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ENGINE PERFORMANCE IMPROVEMENT BY MODELLING OF

AIRFLOW THROUGH INTAKE MANIFOLD.

by

JAN BENJAMIN KRIEL

Submitted in partial fulfilment of the requirements for the degree

MAGISTER TECHNOLOGIAE: MECHANICAL

in the

Department of Mechanical Engineering

FACULTY OF ENGINEERING

TSHWANE UNIVERSITY OF TECHNOLOGY

Supervisor: A. Swarts

Co-Supervisor: C.F. Meyer

May 2008

i

I hereby declare that the dissertation submitted for the degree M Tech: Mechanical, at

Tshwane University of Technology, is my own original work and has not previously been

submitted to any other or quoted are indicated and acknowledged by means of a

comprehensive list of references.

Jan Benjamin Kriel

Copyright Tshwane University of Technology 2002

ii

This study is dedicated to my wife

Janine

and daughter

Elz

for their support and patience throughout the study.

iii

ACKNOWLEDGEMENTS.

I would like to express my sincere appreciation to:

Sasol Technology Research and Development, for financial assistance.

Sasol Technology Fuels Research for using their facility and equipment to carry out the

tests.

Mr S Conradie, my manager for the opportunity to complete the practical tests.

Mr A Swarts, my supervisor for assistance through the project.

Mr C.F Meyer, my co-supervisor for his assistance with the compilation.

iv

ABSTRACT.

This study was conducted to develop an unsophisticated model to predict unsteady gas

flow in the intake system of a four cylinder internal combustion engine. The results of the

model were compared to practical tests that were done in a test cell to evaluate reliability

of the model. The main findings were that the model could predict manifold pressure and

volumetric efficiency and that maximum volumetric efficiency of a specific engine can be

moved to occur at different engine speeds with different intake manifolds. It was

concluded that an unsophisticated model can be used to predict manifold pressure.

Accuracy of the model varied across the engine speed range due to the exhaust cycle that

was not included in the model which influences cylinder pressure at the end of the exhaust

cycle. Pressure amplitudes of the simulation were in phase when it was compared to the

tests done, but varied by up to 9 %, at some engine speeds. Secondly it was found that

different intake manifold dimensions influence the power output of a specific engine as

predicted by the model.

v

EKSERP

Die studie was gedoen om n ongekompliseerde model in Excel te ontwikkel wat

pulserende vloei in die inlaatspruitstuk van n vier silinder binnebrand enjin kan voorspel.

Die resultate van die model was vergelyk met die van praktiese toetse wat in n toets-

kamer onder beheerde toestande gedoen was om vas te stel of die model betroubaar is. Die

hoof bevindings was dat die model die druk in die inlaatspruitstuk en die volumetriese

effektiwiteit van n spesifieke enjin kan voorspel en dat die enjinspoed waar maksimum

volumetriese effektiwiteit voorkom kan wissel as verskillende spruitstukke gebruik word.

Dit is dus moontlik om die druk in die inlaat spruitstuk met n ongekompliseerde model te

voorspel. Akuraatheid van die model het gewissel soos die enjin spoed verander as gevolg

van die uitlaat proses wat nie in die model gesimuleer word en wat die druk in die silinder

beinvloed aan die einde van die uitlaatslag. Druk amplitudes was in fase met die van die

toetse maar het tot 9% verkil met die van die toetse op sekere toeretellings. Tweedens was

daar gevind dat verskillende spruitstukke die drywinguitset van n spesifieke enjin

beinvloed soos dit met die model voorspel was.

vi

CONTENTS

PAGE DECLARATION. i DEDICATION. ii ACKNOWLEGDEMENTS. iii ABSTRACT. iv EKSERP v CONTENTS. vi LIST OF FIGURES............... x LIST OF TABLES... xv

CHAPTER 1 1. INTRODUCTION1 1.1 History on unsteady gas flow..1 1.2 Motivation for investigation on intake manifold design.....2 1.3 Problem statement...3 1.4 Objective of the study.........3

CHAPTER 2 2. LITERATURE STUDY.....5 2.1 Project necessity.....5 2.2 Theory of pulsating flow....6 2.3 Wave formation in intake and exhaust systems..6 2.3.1 Distortion in wave profile during flow..7 2.4 Friction loss during flow.11 2.5 Heat transfer during flow.....11 2.6 Pressure waves traveling through each other...11 2.7 Reflection of pressure waves.......13 2.7.1 Reflection of a compression wave at the open end of a pipe.......................14 2.7.2 Reflection of expansion waves at the open end of a pipe.........14 2.7.3 Flow through bell mouth pipe ends......15 2.7.4 Reflection of pressure waves at sudden area change...15 2.7.5 Reflections of pressure waves in taper pipes...16 2.8 Volumetric efficiency..18 2.9 Modelling of flow....18

CHAPTER 3 3. MODELING OF FLOW.19 3.1 Assumptions used in model.19 3.2 Information required in model.....21 3.2.1 Intake manifold22 3.2.2 Cylinder head.......23 3.2.3 Intake valve..23 3.2.4 Cam shaft.23 3.2.5 Bottom end information...24 3.2.6 General information.24

vii

3.2.7 Volumetric efficiency...25 3.3 Basic procedure of model25 3.4 Mathematical processing26

CHAPTER 4 4. EXPERIMENTAL EQUIPMENT AND PROCEDURE..35 4.1 Method of investigation.............35 4.2 Equipment....37 4.2.1 Engine..37

4.2.2 Fuel used..38 4.2.3 Manifolds.....38

4.2.4 Test cell39 4.2.5 Control room...39 4.2.6 Dynamometer..40 4.2.7 Optical crank angle marker....40 4.2.8 Manifold pressure sensor...41 4.3 Calibration of equipment...42

4.4 Variation in data recorded.........44 4.5 Test procedure.........44 4.6 Data recorded and sample rates........46 4.6.1 Low frequency samples.46 4.6.2 High frequency samples.47

CHAPTER 5 5 RESULTS OBTAINED FROM EXPERIMENTS.....49 5.1 Bracket tests..49 5.2 Statistical evaluation................53 5.3 Removal of cones in air box..54 5.4 Fitment of a larger butterfly valve........55 5.5 Plenum used for tests..56 5.6 Manifolds selected for discussion.57 5.7 Manifold A...58 5.7.1 Torque developed with manifold A.......60 5.7.2 Power developed with manifold A... 62 5.8 Manifold B...62 5.8.1 Torque developed with manifold B....64 5.8.2 Power developed with manifold B... 66 5.9 Manifold G..66 5.9.1 Torque developed with manifold G...67 5.9.2 Power developed with manifold G...........69 5.10 Manifold M..............70 5.10.1 Torque developed with manifold M...........71 5.10.2 Power developed with manifold M......... 72 5.11 Discussion on performance of constant diameter manifolds....74

CHAPTER 6 6. COMPARISON BETWEEN MODEL AND TEST RESULTS.. 76 6.1 Recorded data from manifold pressure sensor 76

viii

6.2 Original manifold..78 6.2.1 Original manifold pressure comparison between model and test..79 6.2.2 Volumetric efficiency ......84 6.3 Manifold A....85 6.3.1 Test and simulated pressure comparison of manifold A..86 6.3.2 Volumetric efficiency of manifold A89 6.4 Manifold B90 6.4.1 Test and simulated pressure comparison of manifold B90 6.4.2 Volumetric efficiency of manifold B93 6.5 Manifold G...93 6.5.1 Test and simulated pressure comparison of manifold G..94 6.5.2 Volumetric efficiency of manifold G...95 6.6 Manifold M..96 6.6.1 Test and simulated pressure comparison of manifold M96 6.6.2 Volumetric efficiency of manifold M...99 6.7 Comparison between tests and model in general....100

CHAPTER 7 7. DISCUSSION.101 7.1 General data recorded...101 7.1.1 Tests done with manifold D..... 102 7.1.1.1 Torque developed by manifold D.... 103 7.1.1.2 Exhaust manifold temperature.... 103 7.1.1.3 Excess-air ratio of manifold D..104 7.1.1.4 Repeatability and reliability of data from manifold D......105 7.2 Data comparison between manifold I and the original manifold.... 105 7.2.1 Comparison of torque developed by manifold I and the original

manifold.106 7.2.2 Comparing fuel flow of manifold I and the original manifold107 7.2.3 Comparing exhaust temperature of manifold I and the original

manifold..108 7.3 Comparing different manifolds tested........ 108 7.4 Power output of manifolds with equal diameters but different lengths.112 7.5 Comparison of dual-diameter manifolds.112 7.6 Manifold L with a re-mapped ECU.....113 7.6.1 Adjustment in ignition timing...114 7.6.2 Change of injection pulse....115 7.7 Conclusion of data recorded..117

CHAPTER 8

8. CONCLUSION ..118 8.1 Summary118 8.2 Future research...120 8.2 Conclusion.120

ix

REFERENCES122

ANNEXURE A: Calculation sheets used in the model. 124

ANNEXURE B: Macros used in the model. 134

ANNEXURE C: Pressures recorded during compression test. 157

ANNEXURE D: Injector flow on test bench. 158

ANNEXURE E: Variation in data recorded at constant engine speeds. 159

ANNEXURE F: Data recorded for different manifolds tested. 165

ANNEXURE G: Kistler pressure sensor calibration verification. 194

ANNEXURE H: Nomenclature. 195

x

LIST OF FIGURES.

PAGE

Figure 2.1 Major components that influence engine performance. 5

Figure 2.2 Distortion of a compression wave as it travels through a pipe. 8

Figure 2.3 Distortion of an expansion wave as it travels through a pipe. 10

Figure 2.4 Two compression waves traveling towards each other. 12

Figure 2.5 Two compression waves traveling through each other. 12

Figure 2.6 Reflection of a pressure wave at a closed end. 13

Figure 2.7 Reflection of an expansion wave at an open end. 14

Figure 2.8 Reflection at a bell mouth entry for inflow. 15

Figure 2.9 Tapered pipe divided into sections. 17

Figure 3.1 A flow diagram of calculations done by the model. 20

Figure 3.2 Original and simulated manifold showing the meshed sections. 22

Figure 3.3 Basic valve dimensions used by model. 24

Figure 3.4 Points on pressure waves before and after time step. 27

Figure 3.5 Intake valve lift simulated in the model. 29

Figure 3.6 Minimum flow area during the intake cycle. 29

Figure 3.7 Variation in piston speed from TDC to BTC at 1250 rpm. 31

Figure 3.8 The input sheet used in model. 33

Figure 4.1 Standard production engine with original intake manifold (Encircled). 36

Figure 4.2 Engine with one of the modified intake manifolds. 37

Figure 4.3 Pressure sensor was installed on the intake manifold behind the injector. 39

Figure 4.4 Position of crank angle marker as installed on the crankshaft. 40

Figure 4.5 Schematic diagram of information recorded by the computer. 41

xi

Figure 4.6 Kistler pressure sensor calibration. 43

Figure 4.7 Production manifold flanged and pressure sensor housing welded in position. 45

Figure 4.8 Engine with original intake system. 45

Figure 5.1 Power versus engine speed for bracket tests. 51

Figure 5.2 Torque versus engine speed for bracket tests. 52

Figure 5.3 The original air box with cones installed. 55

Figure 5.4 Power versus engine speed for standard system, cones in air box removed and 60 mm butterfly with cones removed. 56

Figure 5.5 Testing for pressure variation in plenum. 57

Figure 5.6 Pressure versus crank angle in plenum with the sensor installed in different positions. 57

Figure 5.7 Power versus engine speed for various manifolds. 58

Figure 5.8 Manifold A: The shortest manifold tested with a 34.9 mm diameter. 58

Figure 5.9 Ram tubes installed to minimize entry losses. 59

Figure 5.10 Pressure versus crank angle for manifold A at different engine speeds. 60

Figure 5.11 Torque versus engine speed for manifold A and the original manifold. 61

Figure 5.12 Manifold B: the longest manifold tested. 63

Figure 5.13 Pressure variations in manifold B at different engine speeds. 64

Figure 5.14 Torque versus engine speed for manifold B and the original manifold. 65

Figure 5.15 Pressure versus crank angle for manifold G at different engine speeds. 67

Figure 5.16 Torque versus engine speed for manifold G and the original manifold. 68

xii

Figure 5.17 Pressure versus crank angle for manifold M at different engine speeds. 70

Figure 5.18 Torque versus engine speed for manifold M and the original manifold. 72

Figure 5.19 Power versus engine speed for manifold M and the

original manifold. 73

Figure 6.1 Manifold pressure versus crank angle recorded during bracket tests at 1500 rpm. 77

Figure 6.2 Manifold pressure versus crank angle recorded during bracket tests at 5000 rpm. 77

Figure 6.3 The original manifold showing the tapered primary pipes. 78

Figure 6.4 Pressure versus crank angle for simulated and experimental results at 1500 rpm. 79

Figure 6.5 Pressure versus crank angle for tested and simulated manifold pressure at 2000 rpm. 80








Figure 6.13 Volumetric efficiency versus engine speed of the test and simulation. 85

Figure 6.14 Manifold pressure versus crank angle for recorded and simulated manifold pressure at 1500 rpm. 87

xiii

Figure 6.15 Manifold pressure versus crank angle for manifold A at medium speeds. 88

Figure 6.16 Manifold pressure versus crank angle for manifold A at high engine speeds. 89

Figure 6.17 Comparing volumetric efficiency of the simulation and test. 90

Figure 6.18 Manifold pressure versus crank angle for manifold B

at 1500 rpm. 91

Figure 6.19 Manifold pressure versus crank angle for manifold B at 2500, 3500 and 4500 rpm. 92

Figure 6.20 Comparing simulated and tested volumetric efficiencies. 93

Figure 6.21 Manifold pressure versus crank angle for manifold G

at 1500 rpm. 94

Figure 6.22 Manifold pressure versus crank angle for manifold G at 2500 rpm. 94

Figure 6.23 Manifold pressure versus crank angle for manifold G at 3500 rpm. 95

Figure 6.24 Comparing simulated and tested volumetric efficiencies of manifold G. 96

Figure 6.25 Manifold pressure versus crank angle for manifold M at 1500 rpm. 97





Figure 6.30 Comparing tested and simulated volumetric efficiencies of manifold M. 99

xiv

Figure 7.1 Torque versus engine speed for tests done with manifold D. 103

Figure 7.2 Exhaust temperatures versus engine speed for manifold D. 104

Figure 7.3 Excess-air ratio versus engine speed for manifold D. 104

Figure 7.4 Manifold I consisting of two pipes with different diameters. 106

Figure 7.5 Torque versus engine speed for manifold I and the original manifold. 107

Figure 7.6 Fuel flow versus engine speed for the original manifold and manifold I. 107

Figure 7.7 Exhaust temperatures versus engine speed for the original manifold and manifold I. 108

Figure 7.8 Torque versus engine speed for manifolds tested. 110

Figure 7.9 Fuel flow versus engine speed for manifolds tested. 110

Figure 7.10 Position of lambda sensor used to record excess-air ratio in exhaust. 111

Figure 7.11 Excess-air factor versus engine speed for manifolds tested. 111

Figure 7.12 Power versus engine speed for manifolds with the same diameter but different lengths. 112

Figure 7.13 Torque versus engine speed for dual-diameter manifolds. 113

Figure 7.14 Torque versus engine speed for manifold L with standard

and optimized map. 114

Figure 7.15 Fuel flow versus engine speed for the standard and

optimized fuel map. 116

Figure 7.16 Excess-air ratio versus engine speed for the original and optimised map. 117

xv

LIST OF TABLES.

PAGE

Table 4.1 Sample frequencies used at different engine speeds. 48

Table 5.1 Performance of manifolds tested. 50

Table 5.2 Difference in power output of bracket tests. 51

Table 5.3 Difference in torque developed during bracket tests. 52

Table 5.4 Variation in torque and power of the first bracket test. 53

Table 5.5 Variation in torque and power of the second bracket test. 54

Table 5.6 Comparison of torque developed between manifold A and the original manifold. 61

Table 5.7 Comparison of power between manifold A and the original manifold. 62

Table 5.8 Torque comparison of manifold B and the original manifold. 65

Table 5.9 Power comparison between manifold B and the original manifold. 66

Table 5.10 Torque comparison of manifold G and the original manifold. 68

Table 5.11 Power comparison of manifold G and the original manifold. 69

Table 5.12 Torque comparison of manifold M and the original manifold. 71

Table 5.13 Power comparison of manifold M and the original manifold. 73

Table 7.1 Atmospheric pressure and temperature recorded over a period of time. 102

Table 7.2 Comparing ignition timing of the standard and optimized map. 115

Table 7.3 Injection pulse changed from the original map. 116

1

CHAPTER 1

INTRODUCTION.

Unsteady gas flow is always present in intake and exhaust manifolds due to the pulsation

created by the piston movement and exhaust process. If the pulsation is timed correctly it

can be used to improve engine performance at specific engine speeds.

1.1 History of unsteady gas flow.

The original purpose of the manifold was to get the air/fuel mixture into the engine. It was

important on carburettor models to have a good mixture and as little pulsation as possible

through the manifold. Manifolds were also rough on the inside to give a better mixture and

to prevent droplets sticking to the walls. Manifold diameter was also important to get the

correct velocity which kept fuel suspended in the air stream.

The development of fuel injection resulted in a change in manifold design. Engineers

discovered that pulsating flow can be used to force additional air into the engine making it

more efficient. The level of influence is mainly determined by pipe diameter and length.

Long pipes with small diameters were used to optimise efficiency at lower engine speeds

while bigger diameter manifolds with shorter pipes were used for higher engine speeds

(Venter, 1997:171).

On many of the older cars a single pipe was used to supply more than one cylinder with

air. Later, engineers developed manifolds with a common plenum and equal pipe lengths to

utilize pulsating flow to improve volumetric efficiency. Variable intake manifolds were

2

developed to improve engine efficiency over a wider rpm range. This is done by changing

the length of the pipe via a butterfly in the manifold. At low engine speeds the air had to

travel through an elongated pipe to improve low end torque. At higher engine speeds the

butterfly in the manifold opened to shorten the intake path which increased volumetric

efficiency at high engine speeds (Venter, 1997:172).

Resonance manifolds were also developed for flat and vee-engines. These manifolds make

use of the opposed cylinders to create a mass- spring action which will force extra air into

the cylinders. These manifolds are basically a combination of manifolds and offer

smoothness and improved fuel efficiency across the entire speed range. It also improved

power output from low to high engine speeds. Manufacturers such as Toyota, Mazda,

Porsche and many more have used this with great success [Hatamura, K. et al, 1987:6].

1.2 Motivation for investigation on intake manifold design.

Manufacturers have a pre-determined goal with each engine they build. With this in mind

the manufacturer develops a manifold to perform a specific function. Reliability,

driveability, cost-effectiveness, available space and application of the engine are some of

the considerations taken by manufacturers when designing a new engine. Manifolds can be

used to limit engine performance in order to improve reliability of an engine, or to aim at a

specific market which resulted in this study to be conducted.

Exhaust manifolds and systems are usually the first component to be optimized by after-

market systems to improve engine performance. The most obvious explanation by exhaust

manufacturers of these products is that there is a reduction in friction losses, and secondly

3

that the manifold is used to extract exhaust gas from the engine to improve the next intake

cycle (Venter, 1997:124). The same principle is true during the intake cycle, but with the

objective to force extra air into the engine before the intake valve closes.

A model in Microsoft Excel was developed to predict manifold pressure and volumetric

efficiency of an engine. With the model it was possible to change manifold dimensions and

to determine the efficiency without enormous costs involved. The model limits the number

of tests to be done and decreases development costs and time.

1.3 Problem statement.

Pulsation in the intake manifold creates unsteady mass flow which can result in a high or

low mass flow when waves travel through each other. A pressure sensor can be used to

sample the pressure at a specific location in a manifold but it is not able to indicate mass

flow in the manifold and its direction. It is therefore not possible to determine the influence

of different manifolds on volumetric efficiency from data sampled by a pressure sensor.

The purpose of this investigation is to model air flow through the intake manifold with the

objective of increased power output for specific applications.

1.4 Objective of the study.

The objectives of the study were to investigate the:

Influence of manifold dimensions on the volumetric efficiency of an engine.

The engine speed at which maximum power is developed and whether it can be

moved from that of the manufacturers specification.

4

The presence of pulsating flow in the intake manifold.

The use of a simple simulation to determine manifold pressure and determine

volumetric efficiency.

5

CHAPTER 2

LITERATURE STUDY.

A study on pulsating flow in intake manifolds was conducted to develop an

unsophisticated model that can simulate flow in the intake manifold of an internal

combustion engine. An engine consists of many parts that influence each other and these

parts also need consideration in the development of a model. Figure 2.1 shows the major

parts that influence the intake and exhaust process.

Figure 2.1 Major components that influence engine performance.

2.1 Project necessity.

A model was developed to predict the influence of intake manifold dimensions on

volumetric efficiency and to design manifolds for specific applications. The model saves

1

2

3

4

5

6

7

8

9

10

11

12 1. Air box. 2. Air filter. 3. Secondary pipe. 4. Throttle. 5. Plenum. 6. Primary pipes. 7. Intake and exhaust valves. 8. Camshafts. 9. Cylinder bore. 10. Crankshaft. 11. Exhaust manifold. 12. Exhaust system.

6

development time and costs and reduces the number of manifolds to be tested for

production.

2.2 Theory of pulsating flow.

Pulsating flow in engines is generated by piston movement during the intake cycle and by

valve opening during the exhaust cycle. The type of waves that are present in engines are

as follows.

Sound waves occur as plane waves or spherical waves. Plane waves occur in

straight pipes with constant cross sectional diameter whereas spherical waves are

created by an explosion and are equal in all directions radiating from the point of

explosion. A plane wave acts like a coil spring which is compressed on one side,

and can be seen as it travels in the opposite direction.

Periodic waves are repeated at constant time intervals and are similar to sine

waves.

Waves that cause a significant displacement of the supporting gas are called finite

waves which are the type of waves that are generated in engines. Their velocity is

the sum of the velocities of the supporting medium and its acoustic velocity. Gas

particles of compression pressure waves travel in the same direction as the wave

and gas particles of expansion waves travel in the opposite direction to the

expansion wave (Atherton, 1996:95).

2.3 Wave formation in intake and exhaust systems.

In intake systems the downwards motion of the piston creates a change in volume and a

drop in cylinder pressure which leads to a low pressure wave travelling through the

7

primary pipes towards the plenum where it reflects as a high pressure wave due to a sudden

change in area. If the reflected compression wave reaches the valve when the piston is

close to bottom dead centre it will force extra air into the cylinder before the intake valve

closes and will result in a higher volumetric efficiency.

In exhaust systems the rapid opening of the exhaust valve releases gas from the cylinder

creating a compression wave that travels through the exhaust system until it reaches a

junction where it reflects as an expansion wave. If this expansion wave reaches the valve

when the piston is near top dead centre it will create a vacuum in the cylinder resulting in

more burned gases being extracted from the engine before the outlet valve closes and leads

to more fresh air being drawn in by the engine (Dowds, 2003:70).

2.3.1 Distortion of wave profile during flow.

Different points on a wave profile travel at different speeds and in turn will result in a

change of wave shape as it travels through the pipe. A shock wave forms in the pipe if it is

long enough. Figure 2.2 shows a compression wave that travels through a pipe as described

by Annand and Roe (1974). Step 1 shows the completed wave from a typical exhaust

stroke that is formed in the pipe. The peak of the wave will have a higher propagation

velocity than the toe. In step 2 the peak has moved closer to the toe and the distortion in the

wave profile becomes visible. Step 3 shows that the peak has caught up with the toe and

tries to pass it. At this point a shock wave is formed which travels at a new propagation

velocity )( SH . In water it is possible for the peak to pass the toe of the wave but in

compressible flow this is not possible and this is the point where shock waves are formed.

8

The propagation velocity of a compression shock wave is calculated by formula 2.1 (Blair,

1999:167).

)( 17670 GPGa shsh += (2.1)

Where: shP = Pressure ratio between wave pressure and atmospheric pressure.

sh = Shock wave propagation velocity.

Pipe (L)

Flow direction

Distance (x)

P

Step 2

Po

Distance (x)

P

Step 3

Po

Figure 2.2 Distortion of a compression wave as it travels through a pipe (Annand and Roe, 1974:39).

Distance (x)

P

Step 1

Po

9

Blair (1999) uses G as a shorthand notation for various functions of the ratio of specific

heats for air, for example, the value used for G17 will be one seventh and is calculated as

shown below.

( ) 71

4.1214.1

21

17 =

=

=

G

Also ( ) 76

4.1214.1

21

67 =+

=

+=

G

A compression shock wave will travel at a slower propagation velocity than a normal wave

with the same amplitude ratio, but with almost no difference in particle velocity.

Expansion shock waves form at the tail of the wave and have a greater propagation

velocity than normal expansion waves but with no particle velocity right after the wave.

Figure 2.3 shows the formation of an expansion shock wave in a pipe. Step 1 shows the

completed wave of the intake stroke in the pipe. In step 2 the wave has distorted as it

passes through the pipe and step 3 shows how a shock wave is formed at the tail of the

expansion wave.

The propagation velocity of an expansion shock wave is calculated with the use of formula

2.2 (Blair, 1999:169).

)1(0517670 ++= ishish XaGGPGXa (2.2)

Where: 0a = Reference acoustic velocity.

iX = Initial pressure amplitude ratio.

10

Figure 2.3 Distortion of an expansion wave as it travels through a pipe. (Annand and Roe,

1974:39).

Distance (x)

P

Step 3

Po

Distance (x)

P

Step 2

Po

Distance (x)

P

Step 1

Po

Pipe (L)

Flow direction

11

2.4 Friction loss during flow.

Particle flow creates viscous shear forces in the boundary layer close to the pipe wall. This

creates a pressure loss in the wave in the opposite direction of particle motion and results

in internal heating of the gas particles. According to Blasius the friction factor )( fC is

dependent on the Reynolds number and pipe wall roughness and usually in the range of

0.003 to 0.008 which is calculated by the following formula (Blair, 1999:185).

25.0Re0791.0

=Cf (2.3)

Where: Re = Reynolds number.

2.5 Heat transfer during flow.

Heat can be transferred from or to the wall and although conduction, convection and

radiation are involved it is convection that is the most dominant mode in induction

systems. On a red hot exhaust system radiation should also be considered.

2.6 Pressure waves traveling through each other.

Two waves that travel through each other form a new pressure wave and mass flow rate.

This process is referred to as superposition. Pressure transducers in pipes record the

pressure history but do not indicate what happens to mass flow or its direction. It is

therefore possible to record a high pressure during superposition with no mass flow rate,

12

meaning that there will be no improvement in performance. On the other hand it is possible

to record low pressures with high mass flow rate.

Figure 2.4 shows two compression waves that travel towards each other in a pipe. The

waves are shown as square waves to make the explanation easier. The instant the pressure

waves meet a change in pressure and mass flow rate occurs as indicated in Figure 2.5.

Where: P1 = Rightward pressure wave.

P2 = Leftward pressure wave.

Po = Reference pressure.

Where: Ps = Superposition pressure.

Po

P1 P2

A

B

C

D E

F

G

H

Figure 2.5 Two compression waves traveling through each other.

Ps

Wave 2

Wave 1

Po

P1 P2

A

B C

D E

F G

H

Figure 2.4 Two compression waves traveling towards each other.

Wave 2

Wave 1

13

Sonic particle velocity during wave superposition is unlikely even in racing engines where

high pressure exhaust pulses are present (Blair, 1999:181).

2.7 Reflection of pressure waves.

In engine ducting, reflections occur at any change in cross-sectional area such as junctions,

expansions, contraction, taper sections and even at dead ends such as closed valves. A

pressure wave reaching a closed valve will reflect back into the manifold with the same

amplitude and velocity. At the closed end the particle velocity will be zero but for the rest

of the system it will be the same as that of the initial wave. Figure 2.6 shows the reflection

of a wave that acts the same way as an echo.

Where: Pi = Initial pressure wave.

Pr = Reflected pressure wave.

Cs = Particle velocity.

Pi

Cs = 0 m/s

Pr

Figure 2.6 Reflection of a pressure wave at a closed end.

Closed end

14

2.7.1 Reflection of a compression wave at the open end of a pipe.

A high pressure wave reaching the open end of a pipe will expand and reflect as a low

pressure wave that will travel back through the manifold. In exhaust systems the low

pressure wave is used to help empty the cylinder at the end of the exhaust stroke. It is not

likely to reach sonic particle velocity during a reflection but if this happens a weak shock

wave will be formed (Blair, 1999:200).

2.7.2 Reflection of expansion waves at the open end of a pipe.

In expansion waves gas particles flow in the opposite direction to that of the sound wave.

An expansion wave reaching the end of a pipe creates an inflow of gas particles into the

pipe and reflects as a high pressure wave that travels back through the system. If the

reflected high pressure wave is timed correctly it can be used to force extra air into the

cylinder before the intake valve closes. In the case of a normal open ended pipe with a

sharp entry, the inflow of particles creates a distinct vena contracta as shown in Figure 2.7.

Figure 2.7 Reflection of an expansion wave at an open end.

Pi

Pr

Ps > po

Cs < 0

Flow area

15

2.7.3 Flow through bell mouth pipe ends.

Intake systems with bell mouth entries are used to obtain a better airflow into the pipe.

Tests done on a 22.84 mm diameter pipe at a pressure ratio of 1 to 1.3 proved that there is a

20% increase in flow, compared to a plain pipe end. The radius of the entry was changed

from 4 mm to 5 mm to 6 mm and results show that there is approximately 1% to 2%

increase in flow for each millimetre (Blair, 1999:338).

2.7.4 Reflection of pressure waves at sudden area change.

A sudden area change results in a partial reflection of a pressure wave and this reflection is

dependent on the difference in diameters. Compression waves that travel through a sudden

expansion will experience a drop in pressure but continue through the larger diameter pipe

and with a low pressure wave that will reflect back into the first pipe. The opposite

happens when a low pressure wave reaches a sudden expansion. According to Blair

(1999:206), Benson developed a method called constant pressure solution which is very

simple and gives accurate answers. Benson assumed that the superposition pressure at the

plane of the junction was the same in both pipes at the instant of superposition. This simple

Pi

Cs < 0 Pr

Figure 2.8 Reflection at a bell mouth entry for inflow (Blair, 1999:199).

Ps < Po

16

junction model has its limitations, but according to Blair it is remarkably effective in

practice if the area ratio (Ar) is between 6 and 0.1667 (Blair, 1999:206).

661

17

Where: 1L = Selected mesh length of upstream section.

2L = Selected mesh length of the next section.

1d = Average diameter across the length of the first meshed section.

(Between ad and bd )

2d = Average diameter across the length of the next meshed section.

(Between bd and cd )

= Included angle.

Flow separation from the walls takes place when the Mach number is greater than 0.2 and

is increased significantly if the included angle of the tapered pipe is greater than 7 (Blair,

1999:239).

ad bd cd

1d 2d

1L 1L

Figure 2.9 Tapered pipe divided into sections.

Flow direction

18

2.8 Volumetric efficiency.

Various factors such as pulsating flow, manifold length and diameter entry losses, valve

and port geometry and valve timing influence volumetric efficiency.

Volumetric efficiency can be defined as the volume flow rate of air into the intake system

divided by the rate of volume displacement by the piston [Heywood, 1988:53].

2.9 Modelling of flow.

Software programs are available to simulate compressible flow for steady and unsteady

flow. Numerical models are developed with computational fluid dynamics to analyze

problems that involve fluid flow. Makgata from the University of Pretoria used

computational fluid dynamics and one-dimensional gas dynamics, implemented in the

engine simulation code EngMod4T, to improve the inlet system of a high-performance

rally car.

19

CHAPTER 3

MODELLING OF FLOW

A numerical model was developed in Microsoft Excel to predict the effect of pulsating

flow on the volumetric efficiency of an engine and to show that manifold dimensions can

influence the power output of an internal combustion engine. The input and calculation

sheets are shown in Annexure A. Figure 3.1 gives an indication of the calculations

performed by the model to determine volumetric efficiency and manifold pressure.

3.1 Assumptions used in model.

The model was developed to give a prediction of how a manifold will perform during

testing. The main results of the model are manifold pressure and volumetric efficiency

which in the end will show how a manifold will perform. Assumptions made to keep the

model simple influence the level of accuracy and need to be incorporated in the model if

more accurate answers are required. The following assumptions were made.

Pressure pulses from the exhaust cycle have a major influence on back pressure at

the end of the exhaust stroke and influence the fresh air that will enter the engine.

Back pressure was assumed to be constant at all engine speeds.

Cylinder temperature increases as engine speed and load increases. In the Excel

model it was assumed to be the same for all engine speeds.

During valve overlap flow can occur through both intake and exhaust valves. Only

flow through the intake valve was considered in the model.

20

Figure 3.1 Flow diagram of calculations done by the model.

LOOP: Do for each time step until intake cycle is complete.

Minimum cylinder pressure during intake cycle.

Calculate: 1. Propagation velocity of initial wave. 2. Reflected pressure from open end. 3. Reflected pressure waves propagation velocity.

For each time step calculate: 1. Minimum time through meshed section. 2. Crank angle rotation. 3. Piston travel. 4. Volume change. 5. Valve lift. 6. Throat area. 7. Curtain area.

1. Port length. 2. Manifold length 3. Mesh size.

Total intake length. 1. Divide total intake length into equal mesh sections.

1. Cam timing. 2. Max. Valve lift. 3. Valve OD. 4. Stem diameter. 5. Seat width. 6. Seat angle. 7. Bore. 8. Stroke.

1. Minimum flow area. 2. Volume change.

Exhaust back pressure = initial cylinder pressure

Cylinder pressure during time step.

Manifold pressure at valve.

Valve flow coefficient.

Calculate choked flow and possible mass flow to get the minimum flow that will occur.

Initial mass in cylinder is calculated with exhaust back pressure, cylinder temperature and combustion chamber volume.

Calculate: 1. Energy in cylinder. 2. Cylinder temperature after time step. 3. Cylinder pressure after time step.

Total mass in cylinder.

LOOP: Do for each time step until intake cycle is complete.

Calculate pressure at valve after time step.

Volumetric efficiency.

For each meshed section in intake path, Calculate: 1. Pressure and propagation velocity of pulse traveling away from the valve. 2. Pressure and propagation velocity of pulse traveling towards the valve. 3. The super-positioning pressure. 4. The reflected pressure at the open end of the manifold.

Treat intake valve as a closed pipe for compression, combustion and exhaust cycles to calculate reflection of pressure pulses.

Manifold pressure of complete Otto cycle.

21

A constant valve flow coefficient was assumed from minimum to maximum valve

lift. In reality the valve flow coefficient will vary with valve lift (Stone, 1992:247).

Friction losses were not taken into consideration as smooth pipes were used without

bends.

Port length was taken as an average through the centre of the port to the back of the

valve head.

Influence of the secondary pipe after the plenum was not considered due to a large

plenum that was used (Royo, Corbern, and Prez, 1994:6).

Plenum pressure was assumed to be constant in the model. Tests done showed that

a maximum pressure fluctuation of 5 kPa in the plenum occurred and will be

discussed in chapter 4.

Wave reflection at discontinuities in gas properties was not considered. This occurs

when fresh intake air meets gas from the cylinder with different temperature and

properties when the intake valve opens.

The mesh size selected will determine the time step and will create big pressure

differences if a large mesh size is selected.

Pressure through each meshed section is assumed to be linear.

3.2 Information required in model.

The model deals with air flow into the engine and exhaust effects were not included in the

model. Information required in the model includes the intake manifold, cylinder head and

cylinder dimensions and will be discussed in detail in the following section.

22

3.2.1 Intake manifold.

Information required in the model for the manifold is diameter, length and mesh length.

Mesh length is used to divide the manifold and port into equal lengths as indicated in

Figure 3.2 and is determined by the designer. The Excel model can use three different

diameters in one manifold. This is to enable the model to simulate manifolds with

expansions and for taper manifolds. Taper manifolds can be simulated by taking the

average diameter of a certain length and simulate it as a sudden expansion or contraction

(Blair, 1999:237). The more sections the manifold is divided in, the more accurate the

answers will be.

Figure 3.2 Original and simulated manifold showing the meshed sections.

A B A - Port length. B - Manifold length. C - Plenum D - Throttle X - Meshed sections.

C

D

A B

X D

C

Original manifold

Simulation of manifold.

23

3.2.2 Cylinder head.

The intake port diameter and length is required as it forms part of the intake

manifold adding up to the total length of the manifold.

Compression ratio is required to calculate combustion chamber volume to

determine the amount of exhaust gas left in the cylinder.

3.2.3 Intake valve.

Valve outer diameter, seat width and seat angle is required to calculate valve inner

diameter which is used as the throat diameter.

Valve stem diameter is required to calculate throat area.

Valve flow coefficient is required to calculate air flow through the valve. A flow

coefficient of 0.7 was assumed in the model.

Maximum valve lift is used to determine the curtain area and is compared to the

throat area to get the minimum flow area.

Figure 3.3 show the valve dimensions used by the model.

3.2.4 Cam shaft.

Valve timing was required to determine when the intake cycle started and ended.

The cam shaft is divided into 5 sections to simulate valve lift namely up ramp

angle, main lift angle, dwell angle, main lift down and down ramp. The cam shaft

was measured to determine the angles for each section.

24

Figure 3.3 Basic valve dimensions used by model.

3.2.5 Bottom end information.

To calculate cylinder volume the bore and stroke of the engine are required.

Connecting rod length is also required to calculate volume displacement per time

step.

Number of cylinders was used to calculate volumetric efficiency.

3.2.6 General information.

Atmospheric pressure and temperature were used during calculations to determine

wave amplitudes and numerous other values used in the model.

Exhaust back pressure was assumed as it is needed to calculate mass of air present

in the cylinder when the intake cycle started.

Engine speed is required to calculate time step and gas velocities in the engine.

C

A

B D

E

F A Throat area. B Curtain area. C Seat angle. D Seat width. E Port size. F Valve lift.

25

3.2.7 Volumetric efficiency.

In order to calculate volumetric efficiency the mass flow into the engine is required. In the

model the wave propagation velocity is used to calculate the mass entered through the

valve. The theoretical mass of air that will enter the engine is also calculated using the

swept volume and atmospheric conditions. The ratio of the mass flow calculated in the

model is compared to the theoretical flow and will give the volumetric efficiency (Blair,

1999:54).

atm

elv

m

mmod= (3.1)

3.3 Basic procedure of model.

The model was started by assuming that the lowest pressure created by a downward

moving piston would be 60 kPa and that the manifold pressure is constant and equal to that

of the atmosphere. The assumption of 60 kPa was made from the data collected during

tests as manifold pressure rarely dropped below the selected value. Next the reflected pulse

for the initial wave was calculated and the minimum time step (dt) for the two waves to

travel through each other in a meshed section was calculated. The time step was used to

determine crank angle rotation during such a time step in order to get the vacuum created

in the cylinder. A complete intake cycle followed by the compression, combustion and

exhaust cycle was run to create unsteady flow in the intake manifold. The minimum

pressure created by the piston was then used to calculate a new reflected pressure and time

step. A second complete Otto cycle was then simulated to ensure that unsteady flow in the

intake manifold is true. The third complete Otto cycle simulated was recorded to get the

intake manifold pressure and mass flow into the engine.

26

3.4 Mathematical processing.

Most formulae and theory used is based on studies that were done by G.P. Blair.

Initial manifold pressure is assumed to be atmospheric. The manifold is meshed into equal

lengths from the valve up to the plenum. The pressure amplitude ratio for the initial pulse

was calculated from Formula 3.2.

2

1

= )(0p

pX ii (3.2)

Where: iX = Pressure amplitude ratio of initial wave.

ip = Initial pressure wave

0p = Reference pressure

The reflected pressure was calculated by determining the reflected amplitude ratio at the

bell mouth.

6

26

24

2644 )()22(1)1(

GGGXGGXXG

X iiir+++++

= (3.3)

Where: rX = Pressure amplitude ratio of reflected wave.

64 ,GG = Function of the ratio of specific heat as discussed in chapter 2

section 2.3.1.

Propagation velocity for the initial and reflected pressure waves were calculated to

determine the minimum time (dt) for the fastest wave to travel through a meshed section.

Figure 3.4 shows specific points on a pressure waves as it travels through a meshed

section.

27

Figure 3.4 Points on pressure waves before and after time step.

The minimum time through a meshed section was multiplied with a factor of 0.99 to

ensure that pressures calculated at the nodes do not travel past the next node to ensure

interpolation. New pressures at meshed distances were calculated after each time step.

)1( 24160 = XGXGai (3.4)

Where: i = Propagation velocity of initial wave.

0a = Reference acoustic velocity.

Leftward pressure wave

Rightward pressure wave Pr2

Pl1 Pl2

Pr1

Meshed section

Ps Pipe diameter

Mesh length

P

Dist.

P

Pr2

Pl1 Pl2

Pr1

Dist.

Pipe

Time step x.

Time step x+dt.

Rightward pressure wave Leftward

pressure wave

28

1X = Pressure amplitude ratio of rightward wave.

2X = Pressure amplitude ratio of leftward wave.

)1( 14260 = XGXGar (3.5)

Where: r = Propagation velocity of reflected wave.

totalj

jfastest

Ldt=

=

=

1

99.0

(3.6)

The minimum time step )(dt is used to determine the angular displacement )( per time

step.

dt = (3.7)

Where: = Propagation velocity of reflected wave.

A new sheet was drawn up to accommodate a complete intake cycle. The first column was

used for crank angle from TDC and was filled in with time step increments for a complete

cycle. Another column was used to complete the relative intake valve lift for that angle.

Figure 3.5 shows the valve lift of the simulation which was done using 5 stages. The five

stages were the up ramp, main lift, dwell, main down and ramp down.

29

Intake valve lift.

0

2

4

6

8

10

12

0 25 50 75 100 125 150 175 200 225Crank angle (Deg)

Lift

(mm

)

Valve lif t

Figure 3.5 Intake valve lift simulated in the model.

The curtain and throat areas were determined to get the smallest flow area during the

intake cycle. This was used to verify if choked1 flow occurred. Figure 3.6 illustrates the

minimum flow area that was determined by the model.

Intake valve flow area.

0

0.0002

0.0004

0.0006

0.0008

0.001

0.0012

0 25 50 75 100 125 150 175 200 225

Crank angle (Deg)

Area

(sq

.m

)

Flow area

Figure 3.6 Minimum flow area during the intake cycle.

Piston speed was calculated in order to get volume displacement per time step. The volume

displacement together with the minimum flow area will create a vacuum pulse in the intake

system. Formula 3.8 was used to obtain the piston speed (Hannah and Stephens, 1992: 89).

1 Choked flow occurs when the particle velocity attempts to exceeds the local acoustic velocity.

30

+=

nrV p 2

2sinsin (3.8)

Where: pV = Piston velocity.

= angular velocity.

r = crank radius.

= Crank angle from TDC.

n = Ratio of connecting rod length to crank radius.

Bore, stroke and connecting rod length influence the volume displacement per time step.

The volume displacement per time step is dependant on the distance of the piston from

TDC as this influences piston speed. Piston speed per time step is small when the piston is

close to TDC and BDC and reaches a maximum close to halfway through the stroke length.

Figure 3.7 indicate the piston speed between TDC and BDC that was determined by the

model and it is clear that the maximum speed was reached before the piston reached the

middle of the stroke. Formula 3.9 was used to determine the distance from TDC.

+=

nrx

2sin

cos12 (3.9)

Where: x = Distance from TDC.

31

Piston speed.

0

1

2

3

4

5

6

7

0 20 40 60 80 100 120 140 160 180

Crank angle (Deg)

Velo

city

(m

/s)

Figure 3.7 Variation in piston speed from TDC to BDC at 1250 rpm.

The above information was calculated for each time step to form a complete excel sheet.

The following steps were done during each step to determine cylinder pressure and


The assumption was made that the cylinder temperature was 627 degrees Celsius at

the beginning of the intake cycle.

It was assumed that the temperature remains constant during a specific time step

and the pressure after the time step was calculated. The average pressure for that

step was then calculated using the pressure at the beginning and end of the step.

The next step was to determine when choked flow would occur by means of

Formula 3.10 (Heywood, 1988:226).

( )( ) ( )12/1

2/12/1

0

.

12

+

+=

RTpAC

m ord (3.10)

Where: .

m = Mass flow.

dC = Coefficient of discharge.

rA = Area ratio.

op = Reference pressure.

32

oT = Reference temperature.

R = Gas constant.

In order to calculate mass flow the air density and the acoustic velocity inside the

cylinder was calculated. Mass flow was then determined using Formula 3.11

(Blair, 1999:180).

( ) ( )2121005.

51 XXXXAaGm G += (3.11)

Where: A = Pipe cross sectional area.

The calculated mass flow from Formula 3.11 was compared to the choked flow of

Formula 3.9 and the smallest value was used to calculate the mass that entered the

cylinder during the time step using Formula 3.12.

dtmCdM.

= (3.12)

Where: M = Mass of air.

dC = Coefficient of discharge.

.

m = Mass flow.

dt = Selected time step.

The last step was to calculate the cylinder pressure using the total mass in the

cylinder and to calculate a new cylinder temperature to be used for the next step.

After the completion of the intake stroke the compression, combustion and exhaust strokes

were simulated with a closed intake valve. Pressure waves travelling through the intake

would reflect from the intake valve and return to the open side of the manifold where this

33

action would repeat itself. This ensured that unsteady gas flow was present in the intake

manifold when the next Otto cycle started. Two more complete Otto cycles were run to

ensure that pressures were stable of which the last cycle was then recorded.

Figure 3.8 shows the input sheet used by the Excel model developed for the study. The

calculation sheets used in the model is shown in Annexure A while the macros used is

shown in Annexure B.

34

Engine specifications.

Bottom end Metric unit.

Manifold detail Metric unit.

Bore (D) mm. Diameter mm. Stroke mm. Length (incl. port) mm. Conrod length ( l ) mm. Mesh length mm. Number cylinders Cylinder head Atmospheric conditions Compression ratio Port length mm. Atm. Pressure kPa Valve O.D. mm. Temperature C

Stem diameter ( ds ) mm. Exhaust back pressure kPa

Max. valve lift (in) mm. Seat width mm. Seat angle Deg Valve flow coefficient Cam shaft

Inlet side Open AFTER BTDC Up/ down ramp angle Deg Close ABDC Main lift angle Deg Dwell angle Deg Outlet side Open BBDC Close ATDC

Engine speed Rpm. Max engine speed Rpm. Intervals Rpm. Outlet valve Valve O.D. mm. Stem diameter ( ds ) mm. Max. valve lift (out) mm. Seat width mm. Seat angle Deg Valve flow coefficient

Manifold detail Head detail (from valve) Actual throat diameter mm. Throat diameter (de) mm. Throat length mm. Port diameter mm. Remaining Port length mm. Manifold diameter 1 mm. Manifold length 1 mm. Manifold diameter 2 mm. Manifold length 2 mm. Manifold diameter 3 mm. Manifold length 3 mm. Total length mm.

Figure 3.8 The input sheet used in model.

35

CHAPTER 4

EXPERIMENTAL EQUIPMENT AND PROCEDURE

This chapter discusses equipment used, the calibration of the equipment and the test

procedures. The four extreme manifolds with constant diameters are covered in this

chapter. The reason for tests done with constant diameter manifolds was to investigate the

effect of manifold length and diameter on power developed by an engine.

Most important data required from the experiments were the manifold pressures at specific

crank angles which show the pulsation inside the manifold at various engine speeds.

Pulsating flow could influence power output at different engine speeds and the level of

influence depends on the length and diameter of the primary pipes of the manifold. Power

and torque were also recorded as these give a direct indication of the influence of manifold

dimensions on the performance of the engine. This information was compared to that of

the Excel model to determine reliability of the model. A complete discussion of all data

recorded follows.

4.1 Method of investigation

Tests were performed on a two litre Volkswagen AFW production engine in a test cell,

where different intake manifolds were fitted to the engine. The original intake manifold

shown in Figure 4.1 was flanged after the injectors so that the primary pipes could be

changed to obtain different lengths and diameters.2 The injector part of the production

manifold was used for all tests except for the final manifold which was designed for

2 Photos of manifolds tested can be found in Annexure F.

36

maximum power at high engine speeds. This ensured that injector position did not

influence the results. Figure 4.2 shows one of the new manifolds tested. A plenum with a

volume of 0.0045 m3 was build and used for all tests. Plenums with a greater volume than

that of the engine displacement minimize pulsation effects from the secondary pipe which

connects to the air box. (Royo, Corbern, and Prez, 1994:6) To minimize pressure drop to

the plenum, a 60 mm butterfly was used and the conical inserts inside the air box were

removed as is discussed in section 4.6.3. The rest of the engine remained unchanged.

Figure 4.1 Standard production engine with original intake manifold (encircled).

37

Figure 4.2 Engine with one of the modified intake manifolds.

4.2 Equipment

The equipment used for experiments was checked for calibration and that it was in a good

working condition to ensure reliable results. The following sub-sections provide a

discussion of the equipment used.

4.2.1 Engine.

A 2 litre Volkswagen AFW engine was used for the experiment. At sea level it produces

88 kW at 5000 rpm and 178 Nm at 3750 rpm which is approximately 16 % more than what

it would achieve at an altitude of 1580m where tests were done. The loss in power is due to

a drop in atmospheric pressure as less dense air is drawn into the engine. The engine was

connected to a dynamometer by a custom made driveshaft with rubber couplings and was

aligned within 0.1mm axial and radial. The engine was cooled by a heat exchanger and

control valve which controlled the coolant temperature between 70 and 80C.

38

A complete service was done before any testing started. The rotor, distributor cap, filters,

plug wires and plugs were also replaced to ensure that the engine was in good condition. A

compression test was done on all cylinders to ensure that cylinder pressure was within

specification. Annexure C shows the pressure recorded. Ignition timing was adjusted to

manufacturers specification with compensation for altitude and the fuel injectors were

calibrated and checked for spray pattern and leakage. Annexure D shows the injector flows

recorded on the test bench.

4.2.2 Fuel used.

All tests were done with a 93 octane unleaded reference fuel from the Natref laboratory.

This ensured that the base composition was up to standard and the same for all tests.

4.2.3 Manifolds.

The original manifold was cut and flanged upstream of the injectors to ensure that injector

position did not influence test results. An intermediate section was manufactured to space

out the two middle pipes to get a standard flange from where the new manifolds could be

fitted. Manifolds with different diameters and lengths were fitted to the intermediate

section and tested. Two types of manifolds were considered in this study namely constant

diameter and dual diameter. The reason for testing dual diameter manifolds was to create

an expansion in the pipe to get partial reflection of waves.

39

All tests were done with straight pipes to eliminate effects associated with bends in

manifolds.

Figure 4.3 Pressure sensor was installed on the intake manifold behind the injector.

4.2.4 Test cell.

Air temperature inside the test cell was controlled at 20C with a maximum variation of

2C. During tests an extraction fan was used to extract exhaust gases. A CO-analyzer was

fitted to ensure that air inside the room remained at safe limits. Barometric pressure inside

the test cell was recorded to ensure that data were reliable when tests are compared as

barometric pressure influence power output. Barometric pressure recorded inside the test

cell was between 85 kPa and 87 kPa during experiments.

4.2.5 Control room.

Data from the test cell was sent to a control room via a PLC system and was recorded.

Data such as power, speed, torque, air-fuel ratio, throttle position, water and oil

temperature, inlet temperature, fuel pressure and flow, atmospheric pressure and exhaust

40

manifold temperatures were recorded. A separate computer was installed to record intake

manifold pressure, TDC and crank angle using an Eagle PC 30F data acquisition card.

4.2.6 Dynamometer.

A Shenck W150 eddy current dynamometer was used during experiments. A stain gauge

load cell was installed on the dynamo at a known distance which made it possible to record

torque. Inside the control room a Shenck X-act controller was used to control engine speed

at full throttle conditions.

4.2.7 Optical crank angle marker.

The AVL optical crank angle marker shown in figure 4.4 was used to record top dead

centre and 5 degree crank angle interval of engine rotation. The TDC signal was used to

trigger the manifold pressure sensor to start sampling.

Figure 4.4 Position of crank angle marker as installed on the crankshaft.

41

4.2.8 Manifold pressure sensor.

Manifold pressure was recorded by means of a Kistler model 4075A Piezoresistive

pressure sensor which was installed behind the injector as indicated by Figure 4.3. This

was the closest possible position to the valve and was 140 mm from the valve. The sensor

signal was send to a Kistler model 4603B Piezoresistive amplifier which was connected to

a separate computer system to record samples taken. Sampling in burst mode was triggered

by the crank angle marker. In burst mode the TDC, crank angle and manifold pressure is

sampled after each other at a pre selected frequency until the selected amount of samples

are recorded.

Figure 4.5 indicates the connection between the angle marker, pressure sensor installed on

the intake manifold and the computer.

Com Trig

Junction box

Pressure sensor

Amplifier

Optical crank angle marker

Pulse multiplier

Computer

Figure 4.5 Schematic diagram of information recorded by the computer.

42

4.3 Calibration of equipment.

All equipment used was calibrated at a SANAS approved laboratory (South African

National Accreditation System).

The dynamometer calibration was done by connecting two arms of equal length on

both sides of the dynamometer. The read out from the PLC to the computer was

then set to zero. The next step was to place weights on the hanger on the side were

the load cell was connected. Torque were calculated using the known length from

the centre of the dynamometer to the hanger and was compared to the readout of

the load cell on the Shenck X-act controller.

Fuel flow was recorded via a Micro motion mass flow meter. It was calibrated by

simulating a 4 to 20 milliamps signal to the computer through a Hart

Communicator.

Temperature calibration was done by using a FLUKE 744 meter. Low, high and

intermediate temperatures were simulated by the meter to calibrate the PLC.

Random temperatures were then selected to verify calibration. This was done for

intake, exhaust, water, oil and fuel thermocouples. Thermocouples were supplied

with certificates from the manufacturer ensuring their accuracy.

A Druck DPI 610 pressure simulator was used for fuel and manifold pressure

calibration. Fuel pressure calibration was done up to 1000 kPa and plenum pressure

from full vacuum to 200 kPa.

43

The Kistler Piezoresistive pressure sensor used in the primary pipe was checked

from full vacuum to 600 kPa for calibration. A FLUKE 744 meter was used to

check the voltage output from the amplifier while the pressure sensor was

connected to the amplifier and Druck DPI 610 meter. Figure 4.6 the connection

between the equipment and Annexure G shows the calibration data that were

recorded..

Figure 4.6 Kistler pressure sensor calibration.

The 0 to 5 volt signals from the crank angle marker was checked by connecting an

oscilloscope to the pulse multiplier while the engine was idling.

Calibration of the injectors was done on an ASNU ultrasonic injection cleaner and

flow meter. A leak test was done at 300 kPa to check for proper sealing. The tester

made it possible to simulate low to high engine speeds to check spray patterns and

flow. A constant flow test was also done to check that all injectors delivered the

same flow. Calibration data of the injectors is available in annexure D.

Kistler pressure sensor

Kistler Amplifier

Druck DPI 610

FLUKE 744 meter

44

4.4 Variation in data recorded.

The variation in data sampled was analysed to ensure that data samples were reliable and

consistent. In Annexure E the variation in engine speed, torque, exhaust temperature and

excess-air ratio at constant intervals is discussed. It was concluded that all tests showed

small variations at constant engine speeds.

4.5 Test procedure.

The first test was done with the standard manifold already flanged to be used with other

pipe lengths. Figure 4.7 show the original manifold that was modified with a flange fitted

upstream of the injectors. The first test was done with a standard intake system as indicated

in Figure 4.8. This power curve was used as the reference graph. For other tests the conical

inserts in the air box were removed to minimize pressure drop and the standard 56 mm

butterfly was replaced by a 60 mm butterfly. Manifolds with different diameters and

lengths were tested and the data recorded was compared to the standard system and that of

the model. The final test performed was the same as the first to ascertain if there were any

changes in engine power. A change in engine power would have resulted in data being not

trustworthy.

45

Figure 4.7 Production manifold flanged and pressure sensor housing welded in position.

Figure 4.8 Engine with original intake system.

Testing the manifold only started when the engine oil reached a temperature of 65C to

ensure consistency of tests. The dynamometer was then set to control engine speed at 1250

46

rpm at full throttle condition when the oil reached the correct temperature. The engine was

then given time to stabilise its speed before the two computers were manually set to log

data. The above procedure was repeated in increments of 250 rpm until the test was

complete. Engine speed was then decreased slowly up to idle speed and was kept there

until exhaust temperature was down to at least 450C. This was done to prevent damage to

the intake pressure sensor as it was situated above the exhaust manifold.

4.6 Data recorded and sample rates.

Two computers were used during testing to record data. One of the computers was

connected to the PLC used to record all low frequency data from the engine while the other

computer fitted with the Eagle card recorded high frequency pressure data such as

manifold pressure, TDC and crank angle.

4.6.1 Low frequency samples.

Engine power, pressures and temperatures do not change considerably within a half second

and for this reason a low sample frequency was selected. The abovementioned data were

recorded at a rate of 2 samples per second and a total of 60 samples were recorded after the

engine stabilised.

The following low frequency data was recorded for each test.

Power and torque to indicate the influence of different intake manifolds on


47

Engine speed, to verify data and to check for speed fluctuations as it directly affects

power output.

Exhaust gas temperature was measured inside the outlet manifold close to the head.

A rise in exhaust temperature between different manifolds could indicate a lean

mixture meaning that more air was induced in the engine as the engine had a fixed

fuel map. The electronic control unit on the engine receives no feedback from the

lambda sensor and does not adjust fuel flow if more air is induced into the engine,

thus varying the air/fuel ratio.

The fuel flow rate and pressure.

Atmospheric pressure in the test cell.

Inlet temperature, as this will influence pulsation velocity in the manifold.

Excess air ratio, using a Bosch Lambda sensor and an ETAS LA2 Lambda meter.

Excess air ratio indicates whether the engine runs lean or rich with different

manifolds as the engines ECU does not use a closed loop.

Throttle position, to ensure that it was fully open during tests.

Engine oil and water temperature, to verify that the engine did not reach high

temperatures, thus influencing performance.

4.6.2 High frequency samples.

Manifold pressure, top dead centre and crank angle were sampled at high frequencies in

burst mode as per Table 4.1. The sample frequency did not increase in equal increments as

these were preset on the PC30F card used. At each interval 10000 samples were taken over

3 channels. Recording was triggered by the TDC signal from where the sampling was done

48

until all samples were collected. Between 20 and 25 complete Otto-cycles of data were

available to compare. Taking more than 10000 samples would produce very large files.

Table 4.1 Sample frequencies used at different engine speeds.

Engine speed Sample frequency (rpm) (Hz) 1500 20000 2000 25000 2500 30000 3000 40000 3500 45000 4000 50000 4500 56000 5000 63000 5500 70000

49

CHAPTER 5

RESULTS OBTAINED FROM EXPERIMENTS.

Fourteen manifolds with different diameter and length combinations were tested. Table 4.2

shows how the different manifolds performed. Five manifolds was selected to be discussed

in this Chapter.

5.1 Bracket tests.

The first and last tests were done with a standard intake system to check for repeatability.

These two tests were called bracket tests and were used as reference before and after the

completion of the tests. Tables 5.2 and 5.3 give a comparison of the power and torque of

the two bracket tests. Maximum variation in power was 0.8 kW which occurred at 3750

rpm and 4250 rpm. An average variation in power output of 1.07 % was recorded between

the two bracket tests. Such a small percentage in variation of power shows that there was a

negligible change in power output of the engine. Maximum torque variation occurred at

3000 rpm and was 2.3 Nm while the average variation recorded was found to be 1.08 %.

This shows that engine characteristics did not change during tests and that data recorded

was reliable and could be analysed further.

50

Table 5.1 Performance of manifolds tested.

Manifold Detail Length Diameter Max. Power @

rpm. Max. Torque @

rpm. (mm.) (mm.) (kW @ rpm.) (Nm @ rpm.)

Original Part 1 215 34.9 67.5 @ 5000 142.6 @ 3750 Taper 235 34.9-60

Total length 450

Manifold A Part 1 300 34.4 69.9 @ 5000 141.5 @ 4250

Manifold B Part 1 1280 34.4 53.5 @ 3500 155.4 @ 3250

Manifold C Part 1 700 34.4 65.2 @ 4250 155.1 @3500

Manifold D Part 1 500 34.4 68.5 @ 4750 149.7 @ 3750

Manifold E Part 1 520 34.4 68.3 @ 4500 152.7 @ 3750

Manifold F Part 1 440 34.4 69.4 @ 4750 147.7 @ 3500

Manifold G Part 1 1060 25.4 47.1 @ 3250 146.0 @ 2500

Manifold H Part 1 420 34.4 67.4 @ 4500 151.2 @ 3500 Part 2 180 44

Total length 600

Manifold I Part 1 300 34.4 70.3 @ 5000 146.2 @ 3750 Part 2 180 44

Total length 480

Manifold J Part 1 460 34.4 68.3 @ 4750 154.7 @ 3500 Part 2 180 44

Total length 640

Manifold K Part 1 300 34.4 66.3 @ 4750 151.4 @ 3250 Part 2 160 44

Part 3 160 34.4

Total length 620

Manifold L Part 1 380 40 70.8 @ 5500 142.6 @4500

Manifold M Part 1 280 40 70.8 @ 5750 140.4 @ 3500

Manifold N Part 1 260 40 72.0 @ 5000 144.1 2 4000 Part 2 160 44

Total length 420

51

0

10

20

30

40

50

60

70

80

1000 2000 3000 4000 5000 6000

Engine speed (rpm)

Pow

er

(kW)

Bracket 1Bracket 2

Figure 5.1 Power versus engine speed for bracket tests.

Table 5.2 Difference in power output of bracket tests.

Engine speed Mean power Mean power Variation (rpm) (kW) (kW) %

Bracket 2 Bracket 2 1250 15.1 15.3 1.325 1500 18.2 18.5 1.648 1750 20.9 21 0.478 2000 26.8 27.2 1.493 2250 32.6 33.1 1.534 2500 36.6 37.1 1.366 2750 40.1 40.7 1.496 3000 43 43.7 1.628 3250 46.6 47.3 1.502 3500 51.5 52.1 1.165 3750 55.3 56.1 1.447 4000 58.5 59.2 1.197 4250 62.8 63.6 1.274 4500 65.7 66.1 0.609 4750 67.1 67.5 0.596 5000 67.4 67.5 0.148 5250 65.3 65.3 0 5500 61.6 61.9 0.487

52

80

90

100

110

120

130

140

150

1000 2000 3000 4000 5000 6000

Engine speed (rpm)

Torq

ue

(Nm

)

Bracket 1Bracket 2

Figure 5.2 Torque versus engine speed for bracket tests.

Table 5.3 Difference in torque developed during bracket tests.

Engine speed Mean

Torque Mean

Torque Variation (rpm) (Nm) (Nm) %

Bracket 1 Bracket 2 1250 114.9 117.1 1.915 1500 115.9 117.6 1.467 1750 114 114.5 0.439 2000 127.6 129.4 1.411 2250 138.1 140.1 1.448 2500 139.8 141.8 1.431 2750 139.3 141.1 1.292 3000 136.9 139.2 1.68 3250 136.6 138.6 1.464 3500 140.3 142 1.212 3750 140.6 142.6 1.422 4000 139.7 141.3 1.145 4250 140.9 142.5 1.136 4500 139.1 140.1 0.719 4750 134.6 135.5 0.669 5000 128.6 128.8 0.156 5250 118.8 118.7 0.084 5500 106.8 107.2 0.375

53

5.2 Statistical evaluation.

Engine power and torque were recorded at a sample rate of 2 hertz for a period of 30

seconds at a controlled engine speed. Recording started at 1250 rpm and was increased in

increments of 250 rpm until maximum engine speed was reached. The average coefficient

of variance in power during the first bracket3 test was 0.35 % and 0.41% for the last

bracket test. Spot checks were done on other tests which also indicated that the average

coefficient of variance was in the same region. A manifold that developed more or less

power than that of the mean of the original manifold plus the average coefficient of

variance could be considered to have influenced the power output of the engine. Tables 5.4

and 5.5 give the standard deviation and coefficient of variance in power and torque for the

bracket tests done.

Table 5.4 Variation in torque and power of the first bracket test. Engine speed

Mean Torque

Standard deviation

Coefficient of variance

Mean power

Standard deviation


(rpm) (Nm) (Nm) (%) (kW) (kW) (%) 1250 114.9 0.69 0.6 15.1 0.1 0.69 1500 115.9 0.41 0.35 18.2 0.08 0.42 1750 114 0.57 0.5 20.9 0.12 0.57 2000 127.6 0.45 0.35 26.8 0.12 0.43 2250 138.1 0.65 0.47 32.6 0.16 0.5 2500 139.8 0.26 0.18 36.6 0.08 0.21 2750 139.3 0.19 0.14 40.1 0.06 0.16 3000 136.9 0.26 0.19 43 0.09 0.2 3250 136.6 0.25 0.18 46.6 0.09 0.19 3500 140.3 0.24 0.17 51.5 0.1 0.19 3750 140.6 0.36 0.26 55.3 0.15 0.26 4000 139.7 0.36 0.26 58.5 0.17 0.29 4250 140.9 0.25 0.18 62.8 0.12 0.2 4500 139.1 0.25 0.18 65.7 0.13 0.19 4750 134.6 0.29 0.22 67.1 0.15 0.22 5000 128.6 0.36 0.28 67.4 0.2 0.29 5250 118.8 0.43 0.36 65.3 0.25 0.38 5500 106.8 0.96 0.9 61.6 0.56 0.91

3 To obtain the output of the engine used for the experiments. After completion of all the experiments the

engine was restored to its original state to determine if there was a change in output.

54

Table 5.5 Variation in torque and power of the second bracket test. Engine speed

Mean Torque

Standard deviation


Mean power

Standard deviation


(rpm) (Nm) (Nm) (%) (kW) (kW) (%) 1250 117.1 0.32 0.27 15.3 0.06 0.39 1500 117.6 0.38 0.32 18.5 0.08 0.43 1750 114.5 0.69 0.6 21 0.14 0.67 2000 129.4 0.57 0.44 27.2 0.14 0.51 2250 140.1 0.42 0.3 33.1 0.11 0.33 2500 141.8 0.19 0.13 37.1 0.06 0.16 2750 141.1 0.24 0.17 40.7 0.07 0.17 3000 139.2 0.21 0.15 43.7 0.08 0.18 3250 138.6 0.31 0.22 47.3 0.12 0.25 3500 142 0.27 0.19 52.1 0.11 0.21 3750 142.6 0.42 0.29 56.1 0.17 0.3 4000 141.3 0.3 0.21 59.2 0.14 0.24 4250 142.5 0.87 0.61 63.6 0.39 0.61 4500 140.1 0.81 0.58 66.1 0.39 0.59 4750 135.5 0.44 0.32 67.5 0.22 0.33 5000 128.8 0.55 0.43 67.5 0.3 0.44 5250 118.7 0.57 0.48 65.3 0.32 0.49 5500 107.2 1.12 1.04 61.9 0.66 1.07

5.3 Removal of cones in air box.

The cones on the air box shown in Figure 5.5 were removed to minimize pressure drop in

the box. These cones are designed to minimize breathing noise and to have a minimum

effect on power output. Figure 4.12 shows that there was no gain in power from 1250 rpm

up to 5000 rpm with a 1% gain in the last 500 rpm.

55

Figure 5.3 The original air box with cones installed.

5.4 Fitment of a larger butterfly valve.

The standard 56 mm butterfly was replaced by a 60 mm butterfly to minimize pressure

losses between the air box and plenum. A new test was carried out with the cones in the air

box removed and the 60 mm butterfly installed. The test shows no power gain up to 5000

rpm and a 2.5% increase in power between 5000 rpm and 5500 rpm. Figure 5.4 show the

power output of a standard system compared to a system with the cones removed and a

system with the cones removed and a 60 mm butterfly fitted.

56

0

10

20

30

40

50

60

70

1000 2000 3000 4000 5000 6000

Engine speed (rpm)

Pow

er (kW

)StandardNo cones60 mm butterfly

Figure 5.4 Power versus engine speed for standard system, cones in air box removed and 60 mm butterfly with cones removed.

5.5 Plenum used for tests.

A plenum with a volume of 4.5 litres was used on all manifolds tested. Three tests were

done to measure pulsation in the plenum. The pressure sensor was installed in different

positions to check if pulsation varied inside the plenum. The first position was between the

primary pipes of number 1 and 2 cylinder. For the second tests it was installed next to the

butterfly valve and for the final test it was installed on the opposite side of the butterfly as

shown in Figure 5.5. Pressure fluctuations in all three positions were the same with a

maximum variation of 10 kPa at 5500 rpm. This gives a fluctuation of 5 kPa from the

average plenum pressure. Figure 5.6 indicates the pressure recorded in the three positions.

57

(a) (b) (c)

Figure 5.5 Testing for pressure variation in plenum. (a) sensor installed between piston 1 and 2, (b) next to the butterfly and (c) on the opposite side of the butterfly.

868890

92949698

100102104

0 100 200 300 400 500 600 700

Angle (Deg)

Pres

sure

(kP

a)

Position 1Position 2Position 3

Figure 5.6 Pressure versus crank angle in plenum with the sensor installed in different

positions.

5.6 Manifolds selected for discussion.

Four manifolds with constant diameters and different lengths were selected for discussion

in the following sections. Figure 5.7 compares power output of the standard manifold and

the four manifolds selected for discussion. Data on other manifolds that were tested are

available in Annexure F.

58

0

10

20

30

40

50

60

70

80

1000 2000 3000 4000 5000 6000

Engine speed (rpm)

Pow

er (kW

)

StandardManifold AManifold BManifold GManifold M

Figure 5.7 Power versus engine speed for various manifolds.

5.7 Manifold A

Figure 5.8 Manifold A: The shortest manifold tested with a 34.9 mm diameter. The arrow indicates the position of the pressure sensor.

The manifold tested in this section was the shortest possible manifold which could be

manufactured with the original manifolds injector part. The internal diameter of manifold

59

A was 34.4 mm with a total length of 200 mm. Since it was not possible to open the engine

to measure port length, it was estimated to be 100 mm using a wire that was bent to follow

the centre line of the port up to the back. The total length from the valve to the plenum was

300mm which was 150 mm shorter than the original manifold. The pressure sensor was

installed 140 mm from the valve on the original manifold as shown in Figure 5.8 and

remained in this position for all tests. Figure 5.9 shows the ram tubes that were used in the

plenum to minimize entry losses.

Figure 5.9 Ram tubes installed to minimize entry losses.

Figure 5.10 shows the pressure variation inside the manifold for one complete Otto cycle

from 1500 rpm to 5500 rpm. The green line indicates when the inlet valve opened (IVO)

while the red line indicates valve closure (IVC). The pressure pulses in the intake system

change in amplitude and frequency as engine speed varies and continue to travel up and

down the pipes after the inlet valve closed.

60

1500

2500

3500

450055

00

IVCIVO BDC

60000

70000

80000

90000

100000

110000

0 100 200 300 400 500 600 700Crank angle (Deg)

Pres

sure

(P

a)

1500 rpm2500 rpm3500 rpm4500 rpm5500 rpm

Figure 5.10 Pressure versus crank angle for manifold A at different engine speeds.

5.7.1 Torque developed with manifold A.

Data collected for this test showed the standard deviation on torque was less than 1 Nm

except for 1250 rpm where it reached 1.36 Nm. Manifold A showed a definite loss of

torque from 2000 rpm up to 2750 rpm compared to the original intake. Average loss over

this range is 6.3%. A gain of 6% in torque was visible from 5000 rpm up to 5500 rpm with

a maximum gain of 9.4% at 5500 rpm. Manifold A is shorter than the original one,

therefore the pressure pulses will reach the intake valve in a very short time at low engine

speeds and will thus not force more air into the engine. The shorter manifold will improve

breathing at high engine speeds as the pressure pulses reach the intake valve just before the

valve closes forcing more air into the engine. Figure 5.11 shows the comparison between

the torque developed by manifold A and the original manifold.

61

Table 5.6 Comparison of torque developed between manifold A and the original manifold.

Standard manifold Manifold A Speed Torque Torque Variance (rpm) (Nm) (Nm) (Nm) 1250 117.1 111.7 -5.4 1500 117.6 116.3 -1.3 1750 114.5 120.5 6 2000 129.4 120.6 -8.8 2250 140.1 128.1 -12 2500 141.8 133.6 -8.2 2750 141.1 136.6 -4.5 3000 139.2 137.7 -1.5 325

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