7th Sem Project

8th Semester B.Tech Project National Institute of Technology, Silchar

Department of Mechanical Engineering Page 1

1.1 INTRODUCTION:

A muffler (or silencer) is a device for reducing the amount of noise emitted by

the exhaust of an internal combustion engine. The significant difference in noise

level a muffler can produce can only be appreciated if we ever hear a car running

without a muffler. If vehicles did not have a muffler there would be an unbearable

amount of engine exhaust noise in our environment. Noise is defined as unwanted

sound. Exhaust noise from engines is one of component noise pollution to the

environment [1]. Exhaust systems are developed to attenuate noise meeting

required decibel (dB) levels and sound quality, emissions based on environment

norms. Hence this has become an important area of research and development.

Most of the advances in theory of acoustic filters and exhaust mufflers have been

developed in last two decades.

Sound is a pressure wave formed from pulses of alternating high and low

pressure air. In an automotive engine, pressure waves are generated when the

exhaust valve repeatedly opens and lets high-pressure gas into the exhaust

system. These pressure pulses are the sound we hear. As the engine rpm

increases so do the pressure fluctuations and therefore the sound emitted is of a

higher frequency. Internal combustion engine are typically equipped with an

muffler to supress the acoustic pulse generated by the combustion process. A

high intensity pressure wave generated by combustion in the engine cylinder

propagates along the exhaust pipe and radiates from the exhaust pipe

termination. The pulse repeats at the firing frequency of the engine which is

defined by f = (engine rpm x number of cylinders)/120 for a four stroke engine.

The frequency content of exhaust noise is dominated by a pulse at the firing

frequency [2].

Apart from the exhaust, noise in an automobile is also contributed by intake of

gases, mechanical vibrations in the engine body and transmission.



2.1 TYPES OF MUFFLER DESIGN:

A number of variations of the two main types of muffler designs commonly

used, namely absorptive and reactive. Generally automotive mufflers will have

both reactive and absorptive properties [3].

The reactive (or reflective) mufflers are based on the principle of destructive

interference to reduce noise. This means that they are designed so that the

sound waves produced by an engine partially cancel themselves out in the

muffler. For complete destructive interference to occur a reflected pressure

wave of equal amplitude and 180 degrees out of phase needs to collide with

the transmitted pressure wave. Reflections occur where there is a change in

geometry or an area discontinuity.

2.1.1 Reactive muffler:

A reactive muffler generally consists of a series of resonating and expansion

chambers that are designed to reduce the sound pressure level at particular

frequencies. The inlet and outlet tubes are generally offset and have

perforations that allow sound pulses to scatter out in numerous directions

inside a chamber resulting in destructive interference.

The wide application of reactive mufflers is in car exhaust systems where the

exhaust gas flows and hence noise emission varies with time. They have the

ability to reduce noise at various frequencies due to the numerous chambers

and changes in geometry that the exhaust gasses are forced to pass through

[4].

The drawback of reactive mufflers is that larger backpressures are created,

however for passenger cars where noise emission and passenger comfort are

highly valued reactive mufflers are ideal and can be seen on most passenger

vehicles on our roads today [5].



Fig 1: A layout of a typical reactive muffler

2.1.2 Absorptive muffler:

An absorptive (or dissipative) muffler uses the principle of absorption to

reduce sound energy. Sound waves are reduced as their energy is converted

into heat in the absorptive material. A typical absorptive muffler consists of as

straight, circular and perforated pipe that is encased in a larger steel housing.

Between the perforated pipe and the casing is a layer of sound absorptive

material that absorbs some of the pressure pulses [6]. Absorptive mufflers

create less backpressure then reactive mufflers, however unlike reactive

mufflers they do not reduce noise effectively.

Fig 2: An absorptive muffler

Usually reactive mufflers use resonating chambers that target specific

frequencies to control noise whereas an absorptive silencer reduces noise



considerably over the entire range of spectrum and more so at higher

frequencies.

A muffler is generally designed to work best in the frequency range where the

engine has the highest sound energy. In practice the sound spectrum of an

engine exhaust is continually changing, as it is dependent on the engine speed

that is continually varying when the car is being driven. It is impossible to

design a muffler that achieves complete destructive interference, however

some waves will always get cancelled.

Noise spectrum variation makes muffler design quite difficult and testing is the

only sure way to determine whether the muffler performs well at all engine

speeds. However, as a thumb rule, exhaust noise is generally limited to the

fundamental frequency and the first few harmonics, which can be calculated,

therefore these frequencies should be used as a starting point for preliminary

muffler design [7].

One of the practical ways of determining the frequency range to be controlled

is to measure the unmuffled engine noise. This measured spectrum can then

be used to identify the frequencies, at which the higher noise levels occur. The

high noise level frequencies should be treated with appropriate noise control

to achieve an overall noise reduction.

Design of mufflers are always application specific, however if the designed muffler is practical and achieves the required noise reduction and meets all functional requirements then the designer has succeeded.



3.1 PARAMETERS INVOLVED IN THE DESIGN OF AUTOMOTIVE MUFFLER:

A number of functional parameters that should be considered while designing a muffler for a specific application. Such functional requirements may include adequate insertion loss, backpressure, size, durability, desired sound, cost, shape and style [8]. These functional requirements are discussed below focusing on an automotive mufflers functional requirements.

3.1.1 Adequate Insertion Loss:

The vital function of a muffler is to attenuate or eliminate unwanted sound. An

effective muffler will reduce the sound pressure of the noise source to the

required level. In the case of an automotive muffler the noise in the exhaust

system, generated by the engine, is to be reduced.

The performance of muffler or attenuating capability is generally defined in

terms of insertion loss or transmission loss. Insertion loss is defined as the

difference between the acoustic power radiated without and with a muffler

fitted. The transmission loss is defined as the difference (in decibels) between

the sound power incident at the entry to the muffler to that transmitted by the

muffler.

The designer must determine the required insertion loss so that a suitable style

of muffler can be designed for the specific purpose. As a general principle

when designing an automotive muffler, a reactive muffler with many area

discontinuities will achieve a greater attenuation than one with fewer area

discontinuities. The addition of sound absorptive material will always increase

the attenuation capacity of a muffler, but should be located in an appropriate

place.

3.1.2 Backpressure:

Backpressure refers to the extra static pressure exerted by the muffler on the

engine through the restriction in flow of exhaust gasses. Generally the better a

muffler is at attenuating sound the more backpressure is generated. In a

reactive muffler where good attenuation is achieved the exhaust gases are

forced to pass through numerous geometry changes and a fair amount of

backpressure may be generated, which reduces the power output of the

engine. Backpressure should be kept to a minimum to avoid power losses



especially for performance vehicles where performance is the major

requirement. Every time the exhaust gasses are forced to change direction

additional backpressure is created. Therefore to limit backpressure geometric

changes are to be kept to a minimum, a typical example of this is a straight

through absorption silencer. Exhaust gasses are allowed to pass virtually

unimpeded through the straight perforated pipe.

3.1.3 Size:

Availability of space has a predominant influence on the size and type of

muffler that may be used. A muffler may have its geometry designed for

optimum attenuation however if it does not meet the space constraints, it is of

no use. Generally the larger a muffler is, the more it weighs and the more it

costs to manufacture. For a performance vehicle every gram saved is crucial to

its performance/acceleration, especially when dealing with light open wheeled

race vehicles. Therefore a small lightweight muffler is desirable.

Effectively supporting a muffler is always a design issue and the larger a

muffler is the more difficult it is to support. A mufflers mounting system not

only needs to support the mufflers weight but it also needs to provide

vibration isolation so that the vibration of the exhaust system is not

transferred to the chassis and then to the passenger cabin. This isolation of

vibration is usually achieved with the use of hard rubber inserts and brackets

that isolate or dampen vibration from the muffler to the chassis.

3.1.4 Durability:

The life expectancy of a muffler is another important functional requirement

especially when dealing with hot exhaust gasses and absorptive silencers that

are found in performance vehicles.

Overtime, hot exhaust gasses tend to clog the absorptive material with

unburnt carbon particles or burn the absorptive material in the muffler. This

causes the insertion loss to deteriorate. There are however, good products

such as mineral wool, fibreglass, sintered metal composites and white wool

that resist such unwanted effects [9].

Reactive type mufflers with no absorptive material are very durable and their

performance does not diminish with time. Generally mufflers are made from

corrosion resistive materials such as stainless steel or aluminium. Mild steel or

aluminised steel is generally used for temperatures up to 500C, type 409



stainless steel up to 700 C and type 321stainless steel for even higher

temperatures. Automotive exhaust gas temperatures are usually around 750

C.

3.1.5 Desired Sound:

A muffler is used to reduce the sound of a combustion engine to a desired

level that provides comfort for the driver and passengers of the vehicle as well

as minimising sound pollution to the environment. Muffler designs generally

aim to reduce any annoying characteristics of the untreated exhaust noise such

as low frequency rumble. Muffler modification of a stock vehicle is generally

done for two reasons being performance and sound. Vehicles leave the factory

floor with mufflers generally designed for noise control not optimal

performance. The standard reactive muffler is generally replaced with a

straight through absorption silencer for aesthetics and to minimise

backpressure and therefore improve vehicle performance. Having exchanged

the stock muffler for an absorptive type performance muffler generally means

that exhaust noise is increased, leaving a noticeable deep rumble in the

exhaust system. In most cases this sound is what the owner of the vehicle

desires so that the public is aware of their presence. However in the main

mufflers should be designed so that exhaust noise emission is only barely

audible within the passenger cabin and the appropriate government

regulations are adhered to. Breakout noise from the muffler shell may be a

problem and should be minimised together with flow-generated noise,

especially when designing a muffler for a high insertion loss [10].

3.1.6 Cost:

The most important factor in any component is the cost to the consumer.

Silencers not only have to be effective in performing their task they need to be

affordable otherwise the product will failing the marketplace. The cost is

dependent on the materials used in the construction of the muffler, design

integrity, durability and labour costs.

3.1.7 Shape and style:

Automotive mufflers come in all different shapes, styles and sizes depending

on the desired application. Generally automotive mufflers consist of an inlet

and outlet tube separated by a larger chamber that is oval or round in

geometry. The inside detail of this larger chamber may be one of numerous



constructions. The end user of the muffler usually does not care what is inside

this chamber so long as the muffler produces the desired sound and is

aesthetically pleasing. It is therefore the task of the muffler designer to ensure

that the muffler is functional as well as marketable.

3.2 Possible muffler designs:

There are various types of automotive mufflers currently in the market place

and described below are the key features and benefits of various muffler

designs that may be found on a vehicle. The following types of mufflers have

been widely tested and the general observations from such tests are

described.

Automotive mufflers usually have a circular or elliptical cross section. A circular

shaped cross section is best suited in a vehicle as it delays the onset of higher

order modes.

Most formulas that are used to predict the transmission loss of a muffler

assume plane wave propagation. The properties of the following designs are

only valid up to the cut off frequency, where higher order modes occur.

Generally for all mufflers maximum transmission loss occurs at odd multiples

of a quarter wavelength.

The most basic type of silencing element that may be used for intake and

exhaust mufflers is the expansion chamber. It consists of an inlet tube, an

expansion chamber and an outlet tube. The inlet and outlet tubes may be

coaxial known as a concentric expansion chamber or offset known as an offset

expansion chamber.

Fig 3: A simple expansion chamber

The sudden expansion and contraction in this type of muffler causes sound

waves to reflect back and interfere with each other. Expansion chambers are



efficient in attenuating low frequency sound, which makes them ideal for

automotive applications. They do not attenuate high frequency sound so well

as it beams straight through the muffler.

Fig 4: the incident, reflected and transmitted sound waves caused by a change in cross sectional area.

Expansion chamber mufflers have been widely studied and results show that

the larger the expansion ratio the greater the transmission loss. The length of

the chamber should be at least 1.5 times the diameter. Similar to a standard

expansion chamber is the extended inlet and outlet expansion chamber, where

the inlet and outlet tubes are extended into the expansion chamber. The

benefit of such a design is that part of the chamber between the extended pipe

and the sidewall acts as a side branch resonator therefore improving the

transmission loss.

The greater the protrusion into the muffler the greater the transmission loss

however the inlet and outlet tubes should maintain a separation space of at

least 1.5 times the diameter of the chamber to ensure the decay of evanescent

modes.

Fig 5: Expansion chamber with extended inlet and outlet

Noise can be further attenuated by the addition of porous material inside the

expansion chamber whilst maintaining the same muffler dimensions. Sound

waves loose energy as they travel through a porous medium. The absorptive

material (porous material) causes the fluctuating gas particles to convert

acoustic energy to heat.



Fig 6: Straight through absorptive muffler

Generation of insignificant backpressure is the main benefit of a straight

through absorptive silencer, thus improving vehicle performance. The

perforated tube is used to guide the exhaust flow and avoids the creation of

turbulence as is found in an expansion chamber.

The material used to guide the exhaust flow, yet allow sound waves to escape,

is usually perforated steel with an open area of approximately 20%.An

absorptive silencer produces a more consistent transmission loss (TL) curve.

A more broader and improved attenuation spectrum is achieved with multiple

resonators. Each chamber is designed to reduce a specific frequency being an

odd multiple of a quarter wavelengths apart. Attenuation is increased as the

number of chambers increase although the addition of a third chamber only

provides a small increase in attenuation. If a tube connects the chambers, the

longer the tube the greater the attenuation achieved. This type of muffler is

useful when space is limited and low frequency performance is required.

The volume and shape of the resonating chamber govern its performance

capabilities. Generally as the volume of the resonating chamber increases the

resonant frequency reduces.

Fig 7: A multiple resonative chamber muffler



3.3 Modal Analysis:

The modal analysis of structures is well developed and is a powerful tool in

dynamic analysis. However, a structural body is too complex to be represented

by a single degree of freedom model. Such a body is a continuous body where

all of the three properties namely inertia, damping and stiffness are

continuously distributed from the domain of the system and are inseparable

from one another. However, in our case we have considered the muffler to be

a discrete or lumped system where the inertia, stiffness and stiffness

properties are lumped into specific locations. Such a system is commonly

known as multi-degree of freedom (MDOF) system. Our basic thrust will be to

formulate the required equations of motion, which necessarily addresses all

the possible modes of displacement of the system. Matrix formulation of such

large number of equations, where in general the number of motion equations

is equal to the number of degree of freedom of system, is the only way perhaps

to process them in an elegant way [11].

A mode of vibration is characterized by a modal frequency and a mode shape.

It is numbered according to the number of half waves in the vibration.

Our main aim is to find the natural frequencies of vibration for different modes

and check that whether any of these frequencies so obtained match with the

working frequency range of the engine i.e. from idling to maximum power

operation. This analysis helps in checking the occurrence of resonance in the

muffler and selection of mounting points in the zone of maximum strain

deformation. In our analysis using the Modal Analysis of the Ansys Workbench

14.0, we have considered the muffler to be a 6 DOF body system as the free

vibrations are prominent in the first few harmonics keeping the edge of the

inlet pipe and the edge of the muffler outer chamber as was in the procured

original model.

The following is an analysis for a two degree of freedom system which can be

extended for the analysis of an MDOF system:

Let us consider two equal bodies (not affected by gravity), each of mass, m,

attached to three springs, each with spring constant, k. They are attached in

the following manner:



Here the edge points are fixed and cannot move. We'll use x1(t) to denote the horizontal displacement of the left mass, and x2(t) to denote the displacement of the right mass.

If we denote acceleration (the second derivative of x(t) with respect to time) as , the equations of motion are:

0)(

0)(

2321222

2212111

xkkxkxm

xkxkkxm

Since we expect oscillatory motion of a normal mode (where is the same for both masses), we try:

)sin()(

)sin()(

2

1

tBtx

tAtx

Substituting these into the equations of motion gives us:

0sin)2(

0sin)2(

2

2

tkAkBmB

tkBkAmA

The above two equations are satisfied for every t and therefore,

02

02

2

2

kAkBmB

kBkAmA

And in matrix representation:

02

2det

2

2

mkk

kmk



Solving for , we have two positive solutions:

m

k

m

k 3,

First natural frequency:

BAm

k

The displacement for the first natural frequency is:

A

A

This vector is called mode shape. This indicates that, both masses move in phase and the have the same amplitudes.

Second natural frequency:

BAm

k

3

Mode shape for above frequency:

A

A

Thus, the two masses move in opposite directions.

Displacements of two masses are sums of displacements in the two modes:

)3

sin(1

1)sin(

1

1

)(

)(2211

2

1

t

m

kct

m

kc

tx

tx

,

which is the general expression for vibration of the two-degree of freedom system and c1, c2, 1, and 2, are determined by the initial conditions of the problem.



4.1 Literature Review:

M.L. MUNJAL worked on the ANALYSIS AND DEIGN OF MUFFLERS [1].

This paper describes the cause of noise generation and vibration in

reciprocating engine, motion of the piston of engines and compressors and the

associated intake and discharge of gases are responsible for noise radiation to

the atmosphere that ranks as a major pollutant of the urban environment.

Muffler has been developed based on electro-acoustic analysis and

experimental trial and error. This article basically concern with passive muffler

based on impedance mismatch called dissipative or reactive muffler have been

most common in the automobile industry muffler based on the principle of

conversion of acoustic energy into heat by means of highly porous-fibrous

linings, called dissipative muffler or silencers are generally used in heating,

ventilation and air-conditioning systems. He has been working in the vibro

acoustics of hoses used in automotive climate control systems. The present

paper gives an overview of the research needing in different aspects of active

as well as passive mufflers. The same are related to the contemporary state of

the art finally areas needing further research are indicated. Exhaust noise of

automotive engine is the main component of noise pollution of the urban

environment with the ever increasing population density of vehicles on the

road. This has been an important area of research and development. Most of

the advances in acoustics of ducts and muffler reference article following the

monograph in format briefly reviews the work that is subsequent to the

drafting of monograph.

SHITAL SHAH, SAISANKARANARAYANA K, KALYANKUMAR S. HATTI, Prof.

D.G.THOMBARE, worked on A PRACTICAL APPROACH TOWARDS

MUFFLER DESIGN, DEVELOPMENT AND PROTOTYPE VALIDATION [2].

This paper deals with a practical approach to design, develop and test muffler

particularly reactive muffler for exhaust system, which will give advantages

over the conventional method with shorten product development cycle time

and validation. This paper also emphasis on how modern CAE tools could be

leveraged for optimising the overall system design balancing conflicting

requirements like Noise & Back pressure.



H. BARTLETT, R. WHALLEY worked on MODELLING AND ANALYSIS OF

VARIABLE GEOMETRY EXHAUST GAS SYSTEMS [3].

This paper presents the modelling and analysis of variable geometry exhaust

gas systems. An automotive example is considered whereby the pulsating

exhaust gases flow through an exhaust pipe and silencer are considered over a

wide range of speeds. Analytical procedures are outlined enabling the general

analysis and modelling of variable geometry, exhaust gas systems. Simulation

results show the effect of pulsating gas streams through a vehicle exhaust and

silencer confirming thereby the calculated results.

S. BOIJ, B. NILSSON worked on the REFLECTION OF SOUND AT AREA

EXPANSIONS IN A FLOW DUCT [4].

The object of this paper is to describe the reflection of sound wave of an

analytical model for scattering at area discontinuities and sharp edges in flow

ducts and pipes. A large industrial duct system, where sound attenuation by

reactive and absorptive baffle silencers is of great importance. Such devices

commonly have a rectangular cross-section, so the model is chosen as 2-D. The

modelling of the flow conditions downstream of the area expansion, with and

without extended edges, and its implications for the resulting models are

discussed. Here the scattering problem is solved with the Wiener-Hopf

technique, and a kutta condition is applied at the edge. The solution of the

wave equation downstream of the expansion includes hydrodynamic waves, of

which one is a growing wave. Theoretical results are compared with the

experimental data for the reflection coefficient for the plane wave, at

frequencies below the cut-on for higher order modes. Influence of the

interaction between the sound field and the flow field is discussed. A region

where the reflection coefficient is strongly Strouhal number dependent is

found.

V.P. PATEKAR, R.B. PATIL worked on the VIBRATIONAL ANALYSIS OF

AUTOMOTIVE EXHAUST SILENCER BASED ON FEMAND FFT ANALYSER [5]

This paper describes the first stage in the design analysis of an exhaust system. With the specified properties of a given material, the exhaust system is modelled by using a conventional FEM package. The FEM results are compared with the reading taken on FFT analyser, so as to distinguish working frequency from natural frequency and avoid resonating condition. The silencer natural



frequencies have been calculated by using the ANSYS package and by FFT analyzer. By both the method the natural frequencies are nearly same and that are useful while the design of silencer to avoid the resonance. Though the dynamic performance can be increased by increasing the thickness of different part. Furthermore is to add the support for partition, increase the support etc.

GHAREHBAGHI M, IRAY MAKVANDI R. worked on VEHICLE INTERIOR ACOUSTIC OPTIMIZATION BY USING PASSIVE LAYERS [6].

In this paper, first the effect of structural vibration on the internal noise level of vehicle namely structure-borne noise is investigated. The range of investigated frequency is between 0-250 Hz. Acoustical FE simulation is performed on a three internal space cavity of a truck. The analysis was carried out in frequency response analysis and resultant internal noise level is determined. The results of the above analysis are used for optimizing driver's right ear noise level through location and thickness of damping layers. This method is categorized as a passive noise reduction method and is the most economical way for reducing the noise level for the mid-range vehicles. To achieve this goal, number of simulations is designed using Taguchis method in design of experiments (T-DOE). Then develop a general interpolation function on the sample points using artificial neural network (ANN) method. And finally determine the optimum state by means of a genetic algorithm (GA). Using this method the optimization of parameters (SPL and Weight of Passive Layers) was attained effectively. The results indicate that by proper use of these layers, one can make a reduction of 10-15 dB in sound pressure level.

KAUSHIK RAMCHANDRA GADRE, T. A. JADHAV, SWAPNIL S. KULKARNI worked on the EVALUATING THE DESIGN OF AN AUTOMOBILE SILENCER THROUGH FEA METHODOLOGY FOR MINIMIZING THE VIBRATIONS GENERATED DURING ITS OPERATION [7].

This paper discusses the FEA methodology to be followed in doing the vibrational analysis of an automobile muffler and also the experimental validation of the FEA results. This presents a computational approach for the lifetime assessment of structures. One of the main features of the work is the search for simplicity and robustness in all steps of the modeling, in order to match the proposed method with industrial constraints. The proposed method is composed of a fluid flow, a thermal and a mechanical finite element computation, as well as a final fatigue analysis. The CAE software has intuitive



graphical interface with direct access to CAD geometry, advanced meshing, integration with other compatible software for solving. It is optimized for large scale systems, assemblies, dynamics and NVH simulations. It has graphical interface with direct access to CAD geometry, most suitable for fatigue analysis.

CHANDRASEKHAR BHAT, S.S.SHARMA, JAGANNATH K, N S MOHAN, SATHISHA S G worked on DESIGN AND ANALYSIS OF EXPANSION CHAMBER MUFFLERS [8].

This paper attempts to predict the transmission loss through modal analysis, followed by acoustic analysis using finite element analysis technique for three different configurations of mufflers under different fixing conditions. It was found that three-chamber muffler provides higher attenuation of sound pressure compare to one and two chamber mufflers. And, fixing the muffler at the center enhances sound pressure attenuation. The fact that higher natural frequencies occur inside the chamber and the lower natural frequencies at the free ends, strengthen the need for proper design of muffler. Number of natural frequencies reduces for two and three chamber mufflers. The transmission loss curve appears to be high and broad at higher frequencies for two-chamber muffler and is broad at multiple frequencies for three-chamber muffler.

WANG JIE, DONG-PENG VUE worked on THE MODAL ANALYSIS OF AUTOMOTIVE EXHAUST MUFFLER BASED ON PRO/E AND ANSYS [9].

This paper discusses the fact that in order to improve the design efficiency, resonating of the exhaust muffler should be avoided with its natural frequency. The solid modelling is created by the PRO/E, and modal analysis is carried out by ANSYS to study the vibration of the muffler, so as to distinguish working frequency from natural frequency and avoid resonating. Multi-degrees of freedom, the finite element method of dynamics is the same as structure of static problem, which make objects discrete into finite number of elements body. But considering the unit features in this condition, the load which object is suffered should be considered by many factors such as inertial force and damping.



A. I. KOMKIN worked on OPTIMIZATION OF REACTIVE MUFFLERS [10] The paper explains a new approach to optimization of reactive mufflers, which is based on use of muffler prototype with non-dimensional geometrical parameters and integral criterion of acoustic performance of mufflers, is proposed. Implementation of the approach using the example of chamber mufflers is considered. Optimization of its configuration is important for the performance of a muffler. This permits, on the one hand, determining the configuration of a muffler with maximum acoustic performance under given dimensional restrictions and, on the other hand, estimating the necessary minimum volume of a muffler to ensure required acoustic performance.

POTENTE, DANIEL worked on GENERAL DESIGN PRINCIPLES FOR AN AUTOMOTIVE MUFFLER [11]

This paper discusses the general principles of muffler design and explains the advantages of different configurations of mufflers based on application. When designing a muffler for any application there are several functional requirements which are to be considered, this includes both acoustic and non-acoustic design issues. There will be many possible muffler design solutions for a particular situation and many possible ways to predict a mufflers insertion loss but the design is proven by its performance on the automobile.



5.1 Problem Formulation

Here we have considered the existing muffler model from a Maruti Omni

having a 3-cylider in-line 4-stroke engine for designing and analysis purpose.

The muffler used in this car is a reactive muffler and the exhaust system has

the following specifications:

5.1.1 Specifications:

Type of engine 3-cylinder in-line 4-stroke

Engine RPM 5000

Maximum engine frequency 83.33 Hz

Working frequency range 7.3-83.33 Hz

Type of muffler Reactive type (offset inlet and outlet

pipes)

Type of resonating chamber Multiple (3 chambers)

Cross-section of resonating chamber Circular

Perforations None

Baffles None

Diameter of the resonating chamber 13.32 cm

Length of the resonating chamber 41 cm

Diameter of the inlet pipe 3.3 cm OD; 3 cm ID

Table : 1 Specifications



Fig 8: The existing muffler model of Maruti Omni

The existing model was designed using CATIA V5R20 and the analysis work was

carried out using Ansys 14.0. The figure above shows the cad model of the

existing muffler.

5.2 Design considerations:

There were many scopes for re-designing the muffler considering different

configurations such as-

Changing the lengths of the resonating chambers

Perforating the pipes at specific locations

Tapering of the inlet pipe

Applying absorptive material over the pipes and between the layers of

the resonating chamber

Changing the diameter of the pipes

Changing the cross-section of the resonating chamber (oval or circular)

Providing baffles inside the resonating chamber.

Using an additional outer skin over the outer chamber with a air gap in

between.

Using delta configuration in the circular resonating chamber.

Out of the above options, we have considered the five configurations and

have designed five models keeping the following inlet boundary conditions:



Inlet pressure 3 x 105 Pa

Inlet velocity 200 m/sec

Inlet temperature 800 K

Table : 2 Inlet boundary conditions

The material which we have considered for the three models is 409 Stainless

steel that has a capacity of operating at temperatures above 500C.

5.3 Modelling and analysis

The following is a detailed report on the design and analysis of the three

models including the existing Maruti Omni model:

5.3.1 The existing Maruti Omni muffler model:

The exhaust system was removed from the Maruti Omni and the outer

dimensions such as diameter of the inlet and outlet pipes, diameter of the

resonating chamber and the length of the resonating chamber were taken

using meter tape. Then a part of the resonating chamber was cut using power

saw and angle grinder to take the inner dimensions such as length of the

resonating chambers, thickness of the obstruction plates, the length of the

extension of the pipes inside each chamber etc. The designing of the model

was then done using CATIA V5R20 and the analysis was done in ANSYS 14.0.

Different contours were selected for comparison such as acoustic power level,

static pressure, absolute pressure and the turbulence intensity level.

The models and the analysis of the existing model are shown in the successive

pages.



5.3.1(a) Modelling and meshing

Fig 9: An isometric view of the existing Maruti Omni muffler model

Fig 10: Volume meshing of the existing Maruti Omni model carried out in Ansys 14.0



5.3.1(b) Pressure, turbulence and acoustic stress contours

Fig 11: Absolute pressure contour inside the resonating chamber

Fig 12: Static pressure contour inside the resonating chamber



Fig 13: Turbulence intensity inside the resonating chamber

Fig 14: Surface acoustic power level inside the resonating chamber



Fig 15: Acoustic power level across the resonating chamber

5.3.1(c) Modal analysis and different mode shapes

Fig 16: Total deformation in the 1st mode



Fig 17: Total deformation in the 2nd mode

Fig 18: Total deformation in the 3rd mode



Fig 19: Total deformation in the 4th mode





5.3.2 Model proposed by changing the length of the resonating chamber:

The existing model was redesigned by reducing the length of the last section of

the resonating chamber in view of the fact that if the reflected pressure wave

of approximately same frequency interferes destructively with the incoming

transmitted wave of that frequency then effective attenuation can be

achieved. As the pressure wave travels through the last chamber, longer the

distance travelled by the incoming pressure wave, more is the transmission

loss due to the interference of the wave with the medium. And lesser amount

of destructive interference takes place in the chamber. The improvement in

the attenuation level can be explained from the transmission loss.




Fig 22: An isometric view of the proposed model with reduced length of one of

the segments of the resonating chamber

Fig 23: Volume meshing of the proposed model with reduced length of one of the

segments of the resonating chamber



5.3.2(b) Pressure, turbulence and acoustic stress contour

Fig 24: Absolute pressure distribution inside the proposed model with reduced length

of one of the segments of the resonating chamber

Fig 25: Static pressure distribution inside the proposed model with reduced length of

one of the segments of the resonating chamber



Fig 26: Turbulence intensity inside the proposed model with reduced length of one of

the segments of the resonating chamber

Fig 27: Surface acoustic power level inside the resonating chamber



Fig 28: Acoustic power level across the resonating chamber






5.3.3 Model proposed by tapering the inlet pipe and reducing the length of

the last section of the resonating chamber:

Here the model was modified by gradual tapering of the inlet pipe on account

of the fact that pressure reduction takes place by gradually decreasing the

cross section of a pipe when a fluid flows through it. The earlier model didnt

account for the reduction of back pressure, only attenuation level was

improved. This model takes care of both the parameters i.e. reduction in noise

level and back pressure on the engine




Fig 35: An isometric view of the proposed model with tapered inlet pipe and reduced length of one of the segments of the resonating chamber

Fig 36: Volume meshing of the proposed model with tapered inlet pipe and reduced

length of one of the segments of the resonating chamber




Fig 37: Absolute pressure distribution of the proposed model with tapered inlet pipe and reduced length of one of the segments of the resonating chamber

Fig 38: Static pressure distribution of the proposed model with tapered inlet pipe and

reduced length of one of the segments of the resonating chamber



Fig 39: Turbulence intensity of the proposed model with tapered inlet pipe and reduced length of one of the segments of the resonating chamber

Fig 40: Surface acoustic power level of the proposed model



Fig 41: Acoustic power level of the proposed model






.3.4 Model proposed by tapering the inlet pipe and reducing the length of

the last section of the resonating chamber and providing an outer skin with

air gap in between:

Here the model was further modified by introducing an air gap by providing

an outer skin over the muffler chamber. This air gap acts as an additional

resonating chamber for attenuating noise level so that further reduction of

sound can be obtained. The outer additional chamber acts as a side branch

resonator which attenuates specific frequency band though the attenuation

band is very low. This model is used to treat particular frequency problem in

addition to the main chamber. The air gap also acts as a resistance to

thermal conductivity thereby reducing the heat transfer to the outer layer of

the muffler and as such increases the durability of the mountings.




Fig 48: An isometric view of the proposed model with reduced length of last resonating chamber and tapered inlet pipe and double outer skin with air gap.

Fig 49: Volume meshing of the proposed model with tapered inlet pipe and reduced

length of one of the segments of the resonating chamber and double outer skin with air

gap in between.




Fig 50: Absolute pressure distribution of the proposed model with tapered inlet pipe

and reduced length of one of the segments of the resonating chamber and outer skin

with air gap

Fig 51: Static pressure distribution of the proposed model with tapered inlet pipe and

reduced length of one of the segments of the resonating chamber and outer skin with

air gap



Fig 52: Turbulence intensity of the proposed model with tapered inlet pipe and

reduced length of one of the segments of the resonating chamber and outer skin with air gap

Fig 53: Surface acoustic power level of the proposed model with tapered inlet pipe

and reduced length of one of the segments of the resonating chamber and outer skin

with air gap



Fig 54: Acoustic power level of the proposed model with tapered inlet pipe and

reduced length of one of the segments of the resonating chamber and outer skin with

air gap






.3.5 Model proposed by incorporating delta configuration in the last

resonating chamber:

Modification was done in the existing muffler model by incorporating delta

configuration the last resonating chamber. The delta configuration is normally

being used with an oval type resonating muffler. This design provides

advanced noise cancellation by separating and recombining pulses at precise

phase shifts. It increases the range of frequency over which the noise

attenuation is achieved. Thus more attenuation can be achieved in addition to

the normal range for which the original model was designed. The outlet pipe is

at the middle of the chamber which blocks power-robbing atmospheric

pressure from entering the system which would reduce scavenging capabilities

of the system. The design is rewarded with higher horsepower, torque and fuel

economy and with reduction of exterior sound levels as well as a decrease of

internal resonance inside the vehicle.




Fig 61: An isometric view of the proposed model with delta configuration in the existing

model.

Fig 62: Volume meshing of the proposed model with delta configuration in the

existing model.




Fig 63: Absolute pressure of the proposed model with delta configuration in the

existing model.

Fig 64: static pressure of the proposed model with delta configuration in the existing

model.



Fig 65: Turbulence intensity of the proposed model with delta configuration in the

existing model.

Fig 66: Surface acoustic power level of the proposed model with delta configuration

in the existing model.



Fig 67: Acoustic power level of the proposed model with delta configuration in the

existing model.





6.1 Results, Calculations and Discussion

6.1.1 The existing Maruti Omni muffler model:

The following results can be summarised from the analysis of the existing model:

Back pressure: There is a back pressure at the inlet due to the restrictions provided in the chamber. Absolute pressure at inlet in steady state = 1.48 x 105 Pa Absolute inlet boundary condition pressure = 4 x 105 Pa Back pressure at the inlet = (4- 1.48) x 105 = 2.52 x 105 Pa Turbulence intensity: Percentage of turbulence intensity increases from inlet to the outlet. Inlet pipe: 3.37 x 102

Outlet pipe: 5.88 x 102

Increase in turbulence intensity helps in the proper mixing of pressure pulses improving the attenuation of sound inside the chamber

Acoustic power level: The sound intensity is reduced from the inlet to

the outlet of the resonating chamber due to the muffling effect which can be

shown in the form of transmission loss from inlet to the outlet.

Inlet pipe transmission loss: 78.9 dB

Outlet pipe transmission loss: 94.7 dB

Therefore the acoustic transmission loss through the resonating

chamber = (94.7- 78.9) = 15.8 dB

Modal analysis: The modal analysis for this model was carried out for

the first 6 modes under free vibration condition by fixing the edge of the inlet

pipe and the edge of the muffler outer chamber. Also the region of maximum

strain deformation was identified from various regions of strain deformations

of different modes. This was done in order to select the proper mounting

positions for the muffler. The analysis was carried out using the modal

analysis tool in Ansys Workbench 14.0.



Modal frequency and strain deformation results

Table : 3 Modal frequency and strain deformation

6.1.2 Model proposed by reducing the length of the last section of the

resonating chamber:


Back pressure: There is a back pressure at the inlet due to the restrictions provided in the chamber. Absolute pressure at the steady state = 1.3 x 105 Pa Absolute boundary condition pressure = 4 x 105 Pa Back pressure at the inlet = (4- 1.3) x 105 = 2.70 x 105 Pa Back pressure increases as compared to the original model because as the last chamber is made shorter the pressure pulse will move through a shorter path and as a result it offers more restrictions.

Turbulence intensity: Percentage of turbulence intensity increases from inlet to the outlet. Inlet pipe: 9.09 x 101





Modes Frequency

(Hz)

Strain Deforrmation

1st 511.53 1.5918

2nd 513.47 1.6035

3rd 886.82 3.8924

4th 893.80 3.9054

5th 1373.30 1.3485

6th 1552.60 0



Inlet pipe transmission loss: 70 dB



chamber = (93.2- 70) = 23.2 dB

Here an improved noise attenuation level is achieved due to the fact that the

sound pressure pulses after reflection from the last chamber interferes

destructively with the pressure pulses of same order.


the first 6 modes and the results are listed below. The frequency range of free

vibration is higher than the earlier model and range of strain deformation is

less. Here there is a strain deformation in the 6th mode also.




the last section of the resonating chamber:


Back pressure: There is a back pressure at the inlet due to the restrictions provided in the chamber. Absolute pressure at the steady state = 1.7 x 105 Pa Absolute boundary condition pressure = 4 x 105 Pa Back pressure at the inlet = (4- 1.7) x 105 = 2.30 x 105 Pa

Modes Frequency

(Hz)

Strain Deformation

1st 771.69 1.6349

2nd 774.58 1.6325

3rd 1428.30 4.1214

4th 1470.60 4.4926

5th 1571.30 1.9025

6th 1759.50 4.0227



Back pressure reduces as compared to the second model because tapering of the outlet of the inlet pipe provides a nozzle action which causes a loss in pressure.

Turbulence intensity: Percentage of turbulence intensity increases from inlet to the outlet. Inlet pipe: 9.09 x 101








chamber = (99.6- 75.2) = 23.2 dB

Thus a further improved noise attenuation level is achieved due to more

restrictions provided in the form of tapering of pipe.



vibration is almost of the same order as compared to the earlier model but

range of strain deformation is more than the earlier case. At the 6th mode the

whole structure is becoming fixed with no strain deformation.



Modes Frequency

(Hz)

Strain Deformation

1st 705.37 2.8943

2nd 707.40 2.8124

3rd 1041.40 3.2375

4th 1053 3.2603

5th 1430.1 2.1959

6th 1762.30 0




the last section of the resonating chamber and double outer skin with air gap

in between:

The following results can be summarised from the analysis of the model:

Back pressure: There is a back pressure at the inlet due to the restrictions provided in the chamber. Absolute pressure at the steady state = 4.1 x 105 Pa Absolute boundary condition pressure = 4 x 105 Pa Back pressure at the inlet = (4- 4.1) x 105 = -0.1 x 105 Pa This design accounts for no back pressure on the contrary it creates a suction condition at the inlet which will drive the flow of exhaust gases into the muffler chamber more efficiently.

Turbulence intensity: Percentage of turbulence intensity increases from inlet to the outlet. Inlet pipe: 0.361

Outlet pipe: 8.03




Inlet pipe transmission loss: 13 dB



chamber = (52.2 - 13) = 39.2 dB

Here much greater attenuation level is achieved as there is a side branch

resonator in addition to the main chamber in the form of air gap which also

attenuates specific frequency level.



vibration is almost of the same order as compared to the earlier model but

range of strain deformation is less as compared to the 3rd model. At the 6th

mode the whole structure is becoming fixed with no strain deformation.





6.1.5 Model proposed by incorporating delta configuration in the existing

model:

The following results can be summarised from the analysis of the model:

Back pressure: There is a back pressure at the inlet due to the restrictions provided in the chamber. Absolute pressure at the steady state = 1.17 x 105 Pa Absolute boundary condition pressure = 4 x 105 Pa Back pressure at the inlet = (4 1.17) x 105 = 2.83 x 105 Pa This design accounts for the highest amount of back pressure among all the models which can be explained by the fact that it was a modification in the original model where back pressure was 2.52 x 105 Pa by incorporating three baffles in the shape of cone in the last resonating chamber which provided greater back pressure at the inlet as compared to the original model.

Turbulence intensity: Percentage of turbulence intensity increases from inlet to the outlet. Inlet pipe: 0.361

Outlet pipe: 8.03







chamber = (71.4 14.3) = 57.1 dB

Modes Frequency

(Hz)

Strain Deformation

1st 731.35 2.9524

2nd 733.31 2.9426

3rd 1041.30 3.0527

4th 1058.8 3.1465

5th 1262 1.8512

6th 1561.5 0



This design accounts for the greatest attenuation of noise level among all the

three models which is more than three times of the original one.


the first 6 modes and the results are listed below. The frequencies of free

vibration are lesser than all the models. But range of strain deformation is

less as compared to the 4th model. At the 6th mode there is small amount of

strain deformation in the structure.



Modes Frequency

(Hz)

Strain Deformation

1st 476.24 2.5898

2nd 476.94 2.6136

3rd 659.94 3.5194

4th 660.82 3.5364

5th 1181.06 1.9053

6th 1708.70 0.6416



6.2 Table for comparing among the 5 models

Model

Back

pressure

(Pa)

Acoustic

Trans.

Loss

(dB)

Frequency

(Hz)

Maximum

Strain

Deformation Mode

1 2 3 4 5 6

Original

2.52e+05

15.8

511.53

513.47

886.82

893.8

1373.3

1552.6

3.9054

Model with

reduced length of

last resonating

chamber

2.70e+05

23.2

771.69

774.58

1428.3

1470.6

1571.3

1759.5

4.4926

Model with

reduced length of

last resonating

chamber and

tapered inlet pipe

2.3e+05

24.4

705.37

707.4

1041.4

1053.0

1430.1

1762.3

3.2603

Model with

reduced length of

last resonating

chamber and

tapered inlet pipe

and double outer

skin with air gap

-0.1e+05

39.2

731.35

733.31

1041.3

1058.8

1262.0

1561.5

3.1465

Existing model

with delta

configuration

2.83e+05

57.1

476.24

476.94

659.94

660.82

1181.1

1708.7

3.5364

Table : 8 Comparison of the models



7.1 CONCLUSION

The present work is concerned with the modification of the existing Maruti

Omni muffler model through various design improvements in the resonating

chamber of the muffler. The acoustic and modal analysis using Ansys 14.0

confirms the improvements in the design. The different results obtained from

the analysis gives a comparative study among the various design modifications.

The following conclusions can be drawn from the analysis:

The existing provided for very low acoustic transmission loss of 15.8 dB

and also accounted for a significant amount of back pressure of about

2.52e+05 Pa. The natural frequencies obtained for the first 6 modes

showed no match with the engines working frequency range thereby

nulling the chance of occurrence of resonance.

The second model provided for more attenuation of 23.2 dB but

produced more back pressure of 2.70e+05 Pa as compared to the

existing model. Here also the systems natural frequencies in different

modes didnt match with the engines frequency range.

Modification of second model by tapering of inlet pipes outlet

accounted for reduction in back pressure which was around 2.30e+05

Pa and increased the attenuation level by 1.2 dB. The natural

frequencies didnt match with engines frequency preventing the

occurrence of resonance.

The fourth model accounted for the attenuation of noise at two ranges

one for higher frequency range and another for the lower ranges

thereby causing a significant attenuation as a whole by 14.8 dB as

compared to the third one. Here there was no back pressure but on the

contrary suction condition was created at the inlet by 0.1 Pa below the

atmospheric pressure which enhanced the flow in its defined direction

of flow. This design also showed isolation with respect to resonance

condition.

The fifth model produced the highest attenuation of 57.1 dB as

compared to the other 4 models but accounted for the highest amount

of back pressure of about 2.83e+05 Pa which is not desired. The same

conclusions can be drawn for the modal analysis in this case also.



Thus the model with delta configuration can be used in applications where

high attenuation of noise level is desired but cannot be used where low back

pressure is the requirement. Therefore the fourth model i.e. the model with

reduced last resonating chamber, tapered inlet pipe and with double outer

skin is the best optimised model among the all five models as there was no

problem of back pressure and it produced significant attenuation as compared

to the original one.

7.2 FUTURE SCOPE

Forced vibration analysis of the structure by taking pressure contour as

the input forces for the determination of frequency range of forced

vibration.

Experimental validation of the models in the vibration analysing set up

by a vibration analyser of suitable specifications.

Nodal analysis of the models for the selection of exact mounting points

around a line or a periphery of a cross-section.

Design of muffler mountings according to the values of maximum total

strain deformation.

Selection of suitable material for muffler design through static and

dynamic analysis.



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[2] A Practical Approach towards Muffler Design, Development and

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Engineering, Management and Technology 2 (3) 2014:24-32

[7] Evaluating the Design of an Automobile Silencer Through FEA

Methodology for Minimizing the Vibrations Generated During its Operation.

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BHAT, S.S.SHARMA, JAGANNATH K, N S MOHAN, SATHISHA S. G. Professor, Department of Mechanical & Mfg. Engineering, MIT, Manipal: World Journal Of Engineering,

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7th Sem Project

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