Formula SAE Performance Exhaust Design

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Final Project Report 2011, SEIT, UNSW@ADFA

An Investigation into Formula SAE Performance

Exhaust Design and Analysis

Anthony I. McLeod 1

University of New South Wales at the Australian Defence Force Academy

The Formula SAE competition demands that teams pursue synergistic designs in

creating a competitive high performance vehicle. As a result, the design of each

component fitted to a vehicle, including that of the exhaust, must be undertaken with

a sound foundation of technical understanding in conjunction with creativity and

innovation. Exhaust design is shown to make a significant contribution to engine

performance, economy and noise attenuation. Hence, this work aims to assist the

ADFA Formula SAE team of 2012 develop an understanding of current exhaust

analysis and tuning techniques such that they may be innovatively applied to thedesign of a high performing exhaust system as a part of a holistic engine tuning

approach. Extensive research has been conducted into the mechanisms by which an

exhaust design may enhance engine performance and attenuate noise. In particular,

an exhaust design is understood to effect engine performance via influences upon

engine scavenging. Furthermore, the action of automotive silencers was identified to

be governed by their ability to manage the mass flow rate from the exhaust outlet as

opposed to that of acoustic theory. In addition, research has identified methods such

mechanisms may be analysed and predicted. Engine simulation software Ricardo

WAVE was used to demonstrate and analyse the performance and noise attenuation

implications of exhaust system componentry and their design parameters. A volume

restricted silencer design proposed by Professor Blair of the University of Belfast

formed the basis of further experimental and theoretical analysis of the governing

principles of silencer operation. Specifically, a derivative of this design concept wasmanufactured with in-built variability to enable an experimental investigation of the

design and to also help validate data obtained using WAVE. Finally, WAVE was used

to enable a theoretical analysis which underpinned a design proposal for a high

performing silencer.

Contents

I. Introduction .............................................................................................................................................. 3

A. Motivation ........................................................................................................... ....................... 3

B. Project Aims ........................................................................................................ ....................... 3 C. Project Methodology ........................................................................................... ....................... 3

II. Part A - Literature Review ................................................................ ........................................................ 3

A. Exhaust design for engine scavenging performance ................................................................ ... 4

B. Acoustics, Vehicle Noise and Exhaust Silencing ....................................................................... 5

C. Exhaust Silencing ................................................................................................ ....................... 6

D. Design and Modelling of Exhaust Systems ................................................................................ 9

E. Conclusion ..................................................... ................................................................. ............ 9

III. Part B – Concept Development and Investigation .............................................................. ..................... 10

1 SBLT, School of Engineering & Information Technology, ZEIT4500


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A. WAVE Model Development ............................................................................... ..................... 10

B. Experimental Silencer Parameter Study ................................................................................... 10

C. WAVE Investigation – Exhaust Design Parameters for Engine Scavenging ........................... 13

IV. Part C – Preliminary Design and Design Proposal ................................................................................. 15

A. Design Requirements................................................................................................................ 15 B. Silencing Strategy ................................................................ ..................................................... 15

C. Design Investigation and Definition ......................................................... ................................ 16

D. Design Proposal ........................................................ .................Error! Bookmark not defined.

E. Vehicle Integration of Exhaust System .................................................................................... 18

V. Limitations of Ricardo WAVE .......................................................... ..................................................... 19

VI. Conclusion ........................................................ ................................................................. ..................... 19

VII. Recommendations and Future Work ................................................ ...................................................... 19

Acknowledgements .............................................................. ................................................................. .......... 20

References ...................................................... ................................................................. ................................ 21

APPENDICESAppendix A. Combined theoretical and experimental investigation into exhuast pipe geometry A1

Appendix B. Exhaust pipe optimisation using NSAGA2 and ANSYS Fluent A2

Nomenclature

ADFA = Australian Defence Force Academy

CFD = Computational Fluid Dynamics

SPL = sound pressure level (dB)

Q = volume flow rate(m3/s)

D = pipe diameter (m)

̇ = flow velocity(m/s)


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I. Introduction

A. Motivation

The Formula SAE® competition constitutes a variety of rules and regulations that aim to challenge designteams whilst maintaining fairness and safety. A number of pertinent rules to this study [1] include:

The vehicle‘s engine must be a four cycle piston engine with a maximum swept displacement of 610cc,

An intake restrictor must be fitted with a maximum diameter of 20.0 mm for vehicles operating withgasoline and 19.0 mm for vehicles operating with E85, and

A vehicle‘s measured noise level must be less than 110 decibels.

It is obvious from these rules in particular that teams are challenged to form a competitive advantage via [2]

synergistic vehicle designs by applying technical knowledge innovatively as well as through the application of

advanced performance tuning techniques. In this way, teams may attain the necessary combination of power and

efficiency to be competitive throughout a series of trying auto-cross events.

The concept of ―exhaust tuning‖ has been under development for over 60 years [3]. In this time,

exhaust design has been proven to have a marked influence upon the performance and efficiency of an engine by way of power output, specific fuel consumption, heat production and radiated noise level. It is as a

consequence of the flow on effects of such factors that the implementation of a sound understanding in the

design of an exhaust is crucial in order to obtain a high performing racing vehicle. For the benefit of

competitiveness, it is therefore important for the Formula SAE® team representing ADFA to learn to approach

the design of the exhaust in such a way that maximizes performance of the competition vehicle and ensurescompliancy.

B. Project Aims

The ADFA Formula SAE® team of 2012 has purchased a Yamaha WR450 single cylinder engine to be

integrated into a new competition vehicle. It is therefore the intent of this project to assist the team to understand

the potential benefits of exhaust tuning as well as the methods that are available in the analysis and design of an

exhaust. This project will consist of the validation of theories currently used to enhance engine output.

Furthermore this investigation will be extended to noise generating phenomenon and associated analytical and

prediction techniques. The project will employ a one-dimensional engine simulation software, Ricardo WAVEto then undertake analysis to be validated experimentally, culminating in a final design proposal for a new high

performing silencer.

C.

Project MethodologyThis project utilised a series of methods to carry out an investigation into exhaust systems and their design.

Initially, extensive research constituting a literature review was undertaken to build a knowledge base requisite

of applied analysis and design. The complex trade-offs found to characterise silencer design then motivated an

experimental investigation using a promising silencer design concept that was developed and proven upon a

similar engine as the Yamaha WR450, by Professor Blair of the University of Belfast. An experimental mufflerwas manufactured based upon this design concept which incorporated in-built variability to enable an

experimental parameter study of silencer attenuation. DOE methodology was utilised to conduct this

experimental study which employed an available and operable WR250 motorbike in lieu of the engine testing

rig still under development by the FSAE team. This experiment obtained insertion loss for the silencer within

the frequency domain to such that the governing principles of silencer operation could be identified. An engine

simulation model was then developed using the one-dimensional engine simulation software Ricardo WAVE.This software then underpinned the demonstration and theoretical analysis of performance exhaust tuning and

silencer theories. Exhaust performance aspects investigated include the concept o f ‗tuned length‘, the effect ofstepped pipes and diffuser components as well as the nature of ‗inertial scavenging‘ phenomena. Performance

data obtained is discussed such that these methods become yet another tool for the ADFA FSAE team to utilise

within an integrated engine tuning process. The silencer experiment was duplicated within WAVE to provide a

level of validation of the developed model. Continued silencer analysis employed the WAVE transmission loss

work bench. The conclusions drawn from these analyses then facilitated the development of a design proposal

for a high performance silencer.

II.

Part A - Literature ReviewQuality design of an exhaust system requires a sound understanding of its contribution to both the overall

power output of an engine and to noise attenuation. Furthermore it is important to understand the mechanisms

that enable these contributions as well as their significance. A wide variety of sources were studied to determine

current exhaust theories, design and analysis methods as well as to better understand the restrictions imposed by

Formula SAE noise regulations.


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A. Exhaust design for engine scavenging performancePerformance considerations of exhaust design are a result

of the nature of the gas exchange process in a four stroke

engine. This process includes a period of valve overlap

where both the intake and exhaust valves are open

simultaneously as seen in Fig.(1). Without due regard by the

designer this period could see the induction of exhaust gasesinto the cylinder as shown in Fig.(2), effectively reducing the

amount of fresh combustibles ingested and therefore overall power.

Performance aspects of exhaust design are concerned

with minimising such residual quantities or otherwise stated

as maximising scavenging efficiency of the engine. This is

achieved, one way or another by reducing the exhaust valve

pressure during valve overlap such as to bias this exchange process to achieve this scavenge.

Exhaust scavenging is achieved via two methods. This is

because the exhaust phase of the four stroke cycle consists of

not only the expulsion of a high speed column of exhaust

gases but also a pressure wave. Consequently, scavenging isachieved through techniques known as ‗wave tuning‘ and ‗inertial scavenging‘ depending on which of these

mechanisms we utilise.

The aim of wave tuning is described by Professor Blair of

the Univer sity of Belfast who states that ―the tuned exhaust

pipe harnesses the pressure wave motion of the exhaust

process to extract a greater mass of exhaust gas from the

cylinder during the exhaust stroke and initiate the induction

process during the valve overlap period.‖ This scavenging

effect is possible if a pressure wave originating from the

exhaust valve travel at the local acoustic velocity, over a

tuned length such that it is reflected back to the valve face as

a rarefaction wave, as seen in Fig.(3), in time to assist the gasexchange process during valve overlap.

The phasing of the exhaust valve and the pressure wave

is dependent upon the length over which the waves travel.

Commonly known as the ‗tuned length‘, it is defined by the

length of pipe bounded by the exhaust valve and adiscontinuity in the pipe of an area ratio of 6. This being a

point that significant wave action can operate from.

In addition, scavenging may be achieved via inertial

scavenging. This is a scavenging effect achieved as a result

of the inertia of a high velocity column of gas. It functionsunder the principle that a fixed volume flow rate is achieved

at a certain engine speed and for a fixed volume flow rate,

gas velocity varies inversely with pipe diameter. There thenexists a pipe diameter where the scavenging effect produces a

more than proportionate amount of power than pumping

work required to achieve an effective gas velocity.

In addition to the stated performance exhaust theory there exists a number of complicating factors for the

realistic exhaust system designer. Firstly, the periodic nature of wave phenomenon in the exhaust suggests that

whilst tuning may be carried out for the benefit of power at one engine speed this will inevitably lead to poorer

performance in another [4, 6]. Furthermore, tuning of the exhaust without due regard of the interactions taking

place with other mechanisms such as similar wave action occurring in the intake, has the potential to produce

irregular shapes including troughs and peaks within the power curve. As a consequence the drivabilitycharacteristics of the vehicle could diminish as a result of unpredictable power output behaviour.

Figure 1. Valve timing events showing valve

overlap

Figure 2. Poorly tuned engine ingests

exhaust gas into the cylinder during valce

overlap

Figure 3. Reflection of rarefaction wave at

exhaust pipe end which returns to the exhaust

alve


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Another perspective of the priorities of exhaust system design is provided by a parameter study conducted

by Sammut and Alkidas [11]. This study utilizes the engine simulation software Ricardo WAVE to quantify the

effects of and interactions between exhaust, intake and valve timing parameters. For a constant valve timing and

engine speed, Fig.(4) shows a comparison of the scavenging effect of the intake and exhaust measured in

volumetric efficiency. The data presented firstly shows that the individual contributions of the intake and

exhaust are independent as the contribution made by the exhaust is relatively constant for any intake length.

However, it is important to note that such independence should not be assumed between all parameters. All data presented herein illustrating variation in scavenging as a function of tuned length is obtained for constant valve

timing. Variation in valve timing would inevitably change the characteristics of the overlap period and thereforethe action of exhaust scavenging. The effects of this are well documented throughout literature but assumed

constant for the purpose of this investigation of exhaust design. Secondly, data presented in Fig.(4) also

concludes that the effect of exhaust tuning is relatively small compared to the benefits of intake tuning. As a

consequence of the diminishing significance of exhaust scavenging benefits, minimizing the losses conceded toincreased pumping losses whilst achieving sufficient noise attenuation becomes of relatively high importance if

a maximum amount of power is to be derived from the engine. With the realization that there as much potentialfor an exhaust system design to reduce performance as to improve it, the design of an efficient silencer becomes

crucial to the competitiveness of the vehicle. Moreover, it needs to be integrated within a system with minimal

prejudice towards efforts to attain an effective scavenge by providing low exhaust valve pressure at valve

overlap [6].

B. Acoustics, Vehicle Noise and Exhaust Silencing

Literature was consulted in order to define the problem of vehicle noise as well as to gain an appreciation of

current vehicle noise attenuation techniques such that this design issue could be effectively addressed. A noise

measurement of sound radiated from a vehicle is subject to a variety of sources including mechanical noise,

shell vibration radiated noise and duct noise where duct noise then consists of intake and exhaust tail pipe noise

[13]. Fig.(5) is provided to illustrate the prevalence of intake duct noise, being a source not considered here.

The sound pressure measured at any point in space is

relative to the radius defining the distance between thesource and the point of measurement as well as the

directivity of the source with respect to this radius vector.

Furthermore, an important consequence of the logarithmic

scale of sound measurement is that the sum of SPL from

multiple sources varies little from the maximum SPL [4] asseen in Fig.(5). Consequently, a vehicle silencing strategy

formulated to control sound pressure at a specified location

relative to the vehicle, needs to acknowledge the most

significant source at that location in order to effectively

control the final measurement. Therefore, the following

discussions detailing exhaust tail pipe noise attenuation canonly be effective within spatial regions where this is the

dominant noise source and sound pressure contributions of

Figure 5. Noise level of intake and exhaust duct

noise

Figure 4. Variation of volumetric efficiency with intake and exhaust length


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other sources such as intake duct noise and engine

noise become negligible.

Such conclusions then underpin the next priority

in forming an efficient vehicle noise attenuation

strategy being to recognize the dominant components

of exhaust tail pipe noise. The design of an efficient

muffler should then target these most significant noisecomponents in order to attain the required attenuation

level whilst maintaining low resistance to flow. Tail pipe noise component of duct noise and consists of

[8]:

1) Pulse/Engine noise which describes sound of

frequencies corresponding to harmonics of

the engine firing rate (EFR), as seen in

Fig.(6). The EFR is the rate at which theexhaust valve releases combustion gases

from the cylinder, and

2) Gas flow noise which consists of high

frequency broadband noise resulting from

pressure fluctuations inherent to the turbulentmean gas flow in the exhaust duct.

Fig. (7) [14] indicates that the pulse noise is

dominant at low engine speeds and is superseded by

flow noise as it increases in magnitude with volume

flow rate and engine speed. A silencer design must

therefore incorporate elements that can target the

dominant noise source at the engine speed of interest.

Specifically, silencing at low engine speeds must be

concerned with discrete harmonics of the engine

firing rate while silencing at high engine speeds is

more concerned with high frequency flow noise.

Pang et al [13] shows a direct proportionalcorrelation between flow noise and flow velocity and

therefore exhaust pipe diameter given by Eq (1), where

the volume flow rate is a function of engine speed. The

relationship between flow noise and diameter is seen in

Fig.(8). This shows that at high engine speeds whereflow noise is dominant, a fixed volume flow tranlsates

to greater flow velocity for a smaller diameter and

therefore increased noise emissions.

.

̇ ̇ ̇

(1)

The conversion of flow power to sound power isidentified by Wiemeler, Jauer and Brand [14] to be

relative to an efficiency factor that is proportional to

the flow mach number. They show that a critical flow

velocity mach number of 0.25 represents a transition between flow noise generation mechanisms leading to anicreased efficiency and increased flow noise sound pressure level (SPL).

C. Exhaust SilencingIn order to moderate exhaust tail pipe noise there exists a variety of muffler designs that are commonly

employed. The performance of a silencer may be characterized by its insertion loss defined as the difference in

measured SPL with and without the muffler fitted; its transmission loss which is defined by the difference in

SPL at the inlet and outlet of the muffler; or its effect upon the brake mean effective pressure. An efficient

silencer is defined here by a design that achieves a relatively high ratio of attenuation achieved to reduction in

engine power output. Silencers may consist of a single or a combination of standard silencing componentswhich include reactive, absorptive and resonator types. These components vary in the manner and efficiency

Figure 6. Frequency analysis of radiated vehicle noise

showing noise corresponding to EFR frequencies

Figure 7. Sound pressure with engine speed showing

increasing dominance of flow noise with engine speed

Figure 8. Variation in flow noise with flow

velocity caused by pipe diameter


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with which they enable the viscous dissipation of acoustic energy. Their unique action often makes them highly

effective attenuators in discrete frequency ranges or otherwise less effective over a more general range of

frequencies. As a result hybrid silencers aim to utilise a combination of such components in order to form an

effective broadband attenuator. A summary of the acoustic theory for these common silencer types is attached in

Annex A.

Blair quotes the work of Coates [15] who shows that the sound pressure level at any point in space beyond

the termination of an exhaust system to the atmosphere, is a direct function of the instantaneous mass flow ratefrom the end of the exhaust pipe, the relative distance between source and microphone and the directivity of the

pipe end. The instantaneous mass flow rate was calculated using the Eq.(2) [4].

̇ ̇ (2)

This expression states that the radiated noise is a function of gas temperature, the discharge coefficient of the pipe end, the outlet diameter as well as pressure wave amplitude ratio travelling in the left and rightward

direction. As a result of this direct relationship with the mass flow rate Blair states that silencing is easily

achieved given an unlimited volume able to dampen the pressure and mass flow oscillations. However, when

subject to space restrictions the design of a silencer must conform to the following empirical design guidelines:

A silencer should have a minimum silencer-cylinder volume ratio of ten.

If this cannot be achieved the silencer must choke the exhaust system via a restrictive muffler in orderto sufficiently damp the mass flow rate for effective noise attenuation. (However increased back

pressure will result from increased restriction, therefore a silencer with minimal choke would represent

the most efficient attenuator).

Blair [4] uses this theory to conduct a study into the effectiveness of motorcycle silencers via experimental

and numerical methods. This study tests a plenum, absorption, diffusing and side-resonant type mufflers all with

a constant silencer-cylinder volume ratio of ten. Data shown in Fig.(9) and Fig.(10) illustrates that individually

these mufflers either offer excessive reductions in the BMEP of up to 30% whilst being unable to attenuate

noise sufficiently or offer negligible effect to power and noise.

A novel ‗two- box‘ hybrid silencer design seen in

Fig.(11), comprising an absorption and diffusing

silencer component, is then verfied to result in an

average reduction in BMEP of only 7% whilst notably

attenuating noise. This is seen to be a result of the

effectiveness with which the mass flow rate at theoutlet is reduced as seen in Fig.(11). Comparisons are

shown in Fig.(13) and Fig.(14), of achieved engine

performance and measured noise emission data for this

design as well as its individual constituent components.

Figure 11. Schematic of tow-box silencer

Figure 10. Noise characteristics of muffler varieties

determined by Blair

Figure 9. Torque characteristics of muffler varieties

determined by Blair


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Data illustrate the potential of the volume restricted two-box silencer design as an effective and efficient

attenuator of exhaust tail pipe noise. Noise spectra in Fig.(15) shows the attenuation achieved by the two-box

silencer as well as by individual absorption and diffusing silencer components. This demonstrates the highly

non-linear interaction between the absorption and resonant/diffusing components.Acoustic theory would suggestthat the effectiveness of this particular hybrid silencer represents a combination of the attenuation of the

diffusing silencer at low frequency and the attenuation of the absorption silencer at high frequency which is to a

limited extent demonstrated within Fig.(15). However, acoustic theory is experimentally shown by Blair to be

highly ineffective in accurately predicting the achieved attenuation from a silencing element. Data in Fig.(16)

and Fig.(17) compares the experimentally obtained attenuation with that predicted by acoustic theory which

Figure 13. BMEP with silencer component

Figure 14. Noise attenuation performance of silencer

components

Figure 12. Exhaust outlet mass flow rate for

silencer varieties

Figure 15. Noise spectra at 7500 rpm with two-boxsilencer

Figure 16. Noise attenuation of plenum anddiffusing type silencers Figure 17. Noise attenuation of side-resonant silencer


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III. Part B – Concept Development and Investigation

A. WAVE Model Development

A Yamaha WR450 engine simulation model was developed to underpin the continued concept developmentand analysis of the engine and associated componentry. A view of the WAVE model is provided in Fig.(18).

The model employed a host of user defined inputs available from tabulated data or otherwise from physical

measurements. Due to time and resource constraints an experimental validation could not be undertaken.However, the validity of the model utilised within this study is supported by agreement found between the

generated engine power output prediction and data generated independently by the Cal Poly FSAE team who

managed to conduct an experimental validation of their WR450 WAVE engine model. Power curves obtainedfrom the WAVE model as well as by data from Cal Poly FSAE team are provided in Annex (B).

Figure 18. WR450 Ricardo WAVE model

B. Experimental Silencer Parameter StudyHaving developed an engine simulation model of the Yamaha WR450, an orthogonal experiment was

designed to effect a parameter study of the ‗two- box‘ silencer design. The experiment would be implementedwithin WAVE as well as upon an experimental WR250 engine using a manufactured experimental silencer. The

purpose of this experiment was to investigate the achieved attenuation of the muffler concept, in addition to the

variation in this attenuation as a function of parameter modification. Furthermore, this experiment will be used

to quantify the effectiveness of silencer design theories including that of acoustic theory and of Blair‘s mass

flow rate theory. Upon comparison of theoretical and experimental data, comment will then be made as to thevalidity of the WAVE model (Only partial validation could be obtained as the WAVE model is a simulation of

the team‘s Yamaha WR450 whereas the physical experiment could only be conducted using a Yamaha WR250).

1. Experimental Silencer DesignAn experimental silencer was manufactured as per the CAD model in Fig.(19), based upon the ‗two- box‘

silencer design proposed by Blair.

Figure 19. CAD model of manufactured silencer showing initial discontinuity, absorption component and

expansion/resonator chamber


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In addition to the original concept, the manufactured design incorporates an initial discontinuity of area ratio

equal to six which is conformant to the findings of Blair. This was included to provide a decisive location that

may be used to define the tuned length as well as to decouple design parameters concerned with silencing from

performance aspects of the exhaust. (Blair‘s original design would instead tune from the end of the perforated

pipe. However, an industry SME advises that wave propagation through the perforated pipe enhances wave

degradation and therefore wave tuning effectiveness). This design was subject to some variation from the

original design as it was constrained by the availability of off-the-shelf (OTS) components which were preferredin order to simplify the manufacturing process. The design may be fully disassembled as per Fig.(19), such that

parameters including the choke size, the resonator length and packing density could be varied in accordancewith the experimental intent. Table 1 shows a comparison of the non-dimensional parameters of the Blair design

and the current experimental design.

Table 1. Non-dimensional parameter of Blair silencer and mancufactured silencer

Blair Experimental Design Comments

Inlet Diameter (D) 46.6 mm 51 mm OTS component and recommended

by industry SME

Major Diameter 2.58D 2.49D OTS

Absorption Length 8.58D 9.0D OTS

Resonator Length 4.29D Variable up to 5.88D Custom telescoping component

Perforated Area 19% 25% OTS

Silencer- Cylinder

Volume Ratio

15 14.5 - 19.5

2. Orthogonal Experiment DesignAcoustic theory predicts that the manufactured two-box silencer will achieve broadband attenuation as a

result of the combination of resonant effects of the expansion chamber and viscous dissipation of the absorption

silencer. However, as shown in experimental data obtained by Blair in Fig.(15), the operation of this silencer

does not explicitly conform to predictions underpinned by acoustic theory, nor does data show that attenuation

achieved is linear addition of the attenuation achieved by its constituent components. In contrast, Blair‘s mass

flow rate theory hypothesises that broadband attenuation achieved by this silencer is a direct consequence of themanner with which it damps the magnitude of the mass flow rate from the exhaust outlet. Consequently, an

experiment was conducted to identify the true manner of operation of this silencer so as to provide the means for

the design of an efficient silencer for the ADFA FSAE team.

In order to attempt to validate acoustic theory, noise measurements recorded the noise spectra such that the

insertion loss of the muffler could be calculated. Hence, validation of acoustic theory could be obtained if thisdata was to show agreement with transmission loss data calculated.

The experimental plan was based upon Taguchi Design of Experiment methods [30] for orthogonal

experiments. This experiment was then implemented upon a WR250 engine as well as implemented within

WAVE. Unfortunately, as the WAVE model has been developed to simulate a Yamaha WR450 engine and the

physical experiments were conducted upon a WR250 engine the comparison of these results were not able to

provide conclusive validation of the developed engine model. Instead these two sources of data were simplyused to comment on the nature of operation of the muffler as well as to confirm similarity of trending.

The chosen independent variables include the silencer parameters of resonator length, choke diameter and

packing density. Using the Taguchi L4 orthogonal array the experiment seen in Table 2 was formulated.

Table 2. Orthogonal experimental plan

Experiment Choke Diameter (mm) Resonator Length (mm) Packing Density (g/L)

1 20 100 200

2 20 300 100

3 30 100 100

4 30 300 200

3. Parameter Study Results and Discussion Noise measurements taken at position A and position B for experimental silencer variations as well as for the

standard WR450 muffler is shown in Fig.(C1) and Fig.(C2) in Annex C. WAVE data is also provided in

Fig.(C3) which shows the predicted sound pressure at position A for each of the experimental test silencers. As

expected, comparison with experimental data at position A does not show agreement of sound pressure

magnitude due the difference in engine displacement. However, good agreement is found as to the relativevariation between designs. Measurements obtained of SPL with frequency are provided in Fig.(C4) and


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Fig.(C5). This data shows variation in emitted noise from the unsilenced pipe with engine speed as well as the

broadband attenuation achieved by the manufactured silencer which is seen to be equivalent to a standard

WR450 muffler. Calculated insertion loss is provided in Fig.(C6), Fig.(C7) and Fig.(C8). Data provided is

limited to a frequency of 1000 Hz as consistently high levels of attenuation are achieved for all designs at

frequencies beyond this point. Transmission loss for the each of the silencer configurations was also calculated

using WAVE for comparison, and is seen in Fig.(C9) and Fig.(C10).

The data obtained is generally supportive of the theory of silencer design concerned with the management ofthe mass flow rate from the exhaust outlet as opposed to acoustic theory. For instance, within Fig.(C1), Fig.(C2)

and Fig.(C3) the most choked designs 1 & 3 record significantly lower sound pressure level recordingscompared to the less choked designs. Furthermore, both experimental and WAVE data agree that those designs

with a larger volume will attenuate noise to a greater extent. Plots in Fig.(C11), Fig.(C12), Fig.(C13) and

Fig.(C14) of outlet mass flow rate, generated with WAVE, illustrate the silencing action of the muffler

variations. These plots show data for all tested engine speeds. Prominent features include the higher peaks

recorded for the less choked designs as well as the higher steady mass flow rate recorded for the more choked

designs. This steady flow rate leading up to the peak is much more constant for choked designs, which increasesin magnitude at high engine speeds in comparison to the relatively less choked designs. This behaviour suggests

that designs utilising a 20mm orifice are likely to become aerodynamically choked leading to a rapid increase in

back pressure, but also attests to the effectiveness of a choke for the purpose of exhaust tail pipe silencing under

Blair‘s theory concerned with the outlet mass flow rate. Furthermore, fluctuations within this mass flow rate are

seen to be damped by silencers employing a larger expansion chamber volume.However, with reference to Fig.(C1), Fig.(C2) and Fig.(C3) it could also be argued that higher attenuation is

achieved by those designs with larger expansion chambers due to an increased ability to attenuate low frequency

noise as per acoustic theory for an expansion chamber. This low frequency noise is recorded in Fig.(C4) and

Fig(F5) as a source that is relatively constant as well as relatively elevated in comparison to other regions of the

noise spectra emanating from the unsilenced pipe. To investigate this possibility, the transmission loss for each

of the silencer configurations was attained from WAVE and is shown in Fig.(C9) and Fig.(C10). (This was

conducted within the WAVE transmission loss workbench which employs the well documented two source

method to compare sound power at the inlet and outlet of the silencer). As seen in Fig.(C10), the predicted

transmission loss below 500Hz for both designs 1 and 4 is seen to be consistently up to 5dB greater than for

designs 2 and 3 which employ a smaller chamber. However, it is questionable that this extra achieved

attenuation could be the main reason for these larger designs consistently out performing corresponding designs

with an equal choke diameter and smaller chambers. This doubt is particularly pertinent as transmission loss

data predicts generally lower attenuation achieved by larger designs 1 and 4 at higher frequencies, yetexperimental data states that these larger designs record lower SPL even at high engine speeds where high

frequency flow noise becomes more dominant. Furthermore, calculated insertion loss data does not show any

significant agreement with theoretical attenuation for the silencer represented by the transmission loss data.

Subsequently, the achieved attenuation of a silencer in practice is seen to be more dependent upon the manner in

which the design manages the exhaust outlet mass flow rate to atmosphere than the attenuation predicted byacoustic theory. Results leading to this conclusion are in line with published theory of Blair described

previously.

The value of acoustic theory is not, however, totally diminished as some agreement is found between

transmission loss data and calculated insertion loss data by way of comparative performance. Discrepancies

between these sources may also be exaggerated by inaccurate assumptions and experimental error. This wouldinclude the assumption of nil mean flow during the transmission loss analysis and aliasing within experimental

measurements. Furthermore, the consistently high attenuation achieved at high frequency by the manufactured

design agrees with the acoustic theory of absorption silencer. (The effectiveness of an absorption silencer wasalso experimentally determined by Blair and is shown in Fig.(17)). Therefore whilst acoustic theory and its

implementation may not take into account all non-idealities it may provide a good initial estimate of silencer

performance.

Further comment can be made as to the effectiveness of the manufactured design with reference to Fig.(C4)

and Fig.(C5). These show measured sound pressure with frequency at position B for engine speeds of 3000 and

7000rpm. These plots show that the test silencer offers significant attenuation averaged between 20dB and 30dB

which is also recorded for the standard WR450 muffler. The WR muffler has a silencer-cylinder volume ratio of

5 in its intended role upon a WR450 and a ratio of 7.5 for the test engine and as a result it employs a 10mm

choke in order to meet road authority noise regulations. The trade-off between silencer volume and choke ismade clear as experimental data concludes that the designed silencer of a volume ratio of 4 to 5 times greater

than the standard WR muffler, yet far less choked is able to achieve equivalent attenuation. This data thereby

emphasises the importance of exploiting available vehicle space for a silencer in order to minimise the level of

choke required and therefore minimise power losses. A comparison of these figures also illustrates the


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increasing prevalence of high frequency flow noise at high engine speed, a trend which is also shown by Honda

et al in Fig.(7).

A comparison of all SPL spectra data obtained experimentally is compared with corresponding data attained

from the WAVE model in Fig.(C15), Fig.(C16), Fig.(C17) and Fig.(C18). Obvious discrepancy should be

expected as the data sources are generated by different engines. Despite this, however, fair agreement is found

highlighting the capability of the developed WAVE engine model.

Finally, noise measurement data obtained from position C is given in Fig.(C19) and Fig.(C20). This datashows no clear indication of any significant resonances that are absent from unsilenced data.

4. ConclusionThe base silencer design as proposed by Blair is here shown to have high potential as an effective silencer.

With further design refinement and testing, the efficiency of this concept may also be appreciated. Theexpansion chamber has shown enhanced sensitivity to the non-idealities present within an exhaust silencing

application. As a result, more significant correlation is found between the volume of this component than with

its length as per acoustic theory. However, acoustic theory was demonstrated to generate predictions of

absorption silencer performance with relatively high accuracy.

This experiment established the priority for silencer design as to control the mass flow rate from the exhaust

outlet. Acoustic theory was determined to have limited effectiveness in predicting silencer performance. Theusefulness of acoustic theory within silencer concept development is recognised.

Fair agreement was found with WAVE data obtained despite variation in engine displacement used togenerate the data sets. This agreement was represented mainly by similar trending. As per theory detailed in the

literature review, the value of a one-dimensional software for concept development is shown to be founded in its

ability to quickly describe unsteady gas flow throughout an engine.

C. WAVE Investigation – Exhaust Design Parameters for Engine Scavenging

A literature review identified that exhaust performance considerations are resultant of the nature of the gasexchange process in a four stroke engine. The following investigation was undertaken using WAVE to

demonstrate the potential of the identified exhaust tuning strategies. This included the variation in ‗wave tuning‘

and ‗inertial scavenging‘ with exhaust pipe geometry. Specifically, this study was concerned with identifying

the extent of variation in engine performance possibly accomplished via the design of an exhaust pipe for a

single cylinder engine assuming a constant intake length and valve timing. The purpose of this investigation is

therefore to inform the ADFA FSAE team of the methods commonly incorporated within the design of an

exhaust system, that aim to enhance or merely shape an engine‘s performance characteristic. The followingfindings should therefore act as a tool to be used in conjunction with many other powertrain parameters to

obtain a desired engine performance target.

1. Exhaust Wave TuningAs stated by Professor Blair of the University of Belfast ―the tuned exhaust pipe harnesses the pressure wave

motion of the exhaust process to extract a greater mass of exhaust gas from the cylinder during the exhaust

stroke and initiate the induction process during the valve overlap period.‖ This scavenging effect is possi ble if a

pressure wave originating from the exhaust valve travels over a tuned length such that it is reflected back to the

valve face as a rarefaction wave in time to assist the gas exchange process during valve overlap. The coincident

phasing of valve overlap and the arrival of pressure waves, seen in Fig.(D1), is dependent upon the length over

which the waves travel known as the ‗tuned length‘. Resultant valve mass flows and pressure differentials are

provided in Fig.(D2) and Fig.(D3).

As per Fig.(D4), Fig.(D5) and Fig.(D6), this scavenging effect is characteristic of a certain engine speedwhere the correct phasing occurs. These figures shows the variation of residual gas fraction with tuned length

and the resulting effect upon torque and power output of the engine as a result of an increased delivery of

combustibles. Here relatively small variation in residual gas fraction is seen to have a marked effect upon

torque. Furthermore, when this effect is achieved at high engine speeds a highly significant influence is

exercised over the shape of the power curve. This data therefore demonstrates the significance of the exhaust

wave tuning and resultant scavenging effect achieved. Finally, Fig.(D7) is provided to demonstrate the possibleeffect of interaction between intake and exhaust tuning. As seen, the scavenging ratio (a measure of the quantity

of fresh combustible mixture ingested to the engine) is greater than unity at 3000rpm for a 1450 mm tuned

length. This suggests that at this point wave tuning is interacting with other tuning effects to achieve a greater

scavenge than could be achieved alone. Such interactions are also important when defining the design target for

exhaust tuned length such that these interactions are used to their full potential.

To summarise, a contour plot of residual gas fraction with tuned length and engine speed is provide provided

in Fig.(D8). The region representing the most effective scavenging effect is seen in blue. An inverse relationshipis seen to exist between exhaust length and tuned engine speed. Since pressure waves travel within the exhaust


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system at the acoustic speed which is a function of temperature, the hyperbolic nature of this relationship is

therefore present due to the asymptotic nature of heat transfer from the exhaust pipe.

Stepped Pipe TuningAs a part of an investigation into exhaust

wave tuning techniques, the industry practice

of utilising stepped pipes was considered. Aschematic of a stepped exhaust pipe is shown

in Fig.(20). A stepped pipe offers extra degrees

of freedom in wave tuning practices as there

exists more discontinuities able to create

rarefaction wave reflections. Furthermore thischaracteristic can also lead to varied heat

transfer properties. To illustrate, Fig.(D9) is

provided. This plot is the product of a 1500mm

pipe with a fixed expansion at 500mm and

another expansion whose location in the pipe varies between 510mm to 1490 mm. Wave action from the pipe

end at a tuned length of 1500mm is tuning at 4000 rpm and secondly at 7500rpm shown by two regions of

relatively low residual gas fraction. In addition, wave action from the first expansion occurs at 500mm

enhancing the region of low residual at 7500 rpm. Of note, data shows a noticeable change within each of theseregions as a result of the location of the second expansion. To illustrate Fig.(D10) is provided which shows

variation in exhaust gas temperature (in blue) as well as acoustic velocity (in green) with position for two

stepped pipes with location of the steps indicated. Fig.(D10) specifies that these shifts in the tuning behaviour of

the pipe can be attributed to the effect stepping behaviour of the pipe has on heat transfer, average gas

temperature and the average acoustic speed which are indicated to vary slightly. In addition to this effect,

Fig.(D11) and Fig.(D12) show that for these same two stepped pipe designs, the intermediate expansions are infact also able to reflect rarefaction waves for the purpose of scavenging, despite being of a lesser magnitude.

Specifically, these figures show that for two different stepped pipes (parameters indicated in plot caption), a

variation in the arrival of the first smaller wave is recorded, giving rise to corresponding change in the recorded

valve mass flow rate purely as a consequence of the unique positioning of the intermediate discontinuity.

DiffusersThe most effective exhaust component design for wave

scavenging is that of the diffuser as seen in the schematic provided in Fig.(21). This type of exhaust component is

known to be able to tune over a wider range of engine

speeds offering superior scavenging and engine

performance. For comparison however, a generally

accepted rule-of-thumb states that a third of the length of

the diffuser is used in the calculation of the total effectivetuned length.

To demonstrate the action of the diffuser, Fig.(D13)

shows residual quantity achieved for a diffuser 400mm

long and with a taper angle of 6.34 degrees attached to a

variable length of pipe. What is instantly noticeable is the greater dominance of the scavenging region comparedto that of the straight pipe. Again Fig.(D14), shows residual quantity for a larger diffuser of 600mm in length

with a taper angle of 6.65 deg. Once more this shows a vastly greater scavenging ability than that shown by a

single straight pipe. Finally, Fig.(D15), is showing residual gas fraction for a diffuser larger still, of 900mm in

length with a taper angle of 6.65 degrees, however no significant benefit is seen to be gained by the extra length.

Through the course of this study a diffuser was manufactured in order to obtain experimental validation of

this theory and to promote the diffuser as an innovative technique to enhance overall engine power. However, asa result of the continuing inoperability of the team‘s engine testing rig this validation could not be undertaken.

2. Inertial ScavengingInertial scavenging describes a scavenging effect that is enabled by the inertia of a high velocity column of

exhaust gas escaping from the cylinder. As seen in the simplified schematic in Fig.(22), the interaction between

this column of gas and the gas exhange process during valve overlap may see a build up of pressure energy at

the exhaust valve that acts to assist in engine breathing. This mechanism functions under the principle that afixed volume flow rate is achieved at a certain engine speed and for a fixed volume flow rate through a pipe, gas

Figure 20. Simplified schematic of stepped pipe

Figure 21. Simplified schematic of diffuser


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1. Adequate insertion loss — in accordance with FSAE regulations the noise measurement taken near the

exhaust must not exceed 110dB;

2. Back pressure minimal — to maximize vehicle competitiveness throughout the competition the

implemented silencer should introduce minimal engine power losses by way of back pressure and be

integrated within an exhaust system that provides a desired torque and power output characteristic.

3. Size — the silencer must be able to be easily integrated within the vehicle;

4.

Cost — low cost desirable;5. Durability — high durability desirable.

B. Silencing StrategyIn accordance with findings of the literature review, the formulation of a preliminary design proposal was to

be carried out relative to the identified ‗noise problem‘. In this case ‗noise problem‘ was established from anestimate of the SPL spectra, measured under FSAE conditions, emanating from a 1000mm stepped pipe, which

was obtained from WAVE. This data is provided in Fig.(E1) in Annex I. It shows the predicted sound pressure

relative to a ceiling of 108dB. This target was calculated in accordance with theory detailed in Annex B. This

calculation recognises that the intake is the second most dominant source of noise upon a vehicle, as well as that

any other sources of noise varying from the maximum of more than 15dB offers a negligible addition to the total

SPL measurement. A pessimistic estimation of the intake noise is taken as 100dB, and therefore with theaddition of a 108dB exhaust noise contribution, a total measurement of 109dB would be achieved which is in

accordance with noise level design requirement.Predicted noise spectra for the WR450 at its test engine speed of 7000rpm shows that the noise measured

from the outlet of the stepped pipe constitutes a number of significant contributions from a range of frequencies

corresponding to flow noise as well as to the engine firing rate. Therefore the proposed silencer is required tooffer broadband attenuation of up to 20 dB to satisfy noise the level design requirement.

C. Design Investigation and DefinitionThe definition of a final design proposal comprised a process of systematic analysis and selection, of silencer

components. Each of the significant silencer components including the absorption silencer, expansion chamber

and the choke were investigated individually such that the final silencer assemblage would represent the option

best able to satisfy the design requirements.

In line with findings of experiments conducted, the analysis of these individual components was conducted

relative to their governing principles. Since good correlation was found between acoustic theory and achieved

attenuation for an absorption silencer, the analysis of this component was based upon the predicted acoustictransmission loss, which ignores the non-idealities of an exhaust silencing application. In contrast, the analysis

of the choke and expansion chamber volume was conducted upon the developed WAVE engine model such that

variation in outlet mass flow rate and consequent attenuation could be appreciated.

In accordance with findings of the WAVE enabled investigation into exhaust scavenging mechanisms, the

final silencer design proposal will be implemented upon a 1000mm stepped pipe as this design offers the highest

level of inertial scavenging for a fixed tuned length. Tuning of this integrated system will then be undertaken in

order to demonstrate the process of exhaust tuning relative to a performance target. In this case, the performance

target will be the unsilenced performance trend such that the attainment of this target will help demonstrate the

efficiency of the silencer design proposed by way of minimal back pressure.

1. Absorption Silencer Component

Acoustic theory of an absorption silencer was seen to hold true during experiments. Therefore an analysis of

transmission loss is recognised to represent a reasonable prediction of achieved performance if not merelyrelative performance. Fig.(E2) shows variation in attenuation with increasing diameter assuming the current

51mm diameter perforated pipe is used and holding the length of the component and packing density as

constant. Attenuation is seen to increase linearly with diameter. This result agrees with acoustic theory such that

the increased depth of sound absorbing material increases the viscous dissipation of sound energy via interaction

with particle oscillations. Fig.(E3) shows variation in attenuation with length of the absorption silencer

component. Attenuation is seen to asymptote such that large increases in length are required for only a fractionalincrease in attenuation. Acknowledging that the absorption silencer component will constitute the majority of

the weight of the overall silencer, the attained data was analysed in terms of attenuation achieved per unit mass.

Fig.(E4) shows that for a specified increase in attenuation, an increase in diameter represents a more efficient

means than increasing the length. As a result, the proposed silencer design should incorporate as large a

diameter as possible that still remains conformant to the size constraints specified in the design requirements.

Fig.(E5) shows variation in attenuation with sound absorbing material density. This plot shows the

convergence of attenuation to an asymptote. Consequently, this data suggests that a density greater than 150g/Loffers a negligible increase in attenuation for the extra weight. Finally, Fig.(E6) shows variation in attenuation


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with perforated area of the inner pipe used. Negligible variation is illustrated therefore the selection of

perforated tube used will be constrained the requirement for durability of the silencer, as too high a perforated

area will allow violent exhaust gas flow to degrade or remove sound absorbing material.

2. Choke Diameter

The choke was proven to practice significant control over emitted noise within experiments conducted which

is demonstrated further in Fig.(E7). This shows the ability of the choke to scale the emitted noise levels via practicing direct control over the outlet mass flow rate. Fig.(E8) is also provided in order to emphasise the

silencing action of the choke. This data is generated within the WAVE Transmission Loss Workbench, which

conducts a comparison of sound power at the inlet and outlet of a silencer under nil mean flow conditions. As a

result, the choking of an expansion chamber element is seen to have minimal effect in the absence of mean flow.

In accordance with the design requirement for engine power losses, Fig.(E9) and Fig.(E10) are provided.These plots show that for a choke diameter greater than 26mm a minimal effect upon engine scavenging,

measured in total residual quantity, and brake torque is predicted. Meanwhile a choke of 26mm also achieves an

increase in overall attenuation of up to 5dB making this a highly efficient component for silencing.

3. Expansion ChamberHaving acknowledged the trends in attenuation achieved via absorpion parameters as well as choke, the

volume of the expansion chamber was varied to gain a similar appreciation. Variation in predicted outlet mass

flow rate with expansion chamber length with fixed diameter is provided in Fig.(E11). The maximum mass flowrate recorded is seen to decrease consistently until a length of 150 mm is reached. A negligible change in peakmass flow rate is found beyond this length and instead a phase shift is noted. Similar behaviour is seen in

Fig.(E12) which illustrates the resultant SPL measurement taken under SAE noise test conditions. A consistent

reduction in SPL is recorded up to a length of 150mm at which point the negligible change in the magnitude of

the mass flow rate results in no further reduction in SPL.

B. Design ProposalThe conducted parameter study was used to inform the formulation of a final design concept. Data generated

justified the selection of silencer parameters such that concepts could be verified using the developed enginemodel. So as to minimise size and weight of the silencer, a conservative choke diameter of 30mm was selected

to exercise meaningful control over the outlet mass flow rate without deliberately increasing engine pumping

losses. The diameter of the silencer was identified to represent the most significant factor per unit mass, for

increasing absorption attenuation and silencer volume. As a result, in order to minimise weight of the silencerand with a consideration of space constraints relevant to the current vehicle‘s side pod arrangement, a dia meter

of 175mm was selected. Again to minimise weight, a minimal length was sought for the absorption silencer

component. Data suggested that lengths beyond 300 mm were subject to diminishing returns in terms of

attenuation and so this was selected as the final absorption length. By doing so the requirement for extra

chamber volume would also be minimised. The packing density was selected to be 150g/L as acquired data

suggested diminishing returns beyond this value. Finally, expansion chamber length was increased until a

satisfactorily low outlet mass flow rate was obtained giving a prediction of under 108 dB as per design goal.

The ability of WAVE to accurately calculate the instantaneous mass flow rate from the exhaust outlet, being

the definitive measure of silencer performance, underpins confidence within this design proposal. The

characteristics of the proposed silencer design are provided in Table (4).

Table 4. Design Proposal

Component Parameters Attenuation Characteristics

Resonator Chamber Length: 150 mm

Diameter: 175 mm

Dampen outlet mass flow rate thereby

providing broadband attenuation.

Absorption

Component

Length: 300 mm

Diameter: 175mm

Packing Density: 150 g/L

Target high frequency flow noise

Choke Diameter: 30 mm Scale outlet mass flow rate thereby

providing broadband attenuation.

Silencer-Cylinder

Volume Ratio

24


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Predicted SPL measurement for the proposed silencer configuration is provided in Fig.(I13). Unsilenced

SPL data is also provided on this figure to demonstrate the broadband attenuation characteristics of the design.

Data sets are also seen to diverge slightly at higher engine speeds where the increasing dominance of flow noise

is experienced. This deviation attests to the effectiveness of the absorption silencer component at attenuating

this high frequency flow noise. Performance variation as a result of the implementation of the proposed silencer

design is shown in Fig.(I14). As per design requirements, the efficiency of the proposed silencer design is

defined as its ability to achieve a required level of attenuation with minimal effect upon performance. As such, performance with and without the silencer fitted to a 1000mm stepped pipe is given. The addition of the silencer

was seen to cause a significant change in the mean exhaust temperature over the tuned length of the exhaustsystem leading to a reduction in the effective tuned length of the system. The resultant change in wave tuning is

illustrated in Fig.(I15) which shows a time plot of the inward travelling waves at the exhaust valve. The increase

in exhaust temperature causes the inward wave to arrive earlier than the wave generated by an unsilenced

exhaust system causing a reduction in scavenging effectiveness at this engine speed. In order to investigate a

strategy to retrieve peak torque and power, contours provided in Fig.(I16), Fig.(I17) and Fig.(I18) of residual

quantity, brake torque and power with exhaust length in addition to a 500mm header pipe, were generated.These plots show that an increase of exhaust length by 200mm is able to achieve similar performance obtained

from the unsilenced system without becoming subject to unsteady behaviour seen to exist for longer lengths.

Therefore with an extension of the exhaust length to 1200mm, Figs.(I14), Fig.(I19) and Fig.(I20) show that a

negligible reduction in overall power and torque is achieved for the silenced engine. This data also shows a

number of other performance variations resultant of the addition of the silencer. As per Fig.(I15), whilst theextended exhaust pipe has attained the correct phasing of pressure waves with valve overlap, the intensity of thiswave is seen to be degraded by the extra distance of pipe travelled. The other significant consequence of silencer

addition, has been a marked increase in scavenging and torque at high engine speeds. With reference to

Fig.(I19), Fig.(I20) and Fig.(I21), a large amount of inertial scavenging has been achieved beyond 7000rpm

leading to a significant reduction in pumping losses as well as residual gas fraction.

In conclusion, the exhaust system design process has demonstrated the effectiveness and efficiency of the

proposed silencer as an integrated component within a performance exhaust system. Specifically, the proposed

silencer design is demonstrated to satisfy design requirements in terms of noise attenuation targets and back

pressure. Furthermore this process has demonstrated the tuning effect of silencer addition to an exhaust system

and described a strategy to attain a required engine performance characteristic of the silenced exhaust system.

By no means does this design process identify this particular tuning strategy as the best but instead merely sets

out to demonstrate the implementation of exhaust tuning theories discussed.

D. Vehicle Integration of Exhaust System

The process of vehicle integration of an exhaust system provides numerous constraints to the design

parameters of an exhaust system. In particular, due to limited number of mounting positions of a silencer, tuned

length could be understood to be constrained to discrete values. However, this would merely require the

innovative combination of exhaust techniques discussed such that a desired tuning effect is achieved. For

example, if silencer integration dictates that the exhaust length must differ from that which directly offers the

desired wave tuning effect a number of alternate strategies are available. These may include:

the manipulation of pipe diameter so as to target this performance characteristic with inertial

scavenging,

the use of exhaust pipe insulating wraps or coatings to tailor the mean exhaust gas temperature and

therefore the effective tuned length, or

the implementation of an appropriately sized discontinuity to define the tuned length prior to the

silencer, the use of stepped pipes to target a performance characteristic with a combination of partial wave

reflections and inertial scavenging.

Whilst considering concerns of exhaust integration, it is important to explicitly state that by minimising pipe

bends within the exhaust will act to minimise overall back pressure and engine pumping losses. Consequently,

supplementary reports provided in Appendix A and Appendix B, demonstrates the performance effect of pipe

bends on heat transfer and engine performance as well as the optimisation of pipe geometry for pressure lossusing ANSYS Fluent and the optimisation routine NSGA2.

Finally, it is important to consider the achieved exhaust flow Mach number during the selection of exhaust

pipe parameters. This is in accordance with findings of Wiemeler, Jauer and Brand [14] who direct correlation

between Mach number and flow noise generation efficiency.


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V.

Limitations of Ricardo WAVE As a one-dimensional software WAVE is limited in the accuracy with which it can represent any part of an

engine which inherently consists of three dimensional flow behaviour to some extent. As a result spurious data

can sometimes be obtained from simulations conducted, seen within some of the provided plots as seemingly

random spikes. However, the majority of data is conformant to a definite trend which is for the most part where

the value in this software is derived.

Of particular note, by utilising a one-dimensional simplification the model relies highly upon the quality ofits inputs rather than directly resolving aspects of engine operation. As a result, there exists a central

requirement for the validation of the model whilst it is under development. For example, a flow noise efficiency

factor is user defined input as a part of the process of acoustic acquisitions within the post processing programWAVE-post. Standard values for this input, as with many inputs, were used which may need to be verified

within the process of model development.

VI. ConclusionThe aim of this work is to underpin all future development by the ADFA FSAE team, of a high performance

exhaust system for a single cylinder engine. In accordance with findings presented, the presented silencer design

will be able to provide the required noise management capability without prejudice to engine performance.

Performance exhaust tuning techniques have been discussed and demonstrated such that this design should be

able assist in shaping the torque and power output characteristic of the engine for the benefit of vehicle

competitiveness. Moreover, the importance of conducting of exhaust tuning as a part of an integrated engine performance tuning process has been identified. The presented design and analysis methodology has provided a

meaningful demonstration of the silencer operation and design. Finally this demonstration has culminated in the

proposal of an efficient silencer design.

VII. Recommendations and Future WorkThe complexity of operation and analysis of the exhaust provides a wide scope. This research has attempted

to provide the basic foundations of exhaust and silencer design and analysis however in doing so, depth of

research has been sacrificed for breadth. As a result continued work should hope to explore more specialisedtechniques of exhaust design and analysis to extend upon the basic concepts presented herein. Some topics of

interest would stem from tuning interactions assumed constant within this study. For instance, the simultaneous

tuning of the exhaust and intake as well as valve timing is documented as a significant method of engine

performance optimisation. In addition, tuning methods identified herein suggests potential for innovative

integrated designs that combine the use of inertial scavenging and wave tuning in a synergistic manner that may

be worthy of investigation. For a single cylinder engine there is limited further work that could be undertaken inthe way of exhaust tuning via pipe design. However a hypothesis was formed during the course of this study that

specialised exhaust components such as Helmholtz chambers have been used within industry to not only provide

a means of targeted silencing but also to enhance wave tuning effects via wave interaction. The implementation

of such a component in this way would be expected to enhance performance as well as to justify a lighter

silencer and therefore it seems worthy to recommend an investigation into the feasibility of the idea.

Furthermore, commercial products such as those in

Fig.(39) incorporate components that are unexplained by

this study but may offer extra performance benefit and

therefore may represent another opportunity for further

work.

It was found during the course of this research, that arange of studies into silencer design used other forms of

silencer concepts as the basis of a design optimisation. In

particular a text by Munjal entitled ‗Acoustics of ducts

and mufflers with application to exhaust and ventilation

system design‘ was used by a number of studies who

were concerned with the implementation of reactivesilencers. Ideally, future work conducted by the ADFA

FSAE team would be able to identify whether a reactive

silencer concept would be able to surpass the current

proposed design in terms of attenuation, back pressure

and weight.

Finally, much literature is available as to the implementation of coupled 1D/3D analyses within this topic

area. WAVE openly admits to enhanced inaccuracy when dealing with complex components and thedevelopment of this capability within any aspect of exhaust design would represent a powerful design tool.

Figure39. Commercial exhaust system with novel

components labelled ‘Powerbomb’ and ‘Megabomb’


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AcknowledgementsThe author would like to gratefully thank a variety of important individuals that helped throughout the course of

this study. Thanks goes to thesis supervisor, Dr Warren Smith for providing much needed guidance during the

course of what always seemed to be a grossly under defined problem. To Mr Alan Fien, for your willingness to

offer your vast technical insight. To Mrs Marion Burgess for your patience and understanding despite the

tribulations of this project. Much thanks goes to SEIT workshop staff particularly Doug Collier and Marcos DeAlmeida for their assistance throughout the design and manufacturing process. Thanks to members of the FSAE

team and fellow engineers whose support was invaluable and who at times managed to make this project an

enjoyable process. Finally, to my girlfriend who showed amazing patience over the course of a very long year of

work as well as offering much needed support over the course of this degree. Thanks to my family who are well

deserving of official recognition of all their support over the many years.


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References

[1] SAE, "FSAE Inspection Sheet," ed: SAE, 2011.

[2] SAE-Australiasia. (2011, 07 May). Competition Overview [Internet Web Page]. Available:http://www.saea.com.au/formula-sae-a/competition-overview

[3] G. P. Blair, "Design and Simulation of Engines: A Centruy of Progress," SAE International1999.

[4] G. P. Blair. (1999). Design and Simulationof Four-Stroke Engines.

[5] J. Robinson. (1994). Motorcycle tuning

[6] Smith and Morrison, Scientific Design of Exhaust & Intake Systems, 2009.

[7] A. G. Bell. (2001). Four Stroke Performance Tuning .

[8] D. Winterbone and. R. Pearson, Design Techniques for Engine Manifolds - Wave action methods for IC engines.

London and Bury St Edmunds, UK: Professional Engineering Publlishing Limited, 1999.

[9] M. Ashe, G.Blair, G.Chatfield, D.Mackey, "Exhaust Tuning on Four-Stroke Engine: Experimentation andSimulation," The Queen's Univeristy of Belfast; OPTIMUM Power Technology2001.

[10] Yunquig Li, Jincheng Wang and Peng He, "Study on the exhaust system parameters of a small gasoline engine,"Beihang University 2008.

[11] G. Sammut and. A. Alkidas, "Relative Contributions of Intake and Exhaust - Tuning on SI Engine Breathing - AComputational Study," Oakland University 2007.

[12] J.D. Irwin and E.R. Graf, Industrial Noise and Vibration Control . New Jersey: Prentice-Hall 1979.

[13] J. Pang et al. "Flow Excited Noise Analysis of Exhaust," Ford Motor Company; Gates Coporation2005.

[14] A. Jauer, J. Brand and D. Wiemeler, "Flow Noise Level Prediction Methods of Exhaust System Tailpipe Noise,"Tenneco, Germany 2008.

[15] S. W. Coates, "The Prediction of Exhaust Noise Characteristics of Internal Combustion Engines ", The Queen'sUniversity of Belfast, 1974.

[16] Muthukumar Yadav, Kiran, Tandon and Raju, "Optimized Design of Silencer - An Integrated Approach," TheAutomotive Research Association of India, Pune, India 2007.

[17] Silvestri, Morel, Goerg and Jebasinski, "Modeling of Engine Exhaust Acoustics," Gamma Technologies, BMWAG, J. Eberspacher, GmbH & Co.1999.

[18] Wrtz and Mazzoni, "Application of WAVE in Motorcylce Prototyping," Ducati Motor S.p.A,, Bologna, Italy.

[19] Honda et al, "Honda, Kodama, Wakabayashi,Nakayama, Morimoto and Ueda," Kokushikan University, Japan

2005.

[20] Rose, Marshland and Law, "Optimisation of the Gas-Exchange System of Combustion Engines by GeneticAlgorithm," in 4th International Conference on Autonomous Robots and Agents, Wellington, New Zealand, 2009.

[21] Massey, Williamson and Chuter, "Modelling Exhaust Systems Using One-Dimensional Methods," Flowmaster

(UK) Ltd. ; ArvinMeritor 2002.

[22] Montenegro and Onorati, "A Coupled 1D-multiD Nonlinear Simulation of I.C. Engine Silencers with Perforatesand Sound Absorbing Material," Politecnico di Milano 2009.

http://www.saea.com.au/formula-sae-a/competition-overviewhttp://www.saea.com.au/formula-sae-a/competition-overviewhttp://www.saea.com.au/formula-sae-a/competition-overview


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Annex A

Summary of Acoustic Theory for Automotive Silencers

23


1. Absorptive/Side-Resonant Silencer

The absorptive and side-resonant silencers operate under the principles established for the use of perforated

metals in acoustic treatments. These principals differentiate between the design parameters of the perforated

materials utilised within the design, which in turn specify if the action of the silencer to be through the resonant

action of the perforated material or via viscous dissipation within sound absorbing material placed behind.

Parameters such as the Transparency Index [33] in Eq.(A1) or otherwise Blair‘s empirical relations [4] inEq.(A2) and Eq.(A3) may be used to distinguish between these types of silencer which are concerned with

perforation pattern of the material used. However, since the transparency index measure is only capable of

distinguishing between these variations of silencer beyond 10 kHz, being a frequency fairly well beyond the

significant spectrum present in an exhaust, it is not predominantly used for this purpose within this application.

Aonversely, Blair‘s relations were developed specifically for automotive silencers and are therefore much morerelevant.

(A1)

(A2)

(A3)

An absorption silencer utilises perforated material that shows negligible preference to the transmission of

any region of the frequency spectrum through the material and into the side chamber, otherwise known as the‗transparency approach‘ [33]. By permitting acoustic wave energy within the side chamber it is made to reflect

from the outer shell and constructively interfere with sound waves entering the chamber. This interference then

establishes standing waves characterised by increased amplitude of particle oscillation, within the region

between the perforated pipe and the silencer housing. As seen in Fig.(A1), sound absorbing material fills this

region where wave superposition is predicted to occur.

Consequently, viscous dissipation of sound is achievedas particle kinetic energy is converted to thermal energy

via interaction with the sound absorbing material. As per

Fig.(A2), correlation is shown between the radial

distance between the silencer shell and perforated pipeand the largest wavelength capable of superposition

within the thickness of the sound absorbing material.

Consequently, this figure shows that for an increase in

thickness of the sound absorbing material, a significant

increase in attenuation is achieved for noise of longer

wavelength.

The Transparency Index can however be informative

through the evaluation of the Access Factor, which

represents a measure of the perforated metal‘s ability to

obstruct the entry of acoustic waves and has the effect of

scaling the absorption factor of the silencer [33] (The

absorption factor is defined as the transmission loss

expressed as a fraction of the incident sound energy).Figure A1. Schematic of absorption silencer


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This is therefore a measure of the degradation in the

ability of the silencer to attenuate acoustic energy as a

result of perforation parameters. As per Fig.(A3),

perforated metals with a transparency index of less than

6500 begin to have a noticeable effect on the attenuation

achieved at frequencies below 2000 Hz (frequencies up to

this point are considered significant sources of vehiclenoise).

Literature provides additional design considerationrelevant to the manufacture and implementation of an

absorption silencer are proposed which include the

following:

It is recommended that the perforated pipe be

manufactured with stabbed holes rather than blind holes as

seen in Fig.(A4). This has the effect of increasing thedischarge coefficient for flow into the side chamber from

the central pipe and reducing turbulent eddies produced by

gas flowing over the sharp edges of blind holes. [4]

While taking the transparency approach, it is also

important to consider that an excessive perforated area ofa tube may enable violent exhaust flow through the

silencer to degrade the sound absorbing material and even

attempt to rip it from the side chamber. It is therefore

recommended to use perforates with holes of diameter

between 2 and 3.5mm. [4]. (The addition of a layer of

stainless steel wool is also recommended by industry

SMEs).

An absorptive silencer is most effective at

attenuating high frequency noise. Therefore this

component is recommended to be one of the last within

the exhaust system such that turbulent flow preceding the

absorption silencer has limited opportunity to build up this

high frequency component. The choice of sound absorbing material as well

as the density of the packing will lead to variation in the

achieved transmission loss as per Fig.(A5) [34]. This data

reiterates that the maximum wavelength absorbed

increases with thickness of sound absorbing material. In

addition, an increase in the material density from 100g/L

to 200g/L is seen to accompany a reduction in absorption

achieved. This highlights that a density too high will

restrict the entry of acoustic waves into the side chamber,

while a density too low is also acknowledged to become

less effective in achieving viscous dissipation of acoustic

energy.

In contrast a side-resonant silencer, whilst sharing a

similar form as the absorption silencer, does not utilise

packing material and instead provides attenuation over arelatively narrow band of frequencies as per Fig.(A6). This

is achieved through the resonance of the side cavity at its

natural frequency. The design of this type of silencing

component is governed by Eq.(A4) to Eq.(A7). The design

variables, seen in Fig.(A7), are shown to dictate theresonant frequency of the component as well as the

attenuation achieved at the resonant frequency [4].

Figure A2. Variation in attenuation with

frequency for thickness of absorption silencer

Figure A3. Access Factor vs frequency and

Transmission Index of perforated sheet metal

Figure A4. Schematic of perforated pipe

showing standard blind holes and stabbed

holes


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√ (A4)

(A5)

(A6)

√

(A7)

hole conductivity

Figure A5. Variation in attenuation with density of sound

absorbing material Figure A6. Attenuation predicted for a side-resonant

silencer

Figure A7. Design parameters of side-resonant silencer

element


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2. Diffusing Silencer/Expansion ChamberThe expansion chamber as seen in Fig.(A8), is designed to absorb acoustic wavelengths equivalent to the

natural frequency of the chamber. The transmission loss of an expansion chamber is given by Eq.(A8) to

Eq.(A11).

[ ] (A8)

(A9)

(A10)

(A11)As per Fig.(A9) [12], the attenuation is seen to be periodic with frequency. In addition, the maximum

attenuation is seen to increase with area ratio of the chamber (m). These relations have been experimentally

determined to have value up to a frequency of 1500 Hz. In practical terms, Fig.(A9) suggests that a longer

chamber will offer increased attenuation at lower frequency, hence why this component is employed withinsilencers to address low frequency noise corresponding to the engine firing rate.

Figure A8. Schematic of diffusing si

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Formula SAE Performance Exhaust Design