Wang X. Vehicle Noise and Vibration Refinement

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Body structure noise and vibration reﬁ nement 375

1. Non-resonant path (‘mass law’)

2. Resonant path.

This is shown conceptually in Fig. 15.24 along with the associated dissipa-

tion mechanisms.

When testing a panel for STL, the acoustic damping losses can generally

be ignored, since the source and receiver cavities and the test methods

minimize the impact of these losses on the transmitted noise.

The non-resonant path of the panel is governed only by mass: the more

mass, the lower the vibration and the lower the noise. This comes directly

from Newton’s second law, and is often referred to as the ‘mass law’. The

mass law portion of STL is deﬁ ned as (Raichel, 2006):

∼100 ∼500 ∼10,000

Log frequency (Hz)

Mass law : 6 dB per octave

Resonance

controlled

Mass

controlled

Coincidence

controlled

Coincident

frequency

High damping

Low damping

Transmission loss (dB)

Stiffness controlled

15.23 Sound transmission loss behavior for a ﬂ at plate (Norton, 1989).

Panel

Source

cavity

Receiver

cavity

Power

injection

Power

transmitted

Resonant

path

Non-resonant

path (‘mass law’)

Acoustic

damping

loss

Panel

damping

loss

Acoustic

damping

loss

15.24 Sound transmission mechanisms through panels.

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376 Vehicle noise and vibration reﬁ nement

TL = 20 log(mf) − 47 (15.12)

where m is mass per unit area (kg/m

) and f is frequency.

From eq. 15.12, it can be seen that by doubling the mass of the panel, the

mass-law portion of STL is increased by 6 dB. Therefore, one way to

increase the ability of a body panel to block airborne noise is to increase

its mass. This can be done by adding heavy asphaltic or bitumen patches to

the panel, or by increasing the panel thickness. However, great care should

be taken when increasing panel thickness, since this not only increases the

mass but also increases the stiffness (and therefore the resonant portion)

of the panel radiated noise. This will be covered in detail in the following

section.

The resonant portion of panel noise radiation is governed by the bending

(ﬂ exural) resonances of the panel. Vibration waves traverse across the

panel in bending waves and longitudinal waves. Since bending waves create

motion normal to the panel surface, it is the bending waves which are

dominant in radiating noise.

Resonant bending waves of a panel travel along the panel as dipoles.

Dipoles, by deﬁ nition, have no net noise radiation and so the resonant

radiation of a panel is nearly equal to zero so long as there are no stiffness

discontinuities along the path of the bending wave to disrupt the dipole.

When bending waves encounter a stiffness discontinuity, such as the panel

edge or a stiffening bead, the dipole can be disrupted and converted to a

monopole. A monopole radiates noise in proportion to its surface velocity.

Figure 15.25 illustrates this concept. This means that stiffening beads

and other structural features of the panel can actually increase the high-

frequency noise radiation of a body panel.

Finally, the concept of ‘coincident frequency’ must be discussed, which

is directly related to the resonant part of the noise radiation. The coincident

frequency is the frequency at which the wavelengths of the bending reso-

nances in the body panel exactly match the wavelengths of the propagating

air pressure wave. At this frequency, the panel becomes a very efﬁ cient

radiator and bending waves are easily converted into radiated noise. For

most steel body panels, this phenomenon occurs above 10,000 Hz, but in a

somewhat counterintuitive way, the coincident frequency actually decreases

as the stiffness of the panel increases. This is due to the fact that as the

panel becomes stiffer, the wavespeed of the ﬂ exural modes increases, and

therefore the wavelengths at a given frequency become longer. These

longer wavelengths will then match up with the wavelengths of sound pres-

sure waves in the air at a lower frequency.

Since automotive body panels are far from ﬂ at, and are instead highly

formed, their STL behavior can deviate signiﬁ cantly from ﬂ at panel behav-

ior due to the strong inﬂ uence of the resonant path. This means that any

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Body structure noise and vibration reﬁ nement 377

design strategies which increase the stiffness of a panel for low frequency

will have an adverse effect at high frequencies. This dilemma can be avoided

altogether by introducing sufﬁ cient damping into the panel, which essen-

tially eliminates the bending resonances of the panel (Patil and Goetchius,

2009).

15.7 Body panel design strategies

With the basic principles of body panel dynamics established, a series of

design guidelines can now be discussed, each relating back to one or more

of the panel noise radiation principles covered above. The basic strategy

for body panel design is as follows:

1. Design the surrounding beam structure to minimize the energy trans-

mitted into the panel.

2. Manage individual modes of the panel to avoid excitation sources and

other nearby resonances.

3. Carefully optimize the gauge. Manage the conﬂ icting requirements of

global body static stiffness, modal resonant response and high-frequency

noise radiation.

4. Introduce damping into the panel to minimize resonant noise

radiation.

Reaction force

Line constraint

Plate

Section A-A

++ ++

–– ––

––

–

Radiated sound

(edge or corner

modes)

Radiated sound

(edge or corner

modes)

These regions

cancel each other

Uncancelled volume

velocity

Line constraint

Radiated sound

Subsonic bending

wave (λ

<λ)

15.25 Flexural waves in a panel interrupted by stiffness discontinuity

(Norton, 1989).

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378 Vehicle noise and vibration reﬁ nement

15.7.1 Panel boundary structure

In previous sections, it has been established that providing a strong and

continuous ‘cage’ structure is critical to overall NVH performance of the

body. A strong cage structure will minimize the amount of energy transmit-

ted to the panels and therefore minimize the noise radiating from the

panels. The strategies for accomplishing this have been covered in previous

sections in this chapter.

15.7.2 Panel mode management

In many cases, the resonances of a body panel will coincide with other reso-

nances of the vehicle (acoustic cavity, suspension, etc.) and, more critically,

they can coincide with energy peaks in the forcing function entering the

body. The same rules for global body mode management also apply to body

panels (in order of importance):

1. Manage panel resonant frequencies away from excitation source

frequencies.

2. Manage panel resonant frequencies away from other coupled resonant

systems (i.e. acoustic cavity modes, other nearby panels, etc.).

The most common methods to accomplish this panel mode management

strategy are to change the panel gauge and to introduce bead patterns,

dome patterns and other panel formations to stiffen, and therefore shift the

panel resonant frequencies. Due to the negative impact that panel stiffening

can have on high-frequency radiated noise, it is critical to combine stiffness

increases with sufﬁ cient damping so that the high-frequency resonant

behavior is controlled (Patil and Goetchius, 2009).

Panel beads can be thought of as miniature beams, and as such follow

basic beam theory. Bead design guidelines are as follows (Onsay et al.,

1999):

• Beads should be straight.

• Beads should span the shortest distance between structural

boundaries.

• Beads should not be ‘broken’ or discontinuous across the length of the

bead.

• Beads should terminate close to or at the structural boundaries of the

bead.

• Beads should not cross.

Figure 15.26 shows examples of good panel bead designs and Fig. 15.27

shows examples of less efﬁ cient and generally undesirable panel bead

designs.

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Body structure noise and vibration reﬁ nement 379

15.7.3 Panel gauge optimization

The effect of panel gauge (thickness) on NVH is important. The ﬂ exural

stiffness of a panel can be expressed as bending rigidity, which is given by

(Norton, 1989):

−

()

⎛

⎝

⎜

⎞

⎠

⎟

12 1

(15.13)

where:

B = bending rigidity

E = Young’s modulus

t = panel thickness

ν = Poisson’s ratio.

15.26 Preferred panel bead stiffening examples.

15.27 Inefﬁ cient panel bead stiffening examples.

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380 Vehicle noise and vibration reﬁ nement

Also note that the mass of a panel can be given by:

M = A*t (15.14)

where:

M = mass

t = panel thickness

A = panel area.

Noting that mass changes proportionally to panel thickness and stiffness

changes as the cube of thickness, it can be seen that the net effect of panel

thickness change is a change in the natural frequencies of the panel reso-

nances. This means that optimizing the panel gauge is another tool for panel

modal management.

As always, it is important to understand the implications of changing the

panel gauge (and therefore stiffness and mass) on both structure-borne and

airborne noise. Increasing the panel gauge will shift the resonant energy to

a higher frequency range, which can be useful in managing panel modal

energy which is in the vicinity of excitation forces or other resonances. It

also increases the mass of the panel, which can add extra STL performance

in the mass-law range of the panel. However, if left undamped (or under-

damped), a thicker panel is subject to higher noise radiation in the high-

frequency airborne noise range due to the increased resonant path in the

panel’s STL performance. Therefore, caution must be exercised when

changing panel gauges, and the proper balance between structure-borne

and airborne noise must be maintained.

15.7.4 Panel damping

Introducing damping into a body panel can signiﬁ cantly improve both the

structure-borne and airborne noise performance of the panel. Damping is

often characterized by the term ‘loss factor’. Conceptually, the loss factor

is the ability of a material to dissipate resonant energy, and can range from

0 to 1. A loss factor of 1 means that all resonant energy is completely

damped out (time domain impulse response would have no oscillation), and

a loss factor of 0 means that no resonant energy is damped out (time

domain impulse response would oscillate forever). Table 15.1 shows the

loss factor of several familiar materials.

Achieving high loss factors (high damping) in body panels requires the

introduction of various materials to the panel, since body panel materials

(steel, aluminum, etc.) have very low inherent damping (loss factor less than

0.01). General practice has shown that loss factors above 0.1 provide suf-

ﬁ cient damping for most body panel applications.

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Body structure noise and vibration reﬁ nement 381

There are two kinds of damping mechanisms for sheet metal panels:

1. Unconstrained layer (extensional) damping

2. Constrained layer (shear) damping.

Figure 15.28 illustrates these two types of damping treatments.

Unconstrained damping makes use of a single layer of bitumen or asphal-

tic material which is essentially ‘stretched’ or ‘compressed’ as the bending

waves deform the panel. For this reason, unconstrained damping treat-

ments are often referred to as ‘extensional’ dampers, since the basic strain

mechanism is extensional strain. Extensional strain is much less efﬁ cient in

dissipating energy than shear strain, since the polymer molecule chains do

Table 15.1 Loss factor values for common materials

Material Loss factor

Asphalt 0.055–0.38

Light concrete 0.01

Oak 0.01

Plywood 0.01–0.05

Gypsum board 0.004–0.03

Glass 0.0008–0.002

Dry sand 0.06–0.12

Steel 0.001

Steel with unconstrained damping sheet 0.1

Steel with constrained damping sheet (including

laminated steel)

0.25

Aluminum 0.0001

Lead 0.02–0.03

Source: Cremer and Heckl (1988).

Unconstrained (extensional) damper

Constrained (shear) damper

Viscoelastic layer

Base panel

Constraining panel

Viscoelastic layer

Base panel

15.28 Unconstrained and constrained layer panel damping treatments.

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382 Vehicle noise and vibration reﬁ nement

not slide across each other as much as in the shear strain case. Therefore,

extensional dampers are less effective than constrained layer dampers.

However, extensional dampers are often cheaper and can be easier to apply

than constrained layer damping materials.

Constrained layer damping makes use of a stiff constraining layer and a

soft viscoelastic layer on top of the base sheet metal. The result is a high

level of shear strain imparted on the viscoelastic core when bending waves

deform the panel. Shear strain is highly effective in dissipating vibrational

energy due to the way the polymer molecule chains in the viscoelastic mate-

rial are forced to slide against and across each other. For this reason, con-

strained layer damping treatments are often referred to as ‘shear’ dampers,

since the basic strain mechanism is shear strain. Figure 15.29 shows how

the constraining layer shears the core material under bending. Constrained

layer damping materials can be either added on to the base panel or

included in the base panel itself using a laminated metal approach.

Constrining layer

(1 + ih

)

Damping material

(1 + ih

)

Structure E

(1 + ih

)

Undeformed

constrained layer

Deformed

constrained layer

(a) Constrained-layer treatment

(b) Sandwich panel

Damping material

Constraining layer

Structure

Undeformed sandwich Deformed sandwich

Undeformed multiple

constrained layers

Deformed multiple

constrained layers

15.29 Shear mechanism of constrained layer damper in bending

(Nashif et al., 1985).

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Body structure noise and vibration reﬁ nement 383

In the case of add-on damping treatments (both constrained and uncon-

strained), a minimum coverage area is needed to affect both structure-

borne and airborne noise. For structure-borne noise, add-on damping

treatments can be applied only to the areas of the panel which are exhibit-

ing high modal strain energy (high motion). However, in order to combat

the resonant path in the high-frequency airborne noise range, it is necessary

to cover at least 70% of the panel area which is exposed to the airborne

noise source. In many cases, it can be found that the damping coverage area

needed for structure-borne noise would not sufﬁ ciently cover the panel to

address the resonant path at high frequencies. In these cases, the laminated

metal approach can be an alternative since it has, by deﬁ nition, 100%

damping area coverage.

15.8 Future trends

As with most automotive systems, the drive towards higher fuel economy,

lower emissions and more eco-friendly materials will have a signiﬁ cant

impact on body structure NVH in the future. This will manifest itself in

several ways:

1. Alternative (high efﬁ ciency) propulsion systems will have dramatically

different operational loads on the body structure in terms of both overall

levels but also frequency content. This could fundamentally shift the

basic understanding of the split between structure-borne and airborne

noise contributions.

2. Reducing weight in vehicles has always been a top priority. However,

the recent global emphasis on reducing fuel consumption and emissions

has escalated weight-savings initiatives to the top of most automotive

engineers’ lists. As a result, lighter-weight materials will ﬁ nd their way

into the automotive body structure. Aluminum body structures have

been introduced on several production vehicles, but this practice has

not become mainstream due to cost and manufacturability concerns.

This may change under the current weight reduction pressures. More

‘exotic’ materials (e.g. graphite epoxy composites) commonly found on

racing cars and high-end performance cars may also begin to become

more mainstream. In both of these cases, the challenge for body NVH

control will be daunting since we know that very stiff and very light

structures are also very loud (just ask any loudspeaker designer!). This

also means that some of the more practical and economical solutions

for noise control in the body structure may need to be replaced with

more expensive systems to deal with the lighter-weight materials.

3. Regardless of the materials which will be used in future body structures,

they will all most certainly face tougher requirements with respect to

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384 Vehicle noise and vibration reﬁ nement

environmental impact and recyclability. Certainly, steel and aluminum

are highly recyclable and mature programs have existed for decades for

recycling automotive body structures. However, many of the noise and

vibration treatments applied to the body are not recyclable and neither

are many of the new and exotic composite materials which may ﬁ nd

their way into future body structures.

4. As vehicles get lighter, even more emphasis will be placed on crashwor-

thiness. Much of the crashworthiness of a vehicle comes from the body

structure, and new lightweight materials, along with the different behav-

ior of the aforementioned alternative propulsion systems, will create

new and unimagined challenges for the body structure designer, and

therefore potentially create an even tighter set of criteria within which

body structure NVH must be developed.

5. The use of computer aided engineering (CAE) tools to assist engineers

in designing and reﬁ ning the NVH performance of the body structure

much earlier in the product life cycle will continue to grow. Today, CAE

methods for body NVH are widespread and fairly mature. They are

used extensively at many automotive original equipment manufacturers

(OEMs) early in the design process. However, there are still many

weaknesses in the CAE process which will need to be addressed and

improved in the future, especially as the cost of building prototypes

continues to limit the availability of hardware on which to do develop-

ment testing. These weaknesses include frequency constraints on

structure-borne and airborne noise analysis types, lack of statistical

sampling and probabilistic methods, and lack of a mature understanding

of the forces generated by new and as yet undeveloped propulsion

technologies which act on the body structure, among many others.

15.9 Conclusions

Every aspect of body structure design has an effect on the noise and vibra-

tion level which ultimately reaches the vehicle occupants. As a result, great

care and diligence are required in understanding all of the sources and

paths of noise and vibration energy. Sources and paths for both structure-

borne noise and airborne noise must be well understood, since design

concepts and countermeasures for one often result in poor performance in

the other. The ‘best practice’ design concepts and ‘generic targets’ proposed

in this chapter should be used as the starting point for the evaluation and

design of any automotive body structure. However, realizing that each

vehicle is a unique and complex system, it is critical that the speciﬁ c per-

formance of each body structure be carefully crafted to meet the speciﬁ c

needs of the vehicle of which it is a part. This can be done using a combina-

tion of experience, predictive methods (CAE) and development testing.

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