Formula SAE Performance Exhaust Design

An Investigation into Formula SAE Performance Exhaust Design and Analysis Anthony I. McLeod1

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 the design 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 was manufactured 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.

II.

Introduction .............................................................................................................................................. 3 A.

Motivation .................................................................................................................................. 3

B.

Project Aims ............................................................................................................................... 3

C.

Project Methodology .................................................................................................................. 3

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 1 Final Project Report 2011, SEIT, UNSW@ADFA

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

V.

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

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

VI. Conclusion .............................................................................................................................................. 19 VII. Recommendations and Future Work ...................................................................................................... 19 Acknowledgements ......................................................................................................................................... 20 References ....................................................................................................................................................... 21 APPENDICES Appendix A. Combined theoretical and experimental investigation into exhuast pipe geometry Appendix B. Exhaust pipe optimisation using NSAGA2 and ANSYS Fluent

Nomenclature ADFA CFD SPL Q D ̇

= = = = = =

Australian Defence Force Academy Computational Fluid Dynamics sound pressure level (dB) volume flow rate(m3/s) pipe diameter (m) flow velocity(m/s)

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A1 A2

I.

Introduction

A. Motivation The Formula SAE® competition constitutes a variety of rules and regulations that aim to challenge design teams 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 with gasoline 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 ensures compliancy. 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 WAVE to then undertake analysis to be validated experimentally, culminating in a final design proposal for a new high performing silencer. C. Project Methodology This 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 muffler was 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 of ‗tuned length‘, the effect of stepped 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 Review

Quality 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. 3 Final Project Report 2011, SEIT, UNSW@ADFA

A. Exhaust design for engine scavenging performance Performance 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 gases into 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 Figure 1. Valve timing events showing valve process to achieve this scavenge. overlap 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 is achieved 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 University 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 Figure 2. Poorly tuned engine ingests exhaust valve travel at the local acoustic velocity, over a exhaust gas into the cylinder during valce tuned length such that it is reflected back to the valve face as overlap a rarefaction wave, as seen in Fig.(3), in time to assist the gas exchange 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 a discontinuity 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 functions Figure 3. Reflection of rarefaction wave at under the principle that a fixed volume flow rate is achieved at a certain engine speed and for a fixed volume flow rate, exhaust pipe end which returns to the exhaust gas velocity varies inversely with pipe diameter. There then valve exists 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 drivability characteristics of the vehicle could diminish as a result of unpredictable power output behaviour.

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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 therefore the action of exhaust scavenging. The effects of this are well documented throughout literature but assumed

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

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 to increased 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 potential for 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 the source 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] as seen 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 can only be effective within spatial regions where this is the Figure 5. Noise level of intake and exhaust duct dominant noise source and sound pressure contributions of noise 5 Final Project Report 2011, SEIT, UNSW@ADFA

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 noise components 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 Figure 6. Frequency analysis of radiated vehicle noise Fig.(6). The EFR is the rate at which the showing noise corresponding to EFR frequencies exhaust 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 turbulent mean 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 proportional correlation 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 where flow noise is dominant, a fixed volume flow tranlsates to greater flow velocity for a smaller diameter and therefore increased noise emissions. . ̇

̇

̇

Figure 7. Sound pressure with engine speed showing increasing dominance of flow noise with engine speed

(1)

The conversion of flow power to sound power is Figure 8. Variation in flow noise with flow identified by Wiemeler, Jauer and Brand [14] to be velocity caused by pipe diameter 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 an icreased efficiency and increased flow noise sound pressure level (SPL). C. Exhaust Silencing In 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 components which include reactive, absorptive and resonator types. These components vary in the manner and efficiency 6 Final Project Report 2011, SEIT, UNSW@ADFA

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 rate from 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 order to 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.

Figure 9. Torque characteristics of muffler varieties determined by Blair

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 the outlet 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 10. Noise characteristics of muffler varieties determined by Blair

Figure 11. Schematic of tow-box silencer

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Figure 12. Exhaust outlet mass flow rate for silencer varieties

Figure 14. Noise attenuation performance of silencer components

Figure 13. BMEP with silencer component

Figure 15. Noise spectra at 7500 rpm with two-box silencer

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 suggest that 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 16. Noise attenuation of plenum and diffusing type silencers

Figure 17. Noise attenuation of side-resonant silencer

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indeed shows poor correlation. The inaccuracy of acoustic theory is understood to stem from its specific relevance to waves of infinitesimal amplitude which differ fundamentally from wave phenomenon experienced in exhaust flows being waves are of finite amplitude. Methods of accurately appreciating the true nature of exhaust waves must therefore employ an appreciation of the instantaneous mass flow rate emenating from the outlet which is proven by Coates [15] to an accurate approach. The design of an efficient and effective hybrid silencer is understood to be a highly complex task that may employ acoustic theories as merely the basis of an informed estimate for attenuation in order to commence a design process characterised primarily by experimental trial and error. Luckily, one-dimensional gas dynamic computational programs have the capability of providing accurate estimates of the time-varying mass flow rate from the exhuast outlet and may therefore be used to supplement the experimental silencer design process. D. Design and Modelling of Exhaust Systems Literature was also consulted in order to determine the capability and implementation of current numerical techniques for the design and modelling of exhaust systems. This review identified that as unsteady gas dynamics within an exhaust system are predominantly one–dimensional in nature, a great deal of research, design and optimization is carried out with one-dimensional engine simulation software. These analyses also achieve excellent agreement with experimental data providing the exhaust geometries of concern are free of excessive curvatures or complex silencing components characterized by strong three dimensional turbulence[4]. Additional advantages of employing this simplification include the ability to simultaneously calculate the effect of duct and silencer geometry upon engine performance and noise spectra, as well as to do so across a variety of designs within a reasonable timeframe. Consequently, one-dimensional codes have demonstrated an ability to enhance the efficiency of the design process [16]. Enhanced accuracy is accomplished with the use 3D CFD and coupled 1D/3D analyses however these generally have a far greater computational cost. E.

Conclusion This literature review summarises current theory and practices relating to exhaust system design, performance analysis and optimisation. Specifically, the mechanisms by which an exhaust design contributes to engine performance via scavenging were detailed. The nature of noise measurement, vehicle noise generation and attenuation were described. Acoustic theory of common silencing elements was detailed, and the deficiency of this theory in providing accurate predictions of silencer performance was addressed. The complexity of the silencer design process was established and studies demonstrating successful design methods have been described. Importantly, the outcomes of this literature review enable a more informed requirement definition for a new high performing exhaust system for the Yamaha WR450 engine. The aims of the exhaust design are summarised as an ability to assist in engine scavenging at the desired engine speeds and also to incorporate an efficient silencer design so as to minimise engine pumping work whilst achieving the specified noise target. To demonstrate the implementation of the presented theories within an exhaust design, this project undertook to varying degrees, the stages of concept development and preliminary design of the exhaust system. Activities under the concept development stage stemmed from the manufacture of a ‗two-box‘ hybrid silencer as well as the development of a WR450 engine simulation model using Ricardo WAVE. The conduct of this stage consisted of an experimental investigation of achieved noise attenuation with parameter variations in the manufactured design. Furthermore, the simultaneous conduct of this investigation with a physical engine as well as within WAVE provided a means for comment upon the effectiveness of the design, as well as the quality of the simulation model developed. This experimentation thereby performed both the roles of a parameter study as well as an experimental validation of the developed WAVE model. In addition, the concept development stage included the investigation of each of the scavenging mechanisms detailed, using the WAVE model. (These investigations were also intended to encompass an experimental component utilising an in-house developed engine testing rig, however this could not be accomplished due to technical difficulties). This analysis considered all possible design parameters within a reasonable range such that the ADFA FSAE team would be enabled to make an informed design selection as a part of an integrated engine tuning strategy. Finally, the preliminary design stage was conducted to provide the team with a refined design concept for a high performance silencer based upon that which was manufactured. The proposed prototype design was formulated using engine simulation results as well as data obtained from a silencer parameter analysis assisted by Ricardo WAVE transmission loss work bench. These processes thereby underpinned the aims of this project being to provide the ADFA FSAE team with the means of designing, analysing and implementing a high performance exhaust system.

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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 development and 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 obtained from 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 Study Having 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 implemented within 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 the validity 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 Design An 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 preferred in 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 accordance with 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

Inlet Diameter (D) Major Diameter Absorption Length Resonator Length Perforated Area Silencer- Cylinder Volume Ratio

Blair 46.6 mm

Experimental Design 51 mm

2.58D 8.58D 4.29D 19% 15

2.49D 9.0D Variable up to 5.88D 25% 14.5 - 19.5

Comments OTS component and recommended by industry SME OTS OTS Custom telescoping component OTS

2.

Orthogonal Experiment Design Acoustic 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 the manner 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 this data 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 simply used 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 1 2 3 4

Choke Diameter (mm) 20 20 30 30

Resonator Length (mm) 100 300 100 300

3.

Packing Density (g/L) 200 100 100 200

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 relative variation between designs. Measurements obtained of SPL with frequency are provided in Fig.(C4) and 11 Final Project Report 2011, SEIT, UNSW@ADFA

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 of the 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 recordings compared 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 increases in 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, yet experimental 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 by acoustic 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 would include 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 was also 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 is made 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 12 Final Project Report 2011, SEIT, UNSW@ADFA

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 data shows no clear indication of any significant resonances that are absent from unsilenced data. 4.

Conclusion The 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. The expansion 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. The usefulness of acoustic theory within silencer concept development is recognised. Fair agreement was found with WAVE data obtained despite variation in engine displacement used to generate 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 gas exchange 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 following findings 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 Tuning As 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 possible 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 speed where 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 possible effect 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 relationship is seen to exist between exhaust length and tuned engine speed. Since pressure waves travel within the exhaust 13 Final Project Report 2011, SEIT, UNSW@ADFA

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 Tuning As a part of an investigation into exhaust wave tuning techniques, the industry practice of utilising stepped pipes was considered. A schematic 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 this characteristic can also lead to varied heat transfer properties. To illustrate, Fig.(D9) is provided. This plot is the product of a 1500mm Figure 20. Simplified schematic of stepped pipe 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 these regions 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 in fact 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. Diffusers The 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 effective tuned length. To demonstrate the action of the diffuser, Fig.(D13) shows residual quantity achieved for a diffuser 400mm Figure 21. Simplified schematic of diffuser 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 compared to 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, as a result of the continuing inoperability of the team‘s engine testing rig this validation could not be undertaken. 2.

Inertial Scavenging Inertial 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 a fixed volume flow rate is achieved at a certain engine speed and for a fixed volume flow rate through a pipe, gas 14 Final Project Report 2011, SEIT, UNSW@ADFA

velocity varies inversely with pipe diameter. Consequently, there then exists 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. Under these circumstances an increase in engine power will be realised. To demonstrate Fig.(D16) is provided, which depicts the variation of scavenging effect measured in residual gas fraction with differing pipe diameters for a constant tuned Figure 22. Simplified schematic of inertial scavenging length of 1200mm and for a silenced exhaust system. A constant diameter 35 mm pipe is seen to scavenge well between 4000 and 5000 rpm. The addition of a 42mm step sees a shift in this scavenging effect to 5500rpm and similarly if this step is increased to 48mm in diameter, a further increase in the scavenged engine speed is observed. A pipe consisting of multiple steps and therefore including pipe diameters of 35mm, 42mm and 48mm achieves a scavenging effect which almost represents the addition of scavenging effects for both single step pipes. A consideration of the corrresponding pumping torque data in Fig.(D17), may be used to explain the transfer of energy from the piston, to pressure energy within the exhaust gas and then to its kinetic energy which eventually promises a scavenging effect capable of making proportionately more power than that lost in pumping. A comparison of trending in residual gas fraction and pumping torque for the constant 35mm pipe shows that within the range of engine speed of 4000rpm to 5000rpm, where most significant scavenging is achieved, pumping torque data is recorded as the greatest of all data sets. This increase in pumping torque acknowledges the transfer of energy from the piston to exhaust gas which in turn assists in cylinder scavenge. A similar phenomenon is observed for all other data sets such that a relative increase in pumping torque corresponds to the range of engine speeds where inertial scavenging is achieved. Valve mass flow rate data provided in Fig.(D18) and Fig.(D19), illuminates this theory further. For data at 4000rpm, the highest mass flow rates are attributed to the 35mm pipe. However, for data at 7500 rpm the stepped pipes are seen to record the highest valve mass flow rates. Finally, the subsequent torque and power curves are provided in Fig.(D20) and Fig.(D21). Contours of residual gas fraction for exhaust tuned length and pipe diameter were generated to explore this concept further. Three plots are provided in Fig.(D22), Fig.(D23) and Fig.(D24), one each for the engine speeds of 6000rpm, 7000rpm and 8000rpm. Data is summarised in Table (3). Table 3. Pipe parameters for optimal scavenging

Pipe Diameter (mm) Tuned Length (mm)

6000rpm 724 31

7000rpm 559 31

8000rpm 586 34

These plots draw attention to the minimum residaul gas fraction which is achieved with a specific combination of pipe length and diameter. Comparison of data at 6000rpm and 7000rpm suggests that wave tuning is the dominant mechanism at these engine speeds. This can be concluded as a constant diameter has been maintained, suggesting no significant variation in inertial scavenging achieved between these two engine speeds. Furthermore, a longer pipe is seen to contribute to scavenging at lower engine speed which is in line with previously discussed wave tuning concepts. However, a comparison of data at 7000rpm and 8000rpm shows that inertial scavenging is seen to gain significane again due to devaition from this logic. Now at 8000rpm, peak scavenging is achieved with a longer tuned length than at 7000rpm, which is contrary to wave tuning trends, as well as a larger diameter. This larger diameter can then be understood to underpin a required flow velocity such that inertial scavenging is maximised. The reciprocal exchange of energy from exhaust gas inertia to pressure energy which works to assist in pumping is seen in Fig.(D25) which shows variation in pumping torque with the length of constant 35mm diameter pipe. Here, at low engine speeds longer pipes offering back pressure record higher relative pumping losses. However, at higher engine speeds the maintenance of flow enery within longer pipes works to assist the piston during the exhaust stroke leading to a reduction in pumping losses relative to shorter pipes.

IV.

Part C – Preliminary Design and Design Proposal

A. Design Requirements As per the design of any engineering product, the design of a performance exhaust system must be conducted relative to specified requirements. Therefore before implementing knowledge gathered as a part of the concept development stage, customer requirements were explicitly stated. For a performance exhaust for the WR450 engine, they are stated in order: 15 Final Project Report 2011, SEIT, UNSW@ADFA

1. 2.

3. 4. 5.

Adequate insertion loss—in accordance with FSAE regulations the noise measurement taken near the exhaust must not exceed 110dB; 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. Size—the silencer must be able to be easily integrated within the vehicle; Cost—low cost desirable; Durability—high durability desirable.

B. Silencing Strategy In 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 an estimate 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 the addition 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 to offer broadband attenuation of up to 20 dB to satisfy noise the level design requirement. C. Design Investigation and Definition The 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 acoustic transmission 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 merely relative 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 fractional increase 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/L offers a negligible increase in attenuation for the extra weight. Finally, Fig.(E6) shows variation in attenuation 16 Final Project Report 2011, SEIT, UNSW@ADFA

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 Chamber Having 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 flow rate recorded is seen to decrease consistently until a length of 150 mm is reached. A negligible change in peak mass 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 Proposal The 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 engine model. 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 silencer and with a consideration of space constraints relevant to the current vehicle‘s side pod arrangement, a diameter 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 Resonator Chamber

Parameters Length: 150 mm Diameter: 175 mm

Attenuation Characteristics Dampen outlet mass flow rate thereby providing broadband attenuation.

Absorption Component

Length: 300 mm Diameter: 175mm Packing Density: 150 g/L Diameter: 30 mm

Target high frequency flow noise

Choke Silencer-Cylinder Volume Ratio

Scale outlet mass flow rate thereby providing broadband attenuation.

24

17 Final Project Report 2011, SEIT, UNSW@ADFA

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 exhaust system 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 the extended exhaust pipe has attained the correct phasing of pressure waves with valve overlap, the intensity of this wave 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 loss using 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.

18 Final Project Report 2011, SEIT, UNSW@ADFA

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 of its 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 program WAVE-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.

Conclusion

The 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 Work

The 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 specialised techniques 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 in the 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 a range 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 reactive silencers. 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 Figure39. Commercial exhaust system with novel proposed design in terms of attenuation, back pressure components labelled ‘Powerbomb’ and ‘Megabomb’ 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 the development of this capability within any aspect of exhaust design would represent a powerful design tool. 19 Final Project Report 2011, SEIT, UNSW@ADFA

Acknowledgements The 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 De Almeida 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.

20 Final Project Report 2011, SEIT, UNSW@ADFA

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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's University of Belfast, 1974.

[16]

Muthukumar Yadav, Kiran, Tandon and Raju, "Optimized Design of Silencer - An Integrated Approach," The Automotive Research Association of India, Pune, India 2007.

[17]

Silvestri, Morel, Goerg and Jebasinski, "Modeling of Engine Exhaust Acoustics," Gamma Technologies, BMW AG, 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 Genetic Algorithm," 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 Perforates and Sound Absorbing Material," Politecnico di Milano 2009.

21 Final Project Report 2011, SEIT, UNSW@ADFA

[23]

Montenegro and Onorati, "Modeling of Silencer for I.C. Engine Intake and Exhaust Systems by Means of an Integrated 1D-multiD Approach," Dipartimento di Energetica - Politecnico di Milano2008.

[24]

Zhang and Romzek, "Computational Fluid Dynamics Applications in Vehicel Exhaust System," Eberspaecher North America, Inc.2008.

[25]

J. Middleberg, T. Barber, S. Leong, E. Leonardi and K. Byrne, "Determining the Acoustic Performanec of a Simple Reactive Muffler using Computational Fluid Dynamics," presented at the The Eight Western Pacific Acoustics Conference, Melbourne, Australia, 2003.

[26]

Shah, Kuppili, Hatti and Thombare, "A Practical Approach towards Muffler Design, Developement and Prototype Validation," 2010.

[27]

S. Sen, "Predction of Flow and Acoustical Performance of an Automotive Exhaust System using 3D CFD," TATA Technologies Ltd.2011.

[28]

J. Caradonna, "Advanced Computational Aero-Acoustic Simulation of Complex Automotive Exhaust Systems," Faurecia Emissions Control Technologies2011.

[29]

Lu Lirong, Jin Xiaoxiong, Peng Wei and He Wei, "Application of Flow Field Simulation Technique to the Study of Exhaust Noise of Car," presented at the IEEE Vehicle Power and Propulsion Conference, Harbin, China, 2008.

[30]

W. Y. Fowlkes a. C. M. Creveling, Engineering Methods for Robust Product Design- Using Taguchi Methods in Technology and Product Developement. Reading, Massachusetts: Addison Wesley, 1954.

[31]

Bureau of Meteorology. (05 Nov 11). Sound Attenuation Calculator. Available: http://www.csgnetwork.com/atmossndabsorbcalc.html

[32]

Bureau of Meteorology. Humiditiy Calculator. Available: http://www.bom.gov.au/lam/humiditycalc.shtml

[33]

T. J. Schultz, "Acoustical Uses for Perforated Metals: Principles and Applications," I. P. Association, Ed., ed: Industrial Perforates Association Inc, 1986.

[34]

N. Huff, "Materials for Absorptive Silencer Systems," Owens Cornering Automotive Solutions2001.

22 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers

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] in Eq.(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 more relevant. (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 achieved as 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 pipe and 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 Figure A1. Schematic of absorption silencer expressed as a fraction of the incident sound energy). 23 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers

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

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 vehicle noise). Literature provides additional design consideration relevant 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 the discharge 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 of a 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 a relatively 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 the resonant frequency of the component as well as the attenuation achieved at the resonant frequency [4].

Figure A4. Schematic of perforated pipe showing standard blind holes and stabbed holes

24 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers

Figure A6. Attenuation predicted for a side-resonant silencer

Figure A5. Variation in attenuation with density of sound absorbing material

Figure A7. Design parameters of side-resonant silencer element

√

(A4) (A5)

( √

)

(A6)

⁄

(A7) hole conductivity

25 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers 2.

Diffusing Silencer/Expansion Chamber The 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 within silencers to address low frequency noise corresponding to the engine firing rate.

Figure A8. Schematic of diffusing silencer/ expansion chamber

Figure A9. Plot of theoretical attenuation of expansion chamber with design parameters

26 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers

3. Hershel-Quincke Tube As seen in Fig.(A10), the Hershel-Quincke tube is a device in which sound waves from a common source travel through two tubes of different lengths and recombine, producing reinforcement or cancellation of sound depending on the difference in path length. Unfortunately, no further description can be provided as this concept is poorly documented with regards to acoustic attenuation.

Figure A10. Example of a Hershel Quincke tube silencing component

4.

Helmholtz Resonance Silencer A Helmholtz chamber is seen in Fig.(A11) attached to the header pipe of an aftermarket Akrapovic exhaust system. It is an acoustic filter element that operates under the principles of a spring-mass system where the equivalent mass and spring force components are defined by the structural parameters of the chamber seen in Fig.(A12). The Helmholtz silencer is classified as a band-stop filter which offers attenuation at a specified frequency defined by Eq.(A12) and Eq.(A13). √

(A12) (A13)

The acoustic power transmission coefficient is then defined by Eq.(A14) and shown in Fig.(A13) for a specified chamber. (

(

⁄ ⁄

⁄

) )

Figure A11. Helmholtz chamber

(A14)

Figure A12. Schematic of Helmholtz chamber

27 Final Project Report 2011, SEIT, UNSW@ADFA

Annex A Summary of Acoustic Theory for Automotive Silencers

Figure A13. Acoustic power transmission coefficient for an example Helmholtz chamber

28 Final Project Report 2011, SEIT, UNSW@ADFA

Annex B WAVE model validation data -WR450 Power curves

WAVE Model Power Curve 40

35

30

Engine Power(hp)

25

20

15

10

5

0 1000

2000

3000

4000

5000 6000 RPM

7000

8000

9000

Figure B1. WR450 power curve generated with developed WAVE model

Figure B2. WR450 power curves provided by Cal Poly FSAE team

29 Final Project Report 2011, SEIT, UNSW@ADFA

10000

Annex C Muffler Parameter Study - Experimental Data

SPL Magnitude at Position A with Engine Speed for Parameter Variations and Commercial WR450 Muffler 120

SPL (dB)

115

Design 1

110

Design 2 Design 3 Design 4

105

WR450 Muffler

100 0

2000

4000

6000

8000

10000

Engine Speed (rpm) Figure C1

SPL Magnitude at Position B with Engine Speed for Parameter Variations, Unsilenced and Commercial WR450 Muffler 125 120

SPL (dB)

115 Design 1

110

Design 2 Design 3

105

Design 4 WR

100

Unsilenced

95 90 0

2000

4000

6000

8000

10000

Engine Speed (rpm) Figure C2

30 Final Project Report 2011, SEIT, UNSW@ADFA

Annex C Muffler Parameter Study - Experimental Data

Figure C3

31 Final Project Report 2011, SEIT, UNSW@ADFA

Annex C Muffler Parameter Study - Experimental Data

SPL with Frequency at 3000 rpm at position B 110 100 90 SPL (dB)

Design 1 80

Design 2 Design 3

70

Design 4 60

WR Muffler Unsilenced

50 40 0

200

400

600

800

1000

1200

1400

1600

1800

2000

Frequency (Hz) Figure F4

SPL with Frequency at 7000 rpm at position B 120 110 100 Design 1

SPL (dB)

90

Design 2

80

Design 3

70

Design 4 WR Muffler

60

Unsilenced

50 40 0

200

400

600

800

1000

1200

1400

1600

Frequency (Hz)

Figure F5

32 Final Project Report 2011, SEIT, UNSW@ADFA

1800

2000

Annex C Muffler Parameter Study - Experimental Data

Insertion Loss at 3000 rpm 40

SPL (dB)

30 20

Design 1

10

Design 2 Design 3

0 -10

0

100

200

300

-20

400

500

600

700

800

900

1000

Design 4

Frequency (Hz) Figure F6

Insertion Loss at 4000 rpm 50

SPL (dB)

40 30

Design 1

20

Design 2 Design 3

10

Design 4

0 -10

0

100

200

300

400

500

600

700

800

900

1000

Frequency (Hz) Figure F7

SPL (dB)

Insertion Loss at 7000 rpm 40 35 30 25 20 15 10 5 0

Design 1 Design 2 Design 3 Design 4

0

100

200

300

400

500

600

700

800

Frequency (Hz)

Figure F8

33 Final Project Report 2011, SEIT, UNSW@ADFA

900

1000

Annex C Muffler Parameter Study - Experimental Data

Variation in transmission loss with silencer design at 3000 rpm 90 Design 1 Design 2 Design 3 Design 4

80

70

Transmission Loss (dB)

60

50 X: 2620 Y: 40.65

40

30

20

10

0

0

500

1000

1500 Figure F7 Frequency (Hz)

2000

2500

3000

Figure F9

Variation in transmission loss with silencer design at 3000 rpm 45 Design 1 Design 2 Design 3 Design 4

40

Transmission Loss (dB)

35

30

25

20

15

10

5

0

100

200

300

400 Frequency (Hz)

500

Figure F10

34 Final Project Report 2011, SEIT, UNSW@ADFA

600

700

800

Annex C Muffler Parameter Study - Experimental Data

Figure C11

Figure C12

35 Final Project Report 2011, SEIT, UNSW@ADFA

Annex C Muffler Parameter Study - Experimental Data

Figure C13

Figure C14

36 Final Project Report 2011, SEIT, UNSW@ADFA

Annex C Muffler Parameter Study - Experimental Data Design 1: 4000 rpm

Design 3: 4000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

SPL (dB)

SPL (dB)

100

80

60

40

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

80

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

Design 2: 4000 rpm

1800

2000

100

SPL (dB)

SPL (dB)

1600

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

80

60

40

1400

Design 4: 4000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

800 1000 1200 Frequency (Hz)

80

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Figure F15

Design 1: 3000 rpm

Design 3: 3000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

SPL (dB)

SPL (dB)

100

80

80

60

40

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

Design 2: 3000 rpm

1800

2000

100

SPL (dB)

SPL (dB)

1600

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

80

60

40

1400

Design 4: 3000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

800 1000 1200 Frequency (Hz)

80

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

Figure F16

37 Final Project Report 2011, SEIT, UNSW@ADFA

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Annex C Muffler Parameter Study - Experimental Data

Design 1: 7000 rpm

Design 2: 7000 rpm

150

150 WAVE:SPL at muffler outlet Theoretical resonator chamber transmission loss Experimental data

WAVE:SPL at muffler outlet Theoretical resonator chamber transmission loss Experimental data

SPL (dB)

100

SPL (dB)

100

50

0

50

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

0

2000

0

200

400

600

Design 3: 7000 rpm

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Design 4: 7000 rpm

150

150 WAVE:SPL at muffler outlet Theoretical resonator chamber transmission loss Experimental data

WAVE:SPL at muffler outlet Theoretical resonator chamber transmission loss Experimental data

SPL (dB)

100

SPL (dB)

100

50

0

50

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

0

2000

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Figure F17

Design 1: 8000 rpm

Design 2: 8000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

SPL (dB)

SPL (dB)

100

80

80

60

40

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

Design 3: 8000 rpm

1800

2000

100

SPL (dB)

SPL (dB)

1600

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

100

80

60

40

1400

Design 4: 8000 rpm

WAVE:SPL at muffler outlet Mean SPL measurement Experimental data

120

800 1000 1200 Frequency (Hz)

80

60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

600

Figure F18

38 Final Project Report 2011, SEIT, UNSW@ADFA

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Annex C Muffler Parameter Study - Experimental Data

SPL with Frequency at 3000 rpm at position C 95 90 85 80

Series1

75

Series2

70

Series3 Series4

65

Series5

60 55 50 0

100

200

300

400

500

600

700

800

900

1000

Figure C19

SPL with Frequency at 7000 rpm at position C 100 95 90 85 Series1

80

Series2 75

Series3

70

Series4

65

Series5

60 55 50 0

100

200

300

400

500

600

700

800

Figure C20

39 Final Project Report 2011, SEIT, UNSW@ADFA

900

1000

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D1

Figure D2

40 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D3

Figure D4

41 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D5

Figure D6

42 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D7

Figure D8

43 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D9

Spatial Temperature and Speed of Sound in stepped pipe- Design 1

Spatial Temperature and Speed of Sound in stepped pipe- Design 2

1100

1100 Pipe step 1 Pipe step 2

1000

1000

900

900

Speed of Sound (m/s) / Temperature (K)

Speed of Sound (m/s) / Temperature (K)

Pipe step 1 Pipe step 2

800

700

600

500

400

800

700

600

500

0

0.5

1

1.5

400

0

0.5

Position (m)

1 Position (m)

Figure D10

44 Final Project Report 2011, SEIT, UNSW@ADFA

1.5

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D11

Figure D12

45 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D13

46 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D14

47 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D15

48 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D16

Figure D17

49 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D18

Figure D19

50 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D20

Figure D21

51 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D22

Figure D23

52 Final Project Report 2011, SEIT, UNSW@ADFA

Annex D WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure D24

Figure D25

53 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

SPL at muffler outlet: 3000 rpm

SPL at muffler outlet: 4000 rpm

140

140 SPL prediction at outlet max tolerated SPL: 108 dB

120

100

SPL (dB)

SPL (dB)

120

80 60

40

SPL prediction at outlet max tolerated SPL: 108 dB

100

80 60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

SPL at muffler outlet: 7000 rpm

1400

1600

1800

2000

140 SPL prediction at outlet max tolerated SPL: 108 dB

120

SPL prediction at outlet max tolerated SPL: 108 dB

120

100

SPL (dB)

SPL (dB)

800 1000 1200 Frequency (Hz)

SPL at muffler outlet: 8000 rpm

140

80 60

40

600

100

80 60

0

200

400

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

40

2000

0

200

400

Figure E1

54 Final Project Report 2011, SEIT, UNSW@ADFA

600

800 1000 1200 Frequency (Hz)

1400

1600

1800

2000

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E2

Figure E3

55 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Comparison of Transmission Loss to Weight of Absorption Silencer Ratio increase in Transmission Loss

4 3.5 3 2.5 2

Diameter

1.5

Length

1 0.5 0 0.00E+00 2.00E+00 4.00E+00 6.00E+00 8.00E+00 1.00E+01 1.20E+01 Ratio increase in weight Figure E4

Figure E5

56 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E6

Figure E7

57 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Transmission loss with choke diameter

Transmission loss with choke diameter

30

30 20mm 30mm no choke

20 15 10 5 0 -5

20mm 30mm no choke

25

Frequency (Hz)

Frequency (Hz)

25

20 15 10 5 0

0

200

400

600

800 1000 1200 1400 Transmission Loss (dB)

1600

1800

-5

2000

0

200

400

Transmission loss with choke diameter 30 20mm 30mm no choke

Frequency (Hz)

25 20 15 10 5 0 -5

0

200

400

600

800 1000 1200 1400 Transmission Loss (dB)

1600

1800

2000

Figure E8

58 Final Project Report 2011, SEIT, UNSW@ADFA

600

800 1000 1200 1400 Transmission Loss (dB)

1600

1800

2000

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E9

Figure E10

59 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E11

Figure E12

60 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E13

Figure E14

61 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E15

Figure E16

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Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E17

Figure E18

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Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E19

Figure E20

64 Final Project Report 2011, SEIT, UNSW@ADFA

Annex E WAVE Parameter Study – Exhaust Design for Scavenging Performance

Figure E21

65 Final Project Report 2011, SEIT, UNSW@ADFA