the 1950's (689) and the published
performance data from this engine has been taken as the basis for
prediction.
The resulting engine has six cylinder and a bore of 99 mm (3.89")i
a stroke of 114 mm (4.50") and runs to 45 rev/s. This gives a high
value of 4450mm rev/s (10500 in. rpm) for the bore and speed
parameters which control breathing on a two cycle engine and therefore
gives a penalty in fuel economy. While it would have been possible
to reduce the parameter by employing eight cylinders of 85 mm (3.3611) ,
the very wide scavenge port belts necessary with loop scavenging give
such wide cylinder spacing that the engine would have been excessively
large and heavy although of Vee form.
It is difficult to predict the emission levels of this power unit. Due
to the poor fuel economy and the need therefore to burn more fuel
the levels could be higher than for an equivalent four cycle engine.
It has been assumed however that the NOX primary target can be met
but there is no guarantee of this. Hydrocarbon levels will almost
certainly give problems.
Estimated vehicle fuel consumption levels during CVS-CH are 13-11.5
1/100 km (18-20 mpg).
Predicted noise levels are impossible to quantify with any degree of
confidence but it is estimated that combustion noise levels from this
engine under Californian drive-by conditions will be around 75 dBA.
The arrangement drawing of the engine is shown in Figure 5-22.
The effect of the wide cylinder spacing on engine length is clearly
shown and in Figure 5-21 the performance curves show the rather
poor fuel economy to be expected.
-------
5-27
DESIGN NOTES
V-6 Specification Figures :- 5-22 and 5-23
Bore 99 mm (3.89")
Stroke 114mm (4.50")
Engine Displacement 5.261 (321 CID)
CR 19:1
bhp 96 kW (128 bhp)
bmep 4.7 Bar (68 Ib/in ) at 33.3 rev/s
power/unit piston area .00207 kW/mm2 (1.82 hp/in2)
Piston speed 10.2 m/s (2000 ft/min)
Crankcase Controlling Features
Crankshaft
Pin and journal loadings based on maximum firing pressure of 110 Bar
(1600 Ib/in ).
2
Crankpin 470 Bar (6800 Ib/in ) Gas only
Journals 307 Bar (4450 Ib/in2 )
For crankshaft configuration and pin offset Table 1 (next page) has been
assembled.
From the results a 90 'V engine with a 30 crankpin offset was chosen.
A 120° 'V engine is too wide for serious consideration, and a 60° 'V
is too tall.
Scheme Fig.37 shows the crankshaft layout with flying webs between
adjacent pins.
Connecting Rod and Pistons
Small end eye 745 Bar (10,800 Ib/in2)
Piston bosses 515 Bar ( 7,500 Ib/in2)
Angle split connecting rod with centres of 228 mm (8.96").
<±- - 4)
This high — ratio is caused by piston skirt and the porting of inlet
and exhaust gases. Another effect on this ratio is the angle split
connecting rod.
-------
5-28
Firing
Residual Primary
Couple
120
90
75
* 90
75
90
90
60
0
0
30
15
20
15
60-60-60
90-30-90
75-45-75
60-60-60
60-60-60
70-50-70
75-45-75
0
0
0.53 vert.
0.53 vert.
0.77 vert.
0.35 vert.
0.27 vert.
* This configuration gives even firing and primary couple of low
magnitude that should be removable through correct positioning
of flexible mountings.
-------
5-29
Cylinder
Thick walled, press fit dry liners for compactness.
Cylinder centres fixed at 1.7 x bore due to porting through sides of
adjacent barrels. (In-line version of this engine was rejected for
the reason of excessive engine length).
Oil galleries
Individual lubricating oil and piston cooling oil galleries.
Piston cooling oil is unfiltered.
Water Circuit
Water pump on front face of block, discharging coolant through
both banks of crankcase and to cylinder heads through transfer
pipes.
Cylinder Head Features
Due to large cylinder centres no inlet or exhaust ports; simple
individual head was adopted.
Pre-combustion chamber is therefore centrally mounted.
Manifolding
The Rootes type blower is centrally mounted inside the ' V' and
gear driven from the crankshaft through idler gears.
An inlet manifold casting is attached to both banks of the ' V'
and connected directly to the blower.
Exhaust manifolding is a simple conventional cast iron pipe.
Fuel Pump
The distributor fuel pump is driven through the Rootes type blower,
and the high position allows small fuel pipe lengths,
Auxiliary Drives
Two sets of ' V' belt drives. Alternator and water pump together with
an individual drive to the vacuum pump. Hydraulic pump is gear driven.
-------
5-30
Overall Engine Package
Length 803 mm (31. 60")
Height 692 mm (27.25")
Width 648 mm (25.5")
Box Volume 0.359 m3 (12.7 ft )
Estimated weight 340kg (7601b).
Uniflow Scavenged Two Cycle Direct Injection
The considerable success of the General Motors series of truck
engines of this type suggests that it is a suitable candidate for study.
For good fuel economy, it is desirable to keep the product of
speed and bore size to a modest value and a similar value to
that used on the GM71 series engines was adopted. A value of
10.2 m/s. (2000 ft/min) was chosen for the piston speed.
While there are difficulties in running fuel injection equipment at
engine speed on two cycle engines, the running speed of 45 rev/s
on this size of engine should not give rise to difficulties but a pump
injector system could be used as an alternative.
A direct injection combustion system was chosen due to the known
difficulties of using indirect injection on a two cycle engine and still
obtaining an acceptable fuel consumption. Ricardo do not believe
however that it will be possible to meet the primary NOX target
and the hydrocarbon levels are also in doubt.
As with the loop scavenge engine, the low rated speed of this engine
along with modest bmep levels results in this engine being heavy and
because of the higher torque levels necessary to obtain the target
power at this speed, a heavier transmission may be needed.
Predicted noise level of this engine under drive-by conditions at a
distance of 15.24 m (50ft) is likely to be 75 dBA.
A fuel consumption of about 12.5 - 11.5 1/100 km (19-21 mpg) under
CVS-CH test conditions is predicted assuming optimum performance
conditions. There is insufficient data to enable Ricardo to predict
any performance parameters for this engine in low emissions build.
The comparative arrangement drawings for In-line and V-6 cylinder
versions of an engine of 83 mm (3.28") bore and 114 mm (4.5") stroke
are given in Figure 5-25 with comparative performance data in Figure
5-24.
-------
5-31
DESIGN NOTES
In Line 6 Specification Figures :- 5-25 and 5-26
Bore 83 mm (3.28")
Stroke 114mm (4.50")
Cylinder displacement 3.74 1 (228 CID)
CR 17:1
bhp 96 kW (128 bhp) at 45 rev/s
bmep 6.9 Bar (100 lb/in2) at 33.3 rev/s
power/unit piston area 0.002 9 kW/mm2 (2.52 hp/in2)
Piston speed 10.2 m/s (2000 ft/min)
Crankcase Controlling Features
Crankshaft
Crankshaft pin and journal loadings based on maximum firing pressure
Pmax of 138 Bar (200° lb/in2).
Crankpin 451 Bar (6550 lb/in2)
Journal 342Bar (4970 lb/in2)
Connecting Rod and Piston
Small end eye 7 60 Bar (11,000 lb/in2)
Piston bosses 570Bar ( 8,300 lb/in2)
Angle split connecting rod and cap with centres 254 mm (10")
(± = 4.48 )
Piston compression height controlled by requirement for oil
cooling from fixed crankcase located spray nozzles.
Cylinders
Thick dry sleeve cylinder liners adopted for minimum cost.
Cylinder centre of 1.3 x cylinder bore controlled by water jacket
and air belt between adjacent cylinders.
Camshaft
Single crankcase located shaft with pushrod operated exhaust valves
adopted for minimum engine cost and overall height profile.
Tappets removed through top face of block i.e. head must firstly
be removed from the block.
-------
5-32
Oil Galleries
Individual lubricating oil and piston cooling oil galleries (piston oil
unfiltered, hence smaller filter requirement).
Coolant Pump Location
Located on front face of block, discharging directly into the crankcase
water jacket.
Blower and Fuel Pump
Rootes type blower mounted on the front gear casing and gear driven
from the camshaft. Blower air is directed through the crankcase into
the middle ducting around the cylinder.
The fuel pump is driven through the blower.
The exhaust manifold is a simple cast iron 'Rake1 casting.
Cylinder Head Features
The cylinder head is simplified by only containing the exhaust porting
and injector bosses. Due to cylinder centres of 1.3 x bore a one piece
head was chosen.
Auxiliary Drives
Two sets on twin 'V belts. Alternator and water pump together, with
individual drive to the vacuum pump. Hydraulic pump is directly
mounted on gear casing and gear driven.
Overall Engine Package
Length 946 mm (37.25")
Height 874 mm (34.38")
Width 575 mm (22.63")
Box Volume 0.47 m3 (16.77 ft )
Estimated Weight 365kg (800 Ib)
V-6 Specification Figures :- 5-25 and 5-27
Bore 83 mm (3.28")
Stroke 114mm (4.50")
Engine Displacement 3.74 1 (228 CID)
CR 17:1
bhp 96 kW (128 bhp) at 45 rev/s
bmep 6.9 Bar (100 lb/in2) at 33.3 rev/s
power/unit piston area ,0029 kW/mrp2 (2.52 hp/in2)
Piston speed 10.2 m/sec. (2000 ft/min.)
-------
5-33
Crankshaft Controlling Features
Crankshaft
Pin and journal loadings based on a maximum firing pressure of
138 Bar (2000 lb/in2).
Crankpin 448 Bar (6500 lb/in2)
Journals 290 Bar (4200 lb/in2)
For crankshaft configuration see table prepared for the loop
scavenged 2-stroke V-6 engine.
For this engine application with the possibility of a shorter engine
casing with cylinder centres of 1.6 x bore, a 90° 'V bank angle
with uneven firing but with the advantage of no vertical primary
couples .WHS selected.
(The cylinder centre ratio is lowered when compared to the loop
scavenged engine, by the removal of exhaust gas ducting from the
crankcase to cylinder head).
Connecting Rod and Piston
o
Small end eye 760 Bar (11,000 lb/in )
Piston bosses 570 Bar ( 8300 lb/in2 ) °n y
Angle split rod, with centres of 254 mm (10") ( — = 4.48).
This high •—- ratio is caused by porting of the inlet gases, piston
skirt, angled split rod and small diameter bore.
Piston height controlled by oil cooling requirements.
Cylinders
Thick walled, press fit dry liners for compactness.
Cylinder centres fixed at 1.6 x bore, controlled by bearing require-
ments and porting of inlet gases, through cylinder block around barrels.
Camshaft
Single shaft centrally crankcase mounted with push rod operated
exhaust valves.
Oil Galleries
Individual lubricating oil and piston cooling oil galleries.
Piston cooling is unfiltered.
-------
5-34
Water circuit
Water pump mounted on front face of block, discharging coolant
directly through both banks of crankcase and to cylinder heads
through transfer pipes.
Cylinder Head Features
Due to large cylinder centres and with only the exhaust porting in
the casting, a simple individual head was chosen.
Auxiliary Drives
Two sets of 'V belt drives. Alternator and water pump together
with an individual drive to the vacuum pump. Hydraulic pump is
gear driven.
Internal gear drive was chosen for long life and versatility.
Overall Engine Package
Length 704 mm (27.7")
Height 685 mm (27.0")
Width 768 mm (30.211)
Box Volume 0. 387 m3 ( 13. 66 ft )
Estimated Weight 354 kg (780 Ib)
UNCONVENTIONAL ENGINES
Compound. Direct Injection Engine
Compounding of diesel engines for the attainment of improved power/
weight ratios and thermal efficiency implies the use of air compressors
and exhaust driven turbines with different drive arrangements between
the three components. In the ultimate form the diesel engine is used
solely to produce high temperature, high pressure gas for use in the
gas turbine which forms the power unit - this arrangement is generally
termed a gas generator.
The possible operational variables which can affect the overall perform-
ance and efficiency in varying degrees are very numerous and the
prediction of performance to include the effect of these variables is
very complex.
-------
5-35
The main variables are listed below :-
Engine type - 4 stroke
2 stroke (valve in head, loop
scavenged-opposed piston)
Engine compression ratio
Boost pressure
Cylinder pressure limits
Air fuel ratio
Valve tinning
Turbine efficiency
Compressor efficiency
Use of charge cooler - charge cooler efficiency
Engine /turbo components drive arrangement
Thermal loading limitations
Bore/stroke ratio
Although there is a fairly extensive literature study of the subject,
the little practical experience that exists is concerned with large
engines for which this configuration is most suited. The performance
predictions are also concerned with large engines and are mostly
confined to maximum output conditions without consideration for
part load operation.
In order to arrive at fairly broad estimates for the performance of
a small engine, the available information has been used and modified
where possible to take account of the scaling down of the components.
Engine Configuration
The simplest form of compound engine is that in which the output is
taken from the diesel engine with the turbo-compressor unit geared
to the crankshaft. The gearing has to cope only with the difference
in power between the compressor and the turbine, and there is no
requirement for power equilibrium between these components
such as exist in normal turbocharging.
-------
5-36
With a gas generator system all the power output is taken from the
turbine through reduction gearing. For the relatively low power
requirements of the engine under study the small turbine would
have a very high rotational speed. Although this arrangement has
the theoretical advantage of a high-low speed torque compared with
the compound engine, it is not considered to be feasible for the
present application in view of these high x otational speeds and
the unknown performance of small turbines at high expansion
ratios.
Engine Type
From the available data a 2-stroke engine is seen to offer the best
thermodynamic performance and power/weight ratio with an opposed
piston type being superior to the valve in head form. It also permits
greater control over the division of power between the diesel engine and
the turbine compared with a 4-stroke engine. However, although the
2-stroke engine is basically simpler it is considered that the 4-stroke
might be preferred since it has much lower thermal stresses and
piston heat flow and has greater volume of experience behind its
development.
Estimated Performance
The values tabulated below have been derived from data predicted
for large engines with the following assumptions :-
Boost ratio 4:1- Inlet pressure 4.12 Bar ( 60 lb/in2 abs)
Charge cooler effectiveness 80%
Charge cooling water temperature 85 C
Compressor efficiency 74%
Turbine efficiency 82%
Cylinder pressure limit 207 Bar (3000 lb/in )
The final column shows the engine swept volume for 96 kW (128 bhp)
at an engine speed of 50 rev/s.
-------
5-37
bmep 2 Thermal Displacement
Charge cooled Bar / Ib/in Efficiency * (litres)
2-stroke gas
generator 19 3 280 .29 1.0
2-stroke
compound 22.4 325 .35 0.85
4-stroke gas
generator 24,2 350 .34 1.57
4-stroke
compound 27.6 400 .39 1.38
Uncooled
2-stroke gas
generator 17 -2 250 .32 1.1
2-stroke . , . .-.
compound 19.0 275 .3 1.0
4-stroke gas
generator 18 3 265 .36 2.08
4-stroke
compound 20.7 300 .41 1.63
* Based on fuel energy.
It is pointed out that these figures must be regarded as approximate
and are intended to show the order of magnitude only and the relative
position of the different operating conditions.
Specific output can be increased.significantly with increasing manifold
pressure but will be accompanied by higher thermal and mechanical
stresses. It is for this reason that a pressure ratio of 4:1 has been
chosen in the above table.
From this table it is seen that an engine capacity of 1.0 - 2.0 litres
would be required to produce 96 kW (128 bhp) at 50 rev/s depending
on the operating conditions. There are a number of factors to be taken
into account however in the application of this data to a small engine.
-------
5-38
Firstly, the smallest cylinder size dealt with in the preparation of these
results was 2.8 litres (170 in^), involving a scaling down of about 1/5.
This will obviously have some direct adverse effect on the engine
performance but the most serious problem would be associated with
the compressor and turbine. The dimensions of these would be very
small to match the low gas flows and would have high rotational speeds.
It is difficult to estimate the efficiencies of these components at high
pressure ratios but they would be lower than those used in the predicted
performance and would therefore lead to some increase in engine size.
This problem could be alleviated by the use of two stage operation but
this would add to the complexity and bulk of the power unit. For the
sake of simplicity it would be preferable to dispense with charge cooling
in automotive applications - this also results in an increase in engine
size but has the benefit of increased thermal efficiency. The maximum
cylinder pressure of 20.7 Bar (3000 lb/in2) used in the predicted
performance is very high. If this is reduced to a more reasonable
level of 13.8 Bar (2000 lb/in ) by a reduction in pressure ratio, the
performance will be penalised by possibly 20%.
Alternatively, it could be lowered by retarding the injection timing which
would depress the engine performance , but part of this loss would be
recovered in the turbine due to the increase in exhaust temperature.
This method would also be beneficial from the point of view of NOX exhaust
emissions. It is probable that from the above considerations the engine
size would be increased by 25 - 30% above those quoted in the table.
Choice of Engine Configuration
Assuming that a 4-stroke engine would be preferred on account of lower
thermal and mechanical stresses and superior thermal efficiency, and
that it would be non-cooled to reduce complexity and bulk, the choice
remains between a gas generator and a compound arrangement. The
engine swept volumes would be about 2.6 and 2.0 litres respectively,
both of which fall in the range of large production naturally aspirated
diesel engines. With a bore/stroke ratio of 1:1 the corresponding
bore sizes are 94 mm and 86 mm and piston speeds at 50 rev/s of 9.4
and 8.6 m/s. Although the compound engine is smaller and requires
a lighter gear connection to the engine crankshaft, it has poor low speed
performance and would therefore require a more complex transmission.
It is the question of part load behaviour that poses the greatest difficulty
in prediction owing to the interaction between the three main components,
and the lack of a suitable computer program.
-------
5-39
Emissions
The behaviour of the compound engine from the point of view of exhaust
emissions is also difficult to predict. Exhaust smoke should not be
a problem as it should be possible to operate at an air :fuel ratio giving
low smoke values e.g. 25:1, although low speed, low load boost
characteristics will tend to cause some smoke.
The high combustion temperatures will produce high NOX levels but
these could be reduced with charge cooling. On the other hand, this
would have an adverse effect on HC as would the low compression
ratios essential with high intake pressures, particularly at light load.
These features can only be established by actual engine measurements.
Conclusions
The general conclusions are that a 4-stroke diesel engine of about 2£ litres
would be required to produce 96 kW (128 bhp) at 50 rev/s. Although there
is a vast amount of experience with diesel engines of this size with natural
aspiration, extensive development would be required to produce a robust
enough engine configuration and to establish the performance of high speed
small turbo-units. It is considered that the complexity and durability of
the high reduction gearing would present a major disadvantage and expense.
Starting and light load hydrocarbons will undoubtedly be a problem and
ignition assistance may be necessary at idling and light load as well as
when starting.
Figure 5-29 gives the arrangement drawings for a four cylinder engine
of 93 mm (3.66 inch) bore and stroke, running at 50 rev/s. The good
fuel economy but poor torque curve shape are clear from the performance
data in Figure 5-28.
It should perhaps be stated that in a limited study of this kind, it is
not possible to optimise the design for all variants of compound
engine as it would be possible to design for lower boost ratios.
This would improve the torque curve shape, help starting and idling,
and possibly reduce NOX levels but at the expense of a bulkier and
heavier engine with a worsening of fuel economy.
-------
5-40
DESIGN NOTES
In-Line 4 Cylinder Specification
Bore
Stroke
Cylinder Displacement
CR
bhp
power /unit piston area
Piston speed
Turbine Specification
Wheel tip diameter
Wheel tip length
Exducer diameter
Speed
Compressor Specification
Wheel tip diameter
Pressure ratio
Speed
Crankshaft Controlling Features
Crankshaft
Figures :- 5-29 and 5-30
93 mm (3.66")
93 mm (3.66")
2 = 54 litre (154 CID)
13:1
96 kW (128 bhp) at 50 rev/s
0.00346 kW/mm2 (3.05 hp/in2)
9.3 m/sec (1820 ft/min)
63 mm (2.48")
6.5 mm (0.256")
48 mm (1.89")
2220 rev/s at engine rated spee<
68 mm (2.68")
4:1
2220 rev/s at engine rated speec
Crankshaft pin and journal loadings based on maximum firing pressure
pmax °f 138 bar (2000 Ib/in2)
Crankpins
Journals
Connecting Rod and Piston
Small end eye
Piston bosses
460 Bar (6550 lb/in2) _
342 Bar (4960 lb/in2) GaS °nly
700 Bar (10,000 lb/in )
584 Bar ( 8,300 lb/in2)
•
Angle split connecting rod and cap with centres 175 mm (6.9")
= 3.75 )
r
Piston compression height controlled by requirements for oil cooling
from fixed crankcase located spray nozzle.
-------
5-41
Cylinders
Dry sleeve cylinder liners for overall engine compactness and
minimum cost.
Cylinder centres of 1.23 x cylinder bore controlled by coolant
jacket requirement between adjacent cylinders.
Camshaft
Single crankcase located shaft with pushrod operated valves adopted
for minimum engine cost and overall height profile.
Tappets removed through side of crankcase.
Oil Galleries
Individual lubricating oil and piston cooling oil galleries (piston
cooling oil unfiltered, hence smaller filter requirement).
Coolant Pump: Location .
Located on front face of the block discharging directly into the block
water jacket.
Fuel Pump
High mounted distributor type fuel pump gear driven off.the main timing
gear train located at the front of the engine.
Cylinder Head Features
Uni-sided port layout adopted for casting simplicity and manifold
compactness, turbine on same side.
Pushrod and injector positioned together on opposite side to the ports,
and same side as fuel pump^ hence compact fuel pipe runs to injectors.
TVurbine Drive
Gear train located at flywheel end of engine.
Crankshaft located driving gear T driven through the flywheel via
'cush drive'clutch mechanism.
-------
5 - 42
Gear
Crankshaft driving
Compound Set - Engine side
Turbine side
1st Idler
2nd Idler
Turbine driven
Speed (rev/s)
50
300
300
1200
1200
2220
Pitch Line Velocity
47 m/s
47 m/s
189 m/s
189 m/s
189 m/s
(154 ft/s)
(154 ft/s)
(620 ft/s)
(620 ft/s)
(620 ft/s)
189 m/s (620 ft/s)
Note : - Two idlers necessary to mount the turbine outside the
flywheel profile.
Turbine Location
Separate gear train casing mounted between engine block and flywheel
housing. Turbine housing bolted onto rear face of casing and compressor
housing to the front, thus resulting in a compact exhaust and inlet
manifold arrangement.
Auxiliary Drives
Two sets of twin 'V belt drives. Alternator and water pump together
with individual drive to vacuum pump.
Overall Engine Package
Length
Height
Width
Box Volume
802 mm (31.58")
685 mm (26.96")
584mm (23")
0.32 m (11.3 ft3)
(incl. SAE 4
flywheel housing)
Estimated weight 305 kg (670 Ib)
Rotary (Wankel) Engine
With the considerable interest in the rotary gasoline engine over the
past fifteen years or so, a somewhat smaller interest has been
displayed in the rotary diesel engine. One company $ Rolls Royce,
have gone quite a long way but have now abandoned their efforts
as they do not believe that a commercially attractive engine can be
produced. Due to the difficult combustion chamber shape at diesel
compression ratios, the only engines that have run reasonably
successfully have used two stage compression.
-------
5-43
I
The engine tested by Rolls Royce had a larger 'diameter' first stage
than second but the design schemed for this report has equal
'diameters' with longer first stage rotors to give a lower package
height.
There must, in view of the experience of all who have worked in
this field, be grave doubts on the ultimate success of developing
a rotary diesel engine. Fuel consumption, hydrocarbon and carbon
monoxide emissions, smoke and durability must all have a question mark
over them unless there be a breakthrough in further developments.
Geometric Limitations - Operating Cycle
The 2.3 * or 4 stroke geometry is undoubtedly the most attractive
form for diesel operation. The only alternative would be the 1.2
or 2 stroke form, which, whilst giving scope for much higher
compression ratios, has insufficient space within the rotor to
accommodate an adequate bearing.
Compression Ratio
In order to achieve acceptable cold starting and light load operation
the diesel engine must operate with a high compression ratio; in the
range 18-22:1 depending upon the combustion system used in the type
of application under consideration.
The 2.3* Wankel mechanism with no combustion chamber in either
trochoid or rotor can reach a ratio of about 25:1 with a K ratio
.generating radius of around 14-15. However, as the K ratio
v eccentric radius >
increases and eccentric radius is reduced, the diameters of the
phasing gears are reduced in order to maintain a 1.5:1 ratio and this
limits the diameters of the main and rotor bearings. Consequently a
K ratio greater than 10 would not be acceptable for a diesel engine
from the point of view of bearing loads. This limits the theoretical
maximum compression ratio to 21:1 without any combustion chamber.
(* 2.3 refers to number of trochoid lobes and number of rotor faces),
-------
5-44
Combustion Volume
Whether direct or indirect combustion systems are used, piston engine
experience has shown that the volume outside of the actual chamber
must be kept to a minimum. In the case of small piston engines, about 25%
of the clearance volume is in unavoidable clearances. To maintain this
situation in the rotary engine would mean accepting a compression ratio
of about 5:1. Clearly a compromise must be accepted in which a reasonabl<
compression ratio can be achieved with a minimum of dead space in the
combustion system. Since the compression ratio is going to be well below
that normally used in high speed diesel engines it will be necessary to
utilise two-stage compression and expansion.
As a starting point one might select a first stage compression ratio
of 2:1 and second stage of, say, 10.5:1. This will mean that at
IDC 50% of the total clearance volume will be in the dead volumes.
Again to achieve piston engine conditions a 4:1 first stage and 5.25:1
second stage would be required.
Combustion Chamber Type
The choice here is between direct injection (Dl) or indirect injection (IDI).
It is known that in the DI chambers, air swirl is essential to give good
combustion when wall impingement occurs. Attempts to use DI chambers in
rotary diesel engines have failed because it is quite impossible to generate
adequate air motion in the Wankel rotary engine (unpublished Ricardo
experience).
It is therefore concluded that the combustion system must be of IDI type.
Although the rotary engine may have special requirements as regards
chamber geometry, at this stage, we can only predict combustion
performance from well tried systems such as the Ricardo Comet on
piston engines. IDI systems such as the Comet depend for their mixing
gas motion on having the .bulk of the clearance volume divided into two
by a throat or orifice through which the air is compressed and combustion
gases are expanded. The proportions in the prechamber and main volume
over the piston are fairly critical to performance and should ideally be a
50/50 distribution for the Comet. For prechamber systems such as
Caterpillar, the prechamber volume can be only 25% of the total.
Clearly there must be some dpearture from this ideal because with the
selected compression ratios there is already 40% of the clearance
volume over the rotor before any chamber volume is incorporated
in the rotor similar to those normally present in the piston crown of
conventional engines. At the very least a channel must be provided
in the rotor face to allow gas transfer from trailing to leading
volumes at TDC.
-------
5-45
Piston engine experience shows some fall off in air utilisation at
a volume proportion of 40-60, i.e. 40% in the prechamber, at 20%
in the prechamber the loss of air utilisation is about 33% and the
loss of brake performance about 40%. Presumably the precombustion
chamber will suffer loss in this way although no data is available.
This is the area for maximum growth and where maximum development
should be applied to the diesel rotary engine. We will assume that
some improvement is possible and say at this stage that due to loss of air
utilisation in the Wankel engine the indicated performance will be down
by 20% on the piston engine.
We must consider the relative motoring losses of the rotary and piston
engines. Compared on the basis of Wankel rotational speed = 1.5 x
piston engine rotational speed, there would appear to be an increase of
some .75 bar mep in the case of the Wankel engine. This difference
probably contains increases in heat, mechanical and pumping losses
resulting from the 2 stage arrangement and which are difficult to
separate.
Taking a typical small automotive IDI performance and making allowance
for the above two differences, the performance is predicted as in Figure
5-31, and to achieve the required power the engine will have two high
and two low pressure stages. The swept volume of the low pressure
stages will be 5.35 litre (326 in3) per lobe. The engine will give two
working strokes per revolution of the eccentric shafts and with an output
shaft reduced in the ratio to 3:2 the working cycle frequency will be
as for a 6 cylinder in-line engine. Top eccentric speed will be
100 rev/s and output shaft 66.7 rev/s.
The fuel pump repetition rate of 100 rev/s is too high for one element.
Therefore each injector will be fed by two lines of a modified 4 cylinder
pump siamesed together within the pump.
The layout in Figures 5-32, 5-33 and 5-34 is suggested having a length
of about 752mm(29.6"),width of 6.79mm (26.7") and height about 508 mm
(20") without auxiliaries. There are ample take-off points for driving
auxiliaries, the alternator could be contained within the engine envelope
being on the other end to the fuel pump.
-------
5-46
Low Heat Loss ( ' Adiabatic ' ) Engine
Preliminary calculations were carried out to assess the likely
improvement in fuel economy which would result from large
reductions in the heat transferred from the charge during the
working cycle.
Cycle calculations were carried out using a simplified engine
performance simulation program in which heat is fed into the
cycle by means of an assumed heat release diagram shape. It
was assumed that the heat release diagram shape was unaffected
by the reduction in the heat loss which was simulated by increasing
the assumed temperatures of the combustion chamber walls.
While the results are not of a high order of accuracy, they are
believed to give a reasonable picture as to likely trends . Reducing
the heat transfer during the power stroke by 70% gives a reduction
in fuel consumption of only 4. 5% with a resulting increase in maximum
cycle temperature from 1460°C to 1510° C (2658°F to 2788°F) and
an increase of exhaust release temperature from 600 C to 705°C
to 1300°F).
These calculations assume that the volumetric efficiency is unchanged
which is likely to be an optimistic assumption and any reduction in
this will of course increase the fuel consumption due to a lowering
of mechanical efficiency. The increase in exhaust temperature may
lead to problems with exhaust valves but on the other hand, this
increase could lead to an increase in power recovery in a compound
engine.
NOX emissions are likely to be increased, perhaps dramatically, due
to the strong temperature dependence of NOx formation at these levels
but hydrocarbon levels should be reduced.
Provided that some practical method of reducing heat losses can be
developed, the 'Adiabatic1 principle could be applied to most of the
engines under consideration; on the other hand, for the reasons given in
the preceding paragraphs , the overall result may not be too attractive
and no further analytical studies: have been made at this time.
-------
5-47
Low Ratio, Ignition Assisted Diesel
There has been a considerable interest in recent years in the possibility
of developing a low compression ratio diesel engine. The interest has
been initiated by, amongst others, the following considerations :-
1. For reasons of acceptable cold starting and the prevention of
high speed light load misfire and blue smoke after a cold start,
the compression ratio of current high speed diesel engines is
appreciably higher than the compression ratio for optimum
efficiency. A reduction in ratio would therefore give an
improved fuel consumption.
2. A reduction in peak cycle pressures could give a reduction in
engine scantlings and hence of weight and cost.
3. Reduction in peak cycle temperatures should give reduction in
emissions.
While there is sone doubt as to the validity of the second and third of
these arguments , it would undoubtedly be of interest to run such an
engine to examine its performance and potential.
The major difficulty of such a test arises from the difficulty of
guaranteeing reliable and consistent ignition and it is useful to
consider the alternatives ;-
1. Heater plugs as presently used for a starting aid would have
an unacceptably high power consumption even if they did
give consistent ignition. For cold starting , additional aids
would also be necessary.
2. Inlet air manifold heaters would have a similar power loss if
electrically heated and even if fuel heated, would give an
unacceptable increase in fuel consumption, especially at
light load when the need for an ignition aid will be at its
greatest.
3. Spark ignition must be the most attractive aid due to its low
power consumption. The use of a high energy surface discharge
plug may be necessary however and a repetitive spark system
of the kind used in the TCCS system may be essential. Even
with this , there is no guarantee that consistent and regular firing
will result and high hydrocarbon levels due to misfire may prove
insuperable.
-------
5 - 48
4. Exhaust gas recycle to heat the cylinder charge under light load
conditions could possibly be developed as an effective ignition aid
but Ricardo are already concerned about the effect of high EGR
flow rates on lubricating oil and engine durability when it is applied
at high loads. Increasing the range during which this is used may
aggravate this potential problem.
5. The use of exhaust back pressure to eliminate blue smoke
problems (by increasing cylinder temperatures) is known to
be effective in conventional diesel engines and would
undoubtedly be effective in improving ignition conditions in
a low ratio engine. However, as in the case of inlet manifold
heaters it works by increasing the amount of fuel being burnt
and like intake charge heating is in direct opposition to a high
economy prime mover.
6. Charge air heating by means of the hot exhaust gas could possibly
be employed once the system has warmed up but will not provide
a solution to high hydrocarbon levels due to misfire on start up.
7. Variable compression ratio pistons would limit cylinder pressures,
but can only be applied to quiescent direct injection combustion
systems, would not give improved economy at light load, are
expensive and it is difficult to arrange piston cooling for high
ratings.
In view of these difficulties it was thought that a low ratio, ignition assisted
diesel engine was not a proved and practical proposition at this time but
it is suggested that for experimental purposes it would be worth
attempting to find a solution of the ignition problem in order that the
future potential might be investigated.
-------
mm
Bore
in
Stroke
in
kW
Power ,
hp
Swept litre
Volume in^
Weight j^
Box m3
Volume ft^
Specific power- kW/1
swept volume hp/in
r\
Power/unit kW/mm
piston area hp/in^
Specific power- kW/m^
box volume hp/ft^
Specific weight kg/1
- swept volume Ib/in1-^
Specific weight kg/kW
- power Ib/hp
V8
Gaso-
line
97
3.82
76
3.00
96
128
4.5
275
250
550
:
21.3
0.47
-
-
• 55.6
2.00
2.60
4.30
IL6
Gaso-
line
88
3.46
82
3.22
96
128
2.99
183
186
410
:
32.11
0.70
-
-
60.2
2.19
1.87
3.12
V8
NA
IDI
88
3.46
98
3.86
96
128
4.78
292
320
700
0.32
11.2
20.2
0.44
0.0020
1.70
302
11.4
66.9
2.41
3.31
5.47
V6
TC
IDI
90
3.54
100
3.94
96
128
3.84
234
309
680
0.33
11.6
24.4
0.55
0.0025
2.15
291
11.05
80.9
2.91
3.22
5.31
In Line
6 TC
IDI
90
3.54
100
3.94
96
128
3.84
234
327
720
0.36
12.7
24.4
0.55
0.0025
2.15
265
10.1
85.6
3.08
3.41
5.62
V6
TC
DI
93
3.66
94
3.70
96
128
3.84
234
300
660
0.31
11.0
24.4
0.55
0.0023
2.03
306
11.6
78.5
2.83
3.12
5.16
In Line
6 TC
DI
93
3.66
94
3.70
96
128
3.84
234
310
680
0.34
12.0
24.4
0.55
0.0023
2.03
280
10.7
80.7
2.91
3.22
5.31
V6
2 Stroke
Loop IDI
99
3.89
114
4.50
96
128
5.25
320
340
760
0.36
12.7
18.3
0.40
0.0021
1.82
265
10.1
65.7
2.37
3.59
5.93
In Line 6
2 Stroke
Uniflow DI
83
3.28
114
4.50
96
128
3.74
228
365
800
0.47
16.8
26.0
0.57
0.0029
2.52
201
3.0
98.4
3.55
3.78
6.25
4 Cyl.
Compound
DI
93
3.66
93
3.66
96
128
2.52
153
305
670
0.32
11.3
38.0
0.83
0.0035
3.04
299
11.3
120.6
4.34
3.17
5.23
2 Rotor
2 Stage
Rotary
-
-
96
128
:
227
500
0.26
9.2
-
:
368
13.9
:
2.36
3.91
SUMMARY TABLE OF MAJOR CHARACTERISTICS OF
POWERPLANTSXONSIDERED IN ENGINE CONFIGURATION STUDY
-------
5-49
CVS-CH Fuel Consumption and Emissions Estimation
For rating purposes it was necessary to estimate the expected emission
levels and fuel consumptions when the target vehicle, fitted with the various
engine combinations, is driven over the LA4 cycle.
It is clear that the final accuracy cannot be better than that of the input data
and there are considerable difficulties in giving a definite forecast of the
results to be expected from some of the candidates. The emission levels
to be expected from the compound engine and the rotary engine are clear
examples of this difficulty as is the fuel consumption of the rotary engine.
It was decided therefore to make detailed calculations for the case where
the highest degree of accuracy for the input data could be expected and to
estimate the likely variations from these calculated values for the other
engines.
The detailed calculations were carried out using a cycle synthesis computer
program which has been developed by Ricardoo The program is flexible in
terms of the driving cycle which it will accept but for the purposes of these
calculations, the LA4 cycle was of course employed. The cycle is fed in
in terms of vehicle speed against time and the program calculates the engine
performance which is necessary to drive the cycle.
Engine performance maps showing exhaust constituents, brake specific fuel
consumptions and volumetric efficiency, all against brake mean-effective
pressure and engine speed have also been entered in and the program then
uses this data to calculate the gross fuel consumption and the mass of the
individual exhaust pollutants in 1 second steps for the whole cycle.
As CVS-CH test data is only available, albeit for lower power/weight ratios
and for somewhat lighter vehicles, for the naturally aspirated indirect
injection engines, it was decided to use the computer program for this
engine. The predicted engine performance data and emission maps given
in Figs. 5-35 and 5-36 were used for this purpose and the following estimates
resulted :-
Fuel consumption 11.3 1/100 km (20.7 miles/gallon)
NOX 1.17g/mile
HC 0.46g/mile
A second set of calculations were run using typical ppm levels achieved
from engines of this cylinder size when in optimum performance build.
(Figures 5-37 and 5-38).
-------
5-50
In this case for the 97 kW engine the following emission levels were
predicted :-
NOX 1.9 g/mile
HC 0.7 g/mile
Vehicle fuel economy levels were not calculated with the engine in this
build but a CVS-CH fuel consumption approaching 10.5 1/100 km
(22 mpg) would be expected.
CO levels were not calculated in either case as experience has indicated
that provided the engine is operated at acceptable smoke levels, the
CO levels will be below the target value of 3. 4 g/mile.
The arguments employed in the prediction of the values to be expected
for the other engines are given in the Power Plant Rating section of
this report.
For comparison purposes predicted emission maps were also prepared
for the boosted DI and IDI engines. These are shown in Figures 5-39,
5-40, 5-41 and 5-42. Due to limitations in the availability of air flow
data for turbocharged engines, however, no computer calculations were
carried out for the turbocharged engines.
-------
5-51
ESTIMATED TORQUE CURVE FOR
FIG. No 5-1
Drg Nc D26O84
Date £8-7-74
097x76 mm V8 GASOLINE ENGINE IN LOW EMISSIONS BUILD
BARO 7£0 mm Hg.
A.I.T. 20°c
BARE ENGINE
;>Tbmep-bar
^TT/TfiUlLp FOR. LOW EM 155
- 1 /L' 4-1-14-U.i-l I I I i-L-li-l I I I I i |-rr
5g/mile NOx-CLOSE
OLERANCE SOPWISTICAJED
RBURETTOR MODULATED
AJR INJECTION^:
OXIDATION CATALYST
0-4 a/mi le NOx - AS ABOVE
REDUCIN6 CATALYST
FUEL CONS.-a/KW-h
ENGINE SPEED rev/s.
-------
5-52
ESTIMATED FUEL CONSUMPTION CURVES FOR A
097x76mmV8 GASOLINE ENGINE IN
LOW EMISSIONS Q-Sg/mile NOx) BUILD.
BARE ENGINE
A.I.T. £C°C.
-•5
abo:-t
11
JJi
1 I <
HI
4
4
EL CONSUMPTION LEVEL6 IN*
:: : 0-4q/mile NOx. BUILD ARE LIKELY TO BE 4
• i • • | ! • • •¥• • i - i • • i • • . i t"i ' 11 (tt : r11 • i Trnrt li
--r UP TO 5% HIGHER THAN SHOWN HERE
FIG. No 5-2.
Drg No D 2 40 85
Dat. Zfl-7-74
mm Hg.
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5-53
ESTIMATED TORQUE CURVE FOR A
0 88 x 8£mm G CYLINDER "EUROPEAN TYPE"
GASOLINE ENGINE IN LOW EMISSIONS BUILD
FIG. No 5-3
Drg No D26O86
Date £8-7-74
BARE ENGINE
A.I.T. 20°C.
BARO 7GOmm
"805=^0^
~7&
60-
50'
i:
n
j ioo6T
t >(*
fcr
- -» n
-1-n
t t
"IU '
BUILD FOR LOW EMISSIONS:! f
4-
NOx.- PETROL INJECTION
: j; MODULATED EGR., AIR INJECTION.,,
1 ' '/^vi OATirMci r A.T&V\/eV i ' t ;• • i 4^—1
OXIDATION CATALYST!
;;::T::-U::;:1;;;:1—UuOiuI
^ IT;"~ri ~rr:T~r~^"irrrrn': :!'!'!! i
0-4gj/miiVN'Ox -AS ABOVE-f RE
LAJALYST" SYSTEM"^
,4
i.l-
li
I T | f
t i-h
!1
SPEC. FUEL CONS. (WOT) -fl/KW -h
tr
t; :20 : i!
:;T!i '• ':
i.t M
i •)
rev/min.
aooo
30OO
DUCINGt
-t i H-
T I •*-* '
IT-
;li:-:
4000
::t
-»t-
;1
t;f
i-..
JJ:::
P
r1 I
•4oo :>
300;$
•aoo
te
\-^ H
j: io-!|.
•' T. 1.11 H :
40
50
6O
il
ENGINE SPEED rev/S.
-------
. 5-54
FIG. No 5-4
ESTIMATED LOAD RANGE CONSUMPTION CURVES FOR Dr9 No D£4O87
^ &|X CYUNDER 0 e8x 62mm "EUROPEAN TYPE" Dat* 2fl""7-74
| GASOLINE ENGINE IN l-Sg/mile NOx BUILD.
PETROL INJECTION. EGR, AIR INJECTION + OXIDATION CATALYST.
500
sooo rev/min)
3rev/s (
h 44-- -4-4 ---J4---4-; -44. 4---
Hii. 14 t. *-.-l. l J. jj i 4. u+-, i_|-f- 4-f-H-H'tt
5O rev/s (300O rev/m n)j
33'3 rev/s (abob rev/min]
£•7 rev/s (looo rev
r
i
*
01
i
j
U)
u.
o
UJ
a.
JIN:.!, l}J!i mi'
t r5.}mln
-------
5-55
ESTIMATED PERFORMANCE CURVE FOR
FIG. No. 5-5
Drg. No. DE6O74
Date 28-7-74
NATURALLV ASPIRATED 0 88 x 98mm V8 COMET V ENGINE
BARE ENGINE. LOW EMISSIONS BUILD (l-5g/mile NOx IN A 3500 Ib
PASSENGER CAR) AIT aO°C BARO 7GO mm Hg.
BO-= -fcQ-*
-STEADY STATE
SMOKE - °/o OPACITVrl
BUILD FOR LOW EMISSIONS c
INJECTION TIMING RETARDED 6-
i.; :n ::;. i!::: i:.:; i
tHC'&
-+-H-H ! l.M
ECTION EQUIPEMENT OPTIMISED FO
FULL LOAD FUEL CO
ENGINE SPEED - rev/5.
-------
LOAD RANGE FUEL CONSUMPTION CURVES FOR
FIG. No. 5-G
Drg. No OSLGfO75
Oate £8 -7-74
NATURALLY ASPIRATED #88x98 mm V8 COMET V ENGINE.
BARE. ENGINE. LOW EMISSIONS BUILD (l-Sg/mile NOx IN A
350O Ib PASSENGER CAR) AIT 2O5c. SARO 7^O mm Hg.
I NOTE: INJECTION TIMING RETARDED
Hill!.: J:: :; U:_,.L;I;_LLLI...-; ii LJ-
G-8° FROM OPTIMUM PERFORM
SETTING
li.Sprev/a;
33-3 rev/
bmep- bar
-------
HYDRAULIC
PUMP
6ELL HOUSING
STARTER MOTOR
ftORE:-
STUOKE:-
6.H.P:-
»ME.P :-
r». MAX :-
(88mm)
384." (96mm)
I2B(3> 40OO rpm
»9lb/in» (
-------
i
U1
00
POETS
DEC. N? 30S4/1
E.RA. DIESEL IMPACT' PBO.JECT
PHASE a ENGINE CONFIGUgATION STUDY
N.A. VS I.O.I DIESEL ENGINE
COMPABISON OF CeOSSFLOW/UNISIDED CYLINDEB HEADS
u>
i
-------
crur-CBi mtc
y
I£AI
L ._^ I
t
Ul
DRG.N? 3O84/2
E.PA. DIESEL IMPACT' PROJECT
PHASE HI ENGINE CONFIGURATION STUDY
N.A.-V8-ID.I. DIESEL ENGINE CROSS-SECTIONAL
ARRANGEMENT.
-------
ESTIMATED PERFORMANCE CURVE FOR BOOSTED
FIG. No. 5-IO
Drg. No. D SL&O7G*
Date 28-7-74
SIX CYLINDER 0 90x 100mm I.D.I. ENGINE IN LOW EMISSIONS
BUILD Q-Sq/milfc NOx IN A 2>5OO Ib PASSENGER CAR)
BARE ENGINE AIT 2O°C
ioo
BARO 7(iOmm Hg.
ENGINE SPEED- rev/S.
-------
5-61 /(.
FIG. No 5"H
ESTIMATED LOAD RANGE FUEL CONSUMPTION CURVES Dr9 No OefcO77
FOR A BOOSTED SIX CYLINDER 0 9O x IOO mm I PI
ENGINE IN LOW EMISSIONS BUILD (l-5q/mile NQx
IN A 350O Ib PASSENGER CAR.)
Date
£8-7-74
BARE ENGINE
rSOO
AIT ZO'C.
BARO 7&O mm Hg.
: INJECTION TIMING RETARDED
cr
I
U)
Z
J
UJ
D
U.
•
o
111
Q.
-------
731
GK5 Ak/IP
ALTERNATOR
-SAE *bCs^^
|&aoibC3Q9nq')
&ELL HOUSlMO
-VACUUM PUWP
ORG.No. 3O84/15
STARTER MOTOR
BLOWER ON IN LINE 4, CYLINDER
TUR8OCHARGED IN LIME G> CYL.INDER
&o AMP ALTERNATOR
E.PA. DIESEL 'IMPACT' PROJECT
PHASE HI ENGINE CONFIGURATION STUDY
6 CYLINDER TURBOCHARGED I.D.I. DIESEL ENGINE
NSTALLATION DRAWING SHOWING COMPARISON BETWEEN IN LINE 6/ve
Tl
P
u>
i
io
-------
•Dependent on cylinder liner
and water jacket
DEG. N9 3O84/I4
E.RA. DIESEL IMPACT' pecuecT
PHASS a EMGINF. CO^FIGUgATION STUPV
-------
? 30S4/II
E.=A. ggSEL IMPACT PSQj;CT
PHASE Z EMS'NJc CCSJ'iGUEATlCN 5TLOY
LAYOUT c^ \/6_ g;O° 5A_NK, ANGLE!
n
71-9
-JJSSL
H
-
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i?.i
iUL
1
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'•I i
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T'J
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p
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1
(3.54") E
i 73
i? R91
r
c:
ore Dia.
i
(&*
l-MD
'
Crankshaft Arrangment with
Straight-C~ut Connecting Rod
Block Lenath (18.15"! 461
»tOJD-OT CCTXrlTKi TOO
N? 3O84/I2
ERA. DIHSEL 'IMPACT' P2OJECT
PHASE in ENGINE COMFIGU^TION STUDY
PBELIMINA2Y LAYOUT OP V<5 TIJESOCHAEGSD IMESEL ENGINE
WITH 3 THEOW
-------
Dependent on Crankshaft Design
OEG. N9 3064/13
ERA. DIESEL 'IMPACT' PBOJECT
PHASE ffl ENGINE CONFIGURATION STUDY
TUBBOCHABGED LDI. DIESEL ENGINE
CBOSS & LONGITUDINAL ABgANGEVIENT DBAWINGS.
-------
5-67
ESTIMATED PERFORMANCE CURVES FOR
FIG. No .5-17
Drg. No Oa6O78
r- • as -7-74
BOOSTED 0 95x94mm SIX CYLINDER D.I. IN MINIMUM EMISSIONS BUILD.
(=2: E-5q/mile NOx IN A 35OOIb PASSENGER CAR)
ENGINE BARE AIT 2O°C. BARO 7GO mm
130-
J20-
,
iio-
....
100-
90-
807
HI :
70-
60-
50-
40*
_.-• ..
30-
T '
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ENGINE SPEED- rev/s.
-------
i ESTIMATED LOAD RANGE
FOR A BOOSTED 093x 94
FIG. No 5.19
FUEL CONSUMPTION CURVES Drg. No DZ6O79
mm SIX CYLINDER D.I. ENGINE
D«e 28 - 7-74
MINIMUM EMISSIONS BUILD. (^ a-5q/mile NOx IN A 35OO Ib
PASSENGER CAR)
'Tr
H-- •
.>7-
f
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NOTE: INJECTION
B-IO°t.
MEEDED
• • • • I • • •
: : : 50 rev/s
- - --(•-•- —
::::b3-3
^*^--H
.... i ....
. . .
....
rev/
•~«—
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. : . . : ; .
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4-S4-
337
,—• CO*» PUOW »Of»T
/ CVLIMOCK HCAO
HYOQkUUC
PUMP
(47S)
Ul
I
in
DRC.Mo
E.PA.OIESEL IMPACT PROJECT
• PHASE HI ENGINE CONFIGURATION STUDY
96 KN 6 CYLINDER TURBOCHARGED D.I. DIESEL ENGINE
INSTALLATION DRAWING SHOWING COMPARISON BETWEEN IN-LINE 6/V6
-------
in-line.Crankshaft —
u.
I
' 1.25D I
Dependent on CylinderLiner
I and Watei Passage Core
Crankshaft
L
-4
1.46D
Dependent on Bearing Requirements
DEG. N9 3084/4
E.P.A. DIESEL IMPACT PROJECT
PHASE m ENGINE CONFIGURATION STUDY
6 CYLINQER TUeBOCHARGED D.l. DIESEL ENGINE
CROSS-SECTIONAL A&I2ANGEMENT DBAWING
171
I
Ui
I
-------
5-71
• ESTIMATED PERFORMANCE CURVE TOR
STROKE LOOP SCAVENGE IPI DIESEL ENGINE.
OPTIMUM PERFORMANCE CONDITIONS.
NOx > 2Lg/mile HC > O-4g/mUe..
FIG. No 5-2.1
Drg. No D£6O8O
Date aS-7-74
fEHI.100
rwjtrMiFntt
ENGINE SPEED- rev/S
-------
394
0275")
vn
i
PLATK CLUTCH MAY M
MCtuiMO Oft &AE*3 MIL
HOU«M« WITH UAMtt WA.CLUTCM
»TROKC
«SO" (114 mm)
I2» 9> 2TOO rpm
! (4-7»AR)^000 R.P.M.
CAPACITY:- 321 C.I.D. O-2ta Litre*)
•OX VOL :-
UT WI*MT:- •
E.PA DIESEL'IMPACT PROJECT
PHASE IE ENGINE CONFIGURATION STUDY
96kWLOOP SCAVENGE-I.D.I. 2 STROKE DIESEL ENGINE
INSTALLATION DRAWING
DOG No 30B4/7A
H
5
u»
rO
tO
-------
1.7D
Dependent on Exhaust and
ntake Passage at Cylinders
J
DUG. N? 3084/6
E.PA. DIESEL IMPACT PROJECT
PHASE m ENGINE CONFIGUBATION STUDY
V6-ID.I- ZSTBOKE LOOP SCAVENGE DIESEL ENGINE
CBOSS-SECTIONAL AEEANGEMENT DRAWING
u>
to
-------
5-74
ESTIMATED PERFORMANCE CURVE FOR
STROKE THROUGH SCAVENGE D.I. DIESEL ENGINE
OPTIMUM PERFORMANCE CONDITIONS.
NOx > Zg/mile,1 HC >0-4g/mile.
FIG NO. * 5-24
Drg No DZfrOSl
Date Z8-7-74
t)4OOO
4 < 4 •(• I- m •»•
-------
4 O^hMP .
V6 - 2 STROKE THROUGH SCAVENGE 0 I DIESEL ENGINE
aat, 28?
.. (" »3 .^, fna.'
IN LIME 6>CYUNpm-2»T>OKe-THBOLiaM
SCAVINOC D.I MKML
OR &«e " i aei
WTI LARGER Ol*
CLUTCH.
3084/83
E.PA. DIESEL'IMPACT' PROJECT
PHASE HI ENGINE CONFIGURATION STUDY
3» L.trts)
VEE
13-613 Pt-3
(•387m3)
7BO Ib
(354K3)
IN LJNE
60X VOL.- l«-77 ft 3
(•47mS)
EST. WEIGHT:- 800 Ib
( 365 K9)
s
-------
.DEG.N? 3O64/I6
E.PA. OESEL MPACT PROJECT
PHASE ffl ENGINE CONHGUCATION STUDY
6 CYL. - 2 STBOKE - THROUGH SCAVENGED- D.I. DIESEL ENGINE
CEOSS & LONGITUDINAL AE2AMGEMENT DRAWING
u
m
i
-------
VI
I
DBG. N9 3O84/2I
E.P.A. DIESEL 'IMPACT PBOJECT
PHASE m ENGINE CONFIGURATION STUDY
V& 2 STBOfcE THBOUGH-SCAVENGED D.I. DIESEL ENGINE
CROSS & LONGITUDINAL ABBANGEMENT DRAWING
-------
5.78
ESTIMATED PERFORMANCE CURVE FOR
FOUR STROKE 093x93 COMPOUND Dl. ENGINE.
FIG No. 5-28
Drg No DE6O82
Date ^6-7-74
OPTIMUM PERFORMANCE CONDITIONS
NOx > 2g/mite HC > 0-4g/mile.
mr
I30r
t- tttnitti
SPEC. FUEL
-------
602
(SI-SB )
feOAMP
ALTERNATOR
Ul
I
-a
3-fefc'OIA (93mm)
3-fefe* (93mm)
128 (5) 30OO rpm.
222 Ib/in2 (l53&AR)^!)3OOOrpm
2OOOIb/in2 (3T5API)
153 CID (2'5 Litres)
ll-3cuft- (0-32m»)
feTO Ib (305 l<9 )
WHEEL DIA'.- 2-H8*(S>3mm)
TIP wiOTH:- O-25fe" (fe Smm)
EXOUCER WA:- I 89*(46mm)
SPEED :-
RATED &PEED
COMPRESSOR:- WMCELDIA:- 2-«>a*
PRKSSURE RATIO.'-
CAPACITY :-
BOX. VOL :-
EST. WEIGHT :-
ORG. No 3084/18
96 kN
EPA. DIESEL 'IMPACT' PROJECT
PHASE m ENGINE CONFIGURATION STUDY
COMPOUND IN-LINE 4 CYLINDER D.I DIESEL ENGINE
INSTALLATION DRAWING
CP
-------
1.2S x Bnro _
Dependent on Cylinder Liner
_and Water Jacket
.Nodal Turbine.
Gear Train
DEG. NP 3O84/n
"Cush Drive
E.P.A. DIESEL IMPACT PBOJECT
PHASE m ENGINE CONFIGURATION STUDY
COMPOUND IN-LINE 4 CYLINDER D.I. DIESEL ENGINE
CEO5S & LONGITUDINAL AggANIGEMENT DRAWINGS.
i
00
o
-------
5-81
ESTIMATED PERFORMANCE CURVE FOR
TWO STAGE ROTARY DIESEL ENGINE.
OPTIMUM PERFORMANCE CONDITIONS
NOx > 2g/m»le HC > O-4g/mile
FIG. NO 5.31
Drg. Nc DZ6O83 '
Date Z8-7-74
SPEC. FUEL CONS.-Q/kW-h
F.NfilME SPEED - TCV/S.
-------
LOW PRESSURE ROTOR
DRG No. 3084/10
E.PA DIESEL IMPACT PROJECT
PHASE El ENGINE CONFIGURATION STUDY
SGKKI 2 STAGE/ 2 BANK I.D.I ROTARY DIESEL ENGINE
BMP -
6 M E =>
p
CAPACITY
BOx VOL - 9-ibcof
EST WElGrlT • 5>OO B (
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Ul
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PRO. N? 3084/19
E.PA. DIESEL IMPACT PgQjECT
PHASE a ENGINE CONFIttUgATION STUDY
j STAGE/ta^NK 1.01. BOTAgY OieSEL BNGINE
PgEUMINAgY CBOSS SECTIONAL AggANGtMEMT DgG.
-------
I ,
•H
DEG. N9 3O84/2O
E.PA. OES6L IMPACT PgQJECT
PHASE Bl ENGINE CONPK3UBATION STUDY
2 STAGE II BANK IQI. BOTAIZY DIESEL ENGINE
PgELIMINAEY LONGITUDINAL ABBANGEMENT D&G.
-------
5-85
NATURALLY ASPIRATED 9& kW V8 IDI ENGINE
PREDICTED HC ppm LEVELS IN LOW EMISSIONS
NOTE ;
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NATURALLY ASPIRATED 96 kW V8 IDI ENGINE
Drg. N«. D 14094
DM* lft-7-74
PREDICTED NOx ppm LEVELS IN LOW EMISSIONS BUILD
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-------
5-87
NATURALLY ASPIRATED 9C> KW V8 IDI ENGINE
PREDICTED HC ppm LEVELS USING CURRENT
FUEL INJECTION EQUIPMENT.
'^ NOTE: CVS-CH COMPUTER SIMULATION PROGRAM
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5-89
TYPICAL HC EMISSIONS FROM A
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OPTIMUM PERFORMANCE BUILD.
FIG NO. 5-33
Drg. No. D£GO88
Date 28-7-74
Xo ENGINE RATED SPEED
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-------
5-90
TYPICAL NOx EMISSIONS FROM A
LIGHTLV BOOSTED CONVENTIONAL D.I. ENGINE IN
OPTIMUM PERFORMANCE BUILD.
FIG. Nr 5-4O
Drg No D2
-------
5-91
TYPICAL HC EMISSIONS FROM A
LIGHTLY BOOSTED COMET V ENGINE IN
OPTIMUM PERFORMANCE BUILD.
FIG. No. 5-41
Drg. No DE609I
Date aS-7-74
ENGINE RATED
-------
TYPICAL NOx EMISSIONS FROM A LIGHTLY
BOOSTED COMET V ENGINE
IN OPTIMUM PERFORMANCE BUILD
FIG r i 5-42.
Drg No D2.6O9O
Date £8-7-74
7o ENGINE RATED SPEED
-------
6-1
SECTION 6
POWER PLANT RATING
The quantitative comparison of different power plants was one of the major
aims of the study. This section describes the rating methodology which was
developed, how it was applied to the eleven potentially viable power plants
and the results of that application.
The rating methodology involves the identification of performance aspects or
requirements for a power plant. Each aspect is given a weighting indicating
its relative importance. Each power plant is then given a 'rating1 indicating
how well it met the requirements of each performance aspect. Multiplication
of each 'rating1 by the appropriate 'weighting1 arid summation of the products
then gives a numerical 'overall rating1. Comparison of the 'overall ratings'
allows the power plants to be compared on a numerical basis*
The individual ratings for each performance aspect for each power plant are
listed and the reasons for these ratings are discussed.
The 'overall ratings' indicated that the gasoline engines were some 10%
superior to the diesel engines for an average light duty application in the
primary emissions environment (HC - 0.41 g/mile, CO - 3.4 g/mile, NOX -
1.5 g/mile.) The results, however, indicated that greater weighting on fuel
consumption and certain other economic and durability aspects would make the
diesel at least as attractive as the gasoline engine. Only the four-stroke
IDI diesel power plants were viable, all the other diesel variants being incap-
able of meeting the emissions targets.
Since it was considered that none of the diesel power plants could meet the
secondary emissions targets (HC - 0.41 g/mile, CO - 3.4 g/mile, NOX -
0.4 g/mile), no attempt was made to carry out the rating exercise for this
secondary environment.
-------
6-2
Introduction
One of the major aims of the study was that a methodology should be derived
which would allow a quantitative assessment of the relative merits of various
power plants for light duty vehicle use. Although the study was concerned
only with a comparison of gasoline and diesel configurations the methodology
was developed so that it could be applied to any liquid hydrocarbon power
plant and thus should be of value in other similar studies.
The advantages of such a methodology are that its application will allow a
direct quantitative rating of various power plants and that it should also be
possible to identify those factors and aspects which render a particular
power plant suitable for a particular duty. The second advantage allows an
assessment of changes in a particular area as well as highlighting areas
worthy of effort to make a particular configuration more suitable for use in
a given environment.
Approach of Rating System
The fitness of any power plant for a given duty is a combination of the excell-
ence with which it meets various performance aspects or requirements and the
relative importance of those individual aspects. The application of any rating
system must thus involve five stages:
l) Identification of performance aspects
2) Estimation of relative importance of those aspects
3) Estimation of how well a particular power plant meets a performance
aspect
4) Assessment of overall merit of that power plant
5) Comparison with overall merit of other power plants
Performance Aspects
The fitness of a power plant for light duty use can be assessed under the
following broad headings:
a) Emissions
b) • Package • (Size, weight, etc.)
c) Costs
d) Nature (Driveability)
e) Others (Convenience and minor safety aspects)
-------
6-3
Unfortunately these headings are much too broad for a detailed assessment
of different power plants and a more detailed list of performance aspects
was drawn up such that the various aspects would cover all facets of light
duty vehicle power plants when fuelled by liquid hydrocarbons. The indivi-
dual performance aspects are as follows:
1. Smoke
2. Particulates
3. Odour
4. NOX
5. HC
6. CO
7. S02
8. HC reactivity
9. Evaporative emissions
10. Misc. emissions
11. Drive-by Noise
12. Package volume
13. Package weight
14. Fuel economy
15. Fuel
16. Vehicle first cost
17. Maintenance cost
18. Startability
19. Hot driveability
20. Cold driveability
21. Torque rise
22. Durability
23 . Coolant heat loss
24. Fire risk
25. Idle noise
26. Vibration and torque recoil
It will be seen that although the performance aspects are generally those
studied in the literature survey certain others have been added so that the
rating methodology would be complete.
Relative Importance of Performance Aspects
If the summated importance of all the performance aspects is 100 then each
individual aspect can be assigned a numerical importance or''weighting'.
For this study a committee of six Ricardo personnel was selected and they
were asked to assign a 'weighting1 to each of the performance aspects. Al-
though the committee members were selected so that each member had
experience and knowledge of both the emissions field and the American auto-
-------
6-4
motive situation it was felt that the corporate results may have carried a
bias and it was thus decided to try to eliminate this bias. A further group
of eighteen persons was then selected (again with experience of emissions
and American automotive conditions) and this group of people was asked to
assign weightings in isolation. A comparison of this survey and committee
results (Fig. 6-1) shows that although there was general agreement between
the two, discrepancies were sufficient to indicate that there may well have
been slight committee bias and thus the survey figures were adopted for
this study. These weightings are shown in Table 1 below.
TABLE 1
Final Weightings Used In Study
Aspect Weighting
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
Smoke
Particulates
Odour
NOX
HC
CO
so2
HC reactivity
Evaporative Emissions
Miscellaneous Emissions
Noise (Drive-by)
Package volume
Package weight
Fuel economy
Fuel cost
Vehicle first cost
Maintenance cost
Startability
Hot driveability
Cold driveability
Torque rise
Durability
Heat loss
Fire risk
Idling noise
Vibration and torque recoil
4.48
2.14
4.48
3.92
3.99
3.61
3.48
1.83
1.60
0.98
6.32
2.61
2.59
12.20
5.40
4.65
4.35
4.85
4.48
3.52
1.98
4.80
2.18
3.55
3.83
2.18
-------
6-5
How Well a Performance Aspect or Requirement is Met
It is necessary that a rating scale be devised so that a quantitative assess-
ment of how well a particular power plant meets a given performance aspect
can be made. The above list of performance aspects shows that although
some aspects could be quickly assessed in a numerical fashion, many others
are essentially qualitative and any rating scale should be able to cover all
aspects.
As expected, some difficulty was experienced in relating a purely subjective
impression to a linear quantitative scale, but after some consideration the
following system was adopted as giving the numerical scale several easily
relatable, subjective key points, the numbers without definition being an inter-
polation of the surrounding merit definitions.
Merit rating scale
0 Totally unacceptable
1
2 Bad
3
4 Poor
5 Acceptable
6
7 Good
8
9 Best practical
10 Perfect
Assessment of Overall Merit of the Power Plant
The rating system evolved allows an immediate quantitative assessment of the
overall merit of the power plant and this is accomplished by multiplying each
aspect 'rating1 by its appropriate 'weighting1 and summing all the products.
With a total weighting of 100 and a merit scale of 0-10 as above the maximum
possible is 1000.
The relative merit of various power plants can be assessed immediately by
comparing their total scores, the power plant with the highest score being the
best. An idea of the absolute merit of the power plants can also be obtained
if the score is divided by 100 and the quotient related to the above rating
scale, e.g. a score of 1000/100 = 10 is a 'perfect1 power plant. A score
of 500/100 = 5 is an 'acceptable' power plant.
-------
6-6
Use of the Rating System in the Study
In order to apply the rating system to the power plants considered in this
study a committee: was uscu to assess tucr various ratings. *ne comrriii.cG
consisted of five experienced members of the Ricardo staff and great care
was taken to ensure that the committee had no bias to either diesel or
gasoline power plants. The power plants considered were those described
in the 'engine configuration1 section of the report with the addition of the
two gasoline engines described briefly in the same section.
i.e. l) V-8 'American1 gasoline
2) I.L.-6 'European' gasoline
3) V-8 I.D.I, diesel
4) 6 cyl. I.D.I, diesel turbocharged
5) 6 cyl. I.D.I, diesel with domprex
6) 6 cyl. D.I. diesel turbocharged
7) 6 cyl. D.I. diesel with Comprex
8) 6 cyl. 2-stroke loop scavenged I.D.I, diesel
9) 6 cyl. 2-stroke uniflow D.I. diesel
10) 4 cyl. 4-stroke compound D.I. diesel
11) 2-stage 2 bank rotary diesel
Each of the above power plants was to be considered for the primary and
secondary emission levels of the study.
i.e. Primary Targets
HC 0.41 g/mile
CO 3.4 g/mile
NOx 1.5 g/mile
Secondary Targets
HC 0.41 g/mile
CO 3.4 g/mile
NOX 0.4 g/mile
The above emission levels being measured according to the CVS-CH test
procedure.
-------
6-7
In fact as it was considered that none of the diesel power plants could pos-
sibly meet the secondary targets (in particular the 0.4 g/mile NOX figure)
the rating system was only applied to an environment embracing the primary
targets.
RESULTS OF RATING ASSESSMENT
The following 26 sub-sections give the 'ratings' given to each of the eleven
power plants on each performance aspect.
Each sub-section is preceeded by a summary table giving the numerical ratings
and the sub-section proper then follows giving notes on the derivation of the
various scores. Since a verbatim account of the committee deliberations
which led to each score would be tedious to the reader the first aspect
'smoke1 is described in some detail to illustrate the process while following
aspects are covered briefly unless major contentious points arose.
The final ratings are also shown in Table 2 at the end of the 26 sub-sections.
1. SMOKE
Engine Score
V-8 gasoline 9
IL6 gasoline 8
NA V-8 I.D.I. 6
TC 6 I.D.I. 4.5
Comprex 6 I.D.I. 5
TC 6 D.I. 4.5
Comprex 6 D.I. 5
Loop scav. I.D.I. 4
Uniflow D.I. 5
Compound D.I. 3
2 stage rotary 2
For commercial vehicles the absolute smoke level at which the engine is
limited varies considerably from country to country and even between manu-
facturers , some simply complying with legislative requirements whereas other
companies aim for significantly lower levels. For passenger car use, maxi-
mum smoke levels must be selected on an aesthetic basis in order to avoid
public criticism. From European experience, Ricardo would recommend that
-------
6-8
maximum steady state smoke levels of 5-8% opacity should be aimed for.
The combination of these low smoke levels, with the high power/weight ratio
of the vehicle should result in the exhaust being at least acceptable from
most of the candidate diesel power plants.
V-8 gasoline engine
Although the use of full load mixture enrichment devices can caused visible
black smoke with hard accelerations, the gasoline engine per se is com-
pletely free of visible smoke problems and can be classed in the category
of "best practical". Blue smoke can be formed in a worn gasoline engine
but this is also true of any other conventional internal combustion engine.
Score - 9
6 cylinder gasoline engine
Again the smaller more highly rated gasoline engine will only suffer a
smoke problem under acceleration conditions but its smaller size infers
that mixture enrichment will occur over a larger part of its load range
and therefore for a longer proportion of a given acceleration. Although
marginally inferior to its larger brother the small 180 CID gasoline engine
must still be regarded as extremely good when considering visible smoke
emissions.
Score - 8
NA V-8 I.D.I.
Black smoke is the major criterion which controls the output of any naturally
aspirated diesel engine. The smoke level selected for the performance esti-
mates will result in a very small amount of smoke being visible from the
kerbside, although at levels and for time periods which are better than just
acceptable to the general public. As engine load is reduced, black smoke
disappears rapidly from I.D.I, engines so that this power plant would be
better than acceptable from the point of view of black smoke.
White/blue smoke due to incomplete combustion of fuel should not be a pro-
blem from any of the indirect injection chamber engines.
Score - 6
-------
6-9
TC 6 cyl. I.D.I.
The smoke characteristics from this lightly boosted engine will be very
similar to those of the naturally aspirated V-8 although it will probably
suffer a transient smoke problem. This occurs when the turbocharger
inertia prevents the turbocharger output from keeping pace with the engine's
demands, the resultant over-rich fuel/air ratio leading to black smoke.
For an engine of this rating such a problem would normally be confined to
accelerations from very low speeds where turbine and compressor effici-
encies are low.
The magnitude of this problem in terms of smoke would be slight but would
probably be termed as annoying to the casual observer. Experience has
shown it to be almost insuperable if vehicle driveability is to remain un-
altered. Another annoying aspect of this problem is that it brings the
diesel's smoke right to the city centre and suburbs (the slight smoke problem
of the naturally aspirated diesel occurs at a higher speed/load and will not
normally be apparent during in-city driving).
Score - 4.5
Comprex boosted 6 cyl. I.D.I.
The reputedly instantaneous flow response of the Comprex pressure exchanger
to load and speed should improve the smoke emissions from a diesel engine
considerably. However, much time has been put into development of this
device without a production solution being achieved and the full range per-
formance and durability of Comprex has yet to be proven, At its present
state of development there is some confidence that this power plant will be
better than its turbocharged equivalent.
Score - 5
TC D.I,
Black smoke levels from the turbocharged D.I. should be similar to those
from the turbocharged I.D.I, assuming a practical method of achieving the
required NOX levels is developed (current technology indicates that the D.I.
engine will have great difficulty in achieving 1.5 NOy during CVS-CH) with-
out resorting to excessive amounts of timing retard. Even with small
amounts of retard a blue smoke problem may exist.
Score - 4.5
-------
6-10
Comprex boosted D.I.
Again the theoretical advantage of instant transient response of the pressure
exchanger make this candidate more attractive than the turbocharqed D.I.
and this power plant should be as good as the I.D.I, version.
Score - 5
Loop scavenge 2 stroke I.D.I.
Black smoke response from the 2 stroke I.D.I, should not be any worse than
from a four stroke I.D.I, but blue smoke may be a problem if large quanti-
ties of lubricating oil are burnt - due to the proximity of the inlet and exhaust
ports and their position within the cylinder there is every likelihood of this
occurring. It is this latter aspect which renders it likely that this engine will
be less than acceptable.
Score - 4
Uniflow 2 stroke D.I.
The blue smoke problem associated with the loop scavenge two stroke should
not occur on this engine and providing adequate scavenge can be provided black
smoke should not be a problem even with some timing retard (note it is
extremely unlikely that this engine with a D.I. chamber could achieve the pri-
mary NOx target and some device - e.g. exhaust gas recirculation. would be
required in addition to timing retard). It was considered that this engine
would be capable of achieving an acceptable smoke rating.
Score - 5
Compound D.I.
With its small radial compressor driven direct from the crankshaft, parasitic
losses at low engine speeds in addition to low boost pressures would force the
manufacturer to aim for maximum air utilisation at low speeds giving high
smoke levels. The extremely high boost ratio would require the adoption of
low compression ratios which would introduce light load misfire problems and
white/blue smoke during and immediately after a cold start would almost cer-
tainly be found. This engine would be somewhat worse than poor from the
point of view of smoke.
Score - 3
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6-11
2 stage rotary engine
The successful development of an automotive rotary diesel engine would need
years of research into optimum combustion systems and even then it is un-
likely that competitive air utilisation levels could be achieved. The clear-
ance volume between the rotor and trochoid is far removed from the optimum
as it contains large areas which are too shallow and inaccessible to support
or allow the propagation of combustion. This could lead to smoke problems
at both ends of the load range, blue smoke at light load, and black smoke at
high load with the distinct possibility of a "plateau" smoke problem bridging
the two. There is little doubt that this engine will be the worst of all those
assessed from this aspect and should be rated as 'bad'.
Score - 2
2. PARTICULATES
Engine Score
V-8 gasoline 7
IL6 gasoline 7
NA V-8 I.D.I. 2
TC 6 I.D.I. 2
Comprex 6 I.D.I. 2
TC 6 D.I. 2
Comprex 6 D.I. 2
Loop scav. I.D.I. 2
Uniflow D.I. 2
Compound D.I. 2
2 stage rotary 2
Currently the data is insufficient to come to a final decision concerning this
topic especially as the true reactivity of particulates has yet to be deter-
mined. It was argued that if the gasoline engine running on lead free fuel
were regarded as good, or worthy of seven merit points, the Diesel engine
with particulate levels of 10-20 times that of the gasoline must warrant a
rating of bad or two points. Ricardo do not feel qualified to predict relative
particulate levels between the different Diesel variants and therefore award
the same score to all. If particulate levels are to be limited by legislation,
the Diesel would need hang-on soot or particulate filters and it is in this
area that further work could be usefully carried out. The true health hazards
of various sizes of oarticles should also be determined before leqislation is oassed.
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6-12
3. ODOUR
Engine Score
V-8 gasoline 7
IL6 gasoline 7
NA V-8 I.D.I. 4
TC 6 I.D.I. 4
Comprex 6 I.D.I. 4
TC 6 D.I. 3
Comprex 6 D.I. 3
Loop scav. I.D.I. 3
Uniflow D.I. 3
Compound D.I. 3
2 stage rotary 2
It was generally agreed that the gasoline power plants are good from the point
of view of odour even though the exhaust fumes are noticeable under cold con-
ditions. All the diesel plants, however, are noticeably odorous and although
the modern automotive I.D.I, engines are good as diesel engines go they are
poor by comparison with the gasoline engines. Light load odour is perhaps
the most objectionable while the full load odour at low air/fuel ratios is also
noticeable.
The direct injection four-stroke engines will be worse at both light load and
full load than the indirect injection engines due to the more limited capa-
bilities of their combustion system.
The two-stroke engines will have similar characteristics to the direct injection
four-strokes.
The compound engine may have a problem with odour at low loads but is un-
likely to be much worse overall than the direct injection engines.
The rotary engine will be very odorous due to poor scavenging of end spaces
and the difficulty of arranging for complete combustion over a wide load and
speed range.
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6-13
4. NOX
Engine Score
V-8 gasoline 5
IL6 gasoline 5
NA V-8 I.D.I. 5
TC 6 I.D.I. 5
Comprex 6 I.D.I. 5
TC 6 D.I. (0)
Comprex 6 D.I. (0)
Loop scav. I.D.I. 5
Uniflow D.I. (0)
Compound D.I. (0)
2 stage rotary 5
On performance aspects involving legislative requirements, the manufacturer
will be forced by mainly economic considerations to produce a powerplant
which just satisfies the mandate (unless the engine already betters these
requirements in optimum performance build). Thus for all candidate power
plants where some reduction in pollutants is necessary to achieve the project
targets, a classification of acceptable was awarded even though the potential
of the engine might have been substantially better than regulatory require-
ments (since further reduction would involve other penalties).
Gasoline engines have already demonstrated their ability to achieve 1.5 g/mile
NOX during CVS-CH and it is predicted that a 130 bhp I.D.I, diesel should be
capable of attaining thi s level although with a significant degree of timing
retard. Heavy duty experience indicates that the D.I. engine will not be cap-
able of attaining this level and all variants with this chamber have been
awarded a zero score. When considering the secondary NOX target of 0.4
g/mile, it is known that several gasoline engines have achieved this level at
zero miles but catalyst durability is currently low. Daimler-Benz have demon-
strated a 65 bhp Diesel powered vehicle with NOX levels of 0.4 g/mile but
with the heavier vehicle and higher power/weight ratio of the vehicle for the
study it was felt that this target is not achievable with current technology and
production methods.
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6-14
5. HC
Engine Score
V-8 gasoline 5
IL6 gasoline 5
NA V-8 I.D.I. 5
TC 6 I.D.I. 5
Comprex 6 I.D.I. 5
TC 6 D.I. (0)
Comprex 6D.I. (0)
Loop scav. I.D.I. (0)
Uniflow D.I. (0)
Compound D.I. (0)
2 stage rotary (0)
As demonstrated earlier, it is estimated that with minor development, the
I.D.I, engine should be capable of complying with the 0.41 g/mile require-
ment over CVS-CH. The D.I. in retarded mode (in order to minimise NOX
but still not achieving the target of 1.5 g/mile) is likely to suffer extremely
high HC's and in Ricardo's opinion, none of the D.I. candidates would be
capable of complying with the HC requirement of this project. The poor
combustion chamber shape of the 2 stage rotary engine is hardly conducive
to efficient combustion and unburnt HC's are likely to be very high. Unfort-
unately the rapid drop in exhaust temperature as load is reduced ,on the diesel
engine is likely to restrict catalyst activity to the higher loads only and their
effectiveness during CVS-CH on a high powered diesel vehicle is open to doubt.
Investigations into low temperature light-up catalyst and catalyst matrix design
(for carbon storage and to ensure sufficiently free access of the HC's to the
catalyst) could be an area worth further investigation.
6. CO
Engine Score
V-8 gasoline 5
IL6 gasoline 5
NA V-8 I.D.I. 6
TC 6 I.D.I. 5
Comprex 6 I.D.I. 5
TC 6 D.I. 5
Comprex 6 D.I. 5
Loop scav. I.D.I. 5
Uniflow D.I. 5
Compound D.I. 5
2 stage rotary (0)
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6-15
There should be no problem in achieving the CO targets for any of these
engines apart from the two stage rotary engine where it is felt that the
inefficient combustion will again result in higher emissions.
The naturally aspirated V-8 I.D.I, should achieve the targets without any
modifications and therefore merits 6 points whereas the remaining candi-
dates score 5.
7. S02
Engine Score
V-8 gasoline 7
1L6 gasoline 7
NA V-8 I.D.I. 4
TC 6 I.D.I. 4
Comprex 6 I.D.I. 4
TC 6 D.I. 4
Comprex 6 D.I. 4
Loop scav. I.D.I. 4
Uniflow D.I. 4
Compound D.I. 4
2 stage rotary 3
This pollutant is totally dependent on the sulphur content of the fuel and the
amount of fuel being burnt. In assessing or rating the Diesel forSC>2 it was
assumed that the sulphur content of US light distillate (DFl) is similar to
that found in Europe (i.e. 0.2-0.5%). In theory the sulphur can be completely
extracted from the fuel but economic considerations have ruled this out to date<
The reactions of SC>2 in the exhaust to form sulphates/sulphides should not be
overlooked; the addition of oxidation catalysts would probably increase the
extent of these. The sulphur content of gasoline is very low (<0.1%) and thus
the gasoline engines were rated as good. Diesel engines must rate a merit
number of only 4 apart from the 2 stage rotary engine, where the extremely
high fuel consumption will drop the rating by one further number.
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6-16
8. HC REACTIVITY
Engine Score
V-8 gasoline 5
IL6 gasoline 5
IN A V-8 I.D.I. 7
TC 6 I.D.I. 7
Comprex 6 I.D.I. 7
TC 6 D.I. 7
Comprex 6 D.I. 7
Loop scav. I.D.I. 7
Uniflow D.I. 7
Compound D.I. 7
2 stage rotary 7
This subject is an important aspect of HC emissions and as mentioned earlier,
HC's emitted by the Diesel under high load conditions may be conveniently
wrapped in the soot skeleton to form a coagulum which effectively bars their
escape to the atmosphere or animal tissues. At light load conditions, blue
smoke can be found and here the HC's are freely available. However it has
been demonstrated that HC reactivity from the gasoline engine is 10 times
higher than from the Diesel. Thus if all gasoline variants warrant an accept-
able rating then the diesels should be at least two points better.
9. EVAPORATIVE EMISSIONS
Engine Score
V-8 gasoline 5
IL6 gasoline 5
NA V-8 I.D.I. ' 7
TC 6 I.D.I. 7
Comprex 6 I.D.I. 7
TC 6 D.I. 7
Comprex 6 D.I. 7
Loop scav. I.D.I. 7
Uniflow D.I. 7
Compound D.I. 7
2 stage rotary 7
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6-17
With modern control systems using carbon filter/storage canisters,gasoline is
an acceptable and practical fuel. Using the same systems, mid-distillates
such as used in nigh speed Diesel engines would suffer marginal evaporative
losses and must merit a good rating. This assumes fuels similar to current
No. 1, wider cut fuels may use more of the light end fraction which could
close the gap between diesel fuel and gasoline. One disadvantage of having
a fuel of low volatility is that if a fuel leak occurs the fuel does not evapor-
ate and its odour can be detected for some considerable time.
10. MISCELLANEOUS EMISSIONS
Engine Score
V-8 gasoline 5
IL6 gasoline 5
NA V-8 I.D.I. 5
TC 6 I.D.I. 5
Comprex 6 I.D.I. 5
TC 6 D.I. 5
Comprex 6 D.I. 5
Loop scav. I.D.I. 5
Uniflow D.I. 5
Compound D.I. 5
2 stage rotary 5
At the time of the study none of the power plants was regarded as producing
major quantities of any pollutants other than those already considered so that
all were rated as equal and acceptable.
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6-18
11. DRIVE-BY NOISE
Engine Score
V-8 gasoline 7
IL6 gasoline 6^
NA V-8 I.D.I. 6
TC 6 I.D.I. 6
Comprex 6 I.D.I. 5
TC 6 D.I. 4
Comprex 6 D.I. 3
Loop scav. I.D.I. 6
Uniflow D.I. 6
Compound D.I. 6
2 stage rotary 6
As mentioned earlier, the Diesel powered vehicle is only slightly noisier than
its gasoline counterpart when driven at speed. (Based on experience with
naturally aspirated indirect injection engines). The V-8 gasoline powered
vehicle, being very quiet under drive-by conditions (75-80 dB(A) at 15m
(50 ft)) must be given a drive-by rating of good whereas the 6 cylinder
European engine working with lower drive ratios will be marginally noisier.
Within the Diesel range, the indirect chamber engines will be inherently
quieter than their D.I. equivalents. With all engines operating in a retarded
timing mode, absolute noise levels of the I.D.I, engines will be only margin-
ally higher than the gasoline power plants. Predicted noise levels of the
candidate power plants are listed below. These predictions are for the bare
test bed engine at a distance of 1 m (3.2 ft). To correlate these with
American drive-by noise test conditions one must subtract approximately 20-
22 dB(A).
Bare engine test bed noise levels (l m)
Gasoline V-8 (97 mm bore, 66.7 rev/s, 93 dB(A))
Gasoline 6 (88 mm bore, 82.5 rev/s,, 96 dB(A))
NA V-8 I.D.I. ' 97
TC 6 I.D.I. 96
Comprex 6 I.D.I. 96
TC 6 D.I. 102
Comprex 6 D.I. 102
Loop scav. I.D.I. 93
Uniflow D.I. 96
Compound D.I. 100
2 stage rotary 93
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6-19
The formulae used in predicting the noise levels are empirical only and assume
current technology is used in the crankcase design - significant reductions
could be obtained by substantial modifications to the basic engine design. These
formulae also only predict combustion noise which is normally dominant in
these engines and take no account of other possible sources which could raise
the overall noise levels. When considering the boosted DI and IDI four stroke
engines allowance has to be made for the high sensitivity of the Comprex to
exhaust back pressure and general design restrictions throughout the flow
system (because of this the silencing of both intake and exhaust could not be
made as effective as the silencing of the turbocharged engines and some
additional noise penalty will be incurred). Of the last four candidates, so
little information is known of their noise characteristics, particularly in this
power range, that the degree of confidence in the predicted noise levels must
be low.
12. PACKAGE VOLUME
Scores
Engine Score
V-8 gasoline 7
IL6 gasoline 8
NA V-8 IDI 6
TC 6 IDI 5
Comprex 6 IDI 4
TC 6 DI 5
Comprex 6 DI 4
Loop Scav. IDI 4
Uniflow DI 3
Compound DI 5
2 stage rotary 8
Box volumes as calculated during the design configuration phase are listed as
follows :-
Box Vol.ft3
V-8 IDI 11.2
IL 6 IDI TC 12.7
V-6 IDI TC 11.6
IL 6 IDI Comprex 12.1 (.34)
V-6 IDI Comprex
IL 6 DI TC 12.0 ' (.34)
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6-20
V-6 DITC 11.0 (.31)
IL 6 DI Comprex 12.2 (.34)
V-6 DI Comprex - -
V-6 2 str. Loop Scav- 12.7 (.36)
IL 6 2 str. Uniflow 16.8 (.46)
4 cyl. Compound 11.3 (.32)
2 stage rotary 9.2 (.26)
All are given merit ratings roughly in accordance with their calculated box
volume apart from the Comprex boosted engine where the disadvantage of
taking the exhaust from near the front of the engine is not illustrated.
Since package shape is of at least equal importance to package volume, this
was taken into account in applying ratings. Engine length, for example, was
believed to be important in view of safety regulations.
13. PACKAGE WEIGHT
Scores
Engine Score
V-8 gasoline 6
IL 6 gasoline 7
NA V-8 IDI 5
TC 6 IDI 5
Comprex 6 IDI 5
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 4
Uniflow DI 4
Compound DI 5
2 stage rotary 7
The weight penalty of the small high speed diesel engine is large when compared
with a European type gasoline engine but this penalty is reduced considerably
if a larger American type V-8 is used as basis for comparison. It is Ricardo's
opinion that only the European gasoline engine can warrant a rating of good (7 points)
all other power plants apart from the rotary diesel considered in this exercise
are considerably heavier and must therefore warrant lower scores.
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6-21
Engine Est. weight Ib
V-8 gasoline 550
IL-6 gasoline 410
NA V-8 IDI 700
TC 6 IDI 720
Comprex 6 IDI 680
TC 6 DI 680
Comprex 6 DI 660
Loop Scav. IDI 760
Uniflow DI 800
Compound DI 670
2 stage rotary 500
14. FUEL ECONOMY
Scores
Engine Score
V-8 gasoline 5
IL 6 gasoline 5\
NA V-8 IDI 71
TC 6 IDI 7
Comprex 6 IDI 7
TC 6 DI 8
Comprex 6 DI 8
Loop Scav. IDI 6£
Uniflow DI 7
Compound DI 8
2 stage rotary 5
For this section it was felt that the rating system must be redefined on a quantitat-
ive basis and the following scale was devised.
Fuel economy mpg (1/100 km) Rating (normal subjective scale)
0 (totally unacceptable)
6 (39.4) 1
2 (bad)
10 (23.6) 3
4 (poor)
15 (15.7) 5 (acceptable)
6
20 (11.8) 7 (good)
(contd.)
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6-22
Fuel economy mpg (1/100 km) Rating (normal subjective scale) (contd.)
8
25 (9.45) 9 (best practical)
10 (perfect)
Predicted fuel economy figures were based on estimated and measured fuel
consumption levels from engines of each type. For the rotary engine, little
data are available but from basic combustion considerations and for the
higher friction levels of the rotary engine, specific fuel consumption curves
have been prepared from which comparative vehicle fuel economy levels
could be estimated.
Estimated vehicle fuel consumption and'ecbnorriy levels during CVS-CH are :-
1/100 km mpg
V-8 gasoline 18.2 13
IL6 gasoline 15.7 15
NA V-8 IDI 11.8-10.7 20,22
TC 6 IDI 12.4 - 11.2 19-21
Comprex 6 IDI 12.4-11.2 19-21
TC 6 DI (11.2-10.3) (21-23)
Comprex 6 DI (11.2 -10.3) (21-23)
Loop Scav. IDI (13.1-11.8) (18-20)
Uniflow DI (12.4-11.2) (19-21)
Compound DI (11.2 -10.3) (21-23)
2 stage rotary (16.8-14.7) (14-16)
Fuel economy levels in brackets indicate that although the injection timings of
the engines have been retarded as far as is practical,it is estimated that at
least one primary emissions target (normally NOx - see emissions section)
has not been achieved. The poorer fuel economy of the boosted IDI as compared
with the naturally aspirated V-8 is due to its need for further timing retard to
comply with 1.5 g/mile NOx.
15. FUEL COST
Scores
Engine
V-8 gasoline
IL 6 gasoline
NA V-8 IDI
TC 6 IDI
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6-23
Comprex 6 IDI 5
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 5
Uniflow DI 5
Compound DI 5
2 stage rotary 5
The fuel requirements of all the diesel candidates are similar to current light
distillate (DFl). Due to a lack of real data on the economics of fuel refineries
and because the switch towards greater use of diesel fuels does not require
technological innovation, it is assumed that the fuel situation for both gasoline
and diesel candidates must be regarded as acceptable and a score of 5 has thus
been awarded.
It is estimated that the adoption of lead free gasoline would incur a price
penalty of 10% however, and if sulphur free diesel fuel were stipulated, a severe
cost penalty might arise, but due to the uncertain future situation with regard to
fuel prices no attempt has been made to compare diesel and gasoline engine fuels.
16. VEHICLE FIRST COST
Scores
Engine Score
V-8 gasoline 6
IL 6 gasoline 7
NA V-8 IDI 5
TC 6 IDI 5
Comprex 6 IDI 5
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 4.5
Uniflow DI 4.5
Compound DI 4
2 stage rotary 4
If the existing gasoline powered American car is regarded as slightly better than
acceptable in terms of first cost then the European type engine should give
sufficient advantage to warrant an extra point. The first cost of all diesel
engines is undoubtedly a penalty and all the evidence available indicates that a
cost penalty of some 10% will be incurred with a diesel powered vehicle even
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6-24
after allowing for the cost of catalysts for the gasoline engine. Because of this
the ' conventional' diesel powered vehicles would only be just acceptable in
terms of first cost. (The potential cost savings with the smaller boosted engines
are offset by the cost of the supercharging equipment). The higher development
time and cost for the two stroke engines will impose a surcharge on the vehicle
first cost so that these power plants must be rather less than acceptable. The very
high development, manufacturing and material costs of both the compound and
rotary engines will pull them down to a 'poor' category for this aspect.
17. MAINTENANCE
Scores
Engine Score
V-8 gasoline 5
IL 6 gasoline 6
NA V-8 IDI 6
TC 6 IDI 6
Comprex 6 IDI 6
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 5
Uniflow DI 6
Compound DI 4
2 stage rotary 2
Data extracted from the literature survey and the experience of taxi operators
in Europe all suggest that minor maintenance costs for diesel vehicles are
similar to those of gasoline equivalents but that major engine services for the
diesel are not required within the normal operating life of the vehicle. The
average operating load factor for American passenger car engines is some-
what lower than European vehicles (solely because of their higher power:weight
ratio) and the maintenance periods quoted for European use should be increased
when considering the American situation.
Maintenance costs for gasoline engines are regarded as satisfactory and
European experience indicates that the four stroke IDI engines should be better
than this. The direct injection engine uses plain hole nozzles which require
cleaning at more frequent intervals than the variable orifice pintle nozzle
used in the indirect injection combustion systems. Otherwise maintenance
requirements are the same as for the IDI engine.
Of the two stroke engines, the direct injection uniflow engine has been given a
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6-25
higher reliability rating mainly because of the excellent reputation of GM's
two stroke engines, to which this engine is closely akin.
The compound engine which is the candidate with the highest thermal loading -
0.00353 kW/mm2 (3.04 hp/in2\ and mechanical stresses - 137 bar (2000 lb/in2)
cylinder pressure - is likely to be a much more sophisticated design particularly
around the combustion chamber area,and major cylinder head maintenance
periods are likely to be far shorter than for the other variants.
li; is felt that the 2 stage rotary engine would also need frequent servicing even if
only because the high cylinder pressures would cause high rates of seal wear and
the use of only one injector per bank with very high injection frequencies could
increase servicing demands in this area.
18. STARTABILITY
Sicores
Engine Score
V-8 gasoline 6
IL 6 gasoline 6.5
NA V-8 IDI 5
TC 6 1D1 5
Comprex 6 IDI 5
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 4
Uniflow DI 5
Compound DI 2
2 stage rotary 2
With the ability to start instantaneously under most environmental conditions
experienced in America, the gasoline engine should merit a very high rating
but the hot starting of low emissions vehicles is known to be poor and is not
likely to improve substantially beyond today's levels. Thus the V-8 engine was
awarded 6 points but the 6-cylinder engine with its fuel injection system should
start more easily and was thus awarded 6.5.
The starting characteristics of the naturally aspirated diesel could only be
classified as acceptable even with a programmed start system. The develop-
ment of an instantaneous warm up heater plug which gives a clean start in all
conditions would improve the start up of the die^el to a significant extent.
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6-26
With the low boost pressure ratios used on the 6 cylinder IDI engines it should
be possible to use the same compression ratio as in a naturally aspirated
engine without encountering excessively high cylinder pressures at full load
conditions. Starting characteristics should therefore be unaltered.
Although large DI engines have extremely good starting characteristics, tests
run by Ricardo on a high speed DI with a cylinder capacity of 0.5 1 (30 in^)
have indicated startability to be on a par with the IDI. This is due to higher
heat losses to the cylinder walls (a) because of the higher swirl rates needed
for high speed operation,and (b) because of the higher surface areatvolume
ratio of the smaller cylinder. The compression ratio of the DI was similar
to that of the swirl chamber engine.
It is likely that the loop scavenge IDI will have a low effective compression ratio
at cranking speeds with resultant inferior starting. It is difficult to quantify
this but a rating of 4 has been awarded if only to point out a likely problem area.
The uniflow two stroke should start as well as the four stroke engines.
In order for the compound engine to attain its extremely high outputs without
exceeding its designed maximum cylinder pressures, its compression ratio will
need to be limited to 13:1 and severe starting/misfire/blue-white smoke
problems are foreseen. Methods might be developed for reducing the misfire
and light load smoke problems after starting but with current technology it is
extremely unlikely that a clean, reliable start could be achieved in ambient
American winter conditions. A score of 2 points has therefore been awarded.
With the rotary engine, cranking heat losses between stages will be extremely
high as will be blowby losses at low speeds and it is difficult to envisage its
starting characteristics to be anything better than bad.
19. HOT DRIVEABILITY
Scores
Engine
V-8 gasoline
IL 6 gasoline
NA V-8 IDI
TC 6 IDI
Comprex 6 IDI
TC 6 DI
Comprex 6 DI
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6-27
Loop Scav. IDI 7
Uniflow DI 7
Compound DI 7
2 stage rotary 7
Hot driveability of all candidates apart from the two turbocharged four stroke
engines will in no way be inferior to current generation passenger car engines
(this assumes that all-speed governing used on most European diesel cars
is discarded in favour of simple idling and overspeed governing), and a score
of 7 points has been awarded to all but these types. The turbocharged engines
will suffer a slight lag during acceleration because of tubocharger inertia and
the resultant feeling that the driver has not complete control over the vehicle
can be disturbing. It is claimed that the Comprex does not suffer this delay
problem.
20. COLD DRIVEABILITY
Scores
Engine Score
V-8 gasoline 5
IL 6 gasoline 6
NA V-8 IDI 8
TC 6 IDI 8
Comprex 6 IDI 8
TC 6 DI 8
Comprex 6 DI 8
Loop Scav. IDI 8
Uniflow DI 8
Compound DI 8
2 stage rotary 8.
With flat spots due to carburation problems at cold temperatures the V-8
gasoline engine can barely be regarded as acceptable. The 6 cylinder gasoline
engine, however, will be better than this due to the use of fuel injection. The
diesel engine pulls away immediately without hesitation once a start has been
achieved, the only difference between hot and cold driveability being due to the
higher friction levels of the cold engine. Eight points were awarded therefore
to all the diesel candidates.
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6-28
21. TORQUE RISE
Scores
Engine Score
V-8 gasoline 7
IL 6 gasoline 7
NA V-8 IDI 7
TC 6 IDI 7
Comprex 6 IDI 7
TC 6 DI 7
Comprex 6 DI 7
Loop Scav. IDI 7
Uniflow DI 7
Compound DI 2
2 stage rotary 7
All candidates achieve the torque characteristics demanded by the performance
specifications, assuming a three speed automatic gearbox, with the one
exception of the compound four stroke engine. The extremely poor torque
curve of this engine would need a very sophisticated transmission system to
achieve any degree of acceptance with the general public and only two points
were awarded to this engine.
22. DURABILITY
Scores
Engine
V-8 gasoline
IL 6 gasoline
NA V-8 IDI
TC 6 IDI
Comprex 6 IDI
TC 6 DI
Comprex 6 DI
Loop Scav. IDI
Uniflow DI
Compound DI
2 stage rotary
Durability of 1.5 g/mile NOx American gasoline engines is assumed to be satis-
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6-29
factory on the assumption that catalyst life can be increased. The European type
engine with its higher rating and operational speed range is likely to suffer
slightly higher wear rates but a satisfactory life of 100,000 miles can easily
be achieved.
The life of the diesel powered passenger car engine in Europe is significantly
greater than that of the gasoline engine. The same arguments should apply in
American use although absolute life in terms of mileage will be greater in
America because of higher power: weight ratios (and therefore lov/er relative
load factors). It is therefore Ricard's opinion that the durability of the naturally
aspirated IDI should be classified as good (7 points). The boosted engines, both
DI and IDI, will have marginally lower useful lives because of higher mechanical
and thermal loadings in and around the combustion chamber. However, their
longevity will still be comparable to the American V-8 gasoline engine.
Due to their simplicity the life of the two stroke engines should be equivalent to
or better than the four stroke engines. The excellent reputation of GM 's
uniflow engines cannot be overlooked and for this reason the uniflow DI has been
rated at 8 points.
With considerable development the compound engine could achieve 160,000 km
(100,000 miles) in passenger car service (acceptable - 5 points) but the
durability of the rotary engine is open to speculation and in Ricardo's opinion
it could not be developed to a satisfactory degree using current technology.
23. COOLANT HEAT LOSSES
Scores
Engine
V-8 gasoline
IL 6 gasoline
NA V-8 IDI
TC 6 IDI
Comprex 6 IDI
TC 6 DI
Comprex 6 DI
Loop Scav. IDI
Uniflow DI
Compound DI
2 stage rotary
-------
6-30
Although the heat loss to coolant of the IDI diesel is some 25% higher than that
of the gasoline engine, cooling under full load conditions is not the critical
case in passenger applications and although some increase in radiator size
may be necessary for diesel cars, it certainly would not be a major increase
and it may be possible to use the same radiator (Peugeot use the same
radiator in the gasoline and diesel cars but the diesel engine is of much
lower output). It is therefore assumed that no significant front end changes
are needed for the application of a diesel engine to passenger cars and the
same acceptable rating was awarded to all candidates except the rotary
diesel where the extremely poor fuel utilisation will significantly increase heat
losses - 4 points were awarded in this case.
24. FIRE RISK
Scores
Engine Score
V-8 gasoline 5
IL 6 gasoline 5
NA V-8 IDI 6
TC 6 IDI 6
Comprex 6 IDI 6
TC 6 DI 6
Comprex 6 DI 6
Loop Scav. IDI 6
Uniflow DI 6
Compound DI 6
2 stage rotary 6
Apart from the fact that 0.2% of accidents involve fire, no quantitative data
relating to fire risk was uncovered during the literature survey.
Causes of fires in automotive accidents are inevitably very complex but the
major source of ignition is a high energy electrical discharge or a fire in the
wiring loom due to a dead short. Thus the flash point of the fuel has very little
influence on fire risk but its volatility does in that a highly volatile fuel such
as gasoline can create a large area of mixture with sufficient strength to support
combustion, in this manner it can 'transport' itself effectively to the ignition
source. Under normal climatic conditions the volatility of diesel fuel is too
low to allow the formation of a combustible mixture.
The lack of H.T. ignition systems in diesel vehicles must also lower their fire
-------
6-31
risk (whether an accident occurs or not) .
For these reasons the diesel vehicle has been given one point more on the merit
rating scale than the gasoline vehicle. It is assumed the fire risk of the gasoline
vehicle must be regarded as acceptable (i.e. 5 points).
25. IDLE NOISE
Scores
Engine
V-8 gasoline
IL 6 gasoline
NA V-8 IDI 5
TC 6 IDI 5
Comprex 6 IDI 5
TC 6 DI 3
Comprex 6 DI 3
Loop Scav. IDI 5
Uniflow DI 4
Compound DI 2
2 stage rotary 5
The idling noise of the gasoline engine is low in terms of both objectionability
and overall noise level and must be approaching the best practical level that
a power plant can achieve.. Although idling noise levels of the IDI diesel
engines are low, the characteristics of the noise are subjectively annoying
and have an immediate impact on the casual observer in the street. The overall
noise level from the IDI is sufficiently low to make this just acceptable. With
its greater instantaneous pressure rise the DI engine sounds harsher than the
IDI and noise levels are normally somewhat higher; Ricardo feel it would be
regarded as approaching bad or even unacceptable by the general public. The
two stroke engines were given the same score as the four stroke engines with
the same chamber apart from the uniflow engine which was given the benefit
of one extra point, again mainly because of the reputation of GM's engines.
The compound engine with its very low compression ratio will have very long
delay periods at idling conditions with attendant harsh, loud combustion noises.
Mechanical gear noise in the complex drive may also be a problem because of
torque recoil.
In the case of the two stage rotary engine it is difficult to predict idle noise but
-------
6-32
the shape should limit noise transmission and it was therefore awarded 5 points,
i.e. the same as the reciprocating IDI engines.
26. VIBRATION AND TORQUE RECOIL
Scores
Engine Score
V-8 gasoline 8
IL 6 gasoline 8
NA V-8 IDI 6
TC 6 IDI 5
Comprex 6 IDI 5
TC 6 DI 5
Comprex 6 DI 5
Loop Scav. IDI 5
Uniflow DI 5
Compound DI 4
2 stage rotary 7
The aspect of vibration and torque recoil is undoubtedly a major contributor to
the feel or refinement of a vehicle.
Both the gasoline power plants are so smooth that the user is totally unaware
of any reciprocating motion and they must be regarded as the best practical
expression of power plants.
The unthrottled diesel engines with higher compression ratios are all inherently
rougher than the gasoline engine, and the V-8 would be only good while the 6
cylinder engines would all tend to drop down to an ' acceptable' rating due to the
lower number of cylinders.
The compound engine is essentially a 4 cylinder engine from this aspect and
must be considered poor.
Of all the diesel power plants it was felt that the rotary engine would score best
on this aspect and although not quite as unobtrusive as the gasoline engines, it
would certainly be good due to its inherent good "balance.
-------
6-33
TABLE 2
Final Ratings for Each Aspect
; ;
.
« ' ,,
C ! O
g -a j £•
•— t W ! 1— I
O ^0 ' f"*\
w en »_ ; H
03 Q
01 j; H
00 -^
Aspect ! ' °
^* ^O
1 Smoke 9 8
2 Particulates 7 7
3 Odour 7 7
4 NOx 5 5
l~~!
OO ! ^
1
o
vO
6 4.5
2
4
2
4
5 5
5 HC 555
5
6 CO 5565
1 7 SO2 77
8 HC reactivity 5 5
9 Evap. emissions 5 5
10 Misc. emissions 5 5
11 Noise (drive-by) 7 6.5
12 Package volume 7 8
4 4
7 7
7 7
5
6
6
5
6
5
13 Package Weight 6 7 j 5 5
14 Fuel economy 5 5.5 7.5 : 7
15 Fuel cost 5 5
16 Vehicle first cost 6 7
17 Maintenance cost 5 6
18 Startability 6 6.5
19 Hot drive ability 7 7
20 Cold driveability 5 6
21 Torque rise 7 7
22 Durability 6 5.5
23 Heat loss 5 5
24 Fire risk 5 5
25 Idling noise 8 8
26 Torque recoil 8 8
5
5
6
5
5
6
5 5
7 6
8 8
7
7
5
6
5
6
7
6
5
6
5
5
x !
QJ i
(H
£
0
i
._,
/->,
Q
i
">,
0
vO
.
, 5
2
4
5
5
5
4
7
7
5
5
4
5
7
5
5
6
5
7
8
7
6
5
6
5
5
! 0)
U
H
I— I
Q
Ix
O
SO
4.5
X
Q)
a
o
U
c
0)
CO
o
w
a
r\
,, ) T*S
o
d
'e c
Q 2 3 3
«
^
CJ
5
2 2
3 3
(0) (0)
(0) (0)
5 I 5
4 : 4
7
7
7 7
5 5
4 |3
5 ! 4
5 [5
8 8
5
5
! 5
5
6
8
7
6
5
6
3
j 5
5
5
5
5
7
8
7
6
5
6
3
5
.
SH
W
CM
4
2
3
u i &
fc ! E
1/3 i o
CM O
5 3
2 2
3 t 3
5 (O)j(O)
(0)
5
(o)j(o)
5 5
4 | 4
7
7
5
6
4
4
6.5
7
7
5
6
3
4
7
5 5
4.5 4.5
5
4
7
8
7
7
5
6
5
5
6
5
7
8
7
8
5
6
4
5
4
7
7
5
6
5
5
8
5
4
4
2
7
8
2
5
5
6
2
4
i
i
t'
cc
C
cv
CC
-*-'
CM
2
2
2
5
(0)
(0)
4
7
7
5
6
8
7
5
5
4
2
2
7 !
8.
7 i
2 :
4 !
6
5
7 '
-------
6-34
RESULTS
The product of the weighting and the rating for each performance aspect was
summed for each power plant and the results are shown in Table 3 below:-
TABLE 3
Power Plant Final Score
(Rounded to nearest whole number)
1
2
3
4
5
6
7
8
9
10
11
V-8 gasoline
6 cyl. gasoline
V-8 EDI
6 cyl IDI T/C
6 cyl EDI Comprex
6 cyl. DIT/C
6 cyl. DI Comprex
2 str. loop scavenge
2 str. uniflow
Compound
2 stage rotary
608
620
587
556
554
(500)
(497)
(516)
(515)
(465)
(434)
In order to establish the validity of the 'committee method' of rating two re-
runs were done for the gasoline engine and the IDI diesels,when the final
runs for these power plants were found to be within - 2^% of the above figures
and the relative order of scores was not found to change. Because of this the
final scores from the first complete run shown in Table 3 were taken as
representative of the relative merit of the various power plants.
The rating methodology has the disadvantage that a rating of 0 in any aspect
can be hidden by good ratings for other aspects and the bracketed figures
in Table 3 are for those power plants which scored a 0 (i.e. totally unaccept-
able) in one or more performance aspects.
The results are revealing in that they immediately indicate that the gasoline
power plants are superior to the diesel power plants and that only the indirect
injection 4-stroke diesels are viable for the duty considered. All the other
diesel power plants are unacceptable due to the inability to score better than
zero on one or more performance aspects (generally emissions). With the
weighting adopted, the gasoline engine's superiority for passenger car use
in America is demonstrated but the closeness of the final scores indicates that
only quite minor changes in the weightings would bring the scores equal. It
is apparent that in an emissions and fuel conscious environment there are
-------
6-35
many applications which would change the individual weighting system
sufficiently to make the automotive diesel an attractive alternative to the
gasoline engine.
The smaller 6 cylinder gasoline engine has 12 more points than the V-8
entirely on the strength of its better fuel economy and lower first cost.
The potentially viable IDI diesel power plants scored some 50 points or so
below the gasoline engines. This is well outside the 15 point scatter found
in the re-runs but is sufficiently close to permit their consideration for
many applications.
It can be seen that all the other diesel power plants were unacceptable,
mainly due to their poor HC and NOx characteristics.
If the total score for each contender is divided by 100, this averaged score
can be related to the original merit table. It can immediately be seen that
no plant can be called 'good' although both gasoline engines score better
than 6, i.e. better than 'acceptable' although not quite 'good' .
The IDI diesel plants on this basis are still better than 'acceptable1 although
an average score of 6 is not quite achieved.
Certain aspects beyond the scope of this study, such as depletion of world
stocks of precious metals due to the use of catalysts or the implications
of particulate legislation, cannot be covered by a rating method such as
this but, where possible, these other aspects should be considered before
any major decisions on automotive power plants are made.
CONCLUSIONS
The rating methodology evolved has allowed different power plants to be
compared on a numerical basis.
Its application to the eleven power plants in this study has indicated that
gasoline power plants are some 10%'superior' to the diesel power plants
for an average light duty application.
The viable diesel power plants, four-stroke IDI's, are sufficiently close to
the gasoline engines to allow changes in the importance of some performance
aspects, notably fuel economy, to bring them up to the same effective rating.
-------
6-36
None of the other diesel power plants were viable for the primary emissions
environment.
No diesel power plant would be viable for the secondary emissions environ-
ment without considerable advances in technology.
-------
6-37
FIG.6-1
LIGHT DUTY VEHICLE POWER PLANT
SURVEY RESULTS
1. SMOKE
2. PARTICULATES
3. ODOUR
k. N02
5- . HC
6. CO
7- S02
8. HC REACTIVITY
9. EVAPORATIVE EMISSIONS
10. MISC. EMISSIONS
] 1 . NOISE (DRIVE BY)
12. PACKAGE VOLUME
13. PACKAGE WEIGHT
1/4. FUEL ECONOMY
15. FUEL COST
16. VEHICLE FIRST COST
17. MAINTENANCE COST
18. STARTABILITY
19. HOT DRIVEABILITY
20. COLD DRIVEABILITY
2V. TORQUE RISE
22. DURABILITY
23. HEAT LOSS
2k. FIRE RISK
25. IDLING NOISE
26. TORQUE RECOIL
10 20
°/o WEIGHTING
-------
7-1
SECTION 7
PROGRAMME PLANS
While this study has confirmed that the diesel engine is a viable alternative
power plant for a passenger car and indeed would have advantages in a
number of areas, notably in respect of fuel consumption, it has brought out
other areas where the diesel engine is less attractive than the gasoline
engine. The report has also shown up areas where the technical data is
not as complete as would have been desirable for the purposes of the study.
It would be advantageous therefore to carry out a number of experimental
programmes aimed partly at filling these gaps in knowledge and partly at
overcoming the areas of deficiency in performance.
1. Construction of V-8 naturally aspirated, indirect injection diesel engine
While it is believed that accurate estimates of fuel consumption and exhaust
emission levels have been made, these have involved predictions with vehicle
power:weight ratios much higher than any for which actual test data are avail-
able.
For a convincing demonstration that these performances are possible it is
desirable to fit a suitable diesel engine into a 1600 kg (3500 Ib) passenger
car and obtain a full set of vehicle performance data. This would also pro-
vide a convenient demonstration vehicle.
While the time scale for the design and construction of a new engine is likely
to be in the region of 30-36 months, it would be possible to produce an
engine in a shorter time by carrying out a conversion of an existing gasoline
engine. This would in no way affect the accuracy of the vehicle performance,
fuel consumption or emissions measurements but could lead to a necessity
for compromise in so far as some of the other performance aspects are con-
concerned.
2. Pressure Charging Investigations
More data are necessary to refine the predictions for pressure charged engines.
With a turbocharger for example, it may be difficult to drive a cycle such as
the LA4 with a large number of transients and still have freedom from exhaust
smoke, with resulting excessive exhaust CO levels.
Theoretical simulation is very difficult in this area and there would be much
to be gained by carrying out comparative test bed and vehicle studies using
a turbocharger and as an alternative, the Comprex. It would of course be
necessary to use a smaller engine and once again while it would be possible
to design and develop a new engine, it will probably be preferable either to
-------
7-2
use an existing six cylinder diesel engine such as the Ford "York" or alter-
natively to carry out a conversion of an existing gasoline engine.
3. Improved Fuel Injection System
There are two problems in the area of the fuel injection equipment. The
first results from the high degree of accuracy necessary for acceptable
performance and low emissions. This high accuracy leads to high cost and
hence to the cost of the fuel injection equipment accounting for some 50% of
the difference in cost between a diesel engine and a gasoline engine.
The second arises from the need, for reasons of low emission levels, to
have an injection timing which varies with load and speed. This leads to
additional expense and to a considerable complication with most mechanical
injection systems. Indeed it is this requirement that has spurred work with
piezo, accumulator, electro-magnetic and other injection systems. If it
could be satisfactorily carried out it. might prove advantageous to vary the
pumping rate with either load or speed, to use pilot injection, or even to
vary the pumping rate during each delivery stroke.
There is much to be said therefore for pursuing work in the areas of novel
and lower cost fuel injection equipment.
4. Low Compression Diesel Engines
The interest in the low compression diesel engine lies in the expectation
that it will give a lighter, more efficient engine. If exhaust hydrocarbon
and odour levels are not to be unacceptable however, it is essential to have
a reliable starting and low load running, ignition aid. This might take the
form of a high energy spark, a heater plug, an inlet air heater, exhaust
recirculation or indeed be of some other form.
It may or may not be possible to produce such a device but in view of the
possible advantages it is desirable that work should be carried out to investi-
gate the feasibility of the aid and, if successful, to investigate the actual
advantages of a low compression diesel engine.
5. Exhaust Hydrocarbon Formation
While the chemical reactions leading to the formation of NOX are, at least
partially, understood and it should be possible to meet the primary NOX
emissions target of 1.5 g/mile, further reductions in NOX are made diffi-
cult by resulting rises in hydrocarbon levels. Indeed it may prove necessary
to use a catalyst to hold hydrocarbon levels down low enough with this primary
emissions target. With direct injection engines, it is the rising hydrocarbon
levels which make it impractical to attain 1.5 g/mile NOX-
-------
7-3
It is recommended therefore that a fundamental study should be carried out
into the mechanism of exhaust hydrocarbon formation. Are the hydrocarbons
associated for example with particulate material arising from combustion?
Do they result from flame quenching close to the combustion chamber walls?
Do they arise from slowburning and subsequent quenching of part of the fuel
or alternatively do the last portions of the fuel to be injected have excess-
ively large drops due to changes in the hydraulics of the fuel injection system.
It is possible that some other mechanism may be involved but if the mech-
anism is understood it could lead to a control of hydrocarbon formation and
hence to the development of lower emission diesel engines.
As hydrocarbon levels are always likely to be a problem at the retarded
injection timings which are essential for very low NOX emissions, it is
desirable to carry out work aimed at the development of a catalyst with a
low "light off temperature which would be effective at the low exhaust
temperatures which occur at part load in a diesel engine.
6. Modulation Systems for the Control of Exhaust Gas Recirculation
To avoid excessive derating of naturally aspirated engines, it is necessary to
modulate exhaust gas recirculation quantities with load and it may also prove
desirable to modulate with speed. It may prove to be necessary to reduce
the recirculation at peak torque conditions with a pressure charged engine.
Control devices for such modulation in diesel engines are not, so far as
Ricardo know, available at this time and those being developed for gasoline
engines may suffer problems due to the particulate material in diesel
exhaust. It is suggested therefore that work be initiated to develop a
practical control system.
7. Improved Starting Aids
The delay required to bring heater plugs up to operating temperature is a
considerable subjective problem. The use of a programmed starting sequence
can help and indeed such devices are in use, if not actual articles of com-
merce. "Instant", or more realistically, fast heating plugs would be a con-
siderable advantage however, and a general programme aimed at quick start-
ing under cold ambient conditions is desirable.
8. Demonstration of Diesel Engines for Specialised Applications
This study demonstrates the clear superiority of the diesel engine for taxi
cab and light delivery vehicles. Demonstrations in these applications in
America in the past have failed due to the inadequate power/weight ratio of
the demonstration vehicle or due to the use of unsuitable vehicle/engine com-
binations .
-------
7-4
It is recommended therefore that following the construction or conversion of
a suitable engine and an investigation of its performance in a suitable vehicle,
a programme should be instituted with a number of these vehicles to demon-
strate the advantages and to encourage the conversion of current and future
taxi cabs and light delivery trucks to diesel power.
9. Particulate Levels
The relatively high particulate levels which occur in the exhaust from a diesel
engine even under clean exhaust conditions, could pose a long term problem.
It would be desirable therefore to carry out a fundamental programme aimed
at finding the source of these particulates and hopefully therefore at a method
of controlling their formation.
It would also be desirable to carry out a programme aimed at the development
of a trap which would eliminate, or at any rate reduce, smoke and particulates
from an exhaust which contains excessive levels.
Since neither the 'spectra1 of particle sizes from a given power plant nor
the relative harm which can be caused by particles of a given size and
nature is known, it is apparent that this lack of knowledge should be remedied
before any legislation is finalised. It is thus suggested that work be intensi-
fied in the field of particulate emissions to ensure that legislation is relevant
to the needs of the population.
-------
8-1
SECTION 8
APPENDIX 1 - LIST OF KEYWORDS
1. SMOKE
2. ODOUR
3. GASEOUS EMISSIONS
4. PARTICULATES
5. NOISE
6. VOLUME
7. WEIGHT
8. FUEL ECONOMY
9. FUEL
10. FIRST COST
11. MAINTENANCE
12. STARTING
13. HOT DRIVEABILITY
14. COLD DRIVEABILITY
15. TORQUE RISE
16. DURABILITY
17. COOLANT HEAT LOSSES
18. FIRE RISK
19. VIBRATION AND TORQUE RECOIL
20. IDLE NOISE
21. RELIABILITY
22. ECONOMICS
23. MANUFACTURE
24. PERFORMANCE
25. DESIGN
26. ANCILLIARIES
27. LUBRICATION
-------
8-2
1. SMOKE
Measurement techniques
Legislation
Formation
F.I.E. - timing - injection rate - pilot injection - nozzle
Combustion chamber - D.I. - I.D.I. - compression ratio
E.G.R.
Boosting - turbocharging - comprex - supercharging - ramming
Diesel vs gasoline
Levels vehicle
Levels Test bed
User experience
Public opinion
Endurance
Light load - blue - white - aids
Fuel - quality - dual - additives - alternate
Exhaust system - catalyst - soot filter
Ambient conditions
D.I. vs I.D.I.
Mathematical models
Water injection
-------
8-3
2. ODOUR
Measurement techniques - subjective - meter
Legislation
Formation
F.I.E. - timing - injection rate - nozzle
Combustion chamber - D.I. - I.D.I.
E.G.R.
Diesel vs gasoline
Levels - vehicle
Levels - test bed - subjective - meter
User experience
Public opinion
Endurance
Fuel - quality - dual additives - alternate
Light load aids
Exhaust system - catalyst - soot filter
Ambient conditions
Mathematical models
-------
8-4
3. GASEOUS EMISSIONS
Measurement techniques - heavy duty - light duty - miscellaneous
Legislation - heavy duty - light duty
Formation
F.I.E. - timing - injection rate - nozzle - pilot injection
Combustion chamber - D.I. - I.D.I. - compression ratio
Fuel - quality - dual - additives - alternate
Water injection
E.G.R.
Boosting - comprex
Diesel vs gasoline
Levels - vehicle
Levels - test bed
Endurance
Light load aids
D.I. vs I.D.I.
Ambient conditions
Catalyst
Cost
Mathematical models
Trends
Exhaust reactor
-------
8-5
4. PARTICULATES
Measurement techniques
Legislation
Formation
Combustion chamber - D.I. - I.D.I.
Boosting
Diesel vs gasoline
Levels - vehicle
Levels - test bed
Endurance
Fuel - additives
D.I, vs I.D.I.
Exhaust system - catalyst - soot filter
Ambient conditions
Mathematical models
Toxicity
-------
8-6
5. NOISE
Measurement techniques
Legislation
D.I, vs I.D.I.
Exhaust
Intake
F.I.E. - timing - injection rate - pilot injection - nozzle
Combustion chamber - D.I. - I.D.I. - compression ratio
Fuel - quality - dual - additives - alternate
Boosting
Shields
Enclosures
Structure - isolation - materials
Diesel vs gasoline
Levels - vehicle - interior - drive by - stationary
Levels - test bed
User experience
Public opinion
Air cooled vs water cooled
Engine configuration
Combustion
Mechanical
Endurance
-------
8-7
6. VOLUME
Configuration
Diesel ys gasoline
Vehicle installation..
Air cooled vs water cooled
Specific
-------
8-8
7. WEIGHT
Materials
Configuration
Diesel vs gasoline - engine - vehicle
D.I, vs I.D.I.
Bare engine
Air cooled vs water cooled
Specific
-------
8-9
8. FUEL ECONOMY
D.I, vs I.D.I.
Fuel - dual - additives - alternate - quality
Boosting
C onfig ur ation
Diesel vs gasoline
Vehicle
Test bed
User experience
Economics
Endurance
Light load aids
Thermal efficiency
-------
8-10
9. FUEL
Cost - tax. - prime
Refinery technique
Quality
Alternate
Additives
Diesel vs gasoline
Trends
Production - trends
-------
10. FIRST COST
F.I.E.
Boosting
Legislation - smoke - emissions - odor - noise
Diesel vs gasoline
D.I, vs I.D.I.
Vehicle
Engine - components
User experience
Economics
Air cooled vs water cooled
Specific cost
Tolerances
-------
8-12
11. MAINTENANCE
Engine
F.I.E.
Ancilliaries
Diesel vs gasoline
Methods of improving
I.D.I, vs D.I.
Vehicle
User experience
Economics
Air cooled vs water cooled
Fuel - quality - additives - alternate
Lubrication
-------
B-13
12. STARTING
Temperature
F.I.E. - excess fuel - nozzles - timing
D.I, vs I.D.I.
Fuel - quality - alternate - additives - filter clogging
Boosting
Aids - heater plugs - manifold heaters - ether
Diesel vs gasoline
User experience
Smoke
Compression ratio
Sequence
Battery /starter
Cranking speed
Valve timing
Lubricating oil
-------
8-14
13. HOT DRIVEABIUTY
Diesel vs gasoline
D.I, vs I.D.I.
Vehicle
User experience
Light load aids
F.I.E.
Response time
-------
8--15
14.- COED; EMVEABILITY
Dieaeli vs gasoling-
D.I. vs r.D.,L..
Vehicle
User experience
Light Icrad aids
Odor
Fuel — alternate. - quality
-------
8-16
15. TORQUE RISE
Boosting - comprex
F.I.E.
Diesel vs gasoline
I.D.I, vs D.I.
Valve timing
-------
8-17
16. DURABILITY
F.I.E.
Emissions devices
Boosting
Diesel vs gasoline
I.D.I, vs D.I.
Vehicle
Test bed
User experience
Economics
Air vs water cooling
Combustion chamber
Lubrication
Exhaust system
Wear
-------
8-18
17. COOLANT HEAT LOSSES
Boosting
Diesel vs gasoline
D.I. vs I.D.I.
Vehicle heaters
Test bed
User experience
Specific heat loss
Air vs water cooling
Warm-up time
Thermal loading
Temperature control
Oil cooler
Piston cooling
-------
S-19
18. FIRE RISK
Legislation - fuel tanks - general
Diesel vs gasoline
Vehicle
User experience
Statistics
Exhaust system - catalyst
Fuel
-------
8-20
19. VIBRATION AND TORQUE RECOIL
Engine mountings
Diesel vs gasoline
Vehicle
User experience
Configuration
Engine installation
Damping
-------
8-21
20. IDLE NOISE
Vehicle - absolute - subjective
F.I.E. - pilot injection - injection rate - nozzle - timing
Diesel vs gasoline
D.I, vs I.D.I.
User experience
Public opinion
Air vs water cooling
Configuration
Fuel - quality
Combustion
Mathematical models
Shields
-------
8-22
21. RELIABILITY
F.I.E.
Booster
Diesel vs gasoline
D.I, vs I.D.I.
Vehicle
Test bed
User experience
Economics
Air vs water cooling
Lubrication
-------
8-23
22. ECONOMICS
F.I.E.
Boosting
Diesel vs gasoline
D.I, vs I.D.I.
Vehicle
User experience
Specific costs
Legislation - smoke - emissions - odor - noise
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8-24
23. MANUFACTURE
F.I.E. - economics - technology
Combustion system - economics - methods - tolerance
Diesel vs gasoline - economics - technology
D.I, vs I.D.I.
Air vs water cooling
Configuration
Tolerance
-------
«-25
24- PEaFOBMAiNCE
Vehicle - legislation - boostiiKj - diesel vs gasoline
- D.I. vs I.D.I- - 'durability - torque rise
- road test - F.I.E.
Test bed - fuel additives - F.I.E. - combustion chamber
- combustion - fuel quality - E.G.R. - boosting
- diesel vs gasoline - D.I. vs I.D.I. - specific
- durability - friction - volumetric efficiency
- thermal efficiency - compression ratio - cylinder
- pressure - fuel alternate - dual fuel - swirl
- valve timing - ambient conditions
- component. temperatures
Mathematical models
-------
8-26
25. DESIGN
Configuration
Legislation
Combustion system
Boosting
Diesel vs gasoline
Mechanical - bearings + liners
Tolerances
Air cooled vs water cooled
-------
8-2:7
26. AKCILMARIES
Diesel vs gasoline - cost
Emission' Gorttroilers
Volume
Weight
Servopamp
-------
8-28
27. LUBRICATION
Water injection
E.G.R.
Diesel vs gasoline
D.I, vs I.D.I.
User experience
Durability
Oil change period
Oil consumption
Fuel - quality
F.I.E.
Cooling
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9-1
SECTION 9
APPENDIX 2 - LIST OF REFERENCES
As mentioned in section 4 the scope of the literature review was wide rang-
ing, covering internal Ricardo reports as well as freely available published
articles.
While many of the internal Ricardo reports are the result of Ricardo spon-
sored test programmes, others have arisen from research or development
projects which have been funded by individual clients of Ricardo. These
reports are regarded by Ricardo as confidential between themselves and their
client and because of this the report titles in several of the references have
been modified to ensure this confidential nature.
-------
9-2
1. Diesel emissions as predictors of observed Diesel Odor.
H.E. Dietzmann et al - SAE 720757
2. Noise of small indirect injection Diesel engines.
W.M. Scott - SAE 730242
3. Automotive Diesel engine noise and its control.
M.F. Russell - SAE 730243
4. Design aspects of Low Noise Diesel engines.
S.H. Jenkins et al - SAE 730246
5. Diesel engine exhaust smoke - its measurement, regulation and control.
M. Vulliamy & J. Spiers - SAE 670090
6. Diesel emissions as related to engine variables and fuel characteristics
Marshall & Fleming - SAE 710836
7. Design factors that effect Diesel emissions.
R-C. Bascom et al - SAE 710484
8. Pre-combustion chamber Diesel engine emissions - a progress report.
R.E. Bosecker & D.F. Webster - SAE 710672
9. Diesel exhaust - a European Viewpoint.
B.V. Millington & C.C.J. French - SAE 660549
10. A stratified charge multifuel military engine - a progress report.
E. Mitchell et al - SAE 720051
11. Exhaust characteristics of the automotive Diesel.
R.C. Schmidt et al - SAE 660550
12. Exhaust emission control in medium swirl rate direct injection Diesel
engines.
Parker & Walker - SAE 720755
13. Some effects of fuel injection system parameters on Diesel exhaust
emissions.
R.J. Hames et al - SAE 710671
14. The characterisation of odor components in Diesel exhaust gas.
R.S. Spindt et al - SAE 710605
15. The development, of the small automotive Diesel in western Europe and
its likely role in the U.S.A.
J.H. Pitchford - SAE 215B
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9-3
16. Reductions of emissions from Diesel engines.
C.J. Walder - SAE 730214
17. Measurement and evaluation of Diesel smoke.
R.C. Bascom et al - SAE 730212
18. Some problems encountered in the design and development of high speed
Diesel engines.
C.J. Walder - SAE 978A
19. Diesel fuel properties and exhaust gas - distant relations?
G. McConnell & H.E. Howells - SAE 670091
20. Daimler Benz passenger car Diesel engines - highlights of 30 years of
Development.
E. Eisele - SAE 680089
21. Combustion system parameters and their effect upon Diesel engine exhaust
emissions.
Pischinger & Cartellieri - SAE 720756
22. Towards higher speeds and outputs from the small Diesel engine.
D. Broome - SAE 730149
23. Diesel engine and highway truck noise reduction.
R.M. Law - SAE 730240
24. Cooperative evaluation of techniques for measuring nitric oxide and carbon
monoxide.
J.M. Perez et al - SAE 720104
25. Turbocharged Diesel engine performance at altitude.
J.W. Dennis - SAE 710822
26. Developing a new stratified charge combustion system with fuel injection
for reducing emissions in small farm and industrial engines.
M. Miyake - SAE 720196
27. Combustion Characteristics of rotary engines.
K. Yamamoto et al - SAE 720357
28. A new Diesel combustion chamber - the variable throat chamber.
B. Brisson et al - SAE 730167
29. A preliminary model for the formation of nitric oxide in D.I. Diesel
engines and its application in parametric studies.
S.M. Shahed et al - SAE 730083
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9-4
30. School fleet Dieselisation Program =
R. Davis - SAE 650711
31. Delivering the mail with Diesels - the post office department looks at
Diesel engines„
G.C. Nield - SAE 650710
32. An example of development in automotive small high speed Diesel
engines.
S. Tanaka et al - SAE 978C
33. Pilot injection in Diesel engines.
D.P. 17093 - 1973
34. Review of Combustion Chamber Characteristics.
D.P. 11353A - 1968
35. The automotive Diesel engine in perspective.
Donald - 1st Symposium on low pollution power systems development and
AAPS coordination meeting. Ann Arbor, Oct. 1973.
36. Development of an automotive particulate sampling device compatible
with the CVS system.
Musser & Bernstein - Esso publication - 1973
37. Smoke Measurement - Instruments & comparison of methods.
Dodd & Spiers - I.Mech.E. proc. 1968-69, 183 pt 3E p!57.
38. Influence of fuel properties on Diesel exhaust emissions.
Burt and Troth - I.Mech.E. proc. 1968-69, 183 pt 3E p!71.
39. Comparison of 2 litre D.I. and Comet V automotive engines.
D.P. 12897 - 1970
40. The Turbocharged Diesel as a Road Transport Power Unit.
Holmer & Haggh - I.Mech.E. 1970
41. The Breathing and Combustion Requirements of the Small High Speed
Diesel Engine - D. Broome - I = Mech.E. 1966
42. Where Phillips stands on the Stirling engine.
Automotive Engineering - July 1973 - p37
43. The state of Development of the Air Cooled Diesel engine.
Dr. Ing. M. Weidenmuller - Ric. Trans. 330
44. The Rotary engines of Yanmar outboard motor.
Yamaoka & Tado - SAE 710581
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9-5
45. Experiments with a single cylinder engine to aid development of highly
supercharged high speed Diesel engines.
MIRA Translation 14/72
46. Performance with economy - The RamAire System.
Percival & Ahrens - SAE 670109
47. Low Compression Ratio IDI Diesels with Glowplug added ignition - a
summary of relevant experiments.
D.P. 17555- 1974
48. Dealing with the design problems of present day Diesel engines.
Prof. Dr. H. List - Ric. Trans. 321
49. Results of development of the Daimler Benz OM 360 Diesel engine.
Dipl. Ing. Klinder & Ing. Kern - Ric. Trans. 318
50. Research on diesel engine exhaust pollution control.
Prof. R. Pischinger - ATZ - 1972
51. Influence of exhaust gas composition by means of exhaust gas recir-
culation in a pressure charged swirl chamber diesel engine.
Dipl. Ing. Manfred Fortnagel - MTZ - 1972
52. Some problems of current diesel engine development and the outlook
for other prime movers.
Urlaub - ATZ - 1972
53. Quieting the diesel with structural changes.
Automobile Engineering - 1972
54. Some investigations on cold starting phenomena in diesel engines.
Austen & Lyn - Proc. I.Mech.E. No. 5 - 1959-60
55. Contribution to the problem of starting and operating diesel vehicles
at low temperatures.
Prof. M. Brunner & Dr. H. Ruf - Proc. I.Mech.E. No. 5 - 1959-60
56. Improving temperature control in diesel engined vehicles.
Bisiker & Benford - Proc. I.Mech.E. (AD) No. 8 - 1961-62
57. Some experiences with a differentially supercharged diesel engine.
Dawson et al - Proc. I.Mech.E. Vol. 178, Part 2A No. 6 - 1963-64
58 . Vehicle particulate emissions.
Campbell & Dartnell - I.Mech.E. Conference - Air Pollution Control
in Transport Engines - 1971
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9-6
59. Experiments in the control of diesel engine emissions.
Torpey et al - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
60. Metabolism and dietics of "Moteur diesel compense B" and its result
on air pollution,
Brille & Baguelin - I.Mech.E. Conference - Air Pollution Control in
m , TT* • * f\ *7 -1
iranapoi'L c/uyines — IV il
61. Controlling exhaust emissions from a diesel engine by LPG dual fuelling,
Lyon et al - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
62. Stratification and air pollution,,
Witzky - I.Mech.E., Conference - Air Pollution Control in Transport
Engines - 1971
63. Diesel engine exhaust emissions and effect of additives.
Nesr et al - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
64. The mechanisms of soot release from combustion of hydrocarbon fuels
with particular reference to the diesel engine.
Broome and Khan - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
65. Recent automotive air pollution control legislation in the U.S. and a
new approach to achieve control: Alternative engine systems.
Brogan - I.Mech.E. Conference - Air Pollution Control in Transport
Engines - 1971
66. Prediction of soot and nitric oxide concentrations in diesel engine
exhaust.
Khan et al - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
67. Factors affecting emissions of smoke and gaseous pollutants from
direct injection engines.
Khan & Wang - I.Mech.E. Conference - Air Pollution Control in
Transport Engines - 1971
68. Exhaust emission control system for the rotary engine.
Muroki - I.Mech.E. Conference - Air Pollution Control in Transport
Engines - 1971
69. Symposium on Multi-fuel engines.
Gas & Oil Power 1959
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9-7
70. Compact long life diesel engine.
Timoney - I.Mech.E. Conference - Transport Engines of Exceptionally
High Specific Output - 1968
71. The differential compound engine: Interim test results and assessment
of future development.
Wallace et al - I.Mech.E. Conference - Transport Engines of Except-
ionally High Specific Output - 1968
72. Curtis-Wright's rotating combustion engine - compact, lightweight
power.
Jones - I.Mech.E. Conference - Transport Engines of Exceptionally
High Specific Output - 1968
73. Small high speed diesel engines.
Westwell & Blum - I.Mech.E. Conference - Mechanical design of diesel
engines - 1967
74. Effects of multiple introduction of fuel on performance of a compression
ignition engine.
C.P. Guptor et al - SAE 929A
75. The nature and cause of diesel emissions.
B.W. Millington - I.Mech.E. Conference - Motor vehicle air pollution
control - 1968
76. Some experiments on the mode of action of a diesel smoke suppressant
additive.
B.E. Knight & C.H.T. Wang - I.Mech.E. Conference - Motor vehicle
air pollution control - 1968
77. Natural gas engine for buses meets 1975 emission limits.
Automotive Engineer - June 1972
78. Mid-Range Diesels Mean Savings for Delivery Fleet.
Diesel and Gas Turbine - July 1973
79. Hybrid Diesel - Electric bus may help cities to lick noise-air-pollution
problems.
Automotive Engineering - August 1970 p42.
80. Influence of fuel properties and influence of anti-smoke additives on
Diesel exhaust smoke.
McConnell et al - SAE Journal, March '68 p36.
81. Novel supercharger retains smaller engine economy while providing
larger engine performance. SAE Jni. March '68, iro 62-65.
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9-8
82. Alternative non polluting power sources.
SAE Journal, Dec. '68 p40.
83. Noise in engineering and transportation and its effect on the Community.
T. Priede - SAE 710061
84. The use of specially designed covers and shields to reduce Diesel engine
noise.
G. Thien - SAE 730244
85. Diesel exhaust odor - its evaluation and relation to exhaust gas com-
position.
F.G. Rounds & H.W. Pearsall - SAE 863 - 1956
86. A sampling smokemeter for automotive Diesel engine testing.
E.G. Searle - SAE 436C - 1961
87. CRC Investigation of Diesel smoke measurement.
J.B. Durant - SAE 801A - 1964
88. The measurement and Control of Diesel exhaust smoke emission.
J.D. Savage - SAE 440B - 1962
89. The effect of the Vigom Process on the combustion in Diesel engines.
SAE 929B
90. 1971 European Diesel engine performance survey 10-500 bhp.
D.P. 13718
91. Nissan Emissions Results
Ex E.P.A. - 1973
92. 1DI Diesel Engine - Performance development and piston cooling tests.
D.P. 13679 - 1971
93. 6 cylinder IDI Diesel Engine, Progress Report No. 1, Jan-March 1971.
D.P. 13677 - 1971
94. For Passenger Cars Diesels Must Be Better.
C.J. Walder - SAE Jnl. Nov. '65, Vol. 73, No. 11, p78.
95. A European contribution to lower vehicle exhaust emissions.
(Eindhoven Conf. 1971) - D.P. 13538
96. Work carried out on IDI engine, August '69 - November '70.
D.P. 13277
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9-9
97. EDI engine - effect of valve timing on performance.
D.P. 12921 - 1970
3
98. Note on the effect of exhaust manifold arrangements on a 138 in
Comet V engine.
D.P. 12812 - 1970
99. NOX emissions from single cyl. engine in D.I. and Comet V builds.
D.P. 12429 - 1970
100. Design Developments in European Automotive Diesel engines.
Ricardo & Pitchford - SAE Journal (trans) Vol. 41-3 p405 - 1937
101. Exhaust Emissions Test Program on automotive IDI diesel engine and
dies el car.
D.P. 14442 - 1971
102. The Diesel is Friendly to its Environment.
Garthe - Deutz Publicity - 1971
103. Reducing Hydrocarbons and Odor in Diesel Exhaust by Fuel Injector
Design.
Ford et al - SAE 700734
104. Diesel Fuel Specification and Smoke Suppressant Additive Evaluations.
J.G. Brandes - SAE 700522
105. 4 cylinder Comet Mk V Engine - Progress Report no. 18.
D.P. 4680 - 1958
106. 6 cylinder 5.5 litre, Comet V, Progress Note No. 3. CARB Emissions
Tests.
D.P. 16524 - 1973
107. Performance of a Catalyst Box on a Direct Injection Engine.
D.P. 15867 - 1972
108. Effect of Inlet Air Humidity and Temperature on Diesel Exhaust Emissions,
S.R. Krause et al - SAE 730213
109. Surveying Tests of Diesel Smoke Suppression with fuel additives.
Takeshi Saito - SAE 730170
110. Factors Affecting Smoke and Gaseous Emissions from Direct Injection
Engines and a Method of Calculation.
I.M. Khan et al - SAE 730169
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9-10
111. Diesel Engine Noise Reduction by Combustion and Structural Modifi-
cations .
D.D. Tiede & D.F. Kabele - SAE 730245
112. The Light Duty Diesel Engine.
D.P. 16494 - 1973
113. 6 cylinder 5.5 litre. Comet V. Progress Note No- 4, CARR Emissions
Tests.
D.P. 16660 - 1973
114. 6 cylinder 5.5 litre, Comet V, Progress Note No. 5, GARB Emissions
Tests.
D.P. 16736 - 1973
115. The Diesel
R. Wakefield - Road & Track - Sept. 1973
116. 6 cylinder 5.5 litre, Comet V, Progress Note No. 6, CARB Emissions
Measurement.
D.P. 16878 - 1973
117. Noise Test on IDI engined vehicle.
D.P. 16947 - 1973
118. Engine and vehicle emissions reduction program.
Interim Report No. 1.
D.P. 17095 - 1973
119. NC>2 in Diesels - Ricardo Research.
D.P. 17138 - 1973
120. 6 cylinder IDI Diesel. Progress Note No. 4.
D.P. 14208 - 1971
121. 6 cylinder IDI Diesel. Progress Note No. 3.
D.P. 14068 - 1971
122. Investigation into the Compression Ratio limits of a Typical Comet III
type production engine.
D.P. 4494 - 1957
123. Note on the Effect of Nozzle Heat Shields on a typical Small Comet V
engine.
D.P. 4417 - 1957
124. Diesel car - Ricardo Report - Jan. 1956
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9-11
125. Diesel car fuel consumption with and without fluid drive.
Ricardo GO.3712 - 1956
126. Ricardo Memo - Diesel car - Nov. 1956
127. Comet V automotive diesel. Progress Report No. 5.
D.P. 9715 - 1966
128. Note on the Mechanical behaviour of the IDI diesel engine during test
running.
D.P. 3926 - 1956
129. Comet Mk V Diesel Engine - 6 cylinder.
D.P. 4566 - 1958
130. Progress Report on diesel engine - 4 cylinder. Ricardo Comet Mk V
Combustion System.
D.P. 3963 - 1956
131. Influence of Cetane Number on Compression Ignition Engine Performance.
D.P. 13740 - 1971
132. Diesel Car - Fuel & Oil Consumptions - Internal Ricardo memo 1.10.54
133. Diesel Car.
Ricardo GO.2930 - 1954
134. Diesel Car.
Ricardo GO.3279 - 1954
135. 4 cylinder, Comet Mk V. Report on an investigation into Factors
Affecting Auxiliary Hole Blocking of the Pintaux nozzle.
D.P. 4673 - 1958
136. Diesel fuel rating at high speed - Report on preliminary tests on the
El2 H.S. unit to observe ignition delay.
D.P. 4547 - 1958
137. 4 cylinder, Comet Mk V Progress Report No. 2.
D.P. 4443 - 1957
138. 4 cylinder, Comet Mk V Engines - Progress Report No. 12.
D.P. 4550 - 1958
139. 4 cylinder, Comet Mk V Engines - Progress Report No. 13.
D.P. 4573 - 1958
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9-12
140. 4 cylinder, Comet Mk V Engines - Progress Report No. 15.
D.P. 4617 - 1958
141. Induction Ram.
D. Broome - Automobile Engineer - April, May, June 1969.
142. Report on the Conversion of a Ford Zephyr car to diesel.
D.F. 12093 - 1969
143. V8 IDI Diesel engine. Progress Note No. 10.
D.P. 5761 - 1960
144. Further supercharging tests on a 6 cylinder Comet V engine.
D.P. 5379 - 1960
145. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 1.
D.P. 5084 - 1959
146. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 2.
D.P. 5137 - 1959
147. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 3,
D.P. 5294 - 1959
148. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 4,
D.P. 5353 - 1960
149. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 6,
D.P. 5574 - 1960
150. 4 cylinder automotive Comet Mk V Engine - Progress Note No. 9.
D.P. 5678 - 1960
151. Diesel Car - Conversion to Comet V.
D.P. 3405 - 1955
152. Diesel engine - Modifications to Combustion Chamber.
D.P. 3394 - 1955
153. Report on work carried out on diesel car.
D.P. 3672 - 1955
154. The Ricardo Comet V Combustion System.
D.P. 5834 - 1960
155. Notes on the Nozzle life of the diesel engined vehicle.
D.P. 5545 - 1960
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9-13
156. Diesel engined vehicle - (Effect of mods to combustion chamber).
D.P. 4881 - 1959
157. Further Notes Regarding Nozzle Life on the diesel engined vehicle.
D.P. 5754 - 1960
158. V8 IDI diesel engine - Progress Note No. 2.
D.P. 5020 - 1959
159. V8 diesel engine - Progress Note No. 4.
D.P. 5158 - 1959
160. 4-J litre Comet Mk V - Progress Report No. 1.
D.P. 5169 - 1959
161. V8 IDI diesel engine - Progress Note No. 6.
D.P. 5213 - 1959
162. 4-J litre Comet Mk V six cylinder engine - Progress Report No. 2.
D.P. 5260 - 1959
163. V8 IDI diesel engine - Progress Note No. 7.
D.P. 5323 - 1959
164. Quiet Idling Investigations - Assessment of Diesel Car Fitted with
Quiet Idling Device.
D.P. 11615 - 1969
165. A Technical Appraisal of the Clarke Rotary Piston Machine as a Basis
of an Internal Combustion Engine. - D.P. 16199 - 1973
166. A new class of Rotary Piston Machine suitable for Compressors, pumps
and internal combustion engines.
Clarke et al - SAE Proceedings Vol. 186 - 62/72.
167. 6 cylinder D.I. Engine Performance Calibration.
D.P. 15811 - 1972
168. Results from a short series of CARB Emission Tests run on a Comet
V Diesel Engine.
D.P. 16815 - 1973
169. A Study of the Exhaust Emissions Characteristics of a Turbocharged
Direct Injection Engine.
D.P. 16272 - 1973
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9-14
170. A Study of the Exhaust Emission Characteristics of a Naturally
Aspirated Direct Injection Engine.
D.P. 16470 - 1973
171. A Summary of a Research Project - Diesel Emissions. S.N. 16374.
172. Unclassified data sheet listing Comet engines - 1973.
173. Meeting between a client and Ricardo.
D.P. 17358 - 1973
174. A review of low noise Diesel engine design at ISVR.
Grover & Lalor - Jnl. Sound & Vib. No. 3, 1973 p403.
175. A review of vehicle noise studies - with a reference to some recent
research on petrol engine noise.
J.A. Raff & R.D.H. Perry - Jnl. Sound & Vib. No. 3 1973 p433.
176. Comprex Supercharging - a progress report - Presentation to E.P.A.
by A . Mayer, Brown Boveri.
177. Comprex Supercharger for Passenger Car Diesel engines.
Prof. E. Eisele - D.Benz report to E.P.A. - Sept. '73
178. Notes on a visit to Brown Boveri, re. Comprex, Dec. 1972.
E.P.A. note by C.F. Bachle
179. Characterization of particulates and other non-regulated emissions from
Mobile sources and effects of exhaust emissions control devices on
these emissions. Report for E.P.A, APTD 1567, 29 Oct. '73. Pages
160-163 - Mercedes 220D vehicle.
180. A Report on the Emission Performance of the Ford Stratified Charge
Engine Using the 1975 Test Procedure.
E.P.A. Report 72-2 - August 1971
181. An evaluation of three Honda compound Vortex controlled combustion
(CVCC) Powered vehicles.
E.P.A. Report 73-11 - Dec. 1972
182. Exhaust Emissions Analysis of Two Wankel Powered Cars Furbished by
the U.S.A.T.A.C.
E.P.A. Report 73-10 - Sept. 1972
183. Evaluation of the Texaco Stratified Charge (TCCS) M151 Army Vehicle.
E.P.A. Report 73-27 - June 1973
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9-15
184. Vehicle Test Report No. 3 - Peugeot diesel.
E.P.A. Report - 16th July '73
185. Vehicle Test Report No. 2 - Opel diesel.
E.P.A. Report - 16th July '73
186. Vehicle Test Report No. 8 - '72 EPA Ford.
E.P.A. Report - 7th Sept. '73
187. Vehicle Test Report No. 7 - EPA Ford.
E.P.A. Report - 7th Sept. '73
188. Final Report: Exhaust Emissions from a Mercedes-Benz Diesel Sedan.
E.P.A. Report 73-6 ~ July 1972
189. Emissions from a pick-up truck retrofitted with a Nissan Diesel Engine.
July 1973 - From E.P.A. 29/10/73
190. EPA diesel study status -First symposium on low pollution power
systems development.
Oct. 19th 1973 - Presented by J.J. McFadden - from EPA 29/10/73
191. Exhaust Emissions from three diesel-powered passenger cars.
March 1973 - from EPA 29/10/73
192. M.W. Kellog Study, Ann Arbor, Michigan 48105.
Letter from J. DeKany to E. Stork, MSAPC - from EPA 21/9/73
193. Letter referring to Meeting with Mobil on Impact of Diesel on the oil
industry.
Letter dated 21.8.73 from T. Austin of EPA in reply to AFAE - from
EPA 21/9/73
194. Task 17 final report. Potential increased production of Automotive
diesel fuels. Submitted to EPA office of Air Programs Division of
Control Systems Contract No. CPA 70-68 - from EPA 21/9/73.
Submitted by M.W. Kellog Company, Texas, Dec. 1st 1972
195. Notes on a visit of representatives to Bridge Works on 26th Sept.
1962 - by C.J. Walder - D.P. 6817
196. Notes on the performance of wide cut gas oil in a light duty IDI
diesel vehicle.
D.P. 6820 - 3rd Oct. 1962
197. 6 cylinder 2/4 litre - Comet Progress Report No. 1. - Oct. 10th 1962
D.P. 6839
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9-16
198. Road fuel consumption for the petrol and diesel versions of the Austin
A60 and Morris Oxford saloon cars.
D.P. 7529 - 27th Nov. 1963
199. Notes on the 6-cylinder 6 litre engine.
D.P. 6806 - 26th Sept. 1962
200. Ricardo High Speed Engine Pump Injector Tests. Interim Report.
D.P. 15200 - 1972
201. Report on a program of noise measurements carried out on an IDI
Diesel Engine.
D.P. 15247 - 1972
202. List of Ricardo Comet V Diesel Engines currently in production.
D.P. 14378 - 1971
203. Diesel Engine Exhaust Emissions. Report No. 5.
D.P. 14168 - 1971
204. Client/Ricardo Technical Meeting.
D.P. 14278 - 1971
205. 3 cylinder Comet V.
D.P. 4381 - 1957
206. 4 cylinder Comet Mk V Engines. Report No. 17.
GO File 4661 -1958
207. 6 cylinder 4^ litre engine fitted with Ricardo Comet V cylinder head.
D.P. 5652 - 1960
208. An analysis of some engine test results on an IDI diesel engine.
GO File 4725 - 1958
209. 4 cylinder Comet Mk V Engine.
GO File 4776 - 1958
210. 4 cylinder Comet Mk V Diesel Engines. Report No. 18.
GO File 4817 - 1958
211. Supercharging tests on a Comet V Engine.
GO File 5233 - 1959
212. Report on Diesel Engine 6 cylinder. - Ricardo Comet V head.
GO File 458 - D.P. 4458 - 1958
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9-17
213. Volumetric Efficiency of Comet Engines.
D.P. 8066 - 1964
214. Emission tests on a 2.25 litre Diesel Engined vehicle.
D.P. 16800 - 1973
215. Road test on a diesel car. Unclassified document.
216. Noise tests on a diesel vehicle. 4 cylinder 2 litre Comet Mk V engine.
GO File 4933 - 1959
217. 4 cylinder Comet Mk V diesel engine. Report No. 1.
GO File 4409 - 1957
218. Diesel engine exhaust emissions. Report No. 6.
D.P. 14443 - 1971
219. Estimated friction losses for 6 cylinder Comet Mk V and D.I. engines.
R & Co. Drg. No. D18662 - 1966
220. Hybrid heat engines/electric systems. A summary of a feasibility
study carried out.
D.P. 14670 - 1972
221. Interim report of noise measurements of a BLMC 2.521 engine.
D.P. 14767 - 1972
222. Ford Zephyr car fitted with a 4 cylinder IDI diesel engine.
1975 CVS Emission Tests.
D.P. 14913 - 1972
223. Some notes on the selection of cylinder size and arrangement for small
high speed automotive engines using the Ricardo Comet Mk V combus-
tion system.
D.P. 15108 - 1972
224. Timed inlet and exhaust systems for a 6 cylinder prechamber diesel
engine.
D.P. 15130 - 1972
225. The prospects of the diesel engine to satisfy the U.S. pollution regul-
ations .
D.P. 15448 - 1972 and R & Co. Memo 1973
226. 6 cylinder engined diesel vehicle.
D.P. 15537 - 1972
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9-18
227. 4 cylinder 2 litre Comet V. Performance calibration in modified build.
D.P. 16210 - 1973
228. Future Power Units for Automobiles and Commercial Vehicles.
D.P. 16256 - 1973
229. Opel Rekord Diesel 2100D.
Described briefly in ATZ Vol. 75, No. 1 - 1973 p27.
230. Preliminary study of a low emissions diesel engine for use in passenger
cars.
D.P. 16424 - 1973
231. Summary of Ricardo results showing effect of fuel quality on diesel
engine emissions.
D.P. 16484 - 1973
232. The dynamic injection timing requirements of small high speed diesel
engines using Ricardo Comet combustion chambers.
D.P. 4143 - 1957
233. The unloading mechanism in distributor pumps. Diesel engine with 3 mm
fuel pipes.
D.P. 4170 - 1957
234. Further thoughts on the mechanism of unloading in distributor pumps.
D.P. 4145 - 1957
235. Comet V fuel injection pump tests.
D.P. 3910 - 1956
236. Comet V fuel injection pump tests.
D.P. 3990 - 1956
237. Comet V fuel injection pump tests.
D.P. 3995 - 1956
238. Comet V fuel injection pump tests.
D.P. 3550 - 1955
239- Test to establish the acceptable range of injection timings on the Comet
III diesel engine when fitted with a 6 mm. fuel pump.
D.P. 4255 - 1957
240. Report on the work carried out on a high ratio Comet Mk V engine bet-
ween April - November 1958.
D.P. 4862 - 1958
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9-19
241. Comet Mk III diesel engine - Comparison of exhaust smoke rating with
high and low compression ratio.
D.P. 4899 - 1959
242. Comparison tests on 2 /4 and 2 litre engines using fuels with and
without the addition of ^% by volume isopropyl nitrate.
D.P. 4221 - 1957
243. Ricardo letter to Client - 28th Jan. 1966
244. IDI Diesel engine.
GO File 3971 - 1956
245. Modifications carried out on a diesel car.
GO File 4508 - 1958
246. Report on Comet V Diesel Engine 6 cylinders.
GO File 4718 - 1958
247. 4 cylinder Comet Mk V Engine. Report No. 22.
GO File 5309 - 1959
248. Note on high speed running of an IDI diesel engine.
D.P. 5464 - 1960
249. Notes on condition of nozzles in a diesel car.
GO File 5469 - 1960
250. Comments on indicator diagrams and peak pressure measurements
obtained on an IDI diesel engine.
GO File 5612
251. Diesel Estate Car.
GO File 5297 - 1959 .
252. 3 cylinder Comet V engine cold starting tests.
GO File 4586 - 1958
253. Diesel Estate Car.
GO File 5494 - 1960
254. 4 cyl. IDI diesel engine.
D.P. 5785 - 1960
255. The influence of fuel composition on emissions of carbon monoxide
and oxides of nitrogen.
Carr et al - Univ. of California - SAE 700470
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9-20
256. Investigation of Combustion Phenomena in a Swirl Chamber Compression
Ignition Engine using Schlieren Techniques.,
A<,B. Allen and A. Khin - Univ. of Toronto - SAE 700500
257o The FL413 - A New Series of Deutz Air-Cooled V type Diesel Engines.
Hans-Ulrich Houe - Deutz - SAE 700028
258. Diesel emissions as related to engine variables and fuel characteristics.
Marshall and Fleming - U.S. Bureau of Mines - SAE 710836
259 o Automotive Particulate Emissions and their Control.
Habibi - Du Pont de Nemours - SAE 710638
260o Emission Characteristics of Natural Gas as an Automotive Fuel.
Fleming and Allsup - Bureau of Mines - SAE 710833
261. The variable - Displacement Engine: An Advanced concept powerplarit.
Welsh and Riley - Thermo Mechanical Systems - SAE 710830
262. The effect of exhaust system geometry on exhaust dilution and Odor
intensity.
Colucci and Barnes - G.M. Corp. - SAE 710219
263. Cooperative evaluation of techniques for measuring hydrocarbons in
diesel exhaust.
(A CRC report) Wagner and Johnson - SAE 710218
264. Effectiveness of exhaust gas recirculation with extended use.
Musser et al - SAE 710013
265o Influence of Operating Cycle on Noise of Diesel Engines.
Anderton & Baker - ISVR - SAE 730241
266. Reducing noise from heavy diesel trucks by engine compartment
shielding.
Ronnhult -• Saab Scania - SAE 730682
267. Diesel engine noise reduction hardware for vehicle noise control.
Jenkins & Kuehner - Cummins Engine Co. - SAE 730681
268. Photochemical Reactivity of Diesel Exhaust.
Bureau of Mines ~ RI 7514 - 1971
269 o Diesel emissions reinventoried.
Bureau cf Mines - RI 7530 •- 1971
270. Merits of the Ricardo combustion systems.
Letter to Editor - Gas and Oil Power - 1953
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9-21
271. The small high speed diesel engine.
B.V. Millington - Proc. I.Mech.E. Vol. 18 pt IV - 1965
272. Optimisation of diesel engine parameters at high specific output.
D. Broome - UDK 629. 1-445: 621.436.003.1 - 1973 (Conference -
Belgrade Yugoslavia)
273. Reducing exhaust emissions from diesels.
Automobile Engineer - 1971
274. Looking in on diesel combustion.
W.M. Scott - SAE 690002
275. Recent developments in diesel engine research at the Ricardo labora-
tories .
W.M. Scott - Entropie No. 48 - 1972
276. Some problems encountered in the design and development of high speed
diesel engines - C.J. Walder - SAE 978A - repeat of ref. 18
277. What problems still restrain the small automotive diesel engine.
Pitch ford et al - FISITA Conference 1964
278. Some more light on diesel combustion.
Alcock & Scott - Proc. I.Mech.E. 1962-63 No. 5
279. Development of a High Speed Four Cylinder Diesel Engine under con-
sideration of the existing machine tools equipment for the production
of the gasoline engines.
H. Weitzel - I.Mech.E. Conf. Publication 19 - 1973
280. High Speed Diesel Engines.
H.R. Ricardo - Institution of Automobile Engineers - 1926-30
281. Diesel engines.
H.R. Ricardo - Royal Society of Arts - 1931
282. High Speed diesel engines applied to Motor Vehicles.
R.L. Stafford - Newark Engineering Society - 1932
283. High Supercharging of a C.I. Engine.
C.W.R. Smith - Engineer - 1938
284. Looking at diesel combustion.
Hempson & Scott - New Scientist - Vol. 6 pp 1134-37.
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9-22
285. Diesel engines for road transport.
Sir H.R. Ricardo - Engineering - 1948
286. Ricardo Combustion Systems for High Speed Compression Ignition
Engines. Unclassified publication - 1947.
287. Papers on small high speed diesel engines.
Proc. I.Mech.E. 1954-55 - No. 8 - pp 209-246
288. Road Test Reports - Peugeot 204, Mercedes 220D, 240D & 230/4,
Opel Rekord, and Peugeot 504. Diesel & Gasoline Comparisons
(French) L1 Action - Feb. 1974
289. Some problems arising from a wider use of the small diesel engine.
J.H. Pitchford - The Oil Engine and Gas Turbine - 1953
290. Future of the high speed reciprocating internal combustion engine.
J.H. Pitchford - CME - 1960
291. The high speed diesel engine.
H.R. Ricardo - Diesel engine users association - 1930
292. Diesel smoke suppression by fuel additive treatment.
C.O. Miller - SAE 670093
293. Smoke reduction in diese] engines.
A.W. Carey- SAE 670224
294. Diesel engine exhaust smoke. The influence of Fuel Properties and
the effects of using Barium-containing fuel additive.
D.W. Golothan - SAE 670092
295. The differential compound engine.
F.J. Wallace - SAE 670110
296. Factors influencing diesel emissions.
Marshall and Hurn - SAE 680528
297. Relation of lean combustion limits in diesel engines to exhaust odor
intensity.
Barnes - SAE 680445
298. Effect of design revisions on two stroke cycle diesel engine exhaust.
D.F. Merrion - SAE 680422
299. A flame ionisation technique for measuring total hydrocarbons in diesel
exhaust.
Johnson et al - SAE 680419
-------
9-23
300. Some Notes on Diesel Cold Starting Aids.
D.P. 17750 - 1974
301. Relation between noise and basic structural vibration of diesel engines.
T. Priede et al - SAE 690450
302. Does turbocharging increase diesel engine noise? - observations on the
generation, emission and reduction of diesel engine noise.
R. Wolfgang Hempel. - SAE 680406
303. The noise problem of air cooled diesel engines - measures towards
its reduction with general observations and specific results.
O. Cordier& G. Reyl - SAE 680405
304. Unaided starting of diesel engines.
T.W. Biddulp & W.T. Lyn - SAE 680103
305. Development of Mitsubishi DC2 Series V-Type Diesel Engine.
Kenj Okamura & Kiji Yamanda - SAE 690745
306. Measurement of Automotive Passby Noise.
Ralph K. Hillquist & Richard A. Battis - SAE 725275
307. Motor Vehicle Noise identification and analysis of situations contri-
buting to annoyance.
William J. Galloway & Glenn Jones - SAE 725276
308. Future trends in energy conversion systems.
H -W. Welsh - SAE 660603
309. A method for estimating and graphically comparing the amounts of
air pollution emissions attributable to automobiles, buses, commuter
trains and rail transit.
J.W. Scheel - SAE 720166
310. Factors affecting diesel smoke in highway operation.
W.A. Howe - SAE Golden anniversary diesel engine meeting - 1955
311. Dimensions of diesel fuel performance: Design, depressants and
response.
B.L. Michel and L.D. Fergesen - SAE 660371
312. Conversion of high speed,air cooled diesel engines from precombustion
chamber process to direct injection.
H. Ldnnenkohl - SAE 660010
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9-24
313. Future development of free piston gasifier turbine combinations for
vehicle traction.
F.J. Wallace et al - SAE 660132
314. The role of flow improvers in solving autodiesel winter fuel problems.
K.A. Beyreis et al - SAE 660372
315= Piston Cooling.
C.C.J. French - SAE 720024
316. Noise source definition - interior vehicle noise.
R.J. Vargovick - SAE 720274
317. The effect of fuel and vehicle variables on polynuclear aromatic hydro-
carbon and phenol emissions.
G.P. Gross - SAE 720210
318o Reduction of diesel smoke in California.
M.L. Brubacher - SAE 660548
319. Effect of air swirl on smoke and gaseous emissions from D.I. diesel
engines.
I.M. Khan et al - SAE 720102
320. Comet V automotive diesel engine - Progress Report No. 4.
D.P. 9458 - 1966
321. Analysis of reported tests on a 6 cylinder turbocharged engine.
D.P. 9451 - 1966
322. Comet V automotive diesel engine.
D.P. 9419 - 1966
323. Ricardo Petter high speed 80 x 95 mm D.I.
D.P. 9361 - 1966
324. 1.5 litre diesel car.
D.P. 9319 - 1966
325. Report of some tests carried out on a 1.5 litre diesel car.
D.P. 9254 - 1966
326. Comet V automotive diesel engine.
D.P. 9241 - 1966
327. 4 cylinder IDI diesel engine of about 3 litres.
D.P. 9231 - 1966
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9-25
328. 4 cylinder ID I dies el engine of about 3 litres.
D.P. 9056- 1966
329. Comet V automotive diesel engine.
D.P. 9048 - 1966
330. 6 cylinder indirect combustion diesel engine 5^ litre.
D.P. 10794 - 1968
331. 1.5 litre Comet VA (4 cyl).
D.P. 10700 - 1968
332. PHS 80 x 95 mm engine. A comparison of toroidal D.I. and Comet
Mk V combustion systems.
D.P. 10698 - 1968
333. Comet V automotive diesel engine.
D.P. 10591 - 1968
334. Comet V automotive diesel engine.
D.P. 10485 - 1967
335. Comet Vb engine, 4 cyl, 2^ litres.
D.P. 10191 - 1967
336. Comet Vb engine, 4 cyl, 2£ litres. D.P. 10042 - 1967
337. Comet V automotive diesel engine.
D.P. 10018 - 1967
338. 2 litre Comet V - Heat losses to water and lubricating oil.
D.P. 9942 - 1967
3
339. 4 cylinder 2/4 litre D.I. engine.
D.P. 9931 - 1967
340. Comet Vb engine, 4 cyl, 2^ litres.
D.P. 9924 - 1967
341. Toothed belt timing drives endurance tests on a 2 litre diesel engine.
D.P. 9841 - 1967
342. Comet Mk Vb engine, 4 cyl, 2£ litres.
D.P. 9840 - 1967
343. Comet Mk Vb engine, 4 cyl, 2\ litres.
D.P. 9814 - 1967
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9-26
344. Some notes on the theory of induction ram.
D.P. 9745 - 1967
345. The application of induction ram techniques to four stroke diesel engines,
D.P. 9740 - 1967
3;
346. 4 cylinder 2/4 litre D.I. engine..
D.P. 9690 - 1966
347. The position in regard to small engine D.I. development to Nov. 1966.
D.P. 9634 - 1966
348. 4 cylinder IDI diesel engine of just less than 3 litres.
D.P. 9480 - 1966
349. 4 cyl. diesel Comet V conversion.
D.P. 13320 - 1970
350. 6 cyl. IDI diesel.
D.P. 13811 - 1971
351. Exhaust emissions of an IDI diesel engine.
D.P. 13766 - 1971
352. 1000 hr endurance test of a DI diesel engine.
D.P. 11449 - 1968
353. Notes on work with ram pipes on an IDI diesel engine.
D.P. 11507
354. 1.5 litre Comet VA.
D.P. 11020 - 1968
355. 2.5 litre Comet Mk Vb engine.
D.P. 11105 - 1968
356. Possibilities for induction ram on small high speed diesel engines.
D.P. 11166 - 1968
357. Notes on exhaust manifold layout for four stroke naturally aspirated
engines and its effect on engine performance.
D.P. 11167 - 1968
358. Exhaust emission investigation on a typical prechamber diesel engine.
D.P. 11187/11188 - 1968
-------
9-27
359. Comet V automotive diesel engine.
D.P. 10848 - 1968
360. 4 cyl. 23/4 litre Comet V.
D.P. 10975 - 1968
361. Diesel engine design study.
D.P. 10990 - 1968
362. Turbocharged loop scavenge engine.
S.H. Henshall - Proc. I.Mech.E. Vol. 177 No. 6 - 1963
363. The future of lightweight high-speed diesel engines in the automotive
and agricultural fields.
J.G. Dawson & N.M.F. Vulliamy - Proc. I.Mech.E. Vol. 177 No. 38 -
1963
364. Nitrogen oxides and other toxic gases in diesel engine exhausts.
H. Middledich - Proc. I.Mech.E. Vol. 179 Part 1 1964-65
365. The diesel engine in association with the gas turbine.
E. Chatterton - Proc. I.Mech.E. Vol. 174 No. 10 - 1960
366. The high speed heavy duty diesel engine, its development, design and
application.
E. Schmidt - Proc. I.Mech.E. Vol. 174 - 1960
367. Noise and vibration problems in commercial vehicles.
T. Priede - J. Sound Vib. (1967) 5(l), 129-154.
368. A Summary of Turbocharging Investigations in 1966.
D.P. 17543 - 1974
369. Automotive Piston engine noise and its reduction - a literature survey.
W.W. Soroka & C-S.F. Chien - SAE 690452
370. Engineering know-how in engine design. Part 20. SAE SP-369.1972
Acoustics and noise. Vehicle noise measurement. Noise legislation
on product design. Abatement of structural noise. Practical noise
control.
371. The diesel engine as a source of commercial vehicle noise.
P.E. Waters et al - Proc. I.Mech.E. - 1970
372. Statistical investigation into diesel engine noise.
Dr. Ing. W. Hempel - CIMAC Working Group Paper - given to I.Mar.E.
(r V=-<- V IT,,' -?o%
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9-28
374. The Performance of BP Vanellus S3-20W in an Austin Taxicab.
B.P. Technical Bulletin - 1972
375. Origins of reciprocating engine noise.
D. Anderton et al - ASME 70-WA/DGP-3.
376. Some studies into origins of automotive diesel engine noise and its
control.
T. Priede - FISITA Congress - 1966
377. Identification of mechanical sources of noise in a diesel engine: sound
emitted from the valve mechanism.
B.J. Fielding & J. Skorecki - Proc. I.Mech.E. Vol. 181 Part I No. 19
1966-67
378. Identification of mechanical sources of noise in a diesel engine: sound
originating from piston slap*
B.J. Fielding & J. Skorecki - Proc. I.Mech.E. Vol. 184 Part I No. 46
1969-70
379. Medium speed diesel engine noise.
R. Bertodo & J.H. Worsfold - Proc. I.Mech.E. Vol. 183 Part I No. 6 -
1968-69
380. On piston slap as a source of engine noise.
D. Ross & E.R. Ungar - ASME 65-OGP-10
381. Urban noise legislation.
C. Caccavari - SAE 720902
382. Air swirl on a road vehicle diesel engine.
D. Fitzgeorge & J.L* Allison - Proc. I.Mech.E. (A.D.) No. 4 -
1962-63
383. Particulate sampling.
D.P. 12553 - 1970
384. Single cylinder engine. Investigation into the merits of the D.I. and
Ricardo Comet Va Combustion systems in regard to exhaust emissions.
D.P. 12080 - 1969
385. Tests on a structures research car.
D.Po 12162 - 1969
386. Air utilisation in diesel engines.
D.P. 12205 - 1969
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9-29
387. Slant 4 Comet Mk Vb. Note on light spring diagrams of cylinder and
inlet manifold pressures obtained at 4500 rpm.
D.P. 12288 - 1969
388. Exhaust emission investigation on anIDI diesel saloon car.
D.P. 11603 - 1969
389. Exhaust Control of Diesel Engine (EGR & Water Injection)
M. Tashiro, Isuzu Motors - E.P.A. Report Feb. 1974
390. Small diesel engines for the 1970's.
D.P. 11832- 1969
391. Torque back up in current automotive diesel engines.
D.P. 11284 - 1968
392. Motoring compression pressures.
D.P. 7569 - 1963
393. Note on sound insulation carried out by Ricardo on a compact diesel
sedan.
D.P. 7599 - 1964
394. 4 cylinder Comet V.
D.P. 7723 - 1964
395. Injector Maintenance.
D.P. 7735 - 1964
396. Summary of incoming reports concerning an automotive IDI diesel
engine.
D.P. 7776 - 1964
397. 4 cyl. Comet III diesel engine - Progress Report No. 1.
D.P. 8416 - 1965
398. Report on the calibration tests of a 4 cylinder IDI diesel engine.
D.P. 8465 - 1965
399. Report on the recalibration tests of a 4 cylinder IDI diesel engine with
combustion chamber to latest specification.
D.P. 8639 - 1965
400. 4 cyl. Comet Va (diesel engine) - Progress Report No. 2.
D.P. 8722 - 1965
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9-30
401. Comet V automotive diesel engine.
D.P. 8748 - 1965
402. Diesel car road tests.
D.P. 8770 - 1965
403. Survey of European Diesel engines.
D.P. 8811 - 1965
404. Diesel car road tests.
D.P. 8934 - 1965
405. Note on further induction ram tests on an IDI diesel engine.
D.P. 12083 - 1969
406. 4 cylinder Comet Mk V.
D.P. 7248 - 1963
407. 4 cyl. Comet Vb 2.8 litre (normally aspirated high speed version).
D.P. 7158 - 1963
408. 4 cyl. Comet Va 2.8 litre (normally aspirated high speed version)
D.P. 7159 - 1963
409. 4 cylinder Comet Mk V.
D.P. 7976 - 1964
410. 4 cylinder Comet Mk V.
D.P. 7554 - 1963
411. 4 cyl. 2^ litre diesel engine.
D.P. 7981 - 1964
412. Noise tests of two diesel sedans.
D.P. 8047 - 1964
413. 2 litre Comet V.
D.P. 8220 - 1964
414. California exhaust air pollution test on a Diesel car.
D.P. 8264 - 1964
415. 2 litre Comet V.
D.P. 8298 - 1965
416. Comet Mk V diesel engine.
D.P. 8310 - 1965
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9-31
417. 2 litre Comet V. .
D.P. 8365 - 1965
418. 3 litre" Comet V.
D.P. 8386 - 1965
419. Report on calibration tests of the 4 cylinder in line diesel engine after
conversion to Ricardo Comet V combustion system
D.P. 8839 - 1965
420. Report on the road testing of a diesel car.
D.P. 8840 - 1965
421. 1 litre Comet V engine (4 cyl) low speed.
D.P. 8570 - 1965
422. 4 cyl. diesel engine.
D.P. 8910 - 1965
423. Notes on cold starting of diesel engines.
D.P. 13905 - 1971
424. Comet V automotive diesel engine.
D.P. 10164 - 1967
425. Application of the flash temperature concept to cam and tappet wear
problems.
A. Dyson & H. Naylor - Proc. I.Mech.E. 1960-61 No. 8
426. Measurement of piston ring and cylinder liner wear by using radio-
active tracers.
R.F. Pywell and S.T. Walker - Proc. I.Mech.E. (AD) 1960-61 No. 8
427. Fuel injection system calculations.
B.E. Knight - Proc. I.Mech.E. (AD) No. 1 1960-61
428. Calculations of the effect of heat release on the shape of the cylinder
pressure diagram and cycle efficiency.
W.T. Lyn - Proc. I.Mech.E. (AD) No. 1 1960-61
429. Air cooled automobile engines.
J. Mackerle - Proc. I.Mech-.E. (AD) No. 2 1961-62
430. Relation between fuel injection and heat release in a direct injection
engine and the nature of the combustion process.
A.E.W. Austen & W.T. Lyn - Proc. I.Mech.E. (AD) No. 1 1960-61
-------
9-32
431. Effect of engine structure on noise of diesel engines.
T. Priede, A.E. Austen & E.G. Grover - Proc. I.Mech.E. Vol. 179
Part 2A No. 4 1964-5
432. Automobiles and Petroleum: Past, present & future.
Prof. JoJo Broeze - Proc. I.Mech.E. (AD) No. 7 1953-54
433. Development and application of automotive fuels - diesel and gasoline.
A.E. Felt & G.C. Wilson (1954)
434. Some notes on the design, development and production of high speed
compression ignition engines.
S. Markland & N. Tattersall - Proc. I.Mech.E. 1947-8
435. Symposium on superchargers and supercharging.
Proc. I.Mech.E. No. 6 - 1956-7
436. Performance of vehicles under trans antarctic conditions.
D.L. Pratt - Proc. I.Mech.E. (AD) No. 6 - 1958-59
437. Commercial Vehicle performance and fuel economy.
G.L. Smith - SAE SP 355 (700194)
438. What can the turbocharger do for the engine?
W. Lang - SAE 660473
439. The influence of induction and exhaust system design on power pro-
ducing characteristics of diesel engines.
H.G. Holler - SAE 700535 (SP 359)
440. The turbocharger - a vital part of the engine intake and exhaust systems
W.E. Woolenweber - SAE 700534 (SP 359)
441. Recent developments in variable compression ratio engines.
J.C. Basiletti and E.F. Blackburne - SAE 660344 (SP 280)
442. Diesel combustion at high MEP with low compression ratio.
W.Po Mansfield & W.S. May - SAE 660343 (SP 280)
443. Design and development of a very high output multifuel engine.
R.T. Paluska et al - SAE 670520 (SP 290)
444. Daimler-Benz high output engine - a study in compact design.
O. Herschmann - SAE 670519 (SP 290)
445. Comparative analysis of Stirling and other combustion systems.
S.R. Davis & N.A. Henein - SAE 730620
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9-33
446. Light diesels give promise in mail service.
G. Nield U.S. P.O. Dept. - SAE Jnl., July 1966 Vol. 74 No. 7
447. New diesel combustion chamber gives smoother combustion and cleaner
exhaust.
W. Henny & R. Herrmann (Hispano-Suiza) SAE Jnl. July 1966 Vol. 74
No. 7
448. High speed diesel rates improvement.
SAE Jnl. Dec. 1960 - J.H. Pitchford
449. European small diesel comes of age in light commercial vehicles.
J.G. Dawson - SAE Jnl. - 1961
450. List of manufacturers in current production with engines having Comet
Mk III & Mk V chambers.
D.P. 7271
451~ 4 cyl. Comet V engine - Progress Note No. 2
D.P. 7317
452. 4 cylinder Comet Mk V. Report on tests carried out with two differ-
ent injection-purnps during the period March to June 1963.
D.P. 7290
453. 4 cyl. Comet Vb 2.8 litre normally aspirated high speed version engine
No. 2 - Progress Note No. 5. Noise analysis tests.
D.P. 7340
454. Relation between form of cylinder pressure diagram and noise in diesel
engines *
T. Priede - Proc. I.Mech.E. (AD) No. 1 1960-61
455. An investigation into factors which affect piston ring wear under starting
conditions.
J. Cree & J. Thiey - Proc. I.Mech.E. (AD) No. 8 - 1960-61
456. A comparison of the wear of small gasoline and diesel laboratory
engines.
W.C. Pike & P. Newman - Proc. I.Mech.E. (AD) No. 8 1960-61
457. MWM Diesel features a new pre-combustion chamber-excepts from paper
by H.L. Hockel - SAE Jnl. Sept. 1957
458. Diesel for vehicles under 5000 Ib -
based on discussion and paper by J.S. Bright - F. Perkins (Canada) Ltd.
SAE Jnl. Feb. 1959
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9-34
459. How to step up fuel economy -
based on paper by L.D. Conta & P. Durbetaki - SAE Jnl. March 1959
460. Total energy situation in the United States.
D.R. Oliver - SAE 730514 (SP 383)
461. Impact of automotive emissions regulations on gasoline demand.
D.H. Clewell and W.J. Koehl - SAE 730575 (SP 383)
462. Current and future trends in United States gasoline supply.
E.J. Cahill - SAE 730516 (SP 383)
463. Fuel consumption trends in today's vehicles.
H.C. MacDonald - SAE 730517 (SP 383)
464. Energy and the automobile - General factors affecting vehicle fuel con-
sumption.
G.J. Huebner Jr. & D.J. Gasser - SAE 730518 (SP 383)
465. Alternative powerplants.
J-J. Brogan - SAE 730519 (SP 383)
466. Synthetic fuels for transportation and national energy needs.
D.P. Gregory & R.B. Rosenberg - SAE 730520 (SP 383)
467. Mass transit impact on energy consumption.
R. Husted - SAE 730521 (SP 383)
468. Criteria for evaluating vehicle in use inspection/maintenance impact
on emissions and energy conservation.
J.C. Elston - SAE 730522 (SP 383)
469. Low voltage ignition will start cold diesel.
L.P. Atwell et al - SAE Jnl. May 1960
470. Diesels gaining: Some problems ahead. E.R. Klinge - SAE Jnl., -
May 1960
471. The I.H.C. D-301 Diesel.
W.E. Peterson - SAE Jnl. - Jan. '61
472. Mobile powerplant trends forecast.
H.W. Welsh - SAE Jnl. Aug. 1967
473. Some notes on stratified charge combustion systems.
D.P. 16744 - 1973
-------
9-35
474. A status report on the development of the 1973 California diesel
emissions standards.
R.C. Bascom & G.C. Hass - SAE 700671
475. Ricardo electronic smoke sensor.
D.P. 16547 - 1973
476. Diesel combustion research. A review of 20 years combustion
chamber development.
D.P. 11378 - 1968
477. Fuel consumption of diesel cars.
D.P. 17334 - 1973
478. The development of a variable compression ratio diesel engine which
can greatly reduce the diesels emissions of NOX.
J. Witzky - Unclassified Document - Ex. E.P.A. 1970
479. The light weight diesel engine.
D.P. 17311 - 1973
480. Diesel and gasoline fuel consumption comparisons.
D24923 Ricardo
481. Ethyl Ether best for starting cold diesels.
F. Nelson & C.J., Ulzheimer - SAE Jnl. March 1950 p42
482. Diesel compared with gasoline engine.
F.B. Lautzenhiser - SAE Jnl. June 1950 p73
483. Making diesel engines deliver greater power.
C.R. Maxwell - SAE Jn.l. July 1950 p48
484. How engine was developed for Worlds Fastest Diesel Car.
J.C. Miller & C.R. Boll - SAE Jnl. Oct. 1950 p53
485. Diesel economy stems from higher BTU usage. •
M.C. Horine - SAE Jnl. Oct. 1950 p92
486. Thermal gains seen as Diesel edge over gasoline engine.
F.G. Shoemaker - SAE Jnl. Nov. 1950 p°57
487. Power Booster Fuels for Diesel engines.
E.J. Mclaughlin et al - SAE Jnl. June 1952, p42.
488. Diesels can be started at - 65 F.
D.E. Kiltv et al - SAE Jr',.
-------
9-36
489. Un Diesel dans le monde du silence.
B. Carat - L1 auto-Journal 15 Dec. '73 p26
490. Peugeot 504 diesel.
Quattroruote (Italian Magazine) December '73
491. Ricardo curves D24426/A - 30, 24494 - 1973
492. Nitrogen oxides - diesel villian - gases attack oil and engine parts.
SAE Jnl. Heavy Duty Issue - Sept. 1963 - Article based on material
drawn from paper No. 714B
493. 2 Neue Motoren fur Mercedes-Benz.
Mittelklassewagen - MTZ Sept. '73
494. Results obtained on a single cylinder Comet V.
D.P. 6853 - 1962
495. Notes on preliminary road tests using wide cut fuel.
D.P. 6881 - 1962
496. Comet Vb diesel engine - Progress Note No. 5.
D.P. 6905 - 1962
497. Tests to determine firing requirements to avoid misfire at all con-
ditions in an IDI diesel engine.
D.P. 7038 - 1963
498. Comet V - tests with a distributor type pump.
D.P. 7057 - 1963
499. Comet V diesel engine - Progress Note No. 1.
D.P. 7086 - 1963
500. The operation of compression ignition engines on wide boiling range
fuels.
Donel R. Olson, Nelson T. Meckel, & R.D. Quillian, Jr.
SAE Trans. Vol. 70, 1962, p551.
501. Some steps toward calculating diesel engine behaviour.
A.E.W. Austen & V.T. Lyn - CAV Ltd. presented to I.Mech.E. Nov.
1960
502. GMR 4-4 Hyprex free piston turbine engine.
AoF. Underwood - Head Mech. Dev'ment. G.M. Corp. SAE Summer
Mtg. June 5th '56
503. Hybrid engines.
P.H. Schweitzer & L.J. Grunder - Richfield Oil Corp.
SAE Trans. Vol. 71, 1963, p 541-562.
-------
9-37
504. Charge stratification by fuel injection into swirling .air.
A.W. Hussman et al (work done under contract to U..S. Army Ord-
nance) c.1964
505. 1.5 litre Comet V engine (4 cyl) - Progress Report No. 21 - Nov. 1962.
Effect of changes in water and lubricating oil temperatures on engine
performance and volumetric efficiency.
D.P. 6919 - 29th Nov. 1962
506. 1.5 litre diesel engine.
D.P. 6836 - 6th Dec. 1962
507. 1.5 litre Comet V engine (4 cyl) Progress Report No. 20. Covering work
done - Sept.-Dec. 1962.
D.P. 6990 - 16th Jan. 1963
508. The effect of thermal flow in Comet chamber walls.
D.P. 7036 - 12th Feb. 1963
509. Cold starting in the small high speed diesel engine.
D.P. 17771 - 1974
510. Comet V diesel engine - Progress Report No. 4.
D.P. 7937 - 1963
511. Current small 4 cylinder engines for automotive applications..
Ricardo data sheet - 1974
512. Untitled Ricardo report - D.P. 17357 - 1973
513. Comet V diesel engine - Progress Report D.P. 6419 - 1961
514. IDI diesel engine - Progress Report - D.P. 6421 - 1961
515. Comet V diesel engine - Progress Note No. 5 - D.P. 6480 - 1962
516. Comet V diesel engine - Progress Note No. 7 - D.P. 6559 - 1962
517. Comet V diesel engine - Engine No. 2 - Progress Note No. 2 - 1962
D.P. 6509
518. Comet V diesel engine - Progress Note No. 3 - D.P. 6600 - 1962
519. 4 cylinder Comet Vb - weight of component parts - D.P. 6458 - 1962
520. 6 cylinder IDI diesel engine - weight of major components.
-------
9-38
521. 180 cu.in gasoline engine - weight of engine parts.
Unclassified document 1973
522. 75 cu.in gasoline engine - weight of component parts - 1968
523. 120 cu.in Comet V diesel engine - weight of component parts.
D.P. 6183 - 1961
524. 80 cu.in gasoline engine - weight of component parts.
DoP. 9952 - 1967
525. 100 cu.in gasoline engine - weight of component parts.
Unclassified document
526. 150 cu.in gasoline engine - weight of component parts.
Unclassified document
527. 6 cylinder D.I. weight of component parts.
D.P. 6645 - 1962
528. 4 cylinder Comet V diesel engine - Progress Note No. 8.
D.P. 6607 - 1962
529. 4 cylinder Comet V diesel engine - Progress Note No. 6.
D.P. 6525 - 1962
530o Work carried out on No. 2 engine (IDI diesel) prior to despatch.
D.P. 6739 - 1962
531. IDI diesel engine no. 2 - Progress Note No. 3.
D.P. 6773 - 1962
532. 4 cylinder Comet V - Progress Note No. 1.
D.P. 6057 - 1961
533. 1.5 litre Comet V. Progress Report No.l.
D.P. 6064 - 1961
534. 1.5 litre Comet V - review of temperature tests on the various
piston forms.
D.P. 6122 - 1961
535. Light engine oils for improved subzero operation.
V.G. Raviolo - SAE Trans. AP. 1950 No. 2 pl6l
536. Matching fuels to diesel combustion systems.
Dr. C.G.A. Rosen - 1962 Horning Lecture
-------
9-39
537. Dual fuel combustion of propane in a railroad diesel engine.
J.M. Clark & H.M. Bunch - S.W. Research Institute c.1960
538. Can you afford diesel engines for highway hauling?
R.W. Richardson & E.G. Caputo: Simmonds Precisions Products Inc.
539. Swirl and combustion in divided combustion chamber type diesel
engines -
F. Nagao & H. Kakimoto - SAE Trans. Vol 70 - 1962
540. Gas temperatures during compression in motored and fired diesel
engines.
K.C. Tsao, P.S. Myers and O.A. Uyehara - SAE Int. Congress -
Jan. 1961
541. Advantages of propane as a transit vehicle fuel.
R. Lee - SAE Trans. No. 2 Vol. 6 April 1952
542. I.H. high-speed lightweight diesels.
A. Dewsberry et al - SAE Trans. Vol. 68 - 1960
543. Fumigation kills smoke - improves diesel performance.
M. Alperstein et al - SAE Trans. Vol. 66 1958
544. Subzero winterization of diesel engine power equipment.
P.W. Espenschade et al - presented to SAE Nat. Diesel engine mtg,
Chicago Oct. 30th 1951 - Vol. 6 No. 4 Oct. 1952
545. Compression temperatures in diesel engines under starting conditions.
Pennsylvania State University - W.E. Meyer et al
546. Report concerning turbocharged IDI diesel engine - 1961.
547. Report concerning turbocharged IDI diesel engine - 1962.
548. Report concerning turbocharged IDI diesel engine - 1962.
549. Report on temperature measurement tests in the hot plugs for an IDI
diesel engine - 1967.
550. Weights of component parts of various engines.
120 cu.in diesel engine (Comet V) - D.P. 6183 - llth Sept. 1961
551. Effect of valve size and timing on the torque of small Comet engines.
D.P. 17499 & 17544
-------
9-40
552. Diesel car road tests.
D.P. 8805 - 4th Oct. 1965
553. 4 cylinder Comet Mk V - Progress Report No. 2 - June-Sept. 1961.
D.P. 6216
554. Constant power diesel engine - Gas & Oil Power - April 1964
555. Fuel requirements of the small high speed diesel engine.
G. Barrett et al - Gas & Oil Power - Feb. 1956
556. Investigation into nozzle blockage in high speed diesel engines.
G. McConnell - Gas & Oil Power - April 1960
557. The Volvo dual powerplant for military vehicles.
S. Kronogard - Turbine & Automatic Trans. Div, AB Volvo,
SAE 660017
558. The challenge of pollution.
P. Myers - Jnl of Automotive Engineering - April 1970
559. Towards quieter diesels.
M. Russell - CAV Ltd. Jnl of Automotive Engineering - Dec. 1970
560. Exhaust emission legislation.
A. Aitken - Ford Motor Co. Ltd. - Jnl. of Automotive Engineering -
July 1971
561. Car Maintenance - a reappraisal by Marcus Jacobson - Jnl. of
Automotive Engineering - July 1971
562. Recent developments in high speed oil engines.
Prof. S. Davies - March 1938
563. Oxides of nitrogen in diesel engine exhaust gas: their formation and
control.
G. McConnell - Proc. I.Mech.E. 1963-64
564. Research on the compression ignition engine and its fuels.
P. Vaile - Abstract of I.Mech.E. paper, read before North Western
Branch in Manchester - 24th April 1948
565. Combustion in diesel engines.
H.R. Ricardo - lecture delivered 21st Feb. 1950 - London
566. Producer gas for road transport.
J. Hurley & A. Fitton - 30th July 1947
-------
9-41
567. An experimental investigation into the effects of fuel addition to intake
air on the performance of a compression ignition engine.
W. Lyn - paper accepted by Institution Council for publication 12th March
1953
568. The effect of auxiliary fuels on the smoke limited power output of diesel
engines.
L. Derry et al - paper received at Institution 14th Jan. 1953
569. Comparative studies of methane and propane as engine fuels.
N. Moore - paper accepted by Council for publication 23rd Dec. 1955
570. Diesel engine lubricants. Their selection and utilization with parti-
cular reference to oil alkalinity.
A. Dyson et al - paper received by Institution on 27th July 1956
571. Horizontal diesel engine Comet V. .
Progress Report No. 7 - D.P. 6159 - llth August 1961
572. Comet V 4 cyl. 2.8 1 normally aspirated, high speed version.
Progress Note No. 1.
28th August 1961 - D.P. 6170
573. Diesel engine smoke and pollutants. Diesel engineers and users asso-
ciation - J. Spiers & M. Vulliamy - Publication 342 - 18th Feb. 1971
574. CRC Correlation of diesel smokemeter measurements.
F. Hills, T. Wagner, D. Lawrence. SAE 690493- May 23rd 1969
575. Steady-state correlation of diesel smokemeters.
SAE Task Force Report - A. Carey - SAE 690492 - May 1969
576. The measurement of diesel engine smoke.
A. Dodd & Z. Holubecki. - MIRA Report 1965/10.
577. Curves for 4 cyl, IDI diesel engines.
Dec. 1968 - Unclassified document.
578. Combustion tests on an IDI diesel engine - Jan. 1965
Unclassified document.
579. Curves for EDI diesel engine dated 17.3.64.
Effect of heater plug position in Mk VI chamber. Unclassified' document.
580. Test report dated 1.4.63.
Cam form effect on performance of an IDI diesel engine @ 4500 rpm.
Unclassified document.
-------
9-42
581. Test report dated 20.6.69.
Ram pipe tests on an IDI diesel engine d> 3000 rpm.
Unclassified document.
582. Motoring loss tests on an IDI diesel engine.
c. 1965 - Unclassified document.
583. Test report dated Feb. 1964.
Noise tests on an IDI diesel engine at 4500 rpm. Unclassified document.
584. Tests on an IDI diesel engine Mk VI chamber.
Piston combustion chamber tests at various F.I.E. timings - 27.3.64
Unclassified document.
585. High speed C.I. engine combustion systems - D.P. 327E
586. Performance data, N.A. 4 stroke, blower scavenged 2 stroke with
Ricardo combustion systems - D.P. 199A & B.
587. D.I. emissions. «.
GO.373, 6th March 1948
588. Combustion systems in high speed diesels.
D.P. 380, A, B, C.
589. Relative advantages of aluminium versus cast iron crankcases for
high speed diesel engines.
D.P. 651 - 1st April 1949
590. Note on turbocharging of high speed Comet units.
D.P. 771 - 8th Sept. 1949
591. E12 Comet III/IV.
GO.1044
592. Torque reaction tests.
GO.2136
593. Comet III, Camshaft tests.
GO.2470 - 28.3.55
594. Cold starting diesel engines.
D.P. 2711 - 23rd June 1953
595. Comet comparisons of Mk II, III, IV.
D.P. 2779 - 28th Sept. 1953
596. Diesel car.
D.P. 2904 - 4th March 1954
-------
9-43
597. Tests on 4 cylinder diesel engine.
D.P. 2925 - 30th March 1954
598. Inlet manifold starting device.-
D.P. 3057 - 25.11.54
599. Notes on tests to assess relative performance of Comet III;, V, VI and
inverted V chambers..
GO.3100 - 11.11.55
600. 21 Comet V Progress Report No. 3.
D.P. 3144 - Nov. 1956
601. Diesel engine.
D.P.. 3238 -29th Sept. 1954
602. High speed high power road transport engine^.
D.P. 3267 - 29th Oct.. 1954
603. Diesel engine.
D.P. 3324 - 17th Dec. 1954
604. Diesel engine.
D.P.. 3359 - 29th Jan.. 1955
605. Diesel engined taxi.
D.P.. 3372 - 4th, Feb.. 1955
606. Diesel engine.
D.P.. 3394 - 25th Feb. 1955
607. Diesel tractor engine.
D.P. 3395 - 25th Feb. 1955
608. 90 BHP diesel/petrol comparison.
D.P.. 3400 - 3.355
609. Diesel car.
D..P. 3405 - 7th March 1955
610. Noise reduction.
D.P. 3413 - 16th March 1955
611.- Gold starting tests on 2 /4 litre Comet III. chamber engine.
D.P. 3427 - 30.3.55
612. Small Comet V engines.
D.P. 3460 - 12.5.55
-------
9-44
613. High speed loop scavenge diesel engine.
D.P. 3472 - 23.5.55
614. Comparative weights - diesel/petrol.
D.P. 3474 - 1.6.55
615. Volumetric efficiency of diesel engine.
D.P. 3493 - 25.6.55
616. 2J/4 litre Comet V.
D.P. 3550 - 21.9.55
617. Diesel tractor engine.
D.P. 3700 dated 25.1.56
618. Automotive diesel engine.
D.P. 3710 dated 1.2.56
619. Diesel engine.
D.P. 3714 dated 10.2.56
620. Comet V engine.
D.P. 3724
621. Comet V engine 2 1.
D.P. 3730 - 27.2.56
622. Diesels for road transport.
D.P. 4799 - 4.11.58
623. High speed multi-cylinder Comet V engines of up to 90 mm bore.
D.P. 4847 - 12th Dec. 1958
624. Report of work carried out on a 2 /4 litre high ratio Comet V engine
between April-Nov. 1958. K.A. Atkins - D.P. 4862 - 30th Dec- 1958
625. Tests with hot plug throats of various shapes E12/4.
D.P. 5175 - 27th August 1959
626. The small diesel engine for cars.
D.P. 5284 - 11.11.59
627. IDI diesel engine with Bosch heater plugs.
D.P. 5296 - 23.11.59
-------
9-45
'628. Small high speed diesel.
D.P. 5609 - 4.8.60
629. PHS progress note no. 1.
D.P. 5640 - 10.8.60
630. High speed diesel engine proposals.
D.P. 5703A - 24.11.60
631. The effect of diagram shape on combustion noise in diesel engines
D.P. 5742 - 24.10.60
632. Motoring friction.
D.P. 5844 - 4.1.61
633. 1^1 Comet V engine - Progress Note No. 12'
D.P. 5948 - 8.3.61
634. Comet V diesel engine.
D.P. 6009 - 20.4.61
635. Automotive diesel, nozzle tests.
D.P. 6019 - 28.4.61
636. 1^1 Comet V engine - Progress Note No. 13. D.P. 6026 - 2.5.61
637. 4 cylinder Comet Mk V engine progress note no. 5.
D.P. 6480 - 23rd Feb. 1960
638. The 2 stage rotary engine - a new concept in rotary power.
F. Feller - I.Mech.E. Proc. 1970-71 - Vol. 185 - 13/71
639. Maintenance frequency in certain bus operating undertakings as it
affects injection equipment and combustion zone components.
D.P. 3870 - 1956
640. Classification of diesel fuels.
R.P. Linderan, D.K. Lawrence and T.O. Wagner - SAE 680467 - May
1968
641. Avoidance of diesel fuel filter plugging in winter months.
Esso - Petroleum Technical Service No. 73/29 - 16.11.73
642. The cold filter plugging point: A practical operability task for middle
distillate fuels. Paramins (Esso) RBM 5SU 73 April 5 - 1973
-------
9-4b
643. Minimising low temperature sensitivity of diesel vehicle fuel systems.
(Esso) Paramins RBM 54SU-X 73. - 19.11.73
644. How to run diesel engines in cold weather.
W. Sloan et al - SAE Jnl. - June 1951
645. Fleets, but not private cars, are likely customers for LPG.
SAE Jnl. Dec. 1951
646. 4000 rpm is practical aim for automotive diesels.
N.M. Reiners & R.C. Schmidt - SAE Jnl. Dec. 1951.
647. Daimler Benz Test results - EPA report - 1973
648. The small high speed diesel engine - a power unit for future passenger
cars.
D.P. 17478 - 1974
649. Berliet Pat. Brit. Pat. 1,128,661
650. Berliet Pat. Brit. Pat. 1,213,206
651. Institut Gornogo Dela Pat. 1,270,782
652. Diesel Exhaust Odor - its evaluation and relation to exhaust gas com-
position.
F.G. Rounds & H.W. Pearsall.
SAE Trans. Volume 65, 1957, Pages 608/627.
653. Smoke and Odour Control for Diesel Powered Trucks and Buses.
R.C. Stahman, G. Kettridge, K.J. Springer.
SAE 680443.
654. Investigation of Diesel Powered Vehicle Odor and Smoke.
Part V, SWRI Project No. 11-2340-005. K.J. Springer.
655. Note on a visit to Messrs. Opel, Russelsheim. by C.A. Beard on
November 27th, 1973 - D.P. 17387.
656. Peugeot letter to Ricardo dated 20th December 1973 + attachment.
657. London General Cab. Co. Ltd. letter to Ricardo dated 15th January 1974
+ attachment
658. Note on discussion with Herr U= Alen of Verband Fuer das
Personenwerkehrsgewerke, Hamburg, on May 31st, 1974 - D.P. 18175
-------
9-47
659. Note on visit to Slota Taxis, Paris on 9th May, 1974 - D.P. 18100.
660. CAV letter to Ricardo dated 30th November, 1973.
661. Note on visit to Deutz, Porz, on 28th November 1973 - D.P. 17394
662. Note on visit to Austin Hire & Taxi Service Ltd., Worthing on
19th November 1973 - D.P. 17320.
663. Note on a visit by C.A. Beard to BLMC on 10th January 1974 -
D.P. 17510.
664. Note on discussion with B.P. in London on 21st January 1974 -
D.P. 17551.
665. A new light aircraft engine - the radial 2-stroke diesel.
J.L. Dooley - Ex. EPA data - 1970.
666. Extract from letter sent by W.A. Bareham to Mr. S. Oshika
667. Mass emissions and fuel economy from LDV diesel engines.
1975 FTP - Ex EPA data.
668. Some engine ideas of the Past and Future.
C.F. Bachle
669. Manufacturability and costs of proposed low emissions automotive
engine systems.
National Academy of Sciences report.
3
670. Ricardo Comet Mk V Emission levels from a 416 in truck engine under
turbocharged conditions.
D.P. 15176 - 11 May 1972.
671. McCulloch is developing lightweight aircraft diesel - Automotive
Engineering - Sept. '73, Vol. 79. No. 9, p53.
672. Discussion of diesel operation at. ]ow temperature at Shoreham on
18th January 1974.
D.P. 17540 - Jan. '74
673. Some observations on the nature of blue smoke in engine exhaust.
W.T. Lyn M.I.M.E. - CAV Ltd. - Unclassified document.
674. Exhaust Particulates - D.P. 16471 - April 1973.
-------
9-48
675. Fuel economy and emissions control - EPA report - Nov. '72.
676. A report on automobile fuel economy.
EPA Report - Oct. 1973.
677. The Low Emission Car for 1975 - Enter the Diesel.
K.J. Springer & H.A. Ashby - SAE 739133.
678. Diesel for passenger cars.
F. Bachle - EPA report - Feb. 1974.
679. Nitrogen oxides and variables in pre-combustion chamber type diesel
engines - E.W. Landen - SAE 714B - 1963.
680. Feasibility of meeting the 1975-76 exhaust emission standards in actual
use.
National Academy of Sciences - June 1973
681. Characterization and control of emissions from heavy duty diesel and
gasoline fuelled engines - prepared for EPA by Bartlesville Energy
Research Center, Bureau of Mines - Dec. 1972
682. An evaluation of alternative power sources for low emission automobiles
National Academy of Sciences - April 1973
683. Manufacturability and costs of proposed low emission automotive engine
systems - National Academy of Sciences - Jan. 1973
684. Diesel exhaust odor analysis by sensory techniques.
D.A. Kendall, P.L. Levins & G. Leonardos - SAE 740215 - 1974
685. Chemical analysis of diesel exhaust odor species.
P.L. Levins et al - SAE 740216 - 1974
686. Public opinion of diesel odor.
C.J. Hare et al - SAE 740214 - 1974
687. Progress in diesel odor research.
C.W. Savery, R.A. Matula & T. Asmus - SAE 740213 - 1974
688= Letter from K.J. Springer (Southwest Research Inst) to R.C. Stahman
(EPA) - Dec. 13, 1973.
689. Notes on Turner 2 stroke IDI engine - taken from Commercial Motor &
Engineering - 1953
-------
9-49
690. Higher BMEP's in automotive diesels.
D. Broome - reprint from Automotive Design Engineering - March 1966
691. A Designer's Viewpoint.
H. Barnes-Moss - I.Mech.E. paper - Conference Pub. 19 - No. C343/73
692. Survey of diesel engine emissions - 3rd report.
Project Group CEC-CF 12, June 20th, 1973. Unclassified document.
693. Unaided starting of diesel engines.
T.W. Biddulph & W.T. Lyn - Proc. I.Mech.E. 1966-67, Vol. 181, Part 2A
694. Influence of atmospheric pressure and temperature upon the performance
of the naturally aspirated four stroke C.I. engine.
C.B. Dicksee - I.Mech.E. - 1959
695. Diesel engine combustion processes with a view to their use in pass-
enger cars.
Prof. Dr. Ing. E. Eisele - The Internal Combustion Engines Conference,
Bucharest - 1970, Paper No. 1C.
696. Diesel engines operating in underground mines.
Dr. Ing. G. Reyl - Unclassified document
697. Unclassified document received from EPA - Chapter XVI - Diesel Engine.
698. Letter from K.J. Springer (Southwest Research Inst) to R.C. Stahman
(EPA) - Feb. 15, 1974.
699. Durability of advanced emission controls for heavy duty diesel and
gasoline fueled engines - prepared for EPA by Dr. R.D. Fleming and
T.R. French, Bartlesville Energy Research Center, Bureau of Mines,
Sept. 1973.
-------
10-1
SECTION 10
APPENDIX 3 - GLOSSARY OF TERMS
Some of the terms mentioned in the text of this report may not be fami-
liar to the reader and the following glossary has been compiled to clarify
matters.
CID - engine swept volume in cubic inches.
CVS-CH - 1975 cold/hot start Federal test procedure using CVS (constant
volume sampling) equipment. For diesel powered vehicles HC's
are measured by means of a heated flame ionisation detector,
the gas sample being extracted via a heated line to minimise
condensation of the heavier hydrocarbons (mass emissions at
the end of the test are determined by integrating the continu-
ous trace obtained from the flame ionisation detector). This
is the only difference in the 1975 test procedure between gaso-
line and diesel powered light duty vehicles.
Comprex - belt driven engine boosting device working on a pressure ex-
changer principle. Currently being developed by Brown-Boveri,
Switzerland.
DI - direct injection (open chamber).
IDI - indirect injection (pre-cup or swirl chamber).
EGR - exhaust gas recirculation^
F.IoE. - fuel injection equipment.
getter box - catalytic device fitted upstream of a reducing catalyst to pro-
tect it from oxygen "spikes".
mpg - fuel economy in miles per U.S. gallon.
NA - normally aspirated»
T/C - turbocharged
DP )
- internal identification for classification of Ricardo reports.
SN )
-------
10-2
TECHNICAL REPORT DATA
(Please read Instructions on the reverse before completing)
1. REPORT NO. , 2._
EPA -460/3-74-011
4. TITLE AND SU&TITLE
A STUDY OF THE DIESEL AS A LIGHT DUTY
POWERPLANT
7. AUTHOH(S)
C.C.J. FRENCH, M.L. MONAGHAN, R.G. FREESEo
9. PERFORMING ORGANIZATION NAME AND ADDRESS
RICARDO & CO. ENGINEERS (192?) LTD.,
BRIDGE WORKS, SHOREHAM-BY-SEA, SUSSEX.
BN4.5FG. ENGLAND.
12. SPONSORING AGENCY NAME AND ADDRESS
Environmental Protection Agency, Office of Air & Water
Programs, Office of Mobile Source Air Pollution Control,
Emission Control Technology Division, Ann Arbor,
MiVhirpn 48105-
3. RECIPIENT'S ACCESSIOWNO.
5. REPORT DATE
JULY 1974
6. PERFORMING ORGANIZATION CODE
8. PERFORMING ORGANIZATION REPORT NC
D.P. 18410
10. PROGRAM ELEMENT NO.
11. CONTRACT/GRANT NO.
68-03-0375
13. TYPE OF REPORT AND PERIOD COVEREC
FINAL RKPOKT
14. SPONSORING AGENCY CODE
15. SUPPLEMENTARY NOTES
is. ABSTRACT Tnjs project was carried out to determine whether the diesel engine was a
possible power plant for light duty vehicles in America. The light duty vehicle consid-
ered was a 4/5 seat compact sedan with good acceleration capabilities and was consid-
ered for a primary emissions environment of HC - 0.41 g/mile, CO - 3.4 g/mile, NOX-
1.5 g/mile. A secondary environment of HC-0.41 g/mile, CO-3.4 g/mile, NOX-0.4
g/mile was also considered.
A literature survey was carried out covering existing light duty diesel work and exper
ence throughout the world. This indicated that the diesel engine was a viable power
>lant for light duty use.
An engine configuration study was then carried out when 9 diesel engines developing
96kw were designed. The diesel types covered were indirect and direct injection two-
cycle and four-cycle engines; rotary and compound versions were also included. Two
gasoline engines were also outlined for comparison purposes.
A method of rating the various power plants was devised and this was applied to the 9
diesels and the 2 gasoline engines. It was concluded that although the gasoline engine
was slightly superior to the diesel engine for passenger car use many applications with
an emphasis on fuel consumption and durability would give an equal or better rating to
the diesel.
Only the indirect injection four-cycle diesels were capable of meeting the primary
emissions target with current technology but they did not require any cataly.st.s to do thi:
17. KEY WORDS AND DOCUMENT ANALYSIS
a. DESCRIPTORS
EXHAUST EMISSIONS
DIESEL ENGINES
POWERPLANT RATING METHODOLOGY
ENGINE DESIGN
LITERATURE REVIEW
19. DISTRIBUTION STATEMENT
RELEASE UNLIMITED
b.lDENTIFIERS/OPEN ENDED TERMS
LIGHT DUTY VEHICLES
LIGHT DUTY ENGINES
DIESEL/GASOLINE
COMPARISON
EMISSION CONTROLS
FUEL ECONOMY
19. SECURITY CLASS ( ilns Re/,,'ril
UNCLASSIFIED
20. SECURITY CLASS (This page!
UNCLASSIFIED
c. COSATI Held/Group
-
21. NO. OF PAGE'S
22. PRICE
EPA Form 2220-1 <9-73)
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