Liquid piston engines 9781119323273, 1119323274, 978-1-119-32295-5

Table of contents :
Content: Cover
Title Page
Copyright Page
Contents
Abstract
List of Symbols
1 Introduction
1.1 Background
1.2 Types of Stirling Engines
1.3 Stirling Engine Designs
1.4 Free-Piston Stirling Engines
1.5 Gamma Type Engine
References and Bibliography
2 Liquid Piston Engines
2.1 Introduction
2.2 Objectives
2.3 Brief Overview of Pumps and Heat Engines
2.4 Heat Engine
2.5 Clever Pumps
2.6 History and Development of Stirling Engines
2.7 Operation of a Stirling Engine
2.8 Working Gas
2.9 Pros and Cons of Stirling Engine
2.10 Low Temperature Difference Stirling Engine 2.11 Basic Principle of a Fluidyne2.12 Detailed Working of a Fluidyne
2.13 Role of Evaporation
2.14 Regenerator
2.15 Pumping Setups
2.16 Tuning of Liquid Column
2.17 Motion Analysis
2.18 Losses
2.19 Factors Affecting Amplitude
2.20 Performance of Engine
2.21 Design
2.22 Assembly
2.23 Calculation
2.24 Experiments
2.25 Results
2.26 Comparison Within Existing Commercial Devices
2.27 Improvements
2.28 Future Scope
2.29 Conclusion
2.30 Numerical Analysis
References and Bibliography
3 Customer Satisfaction Issues
3.1 Durability Issues
3.2 Testing of Engines 3.3 Design of Systems3.4 Systems Durability
References and Bibliography
4 Lubrication Dynamics
4.1 Background
4.2 Friction Features
4.3 Effects of Varying Speeds and Loads
4.4 Friction Reduction
4.5 Piston-Assembly Dynamics
4.6 Reynolds Equation for Lubrication Oil Pressure
4.7 Introduction
4.8 Background
4.9 Occurrence of Piston Slap Events
4.10 Literature Review
4.11 Piston Motion Simulation Using COMSOL
4.12 Force Analysis
4.13 Effects of Various Skirt Design Parameters
4.14 Numerical Model of Slapping Motion
4.15 Piston Side Thrust Force
4.16 Frictional Forces 4.17 Determination of System Mobility4.18 Conclusion
5 NVH Features of Engines
5.1 Background
5.2 Acoustics Overview of Internal Combustion Engine
5.3 Imperial Formulation to Determine Noise Emitted from Engine
5.4 Engine Noise Sources
5.5 Noise Source Identification Techniques
5.6 Summary
References and Bibliography
6 Diagnosis Methodology for Diesel Engines
6.1 Introduction
6.2 Power Spectral Density Function
6.3 Time Frequency Analysis
6.4 Wavelet Analysis
6.5 Conclusion
References and Bibliography
7 Sources of Noise in Diesel Engines
7.1 Introduction
7.2 Combustion Noise 7.3 Piston Assembly Noise7.4 Valve Train Noise
7.5 Gear Train Noise
7.6 Crank Train and Engine Block Vibrations
7.7 Aerodynamic Noise
7.8 Bearing Noise
7.9 Timing Belt and Chain Noise
7.10 Summary
References and Bibliography
8 Combustion Based Noise
8.1 Introduction
8.2 Background of Combustion Process in Diesel Engines
8.3 Combustion Phase Analysis
8.4 Combustion Based Engine Noise
8.5 Factors Affecting Combustion Noise
8.6 In Cylinder Pressure Analysis
8.7 Effects of Heat Release Rate
8.8 Effects of Cyclic Variations
8.9 Resonance Phenomenon

Citation preview

Liquid Piston Engines

Scrivener Publishing 100 Cummings Center, Suite 541J Beverly, MA 01915-6106 Publishers at Scrivener Martin Scrivener ([email protected]) Phillip Carmical ([email protected])

Liquid Piston Engines

Aman Gupta, Shubham Sharma, and Sunny Narayan

This edition first published 2017 by John Wiley & Sons, Inc., 111 River Street, Hoboken, NJ 07030, USA and Scrivener Publishing LLC, 100 Cummings Center, Suite 541J, Beverly, MA 01915, USA © 2017 Scrivener Publishing LLC For more information about Scrivener publications please visit www.scrivenerpublishing.com. All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, except as permitted by law. Advice on how to obtain permission to reuse material from this title is available at http://www.wiley.com/go/permissions.

Wiley Global Headquarters 111 River Street, Hoboken, NJ 07030, USA For details of our global editorial offices, customer services, and more information about Wiley products visit us at www.wiley.com. Limit of Liability/Disclaimer of Warranty While the publisher and authors have used their best efforts in preparing this work, they make no representations or warranties with respect to the accuracy or completeness of the contents of this work and specifically disclaim all warranties, including without limitation any implied warranties of merchantability or fitness for a particular purpose. No warranty may be created or extended by sales representatives, written sales materials, or promotional statements for this work. The fact that an organization, website, or product is referred to in this work as a citation and/or potential source of further information does not mean that the publisher and authors endorse the information or services the organization, website, or product may provide or recommendations it may make. This work is sold with the understanding that the publisher is not engaged in rendering professional services. The advice and strategies contained herein may not be suitable for your situation. You should consult with a specialist where appropriate. Neither the publisher nor authors shall be liable for any loss of profit or any other commercial damages, including but not limited to special, incidental, consequential, or other damages. Further, readers should be aware that websites listed in this work may have changed or disappeared between when this work was written and when it is read. Library of Congress Cataloging-in-Publication Data ISBN 978-1-119-32295-5

Cover images: Irrigation, Elswarro | Dreamstime.com. Plant, lofoto | Dreamstime.com Chimneys, Piotr Majka | Dreamstime.com Cover design by Kris Hackerott Set in size of 11pt and Minion Pro by Exeter Premedia Services Private Ltd., Chennai, India

Printed in the USA 10 9 8 7 6 5 4 3 2 1

Contents Abstract

ix

List of Symbols

xi

1

1 1 2 4 6 18 27

Introduction 1.1 Background 1.2 Types of Stirling Engines 1.3 Stirling Engine Designs 1.4 Free-Piston Stirling Engines 1.5 Gamma Type Engine References and Bibliography

2 Liquid Piston Engines 2.1 Introduction 2.2 Objectives 2.3 Brief Overview of Pumps and Heat Engines 2.4 Heat Engine 2.5 Clever Pumps 2.6 History and Development of Stirling Engines 2.7 Operation of a Stirling Engine 2.8 Working Gas 2.9 Pros and Cons of Stirling Engine 2.10 Low Temperature Difference Stirling Engine 2.11 Basic Principle of a Fluidyne 2.12 Detailed Working of a Fluidyne 2.13 Role of Evaporation 2.14 Regenerator 2.15 Pumping Setups 2.16 Tuning of Liquid Column 2.17 Motion Analysis 2.18 Losses 2.19 Factors Affecting Amplitude 2.20 Performance of Engine

v

29 29 32 33 38 42 45 48 53 53 54 55 57 61 61 62 63 64 65 66 67

vi

Contents 2.21 Design 2.22 Assembly 2.23 Calculation 2.24 Experiments 2.25 Results 2.26 Comparison Within Existing Commercial Devices 2.27 Improvements 2.28 Future Scope 2.29 Conclusion 2.30 Numerical Analysis References and Bibliography

3 Customer Satisfaction Issues 3.1 Durability Issues 3.2 Testing of Engines 3.3 Design of Systems 3.4 Systems Durability References and Bibliography 4 Lubrication Dynamics 4.1 Background 4.2 Friction Features 4.3 Effects of Varying Speeds and Loads 4.4 Friction Reduction 4.5 Piston-Assembly Dynamics 4.6 Reynolds Equation for Lubrication Oil Pressure 4.7 Introduction 4.8 Background 4.9 Occurrence of Piston Slap Events 4.10 Literature Review 4.11 Piston Motion Simulation Using COMSOL 4.12 Force Analysis 4.13 Effects of Various Skirt Design Parameters 4.14 Numerical Model of Slapping Motion 4.15 Piston Side Thrust Force 4.16 Frictional Forces 4.17 Determination of System Mobility 4.18 Conclusion

67 70 71 72 74 76 78 79 80 80 83 87 87 88 88 89 89 91 91 93 94 94 95 96 102 104 105 110 114 117 120 131 132 133 133 143

Contents vii 5

NVH Features of Engines 5.1 Background 5.2 Acoustics Overview of Internal Combustion Engine 5.3 Imperial Formulation to Determine Noise Emitted from Engine 5.4 Engine Noise Sources 5.5 Noise Source Identification Techniques 5.6 Summary References and Bibliography

145 145 146

6

Diagnosis Methodology for Diesel Engines 6.1 Introduction 6.2 Power Spectral Density Function 6.3 Time Frequency Analysis 6.4 Wavelet Analysis 6.5 Conclusion References and Bibliography

161 161 162 162 163 164 165

7

Sources of Noise in Diesel Engines 7.1 Introduction 7.2 Combustion Noise 7.3 Piston Assembly Noise 7.4 Valve Train Noise 7.5 Gear Train Noise 7.6 Crank Train and Engine Block Vibrations 7.7 Aerodynamic Noise 7.8 Bearing Noise 7.9 Timing Belt and Chain Noise 7.10 Summary References and Bibliography

167 167 168 168 170 170 171 171 171 172 174 175

8 Combustion Based Noise 8.1 Introduction 8.2 Background of Combustion Process  in Diesel Engines 8.3 Combustion Phase Analysis 8.4 Combustion Based Engine Noise 8.5 Factors Affecting Combustion Noise 8.6 In Cylinder Pressure Analysis 8.7 Effects of Heat Release Rate 8.8 Effects of Cyclic Variations 8.9 Resonance Phenomenon

149 151 154 157 158

179 179 180 183 184 186 187 187 188 189

viii

Contents 8.10 In Cylinder Pressure Decomposition Method 8.11 Mathematical Model of Generation of Combustion Noise 8.12 Evaluation of Combustion Noise Methods 8.13 Summary References and Bibliography

9

Effects of Turbo Charging in S.I. Engines 9.1 Abstract 9.2 Fundamentals 9.3 Turbochargers 9.4 Turbocharging in Diesel Engines 9.5 Turbocharging of Gasoline Engines 9.6 Turbocharging 9.7 Components of Turbocharged SI Engines 9.8 Intercooler 9.9 Designing of Turbocharger 9.10 Operational Problems in Turbocharging of SI Engines 9.11 Methods to Reduce Knock in S.I Engines 9.12 Ignition Timing and Knock 9.13 Charge Air Cooling 9.14 Downsizing of SI Engines 9.15 Techniques Associated with Turbo Charging of SI Engines Boosting Systems

189 192 193 199 199 203 203 204 205 206 207 208 208 213 213 222 223 223 224 225 225

10 Emissions Control by Turbo Charged SI Engines

231

11 Scope of Turbo Charging in SI Engines

233

12 Summary

235

13 Conclusions and Future Work 13.1 Conclusions 13.2 Contributions 13.3 Future Recommendations References and Bibliography

237 237 237 238 240

List of Important Terms

243

Bibliography

247

Glossary

249

Index

251

Abstract Engines and pumps are common engineering devices which have become essential to the smooth running of modern society. Many of these are very sophisticated and require infrastructure and high levels of technological competence to ensure their correct operation. For example, some are computer controlled, others require stable three-phase electrical supplies, or clean hydrocarbon fuels. The first part of the project focuses on the identification, design, construction and testing of a simple, yet elegant, device which has the ability to pump water but which can be manufactured easily without any special tooling or exotic materials and which can be powered from either combustion of organic matter or directly from solar heating. The device, which has many of the elements of a Stirling engine, is a liquid piston engine in which the fluctuating pressure is harnessed to pump a liquid (water). A simple embodiment of this engine/pump has been designed and constructed. It has been tested and recommendations on how it might be improved are made. The underlying theory of the device is also presented and discussed. The second portion deals with noise,vibration and harshness performances of internal combustion engines. Features of various sources of noise and vibrations have been discussed and major focus has been on combustion based noise and piston secondary motion.Various equations of piston motion were solved and effects of various parameters on it were analyzed.

ix

List of Symbols Symbol

Definition

Units

V P T R v I

Volume Pressure Temperature Gas Constant Voltage Current Volume flow Rate

cm3 Bar Kelvin J/K-mol Volt Ampere cm3/s

Heat Absorbed Tube Area Charge Specific Heat Kinematic Fluid Viscosity Frequency Radius Of Tube Fluid Displacement Fluid Density Heat Transfer coefficient Tube length Acceleration due to gravity

Joules cm2 coulomb J/Kg-K m2/S Hz cm cm kg/m3 W/m2-k cm m/s2

Q, V Qe A q Cp η ω Rt X ρ U L,l g

xi

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

1 Introduction

1.1

Background

The Stirling engine system was studied years ago. Such engines have merits the basis of sealings, materials, heat transfer rate, size, and weight issues. During past years the major focus has been on various designs of Stirling engine systems. This engine is based on a heated reciprocating system. The gas receives heat and expands at constant temperature. Rate of transfer is higher, which is a major drawback of these engines. In contrary the internal combustion (IC) engine is operated by combustion of air-fuel mixture which results in higher heat and pressure rise which is converted to useful work. The temperature varies with the combustion and piston motion. As the heat is supplied externally the following varieties of sources can be used: Heat from gaseous, liquid, or solid fuel Solar energy Recycled Waste heat 1

2 Liquid Piston Engines Cooling in a Stirling engine cycle can be done in the following ways: Convection cooling Use of cooling fluids like water, ethylene glycol, or a mixture Reversible nature of Stirling engine differentiates it from IC engines. Combustion outside results in lower emissions as well as less noise and vibration. Solar energy may also be harnessed using parabolic dish. As a smaller number of fuel types or heat sources are available, a Stirling system may be designed as such. This system may use solar heating as the primary heat source, as well as a natural gas burner as an auxiliary unit during nights and cloudy periods.

1.2

Types of Stirling Engines

Using basic concepts of heat engineering many designs of Stirling engines have been proposed over past years. These engines may be classified on the basis of mechanical design features as: Kinematic designs: These engines operate on basis of crankshaft and linkage mechanisms in which the motion of the piston is limited by configuration of linkages. Free-piston designs: In these engines the oscillatory motion of the piston in a magnetic field generate electric power. Pressure gradient cause tuned spring-mass-damper motion of displacer. Such machines are simple to operate but more complex on basis of dynamics and thermodynamics. For cooling purposes, the piston may be driven by a motor. Stirling engines may also have alpha, beta, or gamma configurations which are discussed as follows: Alpha engines which are seen in Figure 1.1 have two separate pistons that are linked and oscillate showing some phase lag. The working gas moves to and fro passing through a cooler, regenerator, and a heater between the cylinders. These engines are kinematic engines which need proper sealings. Beta engines that are seen in Figure 1.2 have a displacer-piston arrangement that are in phase with one another. The displacer pushes the gas to and fro between the hot (expansion area) and cold ends (compression

Introduction 3 Regenerator

Gas motion

Cold end

Hot end

Crank shaft

Figure 1.1 Alpha engines. Hot end Heater Displacer

Regenerator Cooler

Cold end

Power piston

Figure 1.2 Beta engines. Cold end Cooler

Power piston

Displacer Regenerator

Heater

Hot end

Figure 1.3 Gamma engine.

area). As the working gas moves, it passes through a cooler, regenerator, and heater. Beta engines can be either kinematic or free-piston engines. Gamma engines which are shown in Figure 1.3 have a system wherein the displacer and power pistons operate in separate cylinders. The displacer moves the working gas to and fro between the hot and cold ends. The cold

4 Liquid Piston Engines area has cold side of the displacer and power piston. As the gas moves, it passes through a cooler, a regenerator, and a heater. These engines can be either kinematic or free-piston type.

1.3

Stirling Engine Designs

The power piston in the engine is connected to an output shaft by linkages. Kinematic design of the engine has following merits: Coordination of various parts for proper motion during start-up, normal operation, and fluctuations of loads. Some disadvantages of such a design include: Need of lubrication due to rotating parts. Need of more maintenance. Proper sealing needed. Some of the novel designs of kinematic engines are discussed next.

Wobble-plate Mechanisms The wobble-plate that is seen Figure 1.4 has a wobble plate which is in a sliding contact with the crankshaft pivoted by connections to pistons as well as connecting rods. This ensures straight travel inside the cylinder with out rotation. The thrust is transferred to the crank at an offset angle Crank shaft Piston

Wobble plate

Figure 1.4 Gamma engine.

Introduction 5 Swash plate

Power piston

Crank shaft Power piston

Figure 1.5 Swash plate engines.

to wobble plate which acts as a double-acting engine using the power stroke of one cylinder to compress the cold gas for the adjacent cylinder. The power piston for one cylinder is the displacer piston for another cylinder. The Z-crank shape that the same to the wobble plate design has pistons connected directly to the crankshaft. Pivot points are made in order to ensure axial motion of the piston in the cylinder. Such design is more compact as compared to a single-piston Stirling engine. However these engines have certain demerits: Cyclic load and wear of pivots is quick as they are under compression and bendings. Piston-lubrication is a major issue. Oil flow may cause fouling and lesser external heat transfer so reducing the efficiency. Swash Plate Drive mechanisms – This drive has may same features as wobble plate. Bearings are used to connect the swash plate to the crankshaft and rotates with the crankshaft, but the wobble plate which remains fixed is attached to the shaft. This design has many merits: Quiet operation, better sealings with lesser lubrication problems. Design of swash plate may be changed for better stiffness and power transfer. The balancing of swash plate can be done built by adding additional sets of pistons. This in turn increases the power output and reduces the power-to-weight ratio. Rhombic Drive – In this mechanism, yokes connect the power piston and the displacer piston. These are linked to twin crankshafts by means of connecting rods, as seen in Figure 1.6. In this drive mechanism power

6 Liquid Piston Engines Displacer piston

Regenerator

Power piston Connecting rod

Yoke power Gear

Displacer york

Crank disc

Figure 1.6 Rhombic drive engine.

piston and the displacer piston move with constant lag. The rhombic drive has many benefits: The engine has less vibrations due to complete balance of various lateral forces. These engines operate at higher power outputs due to higher pressures. Many units can move at same time in order to provide power to a multi-cylinder engine.

1.4

Free-Piston Stirling Engines

These engines have two oscillating pistons that are not connected as seen in Figure 1.7. The displacer piston has a smaller mass compared to the power piston. The heavier piston moves undamped. Motion of the displacer is simulated by springs or by the compressible working gas. The springs placed between the displacer and the power piston provide harmonic oscillations of the displacer. These oscillations are maintained by temperature difference, and so the system operates at the natural frequency. The power in a free-piston system is generated by a linear alternator. Recently some of the designs have been using a hydraulic drive to run the crankshaft. Use of these hydraulics is good in engines having more torque which reduces the lateral forces in such systems.

Introduction 7 Heater

Hot end

Regenerator Displacer Cooler

Cold end Alternator

Electric output

Figure 1.7 Free piston engine.

Free-piston systems have major advantages: Less lateral forces and lubrication needs due to absence of rotating parts. Less maintenance. Properly sealed units prevent loss of the working gas. These systems have following disadvantages: Need of complex calculations to ensure proper working. Lower response time as compared to kinematic and IC engines. Piston position is an important parameter to control system as oscillations may become unbalanced. The Alpha configuration of engine is the simplest form having two pistons and two cylinders connected by a regenerator. Both these cylinders are normal to one another connected by a flywheel. The hot piston is in contact with located high-temperature source while the cold piston is with the lowtemperature reservoir. The pistons are arranged in a manner that the linear motion is converted to rotatory motion and a constant phase difference is maintained. The pistons are joined at a common point on flywheel. As compared to the other basic designs the alpha type engine has greater volume due to higher compression ratios.

8 Liquid Piston Engines

Figure 1.8 An Alpha Stirling engine.

Cold end

Hot end

Figure 1.9 Alpha engine - Transfer phase.

WORKING OF ENGINES: working of a Stirling engine can be divided into four operations steps that are similar to I.C. engine. Heat is added and removed at constant temperatures. The working of I.C. Engines occurs on the basis of Otto and Diesel cycles, respectively. Mechanisms of these engines is complex as motion is based on movements of multiple pistons. Working of an Alpha Stirling can be analyzed as follows: 1. Transfer of working gas from cold side to hotter side: Flywheel moves clockwise, the hot piston moves towards right hand side towards Dead Centre and the cold piston moves up towards Top Dead Centre (TDC) as seen in Figure 1.9.

Introduction 9 Cold end

Hot end

Figure 1.10 Alpha engine - Power stroke.

The regenerator connects both pistons and operates at hotter temperatures. The pistons move in such a manner that the change in the engine volume is minimum and heat addition occurs at constant volume. Towards end of the process, the working gas will be hotter and the major portion of remains in the hot cylinder. This is similar to suction stroke. 2. Power stroke As flywheel roates by 90°, the majority of the working gas is now in the hotter cylinder and volume of the engine is minimum. The fluid receives heat from a hot source. It expands moving the flywheel further. This is similar to power stroke of the engine and all energy is derived from this stroke. As the hot piston moves towards right side due to gas pressure, the gas expands, with a portion passing through the regenerator. As the heat added to the system at constant temperature it is converted to work, with a little rise in temperature. A perfect isothermal processes will cause a phase change. This may be compared to the power stroke. The working fluid expands to about three times its original volume. The flywheel turns by another quarter rotation and the hot piston starts to move to Dead Centre. The cold piston moves downwards. The regenerator gets heated up as the hot fluid passes by. Heat rejection occurs at constant volume and can be seen as the exhaust stroke of the engine.

10 Liquid Piston Engines

Figure 1.11 Alpha engine - Transfer stroke.

Figure 1.12 Alpha engine - compression stroke.

3. Compression stroke The crank moves by quarter of rotation. The cold piston is at the bottom dead center location and the hot piston moves towards inner dead center. The working gas has major portion in the cold cylinder which cools down rejecting heat to cold reservoir. As the cold piston moves to the top dead center, volume is reduced and the working gas is compressed. During Isothermal Compression the working gas rejects heat and gets compressed. There is minimum change in the internal energy and work needed is also minimum. Towards the end of the process, almost all the gas

Introduction 11 Hot cylinder Cold cylinder

Triangle connecting rod

Figure 1.13 Ross engine. Heater Expansion space

Regenerator

Cooler Compression space

Figure 1.14 Double-acting engine.

is in the cold piston, so volume reduces to about one-third of its original and the cycle goes on. This final stroke may be compared to action of a supercharger or turbocharger. There is no need of compression inside the power piston. This mechanism was first proposed by Andy Ross. This linkage makes design more compact as connecting rods move in a straight line. This in turn reduces the force on the pistons and thus improves performance of the engine. Wear is less due to less friction and life is also increased. The double-acting-engine has four cylinders. The pistons act as the expansion space of one engine and at the same time as compression space

12 Liquid Piston Engines

Pistons

Bearing

Power shaft

Figure 1.15 Rocking yoke.

Pistons

Gears

Power shaft Crank shaft

Figure 1.16 Gear mechanism.

of a neighbouring one. Thus this is the same as four Alpha engines. Sir William Siemens has done major work to develop these engines. The cylinders are connected in a circular manner with cold and hot regions of neighbouring cylinders connected by a reservoir. Hence the outlet of the last cylinder is connected to the first one. So this system is more compact and with high specific power output. All the pistons move at a phase difference of 90°. Maintaining the phase difference between the pistons and harnessing power is complicated. All the above mentioned mechanisms face problems due to excessive side thrust and excessive wear and they have lower life and reliability.

Introduction 13

Heater tubes Swash plate drive

Combustor

Cooler Piston

Regenerator

Figure 1.17 Swash plate mechanism. Expansion space

Heater Regenerator

Cooler Valve

Compression space

Manifold

Turbine

Figure 1.18 Beal engine.

William Beale later designed an engine in which a turbine was used to harness power out. Such mechanism that is seen in Figure 1.18 uses gas compressors to run turbines. Double-acting compressors may be used for more pulses of air per cycle but at lesser specific power of the engine. Uniform loading and lesser thrust force also increases life of the engine. WORKING: Working of the double acting Stirling engine can be understood from the design of Alpha Stirling engine. Various engines in alpha design can have the same stroke. For that the phase lag between any two adjacent pistons must be 90°.

14 Liquid Piston Engines As shown in Figure 1.18, first see the first piston moving downwards. The engine between the last and the first cylinder may be considered as a fourth engine. The first and second piston move downwards at the same time, this transfers the working gas to the hotter side with negligible change in volume. Hence the first engine is working on Isochoric heat transfer in working fluid. The second one is vicinity of the BDC and third piston moves upwards. The second one is in power stroke, and the volume of fluid is maximum. Similarly, other engines are in transfer stroke which moves fluid from hot to cold end. A Beta configuration of engine has a displacer and piston in same cylinder with a 90° phase difference. Robert Stirling was first to invent a Beta Stirling engine that was an inverted beam engine. It was similar steam engine having a beam linkage. These are more suited for space limited applications, but output is lower than other engines. Use of a regenerator is complex in absence of insulation between the hot and cold ends. There is loss of RPM of the engine and hence its output. WORKING: The basic working of a Beta Stirling engine is similar to that of Alpha Stirling engine. The difference lies in the way the working gas

Figure 1.19 Beta engine.

Piston

Compression space

Displacer

Expansion space

Beta engine

Regenerator Cooler

Figure 1.20 Beta engine in working.

Heater

Introduction 15 moves in the cylinder. Displacer causes motion of the working gas and also hinders the use of wire gauss/mesh. For an Alpha engine system, it is easy to see engine strokes. For every quarter rotation of a stroke was finished. However, for a Beta Stirling engine the four strokes may not be easily distinguished. The piston reaches in vicinity of dead center positions near end of strokes. As the transfer strokes are at constant volume, there is a small change in volumes. A vertical configuration of a Beta engine is seen in Figure 1.21. Clockwise quarter movement of crank results in transfer stroke which results in working fluid motion from the cold end towards hot end. Power and compression strokes show a major changes in volumes, whereas transfer stroke show a minimum changes. The displacer is ahead to piston by a quarter of stroke without change in volume. It helps in motion of fluid between the hot and cold ends. As seen in Figure 1.22, the displacer is at dead center while the power piston moves right. The left part of figure shows the position of piston and displacer at the start of the stroke, whereas the right part of figure shows their positions at the termination of the stroke that are opposite at initial and then in same phase. The piston motion is small due to a small changes in volume. Heat addition occurs at constant volume in the Stirling cycle. As a regenerator is not present in cylinder, it is usually placed between the cold and hotter sides of the engine. These devices have lesser efficiency. In Some cases heating is done from conduction of heat from hotter side of cylinder so there is no need of a regenerator.

Transfer of working fluid from hot to cold end

Expansion stroke

Compression stroke

Transfer of fluid from cold to hot end

Figure 1.21 Vertical beta engine.

16 Liquid Piston Engines

Figure 1.22 Working of beta engine.

Figure 1.23 Working of beta engine – expansion.

From the diagrams we can predict the motion of the piston and displacer at the end of each stroke. Towards the end of the transfer stroke as a major portion of the working gas is present in the hotter side of the system and the power piston moves towards dead center. The flywheel turns by 90° and the gas expands at constant temperatures getting heat expands which results in motion of power piston towards dead center. The beta design gives lesser output as compared to alpha one. From the figures it is clear that power piston moves by a larger amount whereas the displacer moves only a short distance which denotes power stroke. 3. Transfer of working fluid from hot end to cold end -towards the end of the expansion phase, the displacer moves towards inner dead center, while the piston is at outer dead center. The flywheel rotates by another 90°. The hot fluid passes in the gaps of the displacer and moves towards the cold side. The motion of the displacer is more than the piston. The fluid loses heat to cylinder walls at constant volume.

Introduction 17

Figure 1.24 Transfer stroke – Beta engine.

Figure 1.25 Compression stroke - beta engine.

Rhombic linkage Flywheel Displacer

Power piston

Figure 1.26 Rhombic drive – Beta engine.

4. Compression stroke After that flywheel moves by another quarter cycle as most of the working fluid is now near the colder section of the cylinder. The piston and the displacer move towards inner dead center as volume is reduced. The fluid is compressed at constant temperatures in contact with the cold reservoir. At the end of the this stroke, the transfer again starts and the cycle repeats. The Rhombic drive has a single cylinder with two separately moving pistons. The Beta design was designed by Rolf Meijer of Philips company, in the 1960s.

18 Liquid Piston Engines Unlike the conventional engines, the connecting rods in this drive are rigid which makes its operation smooth and lesser vibrations. So the engine is quieter as thrust force is reduced resulting in lesser wear. In spite of these merits this linkage is difficult to assemble.

1.5

Gamma Type Engine

A Gamma design of engine has a displacer-type Stirling engine, having power piston in a separate cylinder. This arrangement allows complete separation between the heat exchangers and working space. Displacer cylinder diameter is bigger than the power piston diameter and so larger unswept volumes are seen as compared to Alpha or Beta designs. Though two separate cylinders are present only one is sealed. Power piston remains isolated from the reservoirs. So sealing is easier compared to Alpha designs. Issues of sealing are important as closer tolerance that fit are needed for effective working of the engine. As sealing of the Gamma engine is easier these are easily manufactured. As the heat exchangers are placed on the larger displacer cylinder their placement is easier. Better heat exchangers can be designed as the area available is much larger. Due to this flexibility in the design with use of water-cooled having cooling jackets, have been made. Gamma engines are a major focus of work due to its ease of manufacture. Two types of these engines known as the Ringbom engines and the Low Temperature Engines have been made. The Low Temperature designs are most popular. On the other hand, the Gamma engines have the minimum specific power as during the expansion is incomplete as only a portion of fluid

Figure 1.27 Gamma engine.

Introduction 19 reaches the hot region. As a result, some of portion of expansion occurs in the compression space. As compression ratios are lower so, these engines are used in case need of separate cylinders. WORKING: working of a Gamma Stirling design is similar to that of a Beta Stirling engine. The displacer moves in a separate cylinder, whereas the power piston moves in separate cylinder. Both these cylinders are linked in order to allow passage of the working fluid between them. 2. Transfer stroke At the onset of the stroke, major portion of the working fluid is present in the compression space. The displacer moves up towards the top dead center, whereas the piston moves downwards. The piston motion is lesser when compared to movement of the displacer as heat addition occurs at constant volume.

Displacer

Compression space

Expansion space

Gamma engine Regenerator Piston

Cooler

Figure 1.28 Gamma engine working.

Piston

Displacer

Figure 1.29 Gamma engine – transfer stroke.

Heater

20 Liquid Piston Engines

Figure 1.30 Gamma engine – expansion stroke.

3. Expansion Stroke After the flywheel moves by a quarter of cycle, during the start of transfer stroke most of the working gas is in the expansion area near hot reservoir. At the start of the stroke, the displacer moves up to top dead center, whereas the piston moves upwards. As the gas expands, piston moves upwards during power stroke of the engine. As partial expansion occurs in the compression space, power produced is lesser. The flywheel rotated by another 90˚, the working gas expands and most of it is near the hot reservoir. The displacer is now moving towards the BDC while the piston is moving upwards to its TDC. As the displacer moves down it pushes almost all the working fluid to the compression region of the cylinder. 4. Compression stroke As the flywheel has moved by another 90˚, most of the working fluid is in the compression space. As the piston moves down towards BDC, the effective volume inside the cylinders is reduced. The working fluid which is now in contact with the cold reservoir is compressed. Not all the working fluid is in contact with the cold reservoir as the compression space extends into the power piston cylinder which is not in contact with any thermal reservoir. This increases the work required in compression and the effect is reduced specific power output. If the cold reservoir is the atmosphere, this effect is somewhat negligible.

Introduction 21

Figure 1.31 Low temp engine.

Figure 1.32 Sneft engine.

The compression ratios for a Gamma Stirling engine are much smaller compared to other types as the swept volume of the power cylinder is much smaller compared to the total volume of the engine. In many cases foam is used to make displacer which also has some effects of regeneration. The cylinder of displacer that has smaller diameter

22 Liquid Piston Engines encloses cylinder. Power piston and cylinder are placed on the top plate of the displacer. In such cases air acts as working gas. The power produced is lesser of order of 1W. Such a model seen in Figure 1.32 works at a temperature difference of about 40 °C. Ringbom design of engines are also Gamma type Stirling in which just the piston is connected to the flywheel. The displacer always has a large clearance near dead center positions that is denoted by stops. Such an engine makes a sound in dispalcer and is called as thumper engine. Higher gas pressures pushes the displacer to move up and down. WORKING: In a Ringbom engine, mean operating pressure is same as the atmospheric pressure and works in the absence of a mechanical linkage. The figure shown above is near end of the compression stroke, as pressure inside the cylinder rises which transfers major portion of fluid in the compression area. As the displacer cylinder pressure rises, the force acting on it also increases. The pressure on the top and bottom of displacer is nearly equal. But the forces are not equal as push rod is present near top. This results is an upward force having magnitude equal to pressure acting on pushrod area. As this force is more than the displacer weight, the displacer piston moves up. As the compression goes, the rise in pressure is more which lifts displacer upwards. The piston and displacer motions of Ringbom engine are shown in Figure 1.33. 1. Transfer of fluid from cold to hot end: Displacement of the working fluid occurs towards the hotter side as the displacer piston moves up. Also remember Pmax> Patm > Pmin. 2. Expansion or power stroke At the onset of this stroke, the pressure is maximum. The piston starts moves downwards, which causes gas to expand. As the pressure falls, displacer starts to slow down and comes to stop producing thumping sound. As this stroke ends the pressure falls below the ambient pressure and the displacer moves downwards. 3. Transfer of working fluid from hot to cold end. As the pressure inside the displacer cylinder further drops the transfer stroke begins. The displacer moves with down wards acceleration and the working fluid is moves to the cold end. At the end of this process,

Introduction 23

Guide for displacer push rod

Displacer

Piston

Figure 1.33 Ringbom engine.

Displacer push rod

Piston

Displacer

Figure 1.34 Cut out section.

24 Liquid Piston Engines Power piston

Displacer rod

Displacer piston Displacer cylinder

Heat plate

Figure 1.35 Free-piston engine.

theoretically the pressure will be least and the displacer will be downwards, while the piston starts to move upwards. 4. Compression stroke. As the piston starts to move upwards for the compression stroke, the pressure inside the engine starts to increase. The displacer decelerates and soon stops. As the pressure exceeds the displacer weight, it moves upwards. Free-Piston Stirling engine is the general term given to those which have pistons which are not mechanically connected to a flywheel. There are types that even have liquid pistons or diaphragms to do the job of mechanical pistons. Since there is no connecting rod and flywheel arrangement to convert the linear motion of the piston into rotary motion, some co-axial devices have to be used to harness the produced work. The most commonly used co-axial device is the linear alternator. Another application of the free-piston Stirling engines is to function as a pump, as linear motion of the piston can be easily integrated for that purpose. W. T. Beale has done pioneering work in free-piston type engines which can overcome the lubrication problem. Many of these engines use springs, gravity or inertial masses in order to supply the energy needed for the compression. WORKING: Ringbom Stirling is a form of free piston engine. In these engines, gravity powers the compression stroke. Either springs or inertial masses may be also provided to give necessary energy. Displacer motion is similar to the Ringbom engine.

Introduction 25

(a) Extended rod

(b) Inertial mass

(c) Sprung motor/alternator

(e) Spring to piston

(f) Casing reaction (harwell)

Engine

Heat pump (d) Duplex

Cold

Hot

(g) Fluidyne

Figure 1.36 Various free engines.

Figure 1.37 Transfer stroke – free engine.

(h) Alpha (three to six cylinders)

26 Liquid Piston Engines

Figure 1.38 Expansion stroke – free engine.

Figure 1.39 Transfer stroke – free engine.

Transfer of fluid from cold to hot end: Initially, the pressure inside cylinder is higher and displacer moves upwards. The working gas moves towards hotter end of the engine. 2. Expansion or power stroke: During the onset of the expansion stroke, the pressure inside is highest due to which power piston moves down thus there is an increase in the volume.

Introduction 27

Figure 1.40 Compression stroke – free engine.

The gas expands further causing a fall in pressure. At the end of the stroke the power piston moves down. 3. Transfer of fluid from hot to cold end: As pressure falls near to the atmospheric pressure, the displacer moves downwards and hence the working gas is moved to the cold region. This process continues till the displacer stops at BDC. 4. Compression stroke: The piston mass powers the compression with initial contribution due to atmospheric pressure, but later its effect becomes lesser. During the fall of piston under gravity, the volume reduces and working fluid gets compressed. The piston decelerates towards the end of stroke reaching the bottom dead center at the end of the stroke.

References and Bibliography 1. 2. 3.

Stirling engine assessment, EPRI, paolo alto,CA;2002,1007317. http://www. engr.colostate.edu/~marchese/mech337-10/epri.pdf http://docslide.us/documents/stirling-engine-assessment.html http://www.vineethcs.com/pdf/Stirling%20Engines-A%20Begineers%20 Guide_rev_2.pdf

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

2 Liquid Piston Engines

2.1 Introduction Water is an important civic amenity needed for survival. Oceans are the major source of the water cycle, covering about 71% of the global area, whereas about 3% of freshwater is found as polar ice and glaciers. Some major sources of water on earth and their abundance is shown in the figures below. To account for our basic needs, about 20–50 liters of water are needed every day for the use of a single person. The constant growth of the human population has exposed many citizens in the developing nations of Africa and Asia to an acute shortage of pure and usable water. A United Nations report estimates that about 27% of the population in Africa and 65% in Asia are compelled to use contaminated water unfit for daily use [5]. The graphs presented here clearly show the water stress on human populations. A major source of water available on earth is surface water in the form of freshwater rivers, streams, ponds, etc. But the majority of water bodies are being polluted due to industrial or human activities, leaving the global

29

30 Liquid Piston Engines Fresh 3%

Rivers 2%

Ground water 31.3%

Saline 97%

Swamps 11%

Ice caps and glaciers 68.7%

Total water

Lakes 87%

Fresh water

Fresh surface water

Figure 2.1 Percentage of water sources on surface of earth.

80% 60% 40% 20% 0% 1970

1980

1990

2000

Figure 2.2 Percentage of global population exposed to water shortage. Euurope 2%

America 65%

Africa 27%

Asia 65%

Figure 2.3 Percentage of population exposed to polluted water.

population exposed to water-borne disorders. According to an estimate, about 2 million tons of waste are being dumped into water bodies and about 6,000 children die due to water-borne disorders each day [6]. This situation can be visualized in some of the pictures shown below.

Liquid Piston Engines 31

Figure 2.4 Sources of polluted water in underdeveloped nations.

Renewable groundwater sources on a national basis

Annual withdrawals (cubic kilometers) 0–5 5–20 20–100 More than 100 No data

Figure 2.5 Global groundwater withdrawal.

In such conditions groundwater can be a good way of meeting the needs of the population. Groundwater makes up 31% of the total freshwater supply on this planet. Groundwater is available in porous aquifers. An aquifer is a permeable rock from which the groundwater can be extracted easily. It needs to be pumped out by a suitable mechanical pump running on electricity or a fuel-operated engine [8]. Global usage of groundwater can be seen clearly from the data presented herein. Due to scarcity of fossil fuels, there is a need to look out for alternative sources of energy, including renewables like wind, solar power, geothermal power, etc. Harvesting of solar energy is a good way to meet the future energy demands of our society. The average yearly solar flux reaching earth is 174 PW. About 30% of this is reflected back by clouds, atmosphere and

32 Liquid Piston Engines the surface of earth. The remaining 70% is used to heat water in oceans and on the surface of earth [9]. Solar flux can be used in an indirect or a direct way. The indirect way includes use of bio-mass, wind or solar thermal pond, whereas the direct one includes use of photo voltaic cells or utilization of solar flux for raising vapor cycle to run a heat engine. Alternatively, solar energy can be used to run a liquid piston fluid pump for raising groundwater, thus avoiding the use of polluted water from polluted water bodies. These pumps can also be used for the irrigation of crops. Liquid piston engines are a good choice to tap vast solar power. Their operation dates back to the early 19th century when the Humphrey engine was first used. These engines have many advantages over modern conventional engines. They do not need complex mechanisms like crank or cylinder, are noiseless in operation and simple to construct. There is no need for lubrication in the absence of sliding parts. However, various designs require heat input and output at vastly different temperatures; hence the construction material must be able to resist corrosion as well as high temperatures [10].

2.2 Objectives To design and fabricate a novel heat pump which is affordable in the developing world, hence create the awareness for alternative low-cost energy sources.

Imcoming solar radiation 100% 30% Lost to space

Top of atmosphere 19% Absorbed in the atmosphere

51% Absorbed at surface

Figure 2.6 Solar energy balance.

Liquid Piston Engines 33

2.3 Brief Overview of Pumps and Heat Engines Pumps – convert mechanical energy into fluid energy. Turbines – exactly the opposite, convert fluid energy to mechanical form. Classification of pumps – based on the method by which mechanical energy is transferred to the fluid – Positive-displacement pumps Kinetic pumps Under positive-displacement These pumps discharge a given volume of fluid for each stroke or revolution. Energy is added intermittently Reciprocating action – pistons, plungers, diaphragms, and bellows. Rotary action – vanes, screws, lobes. Types of positive displacement pumps Peristaltic pumps Fluid captured within flexible tube Tube is routed between rollers – rollers squeeze tube and move liquid as parcels Gear Rotary

Vane Screw Progressing cavity

Positive displacement

Lobe of cam Flexible tube (peristaltic) Piston Reciprocating

Plunger Diaphragm

Radial flow (centrifugal) Kinetic

Axial flow (propeller) Mixed flow

Jet or ejector type

Figure 2.7 Pump types.

34 Liquid Piston Engines Vane Vane slot

Cam ring

Suction

Discharge Inlet

Outlet

Drive shaft Rotor

Figure 2.8 Vane pumps.

Discharge manifold Piston

Discharge

Suction Suction manifold (a) Single acting—simplex

(b) Double acting—duplex

Figure 2.9 Simple and double-acting pumps.

Avoids Contact of Liquid with Mechanical Parts Kinetic Pumps Transforms kinetic energy to static pressure – adds energy via rotating impeller Fluid enters through the center of an impeller and is thrown outwards by the vanes

Jet Used for household water systems. Composed of centrifugal pump and jet assembly Suction is created by the jet in the suction pipe

Liquid Piston Engines 35 Discharge

Intake

Discharge

Intake

Discharge

Flow rate

Time

1 Revolution

(a) Single-acting pump—simplex Side #1 Side #2

Discharge Intake

Intake Discharge Discharge Intake

Intake Discharge Discharge Intake

Flow rate

Time 1 Revolution

(b) Double-acting pump—simplex

Figure 2.10 Discharge rates.

(a) Peristalic pump with variable-speed drive system

(b) Peristalic pump with case open to show tubing and rotating drive rollers

Figure 2.11 Peristaltic pumps.

Comparisons between the two types: Characteristic

Positive-displacement

Kinetic

Flow rate

Low

High

Pressure rise

High

Low

Self-priming

Yes

No

Outlet stream

Pulsing

Steady

Works with high viscosity fluids

Yes

No

36 Liquid Piston Engines Outlet Outlet Fluid inlet

Fluid inlet

(a) Radial-flow impeller

(b) Mixed-flow impeller

Outlet Fluid inlet

(c) Axial-flow impeller (propeller)

Figure 2.12 Kinetic pumps.

(a) Pump and motor

(b) Cutaway of pump

(c) Radial-flow impeller

Figure 2.13 Centrifugal pumps.

Pump selection depends on – Discharge Head requirement Horsepower requirements of the pump

Liquid Piston Engines 37 Discharge pipe Motor Impeller

Pressure pipe

Suction pipe Diffuser

Nozzle

Foot valve with strainer

Figure 2.14 Jet pumps. 40

30

Volumetric efficiency Overall

20 efficiency

100 80

Input power

10

60

Input power (hp) Efficiency (%)

Pump capacity (gal/min)

Capacity

40 20

0

0

500 1000 1500 Discharge pressure (psi)

0 2000

Figure 2.15 Performance curves.

As pressure increases there is a slight decrease in capacity due to internal leakage from the high-pressure side. Power needed varies linearly with pressure. Volumetric efficiency = flow rate delivered/theoretical flow rate (90 to 100%). Theoretical flow rate is based on displacement per revolution times the speed of rotation.

38 Liquid Piston Engines 200

60

40 30

100

20 50 0

Total head (m)

Total head (ft)

50 150

10 0

500

0

2000

1000 1500 1000 Pump capacity (gal/min) 4000 6000 Pump capacity (L/min)

8000

2500

10000

Figure 2.16 Centrifugal pump performance curves.

Capacity decreases with increasing head. At “cut-off ” head flow is stopped completely and all energy goes to maintaining the head. Typical operating conditions well below “cut-off ” head. Overall efficiency = power delivered to fluid/power supplied to pump.

2.4

Heat Engine

The onset of industries had led to rapid development of the steam engine. Heat can be transformed into useful work and vice versa. This is the main feature for working of an engine. A refrigerator is also another form of engine since it transfers heat from a cold body to a hot one. In this portion major features of the Carnot engine have been discussed. The first and second law of thermodynamics are as follows: First Law – Energy remains constant in a system i.e.

ΔEth = W + Q. Second Law says that all macroscopic processes are irreversible. Heat can itself move from a hot to a cold body, but reverse process is impossible. The area under p–V denotes done is done by an ideal gas. This work is positive when the energy is out of the system.

Liquid Piston Engines 39 Heat from hot to cold body QH

QC

Copper bar

FIRI

ICI Cold reservoir at Tc

Hot reservoir at Th

Figure 2.17 Heat engine.

Hot reservoir

Th Qh

System

Copper bar as system Qc

Cold reservoir

Tc

Figure 2.18 Energy transfer.

The first law of thermodynamics can be expressed as:

Q + WS = Eth Energy that is transferred inside a system as heat that is stored inside and which causes an increase in thermal energy. A reservoir of energy shows a larger transfer of thermal energy without a significant change of its thermal energy. Hot and cold reservoirs have temperatures that are denoted as TH and TC. The energy transfer diagram of a heat engine has been shown in Figure 2.17. A copper bar is a system that serves as a path of heat transfer between a hotter and a cold body. Amount of heat transferred from the reservoir is denoted by QH, whereas the heat transferred out is denoted by QC. From the previous equation we have

Q = Eth + WS Where Q is the net heat to system. Assuming that the work done is zero. As the bar is at steady state, ΔEth = 0. So QC = QH.

40 Liquid Piston Engines As seen in Figure 2.19, heat is flowing from a body at colder temperature to one at hotter. Such a process is against the second law of thermodynamics. The process may be compared to action of two stones reaching a higher temperature as we rub them together and then place them in water. Work done on the stones causes an increase in their energy equal to the work done. i.e. W = ΔEth. As these stones contact cold water the heat flows to the water, i.e.

ΔEth = QC. This process can be fully efficient as all the energy supplied to the system is ultimately transferred to the water as heat. The reverse cannot take place without need of external force as seen in Figure 2.20. Second law forbids this action Hot reservoir

Th Qh

System

Copper bar as system Qc Tc

Cold reservoir

Figure 2.19 Heat flow.

Hot reservoir

Th Qh

System

Copper bar as system

W Qc

Cold reservoir

Figure 2.20 Work efficiency.

Tc

Liquid Piston Engines 41 The arrangement shown above has a gas placed in an insulated box having a piston with mass M. The base is heated as it has no insulation. The gas expands that lifts piston by distance Δx at constant temperature. The change in internal energy is null due to this fact and so the first law of thermodynamics can be written as WS = Q. As the piston reaches top most position its lift stops. As the system works in a closed cycle, so its efficiency is full as seen in Figure 2.22.

Heat Engines and Refrigerators An engine is a device that gets heat QH from a hot source, does some work, and then rejects heat QC to a cold source. This process is shown in Figure 2.23 as engine connects to both sources. The whole cycle may be written as

Eth net = 0, E = Q + W = Q – WS = 0. or Wout = Qnet + QH – QC Insulation

M

Piston

Gas X

Flame

Figure 2.21 Iso thermal process. Th

Hot system Qh

W out

Figure 2.22 Perfect engine.

42 Liquid Piston Engines Qh is transfered from hot body to system Hot

Th Qh

Heat engine

W out Qc Tc

Cold

Remaining Qc= Qh-Wout is transfered as waste heat to cold one

Figure 2.23 Energy transfer in engine.

The efficiency defined as –

1

Qc Qn

Maximum value of efficiency is 100 percent. But this needs Qc = 0. Such an engine seen in Figure 2.23 is theoretical. Practically engines have lower values of 10–40 percent.

Perfect Heat Engine As seen in Figure 2.24. As the mass reaches its maximum it is removed as piston in locked position with volume of the gas constant. The cooling occurs and the gas is gets compressed to its initial phase hence completing a cycle. Mass is again placed and the process repeats. This process is seen as a pV diagram in Figure 2.25. As zero work is during isochoric process so network output = (WS)12 + (WS)3

2.5

Clever Pumps

Apart from manmade pumps nature also has many clever pumps such as the human heart, capillary action in plants and neuron pumps in the nerve cells of the human cerebrum. a. Human Impulse: All animals have nerve cells known as neurons present in their cerebrum. These cells transmit nerve impulses from and to the brain which form the basis of

Liquid Piston Engines 43 Gas expands lifting the mass M

Heat transfer from fuel to gas

Piston is locked and mass removed M

Pin

M

No heat Gas

Heat

Heat

Isothermal compression restores gas to initial stage

Gas cools at constant volume

Figure 2.24 Ideal engine.

Isotherm

Heating Qh 2

1

Cooling

P

Qc Qc Compression

3

V

Figure 2.25 Ideal engine P-V curve.

human reflexes, movements, emotions, and senses. Neurons have Na+K+ATPase, which is a protein pump present in neurons of the brain. It utilizes energy from ATP molecule stops pump three sodium ions out of a cell and two potassium ions into the cell. This causes a potential difference across cell membranes called resting potential, which is the basic cause of nerve impulses transmitted across neurons in the human

44 Liquid Piston Engines body. These impulses form the basis of human stimuli. Action of this natural pump can be seen in the figure below [12]. b. Capillary action: this effect occurs due to cohesive, adhesive forces or surface tension and plays an important role in transportation of water. Capillary action in trees helps to draw water into roots by xylem tissue cells. Xylem cells are made of cellulose molecules which form a chemical bond with water, hence helping in circulation of water in a tree [13]. c. Human heart – The human heart is an excellent example of a natural pump. It has four valves, namely tricuspid valve, mitral valve, aortic valve, and pulmonic valve. Starting from the right atrium, blood flows through the tricuspid valve to the right ventricle and is sent to the lungs for oxygen enrichment by the pulmonary artery. From the lungs, blood flows through the pulmonary vein to the left atrium and from the left atrium to the left ventricle through the mitral valve. This enriched blood flows to the aorta through the aortic valve, from where it is distributed to the whole body. Each valve has a set of flaps which maintain blood flow through it [14]. Osmosis or reverse Osmosis are other naturally occurring phenomena, which can be used to design a novel pump capable of selective pumping. 3Na+ Outside Cell membrane

Na

Inside

K

Na+K+ATPase

Closed (leak)

ATP ADP+Pl

Closed (leak)

+

2K

Figure 2.26 Human Impulse pump. CH2OH O

H H O

OH H

O H H OH

H

H

OH

OH

H H

H O CH2OH

CH2OH O H H OH O H

Figure 2.27 Water cellulose bonding.

O H H OH

H

OH

OH

H H

H H

O CH2OH

CH2OH O H H O

O

OH

H H

H

OH

H

OH

OH

H H

H H

O CH2OH

O

Liquid Piston Engines 45

2.6 History and Development of Stirling Engines “As you enter the past, you will find direction for the future” —Ivo Kolin [16]

During the industrial revolution of the 18th century, the steam engine became a primary source of power. But this device has its drawbacks. Its

Aorta

Left atrium Right atrium Septum Left ventricle Right ventricle

Blood vessels

Figure 2.28 Workings of a human heart. Hydrostatic pressure

Water

Water

Solution

Osmosis

Semipermeable membrane

Figure 2.29 Osmosis and reverse osmosis.

Solution

Reverse osmosis

46 Liquid Piston Engines maximum efficiency is at the most 2% and there were many accidents involving explosions. This prompted engineers to look for alternative sources of power like Stirling engines. A Stirling engine is a hot-air engine operating on the principle that air expands on being heated and contracts on being cooled. These devices have zero exhaust and are external combustion engines; hence a wide variety of fuels can be used to run a Stirling engine which includes alcohol, bio–products, or waste gases, etc. These engines are suitable for operations which have the following needs [16]: A) Constant power output. B) Noiseless operation. C) Long startup period. D) Low speeds. The development of the Stirling engine is widely attributed to the Scottish scientist Sir Robert Stirling. The first version of this engine, developed in 1815, was heated by fire and air cooled. Figures of some of these early versions are presented in the sections below. Later, in 1864, Erickson invented the solar-powered engine to heat the displacer tube at the hot side. The heat was obtained by use of solar

Figure 2.30 Earliest version of a Stirling engine developed by Stirling brothers.

Figure 2.31 Alpha-type Stirling engine developed in 1875.

Liquid Piston Engines 47 reflectors. The first alpha-type engine was built in 1875 by Rider. Reader and Hooper proposed the first solar-powered heat engine for irrigation purposes in 1908. Following this, Jordan and Ibele designed a 100  W solar-powered engine for pumping of water. In 1983 a low-temperature difference Stirling engine was patented by White, which had an efficiency of about 30%. Colin later presented a design with a low temperature difference of 15 °C, and Senft published specifications of an engine with a very low temperature difference of 5 °C between hot and cold ends [16]. Some of the following events can be considered as important milestones in the design and development of a Stirling engine for use as a pump: 1688: Thomas Savery develops a drainage pump which was a liquid piston machine. 1909: Development of Humphrey pump. 1931: Malone designed and developed an engine with regenerative cycle similar to a Stirling engine. 1965: Philips Company patented a Stirling engine. 1977: The metal box company developed Stirling engine for irrigation purposes in Harwell lab. 1985: McDonnell designed an engine with parabolic reflectors to focus solar energy, thus achieving a high temperature of 1400 °C. A thermal engine is a device which converts heat energy into mechanical energy. The operation of a heat engine can be described by a simple thermodynamic cycle as follows:

Efficiency W Qh Qc Qh Qh

Figure 2.32 McDonnell engine.

48 Liquid Piston Engines High temprature source

Q1 Heat engine

W

Q2

Low temprature sink

Figure 2.33 Heat engine. Energy conversion process Chemical energy Combustion

Heat energy

Machine

Mechanical energy Generator

Electrical energy

Heat engine

Figure 2.34 Energy conversion in a heat engine.

Heat engines can be further classified as external combustion or internal combustion engines. An engine where fuel is burnt outside the engine is an external combustion engine, whereas in the internal combustion engine, the fuel is burnt inside the engine. An engine operating on a Carnot or Stirling cycle is an example of an external combustion engine while one operating on an Otto or Diesel cycle is an internal combustion engine. Comparison of these cycles is presented below:

2.7 Operation of a Stirling Engine “In all places where there exists a difference of temperature, there can be a production of motive power.” —Sadi Carnot, 1824 [16]

In a Stirling engine the fluid is contained in a confined space; hence there are no problems of contamination. In order to reduce the heat losses, the

Liquid Piston Engines 49 Table 2.1 Comparison of various engines. Type of combustion Cycle

Compression

Heat Heat addition Expansion removal

External

Carnot Adiabatic

Isothermal Adiabatic

External

Stirling Isothermal

Isometric

Isothermal Isometric

Internal

Otto

Isometric

Adiabatic

Adiabatic

Isothermal Isometric

Pressure (P)

Carnot (ideal) heat cycle P-V diagram

C

Heat added entropy increases gas expands constants temperature Isothermal expansion

Gas compressed no heat added or lost constant entropy temperature rises Adiabatic compression

D B

Gas expands no heat added or lost constant entropy temperature falls Adiabatic expansion

Gas compressed A heat given out entropy falls constant temperature Isothermal compression

Volume (V)

Temperature (T)

Carnot (ideal) heat cycle entropy diagram Heat added constants temperature expansion (work done by gas)

C

D Expansion no heat added or lost constant entropy no work done

Compressed no heat added or lost no work done

B

Heat extracted constants temperature compression (work done ON gas)

A

Entropy (S)

Heat in

D B

Heat out S=0 compression

A

C B Compression (work in)

A Volume (V)

Figure 2.36 P-V& T-S plot of an Otto cycle.

Heat in

S=0 expansion

Otto heat cycle entropy diagram

V=0 Expansion (work out) Heat out

Otto heat cycle PV diagram C Temperature (T)

Pressure (P)

Figure 2.35 P-V& T-S plot of a Carnot cycle.

D

Entropy (S)

50 Liquid Piston Engines Expansion space Expansion piston Heater Working gas (air) Regenerator Cooler

Compression space Compression piston

Crank Ross yoke linkage

Figure 2.37 Stirling engine.

mass flow rate must be low, which can be maintained by low viscosity fluid or high working pressures. These engines are 30 to 40% efficient in a temperature range of 923–1073 K [7]. A Stirling engine consists of the following components: 1. Heat source – as fuel does not come in direct contact with the working fluid, Stirling engines can work on fluids which may damage parts of a conventional engine. 2. Regenerator – the function of a regenerator is to prevent the waste heat from being lost to the environment by storing it temporarily, thus helping to achieve high efficiencies close to an ideal Carnot cycle. A simple configuration consists of fine mesh of metallic wires. In an ideal Stirling cycle, the connecting space between hot and cold ends acts as regenerator. 3. Heat sink – typically the ambient environment acts as an ideal heat sink; otherwise the cold side can be maintained by iced water or cold fluids like liquid nitrogen. 4. Displacer piston – it causes the displacement of working gas between hot and cold regions so that expansion and contraction occurs alternatively for operation of engine. 5. Power piston – transmits the pressure to the crankshaft. In a Stirling engine, hot air expands when heated and contracts when cooled. This principle of operation was most properly understood by Irish

Liquid Piston Engines 51 scientist Robert Boyle from his results on experiments on air trapped in a J-shaped glass tube. Boyle stated that the pressure of a gas is inversely proportional to its volume and product of pressure and volume occupied is a constant depending on temperature of gas.

Hence PV = NRT Various assumptions are made in this cycle are [16]: 1. 2. 3. 4.

Working fluid is an ideal gas. Conduction and flow resistance is negligible. Frictional losses are neglected. Iso-thermal expansion and contraction.

This cycle can be described by the following stages [16]: 1. Phase C–D: Iso thermal expansion – the working fluid undergoes an iso-thermal expansion absorbing the heat from source. The power piston moves out, hence increasing the volume and reducing the pressure. The work done in expansion of gas is given by:

We

RT log

VD VC

2. Phase D–A: Power piston now reaches the outermost position and stays there so that volume is constant. The working fluid is passed through the regenerator where it gives up heat for use in the next cycle. Hence its temperature and pressure falls. No work is done during this phase. 3. Phase A–B: The power piston starts moving inwards, reducing its volume and increasing its pressure the working fluid gives up heat to cold sink. The work done in compressing the gas is given by:

Wc

RT log

VB VA

4. Phase 2–3: The power piston is at its most inwards point and stays there to keep the volume constant. Working fluid passes

52 Liquid Piston Engines

Heat in

Pressure (P)

Stirling engine heat cycle PV diagram C Regen heat in

T=0 expansion D

Heat out

B

Regen heat out

T=0 compression

A Volume (V)

Heat in

Pressure (P)

Stirling engine heat cycle entropy diagram

C

Expansion D

Regen heat in

V=0

V=0 B Compression

Heat out

Regen heat out

A

Volume (V)

Figure 2.38 P-V & T-S plot of a Stirling cycle.

again through the regenerator, recovering the heat lost in the second phase; hence its pressure and temperature go up.

Wnet

We

Wc

We

R[Th Tc ] log

Vmax Vmin

But

VB = VC, VA = VD

Th Tc Th

4

In a Stirling cycle, two Isochoric processes replace the two Iso-entropic processes in an ideal Carnot cycle. Hence more work is available than a Carnot cycle as net area under P–V curve is more. Thus there is no need for high pressures or swept volumes. This can be seen in the figures presented below.

Liquid Piston Engines 53 p

3

T

pv = C 4C

TH

pvk = C

3

4C

2C

1

4

4 2 TC 2C

2 1 v

S

Figure 2.39 Comparison of Stirling cycle and Carnot cycle.

2.8

Working Gas

It is a gas on which the engine operates. There are several gases that can be used to run a Stirling engine. Lighter gases having atomic mass lesser than that of air have higher specific heat and gas constant and lower viscosity resulting in lesser viscous losses and higher heat storing capacity [16]. This can be seen in the following graph which was obtained by simulation by Philip Brothers.

2.9

Pros and Cons of Stirling Engine

The Stirling engine has some merits as well as demerits, which are discussed below. 1. Merits A. Stirling engines can be run on a wide variety of fuels including solar energy without a need for the fuel to come in contact with operating gas, hence avoiding containment. Hence, even if solar energy is unavailable, alternative fuels can be used for operations. Thus these devices are not susceptible to fuel shortage. B. Low and noiseless operations are possible. Hence suitable for submarines. C. Lower maintenance is needed and combustion of fuel occurs outside the engine.

54 Liquid Piston Engines efficiency %

55 50 45

Hydrogen

Helium

Air

40 35 30

0

20

40

60

80

100

120

140

160

Figure 2.40 Stirling engine efficiency V/S power output for various gas.

Boiler-heat exchanger

Regenerator Heat exchanger

Exhaust gas

Stirling engine Heat consumer

Fuel Combustor Air

Generator

Cooler-heat consumer

Figure 2.41 A CHP Stirling engine.

D. Can be used as a CHP unit. E. No danger of explosion as in steam engines. 2. Demerits A. Commercial feasibility not possible on large-scale manufacturing. B. Takes time to start from cold.

2.10 Low Temperature Difference Stirling Engine These engines can run at a typically low temperature difference of less than 100 °C between hot and cold end. With high temperature difference between hot and cold end, it is necessary to maintain long separation between hot and cold ends, whereas area of heating and cooling is

Liquid Piston Engines 55

High temperature differential engine

Low temperature differential engine

Figure 2.42 Comparison of LTD and HTD engines.

less important. In 1980 Sneft and Kolin developed simple versions of such engines where a cup of hot tea could be used as a heat source. The figures below clearly distinguish between the LTD and HTD engines.

2.11 Basic Principle of a Fluidyne The basic principle of a fluidyne is similar to a Stirling engine. A gas when heated expands, and if its expansion is confined, its temperature rises. This can be understood more easily by the following operations: Initially the displacer piston is at the centre, with half of the gas in the hot side and the other half of the gas in the cold side of the cylinder. The pressure gauge is neutral. As the displacer piston moves towards the cold end, the gas is displaced towards the hot end by the connecting tube, and its temperature and hence pressure goes up, as indicated by the gauge. As the piston moves towards the hot side, the gas is displaced towards the cold end, its temperature and hence pressure falls. The changes in the

56 Liquid Piston Engines

Hot end

Displacer piston

Cold end

Pressure

Figure 2.43 Motion of a displacer piston in cylinder.

Pressure

Hot side

Cold side

Figure 2.44 Motion of displacer piston towards cold side.

Pressure

Hot side

Cold side

Figure 2.45 Motion of displacer piston towards hot side.

displacer pressure can be used to drive another piston known as the power piston. When the gas pressure is high, the power piston moves towards the open end of the cylinder, hence doing some work which can be used to pump water or rotate a crankshaft. But when the gas pressure is low, the power piston returns towards its original position for which work is needed, which is less than the work available from the previous stroke as lesser force is acting on the piston due to low gas pressure. Hence there is an excess of energy that can be used for pumping operation or other tasks.

Liquid Piston Engines 57 Pressure

Hot side

Cold side

Figure 2.46 Motion of displacer piston and power piston.

Pressure

Hot side

Cold side

Figure 2.47 Motion of displacer piston and power piston.

By clever and innovative engineering, some of the power available from the power piston can be used to drive the displacer piston, and so create a variable pressure heat engine.

2.12 Detailed Working of a Fluidyne A fluidyne can be considered a wobbling column of fluid similar to a pendulum or see-saw toy.

Various Phases of Operation of a Fluidyne a. Stage 1 – initially levels of liquid in columns is equal when no heat is applied. b. Stage 2 – as heat is applied at the hot end, the air at that end is heated up and expands, moving towards the cold end through the connecting arm. This pushes the fluid to TDC

58 Liquid Piston Engines

Mean

Extreme 2

1

3

4

5

Figure 2.48 Motion of a see saw and pendulum: Gravity acts as a restoring force to bring back to mean position.

Figure 2.49 Stages of operation of a fluidyne.

at the hot end and BDC at the cold end and the fluid out of the output column. c. Stage 3 – the air comes in contact with fluid at the cold end, cools down and contracts. Once the fluid has reached its extreme positions at both columns of the U tube, at the hot side, the inertia of the weight of the extra risen fluid column tries to bring down the raised level of fluid to its mean position. d. Stage 4 – as this happens, the air is again transferred from the cold end to the hot end through the connecting space, so that the level of fluid overshoots the mean at the hot side and reaches BDC, whereas at the cold end it reaches the TDC and the fluid is again sucked back in the output column. e. Stage 5 – inertia of weight tries again to restore the levels of fluids equal at both ends, so that the cycle starts again.

Liquid Piston Engines 59

TDC Mean position S=length of stroke External heat BDC

Figure 2.50 Stages of operation of a fluidyne.

Initial position of fluid at TDC Mean position S=length External heat of stroke BDC

Inertia of weight

Initial position at BDC

Figure 2.51 Stages of operation of a fluidyne.

Initial position of fluid at TDC Mean position S=length External heat of stroke BDC

Figure 2.52 Stages of operation of a fluidyne.

Initial position at BDC

60 Liquid Piston Engines

Figure 2.53 Stages of operation of a fluidyne. Cold Hot

Displacer

Output

Figure 2.54 General working of a fluidyne. Cold Hot

Displacer

Output

Figure 2.55 General working of a fluidyne.

Analogous to this cycle, a fluidyne operates in the same way with the left end of the U tube acting as a displacer piston, whereas the right end acts as the power piston. Initially most of the air is trapped in the hot side of the engine and the top dead centre of the cold end corresponds to the bottom dead centre of the hot end. The temperature of air rises being in contact with the hot end, hence its pressure rises, which tends to pump fluid out from the output tube.

Liquid Piston Engines 61 Regenerator

Pressure

Displacer

Hot side

Cold side

Figure 2.56 Regenerator.

After half of the cycle most of the air is transferred to the cold side of the machine and so its pressure falls. The cold surface is at bottom dead center and the fluid is pulled back into the U tube.

2.13

Role of Evaporation

Evaporation leads to an increased heat input as the latent heat of evaporation is absorbed by the water present in the displacer column. This leads to a fall in the overall efficiency of the cycle. The evaporation can be suppressed by increasing the pressure of air in the fluidyne. Stirling engines operate at high temperatures of 700–800 °C. However, there are various losses, which bring this temperature down to about 130–300 °C. Evaporation can be suppressed by placing a float in the hot chamber [21].

2.14 Regenerator Though this does not constitute a mandatory part of the engine, use of a regenerator is beneficial. With the use of a regenerator, there is a steady state fall in temperature as the gas gives up heat to the regenerator. Hence by the time the gas goes into the cold chamber it has already been cooled. As the gas moves into the hot chamber, it picks up the heat from the regenerator; thus the regenerator acts as a buffer of heat and increases the efficiency of the cycle. There are several ways to design this heat exchanger. One of the common ways is to increase the heat exchanging, keeping

62 Liquid Piston Engines Air passage

Air flow from hot side to cold side

Heat into regenerator

Tcold

Thot

Air flow from cold side to hot side

Air passage

Heatout of regenerator

Thot

Tcold

Figure 2.57 Heat exchange in a regenerator.

Table 2.2 Comparison of properties of various materials for regenerator [15]. Material

Fibre diameter

Density

Specific heat

MS Steel

30 μm

7.8 g/cc

437 (j/kg°c)

SS Wool

40 μm

7.8 g/cc

510 (j/kg°c)

SS Mesh

100 μm

7.8 g/cc

510 (j/kg°c)

the resistance to flow minimum. The material of the regenerator can be honeycomb, wire meshes, or metallic strips made of high-capacity, heatabsorbing materials. The action of a regenerator and the properties of some materials suitable for use in a regenerator can be seen below:

2.15

Pumping Setups

There are several available setups wherein pressure variations can be used for pumping water. Commonly used pumping configuration involves a T piece at the end of the output tube and two non-return valves. On the

Liquid Piston Engines 63 Gas column

Valves Liquid column Pumping line Gas column

Liquid column

Figure 2.58 Pumping configurations.

outward stroke, the fluid is forced through the upper valve whereas during the inward stroke, the fluid is drawn through the lower non-return valve. However, this setup has some drawbacks. Above a certain pumping head, the work needed to pump the fluid becomes greater than the volume change in the engine. Another configuration uses the pressure variations in the working gas. When the pressure of gas is low, the fluid is drawn up from the lower valve and as it rises, the fluid is expelled.

2.16 Tuning of Liquid Column For any oscillating mechanism, the amplitude of motion is maximum when the frequency of pressure variations on it is equal to the natural resonant frequency of vibrations. The length of the output column has to be adjusted. If the length is too long, then due to the larger mass of the column, working gas would be unable for acceleration. On the other hand, if the length of the column is too small then there would be insufficient pressure built up in the engine. If the displacer is left to oscillate itself, then oscillations in the tube will die due to viscous and other losses. There are several ways to keep the displacer in motion. One of these is the “rocking beam mechanism”. In this mechanism, the whole machine is mounted on a pivot-like spring. As the liquid moves back and forth, its shifting weight causes the whole mechanism to rock like a see-saw.

64 Liquid Piston Engines Heat in

Spring

Figure 2.59 Rocking beam mechanism.

x x

Liquid density P

Cross sectional area A

Figure 2.60 Displacement of fluid in a U tube.

2.17 Motion Analysis Working fluid in our case is air at atmospheric pressure of 100 Kpa, and to get maximum oscillating amplitude of vibrations, flow losses must be minimum. If the fluid raises by a distance X in one column it must fall by the same amount in the other column. Hence the difference between volumes of two columns is given by 2 AX The pressure due to this volume of difference is given by 2 AXg Mass of liquid column is given by AL Hence force acting on column =  ALX

ALX X

2X

2 AXg g l

5

2g radians [21] l sec

Liquid Piston Engines 65

2.18

Losses

There are no moving parts involved in this device, hence no friction. But various other losses in the liquid piston engine which include: 1) Viscous losses in tube – Crandall proposed the resistance coefficient is the ratio of pressure drop per unit length and mean flow velocity and is given by:

)1/2

(2

r

Rt

Viscous power losses are defined as the product of pressure drop and mean velocity of flow, i.e.,

rLV 2 Rt2

PV

LV 2 (2 )‰ 2 Rt Rt 2) Kinetic energy losses – in addition to various viscous losses there are flow losses occurring when the fluid changes its direction or speed at the bends or exits, etc. The pressure drop due to a minor obstruction is given by:

P

K V2 2

Hence the power loss is given by the product of pressure drop and flow velocity.

E

K V 2V PV 2

K V3 2 2 Rt4

Hence total loss is given by summing up individual losses and is given by:

K Va 2 2 Rt4 Where K is a factor depending upon the nature of obstruction.

66 Liquid Piston Engines Table 2.3 Value of kinetic energy loss factor for various configurations [21]. Element

K

90 smooth bend

0.15

90 sharp bend

1

Sharp contraction

0.5

3) Heat losses – This device is a low power output machine compared to conventional Stirling engines, hence various heat losses due to conduction must be minimized. Various components of the system can be viewed as cylinders, for which the heat loss is given by:

Q

2 kL T ln D2 D1

Where k is the thermal conductivity of material. 4) Shuttle losses – Liquid piston when stationary has the temperature equal to that of its adjacent space. When this piston moves to a new region, these are heat losses occurring due to the motion of the displacer piston and are also known as step down losses. They are calculated by formula:

Qs

S 2 k TD 8Lg

Where D – diameter of displacer column L – length of displacer column These losses are dependent on the length of the stroke, which in turn is frequency of oscillations. The more the frequency, the less is the time period available for heat transfer and the more are the heat losses.

2.19 Factors Affecting Amplitude At a particular limit, the liquid piston begins to hit the cylinder top after which the fluid goes into an area of different cross-section. If evaporation is neglected then it is seen that the amplitude of vibrations increase as the

Liquid Piston Engines 67 temperature at hot end goes up. However, it is limited by some of following criteria: 1. Viscous flow losses 2. Poor heat transfer limit 3. Increase in amplitude causes fluid to hit at displacer-air column interface and move into an area of different cross section.

2.20

Performance of Engine

Beals number depending on the temperature of heat source is an important parameter to determine the power output of this device. Beals number is given by ratio Power output Pressure f displaced volume The variation of source temperature with Beals number is shown in graph below.

2.21 Design In this section various factors taken into consideration while designing the setup are discussed.

0.020

Beale no.

0.015

0.010 0.005

300

400

500

Heater temprature in K

Figure 2.61 Variation of Beals number with source temperature.

600

68 Liquid Piston Engines

Collecting cup Wooden frame

Air column

Pumping line

Hot burner

Cold end

Supports Displacer

Reserviour

Figure 2.62 Layout of setup.

The chosen design has the following characteristics: 1. 2. 3. 4.

Easy to assemble. Easy to transport due to small size. Relative low cost. Provision of cheap and ready to use fuel.

Choice of Materials The tables shown in the following sections give an idea about the choice of various design ideas which were evaluated on the basis of various parameters. Red color indicates the most preferred idea, yellow represents a reasonable one, whereas the green color denotes the most preferred choice. The aim of this design was to pump water up to a certain height using a liquid piston engine. Initially a hair dryer was chosen as a source of heat with copper tubes and elbow joints as material for displacer. However, there were several problems with this setup as brazed joints were of poor quality and a hair dryer was found to be insufficient to give the required heat. Hence this setup had to be discarded. There were several delays on account of this, and time and cost were the crucial factors to implement the design. Hence project management was needed so that all accomplished tasks could be done within deadlines. A Grantt chart giving an overall review of various talks related to this project is presented below:

Liquid Piston Engines 69 Table 2.4 Comparison of various design choices. Displacer column material Cost

Availability Ease of working

Glass Plastic Copper tube Heat source

Availability Cost

Required temperature attained

Alcohol Hair dryer

Design Choices Oct–Nov Dec–Jan Feb

March

April

June–July August

Literature Review Initial design Implementation of initial design Final design Search for parts and assembly Testing Writeup

Based on the above criteria, final selection was made for various design parameters as follows: 1. Displacer piston-material must be cheap and corrosion resistant and provide ease of assembly. Plastic tube is ideal for this case, being cheap and readily available. 2. Wooden base for support and robustness. 3. Fuel – methylated spirit and cotton for use in burner due to ease of use. 4. Choice of material for pumping line, burner, air column, cold end: Copper and brass were the choices available. Thermal conductivity of copper is 401 W/m K, whereas for brass it is 109 W/m K, Using copper can cause more heat losses to the

70 Liquid Piston Engines ambient atmosphere. Also, copper is more prone to corrosion. Brass is an alloy of copper and zinc having better corrosion resistance than copper. Hence it is more suitable for use.

Major Components The designed system consists of the following major parts: a. Plastic tube for displacer column of radius pipe 0.63 cm and length 30 cm. It consists of a hot chamber and a cold chamber. b. Pumping Column – Brass column of radius 0.39 cm and height 15 cm. c. Burner for providing the heat. d. Collecting cup. e. Connecting arm of length 6 cm and diameter 3 mm. f. Air column of 5 mm diameter and length 18 cm. g. Supporting wooden base. h. Two balls of mild steel of diameter 5 mm which act as one way valves. i. Brass couple at cold end holding the collecting cup, other end of air column and water returning tube. j. Plastic supports at hot and cold ends. k. 2 hose clips for effective sealing at hot and cold ends.

Cost of Components 1. 2. 3. 4. 5.

Radiation foil – £4.78 Brass pipe – 50 p Brass tap connector – 99 p Copper pipe & brass couple – 50 p Glue – £3

2.22 Assembly Major steps in assembly of components are as following: 1. Machining the surface of brass tap connector (which acts as a burner) and drill to holes for supply of air for burning 2. Joining the two ends of connecting arm by brazing with output column and burner.

Liquid Piston Engines 71 Table 2.5 Volume occupied by various parts. Volume(cm3)

Part list Displacer Tube

29.6

Air column

3.53

Connecting Arm

0.42

Pumping Column

7.19

Pumping Arm

6.93

Connecting Arm

0.42

Air column 8.6%

Pumping column 17.5% Displacer 72% Connecting arm 1.03%

Figure 2.63 Percentage of total volume of system.

3. Machining the plastic supports and making holes equal to the diameter of burner and brass junction to support at hot and cold ends and fixing them by iron nails to the wooden base. 4. Fixing the two ends of displacer plastic tube with the outlets of burner and brass junction. Ensuring tight fitness by fixing the supporting hose clips at the joints. 5. Joining the ends of air column with burner and brass junction. 6. Covering the base with foil to minimize heat losses.

2.23 Calculation Target performance parameters: pumping height = 15 cm, pumping column diameter = 0.78 cm

72 Liquid Piston Engines Available parameters: diameter of displacer pipe = 1.2 cm, Beals number = 0.015 Frequency of oscillations is given by f, Hence

f

f

2

2g l 2

8

1.57 Hz

Time period = 1/f = 0.63 sec Pumping rate = Q

A

2gH  = 8.19 cm3/s = 8.19 × 10 6 m3/s

Power needed to pump water =   × Q × g × H

= 1000 × 8.19 × 10 6 × 9.8 × 0.15 = 0.012 W of pumping arm =

Power needed to pump water input power avaailable from engine

2.24 Experiments Experiments were done to find how different factors affect the work output of the engine. Results from various experiments are discussed in the following sections.

Devices Used a. Thermocouple – it is based on the thermo electric or Seebeck effect which states that a voltage is generated between junctions of two different metals at different temperatures. This voltage is proportional to the temperature difference. b. Manometer – these are direct reading devices which can be used for leak detection, flow measurement, and process monitoring. They are very simple devices and no calibrations are needed. The readings have accuracy of 0.5 m.

Liquid Piston Engines 73

Metal A

+ eAB –

Metal B eAB = seeback voltage

Figure 2.64 A thermocouple.

Figure 2.65 Manometer.

Figure 2.66 Experimental setup for finding pressure and temperature.

1. To study the variation of pressure and temperature of air column with time. Procedure – balance the level of manometer with knob, so that air bubble is at centre, light up the flame and connect

74 Liquid Piston Engines one end of manometer with the gap provided in the hot air column. Note the readings at various time intervals. Plug one end of thermocouple with the opening and other to the meter and note readings at specified time intervals.

2.25 Results Discussion – temperature of air in the engine was found to increase with time as it gains more and more heat from the burning fuel. Pressure was found to fluctuate with time as it moves back and forth from the hot side Table 2.6 Variation of pressure and temperature of air with time. Reading in mm of Hg

Pressure in Bar

Temperature (K)

Time in seconds

730

0.96

296

0

988

1.3

298

300

912

1.2

300

320

1216

1.6

305

340

912

1.2

306

360

1.8

308

380

1

310

400

1444

1.9

311

420

745

0.98

312

440

Temprature K

1368 760

320 315 310 305 300 295 290 0

300

320

340

360 380 Time (s)

Figure 2.67 Variation of temperature with time.

400 420

440

Liquid Piston Engines 75 towards the cold side alternately through the connecting air column. Peak pressure was found to be around 1400 mm of Hg, whereas the peak temperature was found to be around 39 °C, indicating poor heat transfer to the working gas (air).In order to reduce heat losses the connecting column was covered with an insulation covering of PTFE tape. Further, in order to improve the heat transfer rate, bigger connections can be used so that more mass of air is able to gain heat from the burning fuel.

Pressure mm of Hg

Calculation of length of stroke Calculation of stroke of water column was difficult due to quick oscillations; however, it was theoretically found using ideal gas laws and observing temperature and pressure at certain time intervals using manometer, stop watch and thermocouple.

1400 1300 1200 1100 1000 900 800 700 0

300

320

340

360 380 Time (s)

400 420

Figure 2.68 Variation of pressure with time.

S

Hot

Case 1: No heat

Figure 2.69 Osmosis and reverse osmosis.

Case 2: Heat is applied

440

76 Liquid Piston Engines Table 2.7 Variation of stroke length with time. P1 mm of Hg

V1 (cm3)

T1 (K)

P2 mm of Hg

T2 (K)

V2 (cm3)

V1–V2 (cm3)

S (cm) in cm

Time (s) sec

733

22.6

296

988

298

16.8

5.8

2.56

300

988

16.8

298

912

300

18.32

1.52

0.67

320

912

18.32

300

1216

305

13.96

4.36

1.9

340

1216

13.96

305

912

306

18.68

4.72

2.085

360

912

18.68

306

1368

308

12.53

6.15

2.7

380

1368

12.53

308

760

310

22.7

10.17

4.5

400

760

22.7

310

1444

311

11.99

10.71

4.7

420

1444

11.99

311

745

312

23.31

11.32

4.98

440

According to gas law

PV 1 1 T1 Vd

V1

V2

P2V2 T2 2

D2S 4

Where S is length of stroke 2) To study the variation of power output with time

Power

Bn Pf Vt

0.015 1.57 Vt P 0.023 Vt P

[20]

Discussion – the efficiency of the device was found to be in the order of 2–6%, which is very low, due to various poor heat transfer, leakage, viscous and frictional losses. Some measures to improve the efficiency of the system are discussed in the sections below.

2.26 Comparison Within Existing Commercial Devices University of Witwatersrand has investigated a design having a capacity to pump 100 gallons/hour of water. The Bell corparation has also made a

Liquid Piston Engines 77 Table 2.8 Variation of efficiency of pumping column with time. Pressure (Bar)

Time in seconds

Vd (cm3)

Vt (cm3)

Power (W)

0.96

5.8

8.2

0.185

6.4%

300

1.3

1.52

2.14

0.069

2.2%

320

1.6

4.36

6.16

0.22

5.4%

340

1.8

4.72

6.6

0.27

4.3%

360

1.9

6.15

8.69

0.37

3.15%

380

of pump

10.17

14.38

0.56

2.13%

400

10.71

15.14

0.41

2.8%

420

Efficiency (%)

1.7 1.2

6 5 4 3 2 1 0

300

320

340

360 380 Time (s)

400 420

440

Figure 2.70 Variation of efficiency of pumping column with time.

Power (W)

0.6 0.5 0.4 0.3 0.2 0.1 0

300

320

340 360 Time (s)

Figure 2.71 Variation of power output with time.

380

400

420

440

78 Liquid Piston Engines model capable of pumping 30 gallons/hour of water with an efficiency of 18% and head of 4 feet. A working liquid piston fluidyne pump has been developed by Dr. Tom Smith and Dr. Christos Markides at the University of Cambridge. This device is able to pump about 400 litres of water per hour to 1M head when 600 W of heat input is supplied to it [23]. Some other performance parameters of existing fluidyne pumps are presented below [21]: The designed model shows a maximum efficiency close to that achieved practically by existing models. Some more methods to increase the efficiency are discussed in the next section.

2.27 Improvements To increase the engine efficiency some of the following improvements can be made in the current design: 1. Use of bigger diameter displacer tubes – it ensures a greater amount of air flowing between cold and hot side. This can lead to larger amplitude of oscillations due to higher pressure, but smaller compression ratio whereas smaller tubing results in a larger compression ratio. 2. Use of regenerator – The regenerator acts as a thermal sink, releasing and absorbing heat at various stages, hence increasing the efficiency of the engine. The most common method of heat storage is to obstruct the flow of working fluid by use of metallic mesh, porous material, array of tubes, but this may cause flow losses. 3. Better heat exchange – in order to enhance the heat exchange at the hot end, resistance heating can be used instead of burning fuel along with fins for greater heat transfer. Table 2.9 Performance parameters of existing engines. Reference

Flow rate (gallons/hr)

Head (ft)

Efficiency

West (1970)

100

5.3

3.5%

Goldberg (1977)

9.5

2

1%

Mosby (1978)

5.9

1

8%

Reader (1981)

2

3.3

5.2%

Pandey (1985)

2500

10

7%

Current design

5.8 (max)

0.5

6.4% (max)

Liquid Piston Engines 79

2.28

Future Scope

Parabolic mirrors can be used to focus solar energy for operation of a liquid piston engine. Such a device is shown below. Many commercial setups have been built, tested and operated by the team of Dr. Tom Smith and Dr. Markides at the engineering department Sun

Parabolic reflector

Target Pumping column Hot side

Cold side

Reserviour

Figure 2.72 Commercial setups for solar liquid piston engine.

Table 2.10 Comparison of irrigation costs for various methods. Mode of irrigation

Cost or irrigation per hectare per day

Electric pumps

£0.34–£0.55

Diesel pumps

£0.29–£0.17

Photovoltaic pumps

£1.27–£4.07

Liquid piston pumps

£0.29–£1.07

Table 2.11 Comparison of pumping costs and efficiency of various methods. Mode of irrigation

Efficiency

Pumping cost per unit power output

Photovoltaic pumps

20–40%

£3.35–£10.7

Liquid piston pumps

2–4%

£1.50–£3

Table 2.12 Comparison of emissions. Mode of irrigation

CO2 emissions per hectare per day (kg)

Diesel pumps

2.3–3.6

Solar P-V pumps

0.8–1.3

80 Liquid Piston Engines of the University of Cambridge. Typical data for cost, efficiency and CO2 emissions is discussed here, assuming a lift head of 10 m [23]. Going by the reliable data obtained, the future of this technology seems to be bright, and to tap the economic potential there are several organizations currently involved in research in the field of Stirling and liquid piston engines. Some of these are listed below [16]: 1. STM co-operation – holders of various Stirling engine patents and developed a 40 KW engine for use in GE hybrid vehicle. 2. Sun power – founded by Beale, pioneers in development of cryogenic coolers of capacity 35 W–7.5 KW. 3. Infinia – developers of 1 KW free-piston engines and cryogenic coolers. 4. SES – makers of large parabolic dish operating solar power stations of 850 Mw capacities. 5. Thermo fluidics limited – formed in 2006 by Dr Tom Smith of the University of Cambridge and supported by carbon trust, is developing such pumping devices for use in Brazil, India, and Ethiopia.

2.29 Conclusion The most common application of the liquid piston system is in irrigation pumping. Other important applications include drainage pumping, failsafe cooling of nuclear reactors, cooling of combustion engine with waste heat, and circulation of water in remote areas without use of electricity. These devices are simple to construct and can be used easily for demonstrations and teaching purposes.

2.30 Numerical Analysis During charging or suction phase

ZI + q/c = Vin Differentiating bothsides with respect to time we have:

ZC(δI/δt) + (δq/δt) = 0

Liquid Piston Engines 81

ZC(δ2q/δ2t) + (δq/δt) = 0 ZC(δ2V /δ2t) + ( δ2V /δ2t) = 0 Applying Laplace on both sides we have:

ZC(D2V ) + D V = 0 ZC[S2F(S)] + [SF(S) F (S) V (0)

V (0)] = 0 1

1/S+ S

1 ZC

Taking inverse Laplace we have

V (t ) V (0) V (t ) V (0)

t

1 e ZC t

1 e

Where = RC When t =

V (t) = V (0)[1   e 1] = 0.63 V (0) b) During the discharging phase we have ZI + q/C = 0 Z(δq/δt) + q/C = 0

Volume flow rate

0.63 –v (0)

v– (0) Time

Figure 2.73 Suction Phase.

T=

82 Liquid Piston Engines

Z

V t

V /C

0

Taking laplace

ZC[SF(S) F (S)

Volume flow rate –v (0)

0.36 –v (0)

Time period T

Figure 2.74 Discharge phase.

Volume flow rate

0.5 –v (0)

–v (0)

0.693 Time

Figure 2.75 Total flow.

V (0) + F(S)] = 0 V (0) 1 S ZC

Liquid Piston Engines 83

V (t ) V (0)

t

e ZC

When t =

V (t) = V (0) * 0.36 When suction flow = discharge flow rate we have,

V (t ) V (0)

t

1 e ZC

V (0)

t

e ZC

t/ = 0.693 hence t = 0.693 V (t) = V (0) * 0.5

References and Bibliography 1. http://liquidpiston.com/technology/faq-2/ 2. http://liquidpiston.com/technology/how-it-works/ 3. http : / / w w w. p opu l ar m e ch an i c s . c om / m i l it ar y / re s e arch / a 1 5 2 3 3 / liquidpiston-darpa-contract/ 4. http://www.utsc.utoronto.ca/~quick/PHYA10S/LectureNotes/LN-18.pdf 5. http://en.wikipedia.org/wiki/Ground_ 6. waterhttp://en.wikipedia.org/wiki/Water_resources#Sources_of_fresh_water 7. http://www.unesco.org/water/wwap/wwdr/wwdr3/pdf/WWDR3_Water_ in_a_Changing_World.pdf 8. http://www.engin.swarthmore.edu/academics/courses/e90/2005_6/ E90Reports/FK_AO_Final.pdf 9. http://en.wikipedia.org/wiki/Ground_water 10. http://en.wikipedia.org/wiki/Earth’s_energy_budget 11. http://www.humphreypump.co.uk/operating%20cycle.htm 12. Different pumps for irrigation systems, James Dee, FARM Note: 332, Department of Agriculture, January 2009. 13. http://www.biologymad.com/NervousSystem/nerveimpulses.html 14. http://www.cazadero.org 15. http://www.webmd.com 16. http://www.aquatechnology.net/reverseosmosistheory.html 17. http://www.inference.phy.cam.ac.uk/sustainable/refs/solar/Stirling.PDF 18. http://ir.canterbury.ac.nz/bitstream/10092/2916/1/thesis_fulltext.PDF

84 Liquid Piston Engines 19. 20. 21. 22. 23. 24. 25. 25. 26. 27. 28.

29. 30. 31. 32. 33. 34. 35. 36. 37. 38. 39. 40.

41.

http://www.mpoweruk.com/heat_engines.html http://www.animatedengines.com/vstirling.shtml http://www.exergy.se/goran/hig/re/07/stirling.pdf C. D. West, Liquid Piston Stirling Engines, Van Nostrand Reinhold, New York, 1983. http://www.omega.com/temperature/z/pdf/z021-032.pdf http://www.thermofluidics.com http://www.mathpros.com/papers/thermodynamics/Integral_Analysis_for_ Thermo-Fluid_Applications.pdf https://www3.nd.edu/~msen/Teaching/DirStudies/Engines.pdf Qianfan Xin, Diesel Engine System Design (Woodhead Publishing in Mechanical Engineering) 1st Edition,ISBN-13: 978-1845697150. Gupta, Aman, and Sunny Narayan. “Electrical Analogy of Liquid Piston Stirling Engines.” Hidraulica 2, 58, 2016. Narayan, Sunny, and Vikas Gupta. “OVERVIEW OF WORKING OF STRILING ENGINES.” Journal of Engineering Studies and Research 21.4, 45, 2015. Gupta, Aman, and Sunny Narayan. “A Review of Heat Engines.” Hidraulica 1, 67, 2016. Narayan, S. “Analysis of noise emitted from diesel engines.” Journal of Physics: Conference Series. Vol. 662. No. 1. IOP Publishing, 2015. Gupta, Aman, and Sunny Narayan. “Effects of turbo charging of spark ignition engines.” Hidraulica 4, 62, 2015. Narayan, Sunny. “Designing of liquid piston fluidyne engines.” Hidraulica 2, 18, 2015. Narayan, Sunny.  Effects of Various Parameters on Piston Secondary Motion. No. 2015-01-0079. SAE Technical Paper, 2015. Narayan, Sunny. “Analysis of Noise Radiated from Common Rail Diesel Engine.” Tehnički glasnik 8.3, 210–213, 2014. Narayan, Sunny. “TIME-FREQUENCY ANALYSIS OF DIESEL ENGINE NOISE.” Acta Technica Corviniensis-Bulletin of Engineering 7.3, 133, 2014. Narayan, Sunny. “Wavelet Analysis of Diesel Engine Noise.”  Journal of Engineering and Applied Sciences 8.8, 255–259, 2013. Narayan, Sunny, and Vikas Gupta. “Motion analysis of liquid piston engines.”Journal of Engineering Studies and Research 21.2, 71, 2015. Narayan, Sunny. Modeling of Noise Radiated from Engines. No. 2015-01-0107. SAE Technical Paper, 2015. Narayan, Sunny, Aman Gupta, and Ranjeet Rana. “Performance analysis of liquid piston fluidyne systems.” Mechanical Testing and Diagnosis 5.2, 12, 2015. Gupta, Vikas, Sahil Sharma, and Sunny Narayan. “REVIEW OF WORKING OF STIRLING ENGINES.” Acta Technica Corviniensis-Bulletin of Engineering 9.1, 55, 2016. Singh, Amar, Shubham Bharadwaj, and Sunny Narayan. “Review of how aero engines work.” Tehnički glasnik 9.4, 381–387, 2015.

Liquid Piston Engines 85 42. Narayan, Sunny. “A review of diesel engine acoustics.” FME Transactions 42.2, 150–154, 2014. 43. Narayan, S. “Noise Optimization in Diesel Engines.”  Journal of Engineering Science and Technology Review 7.1, 37–40, 2014. 44. Gupta, Vikas, Sahil Sharma, and Sunny Narayan. “REVIEW OF WORKING OF STIRLING ENGINES.” Acta Technica Corviniensis-Bulletin of Engineering 9.1, 55, 2016. 45. Singh, Amar, Shubham Bharadwaj, and Sunny Narayan. “Prikaz rada motora zrakoplova.” Tehnički glasnik 9.4, 381–387, 2015. 46. Klangpraphan, Praphan, Pisit Yongyingsakthavorn, and T. Soontornchainacksaeng. “Development of the Solar Liquid-Piston Stirling Engine: Parameters Affecting the Efficiency of the Engine.” 2013. 47. Winkelmann, Anna, and Eric J. Barth. “Second Generation Controlled Stirling Device.”

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

3 Customer Satisfaction Issues

3.1 Durability Issues Durability is an important aspect as customers need shorter service intervals and longer engine lifetime. Increase in mechanical and thermal loading causes higher engine power density, lower emissions, and longer lifetime which are greater issues in durability. Structural issues are also important in view of system performance. These issues may be classified as: fracture based performance which includes rupture and cracking. buckling under load. Thermal deformations, fatigue, creep, oxidation, and corrosion. Cavitation issues. Wear and oil degradation. EGR cooler fouling, boiling, and corrosion. Deposits leading to fouling. Hydrogen embrittlement.

87

88 Liquid Piston Engines Buckling is failure of parts due to higher compressive stresses. Abrupt rupturing of bubbles formed causes Cavitation. Fouling and deposition may occur due to coking and deposition of soot. As Hydrogen is released embrittlement may also be caused. Shocks and deformation resistance is lessened. Most of the these issues occur in EGR system, turbocharger, injection systems, skirt assembly, cylinder head exhaust manifold, and valve train. Injector choking is also a major cause of concern due to carbon deposition. Engine needs frequent overhauling due to excessive wear, oil consumption or blow-by. The following engine parameters are dependent upon engine durability: Torque produced. Speed of skirt. Heat flux in system. Temperature of compressor and compression ratio. Charge inside air cooler and EGR outlet gas temperatures. Load in values. Piston slap. Engine load.

3.2 Testing of Engines In order to validate the design, testing is needed. This is done in order to see effects of mechanical and thermal loads and fatigue. Some of these tests include full-load test, over-fueling test and load cycle tests. Evaluation of the sealing capacity of the cylinder head gasket by is an important aspect of fatigue testing in which exposing of components is done to high thermal gradients. Thermal shocking is done by changing temperature gradient. Thermal loading may reach its peak at peak torque. Cam stress can maximize with an increase in cranking speed. In field tests includes evaluation by changing temperature, humidity, altitude, speed, fuel and oil consumption, speed and torque, exhaust manifold pressure, turbine outlet pressure, and fluid temperatures.

3.3 Design of Systems Light and durable design is the major objective of design approach that includes increase in stiffness, lowering temperature, higher strains.

Customer Satisfaction Issues 89 Engine structural parameters that need evaluation include: Deflection, plasticity, stress, and strain. Fatigue. Multi-body dynamic vibration and modal analysis. Wear and Cavitation. Transient structural analysis. Fluid-structure interaction. Finite element analysis may be used to analyze complex geometry to identify stress concentrations areas.

3.4 Systems Durability Durability is evaluated using either deterministic approach or probabilistic approach. A FEA simulation software like GT-POWER can analyze the temperature distribution.

References and Bibliography 1. Qianfan Xin, Diesel Engine System Design (Woodhead Publishing in Mechanical Engineering) 1st Edition, ISBN-13: 978-1845697150.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

4 Lubrication Dynamics

4.1

Background

Surface topography may refer to both the shape and the roughness of surface which includes the waviness and the asperity contacts. Topography also affects the oil film thickness and the lubrication regime formed. In the hydrodynamic or mixed lubrication type of lubrication, the influence of roughness of surface on film thickness and pressure distribution is not negligible. The roughness along transverse direction must increase the load bearing capacity and the oil film thickness whereas, the longitudinal one would cause reduction. Topography has an important effect on the load carrying capability. These interactions become more dominant as the thickness-to-roughness ratio decreases. Some of the models related to surface topography include the following: (1) Patir and Cheng (1979) average model, which takes into account the influence of surface roughness (2) Greenwood and Tripp (1971) asperity contact model. Topography also changes with engine operational conditions and changes due to wear. This also affects the friction power of all the sliding

91

92 Liquid Piston Engines surfaces. During break-down, some of the asperities are worn off leading to smoother surfaces. This reduces frictional forces. As the wear progresses, the engine friction reduces and then stabilizes. Taylor analyzed the effects of surface roughness on engine friction and wear. Zhu studied effects of surface roughness on cylinder liner, piston skirt and piston rings. Oil reduces friction, wear and acts as a coolant to removing the excessive heat, and impurities. Additives used include anti-wear agents, friction reducers, viscosity index improvers, detergents, anti-rust agents, and antifoam agents. Some effects of the lubricant on engine friction include cavitation, thermal effects, oil starvation and the changes in viscosity. The viscosity may be used for classification of oils as given by SAE J300 (2004). Grades of viscosity include the SAE5, SAE10, SAE30, SAE40, SAE45, and SAE50. Higher grades of oil mean higher viscosity. The multi grades of oils use viscosity index improvers in order to stabilize the viscosity. The 10W-30 means that the viscosity of the SAE10 at 15°. SAE grades of oil may help to reduce the friction at starting without problems of low viscosity and metal contacts. Newtonian fluids show a linear relationship among stress and shear rate. Under high pressures conditions, as viscosity falls and shear rate increases, the fluids show this nature. Increases of level of soot and dispersant of oil shows an increase in viscosity. Vogel’s equation shows relation among temperature and pressure of oil. The viscosity falls with increasing in temperature. The non-Newtonian can show effects for rough surfaces. Coy and Taylor studied the lubricant rheology, friction and wear. Frictional coefficient decreases with a fall in viscosity in hydro dynamic regime. So, the friction in piston skirt and bearings show lower friction, while the valve train shows an increase. Valve train friction can be reduced by using the oil with friction modifier. An optimum value of viscosity having minimum engine friction exists. Taylor showed the valve train friction losses are higher for in light-duty engines as compared to heavy-duty engines. As the valve train friction is dominant in the boundary lubrication, the modifier additives may be more effective for light-duty engines. Taylor and Kapadia found fuel consumption was less by 5 % by careful choice of lubricants. Reduction in viscosity changes the lubrication regime from hydrodynamic to boundary one. This may increase the wear and scuffing. So, there is a tradeoff between reduction of friction and durability.

Lubrication Dynamics 93

4.2

Friction Features

Engine friction can be analyzed by various methods like motoring and teardown method, indicator diagram, pressurized motoring method, Morse test, Willan’s Line method, torque method. Wakuri and Richardson have analyzed each method in details. As various load and speed vary various engine parts operate at different lubrication regimes. Friction losses changes in design parameters and operating conditions. Stribeck diagrams show effects of duty parameter on viscosity, loading, and sliding velocity. Lubrication regimes depend on load, speed, viscosity, state of break-in, etc. The piston skirt moves in a periodic sliding motion from TDC to BDC positions. At higher speeds during middle stroke, the skirt operates in hydrodynamic or elasto- hydrodynamic lubrication regime. During expansion and compression stroke pistons operate in mixed lubrication. An oil film is developed for surface separation. Near the dead center positions as the sliding velocity reaches zero, there is a probability of formation of mixed lubrication regime. Squeezing action may occur in piston rings. Forces acting on compression rings include the tension and gas pressure. Tension depends on the force needed to compress the ends of the ring. These rings operate in hydrodynamic regime at mid-stroke and in the mixed regime near the dead centers as the gas pressure is high and the viscosity is low. Under such condition, oil film breaks resulting in higher wear. These rings have a tapered face to help in scraping action. The oil ring used has thin rails in order to generate a thin oil film. The friction on the piston skirt and the rings can show a sudden rise and fall near dead center positions as skirt reverses direction which is impulsive and has frequencies that excite the crankshaft resulting in a knocking noise. This is known as ‘stick slip’ noise. Changes in Frictional coefficients may be suppressed using modifiers and coating materials which results in reduction of ‘stick slip’ noise. The bearings in engine have different load bearing capacities at varying speeds. The main bearing operates in the hydrodynamic or elasto hydrodynamic due to constant sliding velocity. The instantaneous force acting on small-end of connecting rod can be null. The resultant force on big end of the connecting rod as well as main bearing of crankshaft is greater than zero. The resultant force on cam shaft is never zero.

94 Liquid Piston Engines The main component of friction force acts between the cam and the follower. The loading on cam is due to spring force and the inertia load. The contact may be sliding or rolling. These contacts are of mixed lubrication or boundary lubrication regime due to high load which acts on a small area. The losses in pump are due to intermeshing friction, bearing friction, internal fluid friction. The torque is proportional to the third power of the gear diameter and speed. Poor lubrication has minimum effects on the friction as sliding velocity is low. The drag force at the mid-stroke, is the major contributor to the friction power.

4.3

Effects of Varying Speeds and Loads

In boundary regime friction depends on speed. In hydrodynamic one, lubrication these force increases with speed. The pumping torque is dependent on square of speed. The boundary lubrication regime dominates in the valve train where the load falls with increase in speed. So as the speed falls, the boundary lubrication friction also increases. As the load or fueling changes gas loading, wall temperature, clearances in the cylinder also rise. Higher loading causes a rise in friction in expansion stroke and the compression stroke near the TDC. The coolant controls the lubricant temperature with higher values at higher load. The higher temperatures cause a fall in viscosity and the viscous force. The clearances due to thermal expansion increase the friction force. During the mid-stroke higher temperatures effects are predominant due to higher velocity of skirt. Forces in bearings, torque and friction in accessories increase with loading. Richardson found that the piston ring friction during motoring condition was lower compared to firing. He found that cylinder friction during firing case was 0–20% higher compared to a motored.

4.4 Friction Reduction Design measures can be adopted to reduce friction in engine components depending upon lubrication regimes. Various factors that must be taken into account for reduction of friction are as follows: Durability and reliability Minimize the load on interface.

Lubrication Dynamics 95 For boundary lubrication regime reducing the normal load acting on the contact is effective. Reduction of friction coefficient For hydrodynamic lubrication regime reducing of lubricated area or the viscosity Use of multi grade oils with higher viscosity indexes For hydrodynamic lubrication regime an increase in the clearance may reduce friction. To increase the oil film thickness profile of skirt is important. For mixed or boundary lubrication type of regime can be changed by changing clearance or profile Use of special coatings on the rings and the skirt can help to reduce friction.

4.5 Piston-Assembly Dynamics From the multi-body dynamics modeling effects, the piston assembly includes study of the piston skirt, piston rings, the piston pin and the connecting rod as defined by SAE J2612. The skirt must have lower distortion, friction and wear, noise, higher cooling and lubrication. Piston rings act as a seal towards gas formed preventing transfer of heat. Piston slap is due to impacting of skirt with liner seen most prevalent during warm-up. This causes vibrations of surface which become dominant speed increases. Design of piston skirt, oil film thickness formed, and slap kinetic energy must be optimized to reduce skirt noise. During an engine cycle, piston moves laterally form thrust force to antithrust side, due to skirt to – bore gap. The reaction force due to connecting rod resists the combined gas and inertia forces. A small tilting of piston pin is also seen due to the moments acting on it. These motions cause a change in lubrication from mixed at the dead center positions to hydrodynamic one at the mid-stroke. Research in piston dynamics during 1960s began without taking into account lubrication to the more complex multi-body dynamics. These motions not only affect the piston slap noise, but also ring operation and wear. These need to be reduced to control blow-by and oil consumption. The piston motions are also dependent on bore polishing. The piston-assembly friction can account for up to 40–55% of the total engine losses, with the contribute of skirt, rings and connecting rod about 15–20%, 15–20%, and the connecting rod 10–15%.

96 Liquid Piston Engines

4.6 Reynolds Equation for Lubrication Oil Pressure Tribology of lubricating oil plays an important role in mechanical losses in skirt assembly of piston in internal combustion engines. About 3–5% of total energy in the engine is dissipated from piston skirt assembly [1]. Figure 4.1 shows a typical breakdown of mechanical losses for a typical diesel engine, wherein it is clear that piston assembly accounts for about 20–30% [24]. Nature of lubrication is hydrodynamic at mid strokes, where sufficient oil film thickness separates skirt from cylinder wall. At dead center positions, oil film thickness is much thinner and surface asperities on skirt and liner make contact. In 1886 Osborne Reynolds proved that hydrodynamic pressure generated can separate two sliding surfaces. For an incompressible fluid with constant density the Reynolds equation gets modified as:

h3 P h3 P x 12 x z 12 z 1 (U 2 U1 )h 1 (W2 W1 )h (V2 V1 ) z 2 x 2

(4.1)

In this relationship, the left hand side term is called pressure term, whereas U U the right hand side terms are called source terms. The terms of & x z Mechanical losses 15%

30% 25% Exhaust losses

Cylinder cooling losses

30%

Figure 4.1 Break up of total dissipation of fuel energy.

Brake power losses

Lubrication Dynamics 97 h h , are x z known as Wedging action. The velocity difference term (V1–V2) is known as squeezing action as shown in Figure 4.2. Assuming that lubrication oil used is Newtonian fluid, flow is incompressible, value of viscosity is constant and neglecting inertial effects, slip, angle of inclination, pressure gradient and stretching action Reynolds equation can be simplified as: in the above relationship are known as stretching action and

p x

h3

x

z

p z

h3

(U 2 U1 )h 12 x

6

h z

(4.2)

In order to estimate the oil pressure distribution, the above given Reynolds equation needs to be solved. One way to do this is to neglect the pressure distribution in one direction and consider it in other direction as shown in Figure 4.3. The numerical method adopted uses non-dimensional U , x h3 x 12

W z P x

V1 – V2 = h t

Stretching action

+

h3 x 12

P z

=

1 (U2 – U1)h x 2

Wedge action (inclinded surfaces)

+ (V1 – V2) + 1 2

(W2 – W1)h x

Squeeze action (bearing surfaces move perpendicular to each other)

h h , x z

Figure 4.2 Interpretation of Reynolds equation. P P

X X

Z Dimension in x-dir is much larger than dimension in z-dir

Z Dimension in z-dir is much larger than dimension in x-dir

Figure 4.3 Variation of pressure along various directions.

98 Liquid Piston Engines analysis of space coordinates and oil pressure developed as given by equation 4.3. Non-dimensionalization can be obtained as follows:

x X y y Y h h C pC 3 p 6 UX 2 tU t C x

(4.3)

Solution of Reynolds equation depends upon geometry of surfaces. Two most common approximations are short and long surfaces. Substituting these non-dimensional values the new equation gets modified as:

x

h

3

X2 p h3 2 Z z z

p x

C h X x

(4.4)

In order to solve this equation, finite element analysis method can be used for which the mating surface needs to be analyzed into number of nodes as shown in Figure 4.4. A mesh was made so that nodes on lubrication zone of skirt correlates with nodes used in finite element analysis to analyze the pressure distribution.

Z

Pi,j+1 Pi–1,j

Pi,j

Pi+1,j

Pi,j–1

X

Figure 4.4 Nodal representation of surface.

Lubrication Dynamics 99 Various gradient terms of Reynolds relationship can be solved using Taylors approximation which yields following results:

h

3

h3

hi3, j

p x

0. 5

pi , j

1

hi3, j

pi

1, j ( hi3, j

0.5

hi3, j

0.5 ) pi , j

0.5 i 1, j ( h 3 i,j

0.5

hi3, j

0.5 ) pi , j

0. 5

x hi3, j

p z

0. 5

pi , j

1

hi3, j

p

z h x

hi

2

1, j hi

2

(4.5)

(4.6)

(4.7)

1, j

2 x

Substituting these relationships and rearranging them we have:

Pi,j = Ai,j Pi,j+1 + Bi,j Pi,j 1 + CI,j Pi+1,j + Di,j pi

1,j

+ Ei,j

(4.8)

As most of values of nodal pressure (Pi,j) are unknown hence iterative loop must be employed to get values of fluid pressure. The process of iterations must be repeated till convergence is satisfied. i.e. n

m

i 1

j 1 i, j

p

iteration k n m i 1

j 1

pi , j

n

m

i 1

j 1 i, j

p

iteration k 1

(4.9)

iteration k

A Matlab code has been developed to analyze the lubrication behavior of oil between piston skirt and liner considering its motion analogous to motion of a journal in bearing. The oil film thickness (h) at a given crank angle (θ) can be calculated from values of nominal clearance (c) and piston eccentricity (e) as:

h = c(1 + e cos )

(4.10)

During analysis the piston is assumed to be as a short bearing and circumferential pressure gradient is neglected as compared to axial pressure. Using these assumptions the Reynolds equation discussed above gets modified as:

z

h3

p z

6

h x

(4.11)

100 Liquid Piston Engines Using boundary conditions of P

, z=

L 2

0, the closed form of

pressure distribution p can be expressed as:

p

3 c2

x

2

L2 4 (1

sin cos )3

(4.12)

The variations in density of lubricant can be expressed in terms of generated oil pressure (P) and density at mean liner temperature ( 0) as [29]:

0

0.6 10 9 P 1 1.7 10 9 P

1

(4.13)

The variations of oil viscosity can be expressed in terms of relationship as given by Reoland [30]: 0

e

(4.14)

Where

1 [log P

0

9.67] 0

138 138

s0

1

P 1.988 108

Z

1

S0 and Z are constants which depend upon temperature and pressure, θ is temperature, θ0 is bulk oil temperature. Hydrodynamic oil pressure distribution was analyzed on piston skirt plane for each 90° crank angle rotation at speed of 2000 RPM and nominal skirt-liner gap (0.05 mm) as shown in Figures 4.5–4.12. During intake stroke the peak value of pressures are close to top part of skirt showing gradual slope. At 90° crank angle the piston is at mid stroke and peak value of oil pressure is witnessed almost at center of skirt. At 180° crank angle towards the end of intake stroke peak pressures are again observed at skirt midpoint with slopes slightly towards the right side. During mid compression stroke the peak hydrodynamic pressure starts to shift towards bottom of skirt at 270° crank angle. At 360° crank angle towards end of compression stroke the pressures shift towards bottom of skirt. During middle of power stroke at 450° crank angle oil pressures

Lubrication Dynamics 101

Oil pressure-Pa

5 4 3 2 1

0 80 60 40 Skirt width-mm 20

0 0

20

40

60

80

Skirt length-mm

Figure 4.5 Oil pressure distribution (90° crank angle).

Oil pressure-Pa

5 4 3 2 1 0 80 60 40 Skirt width-mm 20

0 0

20

40

60

80

Skirt length-mm

Figure 4.6 Oil pressure distribution (180° crank angle).

Oil pressure-Pa

5 4 3 2 1 0 80 60 40 Skirt width-mm 20

0 0

20

40

60

80

Skirt length-mm

Figure 4.7 Oil pressure distribution (270° crank angle).

rise from mid part of skirt towards bottom part. At 540° crank angle same trends in oil pressures were observed. During exhaust stroke peak pressures are generated throughout skirt length having slopes shifting towards bottom part of skirt.

102 Liquid Piston Engines 5 Oil pressure-Pa

4 3 2 1

0 80

60

40

20 Skirt width-mm

0

0

20

60 40 Skirt length-mm

80

Figure 4.8 Oil pressure distribution (360° crank angle).

Oil pressure-Pa

5 4 3 2 1

0 80 60 40 Skirt width-mm

20 0 0

20

40

60

80

Skirt length-mm

Figure 4.9 Oil pressure distribution (450° crank angle).

Oil pressure-Pa

5 4 3 2 1 0 80

60

40

Skirt width-mm

20

0 0

40 20 Skirt length-mm

60

80

Figure 4.10 Oil pressure distribution (540° crank angle).

4.7 Introduction This part describes a mathematical model to simulate the piston secondary motion which is the cause of piston slap noise. Piston slap is caused when forces in connecting rod change direction. The piston impact against

Lubrication Dynamics 103

Oil pressure-Pa

5 4 3 2 1 0 80 60 40 Skirt width-mm

20 0 0

20

40

60

80

Skirt length-mm

Figure 4.11 Oil pressure distribution (630° crank angle).

Oil pressure-Pa

5 4 3 2 1 0 80

60

40 20 Skirt width-mm

0

0

40

20

60

80

Skirt length-mm

Figure 4.12 Oil pressure distribution (720° crank angle).

cylinder liner is a major source of noise and cause wear of liner. Major factors that affect piston slap are [26]: a. b. c. d. e. f. g. h. i. j.

Cylinder Bore Temperature Lubrication Oil Film Thickness Oil Viscosity Engine speed Skirt Profile Skirt Roughness Skirt Waviness Skirt Size Wrist pin offset Piston-Liner gap

Motion of Crankshaft picks up lubrication oil from sump. This oil is then transported along cylinder bore due to motion of piston, piston rings and gravity. Oil is consumed either inside combustion chamber or it returns to sump or is consumed by blow by gases. Piston slap takes place due to

104 Liquid Piston Engines

Hydrodynamic

Mixed Coefficient of friction

Boundary 0.1 Piston

0.01

Valve train

Bearings

0.001 Duty parameter

Figure 4.13 Stribeck lubrication curve.

changes in direction of piston side forces and occurs mainly near top dead center (TDC) position. According to Stibeck curve, the lubrication can be classified into three major types: boundary, hydrodynamic and hydrostatic. In the boundary lubrication zone, the asperities in mating parts come into contact whereas in hydrodynamic zone there is no direct contact and the film of lubricant separates the mating surfaces. The function of piston rings of skirt assembly is to seal pressure in combustion chamber and prevent leakage of oil from crankcase into combustion chamber. Type of lubrication of oil changes with operational conditions of engine. As piston reaches dead center positions, the speed of piston approaches zero and hence boundary lubrication dominates. At mid strokes, where piston speed is at its maximum value, the type of lubrication changes to a hydrodynamic one.

4.8

Background

Increasing demand for noise, vibration and harness comfort levels have led to detailed study of piston dynamics motion as skirt-piston contact plays an important roles in frictional losses in engine [41]. Piston secondary motion is a key concept to understand slapping motion phenomenon. Various forces as well as moments are responsible for lateral displacement motion of piston as well as its rotatory motion about pin axis as depicted in Figure 4.14. Piston impacts occur on either side of liner which are identified as thrust side (TS) and anti thrust side (ATS). These contacting motions cause vibrations in liner which are transmitted from engine surface. In general three major approaches have been identified to study slapping motion. The first one includes study of piston secondary motion without taking into consideration oil lubrication effects and piston rotatory effects. This approach is

Lubrication Dynamics 105 Thrust side L

R

et Rotation

Anti thrust side

Translation eb Lateral motion

Figure 4.14 Piston secondary motion.

known as static method. In another method piston side force is found using solution of lubrication equations as represented by following equation [42]. Oil film thickness is third parameter to study slapping motion. PistonLiner contact occurs when this thickness is minimum towards both thrust as well as anti thrust side [43]. There are two most common methods of studying slapping motion. The first one involves calculation of maximum energy transfer to cylinder wall using force variations and oil film thickness results. The other one involves calculation of rate of minimum oil film thickness. As the slope of this parameter changes, the squeeze action is initiated indicating the instance of piston slap. Another method includes study of initiation of squeezing action and occurrence of minimum oil thickness. This method is known as angular duration method. Up to 16 instances of slapping motion can be identified in a engine cycle, however only 6–10 are practically observed [44]. In order to validate positions and instances of slap, data was measured using accelerometers using various engine operating conditions. Previous works include that of Richmond to capture vibration data [25]. However this data included mixture of piston slap as well as combustion noise since both occur in vicinity of TDC position. Pruvost used spectro filters to separate the two noise sources [26]. Liu and Randall used blind source separation methods to achieve effective separation [27]. Chen has used concept of pseudo angular acceleration to study phase and frequency variations of slapping noise [28].

4.9 Occurrence of Piston Slap Events Figure 4.14 shows the free body diagram of piston in which various forces acting on body include frictional force between liner and skirt (Ff), gas

106 Liquid Piston Engines Fg

My”

b Fh

Mx”

a

J ” Ff dCOG

FL

dP

Figure 4.15 Piston free body.

force (Fg) and oil reaction force (Fh) acting on piston of mass m and pin of mass mp. Piston of moment of inertia J tilts by an angle having pin offset dp and center of gravity offset dCOG. Connecting rod exerts a lateral force Ft on bore wall given by following equation:

Ft = mpX

p

Fhyd = FL sin

[FG

mpYP

Ff] tan

(4.15)

Hydrodynamic reaction force is given by solution of Reynolds equation [29]. Frictional forces can be neglected since they contribute less [30]. Applying moment balance equation we have:

(J + M[dCOG2

(a

b)2])β + Mx (a b) My dCOG Tf + TG + Fhyh

(4.16)

Changes in direction of lateral forces is an important way to diagnose piston slaps. This occurs when Tan ϕ = 0 which is instances of TDC(ϕ = 0) and BDC(ϕ = k , k 0, 1, 2 …) positions. Exact values of ϕ can be determined using following relation in terms of connecting rod length (l), crank radius (r), Crank case offset (C) as:

sin

1

rcos

C l

(4.17)

When gas force is greater than inertial force, slap occurs at thrust side of liner and vice versa.

Lubrication Dynamics 107 14

Dimensionless force

12 G1

10

G2

8

G3

6 4 2 0 sy

–2 –180

–90

+

py

0 Crank angle ( )

90

180

Figure 4.16 Piston force distribution.

Figure 4.16 shows graphical representation of balance between various forces acting on piston. The coincidence points between dimensionless gas forces ( G) and total inertial forces ( sy + py) indicates the instances of slap. As evident from figure number of instances of slap increase with increase of speed. Another approach takes into account elastic deformations of piston and oil film thickness variations to find the lateral forces [29]. Thrust side contact is indicated by negative values of force whereas anti thrust values are indicated by positive values. Hence if lateral force changes into value from negative to positive, slap is considered at anti thrust side and vice versa. Diagnose of slap can be done by consideration of minimum oil film thickness both towards thrust as well as Anti thrust side [30]. Tilting motion of piston causes squeeze action of oil film which indicates pistonliner contact. Next method includes consideration of squeeze velocity. In Reynolds equation the term indicates squeeze velocity of lubricant. Hence change in its sign indicates piston slap instance. The product of oil film force and film displacement yields the energy transferred to liner. When lubricant squeezes, the energy is transferred to cylinder liner. Due to rise in lubrication pressure the squeezing action slows down. The position of maximum energy transfer is assumed to be that of slapping motion. Figure 4.17 shows plots of steady state and transient lateral forces acting on liner for engine running at 2000 RPM. Whenever the side thrust force changes its direction crossing the zero mark, a slap event is expected to occur as depicted by circles in the above figure. Negative values of force depicts contact with thrust side whereas

108 Liquid Piston Engines

Lateral force (N)

1

×105 Steady state Transient

0

–1

0

90

180

270

360 450 Crank angle

540

630

720

Figure 4.17 Piston side thrust force (3000 RPM).

Table 4.1 Summary of slap events (Lateral force method). Slap

Steady state

Transient state

1

70°

60°

2

170°

240°

3

240°

340°

4

320°

390°

5

340°

470°

6

400°

550°

7

470°

720°

8

560°

–

9

720°

–

positive values indicate contact with anti thrust side. Hence when the value of these lateral forces change from negative to positive one, slap is expected to occur at ATS and vice versa. Table 4.1 gives summary of slap events as predicted by the above mentioned methods. Lateral forces are expected to reach their maximum value in vicinity of TDC position when contributions due to gas force is at peak. However there may be some deviations due to crank shaft offset. The concept of oil film thickness has been examined next in figure assuming that skirt-liner gap is fully flooded with lubricant. Four events were identified both at thrust as well as anti thrust side. Locations of instances of slapping motion was found to be [220°, 300°, 420°, 630°] for anti thrust side and [230°, 420°, 540°, 620°] for thrust side. During the intake stroke the film thickness falls and then again rises during compression stroke reaching its maximum value. This indicates lesser piston secondary motion. During

Lubrication Dynamics 109 10–4

Oil film thickness (mm)

1

Thrust side Anti thrust side

0.8 0.6 0.4 0.2 0

0

120

240

360 480 Crank angle

600

720

Figure 4.18 Oil film thickness behavior at 2000 RPM.

1

10–3 Thrust side Anti thrust side

Energy (J)

0.8 0.6 0.4 0.2 0 0

120

240

360 480 Crank angle

600

720

Figure 4.19 Transferred energy behavior at 2000 RPM.

expansion stroke film thickness again drops to minimum value which indicates development of full hydrodynamic lubrication. During expansion stoke the film thickness rises again reaching maximum value at 720 ° crank angle. Locations of instances of slapping motion was on basis of maximum energy transfer method was found to be [120°, 240°, 370°, 500°, 720°] for anti thrust side and for [130°, 200°, 480°, 630°, 710°] thrust side as shown in Figure 4.19. The variations in values of energy are due to fluctuations in in cylinder pressure values. Locations of instances of slapping motion was on basis of analysis of squeezing velocity was found to be [220°, 270°, 420°, 630°] for anti thrust side and for [230°, 420°, 540°, 620°] thrust side. The above mentioned location of events was validated by measuring engine block accelerations both towards thrust as well as anti thrust side and by doing time frequency analysis of filtering signals in 450–3000 Hz range under full load conditions. The events having high energy are seen having wider frequency ranges.

110 Liquid Piston Engines 10–7

Squeeze velocity (m/s)

2.5

Thrust side Anti thrust side

2 1.5 1 0.5 0

0

120

240

360 480 Crank angle

600

720

Figure 4.20 Squeeze velocity of lubricant at 2000 RPM. 10–2 14

10000

Frequency (Hz)

12 5000

10 8 6

0 S1

S2

S3

S4

S6

S5

S9 S8 S7

4 2

–5000

0

120

240

360 Crank angle

480

600

720

Figure 4.21 Time frequency analysis of filtered acceleration signals (Thrust side).

It may be noted that 9–10 events of slapping motion were recorded both at thrust as well as anti thrust side with highest energy level taking place during intake stroke (event S2). There were other traces of energy levels visible which maybe contributed to other cyclic events occurring during operation of engine.

4.10

Literature Review

In engine the small gap between piston skirt and the cylinder liner wall allows a small movement in lateral direction as well as rotational motion

Lubrication Dynamics 111 10–2 14

10000

Frequency [Hz]

12 10

5000

8 6

0 S1 S2 S3

S4

S5

S6

S9 S7 S8

4 2

–5000 0

120

240

360 480 Crank angle

600

720

Figure 4.22 Time frequency analysis of filtered acceleration signals (Anti thrust side).

of piston pin about its axis in addition to reciprocating motion of piston during primary mode of engine operation. Although piston is held in its cylinder bore by piston rings, there is a small gap between ring and grooves and hence piston is free to move within this gap [31]. The existence of piston-liner gap puts a limit on amplitude of piston motion but not on degree of freedom in lateral as well as reciprocating motion [32]. The piston secondary motion is crucial to understand frictional forces in piston assembly which contribute to 30–40% of total mechanical losses and hence a major cause of inefficiency of engine [33, 34]. The piston translates from one side to liner to other side due to changing in direction of side thrust force due to motion of connecting rod [35]. But inertial imbalance as well as piston offset and reciprocating motion guides the piston to move in a direction [34, 36], but piston collides with cylinder walls and generates piston slap noise and engine vibrations. Flores et al. [32] presented a computational methodology for slider crank mechanism dynamics. The existence of translational clearance in the slider crank mechanism makes the system highly nonlinear and the dynamic nature of slider crank trends to be chaotic with gap increase. The coefficient of restitution also plays an important role in dynamics of piston slap. As the coefficient of restitution decreases the motion between the piston skirt and liner transforms from bouncing chaotic to transient and periodic one [37]. This has been confirmed by the study of Farahanchi and Shaw [38] who showed that sliding motion stops as the coefficient of restitution approaches unity. The primary motion of piston in the model has been studied by Mcfadden and Turnbull [39] which is determined by combustion gas pressure.

112 Liquid Piston Engines The secondary motion of piston is suppressed due to hydrodynamic action of oil which plays a role of damper between skirt and liner. A two degree of freedom system developed by Geng and Chen shows a correlation between piston slap and induced vibrations of engine block. The model was used to simulate piston head motion inside cylinder. The slap induced vibration experiment is carried out and results verify the model. However this model was limited to piston lateral motion. Several simulations have been carried out to numerically simulate the two dimensional model of piston slap [40]. Various parameters considered include center of gravity offset [41], skirt profile [42], effects of variable inertial force [43], effect of frictional force [44] and effects of lubricating oil [45]. Another model has been developed which verifies the indirect measurement of piston secondary motion by mounting accelerometers on block surface to measure vibration response and to predict the piston secondary motion from impact force of slapping motion [46]. Several simulations have been carried out to numerically simulate the two dimensional model of piston slap [47]. Various parameters considered include center of gravity offset [17], skirt profile [18,19], effects of variable inertial force [20, 21], effect of frictional force [22] and effects of lubricating oil [23]. Another model has been developed which verifies the indirect measurement of piston secondary motion by mounting accelerometers on block surface to measure vibration response and to predict the piston secondary motion from impact force of slapping motion [5]. There can be several points of contact between liner and skirt as seen from Figure 4.23. Corners 1 or 2 or both of these can come into contact with liner when skirt rotates counter clockwise. Similarly corners 3 or 4 or both can touch liner as skirt moves in clockwise direction. The skirt comes in contact with liner when its lateral displacement is greater as compared to skirtliner gap.

1

2

1

4 (I)

(II)

3 (III)

Figure 4.23 Modes of contact during piston slap.

(IV)

2

3

4 (V)

(VI)

Lubrication Dynamics 113 Various conditions for impacts to occur are enlisted below:

Xc 2 Xc Corner 2 in contact, 2 Corner 1 in contact,

Corner 3 in contact, 0 Corner 4 in contact, 0 Corner 1, 2 in contact,

Xc 2

Corner 3, 4 in contact, 0

Xp

Xc ,

0

Xp

Xc ,

0

Xc , 2 Xc , 2

Xp Xp

Xc ,

Xp Xp

Xc , 2

0 . 0 max

(4.18)

max

Subsequently various modes of piston secondary motion may be classified as in figure given below: These modes can be expressed as: a. Rattling motion-During this motion skirt rotates in clock wise direction before ignition TDC position and turns its direction after TDC position, rotating in clock wise direction with its top part striking anti thrust side of skirt as shown in left part of the above figure. Amplitude of this motion increased with increase of speed and load values.

Rattling Th

Croaking ATh

Th

Figure 4.24 Modes of slapping motion.

Clatter noise ATh

Th

ATh

114 Liquid Piston Engines Fp Fi

Thrust side

FSi

Anti-thrust side

s

FSp

Ms

Mp

Fp Anti-thrust side

Thrust side Fi

Mi

FSp

s

FSi

Mp Mi

Ms

Maximum diameter portion Rattling potion

Center of rotation Croacking motion

Figure 4.25 Force analysis during various modes of piston motion.

During this motion the inertial force component (FSi) of side thrust force acts towards thrust side of liner as shown in force analysis figure shown below, whereas in cylinder gas component of side thrust force (FSp) acts towards anti thrust side of liner. b. Croaking motion-During this motion the top part of skirt strikes thrust side of skirt. The inertial force component (FSi) of side thrust force acts towards anti thrust side of liner, whereas in cylinder gas component of side thrust force (FSp) acts towards thrust side of liner. This mode of motion was found to be least affected by engine speed or load conditions. c. Clatter motion-Bottom part of skirt strikes thrust side of liner which is typical during low piston speeds.

4.11 Piston Motion Simulation Using COMSOL Most prominent instances of piston slap take place in vicinity of dead center positions A multi body dynamics model of piston using COMSOL Multi physics 7 was next used for analysis of dynamic model of piston. The model used here includes the following: a. Distortion of liner under thermal and assembly loads b. Thermal expansion of the piston c. Frictional forces and oil film action Figures 4.26–4.30 shows quarter profiles of piston skirt analyzed using FEA for testing cases enlisted in Table 4.2.

Lubrication Dynamics 115

Case 1

Figure 4.26 FEA Model of piston skirt (Case 1).

Case 2

Figure 4.27 FEA Model of piston skirt (Case 2).

Case 3

Figure 4.28 FEA Model of piston skirt (Case 3).

Results have been obtained for surface velocity of skirt at both thrust as well as anti thrust sides in time as well as frequency domains. As evident from these figures surface velocity of skirt is expected to decrease due to higher hydrodynamic action of oil becoming more dominant.

116 Liquid Piston Engines

Case 4

Figure 4.29 FEA Model of piston skirt (Case 4).

Case 5

Figure 4.30 FEA Model of piston skirt (Case 5).

Table 4.2 Engine parameters. Parameter

Value

S rp l L μ

68 mm 34 mm 121 mm 62.65 mm 0.03 Pa-s 2000 RPM, 3000 RPM 179 g 84 g 100 g 363 g 31.3250 mm 36.9 mm 6.6 10 8 kg-m2 0 mm 0 mm

mpiston mpin msl mt bc ap Ipiston Cp Cg

Lubrication Dynamics 117

4.12 Force Analysis The dynamic model of piston secondary motion as simulated using parameters enlisted in Table 4.2. When gas acts on piston the hydro dynamic oil film force Fh creates a moment Mh about piston pin. The displacement of piston along liner (Z) can be expressed as:

Z

rp cos θ

(l2

Bs2)

(4.19)

This equation may be differentiated to get values of piston velocity as plotted in Figure 4.34.

3

Case 1 Case 2

Piston velocity (m/s)

2 1 0 –1 –2 –3 –4 0

120

240

360 480 Crank angle

600

720

Figure 4.31 Velocity of piston skirt (2000 RPM).

3

Case 4 Case 3 Case 5

Piston velocity (m/s)

2 1 0 –1 –2 –3 –4

0

120

240

360 480 Crank angle

Figure 4.32 Velocity of piston skirt (3000 RPM).

600

720

118 Liquid Piston Engines FIC

y x

Fg bc

FIC Fh

MIC

ap

L

Mh + Mf Ff

Cg

FL Cp

Figure 4.33 Piston skirt force balance.

Piston velocity-m/s

15

2000RPM 3000RPM

10 5 0 –5 –10 –15

0

100

200

300 400 500 Crank angle

600

700

800

Figure 4.34 Piston velocity.

The inertial force acting along X axis (FIC) may be expressed as:

FIC

(mpiston mpin msl )(rpw cos ) (rpw Bs cos )2 l2

rpw 2 (rp cos

Bs 2

l2

2

cos Bs ) Bs 2

(4.20)

Gas force acting on piston top (Fg) may be expressed in terms of in cylinder pressure (Pg) and piston diameter (D) as:

Fg

Pg

D2 4

(4.21)

Lubrication Dynamics 119

Inertial forces -N

40 3000RPM 2000RPM

20 0 –20 –40 0

100

200

300 400 500 Crank angle

600

700

800

Figure 4.35 Variation of Inertial force along X axis.

Inertial forces along Y axis FIC may be expressed as:

FIC

(mpiston mpin msl )

de 2t dt 2

bc de 2t L dt 2

de 2b dt 2

(4.22)

As piston moves along liner the frictional force between liner and skirt causes a shear force (τ) in oil film which can be expressed as [31]:

U h

(4.23)

The hydrodynamic friction force Ff and its moment Mf about the wristpin can be calculated based on the above shear stress and these can be defined as follows:

Ff Mf

R

R

(x , q)dx d

(x , )(R cos

C p )dx d

(4.24) (4.25)

The oil film force Fh and its moment about wrist pin Mh due to the nonlinear pressure distribution, can be calculated by the following integrations [23]:

Fh Mh

R

R

p(x , ) cos dx d

p(x , )(a p

x )cos dx d

(4.26) (4.27)

120 Liquid Piston Engines The rotatory moment about wrist pin MIC can be calculated as:

I piston de 2t L dt 2

M IC

deb2 dt 2

(4.28)

Further various force and moment balance equations for the system may be expressed in form of:

Fg +FIC + Ff + FL cos φ

0

(4.29)

Fh + F

0

(4.30)

IC

+Fr + FL sin φ

Mh + MIC + FIC(ap

bc)+FgCp

FICCg + Mf

0

(4.31)

Which may be consolidated in matrix form as:

bc L

mpis 1 mpis 1 et eb

ap L

mpin 1 (bc a p )

ap L Ipiston L

mpis mpis

ap L

bc L

mpin

(bc a p )

ap L Ipiston L

Fh Fr (FIC Fg Ff )Tan Mh M f Fg CP FIC C g FIC (a p bc )

(4.32)

These equations of motion are nonlinear and stiff differential equations which can be solved using time step Runger Kutta method using initial conditions of et 0, eb 0. The time period of simulation was taken as 2 engine cycle with time step of 0.34s.

4.13 Effects of Various Skirt Design Parameters a) Effects of piston pin offset The piston pin may be offset towards either thrust side or towards the anti thrust side of liner with 0–2 mm amplitude as shown in Figure 4.36. The piston pin offset distance inclines towards anti thrust side when Cp is negative and towards thrust side when Cp is positive. Piston secondary motion equations are defined by piston eccentricities normal to axis of liner. Figures 4.36, 4.37

Lubrication Dynamics 121 Cp

Cp

TS

ATS

TS

ATS

Figure 4.36 Variations of piston pin offset.

Top eccentricity [microns]

0.06 0.05 +1 mm 0 mm –1 mm

0.04 0.03 0.02 0.01 0 –0.01 –0.02

0

120

240

360 480 Crank angle

600

720

Figure 4.37 Variations of Top eccentricities with piston pin offset.

shows the comparisons of dynamic features of piston secondary motion using three different offset distances of 1 mm, 0 mm, 1 mm. The value of side thrust force F = (Fg + FIC) Tan φ falls when φ inclines towards thrust side. i.e. φ = φʹ and vice versa. Hence there is a tradeoff between two positions since in order to reduce slapping motion the piston pin is offset towards thrust side whereas in order to reduce wear it must be offset towards anti thrust side. The amplitude of moment due to gas pressure M increases when offset is towards anti thrust side hence preventing tilt of skirt. b) Effect of skirt-liner gap Piston skirt presses against liner walls at dead center positions which causes slapping motion of piston due to changing inertial forces. The

122 Liquid Piston Engines

Bottom eccentricity (microns)

0.02 0.01 0 –0.01

+1 mm 0 mm

–0.02

–1 mm

–0.03 –0.04 –0.05 –0.06

0

120

240

600

360 480 Crank angle

720

Figure 4.38 Variations of Bottom eccentricities with piston pin offset.

Top eccentricity (microns)

0.6

+1 mm 0 mm –1 mm

0.4 0.2 0 –0.2

0

120

240

360 480 Crank angle

600

720

Figure 4.39 Variations of Top velocities with piston pin offset.

Bottom velocity [m/s]

0.4 +1 mm 0 mm –1 mm

0.2

0

0.2

–0.4 0

120

240

360 480 Crank angle

600

Figure 4.40 Variations of Bottom velocities with piston pin offset.

720

Lubrication Dynamics 123 10–5

Tilt angle (degree)

5 0

–1 mm 0 mm +1 mm

–5 –10 –15 0

120

240

360

480

600

720

Figure 4.41 Variations of piston Tilt angles with piston pin offset.

10–6

3 Tilt velocity (degree/s)

2 1 0 +1 mm 0 mm –1 mm

–1 M

–2 –3

ATS

Ts

–4 –5 –6

0

120

240

360 480 Crank angle

600

720

Figure 4.42 Variations of piston Tilting velocities with piston pin offset.

impact velocity of skirt with liner causes vibrations and hence is important source of noise. Larger skirt-liner gap causes larger impact velocities which leads to a larger impact forces whereas smaller gap causes asperity contact between mating surfaces. Nominal value of clearance is taken as 0.03–0.04 mm. Larger values of clearances increase the piston lateral displacements, however main effecting factor of slapping motion is lateral velocity of piston. The energy of impacts due to piston secondary motion (EI) consists of contributions due to both rotational as well lateral components. i.e.

E

1 MV 2 2

1 J 2

2

(4.33)

124 Liquid Piston Engines The results show that maximum values of velocity of skirt is 0.071 m/s, 0.064 m/s and 0.062 m/s for clearance values of 0.04 mm, 0.05 mm and 0.06 mm respectively and these occur in vicinity of TDC position. From the plots of velocity of skirt it is clear top edge is moving away from anti thrust side of liner and the tilting angle increases with increase in piston-liner gap. Hence slap force caused by lateral component of velocity increases with skirt-liner gap. The reason for this is that load bearing capacity of the oil film increases with a decrease of skirt-liner gap which prevents piston from striking the liner but this in turn increases the wear of skirt. Apart from lateral velocity, rotational component also plays a vital role, hence rotational velocity of skirt must be compared. Figure 4.45 shows results for comparisons of rotational velocity of skirt. The results show that rotational velocity falls with decrease of skirt-liner gap. This is due to the fact that rotational moment about piston pin increases with fall of clearance values which prevents further rotation of skirt about piston pin.

Top eccentricity [microns]

0.05 0.04 0.06 mm 0.05 mm 0.04 mm

0.03 0.02 0.01 0

–0.01

0

120

240

360 480 Crank angle

600

720

Bottom eccentricity [microns]

Figure 4.43 Variations of Top eccentricities with skirt-liner gap. 0.01 0 –0.01

0.06 mm 0.05 mm 0.04 mm

–0.02 –0.03 –0.04 –0.05

0

120

240 360 480 Crank angle

600

Figure 4.44 Variations of Bottom eccentricities with skirt-liner gap.

720

Lubrication Dynamics 125 0.25

Top velocity (m/s)

0.2 0.15

0.06 mm 0.05 mm 0.04 mm

0.1 0.05 0 –0.05 –0.1 –0.15

0

120

240

360

480

600

720

Figure 4.45 Variations of Top velocities with skirt-liner gap.

0.15 0.06 mm 0.05 mm 0.04 mm

Bottom velocity (m/s)

0.1 0.05 0 –0.05 –0.1 –0.15 –0.2 –0.25

0

120

240

360

480

600

720

Figure 4.46 Variations of Bottom eccentricities with skirt-liner gap.

10–5

2

Tilt angle (degree)

0 –2

0.06 mm 0.05 mm 0.04 mm

–4 –6 –8 –10

0

120

240

360

480

Figure 4.47 Variations of Tilting angle with skirt-liner gap.

600

720

126 Liquid Piston Engines 10–6

3

Tilt angle (degree)

2 T3

1

M

ATS

0 0.04 mm 0.05 mm 0.06 mm

–1 –2 –3 –4 –5

0

120

240

360 480 Crank angle

600

720

Figure 4.48 Variations of Tilting velocities with skirt-liner gap.

Top eccentricity (microns)

0.05 0.04 62.67 mm 62.65 mm 62.6 mm

0.03 0.02 0.01 0

–0.01

0

120

240

360

480

600

720

Figure 4.49 Variations of piston Top eccentricities with skirt length.

c) Effects of variations in skirt length Effects of length of skirt has been investigated next. As length of skirt increases, the load compacting area falls which in turn leads to fall in oil pressure. This causes top velocity to fall and bottom velocity to increase. The concentration of oil will move down towards bottom of skirt which leads to higher oil film force Fh and its associated moment Mh. This prevents further rotation of skirt and hence rotational velocity of skirt falls with increase of skirt length. d) Effects of engine speed With an increase in engine speed, combustion features as well as in cylinder pressure trace changes due to increased fuel mass injected. This leads to increased thermal deformations due to higher temperatures.

Lubrication Dynamics 127

Bottom eccentricity (microns)

0.01 0 –0.01 62.67 mm 62.65 mm 62.6 mm

–0.02 –0.03 –0.04 –0.05

0

120

240

360

480

600

720

Figure 4.50 Variations of piston Bottom eccentricities with skirt length.

10–6

Tilt velocity [degree/s]

4 2 0

62.67 mm 62.65 mm 62.6 mm

–2 –4 –6

0

120

240

360

480

600

720

Figure 4.51 Variations of piston Tilt velocities with skirt length.

Figures 4.32, 4.33 shows the variations of axial velocity of piston and piston lateral thrust forces. As speed of piston increases, lubrication type changes to hydrodynamic one as higher fluid pressure will develop in oil film. Figure shows variations of piston lateral forces and velocity for the engine under consideration. It can be seen that lateral forces on skirt increase with an increase in engine load. It can be observed that piston speed is zero at dead center position. Figures 4.52 and 4.53 shows effects of higher speeds (2000 RPM, 3000 RPM). These results show that at higher speeds secondary motion of piston decreases due to higher hydrodynamic action of lubrication oil induced at higher speeds. Effects of wrist pin and piston mass was further analyzed. Figures 4.54–4.55 shows the effects of lighter piston skirt (2/3 rd of original mass) and heavier wrist pin mass (double original mass) on piston

128 Liquid Piston Engines

Top eccentricity (microns)

0.04 0.03

3000 RPM 2000 RPM

0.02 0.01 0

–0.01 0

120

240

360 Crank angle

480

600

720

Bottom eccentricity (microns)

Figure 4.52 Effect of engine speed on Top eccentricities. 0.01 0 –0.01

2000 RPM 3000 RPM

–0.02 –0.03 –0.04 0

120

240 360 480 Crank angle

600

720

Figure 4.53 Effect of engine speed on Bottom eccentricities.

secondary motion. These plots show that these parameters have little affects on piston secondary motion as a slight increase was observed in skirt eccentricities with a decrease in skirt and pin mass. e) Effects of engine load As engine load increases in order to maintain constant speed, the amount of charge to be brought into cylinder must increase. Hence peak in cylinder reached inside cylinder increases. As evident from figure, the axial velocity of piston has no significant increase with increase in load values.

Other Factors Effecting Piston Secondary Motion f) Effects of Inertia of Connecting Rod This factor has pronounced effect on piston lubrication as well as secondary motion especially at high engine speeds.

Lubrication Dynamics 129

Top eccentricity (microns)

0.05 Lighter skirt Normal skirt

0.04 0.03 0.02 0.01 0

–0.01

0

120

240

360 Crank angle

480

600

720

Figure 4.54 Effect of skirt weight on Top eccentricities.

Bottom eccentricity (microns)

0.01 0 Normal skirt Lighter skirt

–0.01 –0.02 –0.03 –0.04 –0.05

0

120

240

360 480 Crank angle

600

720

Figure 4.55 Effect of skirt weight on Bottom eccentricities.

Top eccentricity (microns)

0.05 0.04

Normal pin mass Heavy pin mass

0.03 0.02 0.01 0

–0.01

0

120

240

360 480 Crank angle

Figure 4.56 Effect of pin mass on Top eccentricities.

600

720

130 Liquid Piston Engines g) Effects of lubrication oil supply The oil supply has major effects on friction between liner and piston skirt [27]. The presence of oil film between skirt and liner reduces frictional force hence reducing asperity contact. Presence of oil film does not affect the occurrence of piston slap motion but cushions and reduces piston tiling and bouncing across liner. h) Effects of surface finish Skirt is machined so piston rings can fit in its grooves [26]. These grooves behave as reservoirs of oil for lubrication. The grooves in liner are also honed at an angle to horizon with shallow angles of honing allowing the flow of oil laterally. A smooth piston should need lesser oil for supporting hydro dynamic lubrication as compared to a rough one. I) Shape of skirt Changing flatness of skirt changes boundary and hydro dynamic friction of piston. A flat skirt profile shows a larger wetted area and thicker oil film thickness hence decreasing the boundary contact. The sharp skirt profile experiences boundary friction than a flat profile. J) Size of skirt This is also an important design parameter for piston assembly. Material selection forms an important aspect of skirt design. Smaller skirts must distribute load over a smaller area, hence they tend to have more of boundary lubrication. Skirt ovality can be increased to distribute load over a larger area. In addition to above stated factors, the wear of skirt as well liner during engine operation must also be taken into consideration which sometimes

Bottom eccentricity (microns)

0.01 0 –0.01

Heavy pin mass Normal pin mass

–0.02 –0.03 –0.04 –0.05

0

120

240

360 480 Crank angle

Figure 4.57 Effect of pin mass on Bottom eccentricities.

600

720

Lubrication Dynamics 131 leads to gas blow hence increasing engine emissions. Wear factor can be calculated by taking integral of contact pressure and piston velocity over stroke distance.

4.14 Numerical Model of Slapping Motion The motion of piston skirt can be considered as a model with three degree of freedom system as depicted in Figure 4.58. The skirt can be considered as having dual degree of freedom (Xp, ) with mass (mp) equal to 0.363 kg and moment of inertia (Ip) equal to 7.8540 10−9 kg-m2. The engine block can be considered as a lumped system having single degree of freedom (Xb) with mass (mb) equal to 48.5 kg. The nominal clearance between skirt and liner was taken as 0.05 mm. Motion of skirt can be represented mathematically in matrix form of equation 4.34.

Mp 0 0

0 mb 0

0 0 Ip

Xp Xb q

Cp Cp 0

Kp Kp

Kp Kb K p

0

0

Cp Cb C p 0 Xp Xb

0 0 K

xc/2

ccb

mp

I

cp kcp

0 MZ

ccp

D

l mr y rc mc x

Figure 4.58 Model of Piston secondary motion.

x2

cp

yp

c

(4.34)

kp kcp

ccp

Xp Xb

Fx

xc/2 x1

kp kb

0 0 C

kb mb ccb

132 Liquid Piston Engines

4.15 Piston Side Thrust Force The existence of clearance between skirt and cylinder liner allows the piston to move and rotate freely within the confined region resulting in piston secondary motion and slap. The main driving force is the side thrust force imparted to skirt by connecting rod as shown in Figure 4.59. The frictional forces between piston skirt and cylinder liner as well as between rings and liner act vertically along Y axis. The center of mass of piston assembly is at horizontal offset of LX and at vertical offset distance of Ly from connecting rod position. The force exerted by connecting rod on piston pin can be vertically decomposed along X as well as Y axis. As the angle of connecting rod changes as piston moves from bottom dead center (BDC) to top dead center (TDC) there will be lateral force pushing the piston on to the cylinder liner. The piston side thrust force takes into consideration both inertial forces as well as gas forces as developed by Guzzomi and is given by following equation in terms of crank radius-connecting rod length ratio (K):

Fx = [Fg

mp r 2(cos(θ)+K cos(2θ)]) Dp

Piston grooves

Piston skirt

Ffr FIX

FIY

Lx

Ly Frody

Figure 4.59 Free body Piston diagram.

Frodx

Ff

(4.35)

Lubrication Dynamics 133

Sin2

Where

l

4.16

2

(r sin )2

Frictional Forces

The piston ring friction force is predominant in the total engine mechanical loss and is the highest single contributing factor. According to Zweiri, the piston ring pack friction force can be expressed as the product of the elastic ring tension and the coefficient of friction. As the piston speed increases, the friction coefficient decreases gradually until minimum at mid stroke as hydrodynamic lubrication region is achieved. The frictional forces between liner and skirt (Ff) and piston rings and liner (Ffr) can be written in terms of sliding velocity of piston (V), nominal clearance (h), lubricating oil viscosity ( ), number of piston rings (n) and shear area of contact (As) as:

Ff = μV

(4.36)

Ffr = μV

(4.37)

Where As1 is shear contact area between liner and skirt and As2 is the shear contact area between liner and rings.

4.17 Determination of System Mobility Mechanical mobility (M) can be defined as the ratio of resulting velocity of structure to input force causing excitation. This parameter can be used to find the dynamic mass, stiffness and damping constant of skirt-liner system. In frequency domain the mechanical mobility M(J ) can be expressed as: [26]

M (J )

M (J )

V (J ) F (J )

J ((K M 2 ) JC ) M 2 (K JC )

(4.38)

(4.39)

Where F(J ) is exciting force in frequency domain and V(Jω) is a frequency domain velocity response function.

134 Liquid Piston Engines 3000

Lateral forces -N

2500 2000

Case 2 Case 1

1500 1000 500 0 –500 –1000 –360

–180

0 Crank angle

180

360

Figure 4.60 Piston side Thrust force (2000 RPM).

The measurement of mechanical mobility was carried out to compute the dynamic parameters of skirt-liner system. The response of cylinder block was captured using accelerometer mounted on the top of engine block normal to axis of piston motion, whereas the response of piston skirt was analyzed using COMSOL software. Frequency domain plots of mobility have shown that frequency range K is dominated by dynamic below first anti resonance frequency a m mass of system. Hence the point mobility equation can be approximated as:

M (J )

J m

(4.40) a

Above this anti resonance frequency, response of system is spring dominated and hence the point mobility may be expressed as [26]:

M (J )

J K

a

(4.41)

Figures 4.60, 4.61 shows the plots of piston side thrust forces at 2000 and 3000 RPM conditions. As seen from these plots, side thrust force changes its direction four times in a complete engine cycle indicating instances of piston slapping contact as depicted by circles. COMSOL 7 software was then used to simulate piston secondary motion and hence compute the piston velocity as shown in Figures 4.62, 4.63. As it is seen from these plots, the velocity of piston approaches zero values near top dead center position. Major variations in velocity profile

Lubrication Dynamics 135 3000

Lateral forces -N

2500

Case 3 Case 4 Case 5

2000 1500 1000 500 0 –500 –1000 –360

–180

0 Crank angle

180

360

Figure 4.61 Piston side Thrust force (3000 RPM). 8

Piston velocity -m/s

6 4 2 0 –2 –4

Case 1 Case 2

–6 –8 –10 –360

–180

0 Crank angle

180

360

Figure 4.62 Piston velocity (3000 RPM). 8

Piston velocity -m/s

6 4 2 0 –2 –4 –6

Case 5 Case 3 Case 4

–8 –10 –360

–180

0 Crank angle

Figure 4.63 Piston velocity (2000 RPM).

180

360

136 Liquid Piston Engines

Piston mobility -m/N-s

were observed during exhaust stroke after 630° crank angle position due to variations in inertial forces. Further using equation 5, mechanical mobility of skirt was computed for the given testing condition as depicted in Figures 4.64, 4.65. Variations in values of mobility shows same trends hence confirming that mobility is least affected by change in the engine operational conditions. First anti resonant frequency for skirt was found to be close to 60 Hz range. Block velocities of engine were simulated using numerical integration of accelerometer data as shown in Figures 4.66, 4.67. The values of block velocities were used to plot graphs of block mobility as seen in Figures 4.68, 4.69. At lower speeds gas force acting on skirt is a major factor affecting mechanical mobility, hence despite almost same values of first anti resonant frequency, the values of block mobility were found to be lower as compared with piston mobility. Using the concept of anti-resonant frequency as discussed in previous section, various dynamic parameters of liner-piston were computed for given test conditions. The results can be seen in Table 4.1. 1st anti resonant frequency 10–4 Case 1 Case 2

10–6 10–8 101

102 Frequency-Hz

103

Piston mobility -m/N-s

Figure 4.64 Piston mobility (3000 RPM).

10–4

1st anti resonant frequency

Case 3

10–6

Case 4 Case 5

10–8 101

102 103 Frequency-Hz

Figure 4.65 Piston mobility (2000 RPM).

Lubrication Dynamics 137

Block velocity -m/s

Case 1 Case 2

100

0

–100 –360

–180

0 Crank angle

180

360

0 Crank angle

180

360

Block velocity -m/s

Figure 4.66 Block velocity (2000 RPM). 100 0 Case 3 Case 4 Case 5

–100 –360

–180

Block mobility -m/N-s

Figure 4.67 Block velocity (3000 RPM).

1st anti resonant frequency 10–5 Case 1 Case 2

10–10 101

102 103 Frequency-Hz

Block mobility -m/N-s

Figure 4.68 Block mobility (2000 RPM). 10–5

10–10

1st anti resonant frequency

Case 3 Case 4 Case 5

101

Figure 4.69 Block mobility (3000 RPM).

102 Frequency-Hz

103

138 Liquid Piston Engines Table 4.1 Dynamic parameters of system. Case

Value

Parameter

1

Liner Parameter

Piston Parameter

a

67 Hz

[M(J )] a = 10 m/N-s

[M(J )] a = 2.5

c

c

109330 (kg/s)

k

1.63

42884 (kg/s)

k

4.2

m 2

a

9

2

10 (kg/s )

m

23754 (kg) 67 Hz

65 Hz

[M(J )] a = 3.98

c

c

109330 (kg/s)

k

1

k

4.2

m 3

a

9

2

10 (kg/s )

m

23754 (kg) 63 Hz

a 7

[M(J )] a = 1.99

10 m/N-s

69669 (kg/ s2)

c

k = 1.98

2

a

7

10 m/N-s

c = 69669 (kg/ s2) k = 1.5

a 7

10 m/N-s

c = 69669 (kg/s2) k = 1.63

[M(J )] a = 2.5

107 (kg/s2)

= 100 Hz

[M(J )] a = 1.9

10 5 m/N-s

c = 172750 (kg/s) 2

10 (kg/s )

m = 10105 (kg)

10 5 m/N-s

m = 63 (kg)

= 63 Hz

9

107 (kg/s2)

100 Hz

k = 2.5

10 (kg/s )

[M(J )] a = 2.5

10 5 m/N-s

c = 172750 (kg/s) 2

m = 10105 (kg) a

100 Hz 172750 (kg/s)

a

9

39 (kg)

m = 50 (kg)

= 63 Hz

[M(J )] a = 2.5

107 (kg/s2)

k = 1.98

10 (kg/s )

10 5 m/N-s

[M(J )] a = 3.1623 c

9

m = 11937 (kg)

5

107 (kg/s2)

[M(J )] a = 10 m/N-s 42884 (kg/s)

10 5 m/N-s

63 (kg)

a 7

4

65 Hz

a 7

k = 3.3

107 (kg/s2)

m = 83 (kg)

The rotational motion of skirt about pin axis was simulated by solving dynamic equations of motion and compared with results obtained from COMSOL as seen in Figures 4.70–4.74. Both the trends showed a good correlation. As seen from figures, the angle of tilt of skirt changes at both dead centers due to changing position of connecting rod. Piston was found

Lubrication Dynamics 139 1

Experimental Simulated

Angle, deg

0.8 0.6 0.4

Suction

0.2 0

0

Compression

180

Exhaust

Power

360 Crank angle, degrees

540

720

Figure 4.70 Piston tilting motion (2000 RPM-80% load) 0.8

Experimental Simulated

Angle, deg

0.6 0.4 0.2

Suction

0 –0.2

Compression

0

180

Exhaust

Power

360 540 Crank angle, degrees

720

Figure 4.71 Piston tilting motion (2000 RPM-100% Load). 0.6

Experimental Simulated

Angle, deg

0.4 0.2 0 –0.2 –0.4

Suction

0

Compression

180

Exhaust

Power

360 Crank angle, degrees

540

720

Figure 4.72 Piston tiling motion (3000 RPM-Motored). 0.4

Experimental Simulated

Angle, deg

0.2 0 –0.2

Suction

–0.4 –0.6

0

Compression

180

Power

360 540 Crank angle, degrees

Figure 4.73 Piston tiling motion (3000 RPM-80% Load).

Exhaust

720

140 Liquid Piston Engines

Angle, deg

0.2

Experimental Simulated

0 –0.2 –0.4

Suction

–0.6 –0.8

0

Compression

180

Exhaust

Power

360 Crank angle, degrees

540

720

Block vibration amplitude-mm

Figure 4.74 Piston tiling motion (3000 RPM-100% Load).

5 10–3

4 3 Experimental Simulated

2 1 0

90

180

270 360 450 Crank angle

540

630

720

Figure 4.75 Block vibrations (2000 RPM-80%Load).

to slide for a some crank angle duration before reaching TDC along the cylinder liner. Also piston tilting angle decreases with increase in load. During motion of skirt from TDC to BDC, the piston tilts towards counter clock wise direction corner 1 of skirt touches liner at 135° crank angle during suction stroke. The piston skirt tilts towards clockwise direction during compression stroke during motion from BDC to TDC hence corner 1 touches skirt at around 240° crank angle position. During power stroke, again counter clockwise tilt is observed as corner 2 touches liner at around 410° crank angle position. During exhaust stroke, skirt rotates in counter clock wise direction as corner 4 touches liner at 650° crank angle position. Figures 4.75–4.79 shows the simulated and measured vibratory response of cylinder block as captured by accelerometer. The trend of the simulated vibration response of the cylinder block shows a good agreement with the measured vibration response. The plots show harmonic peaks that are related to fundamental firing frequency of engine. In these figures, the impact of the piston on the cylinder wall results in a sudden increase of the vibration amplitude and this is clearly marked in the diagram where a few impacts occurs. The induced vibration amplitude of the cylinder block

Lubrication Dynamics 141

Block vibration amplitude-mm

5 10–3

4 3

Experimental Simulated

2 1 0

90

180 270 360 450 540 630 Crank angle

720

Block vibration amplitude-mm

Figure 4.76 Block vibrations (2000 RPM-100%Load). 5 10–3

4 3

Experimental Simulated

2 1 0

90

180 270 360 450 540 630 720 Crank angle

Block vibration amplitude-mm

Figure 4.77 Block vibrations (3000 RPM-motored). 5 10–3

4 3 Experimental Simulated

2 1 0

90 180 270 360 450 540 630 720 Crank angle

Figure 4.78 Block vibrations (3000 RPM-80%Load).

measured experimentally is slightly higher than the predicted induced vibration amplitude due to severe oscillations taking place during expansion stroke owing to high combustion gas pressure. Major variations were found at 540° crank angle due to piston impacts with liner. These vibrations

Block vibration amplitude-mm

142 Liquid Piston Engines 5 10–3

4 3 Experimental Simulated

2 1 0

90

180

270 360 450 540 Crank angle

630

720

Figure 4.79 Block vibrations (3000 RPM-100%Load).

Lateral amplitude - microns

20 15 10

Case 2 Case 1

5 0 –5

0

90

180

270 360 450 Crank angle

540

630 720

Figure 4.80 Piston lateral motion (2000 RPM).

gradually decay and are induced upon next instance of impact after some crank angle rotation. The effects of load and speed variations on piston lateral motion were next investigated. The datum for lateral motion of skirt was taken as its lower edge. The range of amplitude of this motion is within 0–0.05 mm which depicts nominal value of skirt-liner gap. As the speed of engine increases, the side thrust force which is dependent upon speed, also increases. An increase in the magnitude of this side thrust force acting on the piston, results in the piston bouncing off the liner more frequently for longer durations and the sliding duration of the piston skirt along liner falls. At low engine speeds, the vibration response of the cylinder block induced by the slapping contact of the piston has a longer duration till decay as compared with higher engine speeds. As evident from these figures, with an increase in engine speed, the secondary motion of piston became more dominant. From these plots it is clear that piston rotates in counter clockwise direction till 270° crank angle

Lateral amplitude - microns

Lubrication Dynamics 143 6 5 Case 3 Case 4 Case 5

4 3 2 1 0 –0.01

0

90

180

270 360 450 Crank angle

540

630

720

Figure 4.81 Piston lateral motion (3000 RPM).

position. During the course of this motion corner a comes in contact near BDC position. During the motion of skirt from BDC to TDC position, piston skirt changes direction of rotation to clockwise sense and corner b comes in contact with skirt near 360° crank angle position. An increase in engine speed causes increase in the side thrust force. This further results in a decrease in sliding motion of skirt along liner as piston bounces off frequently for longer time durations. This result is similar to one discussed in [20]. At a lower values of engine speeds, the vibration response of liner induced by the slapping motion takes longer time to decay.

4.18 Conclusion Piston slapping motion is a major cause of noise and vibrations in engines. In order to understand this motion, a numerical model was presented in the present work. Various dynamic parameters of system were calculated using concept of mobility which were later used to simulate the lateral motion of piston as well as resulting engine block vibrations. The values of first anti resonant frequencies of both skirt and liner were found to be near 60 Hz range and it remains unaffected by variations in the engine operational conditions. Several peaks were found in the simulated block vibrations which were related to firing frequency of engine. COMSOL software was further used to analyze the tilting motion of piston which showed a good match with that simulated by solving dynamic equations of motion. Amplitude of piston secondary motion was found to be maximum value in middle of intake stroke, when lubrication action of oil is minimum. Piston was also found to slide along liner a few crank angle degree before TDC position. Sliding motion is less dominant during power and exhaust

144 Liquid Piston Engines as bouncing motion dominates piston dynamics. The duration of sliding motion of piston along liner was observed to increase with increase in load and speed conditions which is in agreement with previous available literature. This work discusses a two dimensional model of piston secondary motion. Various dynamic parameters of system were calculated using concept of mobility. These parameters were used to simulate piston secondary motion and hence validate the model. Effects of load and speed were also investigated on piston secondary motion. Sliding motion of piston along liner was observed to increase with increase in load and speed which is in agreement with previous data available.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

5 NVH Features of Engines

5.1

Background

Diesel engines constitute a major source of power for various ships, buses, trains as well as road machinery. About one-fifth of total energy consumption in U.S.A. goes towards operating these engines [1] and demand for these engines is fast growing compared with gasoline engine [2]. Sales of diesel based engines reached a peak during the 1980s in U.S.A. due to major oil crises as shown in Figure 5.1 [1]. Various projections at that time had predicted that an increase of about 20% in sales would be achieved at the end of decade [3]. But variations in the costs of diesel, falling petrol prices and problems in operations of diesel engines led to a fall in their sales [4, 5]. Petrol engines use spark ignition for initiation of fuel reactions as compared with diesel engines which are based on compression ignition of fuelair mixture. Diesel engines operate at higher compression ratio as compared with petrol engines allowing more useful work in cycle. Combustion in these engines can be made to occur away from walls, thus helping in overall heat reduction. In addition there are throttling as well as various pumping losses 145

146 Liquid Piston Engines US sales of diesel vehicles 500,000 400,000 300,000 200,000

Autos

100,000 1996

1994

1992

1990

1988

1984

1982

1980

1978

1986

Light trucks

0

Figure 5.1 Sales trend of diesel engine based automobiles in U.S.A.

in petrol engines, which is the reason for their lesser efficiency as compared with diesel engines. Overall fuel efficiency of a diesel engine may be over 40% in medium engines and over 50% for larger engines used in marine propulsions [6]. These factors have renewed interest of automobile companies towards diesel engines. Data about diesel engine automotive sales in Europe have indicated that about a quarter of new automobiles are powered by diesel engines [7, 8]. In France, diesel engines accounted for about half of automotive engine sales [9]. In Japan, the number of diesel engine car sales have tripled in the past [10]. Several commercial vehicle suppliers have started to manufacture their own Diesel Engines for installation in engines. Table 1.1 shows the representation of data about U.S. market share of various diesel engines supplied by automotive manufacturers. Recently several modern technologies like common rail direct injection system (CRDI), exhaust gas recirculation (EGR) and turbo charging are examples of some of modern technologies being introduced for development of diesel engines in modern times [12]. Other methods used include pre-mixed charge compression ignition (PCCI) and homogenous charge compression ignitions (HCCI) systems [13–15]. However higher period of pre-mixed combustion in these technologies leads to higher noise emissions. Hence various merits of using a diesel engine may be lost over noise, vibration and harness performance benchmarks.

5.2 Acoustics Overview of Internal Combustion Engine Vehicle noise and vibrations can have a bad effect on overall performance of automobiles. These aspects also form important benchmarks for

NVH Features of Engines

147

Table 5.1 Supply of diesel engines by various manufacturer, Year-2013. Automotive make

Engine make

Hino

Hino

100%

Freightliner

Cummins

62.3%

Detroit Diesel

37.0%

Mercedes Benz

0.7%

Cummins

7.2%

Navistar

92.8%

Cummins

13.6%

Volvo

86.4%

Cummins

21.2%

Detroit Diesel

78.8%

International Volvo Western Star Mack Peterbilt

Market share

Cummins

6.0%

Mack

94.0%

Cummins

65.2%

PACCAR

34.8%

Mount

Exhaust system Engine Clutch Transmission

Intake system Powerplant

Driveline

Figure 5.2 Power train system.

customers perception of vehicle performance. They are measures of comfort levels and vehicle reliability. In automotive terms noise, vibration and harness (NVH) is used to denote the unwanted sound and vibrations in an automobile [16]. NVH is a term used for branch of engineering related to vehicle refinement in terms of sound and vibration performance experienced by occupants. A typical vehicle system consists of several systems which include chassis system, power train system, HVAC system and electronics system [17].

148 Liquid Piston Engines Vehicle vibrations and sound Vibrations

Engine

Mount

Tire

Drive train

Squeak and rattle

Noise

Body stiffness

Vehicle medes

Miscellaneous

Component resonance

Interior noise

Engine

Tire road

Exterior noise

Wind noise

Passenger noise

Miscellaneous

Suspension Body structure borne

Engine system

Intake/ exhaust

Body air borne

Figure 5.3 Noise and vibration sources in engine.

A schematic arrangement of a typical powertrain system is shown in Figure 5.3 [18]. The powertrain system includes engine, transmission system, clutch, driving system, intake and exhaust systems. The engine system is a major source of vibrations which may be classified as external or internal one. The internal sources are due to variable pressure on piston head and inertia of moving parts. The external sources refer to vibrations due to unbalanced moments and variable engine torque which results in vibrations of whole engine block. The noise sources in engine consists of mechanical, combustion and aerodynamic noise [19]. Mechanical noise which is proportional to engine speed is due to inertial effects of relative motion of parts under air pressure or inertial force that results in impact noise and vibrations. This noise includes noise due to piston motion, bearing noise, cam noise, oil pump noise, timing belt and chain noise as well as structural noise of cover [20]. Aerodynamic noise includes intake noise, exhaust noise and fan motion noise. Combustion noise is generated due to impulsive pressure wave due to combustion process resulting in vibrations of engine block due to impacts on cylinder wall and head [21]. The vibrations due to transmissions and driveline also contribute towards powertrain noise levels. There are also other noise sources due to squeak and rattle of engine body system. The noise experienced by the passengers inside the vehicle not only depends upon various sources but also upon body structure and acoustic transfer function. Various noise and vibration sources have frequency range values. Wind and road tire noise lies in medium frequency ranges.

NVH Features of Engines

149

During the decade of 1970s, increasing attention towards various noise control led to more attention being paid towards the acoustic performance of diesel engines. Priede analyzed the relationship between in cylinder pressure development and radiated noise from engine [21]. Kamal has done finite element analysis of individual engine structure for dynamic analysis of engines [22]. Later on the basis of noise transfer paths, main bearing of connecting rod as a key design factor for controlling the indirect combustion noise [23]. In modern days multidisciplinary analysis approaches are being utilized for NVH performance evaluation of engines. Some of these include modal analysis, Finite element analysis (FEA), Boundary element method (BEM), Statistical energy analysis (SEA), lumped mass approach and transfer path analysis (TPA). Evaluation of noise performance of engines needs to be carried out both objectively as well as subjectively [24]. Each of these methods have specific frequency range over which it is most suited to. e.g. FEA is more suited low frequency ranges, TPA analysis is more suited in medium frequency ranges and SEA is suited for high ranges.

5.3

Imperial Formulation to Determine Noise Emitted from Engine

Figure 5.4 shows plots of in cylinder pressure spectra for two types of engines [25]. A difference of about 20 dB is seen at 1 KHz frequency. Based on mathematical relationships, Anderton has developed mathematical models to quantify combustion noise according to type of engine

Sound pressure level dB

210 200 190

12.2 litre D.I. diesel 130 mm. bore

180 170 160

1.6 litre gasoline 91 mm. bore

150 140 130

100

1000 Frequency-Hz

Figure 5.4 In cylinder pressure spectrum.

150 Liquid Piston Engines [26]. It involves the concept of mechanical impedance Z(f) between force applied at top of piston F(f) and average mean square root velocity v(f) of engine block. i.e.

V 2 (f ) F (f )

Z (f )

(5.1)

The average surface velocity v(f) may be expressed in terms of in cylinder pressure (p) and cylinder bore (B) as:

p B2 4Z ( f )

V (f )

(5.2)

Further the radiated acoustic power (W) from a surface may be expressed in terms of radiation efficiency ( ) and radiated surface area (S) in form of:

W( f )

CSV ( f )

p2 C

(5.3)

Combining the two relationships we have :

W

p B2 SC 4 Z( f )

(5.4)

The intensity of radiated noise I(f) can be written as:

I( f )

C

p B2 4 Z( f )

(5.5)

In order to minimize the effects of speed on Anderton analyzed various in cylinder pressure spectra operating various engines under different conditions and found that the trends of plots were almost a straight line in frequency ranges 0.8 KHz–3 KHz range. The slope of pressure spectrum in this range was defined as combustion noise index. Using further analysis it was shown that in cylinder pressure (p) may be expressed as :

N p (f)~ f 2

z

Antilog (3N )

(5.6)

NVH Features of Engines

151

where N is engine RPM From the above relationships we have:

N I( f ) ~ f

z

I( f )

p B2 Antilog (3N) cS 4Z ( f ) N f

z

S

B4 z ( f )2

(5.7)

(5.8)

Overall Intensity IO can be expressed by integration over given frequency range (f1, f2) values as:

IO

SN z B 4

f2 f1

(5.9)

f z Z ( f )2

Various empirical relationships have been developed at ISVR, University of Southampton for prediction of noise in terms of sound pressure levels for different types of engines. Some of these include :

SPLN.A. Direct Injection Diesel engines = 30

log(n) + 50

log(B)+106 (5.10)

SPLTurbocharged Diesel engines = 40

log(n) + 50

log(B)–135 (5.11)

SPLIndirect injection Diesel engines = 43 SPLPetrol engines = 50

log(n) +60

log(n) + 60

log(B)–176

log(B)-203

(5.12) (5.13)

As compared with diesel engines, a gasoline engine has higher operational speeds, smaller bore and smaller reciprocating mass. Consequently a gasoline engine has lower in cylinder pressure and hence lower sound pressure levels of radiated noise as compared with diesel engines as seen from Figure 5.5.

5.4

Engine Noise Sources

Typical noise sources in a combustion engine are plotted in Figure 5.6 [24]. Combustion noise from engine depends upon the speed of combustion process taking place in combustion chamber, 50% mass fraction burnt (CA50),

152 Liquid Piston Engines

1m noise level (dBA)

110 Diesel engines

100 90

Petrol engines 80 70 60 1000

2000

3000 RPM

4000

5000

6000

Figure 5.5 Variations of sound pressure levels with engine speed.

7 1

8

2 3 4

9 10

5 6

11

Figure 5.6 Schematic representation of various sources of noise (1: Valve train, 2: Chain drive, 3–4: Accessory noise, 5: Piston slap, 6: Bearing noise, 7: Cover noise, 8: Intake noise, 9: Exhaust noise, 10: Combustion noise, 11: Oil pan noise).

peak in cylinder pressure developed and its position in crank angle domain as well as pressure derivatives with respect to crank angle. The intensity of combustion noise is proportional to square of cylinder pressure and it also depends upon engine speed, load and injection delay period. This noise can be further classified as direct combustion noise and indirect combustion noise [27]. Direct combustion noise is directly radiated from engine structure and its transfer function can be calculated by doing an explosion in combustion chamber keeping other engine parts stationary so mechanical noise does not interfere with radiated noise. Indirect combustion noise in engines is portion of noise that is transferred to structure from combustion chamber. Mechanical noise which is proportional to engine speed is due to piston motion, bearing operations, timing belt operation, pump and valve operations. This type of noise can be estimated by running engine under motored condition. Flow noise depends upon turbulence, pressure flow and friction during flow. The effect of tailpipe and radiations of muffler are primary source of exhaust noise.

NVH Features of Engines

153

Table 5.2 Frequency ranges of various noise sources. Noise source

Approximate frequency range

Effecting factor

Combustion Noise

500–8000 Hz

In cylinder pressure

Piston Slap

2000–8000 Hz

Speed, Piston Design

Valve Operation

500–2000 Hz

Valve Type, Engine speed

Fan Noise

200–2000 Hz

Speed, Number of Blades

Intake Flow Noise

50–5000 Hz

Turbulence

Exhaust Flow Noise

50–5000 Hz

Turbulence

Injection Pump operation

2000 Hz

Pump features

Gear Noise

4000 Hz

Speed, Number of teeth

Accessory Belt-Chain Noise

3000 Hz

Engine speed, Misalignment, Number of teeth

Table 5.3 Noise analysis from a V6 engine. Part

dB sound pressure levels

Engine Block

78.7

Cylinder Head

76

Crank Case

79

Engine Base

78

Intake manifold

77

Cam Cover

78

Front Cover

77

Exhaust manifold

74

Oil pan

73

The frequency ranges of various noise sources are enlisted in Table 1.1. The range of frequency not only depends upon engine load and speed but also on configuration of engine. Hence identification and estimation of specific frequency must be done by testing. By comparison of fundamental frequency and harmonics of individual noise sources, contribution of each source can be estimated as seen from Table 5.3 for a V6 engine tested in an anechoic chamber.

154 Liquid Piston Engines

5.5 Noise Source Identification Techniques There are several techniques that can be used to identify various sources of noise in internal combustion engines [28]. Some of these include shielding techniques, surface vibration method and acoustic intensity technique. Of these methods the lead covering method is the most expensive one as well as the time consuming. Other methods are time consuming and need a lot of calculations. These techniques are discussed further in the coming part of this work. a. Lead covering method: it is one of the most reliable methods of acoustic source identification for engines. This method consists of noise emission measurement from engine using selective covering of parts of engine with high transmission loss material which is usually lead. The noise increase is then noted by removing lead cover from component. The procedure is repeated for one by one for all components. Figure 5.7 shows results of such a test done on a 6 cylinder naturally aspirated diesel engine [29]. Total sound power level of this engine was found to be 114 dBA with valve cover, muffler, front gear cover and oil pan cover contributing about 21%, 10%, 8% and 7% respectively. b. Surface vibration method: The A weighted sound power level of engine (Lw[A]) is given in terms of acoustic impedance ( c), surface velocity (u), radiation efficiency ( ) and surface area (S) by [29]:

Lw[A] = 10

log( c) + 10

log(S) + 10

log(σ) + 10

log (u) (5.14)

SPL after uncovering 106

None

1 m SPL dBA

104 Front gear cover

102 100

Valve cover

98 96

Oil pan

94 Muffler

92 90

Average

Above Front Left Micro phone position

Figure 5.7 Noise analysis using lead cover method.

Right

Sound pressure level (dB SPL)

NVH Features of Engines 130 120 110 100 90 80 70 60 50 40 30 20 10 0 –10

155

(estimated) 100 phon 80 60 40 20 (threshold)

10

100

1000

10k

100k

Figure 5.8 Equal loudness contours (grey) (from ISO 226:2003 revision) original ISO standard shown (dark grey) for 40 phons.

Radiation efficiency is the ability of surface vibrations to convert into air borne noise. This is also related to critical frequency of component which may be defined as frequency at which wavelength of vibrations in structure matches with wavelength of radiated vibrations. At frequencies lower than critical one, the radiation efficiency is less than unity and vice versa. The dominant range of critical frequency for components of a diesel engine lies in range 400–800 Hz. The radiation efficiency can be estimated by considering engine as a radiating rigid sphere. This efficiency rises at the rate of 40 d B/decade at frequencies lesser than critical frequency. The value of critical frequency occurs when kr ≈ 4, where k is wave number of sound waves and r is radius of imaginary sphere that has same volume as that of engine. The measurement of surface vibrations is carried out best by use of accelerometers mounted on the engine surface. Positioning of accelerometers must be carefully done, as surface vibrations vary with wall thickness and proper balancing between less and strong sensitive measurement points is necessary. The surface velocity can be calculated by first converting acceleration data into frequency domain using Fast Fourier Transformation (FFT) and then carrying out integration. Figure 5.8 shows the results of experiments of various components towards Sound pressure level (SPL) obtained by surface velocity method [28]. It can be seen that contributions of valve cover, muffler shell, gear cover and oil pan cover is 29%, 20%, 4% and 15% respectively.

156 Liquid Piston Engines Sound power level Lw/dBA

112

1 Valve cover 2 Muffler 3 Front gear cover 4 Oil pan 5 Others

110 108 106 104 102 100

5

3

2

1

4 20 40 60 80 Ratio to total acoustic power/%

0

100

Figure 5.9 Noise analysis using vibrational analysis method. M P

H

W

C

~ C

+

G

E

–

Figure 5.10 Engine noise model. Combustion noise

Cylinder pressure Wiener filter

Residual noise

Noise signal

Figure 5.11 Application of wiener filter for estimation of combustion noise.

c. Use of spectro filters [30]: Diesel engines produce a complex level of noise levels of which combustion based noise is of major interest. The residual noise produced by various other sources is known as mechanical noise. Once in cylinder pressure signal is known these two sources can be separated using Wiener spectro-filters. These filters extract noise sources that are correlated with in cylinder pressure signals hence providing an estimation of combustion noise. The spectro-filter also called Wiener filter is a single input single output system whose impulse response is denoted by H(t). The input and output of system is denoted by P(t) and C(t) respectively corrupted by component M(t). The model is described by following equations: C(t) = P(t)

H(t)

(5.15)

NVH Features of Engines P1

H1

P2

H2

C1

157

M D

C2

Figure 5.12 Dual cylinder engine noise model.

G(t) = M(t) + C(t)

(5.16)

The spectro-filter H(t) can be estimated from following equations:

W( f )

sPD ( f ) sPP ( f )

W(t) = FT 1[W(f)]

(5.17)

(5.18)

In these equations SPP(f) denotes auto spectrum of P(t) whereas SPD(f) denotes cross spectrum of P(t) and D(t). Convolution of input P(t) with W(t) gives an estimate of C(t). i.e. C(t) = P(t)

W(t)

(5.19)

M(t) = D(t)

C(t)

(5.20)

In case of a mono cylinder engine C(t) denotes combustion noise, P(t) denotes in cylinder pressure, M(t) denotes mechanical noise, D(t) denotes total noise emissions and H(t) denotes transfer function between in cylinder pressure and noise emissions. In case of dual cylinder engine, Figure 5.12 depicts the corresponding engine noise model, which is a multiple input–single output (MISO) system. The combustion noise C(t) is now the sum of the two combustion noises produced by each cylinder. i.e. C(t) = C1(t) + C2(t)

(5.21)

5.6 Summary According to an estimate in 2006, the automotive industry had turnover of about 1 Trillion U.S. $ per year with an annual growth of about 6%[24].

158 Liquid Piston Engines Attributes such as durability and serviceability require a vehicle to be in service for a certain period of time. Costs of vibration and noise control are usually very high. e.g. the costs of warranty for brake was about 1 Billion U.S.$. per year during year 2005. There are a wide range of non-intrusive methods that can be used to evaluate NVH performance of combustion engines. In this work results from various experiments done on a water cooled diesel engine have been discussed which can be used to device strategies for control noise emissions from engine as well as its condition monitoring.

References and Bibliography 1. Davis, S.C., Transportation Energy Data Book: Edition 18, Report No. ORNL6941, 1998. 2. De Cicco, J., and Mark, J., “Meeting the Energy and Climate Challenge for Transportation in the United States,” Energy Policy, Great Britain: Elsevier. Vol. 26, No. 5, pp. 395–412, 1998. 3. John, A., Department of Transportation (DOT) Briefing Book on the United States Motor Vehicle Industry and Market, Version 1, Volpe National Transportation Systems Center, Cambridge, 1991. 4. Sperling, D., New Transportation Fuels: A Strategic Approach to Technological Change. Berkeley, Calif.:University of California Press, 1988. 5. Cronk, S., Building the E-Motive Industry: Essays and Conversations about Strategies for Creating and Electric Vehicle Industry. Warrendale, Penn.: Society of Automotive Engineers, 1995. 6. Evangelo, Rakopoulos, “Experimental study of combustion noise radiation during transient turbocharged diesel engine operation”, Energy, Volume 36, Issue 8, pp. 495–499, 2011. 7. Walsh, Michael, P., “Global Trends in Diesel Emissions Control – A 1997 Update”, SAE Paper No. 970179, 1997. 8. Krieger, R.B., Stewart, R., Pinson, K., Gallopoulos, N., Hilden, Monroe, D., Rask, Solomon and Zima, “Diesel Engines: One Option to Power Future Personal Transportation Vehicles,” Proceedings of the Diesel Engine Emissions Reduction Workshop, La Jolla, July 28–31, Washington, 1997. 9. Wang, M., Stork, K., Vyas, A., Mintz, M., Singh, and Johnson, “Assessment of PNGV Fuels Infrastructure, Phase 1 Report: Additional Capital Needs and Fuel-Cycle Energy and Emissions Impacts’’, 1997. 10. Walsh, Michael, P., “Global Trends in Diesel Emissions Control – A 1998 Update.” SAE paper No. 980186, 1998. 11. Sonya, G., Labelle, A., Special report TD Economics, U.S. auto sales basking in their comeback glow, 2014.

NVH Features of Engines

159

12. Kondo, M., Kimura, S., Ηirano, I., Uraki, Y., Maeda, R., ‘’Development of noise reduction technologies for a small direct-injection diesel engine’’, JSAE Review, Vol 21, pp. 327–33, 2001. 13. Torregrosa, AJ., Broatch, A., Novella, R., Monico, LF., “Suitability analysis of advanced diesel combustion concepts for emissions and noise control’’, Energy, Vol 36, pp. 825–38, 2011. 14. Shi, Y., Qiao, X., Ni, J., Zheng, Y., Ye, N., “Study on the combustion and emission characteristics of a diesel engine with multi-injection modes based on experimental investigation and computational fluid dynamics modelling’’, Proceedings of intuition of Mechanical Engineers, Journal of Automobile Engineering, Vol 224, pp. 1161–76, 2010. 15. Mohamad, S., Qatu, Mohamed, Abdelhamid, K., Pang, J., Sheng, G., “Overview of automotive noise and vibration”, Int. J. of Vehicle Noise and Vibration, Vol. 5, No. 1, pp. 1–35, 2009. 16. Genuit, K., “The sound quality of vehicle interior noise-a challenge for NVH engineers’’, Int. J. of Vehicle Noise and Vibration, Vol. 1, pp. 58–68, 2004. 17. Warring, R.H., Handbook of noise and vibration control, Trade and Technical press, Modern, Surry, U.K, 1985. 18. Anderton, D., Baker, J., “Influence of operating cycle on noise of diesel engines’’, SAE paper no. 730241, 1973. 19. Carlucci, P., Ficarrela, A., Laforgia, D., “Study of the influence of the injection parameters on combustion noise in a common-rail diesel engine using ANOVA and neural networks’’ SAE paper no. 2001-01-2011, 2001. 20. Pischinger, F., Schmillen, K., Leipold, FW., “A new measuring method for the direct determination of diesel engine combustion noise’’SAE paper no. 790267, 1979. 21. Priede, T., “Problems and developments in automotive engine noise research”, SAE paper no 7902, 1979. 22. Hickling, R., Kamal, M., Engine Noise – Excitation, Vibration and Radiation, New York, London – Plenum Press, 1982. 23. D’Anna, T., Govindswamy, K., “Aspects of Shift Quality With Emphasis on Powertrain Integration and Vehicle Sensitivity”, SAE N&V Conference 2005, Transverse City, MI, 2005. SAE-Paper 2005-01-2303, 2005. 24. Sheng, G., Vehicle Noise Sound Vibration and Sound Quality, SAE international, 2012. 25. Anderton, D., Noise source identification techniques, ISVR course notes, 2003. 26. Anderton, D., “Relation between Combustion System ad Noise’’, SAE paper no. 790270, 1979. 27. Russell, M., Haworth, R., “Combustion noise from high speed direct injection diesel engines’’, SAE paper no. 850973, 1985. 28. Yuehui, Liu., “Engine noise source identification with different methods”, Transactions of Tianjin University, Vol 8, issue 3, pp. 174–177, 2002.

160 Liquid Piston Engines 29. Grover., Lalor, “A review of low noise diesel engine design at I.S.V.R.’’, Journal of Sound and Vibration, Volume 28, Issue 3, 8 June 1973, Pages 403–428, 1973. 30. Pruvost, L., Leclere, Q., Parizet, E., “Diesel engine combustion and mechanical noise operation using an improved specto filters’’, Mechanical Systems and signal processing 23, pp. 2072–2087, 2009. 31. http://www.fev.com/fileadmin/user_upload/Media/Spectrum/en/spectrum21.pdf 32. http://luis.lemoyne.free.fr/Indicating_Product%20Overview_2009.pdf.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

6 Diagnosis Methodology for Diesel Engines

6.1 Introduction Various events occur in combustion engines cycle that can lead to increased emissions, fuel consumptions and potential damage to engine. A fuzzy based pattern recognition has been used for monitoring of various injection events [1]. Such diagnosis includes analysis of injection pressures and patterns using pressure transducers [2]. Misfires in engines have been analyzed using variations in engine angular speed [3]. Rizzione [4] and Connolly [5] have proposed a torque based algorithm for detection of misfire. Various energy based models have been studied by Tinuat et al [6]. A wavelet based method has been proposed for localization of engine misfires [7]. Engine knock caused due to spontaneous ignition of mixture during combustion leading to chamber resonance is another major event that needs proper diagnosis in diesel engines. A joint time frequency method has been studied in [8] to detect knock process. Suitable knocking index has been defined using band pass filtered cylinder pressure signals [9]. Wavelet analysis [10] and Fourier analysis [11] has also been used for detection 161

162 Liquid Piston Engines of knock. Features of various events can be extracted from vibration signals in time domain [12], frequency-domain [13] and time-frequency domain [14]. The frequency-domain analysis is generally considered the most adequate signal-processing tool for the non-stationary diesel engine vibrations. In this part of work some signal processing methods of diagnosis have been analyzed with focus on power spectral density function and time-frequency function analysis.

6.2 Power Spectral Density Function Power density function (PSD) of a random process provides the frequency composition of data in terms of spectral density of its mean square value [15]. The mean square value of a time sample in frequency range , + can be obtained by passing sample through a band pass filter with sharp cutoff frequency features and computing the average of squared output from filter. The average square value will approach a mean square value as . i.e. T T 2

( ,

) Lim

0 x

x 2 (t )dt T

(6.1)

6.3 Time Frequency Analysis Fourier transformation of a function f(t) is given by: t

f( )

f (t )e

j t

(6.2)

0

This analysis is useful as long as frequency content of signals do not vary with time. Hence time-frequency analysis or wavelet analysis is more suitable for analysis of noise and vibration signals emitted from engine [16]. Time frequency analysis is suitable for noise component having slow frequency changes such as those generated during engine ramp down whereas wavelet analysis is more suited for those signals having fast frequency changes such as those generated during rattle [17]. In the time frequency analysis the signal is windowed into small intervals and then Fourier transformation is taken for each interval [18]. Length of window can be used to change the resolution of output. A shorter window has

Diagnosis Methodology for Diesel Engines 163 high time resolution, but poor frequency resolution and vice versa. High time resolution at higher frequencies of wavelet transformation makes it possible to resolve short consecutive events. The short time frequency analysis is based on expansion of signal into a set of weighted frequency modulated Gaussian functions. It is given by: T

STFT ( , f )

x(t )h * (t

)e

j t

dt

(6.3)

0

where x(t) is input signal & h(t ) is window function. A Wigner Ville function has following quadratic time frequency distribution given as [19]: T

STFT ( , f )

x t 0

2

x* t

2

e

j t

dt

(6.4)

6.4 Wavelet Analysis Wavelet analysis map a signal on time frequency plane and are sensitive towards transient signals. One major drawback of time frequency signals processing methods is that they produce ripples hence making it difficult extract valuable information [20]. In wavelet analysis frequency resolution is better at low frequencies and time resolution is better at higher frequencies. Hence Wavelet analysis results are more accurate [21]. For wavelet transforms the signal is projected onto a family of zero mean functions known as wavelets. These have high time resolution and have no crossterm interference. The power spectral Similar to the short-time-frequency analysis, wavelet transform is a linear time-frequency transformation. The squared wavelet transform is called a scalogram. A single scalogram can easily cover audible frequency range with a time resolution of approximately 0.1 ms for the high-frequency components [22]. This makes the scalogram suitable for such various signals like squeak and rattle noise in an automobile interior, for which a wide range of frequency analysis is needed. Mathematically a complex wavelet transform is defined for a function f(t) as [23]:

CWT (a, b)

f (t )

1 a

(t b) dx a

(6.5)

164 Liquid Piston Engines where ψ(t): Mother wavelet f(t): Analyzed signal a: Scaling factor b: Shifting factor CWT(a, b): wavelet coefficients Mother Wavelet function (t) must satisfy following conditions: a. This function has zero average and decays exponentially to zero. i.e.

(t )dt

0

(6.6)

b. This function and its Fourier transformation must satisfy admissibility condition. i.e. ˆ(t )2

|f |

0

(6.7)

Both dilation as well as translation parameters in CWT are subjected to variations which makes the use of this methodology more complex. Discretization of signals helps to reduce this problem to certain extent. The CWT of a signal discrete signal Xm is defined in terms of sampling time Δt and sample data points m, n as: N 1

CWT

Xm * m 0

(m n) t Xj

(6.8)

where t = m Δt, b = n Δt, m & n varies from 0, 1, 2 … N 1, N

6.5 Conclusion This part of the work investigated the effect of changing injection parameters on noise emissions and combustion pressure. Noise emitted from engine depends on the quantity of fuel injected inside cylinder. Various

Diagnosis Methodology for Diesel Engines 165 ranges of noise signals sources were identified. Time–Frequency analysis showed the onset of various events of engine cycle. Based on the identification of various frequency bands it is possible to filter the signals in order to extract more information about combustion and mechanical based noise events for detailed analysis which is discussed in a later part of this work.

References and Bibliography 1. Sharkey, A. J. C., Chandroth, O., and Sharkey, N., “A multi-net system for the fault diagnosis of a diesel engine”, Neural Computing & Applications, 9(2), pp. 152–160, 2000. 2. Payri, F., et al., “Injection diagnosis through common-rail pressure measurement”. Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering, 220(3), pp. 347–357, 2000. 3. Azzoni, P. M., et al., 1996, “Misfire Detection in a High-Performance Engine by the Principal Component Analysis Approach”, in SAE International Congress & Exposition. Detroit, MI, 1996. 4. Rizzoni, G., “Estimate of indicated torque from crankshaft speed fluctuations: A model for the dynamics of the IC engine”. IEEE transactions on vehicular technology, 38(3), pp. 168–179, 1989. 5. Connolly, F. T., and Rizzoni, G., “Real time estimation of engine torque for the detection of engine misfires”, Journal of Dynamic Systems, Measurement, and Control, 116, pp. 675, 1994. 6. Tinaut, F. V., et al., “Misfire and compression fault detection through the energy model”, Mechanical Systems and Signal Processing, 21(3), pp. 1521–1535, 2007. 7. Chang, J., Kim, K., and Min, K., “Detection of misfire and knock in spark ignition engines by wavelet transform of engine block vibration signals”, Measurement Science and Technology, Vol 13, pp. 1108, 2002. 8. Samimy, B., and G. Rizzoni, “Mechanical signature analysis using time-frequency signal processing: application to internal combustion engine knock detection”, Proceedings of the IEEE, 84(9), pp. 1330–1343, 1996. 9. Borg, J., “Wavelet-based knock detection with fuzzy logic”, IEEE International Conference on Computational Intelligence for Measurement Systems and Applications. Giardini Naxos, Italy, pp. 26–31, 2005. 10. Zhang, Z., and Tomita, E., “Knocking detection using wavelet instantaneous correlation method”, Japan Society of Automotive Engineers (JSAE) review, Vol 23, pp. 443–449, 2002. 11. Lee, J., “A new knock-detection method using cylinder pressure, block vibration and sound pressure signals from an SI engine”, SAE International Fuels & Lubricants Meeting & Exposition, M. I, 1998. 12. Ftoutou, E., Chouchane, M., and Besbès, N., “Internal combustion engine valve clearance fault classification using multivariate analysis of variance and discriminate analysis”, T. I. Meas. Control, 2011.

166 Liquid Piston Engines 13. Carlucci, A. P., Chiara, F., and Laforgia, D., “Analysis of the relation between injection parameter variation and block vibration of an internal combustion diesel engine“, Journal of Sound and Vibrations, Vol 295, pp. 141–164, 2006. 14. Wang, C., Zhong, Z., Zhang, Y., “Fault diagnosis for diesel valve trains based on time-frequency images”, Mech. Syst. Signal Pr., Vol 22, pp. 1981–1983, 2008. 15. Sheng, G., Vehicle Noise Sound Vibration and Sound Quality, SAE international, 2012. 16. Cohen, L., “Time-Frequency Distributions – A Review”, Proceeding of the IEEE, Vol. 77, No. 7, 1989. 17. Ball, A., Gu, F., Weidong, L., “The Condition Monitoring of Diesel Engines using Acoustic Measurements, part 2: Fault Detection and Diagnosis”, SAE Special Publication SP 1501, 2000. 18. Daubechies, I., Ten Lectures on Wavelets. Philadelphia: Society for Industrial and Applied Mathematics, 1992. 19. Chiollaz, M., and Faver, B., “Engine Noise Characterization with WignerVille Time-Frequency Analysis”, Journal of Mechanical Systems and Signal Processing, Vol 75, pp. 375–400, 1993. 20. Chiatti, G., Chiavola, O., Fulvio, P. and Andrea, P., “Diagnostic methodology for internal combustion diesel engines via noise Radiation’’, Energy Conversion and Management, Vol 89, pp. 34–42, 2015. 21. Chiatti, G., and Chiavola, O., “Combustion Induced Noise in Single Cylinder Diesel Engines”, Small Engine Technology Conference Graz, Austria September 27–30, SAE Paper no. 2004-32-0071, 2004. 22. Chiatti, G., Chiavola, O., “Experimental analysis of combustion noise in spark ignition engine”, NVC conference, Tranverse city, Michigan, USA, May 5-8, SAE paper no. 2003-01-1442, 2003. 23. Wu, J., Chen, J., “Continuous wavelet transform technique for fault signal diagnosis of internal combustion engines”, NDT International Vol 39, pp. 304–31, 2006. 24. Wang, J., McFadden, P., “Application of Wavelet to Gearbox Vibration Signals for Fault Detection”, Journal of Sound and Vibration 192 (5), pp. 927–939, 1996. 25. Lin, J., Zuo, M., “Gearbox Fault Diagnosis Using Adaptive Wavelet Filter. Mechanical Systems and Signal Processing”, 17(6), pp. 1259–69, 2003. 26. Smith, T., “The Application of the Wavelet Transform to the Processing of Aeromagnetic Data”, PhD Thesis, The University of Western Australia, 2000.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

7 Sources of Noise in Diesel Engines

7.1 Introduction NVH study of a diesel engine is important from point of view of overall design/planning of system. It includes testing variations under various testing conditions as well as engine-to-engine variations. The theory of NVH applications in diesel engines was introduced by Reinhart et al. [1]. In modern times there are five major simulation tools that have been used in the NVH analysis which includes: 1. 2. 3. 4. 5.

Boundary element method (BEM), Gas dynamics, Multi-body dynamics, Engine cycle simulation and Finite element analysis.

There are several sources of noise in diesel engines some of which are discussed in following sections. 167

168 Liquid Piston Engines

7.2 Combustion Noise Combustion noise is an important concern for design and performance calibration of system design and forms a primary source of noise for direct injection diesel engines. There are three modes of transfer of this noise to surroundings which include from cylinder top, liner walls and connecting rod assembly. It is related to rate of pressure rise in cylinders. Tung and Crocker have studied combustion noise in turbocharged diesel engines [2]. Structural attenuation also affects the combustion noise radiated from engine which is difference between in cylinder pressure spectrum and noise radiated from surface. High structural stiffness of cylinder bore leads to higher values of resonant frequencies. Higher values of resonant frequencies than combustion excitation frequencies can help to attenuate high frequency combustion noise. Knocking or clatter is also an important source of noise in low speed engines. Combustion noise can be reduced either by increasing the structural attenuation of engine or by reducing in cylinder pressure. Reduction in delay period in ignition leads to lower values of in cylinder pressure and hence combustion noise. Other factors which can help reduce the combustion noise include higher compression ratio, increased intake boost pressure, higher exhaust gas recirculation rates and increased structural attenuation of parts. The transfer function of engine structural attenuation has been experimentally determined by Shu [3]. Combustion noise optimization by use of combustion noise meter was done by Wang [4]. Combustion noise assessment by decomposition of cylinder pressure has been done by Torregrosa [5] showing that this method is more accurate as compared to traditional block attenuation curve. Effect of cetane number on engine noise was measured by Machado and De Melo [6]. Engine noise during cold start of engine was analyzed by Alt [7]. Cylinder CFD modeling of combustion noise has been conducted by Blunsdon [8] and Luckhchoura [9]. Cyclic variations of combustion noise variations were studied by Gazon and Blaisot [10]. Analysis of noise emissions from a medium sized diesel engine has been performed in [11, 12].

7.3 Piston Assembly Noise There are three types of noises that occur in piston assembly. These include pin tickling noise, piston slap and piston rattle noise. Piston slap is a major contributor towards mechanical noise sources in diesel engines which is caused due to secondary motion of piston between skirt and liner bore.

Clearance between piston skirt top corners and cylinder bore at thrust side (micron)

Sources of Noise in Diesel Engines

200

169

Piston skirt corner 3 Piston skirt corner 4

150

1

4

100 3

50

Piston in bore

2 Scraping the bore

Slap

0

0

90

180

270 360 450 Crank angle (degree)

540

630

720

Figure 7.1 Simulation of piston secondary motion.

There may be several events of slap in an engine cycle with most prominent occurring just after TDC firing position as shown in Figure 7.1. Piston slap is affected by the following major factors which influence piston secondary motion: a. Piston side thrust force-lower speeds, lower piston assembly mass, lower in cylinder pressures and higher crank radius to connecting rod length ratio can help to reduce side thrust force and hence piston slapping noise [13]. b. Moments about piston pin- lower inertia of assembly, proper piston pin offset and crankshaft offset, proper supply of the lubricant and the piston pin friction force moments can help to reduce the piston slap noise Munro and Parker [14]. c. Allowable distance of travel before hitting liner wall – smaller gap between piston skirt-to-bore can help to reduce the slapping noise at the expense of increased shear frictional force. d. Oil damping force –sufficient supply of oil on the skirt can help to reduce piston slap significantly. Lower tension, longer skirt length and increasing the contact area in the piston rings can help to increase the oil film thickness and hence piston slap [15]. e. The stiffness and the damping of the parts –impact of softer piston skirt causes lesser noise emissions due to due to larger deformations. It is important to increase the gap between top land and bore in order to avoid the contact which otherwise would produce sharp rattling noise.

170 Liquid Piston Engines The slap noise is most common at idle cold start conditions and high load-low speeds. The proper skirt design is an important measure to minimize the piston slap noise The two most commonly used methods include reducing the gap between liner and skirt and offsetting the piston pin.

7.4 Valve Train Noise This type of high frequency noise is a major issue in NVH analysis of engine and includes the following three major excitation sources: a. Cam acceleration-The opening and closure of cam excites high frequency vibrations at high speeds due to inertial forces. b. Valve train impacts-These include impulsive impacts at valve opening between cam and follower, between valve seat and valve at closure of valves and valve train separation and bouncing at high speeds. c. Frictional Vibrations-This noise is dominant at low speeds when asperity contact occurs cam and follower near nose of cam when lubricant velocity becomes zero. Valve train noise identification was done using valve acceleration–cam angle diagram and a time–frequency diagram [16]. Anderton and Zheng [17] found that the valve train vibrations had a major contributions towards total noise at high engine speeds ranges of above 2000–3300 RPM. Savage and Matterazzo [18] has done experiments on a 3.3 L gasoline engine to show the effects of various factors like cam jerk level, valve spring load, tappet-to-bore clearance, valve stem gap and surface finish, rocker arm bearing clearance, valve overlap, and cylinder head mass and damping. Use of high precision manufacturing of cam profile, greater oil film thickness, higher valve train stiffness and smaller tappet-to-bore clearance is very important for reduction in valve train noise [19, 20].

7.5 Gear Train Noise The rattle noise in transmission drive trains is primary cause of concern in NVH development. The engine transmits non uniform torque from crank train to drive train that causes gear rattle noise. Clearances are provided between meshing tooth of gears to account for thermal expansions and

Sources of Noise in Diesel Engines

171

manufacturing tolerances. When gears are lightly loaded at low speeds with strongly oscillating torque, there is a high chance of meshing teeth separation and resulting vibrational impacts. Other type of gear noise include whine noise which is due to tooth deflection under load. Gear train operational noise is dependent upon number of meshing teeth, size of gear train, magnitude of torsional inputs and location of gear train. Detailed investigation of gear train noise was done by Spessert and Ponsa [21] and Zhao and Reinhart [22].

7.6 Crank Train and Engine Block Vibrations The torsional vibrations in crank shaft, the thin sections in the engine block and the covers and the cylinder block are important sources of noise and vibrations. Commercial software like ENGDYN are available to analyze the response of the crank train and engine block system taking into account oil film lubrication models [23]. Detailed analysis of crankcase and engine block vibrations was done by Russell [24], Ochiai and Yokota [25], and Maetani et al.[26].

7.7 Aerodynamic Noise Low frequency Intake noise is due to turbulent fluctuations in flow of air at inlet ducts which depends on intake valve opening area and engine speed. Gas dynamics based design of intake ducts has been done by Silvestri et al. [27]. Exhaust noise is due to pressure variations in exhaust duct due to periodic charging and discharging in engines. The noise due to mechanical vibrations in exhaust pipe is known as shell noise. Pang et al has described exhaust manifold design to control system vibrations [28]. Turbocharging noise also forms an important part of aerodynamic noise.

7.8 Bearing Noise Bearings in crankshaft as well as connecting rod have clearances which are likely to generate noise under action of external excitation forces [29]. One of major noise sources due to bearing effect is rumbling noise which is due to engine torsional and bending resonance induced by clearances. Figure 7.2 shows the key bearing parameters effecting rumbling noise. Key methods to control bearing noise include optimization of

172 Liquid Piston Engines Concord bearing clearance

Thrust bearing clearance

Bearing beam stiffness

Main bearing clearance

Figure 7.2 Various bearing parameters effecting engine noise.

clearances, application of optimal crank shaft damper and application of flexible flywheel design.

7.9 Timing Belt and Chain Noise Major sources of noise for timing chain include meshing impact and polygon effects [30]. The meshing frequency depends upon engine speed and number of meshing teeth. The polygon effect causes elevation and drop of chain element and hence leading to transverse and torsional vibrations of chain. The impact speed of roller(WA) and chain sprocket can be estimated in terms of its pitch (PH), number of teeth (D), pressure angle ( ) and number of sprocket teeth (Z) as:

WA

NPH 360 Sin Z 3600

(7.1)

The impact energy EA can estimated by linear density of chain (υ):

EA

WA2 PH 2000

(7.2)

Use of rubber ring sprocket in chain sprocket can help to reduce noise as shown in Figure 7.3(a). Combustion engines also have transmission belts systems (Figure 3.4) which can exhibit number of modes of vibrations as seen from Figure 7.5.

Sound pressure level (dB)

Sources of Noise in Diesel Engines

0.5p

Tooth

Compression

Rubber

100

(b)

Standard tooth

90 80

Asymmetric tooth

70 60 50

(a)

0

1000 2000 3000 4000 5000 6000 RPM

Figure 7.3 Schematic representation of timing chain and its noise spectra.

1 2 6 5 3 4

Figure 7.4 Timing belt transmission system(1: Sprocket, 2: Tensioner, 3: Fuel pump sprocket, 4: Crankshaft sprocket, 5: Idler sprocket, 6: Water pump sprocket).

(a) Transverse vibration

(b) Axial vibration

C

(c) Torsional vibration

Figure 7.5 Timing belt vibration sources.

173

C (d) Lateral vibration

174 Liquid Piston Engines Due to differences in pitch of tooth belt and sprocket timing belt, the meshing belt creates a meshing impact which is a periodic excitation having frequency (fs) estimated as:

ZD 60

fs

(7.3)

Typical spectrum of belt noise. Meshing noise is dominant at low and medium frequencies, whereas meshing frictional noise is dominant over 6 kHz frequency [31].

7.10 Summary Internal combustion engines have several contributing noise sources, such as noise due to combustion process, fuel injection process, piston slap noise and valve operation noise. It is necessary to separate the contributions due to various sources and then analyze each of them individually. Figure 3.7 Noise amplitude

fe

Time

SPL (dB)

Tz fn 80 60 40 30

fe

2fn fz 63

200

630 2000 Frequency (Hz)

6300

20000

Figure 7.6 Timing belt noise spectra. Noise sources

Vibration generation

Paths of transfer

Radiation of noise

Head vibrations In cylinder pressure

Gas force

Block vibrations

Piston slap Inertia of piston Vibrations in connecting rod

Figure 7.7 Mechanism of noise generation.

Crank shaft vibrations

Noise radiation

Sources of Noise in Diesel Engines 90

175

8

1

Noise level dB (A)

2 80

3

4

5

6 7

70

60

Figure 7.8 The total noise contribution (8) can be decomposed into contributions due to combustion noise (1), contribution due to piston slap noise (2), contribution due to fan noise (3), contribution to gear operation noise (4), contribution due to pump operations (5), valve noise (6), other sources (7).

depicts noise generation mechanism for various sources with a typical contributions of various sources of noise towards sound pressure level of engine measured at a distance of 1m away from engine in an anechoic chamber featured in Figure 7.8. As evident from the above figures the combustion based noise and piston slap contributes a major portion (about 80%) towards noise emissions from engines [31], hence it is necessary to focus on these two important aspects of engine acoustics as discussed in following units of this work.

References and Bibliography 1. Reinhart, T E., “An evaluation of the Lucas combustion noise meter on Cummins B series engines”, SAE paper 870952, 1987. 2. Tung,V., and Crocker, M J., “Diesel engine noise and relationship to cylinder pressure”, SAE paper 820237, 1982. 3. Shu, G., Wei, H., and Han, R., “The transfer function of combustion noise in DI-diesel engine”, SAE paper 2005-01-2486, 2005. 4. Wang, S.,Chalu, C., and Gautier, F., “Optimization of combustion noise of modern diesel engines for passenger cars”, SAE paper 2007-01-2379, 2007. 5. Torregrosa, A J., Broatch, A., Martin, J., and Monelletta, L., “Combustion noise level assessment in direct injection diesel engines by means of in-cylinder pressure components”, Measurement Science and Technology, 18, 2131–2142, 2007. 6. Machado, G B., and De Melo, T C C, “Diesel cetane number versus noise emission”, SAE paper 2005-01-2150, 2005.

176 Liquid Piston Engines 7. Alt, N., Sonntag, H., Heuer, S., and Thiele, R., “Diesel engine cold start noise improvement”, SAE paper 2005-01-2490, 2005. 8. Blunsdon, C A., Dent, J., and Das, S., “Modelling the origins of combustion noise in the indirect injection diesel engine”, SAE paper 952432, 1995. 9. Luckhchoura, V., Won, H., Sharma, A., Paczko, G., and Peters, N., “Investigation of combustion noise development with variation in start of injection using 3-dimensional simulations by applying representative interactive flamelet (RIF) model”, SAE paper 2008-01-0950, 2008. 10. Gazon, M., and Blaisot, J., “Cycle-to-cycle fluctuations of combustion noise in a diesel engine at low speed”, SAE paper 2006-01-3410, 2006. 11. Chiatti, G., Recco, E., Chiavola, O., and Conforto, S., “Acoustic Assessment in a Small Displacement Diesel Engine,” SAE Technical Paper 2014-32-0129, doi:10.4271/2014-32-0129, 2014. 12. Chiavola, O., Chiatti, G., and Recco, E., “Analysis of the Relationship between Noise Emission and in-Cylinder Pressure in a Small Displacement Diesel Engine,” SAE Technical Paper 2014-01-1364, doi:10.4271/2014-01-1364, 2014. 13. Oetting, H., Pundt, D., and Ebbinghaus, W., “Friction in the piston group and new ideas for piston design”, SAE paper 841299, 1984. 14. Munro, R.,and Parker, A., “Transverse movement analysis and its influence on diesel piston design”, SAE paper 750800, 1975. 15. Ryan, J P., Wong, V W., Lyon, R H., Hoult, D P., Sekiya, Y., Kobayashi, Y., and Aoyama, S., “Engine experiments on the effects of design and operational parameters on piston secondary motion and piston slap”, SAE paper 940695, 1994. 16. Suh, I-S., and Lyon, R H., “An investigation of valve train noise for the sound quality of IC engines”, SAE paper 1999-01-1711, 1999. 17. Anderton, D., and Zheng, J H., “A new measurement method for separating airborne and structure borne sound from an IC engine’s valve train mechanism”, SAE paper 931335, 1993. 18. Savage, J., and Matterazzo, J., “Application of design of experiments to determine the leading contributors to engine valve train noise”, SAE paper 930884, 1993. 19. Hanaoka, M., and Fukumura, S., “A study of valve train noises and a method of cam design to reduce the noises”, SAE paper 730247, 1973. 20. H. Kanda, M. Okubo, T. Yonezawa, “Analysis of noise sources and their transfer paths in diesel engines,” SAE Technical Paper 900014, 1990. 21. Spessert, B., and Ponsa, R., “Investigation in the noise from main running gear, timing gears and injection pump of DI diesel engines”, SAE paper 900012, 1990. 22. Zhao, H., and Reinhart, T E., “The influence of diesel engine architecture on noise levels”, SAE paper 1999-01-1747, 1999. 23. Offner, G., Priebsch, H H., Ma, M T., Karlsson, U., Wikstrom, A., and Loibnegger, B., “Quality and validation of cranktrain vibration predictions  – effect of

Sources of Noise in Diesel Engines

24. 25. 26.

27. 28.

29.

30.

31.

177

hydrodynamic journal bearing models”, Multi-Body Dynamics: Monitoring and Simulation Techniques-III, pp. 255–271, 2004. Russell, M.F., “Reduction of noise emissions from diesel engine surfaces”, SAE paper 720135, 1972. Ochiai, K., and Yokota, K., “Light-weight, quiet automotive DI diesel engine oriented design method”, SAE paper 820434, 1982. Maetani, Y., Niikura, T., Suzuki, S., Arai, S., and Okamura, H., “Analysis and reduction of engine front noise induced by the vibration of the crankshaft system”, SAE paper 931336, 1993. Silvestri, J., Morel, T., and Costello, M., “Study of intake system wave dynamics and acoustics by simulation and experiment”, SAE paper 940206, 1994. Pang, J., Kurrle, P., Qatu, M., Rebandt, R., and Malkowski, R., “Attribute analysis and criteria for automotive exhaust systems”, SAE paper 2003-01-0221, 2003. Qatu, M. S., Abdelhamid, M. K., Pang, J., and Sheng, G., “Overview of Automotive Noise and Vibration,” International Journal of Vehicle Noise and Vibrations, Vol. 5, Nos. 1/2, 2009. Young, J. D., Marshek, K. M., Poiret C., and Chevee P., “Camshaft roller chain drive with reduced meshing impact noise levels,” SAE Paper No. 2003-011666, SAE International, Warrendale, PA, 2003. Sheng, G., et al., “A new mechanism of belt slip dynamic instability and noise in automotive accessory belt drive systems,” International Journal of Vehicle Noise and Vibration, Vol 2, pp. 305, 2006.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

8 Combustion Based Noise

8.1 Introduction Combustion noise generated mainly depends on rapid rise of cylinder pressure due to ignition delay period in combustion engines. Design of combustion chamber as well as variations in various injection parameters like injection pressure, amount of fuel injected and its timings also play a crucial role in noise emissions [1]. Depending upon type of engine and various operational parameters, overall noise emissions from a typical engine are in range 80–110 dBA [2]. Anderton has investigated the effects of turbo charging on noise emissions from engines [3]. Split injection using electronic control reduces the premixed combustion and hence is an effective way to reduce overall noise emissions by about 5–8 dBA [4]. Head and Wakes have shown that during transient operational conditions, overall noise levels are 4–7 dBA higher as compared with steady state operations [5]. Cold starting conditions lead to higher ignition delay period which in turn causes more premixed combustion and hence an increase in noise emissions [6]. Quality of fuel also affects combustion noise emissions from engines. It has been seen that reduction of centane number of diesel from 50 to 40 causes a rise of 3 dBA in combustion based noise [7]. In gasoline 179

180 Liquid Piston Engines engines, the ignition delay period is longer due to lesser compression ratio which leads to lower temperature of charge and hence more noise [7]. For a naturally aspirated engine the combustion noise depends upon amount of fuel that mixes with air charge during injection delay period and hence compression ratio of engine also plays a vital role [7].

8.2 Background of Combustion Process  in Diesel Engines Due to high efficiency, diesel engines have been a favorite choice for heavy duty applications including trucks [8]. However they suffer from drawbacks of high noise, weight and vibrations. These engines are of two types: Direct Injection (D.I.) Engines In direct injection engines In the D.I. engines, the fuel is directly injected inside combustion chamber and due to lesser time for mixing, a heterogamous mixture consisting of both rich and lean parts is formed in the chamber. Figure 8.1 shows the three phases of combustion in a conventional diesel engine. The first phase starts with start of injection process and ends with premixed combustion phase. The direct injection of fuel into combustion chamber begins some crank angle degree before top dead center positions depending upon engine operational conditions. As soon as cold jet of fuel penetrates into chamber, it mixes up with hot compressed air. The droplets begin to vaporize forming a sheath of vaporized fuel-air mixture around jet

· ) Injected mass rate (m fuel · Rate of heat release (q) · Cylinder pressure (pcyl)

Phase 1

Phase 2

Phase 3

pcyl

Premixed combustion premixed peak

Ignition delay

q· m· fuel 0

Injection

Figure 8.1 Phases of diesel engine combustion.

20

Crank angle ATDC (deg)

Combustion Based Noise

181

periphery. When temperatures reach around 750 K,the first break down of Cetane fuel occurs. Further reactions produce C2H2, C3H3, C2H4, CO2 and water vapors [9]. Resulting rise in temperature causes complete combustion of fuel-air mixture. This sudden combustion causes rise in heat release rate and high dP which further leads to high temperatures in pre-mixed pressure gradient d zone and NOx production. The premixed combustion consumes all mixture around inner spray region where temperatures of ranges 1600–1700 K are reached and all oxygen available for combustion is consumed [8]. Now various partial burnt particles diffuse to outer layers and are burnt in a thin reaction region at periphery of spray which leads to formation of diffusion flame. This kind of combustion is known as diffusion controlled combustion and is depicted by region 2 and 3 in the above figure. The high temperatures along with lack of oxygen is ideal for soot formation. The diffusion flame is fed with oxygen from surrounding environment and high temperatures of range 2700 K is reached consuming all the soot formed. At outer region of flame there is enough oxygen content for formation of NOx. Figure 8.3 shows soot formation concentration as a function of crank angle. Most of soot produced at early crank angles is consumed later and final exhaust emissions have only a fraction of initial one. As seen from Figure 8.1, the diffusion controlled combustion can be divided into two sub phases. During phase 2 of combustion the burning rate is controlled by mixing of fuel fragments and air and rate of reaction is must faster. During the phase 3 the final oxidation of remaining unburnt

Soot oxidation Rich fuel-air mixture, =4 825 K Diffusion flame CO2, H2O Air 950 K

Cold fuel 350 K Ignition zone

Products of rich combustion, CO, HC soot ~ 1600 K

Figure 8.2 Conventional diesel engine spray formation.

NOX formation ~ 2700 K

Soot concentration

182 Liquid Piston Engines

TDC

Crank angle

Figure 8.3 Rate of soot formation.

So tra ot-N de O -o X ff

EGR Low swirl Reduced Early SOI Intercooling Water injection Variable valve timing Increase of pinj, decrease of hole size Multiple injections (pre- and post-injections) Rate shapping s a Increase boost pressure g t t n s au atme h Ex tre er DeNox aft ...

Diesel particulate filter (DPF)

Soot emission

Late SOI

NOX emission

Figure 8.4 Soot & NOx trade off.

particles and soot takes place, however due to decrease of gas temperature during  expansion stroke and lack of oxygen the reaction rate is much slower. NOx and soot formation in combustion engines show opposite trends as shown in Figure 8.4. In order to reduce NOx, it is necessary that local temperatures must not rise above 2000 K. A possible way to do so is to inject fuel late in cycle inside combustion chamber which shifts combustion phase towards expansion phase and hence significant reduction in chamber temperatures. However consumption of fuel and soot formation increases due to late combustion. Hence modern injection systems use multiple injection techniques in order to reduce both soot as well as NOx emissions as seen from Figure 8.5 [8, 10, 11]. These generally use three phases of injection namely pre-injection Period, Main – Injection Period & Post injection period as seen from

Combustion Based Noise 2c

Injection pressure

2b 3a

2a

183

1 Pre-injection 2 Main-injection 2a rectangular type 2b boot type 2c ramp type 3 Post-injection 3a early 3b late 3b

1

p0 TDC

Crank angle

Figure 8.5 Multiple injection methods adopted for modern diesel engines.

Figure 8.5. There is delay period between the start of ignition process and fuel injected inside diesel engine. The more this ignition delay, more is the temperature during combustion and hence better condition for NOx formation. To shorten the delay period, small amount of fuel is pre-Injection before main injection during the pre-mixed combustion phase. The torque and power produced in engine depends upon main injection period. It is advantageous to vary injected fuel mass with time to reduce the specific fuel consumption. This method is known as rate shaping. Rate shaping may be rectangular, step or boot in shape. Post-injection of fuel is done to reduce soot emissions and in some cases may be useful for Exhaust gas recirculation treatment of gases [12]. It has been reported that post injection reduces soot by about 70% without increasing the fuel consumption [13].

8.3 Combustion Phase Analysis This part of work presents experimental data in which signals from accelerometers and cylinder pressure transducers are used for analysis of combustion behavior of engines. Previous works have shown that engine vibrations are sensitive towards change in engine injection parameters. Accelerometer signals have been able to locate the important features of combustion process in diesel engines [14, 15]. The aim of present part of work is to explore the relationship between block vibrations and in cylinder pressure development process. Experiments were done at various load and speed conditions to explore the sensitivity of vibration data.

184 Liquid Piston Engines

8.4 Combustion Based Engine Noise The aim of this part is to provide an overview of combustion noise generation process. This was done by analyzing the methodology provided in previous works done. Engine combustion noise originates from combustion process taking place in cylinder. When fuel is injected inside the combustion chamber where high pressure air is present, then part of ignitable gas starts to burn causing a rapid rise in pressure as well as temperatures inside combustion chamber. The pressure wave thus generated also strikes the walls of combustion chamber causing resonance of structure. The vibrations are radiated in air through engine structure and are perceived as combustion noise. In actual practice it is difficult to distinguish between combustion noise and piston slap as both coincide near top dead center positions. For sake of convenience it is assumed that combustion noise originates due to pressure vibrations inside engine cylinder and is transmitted to cylinder cover, piston, connecting rod, crank shaft and engine surroundings. Mechanical noise includes noise from piston-liner impact, valve operations, pump operations, injector operations as well as operation of various accessories and valve trains. Generally for indirect injection diesel engines and gasoline engines, the combustion process is less severe as compared with diesel engines, hence the combustion noise is lesser as compared with mechanical noise. The cause of combustion noise is rapid change in cylinder pressure during combustion process. The effects of combustion process consists of high frequency gas vibrations and dynamic load due to rapid pressure change. The intensity of combustion noise (I) is dependent upon the values of maximum pressure value (Pmax) and maximum rate of pressure rise as [20]:

I

dP dt

2

Pmax

(8.1)

max

In a direct injection diesel engine, combustion process occurs in following four phases: retarded, rapid, early and late phase. Combustion noise is mainly generated during rapid phase of combustion however retarded phase has an indirect effect on combustion noise. During this phase of combustion, the firing and fuel propagation results in impulsive pressure wave that gets reflected multiple times after striking walls of chamber. This process causes high frequency oscillations. The frequency (fg) can be estimated from engine bore diameter (D) & Wave propagation speed (Cc) by:

Combustion Based Noise

fg

Cc 2D

185 (8.2)

Further the spectrum plot of cylinder pressure can be obtained from acquired about in cylinder as seen in Figure 8.6. The graph is marked by following three distinct regions [20]:

Cylinder pressure (bar)

a. Region of low frequency-in this region the curve depends upon peak cylinder pressure. Higher the maximum value of cylinder pressure, higher is the peak in low frequency range. b. Medium frequency range-in this part the pressure levels decrease in logarithmic range with slope depending upon rate of cylinder pressure Larger the value of pressure gradient, steeper is the slope of curve. c. High frequency range-in this range rapid evolution of in cylinder pressure occurs due to onset of combustion process which results in high frequency vibrations of cylinder structure having amplitude dependent on second cylinder pressure derivative.

80 dp d max

60

max

40 pd

20 0

–20 –500 –400 –300 –200 –100 0 Crank angle

Cylinder pressure spectrum (dB)

d2p d 2

100

200

300

102.44 2.4

d2 p d 2

pd

10

102.36

dp d max

102.32 101 Frequency-Hz

Figure 8.6 Regions of combustion noise.

102

max

Attenuation (dB)

186 Liquid Piston Engines 100

80

0.5

1

2 5 10 Frequency (kHz)

Figure 8.7 Attenuation curve of engine.

Combustion noise is not only dependent upon cylinder pressure but also upon the structural response and damping effects of engine. The difference between in cylinder pressure and outside noise radiated in engines is characterized by a decay which reflects the structural attenuation of engine structure. Figure 8.7 shows a typical plot of structural attenuation of an engine which is independent of exciting forces and cylinder pressure spectrum. Various operational parameters of engine e.g. load, speed & fuel injection parameters have significant effects on structural attenuation property. This curve can be divided into three distinct regions: a. Below 2000 Hz-in this range the attenuation is very high as most of engine parts have high stiffness with their natural frequencies in low or mid ranges. b. In mid frequency ranges of 2000–5000 Hz, the attenuation is small as most of natural frequencies of engine parts fall in this range. c. Above 5000 Hz – natural frequencies of most of the parts is above the natural frequencies of most of the parts, hence attenuation is quite high.

8.5 Factors Effecting Combustion Noise Controlling the rate of pressure is the key to control combustion noise which mainly depends upon the ignition delay and quantities of combustion gas formed during delay period. A shorter delay period means lesser amount of combustible gas formed in the cylinder and hence lesser combustion noise. Hence delay period must be reduced as much as possible to reduce combustion noise. Structure and layout of engine also plays an important role.

Combustion Based Noise

187

Increase in compression ratio and chamber temperature also shortens the delay period. However increase in compression ratio may cause rise in the piston slap noise. Various parameters of fuel injection system like angle of fuel injection, injection pressure, number of nozzles and fuel supply rate also effect the combustion noise in engines. An increase in the injector pressure leads to an increase in the amount of fuel accumulated in the ignition delay period and hence an increase in the combustion noise. If other factors remain unchanged, an increase in engine speed increases the fuel injection rate and hence greater quantity of fuel is injected in the period of combustion delay which leads to an increase in combustion noise. There are many approaches to control combustion noise. One of these includes reducing cylinder pressure spectrum typically in middle and high frequency ranges. Others include reducing ignition delay period or amount of combustible gases formed during this period. Increasing the stiffness of parts, use of turbo charging process and use of split injection methods in engine cycle have also proved to be effective methods.

8.6 In Cylinder Pressure Analysis The analysis of in cylinder pressure is one of the most common ways to examine the combustion noise in diesel engines. Anderton and Priede were first to observe that abrupt combustion led to high frequency contents of in cylinder pressure spectrum [24]. Crocker suggested that frequency contents up to 300 Hz were related to maximum cylinder pressure [25]. Between 300–2000 Hz they were related to first derivative of in cylinder pressure, whereas above 2000 Hz were related to second derivative of cylinder pressure.

8.7

Effects of Heat Release Rate

Previous works have shown a relationship between peak of combustion noise and overall heat release rate [26]. Rusell has developed a technique based on block attenuation curve which is still most reliable one for study of combustion noise [27]. It was observed that higher slopes of rate of heat release curves led to higher combustion noise irrespective of fuel injection timings [28]. There is a tradeoff between combustion efficiency and noise generated due to combustion [29]. Efficient combustion leads to higher heat release rate near top dead center which gives rise to high frequency components in noise spectrum. Late heat release rates leads to lower in cylinder pressures and hence low frequencies in combustion noise.

188 Liquid Piston Engines Lean slope Sharp slope

Cylinder pressure (bar)

100

Heat release rate 80 150 kJ/kg-deg. 60 100

50

A weighted SPL

40 50 20 0 0 –60°–30° tdc 30° 60° 90° Crank angle

20

50

100 200 500 1k Crank angle

2k

5k

10 kHz

Figure 8.8 Effects of heat release rate on combustion noise.

180 170

dB

160 150 140 130 120 110 3 10

Cycle 1 Cycle 2 Cycle 3 Frequency (Hz)

104

Figure 8.9 Cyclic variations in combustion noise.

It was observed that a 10 dB reduction in sound pressure levels was possible without change in fuel consumption. However, there is a fall in efficiency of cycle and smoke emissions increase if ROHR is not terminated 50° BTDC [27].

8.8

Effects of Cyclic Variations

Combustion process in diesel engines vary with cycle which leads to variations in combustion noise [30]. These variations may be attributed due to different fuel injection rates, compression ratios as well as difference in fuel spray process, mixture formation and flame propagation. Torregrosa has studied cyclic variations in combustion noise as given by Figure 8.9 [31]. Variations may be attributed to resonance phenomenon occurring in combustion chamber.

Combustion Based Noise First circumferential mode (1.0)

Second circumferential mode (2.0)

189

First radial mode (0.1)

Figure 8.10 Various modes of combustion chamber cavity.

8.9 Resonance Phenomenon Resonance process taking place in the combustion chamber effects the noise radiated from engines. Grover has observed high peaks in noise spectrum which may be attributed to this phenomenon [32]. Previous works have shown effects of gas temperature and affects of resonant frequency on high amplitude combustion chamber oscillations [33]. Hickling has observed several peaks in cylinder pressure spectrum by filtering data with cutoff frequencies in range 20–1500 Hz which increase in amplitude with increase of load [34]. The frequencies of these peaks varied inversely with cylinder bore diameter. The resonant frequency (fr) may be defined in terms of cylinder bore (B), axial length (L) and speed of sound (C) as:

fr

K 2L

2

qm,n

2

B

(8.3)

where m, n, k determine circumferential, axial and radial modes.

8.10 In Cylinder Pressure Decomposition Method The signal processing methodology to decompose cylinder pressure signals was first proposed by payri [35]. In this methodology the cylinder pressure signals were decomposed into three parts namely-Combustion pressure, resonance pressure and compression pressure. The combustion art which is strongly dependent on rate of heat release is defined by injection strategy in engines. Using suitable cutoff frequencies motored part of

190 Liquid Piston Engines Total pressure Motored pressure Excessive pressure

40

102.8 Pressure level (dB)

Cylinder pressure -bar

60

20

0

102.7

Total pressure Motored pressure Excessive pressure

–20 –360

–240

–120 0 120 Crank angle

240

360

102

Frequency (Hz)

103

Figure 8.11 Time and frequency decomposition of cylinder pressure signal.

pressure signal was isolated from excessive pressure signals. Figure 8.11 shows results of such a methodology adopted for a cylinder pressure signal. As evident from these figures, motored pressure signal dominates the pressure spectra at low frequency ranges. By subtracting the motored pressure signal from total pressure signal, the excessive part of signal can be obtained. The resulting signal has contributions due to both resonance as well as combustion pressures. Resonance portion is clearly visible with fluctuating peaks in pressure spectra. Hence it can be isolated by filtering excessive pressure signal using high pass filter. The resulting resonance signal can be subtracted from excessive portion to obtain contribution due to combustion process. Figure 8.11 shows total decomposition of cylinder pressure signal. It is evident from the lot that contribution due to combustion portion dominates in mid frequency range whereas the resonance phenomenon dominates at high frequency ranges [36]. The decomposed signals can be used to calculate noise indices defined in terms of ideal engine speed as:

In

N N ideal

I1

I2

n log dp dt

10 log(106

N N ideal

(8.4)

pilot

dp dt

dp dt

motored

main

Presidual 2 dt Pmotored 2

(8.5)

(8.6)

Combustion Based Noise

Cylinder pressure -bar

60

Total pressure Motored pressure Excessive pressure Resonance pressure Combustion pressure

50 40 30 20 10 0 –10 –360

Pressure level (dB)

191

–240

–120 0 120 Crank angle

240

360

Total pressure Motored pressure Resonance pressure Combustion pressure

1028

1027

101

102

103

Frequency (Hz)

Figure 8.12 Total decomposition of cylinder pressure signal.

where dP is maximum pressure gradient during pilot injection dt pilot period, dP is maximum pressure gradient during main injection, dt main Presidual is residual pressure and dP is maximum pressure gradient in motored pressure dt motored signal. These indices can be further used to find overall noise (ON) emitted from engine as:

ON – C0 + C1I1 + C2I2 + CnIn where constants C0, C1, C2 and Cn depend upon size of engine.

(8.7)

192 Liquid Piston Engines

8.11 Mathematical Model of Generation of Combustion Noise In wavelet transformation complex morlet wavelet can be used as mother wavelet. A morlet wavelet may be defined in terms of central frequency fc and bandwidth fb as:

1

(t )

fb

ei 2

fc t

e

t2 fb

(8.8)

Figure 8.13 shows real and imaginary parts of such a morlet wavelet having fb = 1.5 and fc = 1 Figure 8.14 shows a transient model of combustion noise generation in engines. This model can be analyzed by following three process: generation Complex morlet wavelt cmor 1.5-1 0.5 0 –0.5 –8

–6

–4

–2

0 2 Real part

4

6

8

–6

–4

–2 0 2 Imaginary part

4

6

8

0.5 0 –0.5 –8

Figure 8.13 Complex morlet wavelet. Radiation area Rate of decay C(f ) Combustion impact power Wc Piston

Figure 8.14 Noise generation model.

Engine structure

Combustion chamber Vibration Combustion noise energy E(f) power Wn Transmission Rate of rate (f ) radiation b(f )

Combustion Based Noise

193

of vibrational energy in combustion chamber due to combustion process, decay of energy in the engine structure and finally radiation of energy around engine surface. The combustion process in engine generates combustion impact power (Wc) which can be expressed in terms of in cylinder pressure (p), impedance of medium ( c) and cylinder surface area (A) as:

P2

Wc

A

(8.9)

c

The available energy at engine surface can be expressed in terms of transmission rate coefficient η(f) as:

E( f )

t

( f ) Wc dt 0

(8.10)

Differentiating both sides of this equation we have:

d (E( f )) dt

(f )Wc

(8.11)

Taking into account the decay rate this equation gets modified as:

d (E(f )) dt

(f )Wc C(f )E(f )

(8.12)

where the decay constant C(f) may be defined as:

C( f )

d[log(Wn)] dt

d (Wn ) b( f )h( f )Wc C( f )Wn ( f ) dt

(8.13)

(8.14)

Figure 4.85–4.89 shows the plots of combustion impact power plotted for tests enlisted in Table 7.1 using reference value of W0 as 10–12 Watts.

8.12 Evaluation of Combustion Noise Methods A combustion noise measurement technique has been proposed by Lucas [7]. Austin & Priede have shown that combustion noise was most

194 Liquid Piston Engines Structure response function in decibels re 20 mPa for 1 kPa excitation

80

Structure attenuation in decibels

80

70

90

60

100

50

110

40

120

30

130

20

140

10

150

0 100 200 H.S.D.I.s

500 1k 2k 5k Frequency in herts I.D.L

I.1l/cyl.DI.s

10k 20kHz

I.2l/cyl.DI.s

Figure 8.15 AVL Structural response function and structural attenuation.

dominant in the frequency range 800–4 kHz [38]. Acoustic measurements of noise outside the engine may be used for combustion noise analysis only when the engine is operated in such a way that contribution of combustion events towards the noise emissions is maximum. This can be achieved without changing mechanical noise either by either operating the engine with advanced injection timing or by changing Cetane number of fuel used. Russell has used alkyl blended fuel to maximize in cylinder pressure so that combustion noise becomes dominant [7]. Structural attenuation of engine structure also plays a vital role in determination of combustion noise. Austin and Priede were able to determine this function by subtracting spectrum of in cylinder pressure level from the spectrum of noise emissions recorded at a distance of 0.9 m from engine structure [7]. Value of structure response functions fell by 10 dBA in 500 Hz to 5 kHz range [7]. More recently AVL has developed a combustion noise meter which based on analysis of engine indicator diagram on frequency domain [39]. Good correlation was observed when the data from this noise meter has been compared with that computed results from computer programming. Figure 4.95 shows structural response functions of a group of 9 engines as recorded by AVL noise meter. The response of direct injection high speed diesel engines falls by 12 dB over 5 kHz frequency range [7]. Shu has predicted this transfer function by setting an explosive charge in cylinder at a locked crank angle position as seen from Figure 8.16 [40].

Combustion Based Noise

195

Transfer function

0.5 0.4 0.3 0.2 0.1

0

2k

4k

6k F (Hz)

8k

10k

Figure 8.16 Transfer function obtained by explosive charge.

Different functions for various designs of combustion chambers and different explosion charges were compared. All the above discussed methods use expensive and time consuming methodology to analyze the transfer function of combustion noise, consequently an alternative method of analysis has been studied which evolves use of which include Cepstrum analysis. Cepstrum analysis is an important method of signal processing which has wide applications in source separation [41] ]. This methodology has also been used for psychoacoustic analysis of noise emissions from S.I. engine [42]. Acoustic emissions from engine have been used to reconstruct in cylinder pressure using Cepstrum analysis [43]. This methodology has proved effective an way for fault detection in gears [44] and condition monitoring of engines [45]. Mathematically it can be defined as spectrum of logarithmic power spectrum [41]. i.e.

Ca(q) = |F[log Gx(f)]|

(8.15)

where q is frequency in millisecond & F as well as Gx denotes the Fourier transformation of function. Since auto power spectrum density function is even, both its inverse Fourier transformations & Fourier transformations are equal. i.e.

Cx(q) = |F[log Gx(f)]| =F−1[log Gx(f)]

(8.16)

when a noise source x(t) reaches measuring point as a output signal y(t) after passing through a system h(t) the system can be represented by equation:

y(t ) x(t ) h(t )

x( )h(t

)dt

(8.17)

196 Liquid Piston Engines 101 0

Amplitude

10

2000 RPM 1600 RPM

10–1 10–2 10–3 101

102 Frequency-Hz

103

Figure 8.17 Structural attenuation function (Motored).

Taking Fourier transformation we have:

Gy(f) = Gx(f) Gh(f)

(8.18)

Taking logarithm and Fourier transformations on both sides we have:

log(Gy(f)) = log(Gx(f)) + log(Gh(f))

(8.19)

F[log(Gy(f))] = F[log(Gx(f))]+ F[log(Gh(f))]

(8.20)

F-1[log(Gy(f))] = F–1[log(Gx(f))]+F–1[log(Gh(f))]

(8.21)

Cy(q) = Cx(q) + Ch(q)

(8.22)

Or

Structural response function for motored condition was evaluated taking acoustic emissions as output signal and in cylinder pressure as input. For the case of firing conditions, the rate of heat release was taken as input parameter. Figures 8.17–8.19 shows the plots of transfer function obtained by cepstrum analysis method. It is clear from plots that engine noise transfer function for various testing conditions show same trends and the engine shows a higher value of structural attenuation above 1 kHz range. At low frequency ranges, engine parts have high rigidity and radiation efficiency is very low. In

Combustion Based Noise

197

Amplitude

101 100% load 50% load

100

10–1

10–2 101

102 Frequency

103

Figure 8.18 Structural attenuation function (1600 RPM).

Amplitude

101

100

100% load 50% load

10–1

10–2 101

102 Frequency

103

Figure 8.19 Structural attenuation function (2000 RPM).

mid frequency ranges longitude mode of vibrations in piston, connecting rod and crank shaft dominates which gradually increases the structural response of engine. In high frequency ranges the radiation efficiency increases due to Cast Iron parts in engine head. Several engines use same materials for making of various parts, hence engines of different make but having same size show same variations in structural response function. When compared with structural attenuation function of AVL noise meter, the curve obtained by Cepstrum analysis shows same variations in low frequency range with monotonic rise up to 50 Hz followed by a step fall. Significant differences can be seen in high frequency ranges above 1 kHz. These variations may be attributed to differences in the design of the cylinder head, engine block and covers which also play a vital role in determination of structural attenuation function.

198 Liquid Piston Engines

Combustion noise-dB

103

102

50% load 100% load

101

100 101

102 Frequency

103

Figure 8.20 Combustion Noise – 1600 RPM.

Combustion noise-dB

103

102

50% load 100% load

101

100 1 10

102 Frequency

103

Figure 8.21 Combustion Noise – 2000 RPM.

Further neglected flow induced noise, the overall noise emissions (ON) from engine can be expressed as sum of direct combustion noise (CN) and mechanical noise (speed dependent). i.e.

ON = CN(H1) + MN

(8.23)

where H1 is structural attenuation factor of combustion noise Assuming that mechanical noise levels (motored conditions) do not change significantly, the combustion noise levels for the given testing conditions were evaluated using transfer functions previously described as seen in Figure 8.20, 8.21. From these plots it clear that operational speed of engine has significant effect on combustion noise levels, however increase of engine load caused

Combustion Based Noise

199

SOI map SOI

Engine

Speed load Target MBF50

Target MBF50

Vibration PI controller

Actual MBF50

Estimated MBF50

Figure 8.22 Use of vibration signals as a feedback for estimation of MBF50.

an increase as fuel was injected closer to TDC position. These plots are characterized by peaks in high frequency ranges which may be attributed to resonance of engine structure.

8.13 Summary Value of MBF50 affects the cycle efficiency, emissions peak pressure and temperature achieved in an engine cycle. Hence it can be used as a feedback parameter for a closed loop control system to device optimal timings of combustion process as depicted in Figure 8.22 [46]. In this part of work, vibration and noise signals of diesel engine were analyzed using Cepstrum method. As evident from these plots the trends in transfer function remains same in spite of variations in engine running conditions. This function was further used to calculate combustion noise levels from engine. The value of this function is also dependent upon combustion chamber resonance frequencies, hence temperature variations inside chamber also need to be taken into account for analyzing variations in structural attenuation factor. Further suitable changes in engine structure have shown a reduction of 10 d B up to reduction in engine block vibrations [47].

References and Bibliography 1. Rakopoulos, Giakoumis, “Diesel engine transient operation”, London: springer, 2009. 2. Rakopoulos, Giakoumis, “Experimental study of combustion noise radiation during transient turbocharged diesel engine operation”, Vol 36, pp. 4983–4995, 2011.

200 Liquid Piston Engines 3. Anderton, D., Baker, J., “Influence of operating cycle on noise of diesel engine”, SAE paper no. 730241, 1973. 4. Carlucci, P., Ficarrela, A., Laforgia, D., “Study of the influence of the injection parameters on combustion noise in a common-rail diesel engine using ANOVA and neural networks”, SAE paper no. 2001-01-2011, 2001. 5. Head, Wake, J.D., “Noise of diesel engines under transient conditions”, SAE paper no. 800404, 1980. 6. Torregrosa, A.G., Broatch, A., Novella, R., Monico, L., “Suitability analysis of advanced diesel combustion concepts for emissions and noise control”, Energy, Vol 36, pp. 825–38, 2011. 7. Russell., Haworth, R., “Combustion noise from high speed direct injection diesel engines”, SAE paper-850973, 1985. 8. Baumgarten, C., “Mixture formation in internal combustion engines”, Springer Verlag, Heidelberg, 2006. ISBN 13798-3-540-30835-5 9. Flynn, P., Durrett, R., Hunter, G., Loye, A., Akinyem, O., Dec, J., and Westbrook, C., 1999, “Diesel Combustion: An integrated view combining laser diagnosis chemical kinetics and empirical validation”, SAE paper no 1999-01-0509. 10. Hammer, J., Durnholz, M., Dohle, U., “Entwicklungstrends bei Einspritzsystemen fur PKW-Diesel motoren”, Dieselmotorentechnik, pp. 36–52, 2004. 11. Ricaud, J., Lavoisier, F., “Optimizing the multiple injection settings on an HSDI diesel engine, THIESEL conference on Themro and fluid dynamic processes in diesel engines”, 2002. 12. Drake, M., Ratcliffe, J., Blint, R., Carter, C., Laurendeau, N., “Measurements and modelling of flame front NOx formation and super Equilibrium radical concentrations in laminar high pressure premixed flames”, 23rd symposium on combustion, The combustion institute, pp. 387–395, 1990. 13. Duret, P., Gatellier, B., Miche, M., Montreiro, L., Zima, P., Marotaux, D., Blundell, D., Gase, M., Zhao, H., Perozzi, M., Araneo, L., “Innovative diesel HCCI combustion process for passenger cars: European Space light project, EAEC congress, paper no C108, 2003. 14. Arnone, L., Manelli, S., Chiatti, G., and Chiavola, O., “In-Cylinder Pressure Analysis through Accelerometer Signal Processing for Diesel Engine Combustion Optimization” SAE Technical Paper 2009-01-2079, 2009. doi:10.4271/2009-01-2079 15. Arnone, L., Manelli, S., Chiatti, G., and Chiavola, O., “Engine Block Vibration Measures for Time detection of Diesel Combustion Phases, SAE Technical Paper 2009-24-0035, 2009. doi:10.4271/2009-24-0035. 16. Chiavola, O., Chiatti, G., Arnone, L., Manelli, S., “Combustion characterization in diesel engines via block vibration analysis”, SAE Technical Paper 201001-0168, 2010. doi:10.4271/2010-01-0168. doi:10.4271/2009-24-0035 17. Arnone, L., Boni, M., Manelli, S., Chiavola, O., Conforto, S., Recco, E., “Diesel engine combustion monitoring through engine block vibration signal analysis”, SAE Technical Paper 2009-01-0765, 2010. doi:10.4271/2009-01-0765

Combustion Based Noise

201

18. Chiavola, O., Chiatti, G., Recco, E., “Accelerometer measurements to optimize injection strategy”, SAE Technical Paper 2012-01-1341, 2012. doi:10.4271/2012-01-1341 19. Chiatti, G., Recco, E., Chiavola, O., “Vibration processing to optimize pressure development in CR diesel engine”, SAE Technical Paper 2011-01-1560, 2011. doi:10.4271/2011-01-1560 20. Gang, Sheng., “Vehicle Noise, Vibration, and Sound Quality”, SAE International, 2012. ISBN 978-0-7680-7513-7 21. Anderton, D., “Noise source identification techniques”, ISVR course notes, 2003. 22. Anderton, D., “Relation between combustion system and noise”, SAE paper no. 790270, 1979. 23. Anderton, D., “Basic origins of automotive noise”, ISVR course notes, 1990. 24. Schaberg, P., Priede, T., Dutkiewicz, R., “Effects of a rapid pressure rise on engine vibration and noise”, SAE paper no 900013, 1990. 25. Priede, T., Grover, E., Anderton, D., “Combustion induced noise in diesel engines”, Proceedings of diesel engines users association congress, London, 1968. 26. Russell, M., “Reduction of noise emissions form diesel engine surface”, SAE paper no 720135, 1972. 27. Russell, M., Haworth, R., “Combustion noise from high speed direct injection diesel engines”, SAE paper no 850973, 1985. 28. Russell, M., Cavanagh, E., “Establishing a target for control of diesel combustion noise”, SAE paper no 790271, 1979. 29. Lyn, W., “Calculation of the effect of rate of heat release on the shape of cylinder pressure and cycle efficiency”, proceedings of I mech conference on automobiles, pp. 34–62, 1961. 30. Wolschendorf, J., Durnholz, M., and Schmillen, k., “The IDI diesel engine and its combustion noise variations, SAE paper no 910228, 1991. 31. Torregrosa, A., Broatch, A., Maratin, J., Monelletta, L., “Combustion level assessment in direct injection diesel engines by means of in cylinder pressure components”, Measurements science and technology, Vol 18, pp. 2131–2142, 2007. 32. Priede, T., Grover, E., “Noise from industrial diesel engines”, proceedings of symposium on noise from power plant equipment, Southampton, 1966. 33. Torregrosa, A., Broatch, A., Maratin, J., Monelletta, L., “Combustion level assessment in direct injection diesel engines by means of in cylinder pressure components”, Measurements science and technology, Vol 18, pp. 2131–2142, 2007. 34. Hickling, R., Feldmaier, D., and Sung, S., “knock induced cavity resonances in open chamber diesel engines”, JASA, Vol 65, pp. 1474–1479, 1979. 35. Payri, F., Broatch, A., Tormos, B., and Marant, V., “New methodology for in cylinder pressure analysis on DI diesel engines-application to combustion noise”, Measurement Science and technology, Vol 36, pp. 540–547, 2005. 36. Monelletta, L., “Contribution to the study of combustion noise of automotive diesel engines”, Phd Thesis, University polytechnic velencia, 2010.

202 Liquid Piston Engines 37. Nguyen, T., Kai, Y., Miami, M., “Study on combustion noise from a running diesel engine based on transient combustion noise generation model”, International Journal of automotive engineering, Vol 3, pp. 131–140, 2012. 38. Austen., Priede, “Origins of diesel engine noise proceedings of symposium on engine noise and noise suppression”, IMech, London, pp. 19–32, 1985. 39. AVL450 combustion noise meter, AVL manual, August 2000. 40. Shu, “The transfer function of combustion noise in DI diesel engine”, SAE paper 2005-01-2486, 2005. 41. Liang, X., Yang, K., Shu, G., Dong, L., “The Identification of Noise Source in Diesel Engine Based on the Cepstrum Analysis of Sound and Vibration Signals”, SAE Technical Paper 2012-01-0802, 2012. 42. Andrés, Camacho., Gema, Pinero., María de diego, Miguel, Ferrer., “On the use of complex Cepstrum in psychoacoustic evaluations of engine noise”, 11th ICSV, Saint Petersburg, Russia, 2004. 43. Ghamry, Steel, Reuben, Fog, “In direct measurement of in cylinder pressure suing acoustic emissions”, Mechanical Systems and signal processing, Vol 19, pp. 751–765, 2005. 44. Robert B Randall, “A history of Cepstrum analysis and its applications to mechanical problems”,  7th Surveillance International  Conference  , October  29–30, Institute of Technology of Chartres, France, 2013. 45. Chamay, M., Oh, S., and Kim, Young., “Development of a diagnostic system using LPC/Cepstrum analysis in machine vibration”, Journal of Mechanical Science and Technology Vol 27, pp. 2629–2636, 2013. 46. Scafati, Lavorgna, Mancaruso, E., “Use of Vibration signal for diagnosis and control of a four cylinder diesel engine”, SAE Technical Paper 2011-24-0169, 2011. doi:10.4271/2011-24-0169 47. Kanda, Y. and Mori, T., “Diesel Combustion Noise Reduction by Controlling Piston Vibration,” SAE Technical Paper 2015-01-1667, 2015, doi:10.4271/2015-01-1667

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

9 Effects of Turbo Charging in S.I. Engines

9.1 Abstract We all know that one of the prime objectives of any innovation is to achieve maximum output with minimum input. Automobiles also have innovations that aim at achieving maximum mechanical efficiency & fuel-economy, both simultaneously. The following project-paper focuses on one of the most important topics of present day automotive industry “Turbo charging in S.I engines”. The principle objective of turbo charging is to increase the power output per volume and cost of engine. A fact that a turbocharger increases the mass of air in the cylinder and consequently allows more fuels to be burnt, improves the volumetric efficiency of the engine and simultaneously improves engine efficiency by a small but worthwhile amount. Turbochargers are have commonly been used on diesel engines for many years. In contrast only a few petrol engines have been turbocharged until recently and it is unlikely that a large fraction of the world’s petrol engine will be so equipped. Most of the vehicles in the Indian automobile market continue to utilize turbocharged diesel engines compared to gasoline engines. However, experts are of the opinion that turbo growth, in the future would not be confined to diesel sector

203

204 Liquid Piston Engines alone. There would be a tremendous growth in demand of Gasoline downsizing, in next few years. The whole industry is drifting away from naturally aspirated engines & companies like Mercedes & Volkswagen are planning to come up with commercial turbo-charged, downsized gasoline engines. Thus according to experts if automotive sector, as a whole, is growing at 2%, then turbocharger industry is growing at about 10%, with maximum growth rates coming from turbos incorporated in gasoline engines. The present project aims at analyzing the various benefits associated with “Turbo charging in SI engines” and designing it for future automotive applications.

9.2 Fundamentals A. An I.C engine is a device where-in combustion of fuel takes place within the cylinder & corresponding hot gases are used to drive the output shaft. Every such engine consists of an engine cycle comprising of 4 stages: a. b. c. d.

Suction Compression Expansion Exhaust

There are two types of engines which are classified on the basis of their combustion mechanism as follows: A. SPARK IGNITION ENGINE: refers to the category of engine that makes use of gasoline as fuel. It can be of two types i.e. four stroke or two stroke, depending on whether the one engine cycle is completed in four steps or two steps, respectively .The specialty of this engine is we make use of a carburetor to mix the fuel and air as per the stoichiometric ratio, which is then supplied through pumps into the cylinder (intake) at high pressure. This mixture is then compressed and combusted with in the cylinder using a spark plug which supplies the spark to ignite the mixture (as per specific firing order and after a given period of time).After this the hot gases generated as a result of fuel ignition expand (thereby rotating crankshaft and producing output power) and finally get exhausted from the cylinder. B. COMPRESSION IGNITION ENGINE: This category of engines make use of diesel as a fuel and are employed both in heavy as well as small vehicles i.e. four stroke and two stroke, respectively. The specially of engine

E

Effects of Turbo Charging in S.I. Engines 205

Turbo charger

Diesel engine Air cooler

D

AC generator

A

B C

Figure 9.1 Turbocharged C.I engine.

Table 9.1 Comparison: Spark-ignition and diesel engine.

Compression ratio

Relative F/A

Manifold absolute pressure (Bar)

SI Engine

9.5–11.5

1.0

0.5

12

20–25

CI Engine

16–20

0.2–0.7

1 bar

18

30 (aprox)

Engine type

Max. Bmep (Bar)

Part-load efficiency (%)

is it only makes use of fuel injector and atomizer to atomize and spray the fuel into the cylinder that contains highly compressed air at very high temperatures. Its advantages include high compression ratios resulting in high engine net output lesser knock tendency, cheaper fuel.

9.3 Turbochargers A turbocharger basically consists of a compressor and a turbine coupled on a common shaft. The exhaust gases from the engine are directed by the turbine inlet casing on to the blades of the turbine and subsequently discharged to atmosphere through a turbine outlet casing. The exhaust gases are utilized in the turbine to drive the compressor, which compresses the air and directs it to the engine induction manifold, to supply the engine cylinders with air of higher density than is available to a naturally aspirated engine. The higher value of air-pressure achieved using a turbo-unit is

206 Liquid Piston Engines called Boost-pressure. There exist a number of different types of compressors and turbines, but few of these are ideally suitable to form the basis of an exhaust gas driven supercharging system. The combination of a single stage centrifugal compressor and a single stage axial flow or radial flow turbine is almost universally used in turbochargers. The former type is used for medium and large size engines, while the latter type is used for small engines of automotive type. Advantages: 1. Increase the fuel volumetric efficiency by about 30% to 40%. 2. Increase the number of power stroke i.e. increase the final output. Disadvantages: 1. A disadvantage of turbo charger is its resistance to high temperature at high load, which imposes to increase equivalent ratio (enrichment). 2. The main disadvantage in the turbine inertia and the corresponding long response time needed to obtain the supercharging pressure.

9.4 Turbocharging in Diesel Engines Turbochargers are now widely used for truck engines, their output approaching passenger car engine values of 45 KW at one end of the power range and 600 KW at the other for special purpose vehicles. The use of turbochargers on vehicle engine is relatively recent, but the rapid increase in power output demanded by the use of larger trucks and minimum power to weight ratio legislation, has speeded up their introduction even in passenger cars. The factors that limit turbocharged diesel engine performance are completely different to those that limit turbocharged SI engines. The output of naturally aspirated diesel engines is limited by the maximum tolerable smoke emission levels, which occurs at relative A/F ratio values of about 0.7 to 0.8. It is usually constrained by stress levels in critical mechanical components. This limits the maximum cylinder pressure which can be tolerated under continuous operations, though the thermal loading of critical components can become limiting too.

Effects of Turbo Charging in S.I. Engines 207 Turbo charging in CI engines is free from knocking and self - ignition troubles, enabling it to utilize high compression ratios, multi-port direct fuel injections and high cetane number and cheaper exhaust control systems which make it easier and more economical to design.

9.5

Turbocharging of Gasoline Engines

In SI engines, fuel and air are pre-mixed before the air enters the cylinder of petrol engine. Whether a carburetor or manifold petrol injection system is used, cylinder comprises of homogeneous air and fuel mixture, the proportion of fuel being carefully controlled. The homogeneous mixture is ignited by the spark plug. Unlike diesel engine the rate at which combustion proceeds is governed by heat and mass transfer from an area that is burning to an area that is not, and temperature increase due to continued compression thus flame advances across the combustion chamber, from the spark plug until all the fuel is burned. Self ignition is avoided by low compression ratios, enough to hold the temperature of mixture below the self ignition point of the fuel, and by using a fuel having high self ignition temperature. The rate at which the flame progresses is governed by local turbulence, heat transfer between burning and unburned region, compression heating of the unburned gas due to piston motion and expansion of burning mixture, air/fuel ratio and heat transfer to the surrounding walls. Since the unburnt gas that is removed from the advancing flame front is heated by compression

Air filter Air flow meter Intercooler Compressor Turbine shaft

Throttle

Turbine

Intake manifold

Waste gate

Engine Exhaust manifold

Figure 9.2 Sketch of a turbocharged SI-engine.

Catalyst

208 Liquid Piston Engines and to some extent by radiation etc. This gas can reach its self ignition temperature before the flame front arrives, thus increasing the chances of knocking in the end gas region. This extremely rapid combustion generates a high rate of pressure rise in the cylinder, the impulse of force causing the bearing to knock, generally referred to as detonation.

9.6 Turbocharging Turbocharging applied further to S.I engines. As we know, turbo charging is one of the most economic and effective methods used to improve both the volumetric efficiency and thereby improve the overall efficiency of any engine. However, this principle is not very widely used in case of SI engines primarily due to the difference in combustion systems of the two engines (i.e. SI and CI). This difference has been explained in the previous section. The important point is that any measures that might increase the temperature of the mixture towards the end of the compression strokes are undesirable. Unfortunately, turbo charging does just this. By raising the inlet manifold pressure and temperature, the pressure and temperature of the mixture in the cylinder will be raised throughout the compression stroke. Hence, it is possible only to mildly supercharge the engine without inducing knock or self-ignition. Thus a certain margin clear of knock must be maintained by incorporating the following: 1. Low compression ratio 2. Retarded ignition timing or charged air cooling 3. Usage of high octane fuel, having high self- ignition temperature etc. to offset the effect of temperature rise in the compressor.

9.7 Components of Turbocharged SI Engines A. Inlet and Exhaust Manifolds The design of the inlet system may be substantially different from that used in naturally aspirated engine; particularly if carburetor is placed before the compressor. The primary objectives will be to minimize the volume between carburetor and cylinder head (to improve response), encourage good fuel droplet break-up and mixing, avoid fuel condensation,

2600

2400

2200

200

1.

2.

3.

4.

S.No

Engine R.P.M

731

700

658

627

N.A

849

843

785

765

T.CH

Load (N-M)

17.3

15.8

15.3

13.6

N.A

13.1

13.8

13.5

12.9

T.CH

Consumption Of 200 G of fuel (Sec)

208

219

224

232

N.A

236

260

264

279

T.CH

Power in KW

13%

18%

17%

20%

Power

% Increase in

16%

20%

19%

22%

% Increase in load

32%

14%

13%

5%

% Reduction in fuel consumption

Table 9.2 Comparison of naturaly aspirated & turbo-charged BMP C.I engine (15.9 Litres at ambient temperature 28 °C & humidity 61%) Ref:V.R.D.E. MAGAZINE.

Effects of Turbo Charging in S.I. Engines 209

210 Liquid Piston Engines encourage even mixture distribution between cylinders, eliminate swirling air movement and keep the temperature of the mixture down. These objectives are some times achieved by accepting a pressure loss that would not be tolerated on a naturally aspirated engine. The reason is due to the most embarrassingly large amount of energy available in the exhaust system at full speed, enabling boost pressures far above the knock-limited value to be generated. Some flow loss, at full throttle and speed, reduces boost pressure, loss being low when boost pressure and hence mass flow rate are also small. A perforated plate can be used at the entry to the inlet manifold to reduce the swirl and aid mixture distribution, but this method is not to be encouraged since restrictor reduces pressure, but not temperature. Although the potential boost available from the turbocharger is excessive at full engine speed and throttle, the opposite is so at low speeds. Thus the exhaust manifold is to be designed to ensure the maximum utilization of exhaust gas energy at relatively low engines speeds. Thus the pulse turbo charging system should be adopted, with short, narrow exhaust pipes and if possible no more than 3 cylinders connected to each turbine entry. Long pipes have been used to allow cooling of the exhaust gas before it reaches the turbine, but pressure wave action can cause problems. The energy available for expansion through the turbine is reduced at low speeds, engine compartment temperature increases and turbocharger response suffers. Usually a large-bore exhaust pipe will be used from the turbine exhaust onwards, together with low-loss silencer. A small bore system will create flow restriction and back pressure at the turbine exit. Although this need not be a problem because turbocharger can be matched accordingly, it can be dangerous if the exhaust system develops a major leak. The back pressure will be reduced and turbocharger work will increase, possibly resulting excessive boost pressure, combustion knock and engine damage.

B. Turbocharger Boost Pressure Control System The need for boost pressure control system in a turbocharger unit includes: 1. Petrol engine works over a wide speed range, typically 5000 revolutions per minute. 2. Air/fuel mixture must be keep relatively closed to the stoichiometric value rather than varying over a very wide range. 3. Boost pressure must be limited to avoid knock. Problem involved in the matching the turbocharger over the normal

Effects of Turbo Charging in S.I. Engines 211 speed range becomes more severe over the wider speed range of petrol engine. The various ways by which these objectives achieved are:

C. Matching the Turbo Charger for the Desired Maximum Boost Pressure One of the simplest methods to achieve specified boost pressure at maximum speed and load is to match the turbo charger accordingly. The boost pressure can be limited by fitting a large turbine; however the turbine area will be governed by mass flow through it at full power. Since the petrol engine operates over a very wide speed range, mass flow will be very much less at low speeds. Under these conditions, the turbine area will be relatively large and hence the energy available for expansion will be correspondingly small resulting in little or no boost pressure being developed. Aggravating the situation will be low turbo charger efficiency due to operation far from the design point conditions of the compressor. The overall pressure will be a boost pressure and engine torque curve rising rapidly at high speed, with low boost at low speed. Such a curve, with no torque back up is totally unacceptable for the automotive application since continual gear changing is required.

D. Exhaust Waste-Gate systems It is one of the most extensively used pressure control system. It consists of a valve allowing exhaust gas to by pass the turbine. It is an excellent method to control boost pressure, since the exhaust gas energy of the turbo charged engine is excessive at full speed and load. It consists of a diaphragm

Outlet

Inlet

Wastegate

Throttle

Engine

Figure 9.3 Waste-Gate Control System.

212 Liquid Piston Engines on which the boost pressure acts, opening the waste gate when the boost pressure reaches the pre –determined value. The various advantages of this system are: 1. Since not all the exhaust gas passes through the turbine, and no more than air requirement of the engine passes through the compressor, then a smaller turbo charger can be used. This small turbo charger is best able to provide sufficient boost at low speed (when waste gate is closed), and reduces turbo charger lag due to its low inertia, especially if the waste-gate is closed during acceleration. 2. With careful design, the boost curve can be tailored to produce an optimum torque cure within the constraints of knock limit.

E. Selection of Fuel - Supply Systems There are two types of combustion systems used in SI engines, classified on the basis of fuel- preparation systems 1. Spark plug based system: they make use of carburetors to prepare the fuel/air mixture in the correct stoichiometric ratio. There can be two types of such systems based on whether turbocharger is placed before or after the carburetor, which is called sucking-in or blow-through type systems respectively. The various disadvantages associated with the two systems are: a. When turbocharger is placed after the carburetor it makes it more difficult to provide the require the boost to the air/fuel mixture from the carburetor, as the fuel particles are more difficult to compress due to their higher density. Apart from this there might also be flow-separation problems inside the compressor. Air cleaner Compressor Carburetor Turbine Exhaust pipe Exhaust gas

Figure 9.4 Carburetor -based turbocharger.

Effects of Turbo Charging in S.I. Engines 213 b. When turbocharger is placed before the carburetor, it has to handle only air and hence more preferred. However this also results in higher boost pressure losses inside the carburetor. It might also result in excessive rise in fuel / air temperature and pressure inside the carburetor, which both affects the carburetor and also increases the end gas temperature in the combustion camber causing knocking and detonation. 2. Direct-injection systems: they make use of injection nozzles to spray the fuel/air mixture directly into the cylinder at high pressure, just during before the compression stroke begins, such that a homogenous charge mixture is formed which gets ignited using the spark-plug & undergoes uniform combustion. They are more commonly used these days, since they can be more easily controlled for different operating conditions (like part-load, full-load etc.) & can be used in conjunction with turbo-chargers to achieve maximum efficiency with minimum lag at low-speed conditions.

9.8

Intercooler

When intake air is compressed by a turbocharger it is also heated, even more so than when supercharging due to the turbo being heated by the exhaust. Hot intake air is not good for power and will increase the chance of detonation. An intercooler reduces the intake temperature by pushing the air through a heat exchanger (much like a small radiator) that absorbs some of the heat out of the charge. With less heat, less boost pressure is needed to get the desired power and decrease the chance of detonation. Anything that reduces the intake temperature is a big plus in a supercharged engine.

9.9 Designing of Turbocharger Designing a naturally aspirated SI engine of following specifications is compared with a turbocharged SI engine of same specifications.

Specifications of a Naturally aspirated SI Engine: D = 90 mm L = 125 mm

214 Liquid Piston Engines N = 2500 RPM Rp = 7:1 K = Number of cylinders = 3 = Density of gasoline = 800 kg/m3 c.v. = 44000KJ/Kg Maximum speed = 144km/hr Fuel consumed = 14km/litre Ambient pressure = 1 bar= P1 Ambient temperature = 27 °C = 300 K = T1 Calculations: Rate of fuel consumption = mf= Maximum speed/fuel consumed = 10.2 litre/hr Vs = π/4 × D2 × L × N/2 × K = 2.985 m3/minute = 8.22 kg/hr Specific consumption of fuel = mf ˙ = mf Assuming complete combustion of gasoline, the Combustion of gasoline is analyzed to find the mass flow rate of air needed to burn fuel and the mass flow rate of the exhaust formed. Thus by knowing the swept volume the volumetric efficiency is calculated. 2C8H18+25O2 (GASOLINE)

16CO2+18H2O (AIR)

Oxygen needed for combustion of 1 kg of fuel = 0.85(32/12) + 0.15(8) = 3.46 kg/kg of fuel Air needed for combustion of 1 kg of fuel = 3.46/0.233 = 14.87 Hence air flow rate = m˙ a=14.87(8.22) = 122.23 kg/hr = 0.033 kg/s Now P1V1 = m˙ a RT1 V˙air = V1˙=V2˙=. 028m3/s = 1.68 m3/minute v

(Vair /Vs ) 100 = 56%

Hence Oxygen flow rate = m˙O

2

= (0.23)122.23 kg/hr = 28 kg/hr = 0.46 kg/minute Carbon dioxide formation rate= m˙CO =8.22(704/228) =25.38kg/hr 2

Effects of Turbo Charging in S.I. Engines 215 Maximum volume Displacement volume Clearance volume

Absolute pressure

3

2 4

1

0 0

Atmospheric pressure

0

Volume

Figure 9.5 P-V Plot of a Naturally aspirated SI Engine.

Water formation rate= m˙H O=8.22(324/228) =11.68kg/hr 2 Hence m˙exhaust = m˙H O + m˙CO = 0.01kg/s 2 2 Now By finding the pressure and temperatures at states 1,2,3,4 the stroke volume and work done per cycle (area under curve) is calculated which leads to calculation of mean effective pressure and hence theoretical power generated. Hence the brake specific fuel consumption and brake thermal efficiency is calculated P2/P1 = 7 = Rp Hence P2 = 7 Bar T2/T1 = 7(

1/ )

Hence T2 = 300(7)0.4/1.4 = 523K Assuming T4 = 1000 K = Exhaust gas temperature of a S.I. Engine T3/T4 = Rp(

1/ )

T3 = 1743 K = maximum cycle temperature For process 2–3 we have P3/P2

T2 = T3

216 Liquid Piston Engines P3/P2 = 1743/523 = 3.33 P3 = 7(3.33) = 23.33 Bar P3/P4 = 7 P4 = 3.33 Bar = Exhaust pressure of gases Now V˙1 = V˙4 = 0.029 m3/s P3/P4 = (V˙1/V˙3) = (V˙1/V˙2) V˙2 = V˙3 = 7.27

Work done per cycle

(P3V3 P4V4 ) (P2V2 PV 1 1) 1 1 0.160 J

w

Length of indicator curve = V˙1

IHP

10 3 m3/s

V˙2 = 0.02173 m3/s

(IMEP )(L)( A)(N /2)(K ) 60, 000 49.1 HP

36.36 KW

= mechanical efficiency = 90% = BHP/IHP m BHP = 32.96 kW = 44.1 HP BHP = 2 pNT/60,000 T = TORQUE = 125.89 N-m

BSFC

mf /BHP bth

0.2492 Kg/KW-hr

BHP/(mf cv) 32%

Case-2 Turbocharged SI engine without intercooler

Effects of Turbo Charging in S.I. Engines 217 Compressed air flow Engine cylinder

Turbocharger oil inlet Turbine wheel

Compressor

Exhaust gas discharge

Ambient air inlet

Compressor wheel

Wastegate

Oil outlet

Figure 9.6 Turbocharged SI engine. Turbine wheel Inducer diameter

Exducer diameter

Shroud profile

Backface Tip width

Figure 9.7 Turbine wheel. Cb

B

A

F 1

1 = 16°

Inlet velocity triangle

Cr1

Cf1

Ca1 E

Turbine vane Cf2 = Cr2

Cr2 2

2

B

= 90° E

Cb

Outlet velocity triangle

Figure 9.8 Velocity diagrams of compressor vane

TURBINE The turbine wheel is made from a high nickel superalloy investment casting. This method produces accurate turbine blade sections and forms. Larger units are cast individually. For smaller sizes the foundry will cast multiple wheels using a tree configuration.

218 Liquid Piston Engines Specification of turbine wheel: Diameter = D = 270 mm = 0.27 m, N = 50,000 RPM Nozzle angle = 1 = 16 , Type-Radial flow turbine Calculations for turbine wheel: Now the power generated by turbine is calculated by analyzing the velocity triangles for the turbine vane. Cb = πDN/60 = 706 m/s = AB Assuming Turbine efficiency 85%, ρ = 0.7 = Cb/ Ca1 Analyzing the inlet velocity triangle we have Ca1 = 1009 m/s = absolute velocity at inlet Cf1 = Ca1 Tan 16° = 289 m/s = flow velocity at inlet AF = Ca1 Cos 16 = 969 m/s = Cw Let transmission efficiency = 90% Now by calculating the power required to drive the compressor and analyzing the compressor vane, the mass flow rate of air to compressor inlet is found Power generated by turbine = m˙exhaust (CW) (Cb) = 6847 W = 9.17 HP Power supplied to compressor = Pc =0.9(6847) = 6162 W = 8.25 HP

Compressor Compressor impellers are produced using a variant of the aluminum investment casting process. A rubber former is made to replicate the impeller around which a casting mould is created. The rubber former can then be extracted from the mould into which the metal is poured. Accurate blade sections and profiles are important in achieving compressor performance. Back face profile machining optimizes impeller stress conditions. Boring to tight tolerance and burnishing assist balancing and fatigue resistance. The impeller is located on the shaft assembly using a threaded nut.

Inducer dia

Superback profile

Tip width Blade width

Figure 9.9 Comprssor wheel.

Exducer dia

Main blade

Shround profile Spliter blade

Effects of Turbo Charging in S.I. Engines 219 Compressor housings are also made in cast aluminum (cast iron for high-pressure applications). Various grades are used to suit the application. Both gravity die and sand casting techniques are used. Profile machining to match the developed compressor blade shape is important to achieve performance consistency. Specifications of compressor Type-Centrifugal Outer Diameter = D2 = 150mm Inner Diameter = D1 = 100mm N = RPM = 50,000, P = 1.4 σ = 0.94 = Cw2/Cb2 Calculations for compressor wheel Cb2 = pD2N/60 = 392 m/s Cw2 = 369 m/s Pc = m˙ compressed air × Cb2 ×Cw2 m˙ compressed air = 0.0425kg/s = 2.55kg/minute V˙ compressed air = 2.20 m3/minute = 0.036 m3/s Assuming lean mixture formed, A: F = 18:1 m˙fuel = 0.141 kg/minute = 8.5 kg/hr New volumetric efficiency (Vair /Vs ) 100 73% Now assuming compressor isoentropic efficiency as 85%, the rise in temperature and pressure of air by compression is calculated T01 = 27 °C = 300K T02 = Actual Outlet temperature of air from compressor T02˙ = Isoentropic Outlet temperature of air from compressor hc = 85% = (T02 T01)/ (T02 T01) = isoentropic compressor efficiency Now Cp (T02 T01) = P (s) (Cb2)2/gc J Hence rise in air temperature in compressor = (T02 T01) = 130K Outlet temperature of air = T02 = 430K (T02 T01) ( / 1) = P2/P1 = Pressure rise in compressor = 2.99 Now By finding the pressure and temperatures at states 1,2,3,4,8,7,6 the stroke volume and work done per cycle (area under curve) is calculated which leads to calculation of new mean effective pressure and hence theoretical power generated. Hence the new brake specific fuel consumption and the new brake thermal efficiency is calculated.

220 Liquid Piston Engines Cb2

Outlet velocity triangle Cw2

G

D

E =2

Cf2

=3 C2

Cr2 F

Compressor vane C

C1 = Cf1 Cr1 =1

Inlet velocity triangle

A

Cb1

Figure 9.10 Velocity diagram of compressor blade.

p

pch = Charging pressure pa = Atmospheric pressure

3

2 4 pch

8

pa

7

Ap TDC

6 BDC

V

Figure 9.11 P-V curve of a turbocharged S.I.Engine.

V˙ compressed air = V˙1 = V˙4 = 0.036 m3/s A more dense air enters the engine cylinder and hence the compression ratio is assumed to reduce from 7 to 5.7 Now P3/P4 = (V˙4/V3˙) = (V˙1/V˙2) = 5.7 , V˙2 = V3˙ = 0.0103 m3/s T2/T1 = 5.7 ( 1/ ) Hence T2 = 430(5.7) 0.4/1.4 = 707K

Effects of Turbo Charging in S.I. Engines 221 P7 = P6 = 1 Bar=Ambient pressure P1 = P8 = 2.99 Bar Hence P2 = 5.7(2.99) = 17.043 Bar New length of indicator curve = 0.036-0.0103 = 0.0257 m3/s T4 = 850 = Exhaust gas temperature of turbocharged S.I. engine T3/T4 = Rp ( 1/ ) T3 = 1397 K = New maximum cycle temperature For process 2–3 we have P3/P2 T2 = T3 P3/P2 = 1397/707 = 1.97 P3 = 1.97(17.043) = 33.57Bar P3/P4 = 5.7 P4 = 5.89 Bar = Exhaust pressure of gases New work done per cycle W=

P3V3 P4V4 1 0.2418 J

P2V2 PV 1 1 1

P (Stroke volume)

work done per cycle Length of indicator curve 0.2418 / 0.0257 9.4 bar (IMEP )(L)( A)(N /2)(K ) IHP 46.76 KW = 62.68 HP 60.000 = mechanical efficiency = 90% = BHP/IHP m BHP = IHP (0.9) =42 KW=56.41 HP BHP = 2 pNT/60,000 T = 160 N-m IMEP

New BSFC = m˙f/BHP = 0.2428 tvkg/kw-hr NEW bth Brake thermal efficiency BHP/(mf

cv) 40%

Summary of calculations Naturally aspirated SI engine

Turbocharged SI engine

BSFC

0.2492kg/kw-hr

0.2428kg/kw-hr

BHP

32.96kw

42kw

Parameter

222 Liquid Piston Engines TORQUE

125.89 N-m

160N-m

IMEP

7.3bar

9.4bar

32%

40%

bth

9.10

Operational Problems in Turbocharging of SI Engines

The basic efficiency of the thermodynamic cycle on which the engine operates is largely governed by the compression ratio. Thus by reducing the compression ratio to avoid knock, the efficiency of the basic thermodynamic process is reduced. As a result it is probable, but not certain, that the overall efficiency of the engine will suffer. However, the fall in efficiency with low compression ratio is by no means linear. Nearly, 25% reduction in compression ratio reduces the efficiency by only 10%. In addition to the losses occurring in the basic thermodynamic processes, friction in all the bearings and other mechanical debits will reduce the power output to fly wheels. Since the reduced compression ratio partly offsets the increase in cylinder pressure due to turbo charging, the mechanical loads may not change significantly with turbo charging. Thus absolute power loss due to friction will remain steady. E.g.: Consider an SI engine with compression ratio 10:1 producing the power output of 100KW with 46% cycle efficiency and 70% mechanical efficiency. The potential power of 312 KW of which 143 KW arrive at the pistons and 43 KW is then lost due to mechanical inefficiency. If the engine is turbocharged such that 1.5 times as much mixture is trapped in the cylinder, and compression ratio is reduced to 7.5. The potential power becomes 468 Hp and the power delivery from the piston is 196 KW. If mechanical losses remain unaltered at 43KW, the power output at the flywheel is 153 KW and the overall efficiency is 32.7% of the normal engine. (Ref: turbo charging of I.C engines, Watson & Janota) “Thus power output has increased substantially with a small gain in efficiency”.

Effects of Turbo Charging in S.I. Engines 223 Most important is the fact that whatever modifications are made to the engine, the charge temperature should be kept as low as possible, so that the compression ratio can be maintained as high as possible commensurate with freedom from knock. There is also a strict trade-off between boost pressure, charge air temperature ,air/fuel ratio and fuel octane number at the knock limit, with a fixed compression ratio and an optimum ignition timing. As expected octane number has a strong influence on the permissible boost pressure and so does charged air cooling. In addition, richer air/ fuel ratios permit higher boost pressure.

9.11 Methods to Reduce Knock in S.I Engines The various methods used to reduce the likelihood of knock in S.I. Engines are: 1. Low compression ratios : to hold the compression temperature rise to an acceptable limit 2. High self ignition temperature: dependent on many factors like fuel properties, air/ fuel ratio and pressure. 3. High octane rated fuels: have higher self ignition temperature and reduce knocking. 4. Centralized spark plug: to hold down its maximum distance from the extremity of the combustion chamber. 5. High wall surface area to gas volume ratio: to keep the end gases relatively cooler to avoid end gas detonation

9.12 Ignition Timing and Knock Retarding combustion reduces the temperature of the end gas by delaying its heating related to the TDC piston position. Thus the cylinder volume is increasing at the critical time, reducing the compression and temperature rise that would occur, otherwise. Retarded ignition reduces engine efficiency by shortening the effective expansion stroke. However, relative to reducing compression ratio, it is more flexible technique, since it is relatively easy to retard ignition only and the boost pressure is high enough to induce knock.

224 Liquid Piston Engines 1 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 –40

–35

–30

–25

a)

–20 –15 [deg]

–10

–5

0

To avoid an unnecessary fuel consumption penalty with retarded timing, the ignition timing technique should only be used when turbocharger does develop a high boost pressure. Thus at low speeds and part load convectional timing is retained. The simplest method of achieving this requirement is a boost pressure retarded system built into the normal vacuum advance diaphragm of the ignition distributor. One undesirable feature of retarded timing increase in heat rejection to the exhaust system, since the complete combustion and expansion process is delayed. Thus the turbine inlet temperature rises. Although the increase is small, the very high temperature of the petrol engine exhaust gas (upto 1000 °C) is a problem for the turbine manufacturer and can cause oxidation of lubricating oil. Furthermore the potential power increase obtainable by turbo charging with retarded timing alone is limited. Higher boost pressures can be used if compression ratio is also reduced.

9.13 Charge Air Cooling The temperature rise in the compressor and its effect on the knock can be offset by charge air cooler. Charge air cooling to an air temperature of 45 °C enables the knock-limited ignition timing to be advance by 10°. However the low boost pressure of the turbocharged SI engine means that the temperature difference available between ambient air and compressed air is small. A large air cooler is therefore required to achieve a major reduction in temperature and low pressure loss. However, it is doubtful whether the extra cost, complexity and volume are warranted in passenger car applications, other than very expensive sports cars. An additional disadvantage is deterioration in engine response due to the increase in total inlet manifold volume and pressure loss in cooler itself.

Effects of Turbo Charging in S.I. Engines 225

9.14

Downsizing of SI Engines

Downsizing refers to the reduction of engine swept volume without compromising on the potential engine output. It offers enhanced engine efficiency, fuel economy and potential to meet future emission standards. However in order to realize these benefits a number of issues faced by conventional port injected down-sized engines must be resolved. 1. Compression ratio: as previously mentioned, it must be reduced on conventional boosted engines to control knock. 2. Low speed torque: due to air flow and octane requirement constrains, steady state torque is reduced compared to a naturally aspirated engine of equivalent peak torque and power. This adversely affects pull away from rest and performance feel for tip-in 3. Manoeuvres at low engine speed 4. Transient response: although greatly improved with modern boosting systems, turbocharger lag is still an issue for down sized engine. Turbocharger lag compounded by the lack of steady-state torque is a particular issue for low speed performance feel. 5. Economics: any downsizing technology package must be economically justified in terms of a cost benefit analysis compared to other technology packages.

9.15

Techniques Associated with Turbo Charging of SI Engines Boosting Systems

A number of advanced boosting technologies aimed at addressing the steady state low- speed torque and transient response issues are currently in advanced engineering stages as follows: 1. 2. 3. 4.

Gasoline direct injection Variable compression ratio Variable geomety turbine Exhaust gas recirculation

a. Gasoline direct injection (GDI): Direct injection is a key technology for improving fuel economy of down sized engines. As well as the small gain in volumetric efficiency, direct injection allows an increase in

226 Liquid Piston Engines compression ratio of 1 to 1.5 compared with an equivalent port injected engine as a result of charge cooling. Swirl control valve Spark plug

High-pressure fuel injector Hollow cone spray

Swirl air motion

Piston bowl

Direct injection also allows more freedom on choice of valve overlap for boosted engines, as starting to inject after exhaust valve closure can prevent fuel loss. This allows improved scavenging under some conditions, which reduces charge temperature and octane requirement. The ability to have multiple injection events in one combustion cycle allows both increased exhaust temperature from cold start for improve catalyst light-off performance and reduced full-load octane requirement. The increase in exhaust temperature can help to over come the increase exhaust system thermal inertia with turbocharger. Thus the injection strategy must be optimized for each combustion system. b. Variable compression ratio (VCR): In principle, variable compressible ratio is an attractive approach for down sized boosted engines. Low compression ratio would be used at full-load to control knock, allowing very high BMEP to be achieved. As load is reduced; compression ratio will be optimized for best economy. Two variable compression approaches have been proposed for downsized boosted engines: a. In the other, a fixed high geometric compression ratio is used but the effective compression ratio is varied by late closing of inlet valve. This approach is termed as “Miller’s cycle”. b. In one approach, the geometric compression ratio is varied by mechanical means. In practice, variable geometry compression ratio systems have a number of drawbacks. As well as many design and development issues such

Effects of Turbo Charging in S.I. Engines 227 as packaging, friction, transient response, durability and cost to be resolved. The use of low compression ratio at high load will result in poor high load fuel consumption. The resulting high full-load exhaust temperatures will require earlier and increased use of over-fuelling for component protection. The principle is to use high geometric compression ratio, (around 14:1), giving good part load economy, and to control knock by reducing by effective in-cylinder compression ratio by late intake valve closing. As this reduces volumetric efficiency, higher boost pressure are required to achieve same BMEP. However as more of the compression is done in the external compressor and the intercooler can be used, the in-cylinder charged temperature at ignition will be lower than for a conventional boosted engine, even with a same trapped mass. This will benefit the octane requirement while VCT would optimize cam-phasing at each speed. E.g. Miller cycle approach was used in production by Mazda in 1980’s. This engine featured fixed late inlet valve closure and supercharged using lysholm compressor and inter cooled. It has the compression ratio of about 10:1. More recently, this approach has been proposed in combination with VCT (variable cam-shaft timing) as a solution of downsized engines. c. Variable turbine geometry (VGT): By altering the angle of the turbine inlet nozzle both the effective area of the turbine (and hence energy availability) and efficiency characteristic alter. By opening the nozzle at full engine speed and closing them at low speed, exhaust gas energy can

Exhaust gas

Speed sensor

Boost-air outlet

Sliding nozzle ring Turbine

Air inlet Compressor

Pneumatic control cylinder Control air

Figure 9.13 Turbocharger with VGT.

228 Liquid Piston Engines 3

1 2

3

1 5

2

4 7

6

8

(a)

(b)

Figure 9.14 Exhaust gases recirculation system assembly for naturally aspirated (a) and turbocharged (b) - System components: 1 humidity’ separator; 2 booster; 3 EGR valve; 4 single point injection; 5 heat exchanger: 6 compressor: 7 single point injection and equalization box: S turbine.

adjusted to suit low and medium-speed performance while preventing the turbo charger from over-speeding. The difficulty is one of engineering a cheap & reliable variable geometry turbine together with associated control system. As compression ratio increases, modern gasoline engines have exhaust temperature higher and higher. Experts estimated it could exceed 1000°C in the foreseeing future. Perhaps this is why VTG technology for gasoline engines never went into mass production. In terms of single-stage charging units, turbochargers with variable turbine geometry are primarily used for passenger car diesel engines (in addition to modern and economical turbochargers with boost pressure control valves). d. Exhaust gas recirculation(EGR):It is a technique used to cool the flame passing through combustion chamber offsetting knock limit, enabling optimum ignition event and more power at crank shaft also enables addressing full load fuel consumption. Port injection turbo used enrichment to limit peak combustion chamber temperature. Unburned fuel cools the flame and offsets the point where the enrichment starts. EGR also enables reduction in NOx emissions, by reducing the temperature within the combustion chamber and thereby avoiding any chemical reaction of nitrogen present in the air. To do this, a small portion of the exhaust is diverted back into the intake manifold using special flaps in the

Effects of Turbo Charging in S.I. Engines 229 exhaust manifold tubes, which has a cooling effect over the homogenous charge mixture with in the cylinder. Generally, EGR is induced by making use of long and narrow exhaust pipes which reduce the temperature of the exhaust and in association with VGT to get high exhaust velocity to run the turbine and thereby high boost pressures and minimum turbo lag.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

10 Emissions Control by Turbo Charged SI Engines

Emission is one of the important binding factors involved in the design of any engine. The National emissions ceiling directives set binding emissions ceilings to SO2, NOx volatile organic compounds, ammonia etc. In the recent years, NOx emissions have reduce dramatically, some of the credit goes to 3-way catalyst for gasoline engine vehicles and NOx storage catalyst for GDI cars. But increases in the number of diesel cars, makes it difficult to reduce it further. Diesel engine actually accounts for an increase and there has been little development in the truck sector. 2020 objective includes reduction in the emission with 82% SO2 Reduction, 60% NOx, 51% VOC(Volatile organic compounds), 27% ammonia, 59% particulates with respect to 2000 standards. NOx is either NO or NO2 depending upon the temperature within the cylinder. NOx production is mainly due to NO reduction area of high ozone concentration produce more NO2 by reaction with NO, which is harmful to humans. All diesel engines have oxidation catalysts which convert Hydro-Carbons and CO into CO2 and H2O and simultaneously oxidize NO to NO2. Diesel particle filters use some of the NO2 for regeneration, 231

232 Liquid Piston Engines but only occasionally. Some have constant regeneration but won’t use all NO2. Modern cars use DENOx catalyst to adsorb NO2 and NO. Gases are treated and become N2 ad H2O by using H2 in exhaust gas stream. Some NO2 will escape and some released during regeneration. NOx treatment in diesels will be the main focus of EURO 6, although specific targets are not set.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

11 Scope of Turbo Charging in SI Engines

Today, turbo charging is most commonly used on two types of engines: Gasoline engines in high-performance automobiles and diesel engines in transportation and other industrial equipment. Small cars in particular benefit from this technology, as there is often little room to fit a largeroutput (and physically larger) engine. The Porsche 944 utilized a turbo unit in the 944 Turbo (Porsche internal model number 951), to great advantage, bringing its 0–100 km/h (0-60 mph) times very close to its contemporary non-turbo “big brother”, the Porsche 928. In the 1980s, turbocharged cars were difficult to handle. The tuned engines fitted to the cars, and the often primitive turbocharger technology meant that power delivery was unpredictable and the engine often suddenly delivered a huge boost in power at certain speeds. As turbocharger technology improved, it became possible to produce turbocharged engines with a smoother, more predictable but just as effective power delivery. In future, to meet U.S. emission regulations, injector systems will have to be optimized with increased functionalities, multiple injection strategies and increase pressures (up to 2000 bars or more). However improved 233

234 Liquid Piston Engines combustion systems will have to be supplemented with after – treatment techniques like selective catalytic reduction (SCR) for NOx reduction and state of art particulate filters. Turbo charging would be adopted much faster in downsized engines for entry level vehicles, on account of their low cost. Turbo charging in GDI offer most potential, but at a cost. Forced induction of traditional port injected (MPFI) engines is an attractive and economical compromise. While turbo charged GDI can use 15% to 20% less fuel than naturally aspirated variants, turbo MPFI uses 5% to 15% less. Now Volkswagen is taking turbo GDI into mass market with 1.4 liters Turbo charged SI (TSI) engines in the Golf and soon to be launched Polo. But Renault has opted for turbo MPFI for 1.2 liter TCE 100 engine used in Clio and Twingo. Fiat is also taking the same approach with 1.4 liters T-JET engines for Bravo. Twin-staged turbos with two turbo chargers and extra valves is an attractive in long engine with a cross flow design like in V-8 engines and even in sportier and middle sized cars. Turbo chargers with twin scrolls, not only improves scavenging but also turbo efficiency at low speed and inertia i.e. improve transient response. Twin scroll turbine housing, ball bearing and VGT will improves the boost pressures and quicker change from one driving mode to another.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

12 Summary

1. The biggest opportunity for improving the spark ignition engine is boosting and downsizing. 2. Stoichiometric operation enables very low air pollutant emissions. 3. Many other design variables could contribute: E.g. increase compression ratio, variable valve control, and lower friction etc. 4. The major challenge is controlling knock. 5. 20–30% higher part-load efficiency plausible.

235

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

13 Conclusions and Future Work

13.1 Conclusions The present work establishes benefits of using various intrusive as well as non-intrusive methods to analyze noise and vibrations from a dual cylinder diesel engine. Various indices described can serve as good indictors for condition monitoring of engines. However some of these may be sensitive towards interference of background noise or chamber resonance. This drawback may be surpassed using suitable filters. Various Frequency bands in which combustion process or piston slap dominates have been investigated.

13.2 Contributions 1. Investigation of effects of location of various transducers towards signals acquired. 2. Use of Cepstrum analysis to study combustion noise. 3. The application of COMSOL -7 software in conjunction with FEA for analysis of secondary motion of skirt. 237

238 Liquid Piston Engines

13.3

Future Recommendations

The noise, vibration and harness analysis of diesel engines has been an active area of research during past few years. This work has tried to deal with some aspects of this issue. There are many areas in which current further work can be done in future [1]. Some of these include: 1. Quantification of noise emissions. a. Subjective approach Some of indices used for this purpose include: i. Ranking-Various subjects may be asked to rank sound emissions according to annoyance in a scale of 1 to 10. However number of samples must be kept low to avoid complexity [2–4]. ii. Comparison in pairs-In this method various subjects may be asked to evaluate relative judgments on the basis of pairs, however this method can be exhaustive as number of pairs can be large [5, 6]. b. Objective approach-Various psychoacoustic indices that can be used for evaluation include: i. Loudness-It is parameter for intensity evaluation and has unit of phon or sone. Loudness level is SPL of a pure tone plane wave of 1 kHz frequency as perceived by human ears in frontal direction [7]. ii. Sharpness-A 60 dB sound wave of 1 kHz frequency has sharpness of 1 acum. Sharpness of a soundwave can be lowered by either adding low frequency components or by decreasing high frequency components [8]. iii. Roughness-This parameter takes into account modulation of waves. Its unit is aper.1 asper is roughness of a tone of 1 kHz at 60 dB which is modulated by 70  Hz frequency with degree of modulation unity [9]. iv. Impulsiveness-This parameter represents impulsiveness of sound pressure level. It represents the amplitude and frequency of occurrence of peaks. Its unit is Kurt and is most significant during ideal running of engines [8]. 2. Motion of gudgeon pin inside pin hole needs to be taken into account.

Conclusions and Future Work 3. Piston pin is held inside hole either by a full floating system or a semi floating one. For case of full floating system both pin and connecting rod are made of steel, whereas in case of semi floating system piston is made of aluminum alloy and pin is made of steel. Hence a semi floating system is subjected to more noise due to differences in thermal expansion coefficients of different materials used. It has been observed that pin rotates counter clockwise inside hole before strikes the wall of piston vertically in crank angle duration 20° BTDC30° BTDC [10]. Further movement of oil inside pin hole can be visualized by particle tracking velocimeter(PVT). 4. Use of gap sensors/Telemeter device to study piston secondary motion using different skirt profiles. Frictional power losses for different skirt profiles can be evaluated for various engine strokes. Skirt profile having recess at top and bottom part of skirt has shown minimum frictional forces as it has better lubrication load bearing surface [11]. 5. Use of AVL EXITE for modelling of piston motion. This takes into account thermal distortions of liner using GUID (Piston-liner guidance) and EPIL (Elastic piston liner contact) approaches [12]. Surface velocities can be analyzed both in time and frequency domains towards thrust as well as anti-thrust side. At higher speeds, in conjunction with higher inertial forces, piston secondary motion has been found to fall. Hence both approaches have shown almost same results [13]. 6. Investigation into effects of bubble formations, mist and cavitation of lubrication oil during analysis of secondary motion of piston [14]. 7. Use the discussed methodology by varying type of injection or use of EGR or turbo charging. a. Effects of post injection-specific consumption of fuel NOX and emissions can be reduced by increasing the amount of post injected fuel and advanced injection timings.However smoke emissions were found to remain unaffected by post injection [15]. b. Effects of EGR-EGR has been found to reduce combustion noise above 300 Hz range, however excessive use of EGR causes lowering of thermal efficiency and increase in emissions [16]. c. Effects of turbocharging-[17]

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240 Liquid Piston Engines 8. Use of Blind Source Separation (BSS) and Independent Component Analysis(ICA)methods for effective noise source separation.

References and Bibliography 1. Monelletta, L., “Contribution to the study of combustion noise of automotive diesel engines”, Phd Thesis, University polytechnic velencia, 2010. 2. Otto, N.C., Amman, S., Eaton, C., Lake, S., “Guidelines for jury evaluations of automotive sounds”, SAE Technical paper-1999-01-1822, 1999. 3. Guski, R., “Psychosocial methods for evaluating sound quality and assessing acoustic information”, Acta Acustica, 1997. 4. Bisping, R., and Giehl, S., “Psychological analysis of the sound quality of vehicle interior noise: field and laboratory experiments”, proceedings of AVL conference on engine and environment, pp. 65–86, 1996. 5. Hussain, M., Golles, J., Ronacher, A., and Schiffbanker, H., “Statistical evaluation of an annoyance index for engine noise recordings”, SAE paper no 911080, 1991. 6. Kahn, M., Johansson, O., Lindberg, W., and Sundback, U., “Development of an annoyance index for heavy duty diesel engine noise using multivariate analysis”, NCEJ, vol 4, pp. 45, 1997. 7. Zwicker, E., and Fastl, H., “Psychoacoustic-Facts and models, II edition” Springer, 1999. 8. Schiffbanket, H., Brandl, F., and Thien, G., “Development and application of an evaluation technique to assess the subjective character of engine noise”, SAE paper no 911081, 1991. 9. Peluger, M., Holdrich, R, Brandl, F., and Biermayer, W., “Subjective assessment of roughness as a basis for objective vehicle interior noise quality evaluation”, SAE paper no 1999-01-1850, 1999. 10. Kondo, T., and Ohbayashi, H., “Visualization of Oil behavior when piston pin noise occurs”, Honda R&D Technical Review, 2011. 11. Kim, S., Shah, P., “A study of friction and lubrication behavior for gasoline piston skirt profile concepts”, SAE Technical Paper no 09PFL-1163, 2009. 12. Kocaoglu, C., Tabak, M., “Comparison of two modeling techniques for piston-liner interaction in terms of piston secondary motion using AVL exite”, OTEKON 2014, Bursa, Turkey, 2014. 13. AVL EXITE Power unit users guide, Vol 3, pp. 404, 2009. 14. www.SAE.com 15. Park, Y., Bae, C., “Effects of single and double post injection on Diesel PCCI combustion”, SAE Technical Paper 2013-01-0010, doi:10.4271/2013-01-0010, 2013.

Conclusions and Future Work

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16. Shibata, G., Ushijima, H., Ogawa, H., and Shibaike, Y., “Combustion Noise Analysis of Premixed Diesel Engine by Engine Tests and Simulations,” SAE Technical Paper 2014-01-1293, doi:10.4271/2014-01-1293, 2014. 17. Rakopoulos, Giakoumis, “Experimental study of combustion noise radiation during transient turbocharged diesel engine operation”, Vol 36, pp. 4983–4995, 2011. 18. Ochiai, K., and Yokota, K., “Light-weight, quiet automotive DI diesel engine oriented design method”, SAE paper 820434, 1982. 19. Maetani, Y., Niikura, T., Suzuki, S., Arai, S., and Okamura, H., “Analysis and reduction of engine front noise induced by the vibration of the crankshaft system”, SAE paper 931336, 1993.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

Glossary

TDC BDC Isothermal Isobaric Isentropic Isometric LTD HTD

Top dead center Bottom dead center A process in which temperature remains constant A process in which pressure remains constant A process in which entropy is constant A process in which volume is constant, also known as isovolumetric Low temperature difference High temperature difference

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List of Important Terms

1. Stoichiometric fuel/air ratio: A mixture that contains just enough air for complete combustion of all fuel mixture is known as chemically correct or Stoichiometric ratio. 2. Mean effective pressure: Average pressure inside cylinders of an IC engine based on calculated or measured power output is known as mean effective pressure. 3. Twin-scroll turbochargers: A turbocharger comprising a turbine, director, and compressor. The turbine may be formed as a turbine wheel surrounded by at least two scrolls.  The at least two scrolls may direct exhaust gases supplied thereto toward the turbine wheel to cause rotation thereof. The director may control distribution of the exhaust gases between the at least two scrolls to optimize circumferential velocity in the scroll or volute, and thus impingement velocity on the turbine. The compressor may be driven by the turbine. 4. Turbo Lag: A turbocharger uses a centrifugal compressor, which needs rpm to make boost, and it is driven off the

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5.

6.

7.

8.

exhaust pressure, so it cannot make instant boost. It is especially hard to make boost at low rpm. The turbo takes time to accelerate before full boost comes in; it is this delay that is known as turbo lag. To limit lag, it is important to make the rotating parts of the turbocharger as light as possible. Larger turbos for high boost applications will also have more lag that smaller turbos, due to the increase in centrifugal mass. Impeller design and the whole engine combo also have a large effect on the amount of lag. Turbo lag is often confused with the term boost threshold, but they are not the same thing, lag is nothing more the delay from when the throttle is opened to the time noticeable boost is achieved. Turbo Boost: Usually measured in pounds per square inch, it is the pressure the turbocharger makes in the intake manifold. One of the ways to increase airflow through a passage is to increase the pressure differential across the passage. By boosting the intake manifold pressure, airflow into the engine will increase, making more power potential. Boost is also measured in Bar. One Bar equals 14.7 psi. Boost Threshold: Unlike turbo lag, which is the delay of boost, boost threshold is the lowest possible rpm at which there can be noticeable boost. A low boost threshold is important when accelerating from very low rpm, but at higher rpm, lag is the delay that you feel when you go from light to hard throttle settings. Waste gate: The waste gate is a valve that allows the exhaust gasses to bypass the turbine. The waste gate relies on boost pressure to open it. Spliced into the waste gate pressure feed there must be some form of pressure bleed. By bleeding pressure to the waste gate, it is possible to control the amount of boost by reducing the pressure at the waste gate. Turbo Cool Down: A turbocharger is cooled by engine oil, and in many cases, engine coolant as well. Turbos get very hot when making boost, when you shut the engine down the oil and coolant stop flowing. If you shut the engine down when the turbo is hot, the oil can burn and build up in the unit (known as “coking”) and eventually cause it to leak oil (this is the most common turbocharger problem). It is a good idea to let the engine idle for at least 2 minutes after any time you ran under boost. This will cool the turbo down and help prevent coking.

List of Important Terms 9. Multi port fuel injection (MPFI): In the multi point fuel injection system an injector is located in the intake manifold passage. The fuel is supplied to the injectors via a fuel rail in the case of top fed fuel injectors and via a fuel galley in the intake manifold in the case of bottom fed fuel injectors. MPFI systems provide better performance and fuel economy as compared to TBI. Most of the MPFI systems use one injector per cylinder but in certain applications up to two injectors per cylinder is used to supply the required fuel for the engine. 10. Central multi-port fuel injection (CMFI): This is a variation of MPFI system but in this case the injectors (usually one per cylinder) are located in a plastic molded pod and the fuel is distributed to the intake ports via a polymeric hose. To avoid fuel distribution variations a fuel pressure activated poppet valve is installed at the end of the hose. The injectors are activated via the ECU in a similar fashion as in the MPFI fuel systems. 11. Tuned port injection (TPI): A TPI is a fuel/air management system that has a tuned induction system to optimize airflow to each cylinder. This system was developed to obtain the broadest possible torque curve. A single throttle body and one injector per cylinder are used in this configuration. The intake manifold incorporates long runners whose length is tuned to the desired torque curve. For low and mid range torque longer runners are utilized in this application. 12. Direct fuel injection (DFI): In a direct fuel injection system one injector is located in the cylinder head for each cylinder. The high-pressure fuel (single fluid) or low-pressure air/fuel mixture (dual fluid) is metered directly into the combustion chamber when the electromagnetic valve is activated by the ECU. This fuel injection system offers the latest in engine management systems and offers the best in engine performance, low exhaust emissions and fuel economy.

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Bibliography

Watson and Janota, 1984, “Turbo charging of I.C. engines,” Macmillan press, New York, 1983. William Harry Crouse, Donald L. Anglin, “Automotive engines,” Glencoe, 1994, ISBN 9780028010991. J.B. Heywood, 1998, “Internal Combustion Engine Fundamentals,” McGraw-Hill Education. V. Ganesan 1994, “I.C. engines,” Tata Mcgraw hill, New Delhi, ISBN 978-1-25-900619-7. Vasandani, Kumar, 1979, “Heat Engineering: In MKS and SI Units,” Metropolitan . http://www.sae.org/images/books/toc_pdfs/MRSB166.pdf Kant, K., Pati, A., Viswanath, B., and Thiyagarajan, R., “Cyclic Irregularities in Idle and Fuel Delivery Variation of a Rotary Fuel Injection Pump,” SAE Technical Paper 2004-32-0056, 2004, doi:10.4271/2004-32-0056. Han, Z., Henein, N., and Bryzik, W., “A New Ignition Delay Formulation Applied to Predict Misfiring During Cold Starting of Diesel Engines,” SAE Technical Paper 2000-01-1184, 2000, doi:10.4271/2000-01-1184. Brunt, M. and Platts, K., “Calculation of Heat Release in Direct Injection Diesel Engines,” SAE Technical Paper 1999-01-0187, 1999, doi:10.4271/1999-01-0187 AUTO ENGINEER MAGAZINE (EDITION – AUG, 2007) http://ae-plus.com/ AUTO-CAR MAGAZINE www.magazineexchange.co.uk/Autocar-2007

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248 Liquid Piston Engines MTZ Magazine (MAY, 2006) http://www.magazineexchange.co.uk/tractor-magazine-june-2006-issue.html Cummins turbo products magazine https://www.cumminsturbotechnologies. com/sites/g/files/.../f/.../HTi_edition_12.pdf

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan. © 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

Index Aerodynamic noise, 171 Alpha Stirling engine, 2, 3, 46 compression stroke, 10–11 power stroke, 9 transfer phase, 8–9 Beal engine, 13 Bearing noise, 171–172 Beta Stirling engine, 2–3 compression stroke, 17–18 expansion phase, 16 rhombic drive, 17–18 structure, 14 vertical configuration, 15 C.I. See Compression ignition (C.I) engine Carburetor-based turbocharger, 212–213 Cepstrum analysis, 195–197 Clever pumps capillary action, 44 human heart, 44, 45 human impulse, 42–44 CN. See Combustion noise (CN) Combustion based noise, 179–180 cyclic variations effects, 188 in cylinder pressure analysis, 187 in cylinder pressure decomposition method, 189–191

diesel engines combustion process conventional spray formation, 181 multiple injection methods, 182–183 NOx and soot formation, 181–182 phase analysis, 183 three phases, 180–182 evaluation, 193 AVL structural response function and structural attenuation, 194 cepstrum analysis, 195–197 combustion noise (CN) levels, 198–199 transfer function, 194–195 factors effecting, 186–187 generation process, 184–186 heat release rate effects, 187–188 mathematical model, 192–193 MBF50 estimation, 199 resonance phenomenon, 189 Combustion noise (CN), 168, 198–199 Combustion noise index, 149–150 Compression ignition (C.I) engine, 204–205 COMSOL, 114–116 Crank train and engine block vibrations, 171 Croaking and clatter motion, 114

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Index

Customer satisfaction issues durability issues, 87–88 system design, 88–89 system durability, 89 testing of engines, 88 DENOx catalyst, 232 Detonation, 208 Diesel engines combustion process (see Combustion based noise) knock detection, 161–162 noise sources aerodynamic, 171 bearing, 171–172 combustion, 168 crank train and engine block vibrations, 171 gear train, 170–171 noise generation mechanism, 174–175 piston assembly, 168–170 simulation tools, 167 timing belt and chain, 172–174 total noise contribution, 175 valve train, 170 NVH (see Noise, vibration and harness (NVH) features, diesel engines) power density function (PSD), 162 time frequency analysis, 162–163 turbocharging, 206–207 Wavelet analysis, 163–164 Double-acting engine Beal engine, 13 structure, 11–12 working, 13–14 Dual cylinder engine noise model, 157 EGR. See Exhaust gas recirculation (EGR) ENGDYN, 171 Exhaust gas recirculation (EGR), 228–229, 239

Fast Fourier Transformation (FFT), 155 FEA. See Finite element analysis (FEA) FFT. See Fast Fourier Transformation (FFT) Finite element analysis (FEA), 149 Fluidyne evaporation, 61 principle, 55–57 working, 57–61 Free-piston Stirling engines advantages and disadvantages, 7 alpha, 2, 3 compression stroke, 10–11 power stroke, 9 transfer phase, 8–9 beta, 2–3 compression stroke, 17–18 expansion phase, 16 rhombic drive, 17–18 structure, 14 transfer stroke, 15–16 vertical configuration, 15 double-acting-engine Beal engine, 13 structure, 11–12 working, 13–14 structure, 6–7 Gamma type engine, 3–4 compression stroke, 20–21 expansion stroke, 20 ringbom engine, 22–24 ringbom Stirling, 24 compression stroke, 27 expansion or power stroke, 26–27 transfer stroke, 25–26 Sneft engine, 21–22 structure, 18 transfer stroke, 19 Gasoline direct injection (GDI), 225–226

Index GDI. See Gasoline direct injection (GDI) Gear train noise, 170–171 Heat engines energy transfer, 39 heat flow, 40 ideal engine P-V curve, 42, 43 law of thermodynamics, 38–39 refrigerators and, 41–42 work efficiency, 40–41 Human heart, 44, 45 Human impulse pump, 42–44 IC. See Internal combustion (IC) engine In cylinder pressure analysis, 187 In cylinder pressure decomposition method, 189–191 Intercooler, 213 Internal combustion (IC) engine, 1–2 Jet, 34 Kinetic pumps, 33–36 Lead covering method, 154 Liquid piston engines applications, 80 assembly, 70–71 calcuation, 71–72 clever pumps capillary action, 44 human heart, 44, 45 human impulse, 42–44 comparison within existing commercial devices, 76–78 contributions, 237 design characteristics, 68 choices, 69–70 layout, 67–68 major components and costs, 70 materials, 68–69

253

engine performance, 67 experiments, 72–74 factors affecting amplitude, 66–67 fluidyne evaporation, 61 principle, 55–57 working, 57–61 future prospects, 79–80, 238–240 global groundwater withdrawal, 31–32 heat engines energy transfer, 39 heat flow, 40 ideal engine P-V curve, 42, 43 law of thermodynamics, 38–39 refrigerators and, 41–42 work efficiency, 40–41 improvements, 78 liquid column tuning, 63–64 losses, 65–66 motion analysis, 64 numerical analysis, 80–83 objectives, 32 outcomes, 74–76 polluted water percentage and sources, 30–31 pumping setups, 62–63 pumps kinetic and jet, 34 performance curves, 37–38 positive-displacement vs. kinetic, 35 selection, 36 types, 33 regenerator, 61–62 solar energy and, 31–32 Stirling engine history and development alpha-type, 46–47 earliest version, 46 energy conversion, 47–48 engines comparison, 48, 49 important milestones, 47 steam engine, 45–46

254

Index

low temperature difference, 54–55 operation, 48 assumptions, 51 components, 50 stages, 51–53 pros and cons, 53–54 working gas, 53 water sources, 29–30 Lubrication dynamics background, 91–92 friction features, 93–94 friction reduction, 94–95 piston secondary motion simulation and piston slap, 102 COMSOL, 114–116 croaking and clatter motion, 114 factors affecting, 103 force analysis force and moment balance equations, 119–120 inertial force variation, 118–119 piston velocity, 117–118 force distribution, 107 free body, 105–106 frictional forces, 133 modes of contact, 112–113 oil film thickness behavior, 108–109 piston side thrust force, 132–133 rattling motion, 113–114 schematic diagram, 104–105 side thrust force, 107–108 skirt design parameters effects engine load, 128 engine speed, 126–129 inertia of connecting rod, 128 length variations, 126, 127 lubrication oil supply, 130 piston pin offset, 120–123

shape, 130 size, 130–131 skirt-liner gap, 121, 123–126 surface finish, 130 slapping motion numerical model, 131 slider crank mechanism dynamics, 111 squeezing velocity, 109–110 Stribeck lubrication curve, 104 system mobility determination block velocity and mobility, 136, 137 block vibrations, 140–142 dynamic parameters, 136, 138 mechanical mobility, 133–134 piston lateral motion, 142–143 piston mobility, 136 piston side thrust forces, 134, 135 piston tilting motion, 138–140 piston velocity, 134–135 time frequency analysis, 109–111 transferred energy behavior, 109 piston-assembly dynamics, 95 Reynolds equation, lubrication oil pressure fuel energy total dissipation, 96 interpretation, 97 nodal representation, 98 oil pressure distribution, 100–103 pressure variation, 97–98 surface nodal representation, 98–99 varying speeds and loads effects, 94 Manometer, 72, 73 MISO. See Multiple input-single output (MISO) system Mother Wavelet function, 163 Multiple input–single output (MISO) system, 157

Index Noise generation mechanism, 174–175 Noise, vibration and harness (NVH) features, diesel engines automobiles sales trend, 145–146 imperial formulation combustion noise index, 149–150 in cylinder pressure spectrum, 149 sound pressure levels variations, 151, 152 internal combustion engine, 146 analysis approaches, 149 noise and vibration sources, 148 power train system, 147 noise sources frequency ranges, 153 lead covering method, 154 schematic representation, 151–152 spectro filters, 156–157 surface vibration method, 154–156 V6 engine, 153 supply, 146, 147 NVH. See Noise, vibration and harness (NVH) features, diesel engines Osmosis and reverse osmosis, 44, 45 Peristaltic pumps, 33, 35 Piston assembly noise, 168–170 Piston secondary motion simulation and piston slap, 102 COMSOL, 114–116 croaking and clatter motion, 114 factors affecting, 103 force analysis force and moment balance equations, 119–120 inertial force variation, 118–119 piston velocity, 117–118

255

force distribution, 107 free body, 105–106 frictional forces, 133 modes of contact, 112–113 oil film thickness behavior, 108–109 piston side thrust force, 132–133 rattling motion, 113–114 schematic diagram, 104–105 side thrust force, 107–108 skirt design parameters effects engine load, 128 engine speed, 126–129 inertia of connecting rod, 128 length variations, 126, 127 lubrication oil supply, 130 piston pin offset, 120–123 shape, 130 size, 130–131 skirt-liner gap, 121, 123–126 surface finish, 130 slapping motion numerical model, 131 slider crank mechanism dynamics, 111 squeezing velocity, 109–110 Stribeck lubrication curve, 104 system mobility determination block velocity and mobility, 136, 137 block vibrations, 140–142 dynamic parameters, 136, 138 mechanical mobility, 133–134 piston lateral motion, 142–143 piston mobility, 136 piston side thrust forces, 134, 135 piston tilting motion, 138–140 piston velocity, 134–135 time frequency analysis, 109–111 transferred energy behavior, 109 Piston slap, 168–170 Piston-assembly dynamics, 95 Positive displacement pumps, 33, 35 Power density function (PSD), 162 PSD. See Power density function (PSD)

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Index

Pumps kinetic and jet, 34 performance curves, 37–38 positive-displacement vs. kinetic, 35 selection, 36 types, 33 Rattling motion, 113–114 Regenerator, 61–62 Resonance phenomenon, 189 Reynolds equation, lubrication oil pressure fuel energy total dissipation, 96 interpretation, 97 nodal representation, 98 oil pressure distribution, 100–103 pressure variation, 97–98 surface nodal representation, 98–99 Rhombic drive, 5–6, 17–18 Ringbom engine, 22–24 Ringbom Stirling, 24 compression stroke, 27 expansion or power stroke, 26–27 transfer stroke, 25–26 Ross engine, 11 Scalogram, 163 SEA. See Statistical energy analysis (SEA) Sneft engine, 21–22 Sound pressure level (SPL), 155 Spark ignition (S.I) engines. See Turbo charging, S.I engines Spark plug based system, 212–213 Spectro filters, 156–157 SPL. See Sound pressure level (SPL) Statistical energy analysis (SEA), 149 Stick slip noise, 93 Stirling engine history and development alpha-type, 46–47 earliest version, 46 energy conversion, 47–48

engines comparison, 48, 49 important milestones, 47 steam engine, 45–46 low temperature difference, 54–55 operation, 48 assumptions, 51 components, 50 stages, 51–53 pros and cons, 53–54 working gas, 53 Stirling engine system advantages and disadvantages, 4 free-piston (see Free-piston Stirling engines) gamma type, 3–4 (see also Gamma type engine) internal combustion (IC) engine vs., 1–2 mechanical design features, 2 rhombic drive, 5–6 swash plate drive mechanisms, 5 wobble-plate mechanisms, 4–5 Stribeck lubrication curve, 104 Surface vibration method, 154–156 Swash plate drive mechanisms, 5 Thermocouple, 72, 73 Time frequency analysis, 162–163 Timing belt and chain noise, 172–174 TPA. See Transfer path analysis (TPA) Transfer path analysis (TPA), 149 Turbo charging, S.I engines, 203–204 boosting techniques exhaust gas recirculation (EGR), 228–229 gasoline direct injection (GDI), 225–226 variable compression ratio (VCR), 226–227 variable turbine geometry (VGT), 227–228 charge air cooling, 224

Index components direct-injection systems, 213 exhaust waste-gate systems, 211–212 inlet and exhaust manifolds, 208–210 matching turbocharger for desired boost pressure, 211 spark plug based system, 212–213 turbocharger boost pressure control system, 210–211 diesel engines, 206–207 downsizing, 225 efficiency improvement, 208 emissions control, 231–232 fundamentals, 204–205 gasoline engines, 207–208 ignition timing and knock, 223–224 intercooler, 213 knock reduction methods, 223 operational problems, 222–223

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scope, 233–234 turbochargers, 205 advantages and disadvantages, 206 compressor, 218–221 specifications, 213–216 turbine, 217–218 Valve train noise, 170 Variable compression ratio (VCR), 226–227 Variable turbine geometry (VGT), 227–228 VCR. See Variable compression ratio (VCR) VGT. See Variable turbine geometry (VGT) Waste-gate control system, 211–212 Wavelet analysis, 163–164 Wiener filter, 156–157 Wobble-plate mechanisms, 4–5