Noise, vibration and harshness
Slide 2 of 63
AIM
• Introduce the basic concepts and importance of
vibration theory to vehicle design
• Consider the role of the designer in vibration
control
• Demonstrate methods for the control of vibration
to help the elimination of noise and harshness
• Indicate methods by which the designer can
control vibration and noise to create an equitable
driving environment
Slide 3 of 63
Basic concepts
• Vibration sources are characterized by their time
and frequency domain characteristics
• Categorized principally as
– Periodic –
• originate from the power unit, ancillaries or transmission
• simplest form of periodic disturbance is harmonic
• In the time domain this is represented by a sinusoid and in
the frequency domain by a single line spectrum
– Random disturbances
• from terrain inputs to wheels
• only statistical representations are possible
• commonly represented by its power spectrum
Slide 4 of 63
Basic concepts
• All mass-elastic systems have natural frequencies
– For linear system these frequencies are constant
• related only to the mass and stiffness distribution
– Non-linear effects require special treatment
• A few of the lower order frequencies are of
interest because the higher ones are more highly
damped.
• For one frequency, a system vibrates in a
particular way, depicted by the relative amplitude
and phase at various locations - mode of vibration
Slide 5 of 63
Reason for vibration analysis
• Lightly damped structures can produce high levels of
vibration from low level sources if frequency
components in the disturbance are close to one of the
system’s natural frequencies.
• This means that well designed and manufactured subsystems, which produce low level disturbing forces, can
still create problems when assembled on a vehicle.
• In order to avoid these problems, at the design stage it
is necessary to model the system accurately and
analyze its response to anticipated disturbances
Slide 6 of 63
Approach to vibration analysis
• Develop a mathematical model of the system
and formulate the equations of motion
• Analyze the free vibration characteristics
(natural frequencies and modes)
• Analyze the forced vibration response to
prescribed disturbances and
• Investigate
methods
for
controlling
undesirable vibration levels if they arise
Slide 7 of 63
Mathematical models
• Provide the basis of all vibration studies at the design stage.
• Represent the dynamics of a system by one or more differential equations.
• Distributed-parameter approach - distributed mass and elasticity of some
very simple components such as uniform shafts and plates by partial
differential equations.
– not generally possible to represent typical engineering systems (which tend to
be more complicated) in this way.
• Lumped-parameter approach - a set of discrete mass, elastic and damping
elements, resulting in one or more ordinary differential equations.
– Masses are concentrated at discrete points and are connected together by
mass less elastic and damping elements.
– The number of elements used dictates the accuracy of the model
– To have just sufficient elements for natural vibration modes and frequencies
while avoiding unnecessary computing effort.
Slide 8 of 63
Mathematical models
Slide 9 of 56
Formulation of System
• Equations of motion
determined by applying
Newton’s second law to
each free-body
• For complicated
geometry, the
equations can be
formulated by energy
methods
Figure from Smith,2002
Slide 10 of 56
System characteristics
• Equation of motion
mx˙˙ + cx˙ + kx = F(t ).
• Characteristics taken at
free vibration [F(t) =0]
• x = X cos (ωnt – φ) – no
damping
• Similar formulation of a
SDOF system to obtain
response as given by
graph
A(w) =
Slide 11 of 63
Multi Degree of freedom
*M+,x˙˙- + *K+,x- = ,0Assuming {x} = {A} est
([M]s2 + [K]){A} = {0}
The non-trivial solution of these is the characteristic
equation (or frequency determinant)
Set of roots
si2 = – λ = –ω2
λi is the i-th eigenvalue and ωi the i-th natural frequency λ*M]{u} = [K]{u}
It is solved using standard procedures used for vibration
problems
Slide 12 of 63
Multi Degree of Freedom
• Viscous Damping
– In (stable) lightly damped systems the frequency
determinant is | [M]s2+ [C]s + [K] | = 0
– For an n-DOF system this produces n complex
conjugate roots having negative real parts
– providing information about the frequency and
damping associated with each mode of vibration
– Damping must be considered in the analysis of
response of the system when one or more
components of a periodic excitation is at or near to
one of the system’s natural frequencies
Slide 13 of 63
Multi Degree of Freedom
• Forced-damped vibration (harmonic)
– Many of the features of harmonic response analysis of
SDOF systems extend to MDOF ones, e.g.:
• When subjected to harmonic excitation an MDOF system
vibrates at the same frequency as the excitation.
• The displacement amplitudes at each of the degrees of
freedom are dependent on the frequency of excitation and
• The dynamic displacement at each DOF lags behind the
excitation
– The frequency response functions (and hence the
frequency responses) are complex if damping is
included in the analysis.
Slide 14 of 63
Multi Degree of Freedom
• Forced damped vibration (random excitation)
– Random excitation arises particularly from terrain inputs
and is important in the analysis and design of vehicle
suspensions.
– Form of excitation is non-deterministic in that its
instantaneous value cannot be predicted at some time in
the future.
– Some properties of random functions which can be
described statistically.
– The mean or mean square value can be determined by
averaging and the frequency content can be determined
from methods based on the Fourier transform (Newlands,
1975).
Slide 15 of 63
Vibration control
• Control at source
• Engine firing and reciprocating unbalance
combine to produce a complex source of
vibration which varies with engine operating
conditions
• Reciprocating unbalance arises at each
cylinder because of the fluctuating inertia
force associated with the mass at each piston
• no such thing as ‘perfect balance’
Slide 16 of 56
Vibration isolation
• k* = k(1 + ηi)
– where k is the dynamic
stiffness and
– η the loss factor
mx˙˙ + k(1 + i)x =
mer2 sin wt = F f (t )
SDOF vibration isolation model and
free-body diagram
Up to 1.4
Figure from Smith,2002
Slide 17 of 56
Tuned absorbers
• useful for reducing vibration
levels in those systems in
which an excitation
frequency is close to or
coincides with a natural
frequency of the system
• The principles of undamped
and damped tuned
absorbers can be
understood by outlining first
the analysis of the damped
absorber
• Undamped absorber as a
special case
Figure from Smith,2002
Slide 18 of 63
Tuned absorbers
For C = 0 ( undamped) [ideally k2-m2w2=0 ie w1= w2]
Figure from Smith,2002
Slide 19 of 56
Tuned absorbers
A1 = K *X1 / F
A2 = K *X2 / F
• Dimensionless numbers
• With damping
– Wider operating range
– Reduced fatigue of absorber spring
Damped
Figure from Smith,2002
Un damped
Slide 20 of 56
Untuned viscous dampers
• Devices consist of
– an inertia (seismic) mass
– which is coupled to the
original system via some
form of damping medium
(usually silicone fluid)
• Not tuned for particular
resonance
• At infinite damping both
masses move together as
one
Models for analyzing the untuned viscous
damper
Response of an untuned viscous damper (m = 1.0)
Figure from Smith,2002
Slide 21 of 63
Engine Isolation
• Fluctuating torque at the crankshaft
• Shaking forces and moments
• additional dynamic inertial loads arising from vehicle maneuvering and
terrain inputs to the wheels
• The primary components of engine vibration at idling are integer multiples
of engine speed
– Idle speeds for four cylinder engines range from 8–20 Hz producing dominant
frequency components in the range from 16–40 Hz.
– Since the primary bending mode of passenger cars can be less than 20 Hz it is
obvious that it is easy to excite body resonance at idle if engine isolation is not
carefully designed
• The problem
– isolate the chassis from the excitations and
– restrain the engine against excessive movement due to the engine torque
• Select appropriate mounts and position them correctly
Slide 22 of 56
Formulation of the vibration equations
• Denoting the position of the ith mount attachment point
relative to the powertrain in
the equilibrium position as r
• ri 0 (in chassis coordinates) and
the translation and rotation of
the powertrain relative to the
chassis axes as rG and Θ, the
deflection of the mount is
given by
dci = ri – ri0
• ρi is the location of the mount
attachment point relative to
the powertrain axes
Figure from Smith,2002
Slide 23 of 63
Formulation of the vibration equations
• The ‘elastic’ potential energy for a set of
mounts typically having orthogonal complex
stiffness components represented by the
diagonal stiffness matrix [km]i is
• For chassis
Slide 24 of 63
Formulation of the vibration equations
Is a non-diagonal matrix
Kinetic energy of the system is
{h}G is the momentum denoted by
Slide 25 of 63
Formulation of the vibration equations
• [M], the symmetrical mass/inertia matrix, is of
the form
• Lagragne’s Equations
Slide 26 of 63
Mount requirements and types
• a low spring rate and high damping during
idling and
• a high spring rate and low damping for high
speeds, manoeuvring and when traversing
rough terrains.
Slide 27 of 63
Mounting types [1/4]
Simple rubber engine mounts
– These are the least costly and least effective forms
of mount and clearly do not meet all the
conflicting requirements.
– They do not provide the high levels of damping
required at idling speeds.
Slide 28 of 63
Mounting types [2/4]
Hydro-elastic mounts
– These generally contain two elastic reservoirs filled with a
hydraulic fluid.
– Some also contain a gas filled reservoir.
– This type of mount exploits the feature of mass-augmented
dynamic damping which is a form of tuned vibration absorber.
– In operation there is relative motion across the damper which
produces flexure of the rubber component and transfer of fluid
between chambers, thereby inducing a change in mount
transmissibility.
– This type of mount has become common in recent years and
some examples are given in the literature (Kim et al., 1992;
Muller et al., 1996).
Slide 29 of 63
Mounting types [3/4]
Semi-active mounts
– The operation of these mounts is dependent on modifying the
magnitude of the forces transmitted through coupling devices.
– They may be implemented via low-bandwidth low power
actuators which are suited to open-loop control.
– Some forms of hydraulic semi-active (adaptive) mount use low
powered actuators to induce changes in mount properties by
modifying the hydraulic parameters within the mount.
– The actuators may then be on–off (adaptive) or continously
variable (semi-active) types.
– Considerable effort has been devoted to this type of technology
in recent years. Some examples are given in Morishita et al.
(1992) and Kim et al. (1993).
Slide 30 of 63
Mounting types [4/4]
Active mounts
– This type of mount requires control of both the magnitude
and direction of the actuator force used to adjust the
coupling device.
– High-speed actuators and sensors require having an
operating bandwidth to match the frequency spectrum of
the disturbance.
– Power consumption is generally high in order to satisfy the
response criteria.
– Active vibration control is typically implemented by closedloop control.
– An example of active engine mount modelling and
performance is given by Miller et al. (1995).
Slide 31 of 56
Engine mounting system for a fourcylinder diesel engine
• It comprises two mass
carrying mounts (one a
hydramount, the other a
hydrabush, both passive
mounts) and two torque
reacting tie bars.
• The hydramount is linked to
the power unit by an
aluminium bridge bracket.
• Both tie bars have a small
bush at the power unit end
and a large bush at the body
end
Figure from Smith,2002
A torque axis engine mounting system
(courtesy of Rover Group Ltd)
Slide 32 of 56
Engine Mounts
• level
isolation
for
passenger cars.
• The modern car designs
have a trend for lighter
car bodies and more
power-intensive engines.
• Such a weight reduction
and increased power
requirements often have
adverse
effects
on
vibratory
behavior,
greatly increasing the
vibration and noise level
From Yunhe Yu, 2001
Slide 33 of 56
Elastomeric mounts
•
•
•
•
•
•
Since 1930s
represented by familiar Voigt model
compact,
cost-effective
and
maintenance free.
Bonded elastomeric mounts are
known to provide more consistent
performance and longer life
Dynamic stiffness of an elastomeric
mount will be greater at higher
frequencies due to damping
Desirable :
– specific nonlinear characteristics to
obtain constant natural frequency in
a broad weight-load range
– use of materials with high internal
damping
– materials with highly amplitudedependent damping and stiffness
From Yunhe Yu, 2001
Slide 34 of 56
Passive hydraulic mounts
•
•
•
•
•
•
•
All hydraulic mounts reported in the
literature are conceptually similar but
differ in detailed structural design
An elastomeric mount capable of
supporting the load and acting as a
piston to pump the liquid into the
button chamber.
Two separate chambers for fluid
transfer.
An orifice or inertia track to generate
damping.
A fluid medium
Sealing between chamber and the
outside.
Decoupler to permit low amplitude
by-pass of the damping.
From Yunhe Yu, 2001
Slide 35 of 56
Passive hydraulic mounts
• Dynamic stiffness of hydraulic
mount with simple orifice or
inertia track is greater than that
of a comparable elastomeric
mount
• Hydraulic mounts greatly increase
damping at low frequencies, they
also
degrade
isolation
performance
at
higher
frequencies
• A de-coupler functions as
amplitude limited floating piston.
• It makes hydraulic mount
amplitude-dependent at low
amplitude displacement
From Yunhe Yu, 2001
Slide 36 of 56
Semi-active and active Engine mounts
• Semi-active control
mechanism
• Hydraulic type
• Electro-rheological (ER) fluids increase
damping
at
resonance and reduce the
transmissibility
for
shock
excitation
• Theory : One degree sky-hook
system and used for increasing
the damping during the low
frequency shock excitation
• Mainly to improve the system
performance at the low
frequency range
From Yunhe Yu, 2001
Slide 37 of 56
Active engine mounts
• Counteracting dynamic force
created by one or more actuators
to suppress transmission of
system disturbance force
• Passive mounts used to support
engine in event of an actuator
failure
• Classical- hydraulic mount, and an
electromagnetic actuator to
isolate high frequency vibration
• Servo-hydraulic
• Piezo-actuators
-high-speed
response, but displacement very
small and requires suitable
mechanism
to
increase
amplitude
From Yunhe Yu, 2001
Model for active elastomeric mount
Model for active hydraulic mount
Slide 38 of 56
Active engine mounts
• The active mount stiffness is
equivalent to the stiffness of
the passive mount at every
frequency
except
where
engine vibration occurs
• At the disturbance frequency,
a tonal controller commands
the mount to be very soft
• Active
hydraulic
mounts
achieve adequate damping at
engine bounce frequency and
have very low dynamic
stiffness at higher frequency
• considered to be the next
generation of engine mounts
From Yunhe Yu, 2001
Dynamic stiffness of an active elastomeric
mount
Dynamic stiffness of hydraulic active mount
Slide 39 of 56
Crankshaft damping
• Torsional
dynamics
of
crankshafts dependent on the
distribution of their mass and
elasticity and the excitations
arising
from
the
torque/cylinder
• Because the torque contains a
number
of
harmonic
components and engine speed
is variable there is a tendency
to excite a large number of
Torsional
resonances
as
illustrated by the waterfall plot
(Torsional amplitude plotted as
a function frequency for a
range of engine speeds)
Figure from Smith,2002
Waterfall plot for a multi-cylinder engine
(courtesy of Simpson International (UK) Ltd)
Slide 40 of 56
Determination of the mass-elastic
model
Modal analysis of the mass-elastic model
Equivalent model based on first mode of the
mass-elastic model
Mass-elastic model and first torsion
mode of a six cylinder engine
Figure from Smith,2002
Slide 41 of 63
Fundamentals of acoustics
General sound propagation
• In an automotive context this is the surrounding air or the vehicle
body structure, giving rise to the term structure-borne sound.
• Produces a propagating (travelling) wave which has a characteristic
velocity c, the velocity of sound in air.
• At some arbitrary point on the path, the air undergoes pressure
fluctuations which are superimposed on the ambient pressure.
• A sound source vibrating at a frequency f, produces sound at this
frequency.
• The distance between pressure peaks is constant and known as the
wavelength λ.
• This is related to c and f by the equation:
Slide 42 of 56
Fundamentals of acoustics
Plane wave propagation
• Wave motion are most easily understood by
considering the propagation of a plane wave
Elastic deformation
Table from Smith,2002
Slide 43 of 56
Plane wave propagation
Specific acoustic impedance,
Z
• The impedance which a
propagating medium
offers to the flow of
acoustic energy is called
the acoustic impedance.
• It is defined as the ratio of
acoustic pressure p to the
velocity of propagation u.
It can be shown
(Reynolds, 1981) that
Acoustic intensity, I
• This is defined as the time
averaged rate of transport
of acoustic energy by a
wave per unit area
• normal to the wavefront.
It is given by Reynolds
(1981)
Slide 44 of 63
Spherical wave propagation, acoustic
near and far fields
• Spherical waves more closely approximate
true source waves, but approximate to plane
waves at large distances from a source.
• It may be shown (Reynolds, 1981) that the
wave equation in spherical coordinates is
Slide 45 of 63
Spherical wave propagation, acoustic
near and far fields
General solution is
• At large distances from the source (kr >> 1 or r >>
λ/2π), z → ρc. [k – wave number = w/c; z – impedance]
• Then pressure and particle velocity are in phase and we
are in what is called the acoustic far field where
spherical wave-fronts approximate to those of plane
waves and pressure and velocity are in phase.
Slide 46 of 63
Spherical wave propagation, acoustic
near and far fields
• At distances close to the source (kr << 1 or r << λ/2π), z
→ iρckr. Then pressure and velocity are 90° out of
phase and we are in what is called the acoustic near
field.
• The transition from near to far field is in reality a
gradual one, but is normally assumed to take place in
the vicinity of λ/2 π.
• For a harmonic wave at 1 kHz (λ ≈ 1 kHz), r = 50 mm;
while at 20 Hz; r = 2.5 m.
• The far field/near field transition has important
implications for microphone positioning in sound level
measurements
Slide 47 of 63
Acoustic quantities expressed in
decibel form
Sound Power, Intensity and Pressure levels
Slide 48 of 63
Acoustic quantities expressed in
decibel form
• When it is noted that the threshold of hearing
corresponds to a sound pressure level of 0 dB
it may be shown (Reynolds, 1981) that for
normal temperature and pressure (101.3 kPa
and 20 °C). LW, LI and Lp are related as follows
Slide 49 of 56
Effects of reflecting surfaces on sound
propagation
• When an incident wave
strikes a reflecting surface
the wave is reflected
backwards towards the
source.
• In the vicinity of the
reflecting surface the
incident and reflected
waves interact to produce
what is known as a
reverberant field
Sound pressure level as a function of
distance from a simple spherical source
Figure from Smith,2002
Slide 50 of 63
Effects of reflecting surfaces on sound
propagation
• If the sound source is next to a hard reflecting surface, four
idealized cases -– whole space radiation – when there are no reflecting surfaces, i.e. the
source is in free space,
– half-space radiation – when the source is positioned at the centre of a
flat hard (reflecting) surface
– quarter space radiation – when the source is positioned at the
intersection of two flat hard surfaces which are perpendicular to one
another
– eighth space radiation – when the source is positioned at the
intersection of three flat perpendicular hard surfaces. In each case
there is an increase in acoustic intensity.
• The effect can be described by the directivity index DI in terms of
the directivity factor Q.
DI = 10 log Q dB
10
Slide 51 of 56
Human response to sound
• The frequency range
extends from 20 Hz to 20
kHz and the SPL extends
from the threshold of
hearing at the lowest
boundary
to
the
threshold of feeling (pain)
at the highest
• Human ear is most
sensitive between 500 Hz
and 5 kHz and is
insensitive to sounds
below 100 Hz
Figure from Smith,2002
The audible range
Slide 52 of 56
Sound measurement
Sound level meters
• The most basic instrument
for sound measurement is a
sound
level
meter
comprising a microphone,
r.m.s. detector with fast and
slow time constants.
• A-weighting network to
enable measurements to be
made which relate to
human response to noise,
leading to so called A
weighted noise levels LpA,
expressed in dB(A).
Figure from Smith,2002
Because of the frequency sensitivity of the
human ear, the A-weighting network has
the form shown in Figure
The A-weighting curve
Slide 53 of 63
Frequency analysers
• Since the frequency spectrum of noise is closely
related to the origins of its production, frequency
analysis is a powerful tool for identifying noise
sources and enables the effectiveness of noise
control measures to be assessed
• Narrow band frequency analyzers are a necessity.
• Simultaneous filtering in multiple narrow band
filters
Slide 54 of 56
Drive-by noise tests (ISO 362, 1981)
• The procedure is to accelerate
the vehicle in a prescribed way
and in a prescribed gear past a
microphone set up at a height
of 1.2 m above a hard
reflecting surface and 7.5 m
from the path of the vehicle
• When the vehicle reaches A
the throttle should be opened
fully and maintained in this
position until the rear of the
vehicle reaches B
Slide 55 of 63
Noise from stationary vehicles
• Vehicles in the vicinity of the exhaust silencer (ISO
5130, 1978)
• Engine running at 75% of the speed at which it
develops maximum power
• For these measurements, the exhaust outlet and
microphone are in the same horizontal plane with the
microphone 500 mm from the exhaust outlet and at an
axis of 45° to it.
• The background noise level is also measured and the
maximum difference between the vehicle noise and
the background noise is then compared with vehicle’s
specified noise level
Slide 56 of 63
Interior Noise
•
•
•
•
•
No legal requirements
Assessed by experienced assesors
Ad Hoc Criterion like Articulation Index used
200 Hz to 16kHz split into 16 bands
SPL is measured in each band
Slide 57 of 63
Interior noise in vehicles
• Articulation index ( by Greaves)
Where,
• A0 = SPL for zero intelligibility of conversation
• A100 = SPL for 100% intelligibility
• Wf = weighting factor for each third octave band
• Overall AI is measured by adding together 16
different indices
Slide 58 of 63
Engine noise
• Engine noise originates from both the
combustion process and mechanical forces
associated with engine dynamics
• Noise control:
– Controlling pressure variations
– Piston slap – mass of piston, gudgeon pin design,
offset
– Noise shields
– Crankshaft – spoked, damper
Slide 59 of 63
Transmission noise
• Misalignment of shafts, single tooth incorrectly cut or
damaged.
fs1, fs2 = ftm ± fss
fss – shaft speed frequency
• Correct standard tooth profiles
– to account for tooth elasticity effects
– But gear teeth are subjected to variable loading,
– it is impossible to correct for all eventualities
Slide 60 of 63
INTAKE NOISE
• Generated by interruption of airflow at inlet
valves
• Transmitted via air cleaner
• Radiated by air duct
• Noise of 10-15 dB
• Turbo charger compressor noise also radiated
from the air duct
– At blade passing frequency (also higher harmonics)
– Typically 2-4 kHz
Slide 61 of 63
EXHAUST NOISE
• Produced by release of gases as exhaust valves
open and close
• F = engine speed /60 * number of cyls / 2
• Vary with engine load (upto 15 dB)
• Turbo charging reduces engine and exhaust
noise (because of better combustion)
Slide 62 of 63
Attenuation of intake and exhaust
noise
• Devices which minimize flow of sound waves
– Allow gas flow
– Called acoustic filters
• Two types
– Dissipative (absorb acoustic energy)
– Reactive (by intereference)
Slide 63 of 56
Intake and exhaust noise
• Dissipative silencers
– absorptive
material
which
absorbs acoustic energy from the
gas flow
– Produces attenuation at f>500Hz
• Reactive silencers
– when the sound in a pipe or duct
encounters a discontinuity in the
cross-section,
– some of the acoustic energy is
reflected back
– creating destructive interference
– Suitable for attenuating low
frequency noise
– Causes pressure loss
Figure from Smith,2002
Slide 64 of 56
The Helmholtz resonator
• Incorporated in the air
filter
f 
c
2
A
LV
• Produces low frequency
resonance, and
• Attenuation at higher
frequencies
Slide 65 of 63
Catalytic Convertors
• Fitted immediately after exhaust manifold
– For quick heating needed for their functioning
• Have an acoustic attenuation effect
– Gas passes through narrow ceramic pipes
– Attenuation as well as dispersion
• Silencers downstream of catalytic convertors
• Have an acoustic resonant frequency
– Tuned to avoid exciting structural frequencies
• Hence, Often have a double skin and insulation AND
• Isolated from vehicle body by suspending it from
flexible systems
Slide 66 of 63
Intake and exhaust noise
• The following are some of the devices used to
overcome specific silencer tuning problems.
– The Helmholtz resonator – a through-flow resonator
which amplifies sound at its resonant frequency, but
attenuates it outside this range.
– Circumferential pipe perforations – create many small
sound sources resulting in a broadband filtering effect
due to increased local turbulence.
– Venturi nozzles – designed to have flow velocities
below the speed of sound, they are used to attenuate
low frequency sound.
Slide 67 of 63
Aerodynamic noise
• For road vehicles this can be broken down into
three noise generating components:
– Boundary layer distributed over the vehicle body
• Boundary layer noise tends to be random in character
• Absorbent materials
– Edge effects
• noise level higher than boundary layer noise
• Caused by vortices formed at edges
– Vortex shedding (large vortices roll up and break into
smaller ones)
– at various locations on the vehicle body and also at cooling fans
Slide 68 of 63
• Minimizing Aerodynamic noise
– minimizing protrusions from the body surface, making the body
surface smooth and continuous and ensuring that gaps around
apertures such as doors are well sealed.
– Vortices produced at windscreen pillars
• very little which can be done to improve as aerodynamics at A-pillar
contradicts visibility requirements
– Wing mirrors and wheel trims also cause vortices
–
f 
SU
d where S is the Strouhal Number (based on geometry), U is
speed of air, d – obstruction dia
– f = 640 Hz, for d=10 at 70mph
– inlet and outlet apertures are carefully sited and designed
• Should not generate noise and noise from engine compartment should not be
transmitted to the occupant
Slide 69 of 63
Noise from the cooling fan
Blades shed helical trailing vortices
Result in periodic pressure fluctuations when
they strike obstacles
fan rotors are made with unevenly spaced
blades and with an odd numbers of blades
Slide 70 of 63
Tire noise
• Two sources of noise
– Tread pattern excited noise (affected by tire design)
– Road surface excited noise
• Tire designers reduce tire noise
• Chassis designers reduce transmission to occupant
• The mechanism of tire noise generation is due to an
energy release when a small block of tread is released
from the trailing edge of the tire footprint and returns
to its undeformed position.
– Tread patterns designed to control frequencies
– Models of tires with structural dynamic characteristics and
the air contained within them are used at the design stage.
Slide 71 of 63
Brake noise
• Mechanism of noise generation in disc and drum
brakes is still not fully understood
– Complex system of linkages
– Elements of large area held in contact with hydraulic /
friction loading
• Ad hoc / empirical solutions often used
– For low frequency drum brake noise
• Add either a single mass or a combined mass and a visco-elastic
layer applied at anti-nodes of the drum back plate (Fieldhouse et
al., 1996).
– At higher frequencies
• a redistribution of drum mass to eliminate some of the specific
back plate vibration modes.