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Acoustics is the science that studies sound, in particular its production, transmission, and effects. Sound can often be

considered as something pleasant; music is an example. In that case a main application is room acoustics, since the purpose

of room acoustical design and optimisation is to make a room sound as good as possible. But some noises can also be

unpleasant and make people feel uncomfortable. In fact, noise reduction is actually a main challenge, in particular in the

industry of transportations, since people are becoming increasingly demanding. Furthermore, ultrasounds also have applications

in detection, such as sonar systems or non-destructive material testing. The articles in this wikibook describe the

fundamentals of acoustics and some of the major applications.

Fundamentals

Applications

Applications in Room Acoustics

Applications in Psychoacoustics

Musical Acoustics Applications

Miscellaneous Applications

Introduction

Sound is due to any variation in the atmospheric pressure. An easy way to understand how sound propagates is to consider that space can be divided into thin air layers. The vibration (the successive compression and relaxation) of these layers, at a certain velocity, enables the sound to propagate, hence producing a wave. This is the reason why sound waves cannot exist in an incompressible fluid.

In this chapter, we will only consider the propagation of sound waves in an area without any acoustic source, in an homogeneous fluid.

Equation of waves

Sound waves consist in the propagation of a scalar field, acoustic over-pressure, and a vector field, acoustic local velocity. Thus, the propagation of sound waves is governed by the two following equations, which are equivalent:

$\nabla ^2 p - \frac{1}{{c_0 ^2 }}\frac{{\partial ^2 p}}{{\partial x^2 }} = 0$ ; $\nabla ^2 \underline v - \frac{1}{{c_0 ^2 }}\frac{{\partial ^2 \underline v }}{{\partial x^2 }} = 0$

These equations are obtained using the conservation equations (mass, momentum and energy), the thermodynamic equations of state as well as behavior laws (Newtonian fluid, Fourier’s law of conduction). Once viscosity and conduction have been neglected, we consider that all perturbations remain small enough for the previous equations to be linearized (for example, the non-linear term in the momentum equation can be neglected). For specific cases, where acoustic over-pressure becomes too high (sonic boom, etc.), all non-linear terms must be kept, and we then have to deal with non-linear acoustics.

In the propagation equation of sound waves, $c 0$ is the propagation velocity of the sound wave (which has nothing to do with the vibration velocity of the air layers). This propagation velocity has the following expression:

$c_0 = \frac{1}{{\sqrt {\rho _0 \chi _s } }}$

where $ρ 0$ is the density and $χ S$ is the compressibility coefficient of the propagation medium.

Helmholtz equation

Since the velocity field $\underline v$ for acoustic waves is irrotational we can define an acoustic potential $Φ$ by:

$\underline v = grad\Phi$

Using the propagation equation of the previous paragraph, it is easy to obtain the new equation:

$\nabla ^2 \Phi - \frac{1}{{c_0 ^2 }}\frac{{\partial ^2 \Phi }}{{\partial t^2 }} = 0$

Applying the Fourier Transform, we get the widely used Helmoltz equation:

$\nabla ^2 \hat \Phi + k^2 \hat \Phi = 0$

where $k$ is the wave number associated with $Φ$ . Using this equation is often the easiest way to solve acoustical problems.

Acoustic intensity and decibel

The acoustic intensity represents the acoustic energy flux associated with the wave propagation:

$\underline i (t) = p\underline v$

We can then define the average intensity:

$\underline I = < \underline i >$

However, acoustic intensity does not give a good idea of the sound level, since the sensitivity of our ears is logarithmic. Therefore we define decibels, either using acoustic over-pressure or acoustic average intensity:

$p^{dB} = 20\log (\frac{p}{{p_{ref} }})$ ; $L_I = 10\log (\frac{I}{{I_{ref} }})$

where $p r e f = 2.10 - 5 P a$ for air, or $p r e f = 10 - 6 P a$ for any other media, and $I r e f = 10 - 12$ W/m².

Solving the wave equation

Plane waves

If we study the propagation of a sound wave, far from the acoustic source, it can be considered as a plane 1D wave. If the direction of propagation is along the x axis, the solution is:

$\Phi (x,t) = f(t - \frac{x}{{c_0 }}) + g(t + \frac{x}{{c_0 }})$

where f and g can be any function. f describes the wave motion toward increasing x, whereas g describes the motion toward decreasing x.

The momentum equation provides a relation between $p$ and $\underline v$ which leads to the expression of the specific impedance, defined as follows:

$\frac{p}{v} = Z = \pm \rho _0 c_0$

And still in the case of a plane wave, we get the following expression for the acoustic intensity:

$\underline i = \pm \frac{{p^2 }}{{\rho _0 c_0 }}\underline {e_x }$

Spherical waves

More generally, the waves propagates in any direction and are spherical waves. In these cases, the solution for the acoustic potential $Φ$ is:

$\Phi (r,t) = \frac{1}{r}f(t - \frac{r}{{c_0 }}) + \frac{1}{r}g(t + \frac{r}{{c_0 }})$

The fact that the potential decreases linearly while the distance to the source rises is just a consequence of the conservation of energy. For spherical waves, we can also easily calculate the specific impedance as well as the acoustic intensity.

Boundary conditions

Concerning the boundary conditions which are used for solving the wave equation, we can distinguish two situations. If the medium is not absorptive, the boundary conditions are established using the usual equations for mechanics. But in the situation of an absorptive material, we cannot write the equations for mechanics, and therefore, we have to use the concept of acoustic impedance.

Non-absorptive material

In that case, we get explicit boundary conditions either on stresses and on velocities at the interface. These conditions depend on whether the media are solids, inviscid or viscous fluids.

Absorptive material

Here, we do not know the equations for mechanics in the absorptive material and thus, we have to use the acoustic impedance as the boundary condition. This impedance, which is often given by experimental measurements depends on the material, the fluid and the frequency of the sound wave.

Fundamentals of Room Acoustics →

Introduction

Three theories are used to understand room acoustics :

The modal theory
The geometric theory
The theory of Sabine

The modal theory

This theory comes from the homogeneous Helmoltz equation ΔP+k2P=0. Considering a simple geometry of a parallelepiped (L1,L2,L3), the solution of this problem is with separated variables :

P (x, y, z) = X (x) Y (y) Z (z)

Hence each function X, Y and Z has this form :

X (x) = A e - i k x + B e i k x

With the boundary condition $\frac{{\partial P}} {{\partial x}} = 0$ , for x=0 and x=L1 (idem in the other directions), the expression of pressure is :

$P\left( {x,y,z} \right) = C\cos \left( {\frac{{m\pi x}} {{L1}}} \right)\cos \left( {\frac{{n\pi y}} {{L2}}} \right)\cos \left( {\frac{{p\pi z}} {{L3}}} \right)$

$k^2 = \left( {\frac{{m\pi }}{{L1}}} \right)^2 + \left( {\frac{{n\pi }}{{L2}}} \right)^2 + \left( {\frac{{p\pi }}{{L3}}} \right)^2$

where m,n,p are whole numbers

It is a three-dimensional stationary wave. Acoustic modes appear with their modal frequencies and their modal forms. With a non-homogeneous problem, a problem with an acoustic source Q in r0, the final pressure in r is the sum of the contribution of all the modes described above.

The modal density $\frac{{dN}}{{df}}$ is the number of modal frequencies contained in a range of 1Hz. It depends on the frequency f, the volume of the room V and the speed of sound c0 :

$\frac{{dN}}{{df}} \simeq \frac{{4\pi V}}{{c_0^2 }}f^2$

The modal density depends on the square frequency, so it increase rapidly with the frequency. At a certain level of frequency, the modes are not distinguished and the modal theory is no longer relevant.

The geometry theory

For rooms of high volume or with a complex geometry, the theory of acoustical geometry is critical and can be applied. The waves are modelised with rays carrying acoustical energy. This energy decrease with the reflection of the rays on the walls of the room. The reason of this phenomenon is the absorption of the walls.

The problem is this theory needs a very high power of calculation and that is why the theory of Sabine is often chosen because it is easier.

The theory of Sabine

Description of the theory

This theory uses the hypothesis of the diffuse field, the acoustical field is homogeneous and isotropic. In order to obtain this field, the room has to be enough reverberating and the frequencies have to be high enough to avoid the effects of predominating modes.

The variation of the acoustical energy E in the room can be written as :

$\frac{{dE}}{{dt}} = W_s - W_{abs}$

Where $W s$ and $W a b s$ are respectively the power generated by the acoustical source and the power absorbed by the walls.

The power absorbed is related to the voluminal energy in the room e :

$W_s = \frac{{ec_0 }}{4}a$

Where a is the equivalent absorption area defined by the sum of the product of the absorption coefficient and the area of each material in the room :

$a = \sum\limits_i {\alpha _i S_i }$

The final equation is : $V\frac{{de}}{{dt}} = W_s - \frac{{ec_0 }}{4}a$

The level of stationary energy is : $e_{sat} = 4\frac{{W_s }}{{ac_0 }}$

Reverberation time

With this theory described, the reverberation time can be defined. It is the time for the level of energy to decrease of 60 dB. It depends on the volume of the room V and the equivalent absorption area a :

$T_{60} = \frac{{0.16V}}{a}$ Sabine formula

This reverberation time is the fundamental parameter in room acoustics and depends trough the equivalent absorption area and the absorption coefficients on the frequency. It is used for several measurement :

Measurement of an absorption coefficient of a material
Measurement of the power of a source
Measurement of the transmission of a wall

← Fundamentals of Acoustics · Fundamentals of Psychoacoustics →

Due to the famous principle enounced by Gustav Theodor Fechner, the sensation of perception doesn’t follow a linear law, but a logarithmic one. The perception of an object temperature, or the sensation of weight, are good examples of this law. This observation legitimates the use of logarithmic scales in the field of acoustics. A 80dB (10-4 W/m²) sound seems to be twice as loud as a 70 dB (10-5 W/m²) sound, although there is a factor 10 between the two acoustic powers. This is quite a naïve law, but it led to a new way of thinking acoustics, by trying to describe the auditive sensations. That’s the aim of psychoacoustics. By now, as the neurophysiologic mechanisms of human hearing haven’t been successfully modelled, the only way of dealing with psychoacoustics is by finding metrics that best describe the different aspects of sound.

Perception of sound

The study of sound perception is limited by the complexity of the human ear mechanisms. The figure below represents the domain of perception and the thresholds of pain and listening. The pain threshold is not frequency-dependent (around 120 dB in the audible bandwidth). At the opposite side, the listening threshold, as all the equal loudness curves, is frequency-dependent.

Phons and sones

Phons

Two sounds of equal intensity do not have the same loudness, because of the frequency sensibility of human ear. A 80 dB sound at 100 Hz is not as loud as a 80 dB at 3 kHz. A new unit, the phon, is used to describe the loudness of a harmonic sound. X phons means “as loud as X dB at 1000 Hz”. Another tool is used : the equal loudness curves, a.k.a. Fletcher curves.

Sones

Another scale currently used is the sone, based upon the rule of thumb for loudness. This rule states that the sound must be increased in intensity by a factor 10 to be perceived as twice as loud. In decibel (or phon) scale, it corresponds to a 10 dB (or phons) increase. The sone scale’s purpose is to translate those scales into a linear one.

$log(S) = 0,03(L p h - 40)$

Where S is the sone level, and $L p h$ the phon level. The conversion table is as followed :

Metrics

We will now present five psychoacoustics parameters to provide a way to predict the subjective human sensation.

dB A

The measurement of noise perception with the sone or phon scale is not easy. A widely used measurement method is a weighting of the sound pressure level, according to frequency repartition. For each frequency of the density spectrum, a level correction is made. Different kinds of weightings (dB A, dB B, dB C) exist in order to approximate the human ear at different sound intensities, but the most commonly used is the dB A filter. Its curve is made to match the ear equal loudness curve for 40 phons, and as a consequence it’s a good approximation of the phon scale.

Example : for a harmonic 40 dB sound, at 200 Hz, the correction is -10 dB, so this sound is 30 dB A.

Loudness

It measures the sound strength. Loudness can be measured in sone, and is a dominant metric in psychoacoustics.

Tonality

As the human hear is very sensible to the pure harmonic sounds, this metric is a very important one. It measures the number of pure tones in the noise spectrum. A broadwidth sound has a very low tonality, for example.

Roughness

It describes the human perception of temporal variations of sounds. This metric is measured in asper.

Sharpness

Sharpness is linked to the spectral characteristics of the sound. A high-frequency signal has a high value of sharpness. This metric is measured in acum.

Blocking effect

A sinusoidal sound can be masked by a white noise in a narrowing bandwidth. A white noise is a random signal with a flat power spectral density. In other words, the signal's power spectral density has equal power in any band, at any centre frequency, having a given bandwidth. If the intensity of the white noise is high enough, the sinusoidal sound will not be heard. For example, in a noisy environment (in the street, in a workshop), a great effort has to be made in order to distinguish someone’s talking.

← Fundamentals of Room Acoustics · Sound Speed →

The speed of sound c (from Latin celeritas, "velocity") varies depending on the medium through which the sound waves pass. It is usually quoted in describing properties of substances (e.g. see the article on sodium). In conventional use and in scientific literature sound velocity v is the same as sound speed c. Sound velocity c or velocity of sound should not be confused with sound particle velocity v, which is the velocity of the individual particles.

More commonly the term refers to the speed of sound in air. The speed varies depending on atmospheric conditions; the most important factor is the temperature. The humidity has very little effect on the speed of sound, while the static sound pressure (air pressure) has none. Sound travels slower with an increased altitude (elevation if you are on solid earth), primarily as a result of temperature and humidity changes. An approximate speed (in metres per second) can be calculated from:

$c_{\mathrm{air}} = (331{.}5 + (0{.}6 \cdot \theta)) \ \mathrm{m/s}\,$

where $\theta\,$ (theta) is the temperature in degrees Celsius.

Details

A more accurate expression for the speed of sound is

$c = \sqrt {\kappa \cdot R\cdot T}$

where

R (287.05 J/(kg·K) for air) is the gas constant for air: the universal gas constant $R$ , which units of J/(mol·K), is divided by the molar mass of air, as is common practice in aerodynamics)
κ (kappa) is the adiabatic index (1.402 for air), sometimes noted γ
T is the absolute temperature in kelvins.

In the standard atmosphere :

T₀ is 273.15 K (= 0 °C = 32 °F), giving a value of 331.5 m/s (= 1087.6 ft/s = 1193 km/h = 741.5 mph = 643.9 knots).
T₂₀ is 293.15 K (= 20 °C = 68 °F), giving a value of 343.4 m/s (= 1126.6 ft/s = 1236 km/h = 768.2 mph = 667.1 knots).
T₂₅ is 298.15 K (= 25 °C = 77 °F), giving a value of 346.3 m/s (= 1136.2 ft/s = 1246 km/h = 774.7 mph = 672.7 knots).

In fact, assuming an ideal gas, the speed of sound c depends on temperature only, not on the pressure. Air is almost an ideal gas. The temperature of the air varies with altitude, giving the following variations in the speed of sound using the standard atmosphere - actual conditions may vary. Any qualification of the speed of sound being "at sea level" is also irrelevant.

Altitude	Temperature	m/s	km/h	mph	knots
Sea level (?)	15 °C (59 °F)	340	1225	761	661
11,000 m–20,000 m (Cruising altitude of commercial jets, and first supersonic flight)	-57 °C (-70 °F)	295	1062	660	573
29,000 m (Flight of X-43A)	-48 °C (-53 °F)	301	1083	673	585

In a Non-Dispersive Medium – Sound speed is independent of frequency, therefore the speed of energy transport and sound propagation are the same. For audio sound range air is a non-dispersive medium. We should also note that air contains CO2 which is a dispersive medium and it introduces dispersion to air at ultrasound frequencies (28KHz).
In a Dispersive Medium – Sound speed is a function of frequency. The spatial and temporal distribution of a propagating disturbance will continually change. Each frequency component propagates at each its own phase speed, while the energy of the disturbance propagates at the group velocity. Water is an example of a dispersive medium.

In general, the speed of sound c is given by

$c = \sqrt{\frac{C}{\rho}}$

where

C is a coefficient of stiffness

ρ

is the density

Thus the speed of sound increases with the stiffness of the material, and decreases with the density.

In a fluid the only non-zero stiffness is to volumetric deformation (a fluid does not sustain shear forces).

Hence the speed of sound in a fluid is given by

$c = \sqrt {\frac{K}{\rho}}$

where

K is the adiabatic bulk modulus

For a gas, K is approximately given by

$K=\kappa \cdot p$

where

κ is the adiabatic index, sometimes called γ.

p is the pressure.

Thus, for a gas the speed of sound can be calculated using:

$c = \sqrt {{\kappa \cdot p}\over\rho}$

which using the ideal gas law is identical to:

$c = \sqrt {\kappa \cdot R\cdot T}$

(Newton famously considered the speed of sound before most of the development of thermodynamics and so incorrectly used isothermal calculations instead of adiabatic. His result was missing the factor of κ but was otherwise correct.)

In a solid, there is a non-zero stiffness both for volumetric and shear deformations. Hence, in a solid it is possible to generate sound waves with different velocities dependent on the deformation mode.

In a solid rod (with thickness much smaller than the wavelength) the speed of sound is given by:

$c = \sqrt{\frac{E}{\rho}}$

where

E is Young's modulus

ρ

(rho) is density

Thus in steel the speed of sound is approximately 5100 m/s.

In a solid with lateral dimensions much larger than the wavelength, the sound velocity is higher. It is found by replacing Young's modulus with the plane wave modulus, which can be expressed in terms of the Young's modulus and Poisson's ratio as:

$M = E \frac{1-\nu}{1-\nu-2\nu^2}$

For air, see density of air.

The speed of sound in water is of interest to those mapping the ocean floor. In saltwater, sound travels at about 1500 m/s and in freshwater 1435 m/s. These speeds vary due to pressure, depth, temperature, salinity and other factors.

For general equations of state, if classical mechanics is used, the speed of sound $c$ is given by

$c^2=\frac{\partial p}{\partial\rho}$

where differentiation is taken with respect to adiabatic change.

If relativistic effects are important, the speed of sound $S$ is given by:

$S^2=c^2 \left. \frac{\partial p}{\partial e} \right|_{\rm adiabatic}$

(Note that $e= \rho (c^2+e^C) \,$ is the relativisic internal energy density).

This formula differs from the classical case in that $ρ$ has been replaced by $e/c^2 \,$ .

Speed of sound in air

Impact of temperature
θ in °C	c in m/s	ρ in kg/m³	Z in N·s/m³
−10	325.4	1.341	436.5
−5	328.5	1.316	432.4
0	331.5	1.293	428.3
+5	334.5	1.269	424.5
+10	337.5	1.247	420.7
+15	340.5	1.225	417.0
+20	343.4	1.204	413.5
+25	346.3	1.184	410.0
+30	349.2	1.164	406.6

Mach number is the ratio of the object's speed to the speed of sound in air (medium).

Sound in solids

In solids, the velocity of sound depends on density of the material, not its temperature. Solid materials, such as steel, conduct sound much faster than air.

Experimental methods

In air a range of different methods exist for the measurement of sound.

Single-shot timing methods

The simplest concept is the measurement made using two microphones and a fast recording device such as a digital storage scope. This method uses the following idea.

If a sound source and two microphones are arranged in a straight line, with the sound source at one end, then the following can be measured:

The distance between the microphones (x)
The time delay between the signal reaching the different microphones (t)

Then v = x/t

An older method is to create a sound at one end of a field with an object that can be seen to move when it creates the sound. When the observer sees the sound-creating device act they start a stopwatch and when the observer hears the sound they stop their stopwatch. Again using v = x/t you can calculate the speed of sound. A separation of at least 200 m between the two experimental parties is required for good results with this method.

Other methods

In these methods the time measurement has been replaced by a measurement of the inverse of time (frequency).

Kundt's tube is an example of an experiment which can be used to measure the speed of sound in a small volume, it has the advantage of being able to measure the speed of sound in any gas. This method uses a powder to make the nodes and antinodes visible to the human eye. This is an example of a compact experimental setup.

A tuning fork can be held near the mouth of a long pipe which is dipping into a barrel of water, in this system it is the case that the pipe can be brought to resonance if the length of the air column in the pipe is equal to ( {1+2n}/λ ) where n is an integer. As the antinodal point for the pipe at the open end is slightly outside the mouth of the pipe it is best to find two or more points of resonance and then measure half a wavelength between these.

Here it is the case that v = fλ

External links

← Fundamentals of Psychoacoustics · Filter Design and Implementation →

Introduction

Acoustic filters, or mufflers, are used in a number of applications requiring the suppression or attenuation of sound. Although the idea might not be familiar to many people, acoustic mufflers make everyday life much more pleasant. Many common appliances, such as refrigerators and air conditioners, use acoustic mufflers to produce a minimal working noise. The application of acoustic mufflers is mostly directed to machine components or areas where there is a large amount of radiated sound such as high pressure exhaust pipes, gas turbines, and rotary pumps.

Although there are a number of applications for acoustic mufflers, there are really only two main types which are used. These are absorptive and reactive mufflers. Absorptive mufflers incorporate sound absorbing materials to attenuate the radiated energy in gas flow. Reactive mufflers use a series of complex passages to maximize sound attenuation while meeting set specifications, such as pressure drop, volume flow, etc. Many of the more complex mufflers today incorporate both methods to optimize sound attenuation and provide realistic specifications.

In order to fully understand how acoustic filters attenuate radiated sound, it is first necessary to briefly cover some basic background topics. For more information on wave theory and other material necessary to study acoustic filters please refer to the references below.

Basic wave theory

Although not fundamentally difficult to understand, there are a number of alternate techniques used to analyze wave motion which could seem overwhelming to a novice at first. Therefore, only 1-D wave motion will be analyzed to keep most of the mathematics as simple as possible. This analysis is valid, with not much error, for the majority of pipes and enclosures encountered in practice.

Plane-wave pressure distribution in pipes

The most important equation used is the wave equation in 1-D form (See [1],[2], 1-D Wave Equation, for information).

Therefore, it is reasonable to suggest, if plane waves are propagating, that the pressure distribution in a pipe is given by:

$\mathbf{p}=\mathbf{Pi}e^{j[\omega t-kx]}+\mathbf{Pr}e^{j[\omega t+kx]}$

where Pi and Pr are incident and reflected wave amplitudes respectively. Also note that bold notation is used to indicate the possibility of complex terms. The first term represents a wave travelling in the +x direction and the second term, -x direction.

Since acoustic filters or mufflers typically attenuate the radiated sound power as much as possible, it is logical to assume that if we can find a way to maximize the ratio between reflected and incident wave amplitude then we will effectively attenuated the radiated noise at certain frequencies. This ratio is called the reflection coefficient and is given by:

$\mathbf{R}=\left( \frac{\mathbf{Pr}}{\mathbf{Pi}} \right)$

It is important to point out that wave reflection only occurs when the impedance of a pipe changes. It is possible to match the end impedance of a pipe with the characteristic impedance of a pipe to get no wave reflection. For more information see [1] or [2].

Although the reflection coefficient isn't very useful in its current form since we want a relation describing sound power, a more useful form can be derived by recognizing that the power intensity coefficient is simply the magnitude of reflection coefficient square [1]:

$R_{\pi}=\left|\mathbf{R}\right|^2$

As one would expect, the power reflection coefficient must be less than or equal to one. Therefore, it is useful to define the transmission coefficient as:

$T_{\pi}=\left(1-R_{\pi}\right)$

which is the amount of power transmitted. This relation comes directly from conservation of energy. When talking about the performance of mufflers, typically the power transmission coefficient is specified.

Basic filter design

For simple filters, a long wavelength approximation can be made to make the analysis of the system easier. When this assumption is valid (e.g. low frequencies) the components of the system behave as lumped acoustical elements. Equations relating the various properties are easily derived under these circumstances.

The following derivations assume long wavelength. Practical applications for most conditions are given later.

Low-pass filter

Tpi for Low-Pass Filter

These are devices that attenuate the radiated sound power at higher frequencies. This means the power transmission coefficient is approximately 1 across the band pass at low frequencies(see figure to right).

This is equivalent to an expansion in a pipe, with the volume of gas located in the expansion having an acoustic compliance (see figure to right). Continuity of acoustic impedance (see Java Applet at: Acoustic Impedance Visualization) at the junction, see [1], gives a power transmission coefficient of:

$T_{\pi}=\left(\frac{1}{1+\left(\frac{S_{1}-S}{2S}\right)kL}\right)$

where k is the wavenumber (see [Wave Properties]), L & $S 1$ are length and area of expansion respectively, and S is the area of the pipe.

The cut-off frequency is given by:

$f_{c}=\left(\frac{Sc}{\pi L(S_{1}-S)}\right)$

High-pass filter

Tpi for High-Pass Filter

These are devices that attenuate the radiated sound power at lower frequencies. Like before, this means the power transmission coefficient is approximately 1 across the band pass at high frequencies (see figure to right).

This is equivalent to a short side brach (see figure to right) with a radius and length much smaller than the wavelength (lumped element assumption). This side branch acts like an acoustic mass and applies a different acoustic impedance to the system than the low-pass filter. Again using continuity of acoustic impedance at the junction yields a power transmission coefficient of the form [1]:

$T_{\pi}=\left(\frac{1}{1+\left(\frac{\pi a^2}{2SLk}\right)^2}\right)$

where a and L are the area and effective length of the small tube, and S is the area of the pipe.

The cut-off frequency is given by:

$f_{c}=\left(\frac{ca^2}{2SL}\right)$

Band-stop filter

Tpi for Band-Stop Filter

These are devices that attenuate the radiated sound power over a certain frequency range (see figure to right). Like before, the power transmission coefficient is approximately 1 in the band pass region.

Since the band-stop filter is essentially a cross between a low and high pass filter, one might expect to create one by using a combination of both techniques. This is true in that the combination of a lumped acoustic mass and compliance gives a band-stop filter. This can be realized as a helmholtz resonator (see figure to right). Again, since the impedance of the helmholtz resonator can be easily determined, continuity of acoustic impedance at the junction can give the power transmission coefficient as [1]:

$T_{\pi}=\left(\frac{1}{1+\left(\frac{c/2S}{\omega L/S_{b}-c^2/\omega V}\right)^2}\right)$

where $S b$ is the area of the neck, L is the effective length of the neck, V is the volume of the helmholtz resonator, and S is the area of the pipe. It is interesting to note that the power transmission coefficient is zero when the frequency is that of the resonance frequency of the helmholtz. This can be explained by the fact that at resonance the volume velocity in the neck is large with a phase such that all the incident wave is reflected back to the source [1].

The zero power transmission coefficient location is given by:

$f_{c}=\left(\frac{c}{2\pi}\right)\sqrt{\left(\frac{S_{b}}{LV}\right)}$

This frequency value has powerful implications. If a system has the majority of noise at one frequency component, the system can be "tuned" using the above equation, with a helmholtz resonator, to perfectly attenuate any transmitted power (see examples below).

Helmholtz Resonator as a Muffler, f = 60 Hz

Helmholtz Resonator as a Muffler, f = fc

Design

If the long wavelength assumption is valid, typically a combination of methods described above are used to design a filter. A specific design procedure is outlined for a helmholtz resonator, and other basic filters follow a similar procedure (see [1]).

Two main metrics need to be identified when designing a helmholtz resonator [3]:

Resonance frequency desired: $f_{c}=\frac{c}{2\pi}\frac{\sqrt{C_{o}}}{V}$ where $C_{o}=\frac{S}{L}$ .
- Transmission loss: $\frac{\sqrt{C_{o}V}}{2S}=const$ based on TL level. This constant is found from a TL graph (see HR pp. 6).

This will result in two equations with two unknowns which can be solved for the unknown dimensions of the helmholtz resonator. It is important to note that flow velocities degrade the amount of transmission loss at resonance and tend to move the resonance location upwards [3].

In many situations, the long wavelength approximation is not valid and alternative methods must be examined. These are much more mathematically rigorous and require a complete understanding acoustics involved. Although the mathematics involved are not shown, common filters used are given in the section that follows.

Actual filter design

As explained previously, there are two main types of filters used in practice: absorptive and reactive. The benefits and drawback of each will be briefly explained, along with their relative applications (see [Absorptive Mufflers].

Absorptive

These are mufflers which incorporate sound absorbing materials to transform acoustic energy into heat. Unlike reactive mufflers which use destructive interference to minimize radiated sound power, absorptive mufflers are typically straight through pipes lined with multiple layers of absorptive materials to reduce radiated sound power. The most important property of absorptive mufflers is the attenuation constant. Higher attenuation constants lead to more energy dissipation and lower radiated sound power.

Advantages of Absorptive Mufflers [3]:

(1) - High amount of absorption at larger frequencies.

(2) - Good for applications involving broadband (constant across the spectrum) and narrowband (see [1]) noise.

(3) - Reduced amount of back pressure compared to reactive mufflers.

Disadvantages of Absorptive Mufflers [3]:

(1) - Poor performance at low frequencies.

(2) - Material can degrade under certain circumstances (high heat, etc).

Examples

Absorptive Muffler

There are a number of applications for absorptive mufflers. The most well known application is in race cars, where engine performance is desired. Absorptive mufflers don't create a large amount of back pressure (as in reactive mufflers) to attenuate the sound, which leads to higher muffler performance. It should be noted however, that the radiate sound is much higher. Other applications include plenum chambers (large chambers lined with absorptive materials, see picture below), lined ducts, and ventilation systems.

Reactive

Reactive mufflers use a number of complex passages (or lumped elements) to reduce the amount of acoustic energy transmitted. This is accomplished by a change in impedance at the intersections, which gives rise to reflected waves (and effectively reduces the amount of transmitted acoustic energy). Since the amount of energy transmitted is minimized, the reflected energy back to the source is quite high. This can actually degrade the performance of engines and other sources. Opposite to absorptive mufflers, which dissipate the acoustic energy, reactive mufflers keep the energy contained within the system. See [Reactive Mufflers] for more information.

Advantages of Reactive Mufflers [3]:

(1) - High performance at low frequencies.

(2) - Typically give high insertion loss, IL, for stationary tones.

(3) - Useful in harsh conditions.

Disadvantages of Reactive Mufflers [3]:

(1) - Poor performance at high frequencies.

(2) - Not desirable characteristics for broadband noise.

Examples

Reflective Muffler

Reactive mufflers are the most widely used mufflers in combustion engines[1]. Reactive mufflers are very efficient in low frequency applications (especially since simple lumped element analysis can be applied). Other application areas include: harsh environments (high temperature/velocity engines, turbines, etc), specific frequency attenuation (using a helmholtz like device, a specific frequency can be toned to give total attenuation of radiated sound power), and a need for low radiated sound power (car mufflers, air conditioners, etc).

Performance

There are 3 main metrics used to describe the performance of mufflers; Noise Reduction, Insertion Loss, and Transmission Loss. Typically when designing a muffler, 1 or 2 of these metrics is given as a desired value.

Noise Reduction (NR)

Defined as the difference between sound pressure levels on the source and receiver side. It is essentially the amount of sound power reduced between the location of the source and termination of the muffler system (it doesn't have to be the termination, but it is the most common location) [3].

$NR = \left(L_{p1}-L_{p2}\right)$

where $L p 1$ and $L p 2$ is sound pressure levels at source and receiver respectively. Although NR is easy to measure, pressure typically varies at source side due to standing waves [3].

Insertion Loss (IL)

Defined as difference of sound pressure level at the receiver with and without sound attenuating barriers. This can be realized, in a car muffler, as the difference in radiated sound power with just a straight pipe to that with an expansion chamber located in the pipe. Since the expansion chamber will attenuate some of the radiate sound power, the pressure at the receiver with sound attenuating barriers will be less. Therefore, a higher insertion loss is desired [3].

$IL = \left(L_{p,without}-L_{p,with}\right)$

where $L p, w i t h o u t$ and $L p, w i t h$ are pressure levels at receiver without and with a muffler system respectively. Main problem with measuring IL is that the barrier or sound attenuating system needs to be removed without changing the source [3].

Transmission Loss (TL)

Defined as the difference between the sound power level of the incident wave to the muffler system and the transmitted sound power. For further information see [Transmission Loss] [3].

$TL = 10log\left(\frac{1}{\tau}\right)$ with $\tau =\left(\frac{I_{t}}{I_{i}} \right)$

where $I t$ and $I i$ are the transmitted and incident wave power respectively. From this expression, it is obvious the problem with measure TL is decomposing the sound field into incident and transmitted waves which can be difficult to do for complex systems (analytically).

Examples

(1) - For a plenum chamber (see figure below):

$TL = -10log\left(S\left(\frac{cos\theta}{2\pi d^2}+\frac{1-\alpha}{\alpha S_{w}}\right)\right)$ in dB

where $α$ is average absorption coefficient.

Plenum Chamber

Transmission Loss vs. Theta

(2) - For an expansion (see figure below):

$NR = 10log\left[ \frac{1}{2}\left| e^{-ikx_{s}}+\left( \frac{1-S}{1+S} \right)e^{ikx_{s}} \right|^2\left( 1+S \right)^2 \right]$

$IL = 10log\left[ \frac{\left( 1+S \right)^2}{4} \right]$

$TL = 10log\left[ \frac{\left( 1+S \right)^2}{4S} \right]$

where $S=\left( \frac{A_{2}}{A_{1}} \right)$

Expansion in Infinite Pipe

NR, IL, & TL for Expansion

(3) - For a helmholtz resonator (see figure below):

$TL = 10log\left[ 1+\left( \frac{\left( \frac{c}{2S_{b}} \right)}{\omega LS - \left( \frac{c^2}{\omega V} \right)} \right)^2 \right]$ in dB

Helmholtz Resonator

TL for Helmholtz Resonator

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References

Fundamentals of Acoustics; Kinsler et al, John Wiley & Sons, 2000
Acoustics; Pierce, Acoustical Society of America, 1989
- ME 413 Noise Control, Dr. Mongeau, Purdue University

← Sound Speed · Flow-induced Oscillations of a Helmholtz Resonator →

Introduction

The importance of flow excited acoustic resonance lies in the large number of applications in which it occurs. Sound production in organ pipes, compressors, transonic wind tunnels, and open sunroofs are only a few examples of the many applications in which flow excited resonance of Helmholtz resonators can be found.[4] An instability of the fluid motion coupled with an acoustic resonance of the cavity produce large pressure fluctuations that are felt as increased sound pressure levels.

Passengers of road vehicles with open sunroofs often experience discomfort, fatigue, and dizziness from self-sustained oscillations inside the car cabin. This phenomenon is caused by the coupling of acoustic and hydrodynamic flow inside a cavity which creates strong pressure oscillations in the passenger compartment in the 10 to 50 Hz frequency range. Some effects experienced by vehicles with open sunroofs when buffeting include: dizziness, temporary hearing reduction, discomfort, driver fatigue, and in extreme cases nausea. The importance of reducing interior noise levels inside the car cabin relies primarily in reducing driver fatigue and improving sound transmission from entertainment and communication devices.

This Wikibook page aims to theoretically and graphically explain the mechanisms involved in the flow-excited acoustic resonance of Helmholtz resonators. The interaction between fluid motion and acoustic resonance will be explained to provide a thorough explanation of the behavior of self-oscillatory Helmholtz resonator systems. As an application example, a description of the mechanisms involved in sunroof buffeting phenomena will be developed at the end of the page.

Feedback loop analysis

As mentioned before, the self-sustained oscillations of a Helmholtz resonator in many cases is a continuous interaction of hydrodynamic and acoustic mechanisms. In the frequency domain, the flow excitation and the acoustic behavior can be represented as transfer functions. The flow can be decomposed into two volume velocities.

qr: flow associated with acoustic response of cavity

qo: flow associated with excitation

Acoustical characteristics of the resonator

Lumped parameter model

The lumped parameter model of a Helmholtz resonator consists of a rigid-walled volume open to the environment through a small opening at one end. The dimensions of the resonator in this model are much less than the acoustic wavelength, in this way allowing us to model the system as a lumped system.

Figure 2 shows a sketch of a Helmholtz resonator on the left, the mechanical analog on the middle section, and the electric-circuit analog on the right hand side. As shown in the Helmholtz resonator drawing, the air mass flowing through an inflow of volume velocity includes the mass inside the neck (Mo) and an end-correction mass (Mend). Viscous losses at the edges of the neck length are included as well as the radiation resistance of the tube. The electric-circuit analog shows the resonator modeled as a forced harmonic oscillator. [1] [2][3]

Figure 2

V: cavity volume

$ρ$ : ambient density

c: speed of sound

S: cross-section area of orifice

K: stiffness

$M a$ : acoustic mass

$C a$ : acoustic compliance

The equivalent stiffness K is related to the potential energy of the flow compressed inside the cavity. For a rigid wall cavity it is approximately:

$K =(\frac{\rho c^2}{V})S^2$

The equation that describes the Helmholtz resonator is the following:

$S \hat{P}_e =\frac{\hat{q}_e}{j\omega S}(-\omega ^2 M + j\omega R + K)$

$\hat{P}_e$ : excitation pressure

M: total mass (mass inside neck Mo plus end correction, Mend)

R: total resistance (radiation loss plus viscous loss)

From the electrical-circuit we know the following:

$M_a = \frac{L \rho}{S}$ $C_a = \frac{\pi V}{\rho c^2}$ $L ' = \ L + \ 1.7 \ re$

The main cavity resonance parameters are resonance frequency and quality factor which can be estimated using the parameters explained above (assuming free field radiation, no viscous losses and leaks, and negligible wall compliance effects)

$\omega_r^2 = \frac{1}{M_a C_a}$ $f_r = c 2 \pi \sqrt{\frac{S}{L' V}}$

The sharpness of the resonance peak is measured by the quality factor Q of the Helmholtz resonator as follows:

$Q = 2 \pi \sqrt{V (\frac{L'} {S})^3}$

$f r$ : resonance frequency in Hz

$ω r$ : resonance frequency in radians

L: length of neck

L': corrected length of neck

From the equations above, the following can be deduced:

The greater the volume of the resonator, the lower the resonance frequencies.
If the length of the neck is increased, the resonance frequency decreases.

Production of self-sustained oscillations

The acoustic field interacts with the unstable hydrodynamic flow above the open section of the cavity, where the grazing flow is continuous. The flow in this section separates from the wall at a point where the acoustic and hydrodynamic flows are strongly coupled. [5]

The separation of the boundary layer at the leading edge of the cavity (front part of opening from incoming flow) produces strong vortices in the main stream. As observed in Figure 3, a shear layer crosses the cavity orifice and vortices start to form due to instabilities in the layer at the leading edge.

Figure 3

From Figure 3, L is the length of the inner cavity region, d denotes the diameter or length of the cavity length, D represents the height of the cavity, and $δ$ describes the gradient length in the grazing velocity profile (boundary layer thickness).

The velocity in this region is characterized to be unsteady and the perturbations in this region will lead to self-sustained oscillations inside the cavity. Vortices will continually form in the opening region due to the instability of the shear layer at the leading edge of the opening.

Applications to Sunroof Buffeting

How are vortices formed during buffeting?

In order to understand the generation and convection of vortices from the shear layer along the sunroof opening, the animation below has been developed. At a certain range of flow velocities, self-sustained oscillations inside the open cavity (sunroof) will be predominant. During this period of time, vortices are shed at the trailing edge of the opening and continue to be convected along the length of the cavity opening as pressure inside the cabin decreases and increases. Flow visualization experimentation is one method that helps obtain a qualitative understanding of vortex formation and conduction.

The animation below, shows in the middle, a side view of a car cabin with the sunroof open. As the air starts to flow at a certain mean velocity Uo, air mass will enter and leave the cabin as the pressure decreases and increases again. At the right hand side of the animation, a legend shows a range of colors to determine the pressure magnitude inside the car cabin. At the top of the animation, a plot of circulation and acoustic cavity pressure versus time for one period of oscillation is shown. The symbol x moving along the acoustic cavity pressure plot is synchronized with pressure fluctuations inside the car cabin and with the legend on the right. For example, whenever the x symbol is located at the point where t=0 (when the acoustic cavity pressure is minimum) the color of the car cabin will match that of the minimum pressure in the legend (blue).

The perturbations in the shear layer propagate with a velocity of the order of 1/2Uo which is half the mean inflow velocity. [5] After the pressure inside the cavity reaches a minimum (blue color) the air mass position in the neck of the cavity reaches its maximum outward position. At this point, a vortex is shed at the leading edge of the sunroof opening (front part of sunroof in the direction of inflow velocity). As the pressure inside the cavity increases (progressively to red color) and the air mass at the cavity entrance is moved inwards, the vortex is displaced into the neck of the cavity. The maximum downward displacement of the vortex is achieved when the pressure inside the cabin is also maximum and the air mass in the neck of the Helmholtz resonator (sunroof opening) reaches its maximum downward displacement. For the rest of the remaining half cycle, the pressure cavity falls and the air below the neck of the resonator is moved upwards. The vortex continues displacing towards the downstream edge of the sunroof where it is convected upwards and outside the neck of the resonator. At this point the air below the neck reaches its maximum upwards displacement.[4] And the process starts once again.

How to identify buffeting

Flow induced tests performed over a range of flow velocities are helpful to determine the change in sound pressure levels (SPL) inside the car cabin as inflow velocity is increased. The following animation shows typical auto spectra results from a car cabin with the sunroof open at various inflow velocities. At the top right hand corner of the animation, it is possible to see the inflow velocity and resonance frequency corresponding to the plot shown at that instant of time.

It is observed in the animation that the SPL increases gradually with increasing inflow velocity. Initially, the levels are below 80 dB and no major peaks are observed. As velocity is increased, the SPL increases throughout the frequency range until a definite peak is observed around a 100 Hz and 120 dB of amplitude. This is the resonance frequency of the cavity at which buffeting occurs. As it is observed in the animation, as velocity is further increased, the peak decreases and disappears.

In this way, sound pressure level plots versus frequency are helpful in determining increased sound pressure levels inside the car cabin to find ways to minimize them. Some of the methods used to minimize the increased SPL levels achieved by buffeting include: notched deflectors, mass injection, and spoilers.

Useful websites

This link: [1]takes you to the website of EXA Corporation, a developer of PowerFlow for Computational Fluid Dynamics (CFD) analysis.

This link: [2] is a small news article about the current use of(CFD) software to model sunroof buffeting.

This link: [3]is a small industry brochure that shows the current use of CFD for sunroof buffeting.

References

Acoustics: An introduction to its Physical Principles and Applications ; Pierce, Allan D., Acoustical Society of America, 1989.
Prediction and Control of the Interior Pressure Fluctuations in a Flow-excited Helmholtz resonator ; Mongeau, Luc, and Hyungseok Kook., Ray W. Herrick Laboratories, Purdue University, 1997.
Influence of leakage on the flow-induced response of vehicles with open sunroofs ; Mongeau, Luc, and Jin-Seok Hong., Ray W. Herrick Laboratories, Purdue University.
Fluid dynamics of a flow excited resonance, part I: Experiment ; P.A. Nelson, Halliwell and Doak.; 1991.
An Introduction to Acoustics ; Rienstra, S.W., A. Hirschberg., Report IWDE 99-02, Eindhoven University of Technology, 1999.

← Filter Design and Implementation · Active Control →

Introduction

The principle of active control of noise, is to create destructive interferences using a secondary source of noise. Thus, any noise can theoretically disappear. But as we will see in the following sections, only low frequencies noises can be reduced for usual applications, since the amount of secondary sources required increases very quickly with frequency. Moreover, predictable noises are much easier to control than unpredictable ones. The reduction can reach up to 20dB for the best cases. But since good reduction can only be reached for low frequencies, the perception we have of the resulting sound is not necessarily as good as the theoretical reduction. This is due to psychoacoustics considerations, which will be discussed later on.

Fundamentals of active control of noise

Control of a monopole by another monopole

Even for the free space propagation of an acoustic wave created by a punctual source it is difficult to reduce noise in a large area, using active noise control, as we will see in the section.

In the case of an acoustic wave created by a monopolar source, the Helmholtz equation becomes:

Δ p + k 2 p = - j ωρ 0 q

where q is the flow of the noise sources.

The solution for this equation at any M point is:

$p_p (M) = \frac{{j\omega \rho _0 q_p }}{{4\pi }}\frac{{e^{ - jkr_p } }}{{r_p }}$

where the p mark refers to the primary source.

Let us introduce a secondary source in order to perform active control of noise. The acoustic pressure at that same M point is now:

${\rm{p(M) = }}\frac{{{\rm{j}}\omega \rho _{\rm{0}} q_p }}{{4\pi }}\frac{{e^{ - jkr_p } }}{{r_p }} + \frac{{{\rm{j}}\omega \rho _{\rm{0}} q_s }}{{4\pi }}\frac{{e^{ - jkr_s } }}{{r_s }}$

It is now obvious that if we chose $q_s = - q_p \frac{{r_s }}{{r_p }}e^{ - jk(r_p - r_s )}$ there is no more noise at the M point. This is the most simple example of active control of noise. But it is also obvious that if the pressure is zero in M, there is no reason why it should also be zero at any other N point. This solution only allows to reduce noise in one very small area.

However, it is possible to reduce noise in a larger area far from the source, as we will see in this section. In fact the expression for acoustic pressure far from the primary source can be approximated by:

$p(M) = \frac{{j\omega \rho _0 }}{{4\pi }}\frac{{e^{ - jkr_p } }}{{r_p }}(q_p + q_s e^{ - jkD\cos \theta } )$

Control of a monopole by another monopole

As shown in the previous section we can adjust the secondary source in order to get no noise in M. In that case, the acoustic pressure in any other N point of the space remains low if the primary and secondary sources are close enough. More precisely, it is possible to have a pressure close to zero in the whole space if the M point is equally distant from the two sources and if: $D < λ / 6$ where D is the distance between the primary and secondary sources. As we will see later on, it is easier to perform active control of noise with more than on source controlling the primary source, but it is of course much more expensive.

A commonly admitted estimation of the number of secondary sources which are necessary to reduce noise in an R radius sphere, at a frequency f is:

$N = \frac{{36\pi R^2 f^2 }}{{c^2 }}$

This means that if you want no noise in a one meter diameter sphere at a frequency below 340Hz, you will need 30 secondary sources. This is the reason why active control of noise works better at low frequencies.

Active control for waves propagation in ducts and enclosures

This section requires from the reader to know the basis of modal propagation theory, which will not be explained in this article.

Ducts

For an infinite and straight duct with a constant section, the pressure in areas without sources can be written as an infinite sum of propagation modes:

$p(x,y,z,\omega ) = \sum\limits_{n = 1}^N {a_n (\omega )\phi _n (x,y)e^{ - jk_n z} }$

where $φ$ are the eigen functions of the Helmoltz equation and a represent the amplitudes of the modes.

The eigen functions can either be obtained analytically, for some specific shapes of the duct, or numerically. By putting pressure sensors in the duct and using the previous equation, we get a relation between the pressure matrix P (pressure for the various frequencies) and the A matrix of the amplitudes of the modes. Furthermore, for linear sources, there is a relation between the A matrix and the U matrix of the signal sent to the secondary sources: $A s = K U$ and hence: $A = A p + A s = A p + K U$ .

Our purpose is to get: A=0, which means: $A p + K U = 0$ . This is possible every time the rank of the K matrix is bigger than the number of the propagation modes in the duct.

Thus, it is theoretically possible to have no noise in the duct in a very large area not too close from the primary sources if the there are more secondary sources than propagation modes in the duct. Therefore, it is obvious that active noise control is more appropriate for low frequencies. In fact the more the frequency is low, the less propagation modes there will be in the duct. Experiences show that it is in fact possible to reduce the noise from over 60dB.

Enclosures

The principle is rather similar to the one described above, except the resonance phenomenon has a major influence on acoustic pressure in the cavity. In fact, every mode that is not resonant in the considered frequency range can be neglected. In a cavity or enclosure, the number of these modes rise very quickly as frequency rises, so once again, low frequencies are more appropriate. Above a critical frequency, the acoustic field can be considered as diffuse. In that case, active control of noise is still possible, but it is theoretically much more complicated to set up.

Active control and psychoacoustics

As we have seen, it is possible to reduce noise with a finite number of secondary sources. Unfortunately, the perception of sound of our ears does not only depend on the acoustic pressure (or the decibels). In fact, it sometimes happen that even though the number of decibels has been reduced, the perception that we have is not really better than without active control.

Active control systems

Since the noise that has to be reduced can never be predicted exactly, a system for active control of noise requires an auto adaptable algorithm. We have to consider two different ways of setting up the system for active control of noise depending on whether it is possible or not to detect the noise from the primary source before it reaches the secondary sources. If this is possible, a feed forward technique will be used (aircraft engine for example). If not a feed back technique will be preferred.

Feedforward

In the case of a feed forward, two sensors and one secondary source are required. The sensors measure the sound pressure at the primary source (detector) and at the place we want noise to be reduced (control sensor). Furthermore, we should have an idea of what the noise from the primary source will become as he reaches the control sensor. Thus we approximately know what correction should be made, before the sound wave reaches the control sensor (forward). The control sensor will only correct an eventual or residual error. The feedforward technique allows to reduce one specific noise (aircraft engine for example) without reducing every other sound (conversations, …). The main issue for this technique is that the location of the primary source has to be known, and we have to be sure that this sound will be detected beforehand. Therefore portative systems based on feed forward are impossible since it would require having sensors all around the head.

Feedforward System

Feedback

In that case, we do not exactly know where the sound comes from; hence there is only one sensor. The sensor and the secondary source are very close from each other and the correction is done in real time: as soon as the sensor gets the information the signal is treated by a filter which sends the corrected signal to the secondary source. The main issue with feedback is that every noise is reduced and it is even theoretically impossible to have a standard conversation.

Feedback System

Applications

Noise cancelling headphone

Usual headphones become useless when the frequency gets too low. As we have just seen active noise cancelling headphones require the feedback technique since the primary sources can be located all around the head. This active control of noise is not really efficient at high frequencies since it is limited by the Larsen effect. Noise can be reduced up to 30dB at a frequency range between 30Hz and 500Hz.

Active control for cars

Noise reduction inside cars can have a significant impact on the comfort of the driver. There are three major sources of noise in a car: the motor, the contact of tires on the road, and the aerodynamic noise created by the air flow around the car. In this section, active control for each of those sources will be briefly discussed.

Motor noise

This noise is rather predictable since it a consequence of the rotation of the pistons in the motor. Its frequency is not exactly the motor’s rotational speed though. However, the frequency of this noise is in between 20Hz and 200Hz, which means that an active control is theoretically possible. The following pictures show the result of an active control, both for low and high regime.

Low regime

Even though these results show a significant reduction of the acoustic pressure, the perception inside the car is not really better with this active control system, mainly for psychoacoustics reasons which were mentioned above. Moreover such a system is rather expensive and thus are not used in commercial cars.

Tires noise

This noise is created by the contact between the tires and the road. It is a broadband noise which is rather unpredictable since the mechanisms are very complex. For example, the different types of roads can have a significant impact on the resulting noise. Furthermore, there is a cavity around the tires, which generate a resonance phenomenon. The first frequency is usually around 200Hz. Considering the multiple causes for that noise and its unpredictability, even low frequencies become hard to reduce. But since this noise is broadband, reducing low frequencies is not enough to reduce the overall noise. In fact an active control system would mainly be useful in the case of an unfortunate amplification of a specific mode.

Aerodynamic noise

This noise is a consequence of the interaction between the air flow around the car and the different appendixes such as the rear views for example. Once again, it is an unpredictable broadband noise, which makes it difficult to reduce with an active control system. However, this solution can become interesting in the case an annoying predictable resonance would appear.

Active control for aeronautics

The noise of aircraft propellers is highly predictable since the frequency is quite exactly the rotational frequency multiplied by the number of blades. Usually this frequency is around some hundreds of Hz. Hence, an active control system using the feedforward technique provides very satisfying noise reductions. The main issues are the cost and the weigh of such a system. The fan noise on aircraft engines can be reduced in the same manner.

← Flow-induced Oscillations of a Helmholtz Resonator · Rotor Stator Interactions →

Applications in Room Acoustics

Introduction

Acoustic experiments often require to realise measurements in rooms with special characteristics. Two types of rooms can be distinguished: anechoic rooms and reverberation rooms.

Anechoic room

The principle of this room is to simulate a free field. In a free space, the acoustic waves are propagated from the source to infinity. In a room, the reflections of the sound on the walls produce a wave which is propagated in the opposite direction and comes back to the source. In anechoic rooms, the walls are very absorbent in order to eliminate these reflections. The sound seems to die down rapidly. The materials used on the walls are rockwool or glasswool, which are materials that absorb sound in relatively high frequencies. Cavities are dug in the wool so that the large wavelength corresponding to bass frequencies are absorbed too.

Anechoic rooms are used in the following experiments:

Intensimetry: measurement of the acoustic power of a source.

Study of the source directivity.

Reverberation room

The walls of a reverberation room consist of concrete and are covered with reflecting paint. The sound reflects a lot of time on the walls before dying down. It gives the same impression of a sound in a cathedral. Because of all these reflections, a lot of plane waves with different directions of propagation interfere in each point of the room. Considering all the waves is very complicated so the acoustic field is simplified by the diffuse field hypothesis: the field is homogeneous and isotropic. Then the pressure level is uniform in the room.

Several conditions are required for this approximation: The absorption coefficient of the walls must be very low (α<0.2) The room must have geometrical irregularities to avoid nodes of pressure of the resonance modes.

With this hypothesis, the theory of Sabine can be applied. It deals with the reverberation time which is the time required to the sound level to decrease of 60dB. T depends on the volume of the room V, the absorption coefficient αi and the area Si of the different materials in the room :

Reverberation rooms are used in the following experiments:

measurement of the ability of a material to absorb a sound

measurement of the ability of a partition to transmit a sound

Intensimetry

← Interior Sound Transmission · Basic Room Acoustic Treatments →

Introduction

Many people use one or two rooms in their living space as "theatrical" rooms where theater or music room activities commence. It is a common misconception that adding speakers to the room will enhance the quality of the room acoustics. There are other simple things that can be done to increase the room's acoustics to produce sound that is similar to "theater" sound. This site will take you through some simple background knowledge on acoustics and then explain some solutions that will help improve sound quality in a room.

Room sound combinations

The sound you hear in a room is a combination of direct sound and indirect sound. Direct sound will come directly from your speakers while the other sound you hear is reflected off of various objects in the room.

The Direct sound is coming right out of the TV to the listener, as you can see with the heavy black arrow. All of the other sound is reflected off surfaces before they reach the listener.

Good and bad reflected sound

Have you ever listened to speakers outside? You might have noticed that the sound is thin and dull. This occurs because when sound is reflected, it is fuller and louder than it would if it were in an open space. So when sound is reflected, it can add a fullness, or spaciousness. The bad part of reflected sound occurs when the reflections amplify some notes, while cancelling out others, making the sound distorted. It can also affect tonal quality and create an echo-like effect. There are three types of reflected sound, pure reflection, absorption, and diffusion. Each reflection type is important in creating a "theater" type acoustic room.

Reflected sound

Reflected sound waves, good and bad, affect the sound you hear, where it comes from, and the quality of the sound when it gets to you. The bad news when it comes to reflected sound is standing waves.

These waves are created when sound is reflected back and forth between any two parallel surfaces in your room, ceiling and floor or wall to wall.

Standing waves can di

Acoustics/Print version

From Wikibooks, the open-content textbooks collection

Table of contents

Fundamentals

Applications

Applications in Room Acoustics

Applications in Psychoacoustics

Musical Acoustics Applications

Miscellaneous Applications

Introduction

Equation of waves

Helmholtz equation

Acoustic intensity and decibel

Solving the wave equation

Plane waves

Spherical waves

Boundary conditions

Non-absorptive material

Absorptive material

Introduction

The modal theory

The geometry theory

The theory of Sabine

Description of the theory

Reverberation time

Perception of sound

Phons and sones

Phons

Sones

Metrics

dB A

Loudness

Tonality

Roughness

Sharpness

Blocking effect

Details

Speed of sound in air

Sound in solids

Experimental methods

Single-shot timing methods

Other methods

External links

Introduction

Basic wave theory

Plane-wave pressure distribution in pipes

Basic filter design

Low-pass filter

High-pass filter

Band-stop filter

Design

Actual filter design

Absorptive

Examples

Reactive

Examples

Performance

Noise Reduction (NR)

Insertion Loss (IL)

Transmission Loss (TL)

Examples

Links

References

Introduction

Feedback loop analysis

Acoustical characteristics of the resonator

Lumped parameter model

Production of self-sustained oscillations

Applications to Sunroof Buffeting

How are vortices formed during buffeting?

How to identify buffeting

Useful websites

References

Introduction

Fundamentals of active control of noise

Control of a monopole by another monopole

Active control for waves propagation in ducts and enclosures

Ducts

Enclosures

Active control and psychoacoustics