BASICS OF HEAVY DUTY
DIESEL NVH
November 7, 2006
Ed Green
Roush Industries, Inc.
734-779-7421, ergree@roushind.com
Course Outline
Introduction (8:00 AM - 8:15 AM)
Roush Industries

Ed Green

Course objectives

Pre-Quiz
II. Fundamentals of noise and vibration (N&V) (8:15 AM – 11:30 AM)
A.
Time and frequency domain analysis (8:15 AM – 8:45 AM)
B.
The single degree of freedom system (SDOF) (8:45 AM – 9:30 AM)
C.
Transmissibility and isolation (10:00 AM – 10:45 AM)
D.
The source/path/receiver model (10:45 AM – 11:00 AM)
E.
Summary (11:00 AM – 11:05 AM)
F.
Discussion (11:05 AM – 11:30 AM)
III. Vibration (12:30 PM – 3:00 PM)
A.
Sensitivity to vibration (12:30 PM – 12:45 PM)
B.
Forced vs resonant vibration (12:45 PM – 12:50 PM)
C.
Rotating machinery vibration, orders, and critical speed (12:50 PM – 1:00 PM)
D.
Vehicle driveline issues (1:00 PM – 2:00 PM)
E.
Summary (2:30 PM – 2:35 PM)
F.
Discussion (2:35 PM – 3:00 PM)
IV. Sound and sound quality (3:00 PM – 3:30 PM)
A.
Decibels and A-weighting (3:00 PM – 3:05 PM)
B.
Beating, fluctuation, roughness, loudness, sharpness, and knock (3:05 PM – 3:20 PM)
C.
Passby noise procedure and stationary noise requirement (3:20 PM – 3:30 PM)
V. N&V features of DDC engines (3:30 PM – 3:45 PM)
A.
Engine mounting system
B.
Isolated oil pan
C.
Isolated rocker cover
D.
Crankshaft damper
E.
Lined bellows exhaust manifold
VI. N&V sensors (3:45 PM – 4:00 PM)
A.
Accelerometers
B.
Microphones
C.
Rotational speed transducers
D.
Transducer calibration
•
Quiz and Review (4:00-4:30)
I.

1
I. Introduction

Roush Industries
 Detroit based automotive engineering service company.
 Associated with famous Roush NASCAR Racing Teams.
 2200 employees total.
 45 employees dedicated to noise and vibration engineering and
products.
 www.roushind.com
2
I. Introduction

Ed Green
 Ph.D. from Purdue University
 1983-1984, earthquake certification of lightning
arrestors
 1986-1990, modeling of ultrasonic sound waves for
NRC, ultrasonic inspection
 1994-present, automotive N&V at Roush
 Vehicle N&V engineering
 Noise-control-material engineering
 Vibration dampers
 Intake and exhaust systems
 Passby noise
Station Class
Lightning Arrestor
3
I. Introduction

Course Objectives
 Give tools and training to improve efficiency of N&V issue field
response
 Provide an introduction to N&V concepts
 Teach “language” of N&V
 Not to make experts
4
Pre-Quiz
5
II. Fundamentals of Noise and Vibration (N&V)
A.
B.
C.
D.
E.
F.
Time and frequency domain analysis
The single degree of freedom system (SDOF)
Transmissibility and isolation
The source/path/receiver model
Summary
Discussion
7
II.A. Time and Frequency Domain Analysis


Microphones and accelerometers
produce time domain signals
The Fast Fourier Transform (FFT) is
a technique used for conversion to
the frequency domain (for display by
spectrum analyzers like the
MTS4100)
8
II.A. Time and Frequency Domain Analysis

With the FFT:
 Any time domain signal can be
approximated by a summation of
sinusoids
 A unique set of sinusoids is
defined by magnitude, phase, and
frequency (MTS4100 displays only
magnitude vs frequency)
 The magnitude vs frequency is
called the “Spectrum”
 Based on the frequency range, the
MTS automatically filters the data
and sets the sample rate.
9
II.A. Time and Frequency Domain Analysis

Animation of FFT applied to a square wave.
10
II.A. Time and Frequency Domain Analysis

Animation of FFT
applied to idle noise
signal from a Series
60 engine.
Click to play sound
11
II.A. Time and Frequency Domain Analysis
About the Fast Fourier Transform – The Fourier Transform is a
mathematical function that finds the best least squares fit of a function to a
set of basis functions based on the sine and cosine functions.
The Fourier Transform is named for the French Mathematician and
Physicist, Jean Baptiste Joseph (1768-1830).
The Fourier Transform
(http://mathworld.wolfram.com/FourierTransform.html) is used to find the
spectral terms of a closed form expression (such as, y(t) = cos(at2+bt+c)).
The Discrete Fourier Transform (DFT)
(http://mathworld.wolfram.com/DiscreteFourierTransform.html) is used to
calculate the spectral terms of a digitized signal.
The Fast Fourier Transform (FFT)
(http://mathworld.wolfram.com/FastFourierTransform.html) is a specialized
implementation of the DFT that requires that the set of digitized data have
a radix-two length (i.e. 512, 1024, 2048, etc.). The FFT is much faster
than the DFT, and the FFT is by far the most common implementation of
the Fourier Transform.
The FFT was first published by Cooley and Tukey (J.W. Cooley and O.W.
Tukey, “An Algorithm for the Machine Calculation of Complex Fourier
Series,” Math. Comput. 19, 297-301, 1965).
12
II.A. Time and Frequency Domain Analysis
FFT Signal Processing Issues - For the task at hand (using the MTS 4100 to
diagnose N&V field issues), an advanced knowledge of the FFT is not
necessary, but some participants may be interested.
Digitized Signal
When signals are digitized, issues arise. These are aliasing, leakage, and
quantization error.
Aliasing – The first step in calculating the FFT is to digitize the signal as
shown in the figure. The signal is sampled at intervals of Dt. The sample
frequency is defined as 1/Dt. The signal is digiitized by the A/D (analog to
digital) circuit of the MTS 4100. If a signal entered the A/D circuit that had a
frequency much faster than the sample frequency as shown in the figure, the
digiitzed signal would have a lower frequency than the actual frequency as
shown. This is called aliasing because a line appears in the spectrum that is
not at the true frequency. To prevent aliasing, the sample frequency must be
at least two times faster than the signal frequency. Put another way, the
“Nyquist frequency” (equal to half the sample frequency) must be greater than
the signal frequency.
Example of Aliasing
To prevent aliasing, powerful, analog, low-pass filters called “anti-aliasing
filters” or “brick-wall filters,” are applied to the analog signal before the A/D
circuit. Anti-aliasing is automatically performed by the MTS 4100. Note that
anti-aliasing filters must be analog filters and that aliased data cannot be
repaired.
13
II.A. Time and Frequency Domain Analysis
FFT Signal Processing Issues – Cont.
When a set of digitized samples of length 2n is used to calculate
the FFT, it is implicit in the operation that the pattern represented
in the set repeats through out the data (i.e. it is periodic) as
shown in the upper group of plots in the figure. However,
usually the pattern is not periodic as shown in the lower group of
plots in the figure. Thus, the assumed input for the FFT is
different from the actual input.
The assumed input for the FFT is the sine wave with transients
added at the beginning and end of the data “window.” The
transients cause artifacts in the spectrum called “leakage.” In
the figure, the leakage is spectral content at frequencies other
than the signal frequency. Often leakage is in the form of
spectral “sidebands.”
To prevent leakage, a “window function” is applied to the data so
that the windowed input is periodic. This eliminates the
sidebands as shown in the figure..
A “Hanning Window” like that shown in the figure is automatically
applied to the data by the MTS 4100 to eliminate leakage.
14
II.A. Time and Frequency Domain Analysis
FFT Signal Processing Issues – Cont.
When the signal is digitized, the value at a
sample time is assigned the closest
available digital value as shown in the
figure. This digital “rounding off” is called
quantization error.
The most obvious effect of quantization
error is the limitation of the “resolution” or
“dynamic range” of the instrument. The
“dynamic range” is the difference between
the strongest measurable signal (without
overload) and the weakest measurable
signal. As shown, for a 4-bit A/D values are
available from 0 to 15. The “dynamic
range” or “resolution” of a 4-bit A/D is
20Log10(15) = 24 dB. Similarly, the
dynamic ranges of 8-bit, 12-bit, 16-bit, and
24-bit A/Ds are 48 dB, 72 dB, 96dB, and,
144 dB, respectively.
To maximize its dynamic range, the MTS
4100 is set for input signals typical of
vehicle acceleration and microphone
measurements; however, the MTS 4100
does not allow the inputs to be
automatically ranged for getter dynamic
range.
15
II.B. The Single Degree of Freedom System (SDOF)




Demo Description – An elastic cord is driven
by a sine wave signal at its first four natural
frequencies. A strobe is used to show the
mode shapes.
This demonstration shows that the motion of
an elastic structure is made up of specific
shapes (called mode shapes or eigenshapes)
associated with specific frequencies (called
natural frequencies or eigenvalues).
The demonstration also shows that a strobe
can be used to determine the natural
frequency and mode shapes of structures
undergoing relatively large displacements.
This same behavior could be shown on a truck
frame, steering wheel, or seat while the truck is
under operation.
Setup
17
II.B. The Single Degree of Freedom System (SDOF)
Second Mode
Setup
First Mode
Third Mode
18
II.B. The Single Degree of Freedom System (SDOF)

Many of the characteristics of a
vibrating system can be examined
using the SDOF System consisting
of a mass, spring, and damper. The
model is also known as the
mass/spring/damper model.

The SDOF System model can be
used to understand the behavior of
engine mounts, cab mounts, steering
columns, frames, and other
important dynamic systems of a
truck.

Even though all dynamic systems of
trucks have multiple degrees of
freedom, the performance of each
mode is represented by the SDOF.
19
II.B. The Single Degree of Freedom System (SDOF)
Why study the SDOF? – True SDOF systems are
rare. For example, a truck engine on flexible engine
mounts will have six different rigid-body natural
frequencies.
As we saw in the previous demonstration with the
elastic cord, the behavior of multiple DOF systems
and continuous systems is the sum of an infinite
number of modes.
The mathematical process of describing the
response of a complex dynamic system as the sum
of modes is called modal superposition. Modal
superposition shows that the behavior of the
structure for each natural frequency and mode
shape is described by the behavior of the SDOF
system.
Associated with each mode is a modal mass, modal
stiffness, and modal damping; but these properties
will be different from the global mass, stiffness, and
damping.
20
II.B. The Single Degree of Freedom System (SDOF)

The frequency domain transfer
function (output displacement
normalized by input force as a
function of frequency) is:
X  
1

F     2 m  jc  k

X() and F() are the complex amplitudes of the response displacement and
the input force, respectively. The variable “” is the frequency in
radians/second, and “j” is the square root of minus one.
A Note Regarding Equations – Equations are used throughout this presentation to
show that the conclusions stated are strongly supported by well established theory
and that the conclusions are not opinions. An understanding of the equations is not
required to develop a strong understanding of vibratory systems, but some
participants may feel more comfortable with the material after examining the
equations.
21
II.B. The Single Degree of Freedom
System (SDOF)

The animation below shows the behavior of
a SDOF system with m=1000 kg, fn=10 Hz,
and 10% damping:
22
II.B. The Single Degree of Freedom System (SDOF)

At low frequency (stiffness controlled region):

Near the natural frequency (damping controlled
region):
X   0 1

F   0
k

F  
 1
jc
k m
X  k m

At high frequency (mass controlled region):
X    
1

2
F    
 m
23
II.B. The Single Degree of Freedom System (SDOF)



At low frequency, the response of a forced
resonant system is a function of stiffness.
Near the natural frequency, the response is
very strong and is a function of damping.
At high frequency, the response weakens and
is a function of mass. For example, this would
be the desirable operating condition of truck
engine mounts, etc. The natural frequency
would be below the engine firing frequency.
24
II.B. The Single Degree of Freedom System (SDOF)



Demo Description – To simulate an engine on mounts an electric motor is
suspended from elastic cords. A tap test and the MTS 4100 was
previously used to determine the natural frequencies. The motor is
operated discrete speeds and the peak-hold spectrum is observed using
the MTS 4100.
As the motor speed passes through the natural frequencies of the dynamic
system, the response increases significantly as seen on the plot.
This behavior is similar to what would be seen in an operating engine
except that the engine excites more frequencies than the rotational speed.
These additional engine “orders” are discussed later in the presentation.
Demonstration Setup
Electric Motor with Imbalance
Springs
Base
25
II.B. The Single Degree of Freedom System (SDOF)



Demo screen shot from
MTS 4100.
Resonances at low speed
(frequency).
Response due to
centifugal force at high
speed proportional to
speed squared.
26
II.B. The Single Degree of Freedom System (SDOF)


SDOF Systems can represent torsional resonances also.
Mass (kg), stiffness (N/m), and damping (N*s/m) are analogous to moment
of inertia (kg*m2), rotational stiffness (N*m), and rotational damping (N*m*s).
Linear System
m
k
c
Torsional System
I
k
c
27
II.B. The Single Degree of Freedom System (SDOF)

The animation shows the resonance of a shaft system in a hybrid powerplant
as it passes through a torsional resonance.
28
II.C. Transmissibility and Isolation

Many of the characteristics of an isolated mass (e.g. an engine on rubber
mounts) can be examined using a slight variant of the previous SDOF System:
30
II.C. Transmissibility and Isolation

The transmissibility (mass response normalized by base motion as a
function of frequency) is:
X  
jc  k

2


Y
  m  jc  k



Lower transmissibility is better.
Isolation is the reciprocal of
transmissibility.
Higher isolation is better.
An understanding of the equations is not
required to develop a strong understanding of
vibratory systems.
31
II.C. Transmissibility and Isolation

At low frequency:

Near the natural frequency:
X   0
1
Y   0

F  
  1  mk
c
k m
X  k m

2
At high frequency:
X    
 jc

F    
m
32
II.C. Transmissibility and Isolation




At low frequency, the mass moves in phase
with the base, and there is no isolation.
Near the natural frequency, the
transmissibility is greater than one (gain
rather than isolation). The gain is controlled
by the ratio of the damping squared to the
product of the mass and the stiffness (the
damping ratio squared).
At high frequency, the transmissibility is less
than one, and the gain is controlled by the
ratio of damping to mass.
Overall, the goal of an effective isolation
system is to have the natural frequency lower
than the excitation frequency. Damping is
good to have near the natural frequency, but
decreases isolation at higher frequencies.
33
II.C. Transmissibility and Isolation





When motion of the engine mounts causes motion at the frame-side
attachments, vibration energy is transmitted into the truck chassis.
To minimize this effect for automobiles, the static stiffness of the frameside attachments is made relatively large (10 times).
For heavy trucks, this is not possible because the mounts are much stiffer
and the frame is not relatively stiff (maybe 2 times).
A secondary isolation system (cab mounts) is necessary to keep vibration
at a reasonable level.
Field experience will establish what are good values for engine and cab
mount transmissibility.
34
II.D. The Source/Path/Receiver Model

Sources








Engine
Wind
Tire/Road
Exhaust
Intake
Driveline
Axle
Traffic
HVAC
Paths




Receiver
Airborne Paths
Structure-Borne Paths
SPL at Front/Rear Seat Ear Positions
Vibration at Floor, Pedals, Steering
Wheel, and Seat
36
II.D. The Source/Path/Receiver Model

All truck vibration issues will be due to engine,
driveline, or tire orders.

This indicates that the engine (etc.) is the source
of the vibration, but not necessarily the root cause
because the path may be inadequate.

Inadequate paths include bad engine mounts, cab
mounts, and exhaust mounts.

Inadequate paths can also include strong (or
multiple) resonances of the truck structure.
Sources
Paths
Receiver
37
II.D. The Source/Path/Receiver Model
A Case Study – Several years ago Roush worked on a medium duty
truck that had unacceptable vibration at a particular frequency. It was
found that the truck had four major subsystems with that same natural
frequency including: (1) the back panel of the cab, (2) the second
acoustical mode of the cab, (3) the exhaust system, and (4) the clutch
system. Presumably, the same natural frequency had been used as a
minimum design target for the subsystems. The natural frequency of
these subsystems was excited by the first strong torsional “order” of the
engine.
In this case, the unacceptable vibration was excited by the engine, but
the engine was not the root cause of the vibration issue. There was
nothing wrong with the design of the cab, the exhaust system, or the
clutch system. The issue was the NVH design integration. This is an
example of the path being the root cause of the issue without any
specific part in the path being the cause.
Sources
Paths
Receiver
When designing vehicles, many-but-not-all vehicle developers check
the modal alignment of the vehicle. A chart is made of all the natural
frequencies of all the major subsystems of the vehicle. When natural
frequencies align, the design is checked and usually changed.
38
II.D. The Source/Path/Receiver Model

For trucks without N&V issues, low frequency
noise is dominated by structural paths and high
frequency noise is dominated by airborne paths.
39
II.E. Summary


•
•
•
The magnitude of sound or vibration vs frequency is called the
spectrum and it is calculated using the FFT (Fast Fourier
Transform).
For a resonant SDOF (Single Degree of Freedom) system:
Low frequency is stiffness controlled.
Near the natural frequency is damping controlled.
High frequency is mass controlled.
41
II.E. Summary



•
•
•

Transmissibility is the reciprocal of isolation.
Lower transmissibility is better.
For the modified SDOF model transmissibility:
Low frequency, transmissibility is approximately one.
Near the natural frequency, transmissibility is higher than one
and controlled by the damping ratio squared (c2/km).
High frequency, transmissibility is less than one and controlled
by damping and mass (c/m).
Frame flexibility reduces the effectiveness of engine mounts.
42
II.E. Summary



The source/path/receiver model shows that the root cause of a
N&V issue may be the path rather than the source.
Paths may be mounts or structural resonances.
Low frequency noise is normally dominated by structural paths,
and high frequency noise is normally dominated by airborne
paths.
43
II.F. Discussion

What should the natural frequency of a truck steering column be
and why?
Typically steering columns (including all attachments and the steering wheel) have
relatively little damping (approximately 1 %). Thus, it is important that the natural
frequency is not in a frequency range where engine vibration is strong.
This would be most easily achieved by making the natural frequency very low (e.g.
10 Hz. This low frequency would require that the steering column be mounted
relatively softly, potentially compromising steering feel and control. Also, at low
frequency, excitation from the road (bumps, dips, and potholes) is strong.
The steering column is usually tuned to have a natural frequency just above idle
“firing frequency” because this is a condition where the engine is not typically
operated continuously. For example, the firing frequency at idle for a Series 60
engine at 600 rpm is 30 Hz (= 600 revs/min * 3 fires/rev ÷ 60 sec/min). The
steering column could be designed to have a natural frequency of 36 Hz which
would correspond to the firing frequency at 720 rpm.
44
II.F. Discussion

Steering wheel vibration in a truck is excessive at the engine firing
order. Name two reasons why excessive engine vibration may not be
the root cause. Hint – source/path/receiver model.
The response of the steering wheel is the product of the input forcing function
(vibration from the engine source) and the vibration transfer function. If the engine
vibration (source) is within normal limits, the transfer function (path) of the steering
wheel is too large at the engine firing frequency.
(1) The natural frequency of the steering column might be lower (and closer to a
“significant engine-vibration-order”) than intended due to loose fasteners or a poor
fit at the attachment points.
(2) The steering wheel is mounted to the cab, so deficient cab mounts would
contribute to excessive steering wheel vibration. In this case, vibration levels would
also be higher than normal in other parts of the cab. Likewise, deficient engine
mounts would contribute to excessive steering wheel vibration with higher than
normal vibration levels in other parts of the truck.
45
II.F. Discussion

An off-road truck manufacturer increased the stiffness of the
engine mounts to increase durability. Vibration at idle is
significantly worse. Name two reasons why. Hint –
transmissibility.
Normally engine mounts are designed so that the rigid-body natural frequencies of the
engine/mount system are lower than the engine firing frequency at idle. This is a fairly low
frequency (less than 30 Hz at 600 rpm idle). Thus, the excursion of the mounts under the
influence of road inputs can be significant. This negatively impacts the durability of the mounts
and can also lead to perceivable “after-shake” of the engine. (1) Vibration at idle is worse
because the engine mounts were made stiffer, and this moved the rigid-body natural
frequencies closer to the idle frequency. Also, the stiffer mounts would make the ratio of mount
stiffness to frame stiffness worse.
Engine mounts can be tuned to “decouple” the modes of the engine. This means that the “roll
mode” of the engine is not associated with engine bounce, pitch, etc. Likewise, the pitch mode
is not associated with engine roll, bounce, and pitch; and so on. (2) Vibration at idle is worse
because changing the mount rates increased the coupling of rigid-body modes of the
engine/mount system.
46
III. Vibration
A.
B.
C.
D.
E.
F.
Sensitivity to vibration
Forced versus resonant vibration
Rotating machinery vibration, orders, and critical speed
Vehicle driveline issues
Summary
Discussion
48
III.A. Sensitivity to Vibration

Human sensitivity to vibration depends on a large number of
factors including the part of the body, body position, the nature of
the vibration (shock or continuous), and direction of vibration.

A sampling of vibration sensitivity data is presented.
49
III.A. Sensitivity to Vibration



The perception of vibration is approximately linear (i.e. twice as much
acceleration is perceived as twice as much acceleration).
The figure (Cyril M. Harris and Charles E. Crede, Shock and Vibration
Handbook, Second Ed., McGraw-Hill, 1976) shows the limit of perception,
unpleasantness, and intolerability as a function of frequency.
The difference between barely perceivable and intolerable is approximately
100X. (3,000,000X for human hearing).
Sensitivity to vibration declines steeply above 200 Hz.
Vibration Perception
1
Intolerable
Note that 1 g = 9.8 m/s2.
Acceleration (g, peak)

Unpleasant
0.1
Threshold of
Perception
0.01
0.001
1
10
Frequency (Hz)
100
50
III.A. Sensitivity to Vibration

ISO 5349-1 provides exposure limits
for hand vibration.

The standards are for the vector
addition of all directions.
Weighting Factor for Hand Vibration Data



The accelerations are multiplied by the
weighting factor shown in the figure,
so the highest susceptibility is around
12 Hz.
The 8-hour levels are 26 m/s2, 14
m/s2, 7 m/s2, and 3.7 m/s2 for 1, 2, 4,
and 8 years of exposure. These
levels may be expected to produce
episodes of finger blanching in 10% of
operators exposed.
Weighting Factor
1.00
0.10
0.01
1
10
100
1000
Frequency (Hz)
Symptoms are rare for operators with
an 8-hour level of less than 2 m/s2 and
unreported for 1 m/s2.
51
III.A. Sensitivity to Vibration
10
Vibration Exposure Criteria for Fatigue, Vertical, Seated,
Whole-Body
24 hour
16 hour
8 hour
4 hour
2.5 hour
1 hour
25 min
16 min
1 min
1
80.00
63.00
50.00
40.00
31.50
25.00
20.00
16.00
12.50
10.00
8.00
6.30
5.00
4.00
3.15
2.50
2.00
1.60
0.1
1.25

1.00

Whole body vibration can cause operator fatigue.
The figure (Jens Trampe Broch, Mechanical Vibration and Shock
Measurement, Bruel and Kjaer, 1980) shows vibration exposure
versus exposure time.
The highest susceptibility is between 4 Hz and 8 Hz.
Acceleration (m/s^2, rms)

Third-Octave-Band Frequency
52
III.A. Sensitivity to Vibration

Some companies use DVD (vibration dose value) to access ride
harshness.

From Ken Horste, Ford Motor Company, SAE Paper 95373, 1995.

Primarily used for shift quality and driving over impact strips.

To calculate DVD, the acceleration data is bandpass filtered between
1 Hz and 31.9 Hz. Then the filtered acceleration is taken to the
fourth power and integrated numerically.
53
III.B. Forced vs Resonant Vibration
Vibration in structures can be induced without resonance. For
example, low-frequency rigid body motion, and motion of nonresonant structures (analogous to a pile of sand).
The figure shows both forced and resonant vibration behavior.
(Accelerometer mounted on Series 60 engine).


2020
3.00
50
1950
30
1900
20
1850
1800
10
1750
Forced Vibration
8
1700
6
5
1650
2
1500
Log
rpm
Tacho1 (T1)
3
1550
(m/s2)
4
1600
1450
1400
1
1350
800e-3
1300
600e-3
500e-3
1250
400e-3
1200
300e-3
1150
Strong Resonant
Vibration
200e-3
1100
494.6
1020
0
50
100
150
200
250
300
350
400
450
500
100e-3
550
600
650
700
750
800
850
900
950 1000
Hz
Front Mount Y:+Y (CH25)
55
III.B. Forced vs Resonant Vibration

It is important to distinguish forced and resonant vibration because the
method of treatment is fundamentally different.

Resonant vibration is treated by adding damping or changing the
resonant frequency.

Forced vibration is treated with isolation or reducing the forcing amplitude
(e.g. improved balancing).
2020
3.00
50
1950
30
1900
20
1850
1800
10
1750
8
1700
6
5
1650
2
1500
Log
rpm
Tacho1 (T1)
3
1550
(m/s2)
4
1600
1450
1400
1
1350
800e-3
1300
600e-3
500e-3
1250
400e-3
1200
300e-3
1150
200e-3
1100
494.6
1020
0
50
100
150
200
250
300
350
400
450
500
100e-3
550
600
650
700
750
800
850
900
950 1000
Hz
Front Mount Y:+Y (CH25)
56
III.C. Rotating Machinery Vibration, Orders, and Critical Speed

Rotating equipment includes truck engine, driveline, and tires.

If an event happens once per revolution it is called “first order.”

Imbalance is an example of a first order vibration source.

If an event happens twice per revolution it is “second order.”

Vibration due to driveline “Cardan Joints” (U-Joints) is an example
of a second order vibration source.

Orders can be non-integer. For example, due to FEAD pulleys,
transmissions, differentials, etc.
58
III.C. Rotating Machinery Vibration, Orders, and Critical Speed
The truck engine produces vibration and noise at many orders.
Orders appear as straight, diagonal lines in the “Waterfall Plot” or “Campbell
Diagram.”
Resonances appear as straight vertical lines.



3rd
6th
9th
Order Order Order
2020
3.00
50
1950
30
1900
20
1850
1800
10
1750
8
1700
6
5
1650
2
1500
Log
rpm
Tacho1 (T1)
3
1550
(m/s2)
4
1600
1450
1400
1
1350
800e-3
1300
600e-3
500e-3
1250
400e-3
1200
300e-3
1150
200e-3
1100
494.6
1020
0
50
100
150
200
250
300
350
400
450
500
100e-3
550
600
650
700
750
800
850
900
950 1000
Hz
Front Mount Y:+Y (CH25)
59
III.C. Rotating Machinery Vibration, Orders, and Critical Speed
Why does the Series 60 engine have vibration at certain orders?
The Series 60 is an “even fire” six cylinder engine. Every revolution of the engine takes the crankshaft, rod,
and pistons through a complete cycle. Thus, crankshaft imbalance produces first order vibration. Other
multiples of first order (2nd, 3rd, etc.) are also produced by the rotation of the crankshaft due to the “nonsinusoidal” motion of the pistons and rods (Charles Fayette Taylor, The Internal Combustion Engine in Theory
and Practice, Volume 2, MIT Press, 1979).
Every two revolutions the cam/injector system goes through a cycle. Thus, the cam drive produces half order
vibration. Other multiples of half order (1.0, 1.5, 2.0, etc.) are also produced by motion of the cam drive due to
the non-sinusoidal torque load. The 4.5 order is strong.
Every two revolutions (720 degrees) a cylinder goes through a complete combustion cycle (intake,
compression, ignition, and exhaust), but this is an even fire engine, so every 120 degrees (720/6) the
combustion process repeats. Thus, combustion produces 3rd order vibration and multiples (6th, 9th, 12th order,
etc.).
The Series 60 engine has an air pump that produces vibration at 1.2 engine order because it is driven by a
gear train at 1.2 times the speed of the crankshaft.
60
III.C. Rotating Machinery Vibration, Orders, and Critical Speed

Shafts have a first bending natural frequency. When the rotational
speed of the shaft is at this natural frequency the imbalance
excites this natural frequency, and the amplitude of vibration is
only limited by the damping. Damping is usually very small
because the shaft is like a rotating banana, so there is no
alternating stress. The shaft speed at which the rotational speed
matches the natural frequency is called the “critical speed.”

Shaft systems (like a heavy truck driveshaft) are almost always
designed so that the shaft never operates at the critical speed.

At the critical speed, the shaft deformations become larger which
produces even larger imbalance. The situation (deformation
producing larger imbalance and imbalance creating larger
deformation) can become unstable, and the shaft can experience
catastrophic failure.
61
III.C. Rotating Machinery Vibration, Orders, and Critical Speed
Sometimes shafts must spin at speeds greater than critical speed. For example, high
speed race cars and steam turbines at electrical power plants.
For the race car, the driveshaft critical speed is moved to a lower frequency, so the
driveshaft quickly accelerates through the critical speed and never dwells at the
critical speed.
For the steam turbine, the natural frequencies are all known in advance. When the
turbine is put into service from a stop, the turbine is quickly accelerated through the
critical speeds to prevent damage. The steady-state speed of the turbine is not
located near a critical speed.
Note that vehicles with one-piece driveshafts are usually speed limited to prevent the
vehicle from operating at the critical speed of the driveshaft. This is approximately
100 mph for pickup trucks and 130 mph for RWD sports cars.
62
III.D. Vehicle Driveline Issues

Vehicle driveline issues usually occur at the first and second order
driveline frequency (i.e. once or twice per driveshaft revolution).

First order disturbances are associated with “imbalance” and
“runout.”

Second order disturbances are associated with “Cardan Joints”
and “Driveline Angles.”
64
III.D. Vehicle Driveline Issues
What is a flexible powertrain mode?
(Mode indicates the shape of the object while vibrating at resonance)
Flexible modes occur
at higher frequencies
usually over 50 Hz
First Mode P/T Bending
(automotive)
(from: SAE Universal Joint and Driveshaft Manual)
Mode shapes are usually displayed
as if frozen at maximum deflection
or are animated.
Second Mode P/T Bending
65
III.D. Vehicle Driveline Issues
Driveshaft Critical Speed Calculation
(Pin-Pin)
General
Equation:
EIg
Nc  30
3
WL
where:
Nc= Revolutions per minute, rpm
E = Modulus of elasticity, psi
I = Area moment of inertia, in4
g = Acceleration due to gravity, in/s2
W = Total weight of shaft, lb
L = Shaft length between supports, in
Round Tubular
Steel Shaft:
do  di
N c  4,705,000
L2
2
2
where:
do= outside diameter, in
di = Inside diameter, in
L = Shaft length, in
note: for solid shaft, di = 0
(Hollow shaft gives higher Nc)
Equations are from the SAE Universal Joint and Driveline Design Manual
66
III.D. Vehicle Driveline Issues
Critical Speed - Calculated vs. As-Installed


Calculated critical speed, Nc is based on pin-pin end conditions
Nc drops by 10 to 15% when installed in a RWD single piece driveshaft
vehicle (flexibility at attachments).
 Nc can drop another 5% when the vehicle is operating, due to nonlinearity (like slip spline).
 For cars & light trucks, good practice requires 10% margin between
the driveshaft resonance as operating in the vehicle, and the maximum
driveshaft speed.
Total 25-30%
 For medium & heavy trucks, it is recommended that the maximum
driveshaft operating frequency is > ½ Nc.
(This is to stay away from resonances excited by 2nd order U-joint
excitation.)
67
III.D. Vehicle Driveline Issues
Driveline Roughness, Drumming/Boom, & Shudder








Definitions
Driveline arrangements
Excitation
Path
Unbalance and runout
Pitch line runout
Transverse AWD
Driveline angles
68
III.D. Vehicle Driveline Issues
Driveline Disturbance Definitions:

D/L Roughness – mid frequency vibration (50–80 hz), usually at
vehicle speeds above 50 MPH (automotive), and can have
boom/drumming associated with it.
69
III.D. Vehicle Driveline Issues
On-Highway Driveline Components
Courtesy of
Rockwell Int’l
70
III.D. Vehicle Driveline Issues
Driveline Excitations






Unbalance
Runout in driveline
Slop in driveline
Driveline angles
Pinion pitch line runout (first order gear rotation error, explained
in gear section)
Path is usually structure-borne
Note that “Unbalance” and “Imbalance” are the same.
71
III.D. Vehicle Driveline Issues
Potential Contributors to 1st Order Driveline Excitation
72
III.D. Vehicle Driveline Issues
System Related Issues
73
III.D. Vehicle Driveline Issues
Runout Measurement
PART CENTER
TIR
Runout 
 E
2
TIR stands for Total Indicator Reading
74
III.D. Vehicle Driveline Issues
10,000 RPM !!!
Effect of Unbalance
W
F  R 2
g
where:
W = Weight
g = gravity
R = radius of mass center
 = angular velocity
Typical
passenger
vehicle
spec: <.4
(from: SAE Universal Joint and Driveshaft Manual)
Note that “Unbalance” and “Imbalance” are the same.
75
III.D. Vehicle Driveline Issues
Runout Results in Unbalance
Examples:

A driveshaft weighing 20 lb has a perfect component balance.
 We put it into a vehicle with companion flange runout of .0045 inch (.009
TIR).
The resulting unbalance from runout at that end
1/2 Shaft Weight
Offset
Unbalance
= 10 lb x 16 oz/lb x .0045 in = .72 in-oz
(typical driveshaft component balance specification is < .4 in-oz)

If we have total spline slop of .035, resulting unbalance is
= 10 lb x 16 oz/lb x .035 in /2 = 2.8 in-oz
(Thus, take care that splines are tightly controlled)
76
III.D. Vehicle Driveline Issues
Contributors to
Unbalance & Runout

Spline clearance/slop

Bent shafts (Driveshaft or transmission)

Transmission unbalance contribution (have seen up to 10 in-oz in large automatic,
that’s a lot!)

System balancing – match mounting of components

Transmission vs. Axle sensitivity - the end attached to the axle (including AWD) is
usually 2 or more times as sensitive, because of mass ratio & path
77
III.D. Vehicle Driveline IssuesTire/Wheel
Inputs
1st Order
Imbalance
LF
RF
LR
RR
Runout
Force Variation
2nd Order
Oval Shape
3rd Order
Out-of-Round
Force Variation
Force Variation
78
III.D. Vehicle Driveline Issues
Tire Frequency Calculation
The following empirical formula was developed (by General Motors):
rev
20,850
20,850


mile tire diameter wheel diameter  2section width  aspect ratio 
where:
tire diameter  wheel diameter  2tire section height 
tire section height = section width x aspect ratio
Example for a P205/60R15 tire:
then:
rev

mile
20,850
in


15in  2   205mm 
 .60 
25.4mm


rev
 MPH
rev mi
hr
mile
RPM 
 

mile hr 60min
60
and:
frequency(hz ) 
 844
RPM
60
Notes:
Tire load and pressure - have little effect on revs/mile due to the fact that the radial belts and tread have a very high hoop strength, acting like a track.
Tread wear also has little effect since tire cords tend to stretch with age as tread wears. Usually <2% change with pressure, load, and/or wear.
% slip - due to power transferred and road conditions ( wet, dry, sand, ice, etc) make more difference in revs/mile than any of the above.
79
III.D. Vehicle Driveline Issues
Driveline Disturbances


Hardware comparison
 Cardan or Hooke Universal Joints
 CV joints – not covered
 Flex couplings – not covered
Discussion
 Rotational errors
Cardan Joints
80
III.D. Vehicle Driveline Issues
Cardan or Hooke Universal Joint
(from: SAE Universal Joint
and Driveshaft Manual)
81
III.D. Vehicle Driveline Issues
Cardan Joint Developing Non-Uniform Motion
Courtesy of
Dana Corp.
82
III.D. Vehicle Driveline Issues
Measurement of Driveline Angles
This analysis assumes
no offset from top view.
83
III.D. Vehicle Driveline Issues
Driveline Angle Change
84
III.D. Vehicle Driveline Issues
Driveline Angle Change
Proper phasing of the joints results in some
cancellation of the torsional transmission errors.


residual


front  rear
2
2
or residual angles
(from: SAE Universal Joint and Driveshaft Manual)
85
III.D. Vehicle Driveline Issues
Residual Angle at Front/Rear Angle Differences of 0.5, 1.0, and 2.0 Degrees
2.0 Degree
Difference
Front to Rear
R O U S H
A N A T R O L
Generated Residual Angle - Deg.
7.00
 residual 
6.00


front  rear
2
2
1.0 Degree
Difference
Front to Rear
5.00
4.00
0.5 Degree
Difference
Front to Rear
3.00
2.00
Probable Limit
For Acceptance
Drive-Line
1.00
0.00
2
3
4
5
6
7
8
9
10
11
12
Front/Rear Average Working Angle - Deg.
13
Mark Jackson
Roush Industries
February 1998
86
III.D. Vehicle Driveline Issues
Rotational Inertia Effects
1= 2 & 3 - The equal
angles cancel speed errors,
but the driveshaft inertia is
being accelerated and
decelerated twice per
revolution. This results in
cyclic torque disturbances
within the driveline system.
Forces generated can be
significant, especially for
large angles & heavy shafts.
87
III.D. Vehicle Driveline Issues
Two Piece Driveline
& Center Bearing
CVertical
C1
Center Support 3
C2 Bearing
Lateral
SAE Universal Joint
and Driveshaft Manual


Couples generated at the joints result in take-off shudder
The rotating secondary couples, Cn are a function of the driveline torque and the
angles at the Cardan joints:
Cn = T tann




where: T = driveline torque
n = joint angle
The forces generated are dependant on shaft lengths
fn usually 12-18 hz, 12-18 mph
Low speed - you go through fn quickly under high torque
Low torque moment when steady state at fn
88
III.D. Vehicle Driveline Issues
2 Piece Driveline & Center Bearing
Cn = T tann
Center tan 
Angle
0o
.000
2o
.035
4o
.070
6o
.105
8o
.140
10o
.176
15o
.268
Keep the angles small!
Couples
C1
C3
C2
1
2
3
65 hz
(from: SAE Universal
Joint and
Driveshaft Manual)
18 hz
Typical shaft support bearing
transmissibility characteristics
due to unbalance
89
III.E. Summary

The perception of vibration is approximately linear (i.e. twice as
much acceleration is perceived as twice as much acceleration).

The difference between barely perceivable and intolerable
vibration is approximately 100X.

Sensitivity to vibration declines steeply above 200 Hz.

ISO 5349-1 provides exposure limits for hand vibration. Limits are
specified for various vibration levels and exposure times.
Symptoms are rare for operators with an 8-hour level of less than
2 m/s2 and unreported for 1 m/s2.
91
III.E. Summary

Vibration can be either forced or resonant. Resonant vibrations are lines
at specific frequencies on the Campbell Diagram.

Forced vibrations are treated by reducing the forcing levels and adding
isolation.

Resonant vibrations are treated by adding damping or moving the
resonant frequency.

If an event happens once per revolution (like rotating imbalance), it
produces first order vibration.

If an event happens twice per revolution (like a Cardan Joint), it produced
second order vibration, etc.

Orders appear as angled lines in the Campbell Diagram.
92
III.E. Summary

The Series 60 engine produces vibration at many orders including: 1st
order (engine imbalance); 1.2 order (air pump); 4.5 order and various half
orders (cam/injector drive); and 3, 6, 9, … order (engine combustion).

The speed at which a shaft (like a driveshaft) is rotating at its first
bending natural frequency is called the critical speed.

Catastrophic failure can occur at the shaft critical speed.

Driveline imbalance and runout produce first order driveline vibration.

Driveline angles must be properly matched to minimize second order
vibration due to the Cardan joints. Suspension movements usually cause
driveline angles to change.

Tires produce vibration because of imbalance, runout, out-of-round, and
force variation.
93
III.F. Discussion

The MTS 4100 automatically “sorts” and “prioritizes” vibration
using its Principal Component Display (PCD). How does it do
this?
The MTS 4100 continuously monitors engine speed and vehicle speed from the interface bus. The user selects the tire size, axle
ratio, and ratio information from user defined sources. From this information, the orders associated with each frequency spectrum
peak are calculated. These are grouped into components and summed for the component. For example, orders of the engine (0.5,
1, 2, and 3) are added to calculate the engine contribution, etc.
Consider the following example:
The truck has a seat track vibration at 40 Hz, 2000 rpm, and 70 mph. The tires are P295/75R22.5, and the axle ratio is 4.11. The
pulley ratio for the air pump (user defined) is 1.2.
From the formula presented earlier, the tire produces 522.3 revs/mile = 20850/[22.5+2(295*0.75/25.4)]. The tire speed at 70 mph is
609.3 rpm = 522.3*70/60, and the first order tire frequency is 10.16 Hz = 609.3/60. The frequencies for second, third, fourth, and
fifth order tire vibration are 20.3, 30.5, 40.6, and 50.8 Hz, respectively. None of these frequencies correspond to the 40 Hz
vibration, but 40.6 is close.
The axle ratio is 4.11, so the driveshaft spins 4.11 times faster than the tires. Thus, the first order driveshaft frequency is 41.7 Hz =
10.16 *4.11.
The first order engine frequency is 33.3 Hz = 2000/60. The second order engine frequency is 66.7 Hz. These frequencies do not
correspond to the 40 Hz vibration.
The first order air pump frequency is 40 Hz = 1.2*2000/60.
The MTS 4100 would go through the process above to identify the air pump as the component producing the vibration.
94
III.F. Discussion

Is the vibration only observed at one vehicle speed or one engine
speed (resonant vibration), or is it spread over a wide speed
range (forced vibration)?
If there is a bad engine mount, there might be one engine speed at which the vibration
is worst, but higher levels of vibration will be obvious at other engine speeds as well.
This would be an example of a forced vibration issue.
Similarly, if the driveline angles are not right, there will be vehicle speeds at which the
vibration is worst, but higher levels of vibration will be observed at other vehicle
speeds.
95
III.F. Discussion

If the same vibration is observed in different gears, does it follow
speed (tire/driveline) or engine speed (engine)?
Engine vibration issues track with engine speed even in different gears. Tire and
driveline issues track with vehicle speed. Note that engine loads changes in different
gears, and there may be a combination of engine and tire/driveline issues, so data
from the MTS 4100 (rather than seat-of-pants) may be necessary.
96
III.F. Discussion

Is the vibration engine-torque sensitive?
Vibration issues can sometimes be identified by their sensitivity to engine torque.
For example, axle whine is very sensitive to torque, driveline issues are moderately
sensitive to torque, and tire issues are relatively insensitive to torque.
97
III.F. Discussion

Does the vibration produce a beating that comes and goes,
especially during sweeping turns?
If there are vibration issues with multiple tires, very small speed differences between
the tires will cause a beating vibration (and noise) that will slowly come and go. If
the tires are on opposite sides of the vehicle, the beating will come and go more
quickly during sweeping turns.
98
IV. Sound and Sound Quality
A.
B.
C.
Decibels and A-weighting
Beating, fluctuation, roughness, loudness, sharpness, and knock
Passby noise procedure and stationary noise requirement
100
IV.A. Decibels and A-Weighting


Human hearing is non-linear and is a function of both frequency
and loudness.
The variation of frequency/loudness dependence of perceived
loudness (acute hearing) is shown in the figure (Robinson and
Dodson, British Journal of Applied Physics, 7, 166(1956) and ISO
R226-1961).
Equal Loudness Curves
101
IV.A. Decibels and A-Weighting
To construct instruments that would measure perceived sound levels, A-,
B-, and C-Weighting filters were developed.
The filters are applied so that measured values are approximately equal
to perceived values according to the equal loudness curves.
Originally, the A-Weighting was used for levels below 55 dB, B-Weighting
was used for 55-85 dB, and C-Weighting was used above 85 dB. For
various reasons, A-Weighting (dBA) is the primary metric used today.



A, B, and C Weighting Filters
Equal Loudness Curves
0
-10
A-Weighting
-30
B-Weighting
C-Weighting
-40
-50
-60
16000
10000
6300
4000
2500
1600
1000
630
400
250
160
100
63
40
25
16
-70
10
Gain (dB)
-20
Third-Octave-Band Frequency (Hz)
102
IV.B. Decibels and A-Weighting
Decibel Facts
• The decibel is named for Alexander Graham Bell.
• Perceived doubling of loudness is 10 dB, while actual doubling of sound pressure
level is 6 dB. Note: “deci” + “bel”
• Smallest perceivable change in sound is 1 dB.
• Quietest perceived sound is 0 dBA.
• Threshold of pain is approximately 130 dBA.
• Dynamic range of human ear (0 to 130 dBA) is 3,000,000 times.
L(dB)  20 Log ( Prms / 20e  6Pa)
103
IV.B. Beating, Fluctuation, Roughness, Loudness,
Sharpness, and Knock



With digital computers, the perceived loudness can be calculated by
applying the equal loudness curve correction directly to recorded data
(rather than using the A-, B-, or C-Weighting curves).
Loudness can be expressed in dB Phones or in Sones.
Sones are arranged so that sound which is perceived as twice as loud
has twice the Sones value. Heavy truck interior levels at highway speed
are probably approximately 10 Sones while underhood levels are
approximately 100 Sones (sounds ten times louder).
Note 40 dBA is approximately
equal to 1 Sone.
105
IV.B. Beating, Fluctuation, Roughness, Loudness,
Sharpness, and Knock
500 Hz, 70 Hz Roughness
“Roughness”
describes a
signal with
amplitude
modulation of 10
Hz to 200 Hz.
Worst at 70 Hz.
0.5
0.0
-0.5
500 Hz
-1.0
0.0
0.1
0.2
0.3
0.4
0.5
0.4
0.5
Time (s)
500 Hz, 70 Hz Roughness
1.0
500 Hz
with 4 Hz
Modulation
Signal (volts)

“Beating” or
“Modulation”
describes a
signal with
amplitude
modulation of 1
Hz to 10 Hz.
Worst at 4 Hz.
0.5
0.0
-0.5
-1.0
0.0
0.1
0.2
0.3
Time (s)
500 Hz, 70 Hz Roughness
1.0
500 Hz
with 70 Hz
Roughness
Signal (volts)

Signal (volts)
1.0
0.5
0.0
-0.5
-1.0
0.0
0.1
0.2
0.3
Time (s)
0.4
0.5
106
IV.B. Beating, Fluctuation, Roughness, Loudness,
Sharpness, and Knock

“Sharpness” describes how sharp a signal sounds. It is a
combination of spectral shape and loudness.
Noise
With Sharpness
Reduced
Sharp Noise
Produced By
Exhaust Flex
Coupling
107
IV.B. Beating, Fluctuation, Roughness, Loudness,
Sharpness, and Knock

“Knock” describes the degree of knock. Usually a metric is selected based on
the particular nature of the noise.

The ratio of the peak transient loudness to the average loudness can be used.
Kurtosis can sometimes be used.

The character of the knock can vary by application. Knock and terms like it
are called “Semantic Differentials.”
Series 60
At Idle
Electric Air
Pump with
Light Knock
Electric Air
Pump with
Strong Knock
108
IV.C. Passby Testing



The EPA standard for external heavy truck noise is the Code of
Federal Regulations 40CFR205 which requires noise passby levels of
80 dBA or less when tested per SAE J366.
The standard involves full throttle acceleration of the heavy truck
tractor through rated engine speed while peak sound levels are
measured at 50 ft.
The noise levels are 6 dBA louder than those allowed for cars in the
US and 12 dBA louder than those allowed for cars in Europe.
Engine - Other Noise
Sources
20%
Engine - Front of Engine
44%
Passby Noise Sources
for a Class 8 Truck
Powered by a Series
60 Engine
Engine - Oil Pan
36%
110
IV.C. Passby Testing

Truck is accelerated from the acceleration line to the end line. Gears
selected so that approach engine speed is no more than 67% of
maximum rated speed (MRS) and MRS is achieved in the indicated
zone.
100 ft
60 ft
50 ft
Reach Max
Speed in
Zone
Truck
Acceleration
Line
End
Line
Microphones
111
IV.C. Stationary Testing

DDC N&V engineers use stationary testing for auditing and
competitive benchmarking purposes. Stationary noise is not an EPA
requirement.

Testing is performed per the SAE J1096 standard.

The engine is accelerated (WOT) from idle to governed speed,
held steady to stabilize, and then decelerated to idle with no
throttle.

Testing is performed in a large paved area free from obstructions
for 100 ft.

Microphones are placed 50 ft from the centerline of the truck even
with the rear of the cab.

The maximum A-weighted sound level is recorded.

Additional microphones are sometimes used at the front and rear
of the truck to help identify noise sources.
112
V. N&V Features of DDC Engines – Current and 2007
A.
Engine Mounting System
B.
Isolated Oil Pan
C.
Isolated Rocker Cover
D.
Crankshaft Damper
E.
Lined Bellows Exhaust Manifold
114
V.A. Pre-2007 Engine Mounts
Rear engine mounts
115
V.A. 2007 Engine Mounts
Rear engine mount
Front engine mount
116
V.B. Isolated Oil Pan
-isolating gasket
-composite oil pan with isolating/sealing gasket
-oil pan sleeve/bolt assembly with grommet
117
V.C. Isolated Rocker Cover
-isolating gasket
-composite rocker cover with isolating/sealing gasket
-rocker cover sleeve/bolt assembly with grommet
118
V.D. Crankshaft Damper
-function of the crankshaft viscous damper is to minimize torsional
vibration present in broad operating frequency ranges
119
V.E. Lined Bellows Exhaust Manifold
NVH ISSUE
Without internal Bellows
liner, “chirp” noise present
during engine operation
-current S60 exhaust manifold configuration with fey ring design
-2007 S60 exhaust manifold configuration with lined Bellows design
120
V.E. Lined Bellows Exhaust Manifold Series 60 ‘Chirp’
Noise – Sound Quality Comparison
Measurement: APS
Amplitude Type: rms
dB
Graphic: Sum level
Statistical Parameter:
Truck 904 Original Configuration Neutral Run Up 2_pp Interior Mic_DRE S
Truck 904 With Lined Bellows Neutral Run Up_pp Interior Mic_DRE S
75
SOUND
PRESSURE
LEVEL
HIGHER ON
ENGINE
WITH BELLOWS
LINERS
70
65
700
800
900
1000
1100
1400
1500
1600
1700
1800
1900
2000
2100
RPM
Graphic: Instationary loudness
Statistical Parameter:
Truck 904 Original Configuration Neutral Run Up 2_pp Interior Mic_DRE S
Truck 904 With Lined Bellows Neutral Run Up_pp Interior Mic_DRE S
40
WITH
LINERS
1300
Measurement: Instat. loudness
Amplitude Type: lin
sone
WITHOUT
LINERS
1200
OVERALL
LOUDNESS
IS ALMOST
EQUAL WITH OR
WITHOUT LINERS
35
30
25
20
700
800
900
1000
1100
1300
Measurement: Sharpness
Amplitude Type: lin
acum
1.35
1200
1400
1500
1600
1700
1800
1900
2000
2100
RPM
Graphic: Sharpness
Statistical Parameter:
Truck 904 Original Configuration Neutral Run Up 2_pp Interior Mic_DRE S
Truck 904 With Lined Bellows Neutral Run Up_pp Interior Mic_DRE S
SHARPNESS
CLEARLY
HIGHER ON
ENGINE
WITHOUT
LINERS
1.30
SHARPNESS
1.25
1.20
1.15
1.10
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1800
1900
2000
2100
RPM
121
V. Proposed NVH Features

Camshaft Anti-Lash Gear
 Split Gear
 Pendulum Damper Gear
 Viscous Damper

Gear Case Cover
 Constrained Layer Damping

Bedplate Block Stiffener

Acoustically Treated Oil Pan
122
V. Passby Noise Engine Sources
Engine - Other Noise
Sources
20%
Engine - Front of Engine
44%
Engine - Oil Pan
36%
123
V. Gear Train
Camshaft Gear NVH Improvement Methods
Split gear
-reduces backlash between
gears through absorption of
torque reversal by the spring
mechanism
Viscous damper
-reduces torque reversals in
gear train via viscous fluid
internal to damper
Compact gear train
-most dominant noise source of the S60 engine
124
V. Gear Case Cover
Constrained Layer Damping Treatment
-function of treatment is to dampen the resonant vibrations of the cover in areas of higher amplitude
125
V. Acoustically Treated Oil Pan
-acoustical barrier intended to reduce amount of radiated noise emitted from oil pan surfaces
-preliminary results show a reduction of 2 dBA in passby noise compared to production oil pan
126
V. Bedplate Block Stiffener
-bedplate is intended to stiffen crankcase to reduce structural vibration and shift natural frequencies out
of normal operating engine frequency range (> 200 Hz)
127
VI. N&V Sensors
A. Accelerometers
B. Microphones
C. Rotational speed transducers
D. Transducer calibration
129
VI.A. Accelerometers

Piezoelectric element measures the motion of a seismic mass.

Some utilize an external amplifier (like those supplied with the MTS
4100) while others have a miniature amplifier built in. These ICP
accelerometers do not work with the MTS 4100
130
VI.A. Accelerometers

Accelerometers may be
either longitudinal or shear
type.
131
VI.B. Microphones

The DDC MTS 4100 kit contains a microphone. The microphone has a
piezoelectric sensing element, and it uses an ICP (Integrated Circuit Power)
conditioning unit. This type of microphone is very durable and temperature
resistant.

The piezoelectric element produces an electrical charge proportional to the
applied acoustic pressure. This charge is amplified by a small circuit (called
a preamp) located near the piezoelectric element. The preamp circuit is
powered by a DC voltage that is carried over the same wires as the signal
between the signal conditioner and the preamp. The signal conditioner
supplies the DC voltage (ICP) and separates the AC signal from the DC
voltage using a high pass filter.

Other types of microphones (including condenser microphones, dynamic
microphones, and crystal microphones) are available from local sources.
These all have different signal conditioning and impedance matching
requirements, and they might not be compatible with the MTS 4100.
132
VI.C. Rotational Speed Transducers

DDC N&V engineers measure
torsional vibrations by mounting
a toothed gear and using a
magnetic pickup (mag-probe) to
produce a pulse every time a
tooth passes the probe.

The data is processed by
examining the frequency
fluctuation of the signal – the
frequency modulation (FM). The
MTS 4100 analyzer does not
currently support this function.
133
VI.C. Transducer Calibration

Calibration values are provided for the accelerometers and
microphone provided with the MTS 4100. These are entered using
the procedures in the next major section of the course. (for
example, Setup Menu>Calibrate Sensors>Microphone A –
sensitivity)

For microphones and accelerometers, standard calibrators are
available so that the transducer calibration can be checked before
and after each use. These are not included in the MTS 4100 kit.

Depending on company ISO requirements, the MTS 4100 and its
transducers should be returned to ETAS periodically for calibration.
Probably once per year.
134
Quiz
1. The MTS 4100 and almost all analyzers calculate the frequency
spectrum (magnitude vs frequency) using:
A. Fourier Transform
B. Discrete Fourier Transform (DFT)
C. Fast Fourier Transform (FFT)
D. Third-Octave Band Analyzer
2. Based on the single degree of freedom (SDOF) model, what property
has the strongest effect on the amplitude of vibration at the natural
frequency:
A. Mass
B. Stiffness
C. Damping
136
Quiz
3. The dynamic behavior of which of the following physical systems can
be better understood using the SDOF model:
A. Engine on mounts
B. Cab on mounts
C. Steering column
D. Exhaust system
E. Acoustic modes of cab
F. Rear cab panel
G. A and B because the model is only useful for rigid body modes
H. All of the above
137
Quiz
4. Based on the single degree of freedom (SDOF) model, mount isolation is
best:
A. Below the natural frequency
B. Above the natural frequency
C. At the natural frequency
5. Which of the following is NOT a source of vibration per the
source/path/receiver model
A. Resonance of the exhaust system
B. Engine vibration
C. Road inputs to the suspension
D. Driveline imbalance
E. Tire force variation
F. Wind noise
138
Quiz
6. Why is noise data A-weighted:
A. To correct for the change in the speed-of-sound due to varying
air temperature/density.
B. To approximately mimic the frequency sensitivity of the human
ear so that measured sound levels correlate to perceived sound
levels.
C. To approximately mimic the change in frequency sensitivity of
the human ear as loudness changes.
7. Which increase in noise level is perceived as twice as loud:
A. 40 dBA to 50 dBA
B. 40 dBA to 80 dBA
C. 40 dBA to 46 dBA
139
Quiz
8. The difference between barely perceivable sound and painful sound
is 3 million times. The difference between barely perceivable and
intolerable vibration is:
A. 3 million times
B. 10 times
C. Approximately 100 times
9. The Series 60 engine is a six cylinder engine with even firing
intervals. What is the order of the engine firing frequency:
A. 6th order
B. 3rd order
C. 1st order
140
Quiz
True or false:
10. Driveline angles don’t matter because the Cardan Joint is
flexible.
11. In the Campbell Diagram, resonances are lines at a particular
frequency.
12. In the Campbell Diagram, orders are lines at a particular
frequency.
13. Critical speed is the nominal design condition for a rotating
shaft.
14. The MTS 4100 reads vehicle speed information from the
computer bus to calculate driveline rotation speed.
15. Cardan joints produce 2nd order driveline vibration.
16. Driveline imbalance produces 2nd order driveline vibration.
141
Quiz Answers
1C. The FFT is used to calculate the frequency spectrum due to its
computational speed.
2C. While the mass and stiffness determine the natural frequency, the
amplitude is determined by the damping.
3H. All dynamical systems that are approximately linear can be better
understood using the SDOF model. All these systems have natural
frequencies with associated mass, spring, and damping properties.
4B. Isolation is best above the natural frequency. As the frequency
approaches zero, the isolation is one. At the natural frequency,
there is gain (for lightly damped systems which are most common),
so the isolation is poor.
5A. Resonance of the exhaust system is a path issue, all the others are
sources of vehicle vibration.
6B. A-weighting corrects for the frequency response of human hearing
at lower sound levels. B-weighting corrects for intermediate sound
levels, and C-weighting corrects for high sound levels.
142
Quiz Answers
7A. An increase in loudness of 10 dBA is perceived as twice as loud.
An increase of 6 dBA corresponds to a doubling of sound pressure.
Increasing the sound level by 40 dBA would be perceived as being
16 times louder (= 2x2x2x2).
8C. The difference between barely perceivable and intolerable vibration
is approximately 100 times. Vibration perception is approximately
linear, while sound perception is compressed logarithmically.
9B. A single-cylinder, four-stroke engine has a firing order of ½ because
the cycle of intake-compression-combustion-exhaust takes two
revolutions to complete. This process takes place six times faster
for a six cylinder engine with even firing intervals, so the order of the
firing frequency is 3rd order (= 0.5*6).
10.F
11.T
12.F
13.F
14.T
15.T
16.F
143