Internal Combustion Engines

2.61 — Spring 2004 Internal Combustion Engines

John HeywoodWai Cheng

2.61 — Spring 2004 Internal Combustion Engines

John HeywoodWai Cheng

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Contents

Highlights of this Course ...................................................................... 1

Course Description .............................................................................. 1

Syllabus ............................................................................................ 2

Calendar............................................................................................ 3

Readings ........................................................................................... 5

Lecture Notes ................................................................................... 13

Overview and Basics of Engine Operation Engine Geometry and Performance Parameters Ideal Cycle Analysis and Combustion Stoichiometry Gas Exchange: 4-Stroke and 2-Stroke

Spark-ignition Engine Combustion and Flame Propagation and Structure KnockSI Engine Emissions Emission Control Technology and Review Mixture Preparation and Engine Friction Engine Heat Transfer and 2-Stroke SIE Performance 4-Stroke SIE Performance Variable Valve Control and Gasoline DI Engines Diesel Overview and Diesel Combustion Future Engine Technology and Discussion

Labs...............................................................................................110

Lab 1 Lab 2

Assignments ...................................................................................117

Problem Set 1 Problem Set 2 Problem Set 3 Problem Set 4 Problem Set 5 Problem Set 6

Exams............................................................................................141

2002 Quiz 1 2002 Final Exam

Project ...........................................................................................153

Project Description Exhaust Temperature Data, with EGR Emissions Data One of Many Possible Solutions

Related Resources............................................................................174

Highlights of this Course

This course features homework assignments, labs and an extensive reading list.

Course Description

This course elaborates on the fundamentals of how the design and operation of internal combustion engines affect their performance, operation, fuel requirements, and environmental impact, study of fluid flow, thermodynamics, combustion, heat transfer and friction phenomena, and fuel properties, relevant to engine power, efficiency, and emissions, examination of design features and operating characteristics of different types of internal combustion engines: spark-ignition, diesel, stratified-charge, and mixed-cycle engines. The project section details the Engine Laboratory project. We have aimed this course for graduate and senior undergraduate students.

Course Meeting Times Lectures: Two sessions / week 2 hours / session

LevelGraduate

1

Syllabus

Text

Heywood, John B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, 1988. ISBN: 007028637X.

Grading

Grade will be based on

ACTIVITIES PERCENTAGES

Quiz + Exam 45%

Homework 25%

Lab Report 20%

DesignProblem

10%

Total 100%

Homework

Homework Policy

1. The purpose of the homework and problems is to get each of you to think about and use the material we discuss in class.

2. Obviously, I want each of you to make a serious try at each problem. If discussions with other students help you get started, then contribute to and benefit from such discussions.

3. However, I expect that each of you will work independently on the details of the problem solutions which you hand in as yourwork. I regard it as dishonest to copy from any previously circulated solutions and present such work as your own.

4. If you have any questions about the above, please discuss them with me.

2

Calendar

SES # TOPICS KEY DATES

1Overview

Basics of Engine Operation Problem set 1 out

2Engine Geometry

Performance Parameters

3Engine Disassembly

In Lab

Problem set 1 due

Problem set 2 out

4Ideal Cycle Analysis

Combustion Stoichiometry

5Fuel-air Cycle Model

Fuel-air Cycle Results

Problem set 2 due

Problem set 3 out

6Gas Exchange: 4-Stroke

Gas Exchange: 2-Stroke

7Spark-ignition Engine Combustion

Flame Propagation and Structure

8 RecitationProblem set 3 due

Problem set 4 out

9 Knock

10 SI Engine Emissions

11Emission Control Technology

Review

3

12Quiz (Duration: 1-1/2 hours) Open Book

Problem set 5 out

13Mixture Preparation

Engine Friction

14Engine Heat Transfer

2-Stroke SIE Performance

Problem set 5 due

Problem set 6 out

15 4-Stroke SIE Performance

16Variable Valve Control

Gasoline DI Engines

17Diesel Overview

Diesel Combustion

Problem set 6 due

Design problem out

18Diesel Emissions

Emission Control

19Turbocharging

Diesel Performance

20 1-hour Lab Preparation Design problem due

21 2-hour Lab Session

22Engine Noise

Engine Dynamics

Lab report due 1 week after session 22

23Future Engine Technology

Discussion

24Final Exam (Duration: 3 hours) Open Book

4

Readings

The readings referred to the table below are recommended readings from the text:

Heywood, John B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, 1988. ISBN: 007028637X.

SES # TOPICS READINGS

1 Overview

Basics of Engine Operation Essential

Chapter 1

2 Engine Geometry Essential

Section 2.2

Performance Parameters Essential

Chapter 2

3 Engine Disassembly

In Lab

4 Ideal Cycle Analysis Essential

Sections 5.1-5.4

Combustion Stoichiometry

Essential

Sections 3.1-3.4

Sections 3.7, 4.1, 4.3 (Working fluid properties)

Helpful

Section 3.5

5

Sections 4.2, 4.4, 4.9.1 (Working fluid properties)

5 Fuel-air Cycle Model

Essential

Sections 4.5, 4.6, 5.5

Helpful

Section 4.7

Fuel-air Cycle Results

6 Gas Exchange: 4-Stroke

Essential

Sections 6.1, 6.2, 6.5

Helpful

Section 6.3

Gas Exchange: 2-Stroke

Essential

Section 6.6

Sections 8.1, 8.2.1 (In-cylinder flows)

Helpful

Section 6.7

Sections 8.3-8.5 (In-cylinder flows)

7Spark-ignition Engine Combustion

Essential

Sections 9.1, 9.2.2, 9.2.3, 9.3.1, 9.3.2, 9.4.1, 9.4.3

Helpful

6

Sections 9.2.1, 9.3.3, 9.3.4, 9.4.2, 9.5.1

Flame Propagation and Structure

8 Recitation

9 Knock

Essential

Section 9.6.1

Helpful

Sections 9.6.2, 9.6.3

10 SI Engine Emissions

Essential

Sections 11.1, 11.2.1, 11.2.3, 11.3, 11.4.1, 11.4.3, 11.6.1, 11.6.2, 7.4

SI Engine Emissions (cont.)

Essential

Sections 11.1, 11.2.1, 11.2.3, 11.3, 11.4.1, 11.4.3, 11.6.1, 11.6.2, 7.4

11Emission Control Technology

Review

12Quiz (Duration: 1-1/2 hours) Open Book

13 Mixture Preparation

Essential

Sections 7.1, 7.3.1, 7.6

Table D4, p. 915 (Fuels)

Helpful

7

Section 7.5

Engine Friction

Essential

Sections 13.1 – 13.3, 13.5

Helpful

Section 13.6

14 Engine Heat Transfer

Essential

Sections 12.1, 12.2, 12.3

Helpful

Sections 12.4, 12.7

2-Stroke SIE Performance

Essential

Sections 15.1, 15.2

Helpful

Sections 15.3, 15.4

15 4-Stroke SIE Performance

Essential

Sections 15.1, 15.2

Helpful

Sections 15.3, 15.4

4-Stroke SIE Performance (cont.)

Essential

Sections 15.1, 15.2

Helpful

Sections 15.3, 15.4

8

16 Variable Valve Control

Gasoline DI Engines

17 Diesel Overview

Diesel Combustion

Essential

Sections 10.1, 10.2, 10.3, 10.7

Helpful

Sections 10.5, 10.6.1, 10.6.3

Sections 10.6.2, 10.6.5 (Diesel fuels)

18 Diesel Emissions

Essential

Sections 11.2.4, 11.4.4, 11.5.2

Helpful

Sections 11.5, 11.6.4

Section 12.5 (Diesel heat transfer)

Emission Control

19 Turbocharging

Diesel Performance

Essential

Sections 15.5, 15.7

Helpful

Section 15.6

20 1-hour Lab Preparation

9

21 2-hour Lab Session

22 Engine Noise

Essential

Section 7.6, "Engine Noise." In Heywood, J. B. and E. Sher, The Two-Stroke Cycle Engine. Philadelphia PA: Taylor and Francis, 1999.

Engine Dynamics

23 Future Engine Technology

Discussion

24Final Exam (Duration: 3 hours) Open Book

Corrections to the text Internal Combustion Engine Fundamentals

Chon, Dale M., and John B. Heywood. "Performance Scaling of Spark-Ignition Engines: Correlation and Historical Analysis of Production Engine Data."SAE (Society of Automotive Engineers) Technical Paper 2000-01-0565 (2000).

Berckmüller, M., et al. "Potentials of a Charged SI-Hydrogen Engine." SAE (Society of Automotive Engineers) Technical Paper 2003-01-3210 (2003).

10

11

Corrections to text (current printing): (JBH 10/29/01)

p. 77 Table 3.2. The enthalpies of formation for C8H18 are for n-octane. For isooctane they are –224.1

and –259.3 MJ/kmol for gas and liquid C8H18, respectively.

p. 89: Middle of page: xCO 2, xCO and x02

should be x̃CO 2, x̃CO, and x̃O2

.

p. 122: Figure 4-10 is a repeat of Fig. 4-3 due to an editing error, though Fig. 4-10 is correctly labeled “burned mixture properties.” A correct Fig. 4-10 is attached. It is only slightly different: e.g., at 1000 K the burned mixture u for 1.2 is 4% lower than the unburned mixture value, and h iss s

1% lower than the unburned mixture value. These differences scale, approximately, with .

p. 151 Underneath Eq (4.65) insert:

K is given by Eq. (4.63)

p. 152: Line 5. CmHnOr should be CnHmOr .

p.188 In Eq. (5.66c), m is omitted. It should read:

S3b S2 mcv lnTT3

2

a mcp lnT3b mcv ln mcp ln T3a

p. 306: Equation (7.18): The sign at the beginning of the second line of the equation (a minus sign) should be a plus sign.

p. 388: Equation (9.27). The sign in front of the third term in the square bracket should be , not + :

T T 1 1i.e., Tw ( 1)

lnTw bTw 1

p. 553: Equation (10.37). There should be a + sign between the two round brackets within the square bracket., i.e.,

1 1 21.20.63

(CA) (0.36 0.22Sp )exp EA R T 17,190 p 12.4id ˜

p. 620: The reference for Fig. 11-33 should be Yu, R.C., Wong, V.W., and Shahed, S.M., “Sources of Hydrocarbon Emissions from Direct Injection Diesel Engines,” SAE paper 800048, SAE Trans., vol. 89, 1980. (This is a new reference; make it reference 87 and add it to p. 667.)

p. 679: In the inserted graph in Figure 12-5, the scale for thermal conductivity kg is not correct. The values

should be multiplied by 5 x 105: e.g., the peak value of 10 x 10-8 = 10-7 W/m.K should be 10-7 x (5 x 105) = 5 x 10-2 W/m.K.

p. 880 In Fig. 15-45, the units for pressure (middle left) should be kPa and not MPa.

12

Lecture Notes and Topics

Lecture Session Topic

1 Overview and Basics of Engine Operation 2 Engine Geometry and Performance

Parameters3 4 Ideal Cycle Analysis and Combustion

Stoichiometry56 Gas Exchange (4-Stroke and 2-Stroke) 7 Spark-ignition Engine Combustion and

Flame Propagation and Structure 8 No Lecture Available 9 Knock10 SI Engine Emissions 11 Emission Control Technology and Review 12 No Lecture Available 13 Mixture Preparation and Engine Friction 14 Engine Heat Transfer and 2-Stroke SIE

Performance15 4-Stroke SIE Performance 16 Variable Valve Control and Gasoline DI

Engines17 Diesel Overview and Diesel Combustion 18 No Lecture Available 19 No Lecture Available 20 No Lecture Available 21 No Lecture Available 22 No Lecture Available23 Future Engine Technology and Discussion

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________________________________________________________________________

________________________________________________________________________

2.61 Internal Combustion Engines

Lab session: Engine disassembly

Tuesday, February 10, 2004

2:00 – 4:00 P.M.

Please report to Ferran Ayala and Thane Dewitt in the Laboratory area of the Sloan

Automotive Lab (first floor of Building 31).

The purpose of this session is to get some hands-on experience on the mechanical aspect

of the engine. You should learn the mechanical construction of the engine, how the

different components are arranged, and get a feel for the size and weight of the

components.

There are three engines. Two of them are Ford 4-cyliner, 16-valve engines for the

compact size vehicles. The more modern one has a plastic intake manifold to reduce cost

and NVH (Noise, Vibration and Harshness). The third engine is a Peugeot engine. The

class will be divided into three groups; each group is responsible to disassemble one

engine. However the whole class should participate in the initial ‘looking’ at each of the

whole engine and the final study of all the parts for the three engines.

Before disassembly, take a look at the engines as a whole, note the arrangement of the

different components, the gas exchange circuit: intake, exhaust, EGR (Exhaust Gas

Recirculation), and PCV (Positive Crankcase Ventilation) system, the coolant circuit and

the fuel flow circuit. Understand the function of these components.

When the engine is open, look at the valve train arrangement, the lubrication circuit, the

EGR route, the coolant passage, and the piston/ crank/ balance weight arrangement.

Record the measurements indicated on the next page. The measurements may be shared

by the group, but the comments and calculations should be done by each individual.

110




Engine

Measurements

Mean intake port length

Mean exhaust port length

Minimum intake port cross sectional area

Minimum exhaust port cross sectional area

Mean intake manifold runner length

Mean exhaust manifold runner length

Mean intake manifold cross sectional area

Mean exhaust manifold cross sectional area

Intake valve and minimum seat diameter

Exhaust valve and minimum seat diameter

Intake and exhaust valve stem diameter

Maximum intake and exhaust valve lift

Cylinder bore (at TDC and BDC)

Stroke

Connecting rod length (pin center to pin center)

Piston top land diameter

Piston top land height

Piston middle land diameter

Piston middle land height

Overall piston height (max and min)

Piston diameter (skirt center to skirt center)

Piston top ring dimensions (thickness x height x outer diameter)

Piston top ring groove diamensions (inner diameter x height)

Piston middle ring dimensions (thickness x height x outer diameter)

Piston middle ring groove diamensions (inner diameter x height)

Calculations

Cylinder displaced volume

Bore/Stroke

Connecting rod length/Crank radius

Intake valve minimum seat diameter/Bore

Exhaust valve minimum seat diameter/Bore

Maximum intake valve lift/Intake valve diameter

Maximum exhaust valve lift/Exhaust valve diameter

Volume of one intake port plus one intake manifold runner/

displaced volume of one cylinder

Volume of one exhaust port plus one exhaust manifold runner/

displaced volume of one cylinder

Clearance Volume (Vc)

Crevice Volume V1

Crevice Volume V2

Crevice Volume V3

Crevice Volume V4

Connecting rod lower end and upper end mass

111




Crevice Volumes

V1

V2

V3

V4

Top of Piston

3Determine V1, V2, V3, and V4 in mm , and as a percentage of the clearance volume

112


Lab session: Engine disassemblyTuesday, February 10, 2004

Balancing of Connecting RodThe connecting rod can be considered as equivalent to two masses concentrated at its

ends (see diagram below), such that the sum of the masses is equal to the mass of the rod

(Taylor).

When doing a force balance of the piston motion, the mass of all the parts which are

considered to reciprocate with the piston must be taken into account. These include the

piston, the piston rings, the piston pin, and the equivalent mass of the upper end of the

connecting rod.

Roughly find the center of gravity of the connecting rod by balancing it, and then find the

corresponding mass of its lower and upper ends by summing moments. What proportion

of the total mass of the rod should be included in the piston mass calculation?

(A) W1

W1

W2

hj

c g

c g

W2

(B)

113


Laboratory Session

Tuesday, April 27, 2004

Background

This laboratory is designed to provide direct experience with the experimental methods used in internal combustion engine research and development. During the laboratory a set of engine performance and emissions data will be taken from an operating spark-ignition engine on a test stand. The data will be processed, analyzed and interpreted in the context of the course lecture material.

Students will work in groups. Each group should process their experimental data as a group. Each individual will then write up a laboratory report as described below. Please read these notes carefully before you come to the laboratory.

Objective

The experiments are designed to examine the effect of changes in spark timing and relative AFR on spark-ignition engine performance and emissions characteristics. The results are going to be compared with those computed from the GMR simulation code.

Experimental Method

To eliminate effects of extraneous variables we will hold constant:

Design variables (compression ratio, valve timing, etc).

Inlet pressure (load) and speed.

Inlet temperature and exhaust pressure.

Oil and coolant temperature.

The test procedure is the following:

1. Maintain engine speed at 1500 rpm and inlet pressure at 0.7 bar. Run the engine until it reaches steady state conditions (coolant temperature is approximately 600 C).

2. Set the relative AFR to stoichiometric. Retard and advance the spark timing about the value that gives the MBT according to the lab sheet and record all data listed on it. To determine GIMEP, COV and location of maximum pressure, record in-cylinder pressure data for approximately 150 cycles for each case.

Page 1 of 3 114

3. Repeat these measurements for different relative AFR’ s, both on the lean and rich side according to the lab sheet. Use the specified MBT spark timing for each value of the relative AFR.

Analysis

The following tasks are to be addressed in the lab report:

1. Effect of changes in spark timing:

Plot GIMEP, COV and location of maximum in-cylinder pressure versus spark advance and explain the shape of the curves. What is the location of maximum in-cylinder pressure when MBT spark timing is applied?

Plot NOx, HC emissions and exhaust gas temperature versus spark advance. Explain the effect of spark timing on each of them.

2. Effect of changes in relative AFR:

Plot the MBT spark timing variation with the relative AFR. What does the MBT spark-timing data tell you about the variation in burn rate with the relative AFR? Is there any significant variation in the location of maximum in-cylinder pressure when the relative AFR changes but the spark timing is always set to be the optimum?

Plot GIMEP and COV versus the relative AFR. Explain the effect of AFR on combustion variability. At what AFR is combustion most stable, and why?

Determine the indicated values for specific fuel consumption and fuel conversion efficiency and plot them versus the relative AFR.

Plot HC, CO, CO2 and NOx emissions versus the relative AFR. Calculate H2O and estimate H2 concentrations in the exhaust stream and include them also in graphs. Specify the background moisture for the measurement of each species. Plot also the exhaust gas temperature versus the relative AFR. Explain the trends of all curves you plot.

Determine the relative AFR from the exhaust gas composition measurements and plot it versus the measured value as recorded by the UEGO sensor.

Calculate the combustion efficiency for each point and explain the trend.

3. Comparison of experimental data with the fuel-air cycle results and the GMR simulation code (change geometry when possible according to given engine specs):

Plot on the same graph GIMEP based on the experimental data, the fuel-air cycle results and the GMR simulation code versus the relative AFR.

Draw the corresponding graph for the indicated fuel conversion efficiency.

Page 2 of 3 115

Compare the experimental data with the corresponding calculated values based on the GMR simulation code of CO, CO2, NOx and H2O. Be careful to have the same background moisture for each species for compatible comparisons.

Laboratory Report

The laboratory report is due on Thursday, May 6, 2004 by 5 pm in Ferran Ayala’s office. We recommend that you process your data with other members of your laboratory group. Reports should be written separately by each individual, however. The report should include the following:

Brief introduction, which reviews the goals of the experiment and procedure.

Analysis.

Results.

Discussion and interpretation.

Experimental data (attached as one or more appendices).

The questions the report should address are listed in the previous section.

Word processor or neatly handwritten reports are requested. Excessive length and repetition of material in the laboratory handout are undesirable. Best marks are obtained with a logical report organization, a concise summary of what was done, and good physical explanations of the trends observed in the data. There may well be inconsistencies in the data; these should be identified and explained where possible.

Page 3 of 3 116


Problem Set 1


Due: Tuesday, February 10, 2004

1. Explain briefly the following differences between a standard automobile spark-ignition (SI) engine

and a truck diesel engine:

(a) where the fuel is injected and why

(b) how the load is varied at fixed speed

(c) how the combustion process starts, develops and ends

(d) how the in-cylinder pressure varies as a function of crank angle (draw qualitatively the

pressure traces for both engines in the same graph showing their relative magnitudes and

when approximately the combustion starts and ends in each case).

(e) how the fuels are different and why

2. At 450 after top dead center (ATDC) on the expansion stroke, the pressure in a SI engine cylinder

is 1000 kPa. The bore and the stroke are 80 mm, the ratio of the connecting rod length to the crank

radius is 3.5, the piston mass (including the pin and half the connecting rod) is 0.57 kg, the

crankcase pressure is 100 kPa, the axial friction force on the piston (due to rings sliding against

the cylinder liner) is 65 N and the engine speed is 2500 rpm.

(a) Draw the piston and connecting rod free body diagrams, calculate all forces acting on them

and compute the torque delivered to the crankshaft.

(b) Answer part (a) for the same crank position during the intake stroke if the engine is

working at half load conditions assuming that the friction force is the same. Hint: a good approximation to piston acceleration is the following:

12 2a = ?(cosLNp 2 +

R)?2cos

N … engine speed

L … stroke

R … connecting rod length to crank radius ratio

theta … crank angle measured from TDC

(Bosch Automotive Handbook, 4th

Edition, pp. 404 – 405)

A more accurate expression can be found be differentiating twice the distance between the crank

axis and the piston pin axis.

3. An automobile engine should provide enough power to a car in order to overcome its resistances.

These resistances consist of the rolling resistance arising from the friction of the tires with the

road, the aerodynamic drag, the inertia as the car accelerates and the gravity if it is travelling up a

hill.

Ford Taurus (4-door) weights 1515 kg, and has a relatively flat torque – engine speed curve of 247

Nm for medium speeds. Estimate:

(a) the road load power (power required to drive a vehicle on a level road at steady speed)

when the vehicle speed is 50 and 100 km/h. Comment on the relative importance of the

resistances as vehicle speed increases

(b) the power required to drive the car up to 20% gradient hill at steady speed of 50 km/h

Page 1 of 2117

(c) the minimum time required to accelerate the car from 40 to 80 km/h on a level road with

the 3rd

gear. You can assume that the ratio of the engine to the wheel speed is

approximately 1x3.27 (which is the transmission torque ratio with the 3rd

gear times the

axle ratio).

Make appropriate estimations for road roughness and car geometry (drag coefficient, frontal area

and wheel radius).

4. A 4-cylinder, 4 stroke diesel engine is being designed. A bore of 100 mm and a stroke of 120 mm

have been selected and the operating speed is to be 1500 rev/min. A turbocharging system is

envisaged which will supply inlet manifold air at 2.0 bar, 380 K. The volumetric efficiency is

expected to be 90%. The indicated fuel conversion efficiency has been estimated at 55%, and an

air-fuel ratio of 28:1 is to be used. The friction mean effective pressure is expected to be 2.2 bar.

The calorific or heating value of the fuel is 42.5 MJ/kg.

Estimate:

(a) Mass flow rate of air into the engine (kg/s)

(b) Mass of fuel burned per cylinder per cycle (mg)

(c) The indicated work done per cylinder per cycle (kJ)

(d) The brake mean effective pressure (kPa)

(e) The engine shaft power output (kW)

(f) The brake specific fuel consumption (g/kWhr)

(g) The brake fuel conversion efficiency

Page 2 of 2118


Problem Set 2


Due: Thursday, February 19, 2004

1. Several velocities, time, and length scales are useful in understanding what goes on inside engines.

Make estimates o the following quantities for a 1.6-liter displacement four-cylinder spark-ignition

engine, operating at wide-open throttle at 2500 rev/min.

(a) The mean piston speed and the maximum piston speed.

(b) The maximum charge velocity in the intake port (the port area is about 20 percent of the

piston area).

(c) The time occupied by one engine operating cycle, the intake process, the compression

process, the expansion process, and the exhaust process. (Note: The word process is used

here not the word stroke.)

(d) The average velocity with which the flame travels across the combustion chamber.

(e) The length of the intake system (the intake port, the manifold runner, etc.) which is filled

by one cylinder charge just before the intake valve opens and this charge enters the

cylinder (i.e., how far back from the intake valve, in centimeters, one cylinder volume

extends in the intake system).

(f) The length of exhaust filled by one cylinder charge after it exits the cylinder (assume an

average exhaust gas temperature of 425 C)

You will have to make several appropriate geometric assumptions. The calculations are

straightforward, and only approximate answers are required.

2. Evaluate the maximum bmep, the bmep at maximum power and the maximum value of the mean

piston speed (where applicable) for the engines described in the attached sheets:

Gasoline (natural aspirated and turbocharged)

(a) General Motors Vortec 4.2L DOHC I-6

(b) Audi AG 4.2L DOHC V-8

(c) BMW AG 3.2L DOHC I-6

(d) K13C with common rail fuel injection

(e) Subaru 2.5L H-4 turbocharged

Diesel (natural aspirated and turbochared)

(f) 5.9L Cummins 600 OHV l-6 turbodiesel

(g) Isuzu new V12 PE1-S

(h) Mitsubishi 2.8L

Specialty (hybrid, rotary, formula 1)

(i) Mazda 1.3L Renesis rotary

(j) Toyota 1.5L DOHC I-4 Hybrid

(k) Formula 1 engine: naturally aspirated, 3.0 liter, 10-cylinder, stroke to bore ratio is 0.48,

maximum brake power at 17000 rpm is 800 hp and the maximum brake torque at 13000

rpm is 350 Nm

Page 1 of 12 119

Explain briefly the differences in performance parameters between the various engine

configurations:

- SI: 2 valves per cylinder, 4 valves per cylinder, F1 racing engines

- CI: naturally aspirated, turbocharged engines

Also using the bsfc maps for the Isuzu H-Series diesel, the Audi 5-cylinder turbo diesel and the

Toyota SI 3L V-6 engine, calculate the brake fuel conversion efficiencies at the following

conditions:

- Best efficiency

- Typical automotive light load operation (half speed and 25% load)

3. The Isuzu V12 DI diesel 12PEI-S engine described in the attached sheets is operating at 2000 rpm,

full load. At this operating condition the total friction power (consisting of rubbing friction,

pumping and accessory power) is 50 kW. Making an appropriate assumption for the relative air-

fuel ratio (remember that diesels operate lean overall) and using the manufacturer’s specifications,

calculate the following:

(a) mechanical efficiency

(b) brake fuel conversion efficiency

(c) indicated fuel conversion efficiency

(d) volumetric efficiency

The engine is working with light diesel fuel, which has lower heating value 43.2 MJ/kg and

stoichiometric air-fuel ratio 14.5.

4. Using the ideal constant volume combustion cycle model, draw an accurately proportioned

cylinder pressure versus cylinder volume diagram for a lower compression ratio engine (say 8) and

a higher compression ratio engine (say 12) that shows why the higher compression ratio engine

has a higher indicated fuel conversion efficiency. In making this comparison on a p-V diagram, a

critical question is “what should be held constant?” Possible choices are one or more of these:

maximum cylinder volume, displaced cylinder volume, mass of air in cylinder, mass of fuel in

cylinder, relative air/fuel ratio of mixture in cylinder, pressure in cylinder at start of compression,

temperature of gas in cylinder at start of compression.

(a) Explain your choice of which of the above you will hold constant in this comparison.

(b) Carefully draw the two p-V diagrams on the same graph, indicating the relative magnitude

of pressures and temperatures at start and end of compression, at start and end of

expansion.

(c) Using these p-V diagrams explain why the higher compression ratio engine is more

efficient

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Problem Set 3

Thursday, February 19, 2004

Due: Tuesday, March 2, 2004

1. A modern naturally-aspirated spark-ignition engine has an indicated fuel conversion efficiency

i f of 38 percent over its operating range. At 2000 rev/min, the mechanical friction mean ,

effective pressure (mfmep) is 90 kPa and the accessory (water, oil and fuel pump) friction (afmep)is 20 kPa; both are independent of load. The engine operates at stoichiometric and MBT spark timing at all intake manifold pressures. The exhaust pressure pe is 1 atmosphere.

(a) Develop an expression for the brake fuel conversion efficiency as a function of i f ,,

mfmep, afmep, intake manifold pressure pi, exhaust manifold pressure pe, fuel heating

value QHV, fuel/air ratio, intake system air density i a , just downstream of the throttle and ,

volumetric efficiency based on i a .v ,

(b) Evaluate the engine’s mechanical efficiency and brake fuel conversion efficiency, at an intake manifold pressure of 0.4, 0.7, and 1.0 bar

(c) Assuming that the mechanical and accessory friction values remain the same at idle, determine the intake manifold pressure at idle

2. An internal combustion engine is being operated with a pure alcohol as fuel. The fuel molecule has 7 carbon atoms. A detailed analysis of the exhaust gas constituents on a wet molar basis (i.e., with the H2O present) in the exhaust manifold yielded the following:

CO2: 13%, CO: 5.2%, H2O: 9.1%, H2:1.3%, and N2: 71.3%

(a) Determine the elemental composition of the fuel molecule (b) Determine the relative air-fuel ratio. Is this lean stoichiometric or rich? (c) The engine is a 4-valve per cylinder naturally aspirated SI engine. It has a displaced volume

of 3 liters, the airflow is 0.06 kg/sec and the engine speed is 3000 rpm. Estimate the engine load (i.e., the approximate ratio of the actual brake torque to the maximum brake torque at these conditions. The friction mean effective pressure is 200 kPa at this speed and is essentially independent of load at the loads considered here. The gross indicated fuel conversion efficiency is also independent of load.

Hints: you may need to make appropriate assumptions concerning engine’s breathing characteristics and normalized torque, according to the given engine technology.

3. Calculate the following parameters for a constant-volume combustion fuel-air cycle (Fig. 5-2a from text): (a) the pressures and temperatures at states 1, 2, 3, 4, 5 and 6 (b) the indicated fuel conversion efficiency (c) the imep (d) the residual gas fraction (e) the volumetric efficiency The engine is working at full load and the compression ratio is 8.

Page 1 of 2 131

Make appropriate assumptions for the inlet pressure and temperature, exhaust pressure and air-fuel ratio.Hint: start the calculations at point 1 using a residual gas mass fraction of 3% and a mixture temperature at start of compression of 343 K. One cycle calculation should suffice.

4. Estimate from fuel-air cycle results the indicated fuel conversion efficiency, the imep, and the maximum indicated power at wide open throttle of two 4-stroke SI engines with the following specs:(a) six-cylinder, bore 9.2 cm, stroke 9 cm, compression ratio 7 and equivalence ratio 0.8. (b) six-cylinder, bore 8.3 cm, stroke 8 cm, compression ratio 10 and equivalence ratio 1.1. Assume that the actual indicated efficiency is 0.8 times the appropriate fuel-air cycle efficiency. The maximum permitted value of the mean piston speed is 15m/s. Explain also briefly why: - the efficiency of these two engines is approximately the same despite their different

compression ratios. the maximum power of the smaller displacement engine is approximately the same as that of the larger displacement engine.

5. Carbon dioxide (CO2) emissions from engines are a matter of increasing concern due to their contribution to the so-called “greenhouse effect”. Both the characteristics of the hydrocarbon fuel and engine factors determine an engine’s CO2 emissions levels. (a) Derive an expression for the mass of CO2 produced per unit brake work output from the

engine in terms of the fuel H:C ratio (assume a fuel of composition (CHy)n), the fuel’s lower heating value QLHV, and the appropriate engine efficiencies.

(b) Compare the CO2 producing potential of the following two fuel/engine combinations: - A methane (CH4) fueled spark-ignition engine with a compression ratio of 10 operating

with a stoichiometric mixture. - A diesel (CH2)n fueled compression-ignition engine with a compression ratio of 16

operating with a fuel-air equivalence of 0.4.

At the relevant load, the mechanical efficiency of the first engine is 0.6 and of the second 0.7. Assume the results of available fuel-air cycle calculations adequately describe the effects of compression ratio and equivalence ratio changes for both these fuel/engine combinations.

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Problem Set 4

Tuesday, March 2, 2004

Due: Thursday, March 11, 2004

1. Hydrogen is a possible future fuel for SI engines, considering both emissions (only water vapor is present in the exhaust) and consumption (high heating value). A disadvantage of hydrogen fuel in SI engines is that the partial pressure of hydrogen in the mixture with air reduces engine’s volumetric efficiency, which is proportional to the partial pressure of air. (a) Find the partial pressure of air in the intake manifold downstream of the hydrogen fuel

injection location at WOT (the engine is working with stoichiometric mixture). (b) Estimate the ratio of the fuel energy per unit time entering a hydrogen-fueled engine

operating with a stoichiometric mixture to the fuel energy per unit time entering an identical gasoline-fueled operating at the same speed with a stoichiometric mixture.

(c) How does the power output compare for both engines? What can you do to ensure that the output of each engine is the same? What are your limitations? (for reference see BMW paper: SAE 2003-01-3210, on 2.61 webpage)

2. A conventional spark-ignition engine operating with gasoline will not run smoothly (due to

incomplete combustion) with an equivalence ratio leaner than about =0.8. It is desirable to extend the smooth operating limit of the engine to leaner equivalence ratios so that at part throttle operation (with intake pressure less than 1 atmosphere) the pumping work is reduced. Leaner than normal operation can be achieved by adding hydrogen gas (H2) to the mixture in the intake system. The addition of H2 makes the fuel-air mixture easier to burn. (a) The fuel composition with “mixed” fuel operation is H2 + C8H18 --- one mole of hydrogen

to every mole of gasoline, which is assumed the same as isooctane. What is the stoichiometric air/fuel ratio for the “mixed” fuel?

(b) The lower heating value of H2 is 120 MJ/kg and for isooctane is 44.4 MJ/kg. What is the heating value per kilogram of fuel mixture?

(c) Engine operation with isooctane and the mixed (H2 + C8H18) fuel is compared in a particular engine at a part-load condition (brake mean effective pressure of 275 kPa and 1400 rev/min).

You are given the following information about the engine operation:

Fuel C8H18 H2 + C8H18

Equivalence ratio 0.8 0.5Gross indicated fuel conversion efficiency 0.35 0.4Mechanical rubbing friction mep 138 kPa 138 kPa Inlet manifold pressure 46 kPa ?Pumping mep 55 kPa ?

Estimate approximately the inlet manifold pressure and the pumping mean effective pressure with (H2 + C8H18) fuel. Explain your METHOD and ASSUMPTIONS clearly. Note that mechanical

efficiency m is defined as

m = bmep/imepg = bmep/(bmep + rfmep + pmep)

Page 1 of 3 133

3. For four stroke engines, the inlet and exhaust valve opening and closing crank angles are typically: IVO: 15 CAD BTDC IVC: 50 CAD ABDC EVO: 55 CAD BBDC EVC: 10 CAD ATDC

(a) Explain why these valve timings improve engine breathing relative to valve opening and closing at the beginnings and ends of the intake and exhaust strokes.

(b) Mention at least one modification you would do to the above valve timings if the engine you are designing was - intended for a race car - turbo-charged- working at low speeds

What other additional design issues are important?

4. In a spark-ignition engine, a turbulent flame propagates through the uniform premixed fuel-air mixture within the cylinder and extinguishes at the combustion chamber walls. (a) Draw a carefully proportioned qualitative graph of cylinder pressure p and mass fraction

burned xb as a function of crank angle for –90 < < 90 for a typical SI engine at wide-open throttle with the spark timing adjusted for maximum brake torque. Mark in the crank angles of spark discharge, and of the flame development period (0 to 10 percent burned)

and end of combustion, on both p and xb versus curves relative to the top-center crank position.

(b) Estimate approximately the fraction of they cylinder volume occupied by burned gases when the mass fraction burned is 0.5 (i.e., halfway through the burning process).

(c) A simple model for this turbulent flame is shown below in Fig. 1. The rate of burning of the charge dmb/dt is given by

dmb Af STudt

Where Af is the area of the flame front (which can be approximated by a portion of a

cylinder whose axis is at the spark plug position), u is the unburned mixture density, and ST is the turbulent flame speed (the speed at which the front moves relative to the unburned mixture ahead of it). The rate of mass burning is influenced therefore by combustion chamber geometry (through Af) as well as those factors that influence ST (turbulent intensity, fuel/air ratio, residual gas fraction, and EGR). Compare combustion chambers A and B shown in Fig. 1. Sketch the approximate location of the flame front when 50 percent of the mass has been burned. (A careful qualitative sketch is sufficient; however, provide a quantitative justification for your sketch.) Sketch the mass fraction burned versus crank angle curves for these two combustion chambers on the same graph, each with its spark timing set for maximum brake torque. You may assume the value of ST is the same for A and B.

(d) Compare combustion chambers A and C in the figure below, which have the same flame travel distance but have different chamber shapes. Which chamber has - the faster rate of mass burning during the first half of the combustion process; - the faster rate of mass burning during the second half of the combustion process; - the more advanced timing for maximum brake torque? Explain your answers

Page 2 of 3 134

Figure 1

5. An approximate way to calculate the pressure in the end gas (the unburned gas ahead of the flame)just after knock occurs is to assume that all the end gas burns instantaneously at constant volume. We assume that the inertia of the burned gases prevents significant gas motion while the end gas autoignites. Use the data in Figure 9.5 from the text and assume that autoignition occurs at 10 CAD ATDC. You may also need to use the equilibrium charts of Chapter 4 from the text. (a) Determine the maximum pressure reached in the end gas after knock occurs (b) Estimate the volume occupied by the end gas as a fraction of the cylinder volume just

before autoignition occurs

6. Use the GM Vortec 4.2L DOHC engine, which appeared on problem set 2 to calculate the ratio of actual gross indicated engine performance to the equivalent fuel-air cycle predictions in Figures 5-9 and 5-10. (a) Maximum power, assume total friction mep = 200 kpa (b) Maximum torque, assume total friction mep = 150 kpa (c) Best brake specific fuel consumption: the best measured value is 250 g/kWh at part throttle

and total friction mep = 140 kpa. At WOT the engine operates rich with fuel-air equivalence ratio of 1.2. The engine operates stoichiometric at other load points for emission control reasons. Explain briefly the reasons why the fuel-air cycle predicts better performance than measured

Page 3 of 3 135


Problem Set 5


Due: Tuesday, March 30, 2004

1. This problem requires you to use an Engine Simulation Program developed by General Motors

Research Laboratories, which they have made available for free to academic institutions.**

Unless otherwise stated, use the preset values of the variables, which define engine geometry and operating conditions. Assume burn duration of 50 CAD, unless otherwise stated. (a) At WOT, 2000 rpm and stoichiometric mixture conditions vary the spark timing in 10

CAD intervals from 50 to 10 CAD BTC

Plot the gross indicated mean effective pressure versus spark advance and determine the MBT spark timing. Explain briefly the shape of the curve.

Plot in the same graph max P and burned %50 versus spark advance. At MBT spark

timing, what are the values of max P and burned %50 ?

Plot the indicated specific fuel consumption (isfc) versus spark advance. What spark advance gives a minimum isfc? How does it compare to MBT spark timing? Are there any other factors that you would take into account when setting the spark timing in an engine apart from efficiency and torque output?

Plot Pmax versus spark advance. Why is Pmax important? (b) At 2000 rpm and stoichiometric mixture conditions vary the inlet pressure in 0.25 bar

intervals from 0.25 to 1 bar (recall that the inlet pressure approximately scales with the load). The burn durations are: Pi=0.25 bar, 70˚; Pi=0.5 bar, 50˚; Pi=0.75 bar, 44˚; Pi=1.0 bar, 38˚. Be sure to have the MBT spark timing for every point.

Plot the MBT spark timing versus the inlet pressure. Explain briefly the shape of the curve.

Plot in the same graph max P and burned %50 versus the inlet pressure. Do you see any

variation of these values with the inlet pressure when the spark timing is always set to MBT? How do they compare to the values you got in part (a)?

Plot in the same graph the gross indicated, the net indicated and the brake mean effective pressures versus inlet pressure. Show on the graph the pumping and the mechanical friction losses.

Plot in the same graph the gross indicated, the net indicated and the brake fuel conversion efficiencies versus the inlet pressure. In the same graph include also the mechanical and volumetric efficiencies. Explain briefly the dependence of the mechanical efficiency on the inlet pressure. Does the gross indicated fuel conversion efficiency vary significantly with the inlet pressure?

(c) At WOT and stoichiometric mixture conditions vary the engine speed in 1000 rpm intervals from 2000 to 5000 rpm. The MBT spark timing for each speed is 2000 rpm, 28˚BTC; 3000 rpm, 32˚ BTC; 4000 rpm, 36˚ BTC; 5000 rpm, 40˚ BTC. Determine the 0-100% burn duration at each of these speed conditions (that corresponds to MBT). Answeragain part (b), having now the engine speed as a parameter.

**See GM Research Publication GMR-5758, "A User's Guide for the GM Engine-Simulation Program," March 4, 1987

Page 1 of 2 136

(d) At 2000 rpm and 0.5 bar inlet pressure, vary the fuel equivalence ratio in 0.1 intervals from 0.8 to 1.2. Be sure to have the MBT spark timing for every point. The burn durations are

=0.8, 65˚; =0.9, 54˚; =1.0, 50˚; =1.1, 48˚; =1.2, 47˚; Answer again part (b), having now the fuel equivalence ratio as a parameter.

(e) Compare the simulation and the fuel-air cycle results at the operating conditions defined in part (d).

Plot on the same graph the gross indicated mean effective pressure versus the fuel equivalence ratio for both cases.

Plot on the same graph the gross indicated fuel conversion efficiency versus the fuel equivalence ratio for both cases.

2. In this problem you are asked again to use the GMR Engine Simulation Program. We are now interested in heat losses and emissions. Keep the combustion duration at 60 CAD for all cases. (a) Using MBT spark timings, sketch the following graphs:

absolute levels of total heat losses with respect to speed and load (intake pressure) for stoichiometric mixture. Use the same speed and load increments as in the last problem set and draw the total heat losses as a function of speed with contours of constant intake pressures.

relative levels of total heat losses (as a fraction of the fuel energy delivered into the engine) with respect to speed and load for stoichiometric mixture.

(b) Answer part (a) having only the fuel equivalence ratio as a parameter. The operating conditions are: WOT, 2000 rpm and MBT spark timing for each point.

(c) Explain briefly the dependence of total heat losses (both absolute and relative) on speed, load and mixture composition.

(d) At WOT, 2000 rpm and MBT spark timing for each point, plot the molar concentration of CO2, CO, H2O, O2, H2 and NO in the exhaust stream as a function of fuel equivalence ratio. Plot also the exhaust gas temperature as a function of the equivalence ratio. Explain briefly the shape of the curves you obtain.

(e) As already discussed in the lecture the Exhaust Gas Recirculation (EGR), is a widely used emission control technology for decreasing exhaust NO emissions.

Explain briefly why recycling exhaust gases reduces NO emissions. Is there any impact on the combustion process? At what loads is EGR used and why?

Plot NO emissions as a function of EGR mass fraction using the following operating conditions: 2000 rpm, stoichiometric mixture, spark timing set at 28 CAD BTDC and intake pressure such that each point has the same gross indicated mean effective pressure. Use as a baseline case intake pressure 50 kPa and EGR 0% and increase EGR until 30% with an increment of 10%. Plot also the intake pressure and the exhaust gas temperature as a function of EGR.

Do you think that the fraction of exhaust gases recycled must be higher or lower for diesel engines in order to have approximately the same relative impact on NO and why?

Suggest another way to affect the combustion process and NO emissions in the same way as EGR? Is there any way to control it?

(f) It is also interesting to see during what part of combustion NO is formed. Therefore, plot NO emissions on a CAD basis from –120 to 120 CAD ATDC. Also plot the in-cylinder pressure and the mass fraction of fuel burned. Use the baseline operating conditions: WOT, 2000 rpm, stoichiometric mixture and MBT spark timing.

Page 2 of 2 137

3. Explain, using your understanding of the fundamentals of spark-ignition engine combustion, the causes of the observed variation in cylinder pressure cycle by cycle. What impacts do these cylinder pressure variations have on engine operation? Do you expect diesel engines to have higher or lower cycle by cycle variability?

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Problem Set 6


Due: Thursday, April 8, 2004

1. Compare the three-valve two-plug combustion chamber shown, with a four-valve chamber with one spark plug on the cylinder axis. Make appropriate simple assumptions about the flame shape, as it grows outward from the spark(s). (a) For the four-valve, center plug, plot flame envelope area (flame outer boundary)

Af 2 h r against rf r0 through the combustion process. On the same graph, sketch 0

carefully the equivalent curve for the three-valve two-plug design assuming both plugs fire at the same time: simple geometric calculations are required to short out the physics.

(b) Plot the mass fraction burned curves versus CAD for these two combustion chambers on a qualitative but carefully proportioned graph from 45 CAD BTDC to 45 CAD ATDC. Show on the graph the spark discharge location, the TDC, the location of maximum in-cylinder pressure and the approximate end of combustion, with the spark timing for both chambers set at MBT timing.

(c) With the two-plug chamber, the burn rate can be slowed down by firing one plug later than the other. Estimate approximately the mass fraction burned after which further delaying the second plug firing will have no impact on the burn rate.

Intake Intake

Spark

plug

r0/2

r0

1200

Exhaust

h

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2. You are designing a high power density (high maximum power per unit displaced cylinder volume) port fuel injected naturally aspirated spark-ignition engine to have a maximum power of 150 kW for an automobile. Address the following questions: (a) Explain how you will achieve higher than “average” power density. (b) Explain how you will deal with knock. You may choose regular gasoline (Research octane

no. 92, Motor octane no. 82) or premium gasoline (Research octane no. 98, Motor octane no. 88) which is $0.20 a gallon more expensive

(c) Explain your emission control strategy and emission control system; you must meet strict standards for hydrocarbons, carbon monoxide, and oxides of nitrogen

(d) Based on your choices in (a), (b) and (c) estimate the required displaced volume, bore, stroke, number of cylinders, and compression ratio.

(e) Provide a description of your combustion system, explaining the objectives and logic behind your choices.

3. In the disked-shaped combustion chamber of a spark-ignition engine, the flame propagates radially outward from the spark plug at the center (on the cylinder axis). The flame can be modeled as a thin smooth sheet at the location within the wrinkled turbulent flame, which on average separates the unburned and burned gases.

(a) Derive an equation for the mass fraction burned x b mb m as a function of rf ro (rf :

flame radius and ro : cylinder radius) and the density ratio rd u b . Plot this curve for

an appropriate value of rd.(b) Draw a carefully proportioned graph of xb versus CAD for the engine operating at part load

with MBT spark timing (30 CAD BTDC), indicating the approximate mass fraction burned

at TDC and 10 CAD ATDC. Note that xb at max P is about 0.75.

Make an appropriate estimate of Pmax. The pressure in the cylinder at IVC (50 CAD ABDC) is the inlet pressure, i.e. 0.5 bar. The cylinder volume ratio, volume at IVC to volume at Pmax is 7.3. The mixture is stoichiometric and the compression ratio is 10. Make reasonable assumptions of any other quantities you may need.

4. A spark-ignition engine driving a car uses, on average, 120 grams of gasoline per mile traveled. The average emissions from the engine (upstream of the catalyst) are 1.5, 2, and 20 grams per mile of NOx (as NO2), HC (interpret HC measurement as hydrocarbons with H/C ratio of 1.85), and CO, respectively. The engine operates with a stoichiometric gasoline-air mixture. (a) Find the average concentrations in parts per million of NOx, HC (as ppm C1), and CO in the

engine exhaust. (b) Calculate the average combustion inefficiency associated with the given emissions levels.

Include any hydrogen you estimate would be present in the exhaust stream. (c) Assuming the engine has a three-way catalytic converter with efficiency of 95%, find the rise

in temperature in the catalyst, after the pollutants are removed. Neglect heat losses.

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141

142

143

144

145

146

147

148

149

150

151

Rating rev

min

mep

bar

cm

m/s

A

A

B

C

D

E

F

tonnes 555 655 770 900

90

94

13.79

14.79

5.70

5.95

bw

bhp

kw

bhp

96 13.42 5.21 kw

bhp

102 14.31 5.16 kw

bhp

Number of cyls

RLB 90

Stroke 1900 mm Scre 900 mm

Net weight, without water and oil

10 780

11 270

4 000

1 800

9 250

1 520

13 150

12 500

12 550

4 000

1 800

9 250

1 520

13 150

14 220

14 710

4 000

1 800

9 250

1 185

13 150

4 000

24 360

1 800

9 250

1 520

13 150

4 000

1 800

9 250

1 520

13 150

4 000

1 800

9 250

1 520

13 150

4 000

1 800

9 250

1 520

13 150

4 000

1 800

9 250

1 185

13 150

1 005 1 120 1 255

ERP 2

MCR 2

10 000

13 600

11 200

15 200

12 500

17 000

14 000

19 000

15 000

20 400

16 800

22 800

17 500

23 800

19 600

26 600

20 000

27 200

22 400

30 400

22 500

30 600

25 200

34 200

25 000

34 000

28 000

38 000

30 000

40 000

33 600

45 600

ERP 1 10 500

14 400

13 250

18 000

15 900

21 500

18 550

25 200

21 200

23 800

23 850

32 400

25 500

36 000

31 800

43 200

MCR 1 11 760

16 000

14 700

20 000

17 640

24 000

20 580

28 000

21 520

32 000

25 460

36 000

29 400

40 000

35 280

48 000

1 470

15 780 17 480 19 200 20 320

Typical performance diagram

SULZER RLB 90 2-STROKE ENGINE

(TURBOCHARGED)

Outputs, weights and dimensions

152


Design Project – Corrected Version Number 2

Wednesday, April 14, 2004

Due: Thursday, April 22, 2004

Heavy duty diesel engine with EGR and particulate trap

For this project you need to design a heavy-duty truck diesel engine. The engine is 11 liter, 6 cylinder, with target maximum power of 360 kW and target bsfc of 185 g/kW-hr.You will also incorporate an emissions strategy to meet current emissions standards.

1. Base Engine Begin by sizing the base engine (no turbocharger), which will be 11 liter, 6 cylinder.Assume a maximum mean piston speed (Sp) of 10 m/s. Use data on Figure 13.7 on page 722 in the text for estimates of mechanical/rubbing plus auxiliary mep. Making any other reasonable assumptions necessary, determine the following parameters:

(a) Bore and stroke (b) Compression ratio (c) Connecting rod length (d) Brake mean effective pressure, at maximum torque and maximum power (e) Maximum torque and maximum power (f) Maximum engine speed, at maximum power

2. Boost, turbomachinery, and intercooler Design the required turbomachinery and intercooler for the engine by addressing the following points:

(a) Based on your calculations in part 1, calculate the amount of boost pressure required at maximum speed to produce the target power.

(b) Use a turbocharger to produce this boost, and define the main operating parameters of the required turbomachinery. For both turbine and compressor, provide values for mass flow rate, pressure ratio, inlet and exhaust temperatures

and pressures. Use typical values for isentropic efficiencies ( t=0.85, c =0.80) and assume the exhaust temperature is 900K.

(c) Include an intercooler to lower the temperature of the air coming out of the compressor. Provide the inlet and outlet temperatures of the air, as well as the coolant’s inlet and outlet temperatures and mass flowrate. Assume a counter-flow heat exchanger with effectiveness of 0.8. (Heat exchanger effectiveness is the ratio of the actual heat transfer to the heat transfer that would occur if the stream with the minimum capacity rate were heated (or cooled) from its inlet temperature to the inlet temperature of the other stream).

(d) Draw a schematic of your system (Hint: You will need to include the effect of turbo-charging on pumping work)

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3. Brake efficiency Evaluate the brake fuel conversion efficiency of the design at half maximum speed and full load at that speed. Does it meet the target bsfc? If not, what changes in engine design would bring it closer?

4. Emissions – NOX The engine you are designing must comply with 2004 EPA NOx requirements which are set at 2.5 g/bhp-hr (assume for simplicity that the NOx requirement must be met at all operating points); EGR has been chosen as the technology to reduce NOx levels.Address the questions below using available data, and an appropriate safety factor:

(a) Draw a schematic of your system, showing clearly how you will drive EGR from exhaust to intake. There are a few possibilities for doing this (the reference paper on EGR systems might be helpful).

(b) Find the required amount of EGR to run at low load (25% max torque and 1600 rpm).

(c) Find the required amount of EGR to run at maximum power. (d) Similar to part 2, recalculate the boost pressure at maximum power. Resize the

turbomachinery and intercooler to reach the stated target or best case power output; use available data to determine the exhaust temperature. Also, include an EGR cooler to lower the re-circulated gas temperature before it enters the engine; using the same assumptions as part 2c, provide operating temperatures and flowrates.

As a safety measure, it is common standard to reduce the amount of NOx, by an additional 20% to 40% of the required EPA standard. For this design please use a safety factor of 30% (i.e., reduce NOx to 1.75 g/bhp-hr, 30% below required standard). Assume that the equivalence ratio based on the mass of fresh fuel and fresh air must stay below the smoke limit of 0.7; for simplicity once you have selected the level of EGR, assume NOx levels remain constant, in spite of additional boosting (this is not the actual case). Also assume that beyond 8 CAD BTDC, for every additional CAD delay in injection you lose 0.25 percentage points in indicated fuel conversion efficiency. (Note: Watch the units in the emissions data)

5. Emissions – particulate The engine you are designing must also comply with 2004 EPA particulate requirements. Diesel Particulate Filter (DPF) has been selected as the technology to achieve the target PM levels of 0.05 g/bhp-hr. Assume that current DPF technology can reach 99% efficiency (i.e., 99% removal of particulates). Using a PM emissions safety factor of 50%, answer the following questions, all for maximum power conditions

(a) What is the approximate level of PM coming out of the engine?

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(b) Size the trap and calculate the average pressure drop. Assume a space velocity in the range of 10,000 – 28,000 hr-1 that minimizes the physical size of the particulate trap (see SAE 2003-01-0047 for reference).

(c) What impact does the particulate trap have on the performance of the turbocharged engine (be quantitative). What changes in boost pressure and turbomachinery operating conditions are needed to keep best case output?

6. 2007 Emissions requirements: Below is a table showing EPA Diesel engine emissions requirements for 2007.

NOx (g/bhp-hr) PM (g/bhp-hr)

0.20 0.01

(a) Based on engine out NOx levels of part 4, how efficient a NOx catalyst is needed to meet 2007 emissions levels? Assume that the catalyst is used in conjunction with EGR.

(b) Is it possible to achieve these levels of PM with the trap described in the SAE 2003-01-0047 paper?

(Note: Keep the same safety factors as in parts 4 and 5)

Page 3 of 2 155

Ex

ha

us

t T

em

pe

ra

ture

vs

. E

GR

(F

or D

iffe

re

nt

Sta

rt

of

Injec

tio

n @

Ma

xim

um

Po

we

r)

80

0

85

0

90

0

95

0

10

00

10

50

5%

10

%1

5%

20

%2

5%

30

%3

5%

% E

GR

Temperature (K)

0.5

CA

D B

TD

C7

.5 C

AD

BT

DC

3.5

CA

D B

TD

C2

.5 C

AD

AT

DC

6.5

CA

D B

TD

C

156

25

% M

ax t

orq

ue,

16

00

RP

M

bsN

Ox

vs. S

tart

Of M

ain

Inje

ctio

nLo

w-L

oad

Con

ditio

n

0.00

0

0.50

0

1.00

0

1.50

0

2.00

0

2.50

0

3.00

0

3.50

0

4.00

0

4.50

0

5.00

0

-8-6

-4-2

02

46

810

Star

t of M

ain

Inje

ctio

n [C

A F

rom

TD

C]

bsNOx [g/kW-hr]

Incr

ease

d EG

RSt

ock

EGR

Red

uced

EG

R

Incr

easi

ng E

GR

23.8

%

10.9

%

11.0

%

10.8

%

9.1%

8.6%

31.8

%32

.1%

32.3

%

32.2

%

32.7

%

23.6

%24

.4%

24.2

%

24.9

%

157

bsN

Ox

vs. S

tart

Of M

ain

Inje

ctio

n (M

axim

um P

ower

)

0.00

0

1.00

0

2.00

0

3.00

0

4.00

0

5.00

0

6.00

0

7.00

0

8.00

0

9.00

0 -10

-8-6

-4-2

02

46

8

Star

t of M

ain

Inje

ctio

n [C

A F

rom

TD

C]

bsNOx [g/kW-hr]

Incr

ease

d E

GR

Sto

ck E

GR

Red

uced

EG

R

Incr

easi

ng

19.9

%20

.6%

20.4

%

20.2

%

20.0

29.5

30.9

29.6

%

29.5

8.5%

9.4%

7.5%

7.4

7.6%

158

bsPM

vs.

Sta

rt O

f Mai

n In

ject

ion

Low

-Loa

d C

ondi

tion

0.00

00

0.02

00

0.04

00

0.06

00

0.08

00

0.10

00

0.12

00

0.14

00

0.16

00

0.18

00

-8-6

-4-2

02

46

810

Star

t of M

ain

Inje

ctio

n [C

A Fr

om T

DC

]

bsPM [g/kW-hr]

Incr

ease

d EG

RSt

ock

EGR

Red

uced

EG

R

Incr

easi

ng

G

20.0

%

22.2

22.9

23.8

23.8

%31

.8

32.1

32.3

32.2

32.7

%

10.9

%

11.0

10.8

9.1%

8.6%

25

% M

ax t

orq

ue,

16

00

RP

M

159

bsP

M v

s. S

tart

Of M

ain

Inje

ctio

n (M

axim

um P

ower

)

0.00

0

0.05

0

0.10

0

0.15

0

0.20

0

0.25

0

0.30

0

0.35

0

0.40

0

0.45

0

0.50

0 -10

-8-6

-4-2

02

46

8

Sta

rt o

f Mai

n In

ject

ion

[CA

Fro

m T

DC

]

bsPM [g/kW-hr]

Incr

ease

d E

GR

Sto

ck E

GR

Red

uced

EG

R

Incr

easi

ng

8.5%

9.4%

7.5

7.4%

7.6%

29.5

%

30.9

29.6

29.5

%

20.0

%20

.2%

20.4

19.9

%

20.6

%

160

Displacement (m3) 11

Cylinders 6

Bore (m) 0.1326

Stroke (m) 0.1326

Compression Ratio 18

Connecting Rod Length (m) 0.3315

Bmep @ max torque (kPa) 895

Bmep @ max power (kPa) 835

Maximum torque (N-m) 783

Maximum power (kW) 173

Maximum engine speed at maximum power(RPM) 2261


Design Project Solution

Here is a possible solution for the design problem.

1. Base Engine

Table 1 below summarizes the main parameters of the base engine

Table 1 Base Engine Summary

There are two possible methods to size the engine, and they should be consistent with each other:

Method 1:

2

/,, stoichoaHVdvifm AFQVNP (1)

Method 2:Assume a bmep based on practical limits and fuel-air cycle charts, and solve for the power output:

2000

NVdbmepP (2)

161

Using the first method, we must determine N,,,,, ,, oavifm

Calculate Engine Speed (N)

L2

pSN max find L (3)

Assume B ~ L, so

4

L6

4

B6V

32

d

Lcylinders (4)

mL 1326.06

(0.011)4

6

V4 3

1

3

1

d

so:

RPM2260or,sec/71.37)1326.0(2

10m/sN revs

Determine , and f,i

Chose rc=18 (maybe a bit high), and =0.7 (smoke limit, maximum possible fuel we can

get in per mass of air). Using Fuel-air cycle results (Fig. 5-9, Heywood p. 182), then f,i=

0.575. Applying a correction factor of around 80%, actual f,i= 46%. The correction factor can be between 80% and 85%; For this case, I chose 80% so that Method 1, and 2, as explained above, are consistent with each other.

Determine IMEP

For phi=0.7, and rc=0.8, we get

5.10imepso5.10imep

1PPi

, (5)

Note that Pi is not atmospheric pressure. At WOT, there is a pressure loss in the intake system, due to frictional losses that scale with speed. Pi will be less than atmospheric. Likewise, the exhaust pressure (Pe) is not atmospheric; a higher than atmospheric pressure is needed to pump the gases through the exhaust system. Once the gases leave the exhaust system and reach ambient conditions, they will expand to atmospheric pressure. Additionally, depending on the opening timing of the exhaust valves, the gases might exit at a higher pressure than what is required to overcome the pumping loss in the

162

exhaust system. To get an idea, of the value of Pi, look at Figure 13-13 in the text (Heywood P. 725). For a piston speed of 10 m/s,

(6)

Now allocate this pumping loss between Pe and Pi. At high speeds around 18% of the loss is on the intake side, and the remaining 82% on the exhaust side. This will be consistent with volumetric efficiency as explained below. So:

Pi= 101 kPa – 0.18(40 kPa) = 93.8 kPa Pe=101 kPa + 0.82(40 kPa) = 133.8 kPa

We can now calculate an imep:

9kPa.9845.10kPa8.93imep

Determine Mechanical Efficiency m

;1imeptfmep

imeptfmepimep

m (7)

wherepmepfmeptfmpe mep)auxiliaryandfriction(rubbingmepfriction total

From figure 13-7 (Heywood p 722), fmep for a fired engine at 2260 rpm 140 kPa. So

%7.81985

1801

and;18040140

m

kPakPatfmpe

Determine Volumetric Efficiency and oa,

Using figure 6-8 (Heywood p. 217), assume a volumetric efficiency of 90% for a piston speed of 10 m/s. Note that this volumetric efficiency measures the efficiency of the entire intake system. Also note that we have chosen the right pressure loss allocation for the intake system (as calculated in the imep section), consistent with volumetric

efficiency. The air density oa, , is just calculated from ideal gas law, at ambient

conditions. The value is 1.17 kg/m3

40)10(4.0p)S(4.0Pi)-(Pe 22xpmep

163

Fuel-to-Air Ratio & Heating Value

From table D.4 in the text (Heywood p. 915) we get the stoichiometric Fuel-to-Air ratio of gasoline as 0.0697, and its heating value of 43.2 MJ/kg

Power calculation

With the estimates for each value, we can now calculate the power

kWP

mkgkgkJemmP

173

2

)0697.0)(7.0(/17.1/32.43)011.0sec)(/7.37)(90.0)(46.0)(817.0( 3

We also use method 2 to check for consistency. Rearranging equation 2.19b (Heywood p50), we get:

kWrevdmkPaNVdbmepP 1672000

sec/7.37311805

2000

the methods are close

For low loads, follow the same procedure, with lower pumping loss, due to lower speed (see figure 13-13, Heywood), and lower rubbing and auxiliary friction (see figure 13-7 Heywood); additionally, the allocation of pressure losses is different, and must be consistent with volumetric efficiency.

2. Boost, Turbo-machinery and Intercooler

Boost pressure:

To find the boost pressure required, we use equation 1, and replace the volumetric efficiency for the entire inlet system with the volumetric efficiency for the valves only

( v ~ 94%). We also replace the ambient air density with the air density right before the

valves, ia , . This density can be determined from the ideal gas law, knowing the pressure

(which is approximately cylinder pressure divided by volumetric efficiency), and the temperature (about the same as the cylinder temperature). Thus, we can vary the cylinder pressure until we get the required power level, as defined by equation 1.

2

/,, stoichiaHVdvifm AFQVNP

Note that as we vary the cylinder pressure, and consequently the density, the mechanical efficiency (as defined by equation 7 above) will also change because the pumping loss will change.

164

Pmep= Pexhaust - Pintake

Thus the solution to this problem is iterative, and can easily be done with a spreadsheet.After varying the cylinder pressure, determining the corresponding air density at the valves through the ideal gas law, and calculating the mechanical efficiency, we get the following target power:

kWP

mkgkgkJemmP

360

2

)0697.0)(7.0(/065.2/32.43)011.0sec)(/7.37)(944.0)(46.0)(918.0( 3

For this case the pressure that gives a density of 2.065 is 176 kPa, as dictated by the ideal gas law:

)944.0(314/97.28

/314.8/065.2* 3

_)()(, Kkmolekg

kmoleKkJmkgRTP valvesvesbeforevalvesbeforevalviacylinder

which gives Pcylinder=176 kPa. The pressure that must come out of the compressor is approximately:

kPakPaPP

v

cylindercomp 186

944.0

176

Thus the desired boost is 85 kPa. That is we have to compress 85 kPa above atmospheric. Note that to relate pressure before the valves, and after the valves, as a first approximation I have used the volumetric efficiency.

Turbo-machinery

Knowing the desired boost, the turbo-machinery can now be sized to generate the required pressure. This is done using the insentropic relationships for the compressor and turbine. First we must size the compressor by finding the work required to compress the gas to the desired pressure. Second, we must size the turbine to produce the work that drives the compressor.

To determine the amount of work that is required to compress the gas we do an energy balance assuming an adiabatic compressor:

)( 12 TTCmW apc (8)

where,

T2a= Actual compressor exit temperature T1= Compressor inlet temperature (300K)

165

We calculate T2a using the compressor efficiency, and isentropic relationships:

112

2 TTTTc

sa (9)

and:1

2

112 P

PTT s (10)

T1, P1, P2, , and c are all known, so T2a, and consequently the compressor work can be calculated.

Knowing the compressor work, we now size the turbine using the following equations:

m

ct

WW (11)

where m is the mechanical efficiency for the turbine and compressor system. 95% is

reasonable estimate for this number

CpW

TT ta 45 (12)

where:T5a=Actual turbine exhaust temperature T4 = Turbine inlet temperature (engine exhaust, given at 900K)

To find the required turbine pressure ratio:

1

4

5

5

4

TT

PPP s

r (13)

where:

445

5 TTTTt

as (14)

Thus, enough equations for enough unknowns. Values for the temperatures, pressures, and compressor work, are show in table 2.

Intercooler

Adding an intercooler to lower the intake temperature, will increase the density of the gas, and consequently decrease the required boost, as reflected in table 2. For a given pressure rise we get a higher change in density (due to lower gas temperatures going into the engine). To size the intercooler you can select a coolant, and based on adequate

166

compressor turbine

Heat Exchanger Engine

1

2 3

4

5

67

estimates for inlet and outlet coolant temperatures, you can determine the required mass flow-rate that is needed to achieve a certain temperature change in the air. You must use the definition for heat exchanger to determine the allowed change in air temperature: For the coolant I used water (Cp=4.2 kJ/kg K), and assumed that it goes in at 300 K, and I want it to leave at 380 K.

To find the exit temperature of the air, and to determine the required flow rate of water, I use the definition for heat exchanger effectiveness in conjunction with an energy balance:

For effectiveness we have:

)(

)(

max_min

airor,

outin

coolantoutinairorcoolant

TTCpmTTCpm

and for the energy balance we have:

aircoolant TCpmTCpm

For this case, I chose the air and water to have about the same capacitance (mCp). Using the effectiveness equation I can solve for Tout air. Note that the capacitances will cancel out in the equation, and the maximum change in temperature occurs when Tout air= Twater in, thus:

)()(

)(____

__

irrinwaterinairinairoutair

inwaterairin

aoutin TTTTTT

TT

Assuming, water temperature increases from 300 to 357, then Tair in is 314 K. Values for the heat exchanger temperatures and flow-rates are also shown in table 2.

Figure 1

Schematic of Turbocharged Engine with Intercooler

167

State Temperature (K) Pressure (kPa) Temperature (K) Pressure (kPa)1 300 101 300 1012 401 233 372 1863 401 233 314 1864 900 223 900 1885 801 127 833 1276 N/A N/A 101 3007 N/A N/A 101 380

Work Compressor 40 kW Work Compressor 29 kW

Intercooler m_dot water 0..097 kg/sec

m_dot air 0.404 kg/sec

No intercooling With intercooling

Table 2

Turbocharged Engine with Intercooler: Operating Parameters

3. Brake Efficiency

At half maximum speed, and full load at that speed, I kept the same boost, but lowered the fmep, per figure 13-7 in the text. The BSFC came out to be 188 g/kw-hr. This number is actually quite good for industry standards. Other people perhaps got lower (around 175), however, as I previously explained, I was more conservative in my efficiency estimate from fuel air cycle tables, to be consistent with different ways of calculating power. My calculation is shown below:

hrkWgkgghrkgkWPowerhrgm

bsfc f/188

193

)/1000)(sec/3600sec(/010078.0

)(

)/(

Note that if we directly use break engine efficiency, we should get the same answer:

hrkWgQQ

bsfcHVifmHVbf

/188)2.43)(460.0(96.0

1

)(

1

)(

1

,,

Ways to decrease bsfc include raising compression ratio, and reducing frictional losses.

4. Emissions – NOx

168

compressor turbine

Inter-

coolerEngine

1

2 3

5

6

Inter-

coolerVenturi/

Mixer

4

7

Engine Mode NOx Standard g/bhp-hr

NOx safety target g/bkW-

hr

NOx Safegy target g/bkW-

hrEGR Timing (CA

from TDC)

Hit in Fuel economy

(percentage points)

25% max torque, 1600

rpm2.50 1.75 2.35 24% 1 2%

Maximum power 2.50 1.75 2.35 24% 0.5 2%

Figure 2

Schematic of Turbocharged Engine with Intercooler

The schematic shows how EGR will be driven from the engine. There are a few ways of doing this; one way is to use a Venturi system, as shown above. Another way is to optimize the system so that the pressures at the air and EGR intersection are about the same. It is necessary for these pressures to be equal, otherwise there will be backflow in the direction of lower pressure. Overall, however, the addition of EGR will impact the fuel economy of the engine. This is the price that we must pay to have lower emissions.

The first step of this problem is to define the amount of EGR that is needed to meet EPA emissions levels. The emissions requirements along with their safety levels are shown in table 3 below.

Table 3

169

Must operate

below dashed

line

Using the figures provided, there are various possibilities for selecting EGR, depending on the hit on fuel economy. Figure 3 below is an example that shows that there is a range of timings and EGR levels that will give the proper amount of NOx

Figure 3

Acceptable operating area for low load

Once EGR has been calculated, the loss in engine efficiency can be assessed, as well as the required boost. Again this is an iterative process. There are many variables affecting engine power, and they are all related as well, thus at least a spreadsheet must be setup. For example, boost affects engine power, but it also affects mechanical efficiency, which in turn affects engine power, thus all these variables must be connected when solving the system.

One important implication of adding EGR, is that the pressure in the cylinder chamber must increase if we are to maintain constant mass of fresh fuel and air; this is what we should desire if we are to maintain the same power output from the engine as the case without EGR.

The total pressure is equal to the sum of the partial pressures of air and the EGR:

170

State Temperature (K) Pressure (kPa)1 300 1012 434 2953 327 2954 349 2795 990 2416 864 1217 438 241

With intercooling

Intercooler Mass flowrate (kg/sec) Tin (K) Tout (K)

EGR 0.171 300 380

Compressor 0.131 300 380

EGRairT PPP

Assuming the molecular weights of both Air and EGR are about the same, then the mole fraction is approximately equal to the mass fraction of each mixture, and PT can be expressed as:

EGRPP air

T1

Additionally, since there is a pressure loss of around 16 to 20 kPa associated with the venturi, a higher boost is still needed. To reach the target power output, a total boost of 194 kPa is required, for total pressure of 295 kPa. This is a high boost, higher than industry standard for this size engines. Perhaps a more practical boost is 150 kPa (PT

=251 kPa), or less. However this limits the maximum power to 310 kW. If you recognized the practical limitations, this is a perfectly acceptable answer.

Table 4

Table 5

171

5. Emissions Particulates

(a) The particulate emissions corresponding to the chosen EGR level, can be obtained from the data provided (PM levels vs injection timing for various EGR fractions). At 24% EGR, the PM coming out of the engine is approximately 0.2 g/bkW-hr

(b) To size the trap, use the space velocity that will minimize the volume of the trap, in this case 28,000 hr-1. This is evident from the relationship for space velocity:

velocitySpaceVV

_

where V is the volume flow-rate of the gases going through the trap, and V is the volume

of the trap. For a smaller trap volume we get a higher space velocity.

Solving for the volume of the trap we get

velocitySpaceVV_

Using the ideal gas law to solve for V

sec/10.1)(

3mP

RTmmmV

exhaust

exhaustegrfuelair

Solving for Volume

LmmV 141141.0sec3600/000,28

sec/10.1 33

Values used are shown in table 6 below

Table 6

As shown in figure 4 of SAE 2003-01-0047, The maximum pressure loss through the trap, at a space velocity of 28,000/hr, is 6 kPa. This is a very small percentage of the total exhaust pressure (~2.5%), and the effect on turbo-machinery is small.

Mdot_air&fuel (kg/sec)

Mdot_egr (kg/sec) Texhaust (K) Pexhaust (kPa) Space Velocity

(1/sec)

Volume flowrate (m3/sec)

Trap Volume (L)

0.43 0.10 864.74 121.00 7.78 1.10 140.87

172

6. 2007 Emissions requirements

Using the same safety factor as in part 6, the table below shows the new emissions that must be met:

Table 7

Where catalytic converter efficiency is defined as:

intpollu

outtpollucat m

m

,tan

,tan1

As shown in table 7, a catalytic converter with 92% efficiency will be needed to meet 2007 NOx emissions requirements.

The required particulate trap efficiency is fairly high (97%) but the trap presented in the Ford paper seems to have efficiencies of around 99%, so it should work fine for 2007 emissions requirements.

NOx Standard g/bhp-hr

NOx safety target

g/bhp-hr

NOx Safegy target

g/bkW-hr

Current Engine out g/bkW-hr

Required efficiency (catalytic)

PM Standard g/bhp-hr

PM safety target

g/bhp-hr

PM Safegy target

g/bkW-hr

Current Engine out g/bkW-hr

Required Efficiency

(trap)

0.2000 0.1400 0.1879 2.3490 0.9200 0.0100 0.0050 0.0067 0.2000 0.966

173

Bibliography on Internal Combustion Engines

Obert, Edward F. Internal Combustion Engines and Air Pollution. New York: Intext

Educational Publishers, 1973 edition.

A good basic text on engines from the 1950s with modest updating in 1968; much

excellent descriptive material.

Taylor, C. Fayette, and Edward S. Taylor. The Internal Combustion Engine. International

Textbook Company, 1961.

A basic text now out of print and somewhat dated.

Taylor, C. F. The Internal Combustion Engine in Theory and Practice. Vol. 1, and 2.

Cambridge, MA: M.I.T. Press, 1966 and 1968. Reissued in paperback in 1977, and in

1985 as Second Edition with minor modifications.

A much expanded version of reference 2; an advanced text with extensive material on

engine design practice of the 1950s and 60s.

Rogowski, A. R. Elements of Internal Combustion Engines. New York: McGraw-Hill,

1953.

An elementary text used primarily for undergraduate teaching.

Lichty, L. C. Combustion Engine Processes. 6th ed. New York, : McGraw-Hill Book

Company, 1967. ISBN: 0070377200.

A good basic text on all types of combustion engines, now somewhat dated.

Khovakh, M., (general editor). Motor Vehicle Engines. English translation form Russian.

Moscow: MIR Publishers, 1976.

A Russian text with an excellent ordering of subject material.

Patterson, D. J., and N. A. Henein. Emission from Combustion Engines and their Control.MI: Ann Arbor Science Publishers, Inc., 1972.

A comprehensive text on engine emissions; now somewhat dated.

Ayres, Robert U., and Richard P. McKenna. Alternatives to the Internal Combustion Engines. Baltimore: Johns Hopkins University Press, 1972. ISBN: 0801813697.

A fundamental text on the alternative engines to the internal combustion engine.

Nunney, M. J. The Automotive Engine. London: Newnes-Butterworths, 1974.

A book which reviews modern automotive engine practice; contains descriptions of

design and operation of engines and engine components.

Yamamoto, Kenichi. Rotary Engine. Mazda: Toyo Kogyo Co., Ltd., 1969.

Excellent text on the design and operation of Wankel engines.

174

Ansdale, R. F. The Wankel RC Engine: Design and Performance. London: Iliffe Books,

Ltd., 1968.

Contains much technical and historical information on the Wankel engine.

Springer, G. S., and D. J. Patterson, eds. Engine Emissions: Pollutant Formation and Measurement. New York, and London: Plenum Press, 1973.

A set of contributed chapters on different emissions topics; some chapters are still useful.

Sitkei, G. Heat Transfer and Thermal Loading in Internal Combustion Engines.Budapest: Akademiai Kaido, 1974.

A monograph on heat transfer in spark-ignition and diesel engines and temperature

distributions in engine components.

Annand, W. J., and G. E. Roe. Gas Flow in the Internal Combustion Engine. Haessner

Publishing, Inc., 1974.

A review of selected topics related to gas flow in IC engine intake and exhaust systems.

"Should We Have a New Engine?" An Automobile Power Systems Evaluation. Vol. 1.

Summary, Jet Propulsion Laboratory, California Institute of Technology, JPL SP 43-17,

August 1975.

Popular summary of study which evaluates the internal combustion engine and its

alternatives.

Goodger, E. M. Hydrocarbon Fuels; Production, Properties and Performance of Liquids and Gases. London: Macmillan, 1975.

Useful review of fuels, automotive and non-automotive.

Cummins, Lyle. Internal Fire: The Internal Combustion Engine 1673 - 1900 Revised Edition. 2nd ed. Warrendale, PA: Society of Automotive Engineers, 1976.

Excellent and readable history of the internal combustion engine by the son of the

founder of the Cummins Engine Company.

A History of the Automotive Internal Combustion Engine. Warrendale, PA: Society of

Automotive Engineers special publication, SP-409, 1976.

A set of four SAE papers reviewing the history of IC engine developments.

Blackmore, D. R., and A. Thomas. Fuel Economy of the Gasoline Engine. New York,

NY: John Wiley & Sons, 1977.

A useful introduction to how fuel properties affect spark-ignition engine operation.

Thomson, W. Fundamentals of Automotive Engine Balance. London: Mechanical

Engineering Publications, Ltd., 1978.

A short straightforward monograph on the balancing of various arrangement

reciprocating engines.

175

Benson, R. S., and N. D. Whitehouse. Internal Combustion Engines. Vol. 1, and 2.

London: Pergamon Press, Inc. 1979.

A modern text, limited in scope, with special emphasis on computer simulations of

engine flow and combustion processes.

Watson, N., and M. S. Janota. Turbocharging the Internal Combustion Engine. New

York: John Wiley & Sons, 1982.

An extensive and excellent professional reference text on turbochargers, and

turbocharged engine performance.

Benson, R. S. The Thermodynamics and Gas Dynamics of Internal Combustion Engines.Vol. 1. Edited by J. H. Horlock, and D. E. Winterbone. Oxford: Clarendon Press, 1982.

Extensive and detailed monograph on unsteady engine intake and exhaust flow processes.

Horlock, J. H., and D. E. Winterbone, eds. The Thermodynamics and Gas Dynamics of Internal Combustion Engines. Vol. 2. Oxford: Clarendon Press, 1986.

Extensive and detailed monograph on in-cylinder engine processes and methods of

analysis.

Hilliard, J. C., and G. S. Springer, eds. Fuel Economy in Road Vehicles Powered by Spark Ignition Engines. New York, and London: Plenum Press, 1984.

A set of contributed chapters on engine and vehicle factors which affect fuel economy;

some are excellent.

Stone, R. Introduction to Internal Combustion Engines. MacMillian Publishers, Ltd.,

1985. 2nd ed. 1992.

An introductory text appropriate to a survey undergraduate course on engines.

Ferguson, C. R. Internal Combustion Engines--Applied Thermosciences. New York: John

Wiley & Sons, 1986.

A new text focusing primarily on Thermal/Fluids Science aspects of engine operation.

Bosch. Automotive Handbook. 5th ed. Published by Robert Bosch GmbH. Warrendale,

PA: Distributed by SAE, 2000.

A concise and useful summary of technical data on engine and vehicle components and

systems.

Heywood, J. B. Internal Combustion Engine Fundamentals. London: McGraw-Hill,

1988.

An extensive text and professional reference on the fundamentals behind engine

operation and design.

Bosch. Automotive Electric/Electronic Systems. Published by Robert Bosch GmbH.

Warrendale, PA: Distributed by SAE, 1988.

A practical guide to and description of automotive electrical systems.

176

Arcoumanis, C., ed. Internal Combustion Engines. London; San Diego: Academic Press,

1988.

A collection of contributed chapters on gasoline and diesel engines, turbocharged engines

and automotive fuels; some are good.

Blair, G. The Basic Design of Two-Stroke Engines. Warrendale, PA: Society of

Automotive Engineers, 1990.

A monograph with simple programs focused on two-stroke gasoline engine design issues

and their underlying principles.

Owen, K., and T. Coley. Automotive Fuels Handbook. Warrendale, PA: Society of

Automotive Engineers, 1990.

An extensive compilation of information on gasolines and diesel fuels and their effects on

engine operation.

Newton, K., W. Steeds, and T. K. Garrett. The Motor Vehicle. 11th ed. London; Boston:

Butterworth, 1989.

A useful source of practical information on engines, transmissions and vehicles.

Lenz, H. P. Mixture Formation in Spark-Ignition Engines. New York, NY: Springer-

Verlag, 1990.

A resource for detailed information on gasolines, carburetors, fuel injection systems, and

the mixture formation process.

Ramos, J. I. Internal Combustion Engine Modeling. New York: Hemisphere Publishing

Co., 1989.

A review and useful introduction to the various models now available for engine

processes.

Heck, R. M., and R. J. Farranto. Catalytic Air Pollution Control. New York: Van

Nostrand, Reinhold, 1995.

A readily understandable review of catalyst fundamentals and application to vehicles.

Blair, G. P. Design and Simulation of Two-Stroke Engines. Warrendale, PA: SAE, 1996.

An update and extension of Blair’s earlier book; extensive information on small high-

performance two-stroke spark-ignition engines.

Sher, E., ed. Handbook of Air Pollution from Internal Combustion Engines: Pollutant Formation and Control. San Diego, CA: Academic Press, 1998.

An extensive set of chapters, by different authors, on four-stroke and two-stroke cycle

sparkignition and diesel engine operation and emissions, and fuel effects.

Pulkrabek, W. W. Engineering Fundamentals of the Internal Combustion Engine. New

York: Prentice-Hall, Inc., 1997.

An introductory text on IC engine fundamentals.

177

Borman, G. L., and K. W. Ragland. Combustion Engineering. WCB McGraw-Hill, 1998.

A valuable reference volume on combustion processes in different practical systems,

including IC engines, with extensive information on fuels.

Heywood, J. B., and E. Sher. The Two-Stroke Cycle Engine: Its Development, Operation and Design. Warrendale, PA: SAE, Taylor & Francis, 1999.

A comprehensive summary of the technical literature on two-stroke cycle engine

processes which govern its operation and its design.

Flagan, R. C., and John H. Seinfeld. Fundamentals of Air Pollution Engineering.Englewood Cliffs, CA: Prentice-Hall, Inc., 1988.

A review of air pollutant formation processes and sources, and control approaches.

Lenz, H. P., and C. Cozzarini. Emissions and Air Quality. Warrendale, PA: SAE, 1999.

A concise handbook with data on transportation emissions, their impact, and ways to

control their magnitude.

Challen, B., and R. Baranescu, eds. Diesel Engine Reference Book. 2nd ed. Warrendale,

PA: SAE, 1999.

An extensive handbook on the theory, design, and applications of diesel engines.

Bosch. Gasoline-Engine Management. 1st ed. Published by Robert Bosch GmbH.


A handbook with extensive practical details on gasoline spark-ignition engines and their

management and control.

———. Diesel-Engine Management. 2nd ed. Published by Robert Bosch GmbH.


A handbook witih extensive practical details on diesel engines, their emissions, and their

management and control.

Stan, C., ed. Direct Injection Systems for Spark-Ignition and Compression-Ignition Engines. Published by Springer-Verlag, Berlin, Heidelberg. Warrendale, PA: Distributed

by SAE, 1999.

Multi-author volume on direct injection gasoline and diesel engines, focusing on the

different practical approaches to direct injection of liquid fuel into the cylinder.

Winterbone, D. E., and R. J. Pearson. Theory of Engine Manifold Design. Warrendale,

PA: SAE, 2000.

A text on the theory and methodology for analyzing unsteady gas flows in engine

manifolds.

———. Design Techniques for Engine Manifolds. Warrendale, PA: SAE, 1999.

A comparison text to #49, focusing on application of unsteady gas flow analysis tools to

engine manifold design.

178

Blair, G. P. Design and Simulation of Four-Stroke Engines. Warrendale, PA: SAE, 1999.

A description of engine simulations, largely developed in the author’s laboratory, and

their application to four-stroke engine performance prediction and design.

Ferguson, C. R., and A. T. Kirkpatrick. Internal Combustion Engines Applied Thermosciences. 2nd ed. NY: John Wiley & Sons, Inc., 2001.

A new edition of #27: An introductory text focusing on the thermal science processes

important to internal combustion engine operations.

Nuti, M. Emissions from Two-Stroke Engines. Warrendale, PA: SAE, 1998.

A monograph on two-stroke cycle gasoline engines, the origins of their emissions and

methods of control.

Eastwood, P. Critical Topics in Exhaust Gas after-Treatment. Research Studies Press

Ltd., 2000.

A detailed monograph on engine exhaust gas treatment—catalysts, particulate filters—as

well as exhaust treatment system issues.

Makartchouk, A. Diesel Engine Engineering: Thermodynamics, Dynamics, Design, and Control. New York, and Basel: Marcel Dekker, Inc., 2002.

Analysis based text, focused primarily on engine dynamics, structural design, and

automated diesel engine control.

Zhao, F., D. L. Harrington, and M-C. Lai. Automotive Gasoline Direct-Injection Engine.Warrendale, PA: SAE, 2002.

An extensive review of the literature on GDI engine performance, combustion,

efficiency, and emissions, and the state of GDI engine development.

179

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Internal Combustion Engines