The document is a guide on interpreting vibration frequency and time spectrums for condition monitoring in machinery, utilizing case histories from 35 years of the author's experience. It focuses on teaching readers to recognize faults through vibration analysis rather than how to fix them, emphasizing the combination of science and intuition in the field. The publication includes practical guidelines for setting up monitoring programs, data collection strategies, and insights on various diagnostic technologies.

1. 1.1.1.
A Guide to the Interpretation of
Vibration Frequency and Time Spectrums
Summary
The purpose of this publication is to teach the reader how to read and understand
frequency and time domain spectrums collected in a condition monitoring vibration
program. This will be accomplished using case histories collected by the author over the
past 35 years in hundreds of locations around the world. Note that this book will not teach
you how to balance a machine or perform a precision alignment, instead it will provide you
the understanding necessary to recognize that these faults exist. Just as you study a
language to learn the various sounds and their meanings (FFT spectra being a visual
language), with this book you will learn a new language and what those wiggly FFT lines
mean.
SKF @ptitude Exchange
SKF Reliability Systems
5271 Viewridge Court
San Diego, CA 92123
United States
tel. +1 858 496 3400
fax +1 858 496 3511
email: info@aptitudexchange.com
Internet: http://www.aptitudexchange.com
SKF-SDG-DS01
Dr. Robert M. Jones
118 Pages
Published Sept 2011
A Guide to the Interpretation of Vibration Frequency and Time Spectrums © 2011 SKF Group

2. A Guide to the Interpretation of
Vibration Frequency and Time Spectrums
by
Dr. Robert M. Jones
January 10, 2011

3. 2

4. 3
Table of Contents
Table of Contents 3
Preface 5
Author’s Notes to the Reader 7
Introduction 11
Bearings 17
Balance Conditions 45
Misalignment 49
Gear Boxes 53
Electric Motor Analysis 61
Resonance 71
Enveloped Acceleration (gE) 81
Spectral Emitted Energy (SEE) Technology 85
Mechanical Looseness 91
Leak Detection 95
Detecting Rubs 101
Blade Pass 105
Unusual Situations 109
Conclusion 117

6. 5
Preface
The purpose of this publication is to teach the reader how to read
and understand frequency and time domain spectrums collected in
a condition monitoring vibration program. This will be
accomplished using case histories collected by the author over the
past 35 years in hundreds of locations around the world. Note that
this book will not teach you how to balance a machine or perform
a precision alignment, instead it will provide you the
understanding necessary to recognize that these faults exist. Just
as you study a language to learn the various sounds and their
meanings (FFT spectra being a visual language), with this book you
will learn a new language and what those wiggly FFT lines mean.
Vibration Analysis is a mixture of science and art. The math and
calculations are cut and dry, but the interpretation requires some
amount of intuition and experience. Three important points you
should remember and live by are:
 Learn how to say “I don’t know”
 Never say “Never”
 Never say “Always”
First, realize there will probably be some unknown factor in your
analysis. If you don’t tell the truth and instead make up some
answer that is later found to be false, you will lose your
creditability and your program will never recover. Second and
thirdly, there are nearly always exceptions to problem solutions,
and you shouldn’t block out that possibility. Sooner or later that
exception will appear at the most unexpected and unfortunate
time.
Vibration analysis is an interesting field. It requires a person with a
large amount of curiosity who enjoys problem solving and has a
willingness to keep looking when the answer isn’t readily available.

7. 6
The intent of these examples, insights and suggestions is to help
you in your job.
You will notice that there are not a lot of mathematical formulas
used in this text. Those may be found in many other publications,
you only have to use the internet with a good search engine to find
whatever formula you are seeking.

8. 7
Author’s Notes to the Reader
When I first started collecting vibration data, nearly everything was
measured in displacement (mils). The big problem with
displacement is that it is speed dependent for severity
determination (i.e., 6 mils at 300 RPM wouldn’t be a problem, but
6 mils at 3600 RPM would be a serious problem). So, just knowing
the amplitude didn’t provide enough information for a complete
analysis. Then, in the 1980’s, velocity (IPS) became the
measurement of choice because it is not speed dependent. A
vibration of 0.35 IPS is considered rough at 300 RPM or at 3600
RPM. The problem then became one of communication with the
machinists, because they could understand 6 mils, they knew that
meant the part was moving 0.006” and could be read on a scale.
However, there was no way they could read the same problem in
velocity. They knew that the part was moving at 0.35 inches per
second and that was bad, but they didn’t have a scale in their
pocket that would explain that information. Time and teaching
solved the communication problem and we slowly entered a new
era of vibration analysis.
As accelerometers became cheaper, smaller and more reliable, the
big velocity pickups filled with oil and internal moving parts, and
which had both low and high frequency limits, became rare in our
toolkits. Since Force = (mass)(acceleration), we now had an
accurate way to measure the forces taking place. For example,
forces inside a gearbox between the bull gear and the pinion gear.
However, measuring acceleration introduced another
measurement that was not intuitive and we had another
communication problem since our machinists didn’t have a visible
scale to measure acceleration. Again teaching and experience
solved the problem.
As we continued to ply our trade, the machinists learned to trust
us when we looked at the wiggly lines on our screens and said that
the bull gear was damaged, or that even though they had just
aligned the pump and motor, it still was not good enough and

9. 8
needed to be done again or the life of the bearings would be
adversely influenced. Also, our equipment became more compact,
more accurate and finally a “portable” tool that was really portable.
The advent of digital storage allowed us to carry the data back to
the office and make hard copies at our leisure instead of out on
the factory floor. Life was good.
Then in the early 1990’s, SKF released two more field
measurement tools, enveloped acceleration (gE) and spectral
emitted energy (SEE). I was working with the research and
development group and took the first data collectors out for field
testing. The learning curve was rather steep, and as we went
along we had to rewrite some of the rules we developed back in
the lab.
gE was another communication problem, not only did it involve
acceleration signals, it took that information and mathematically
processed it into an artificial display that required some knowledge
of signal processing to be able to explain the results. The challenge
was met and today it is the premier method of bearing and
gearbox analysis.
SEE technology is another mathematical based tool developed by
SKF. SEE technology uses an acoustic accelerometer that “hears”
the acoustic signals generated by the rolling elements in bearings
as they pass over damaged areas - and then processes the sounds
into a display similar to the familiar FFT. SEE and gE are unique
technologies that DO NOT relate or compare to any other
measurement methodology, and do not convert to any other
measurement types.
As you will note as you use this training tool, most of the case
histories in this book use velocity, acceleration and enveloped
acceleration for analysis. Other tools are available, but these are
the prime tools in use today. Also, the FFTs that are displayed as
field examples have been collected over many years using a
number of data collectors and software programs. There will be
some variation in their appearance, but the information remains

10. 9
valid. You may also notice that some important suggestions and
tools are repeated in several sections for emphasis.
Dr. Bob
bobnewnan@msn.com

11. 10

12. 11
Introduction
If you are just starting a vibration program and taking data for the
first time, you may feel like there is just too much to learn and do.
Well don’t feel alone, when I started some thirty years ago I had
the same feelings. What you want to do is start small and add a
few machines every month or so. As you learn to trust your
equipment and your judgment, you will find that life gets easier.
We started with the sixty machines mandated by the Nuclear
Regulatory Commission (NRC), and when I left the company nine
years later we were monitoring over 400 pieces of equipment and
had written acknowledgement from the NRC that we had one of
the best vibration monitoring programs in the U.S. nuclear power
industry.
In some instances, the vibration program has management
support, which is the best situation. In others you will find yourself
doing a job because some industry regulation requires it or
because it has been there for years and you are the new guy in
town, so you get to do it. If that is the case, you have two jobs, one
to work the vibration program and the other to show management
that, if done properly, the vibration program can contribute in a
positive way to bottom line profit. In my last year as the vibration
and acoustic engineer in a nuclear plant, independent auditors
documented that the vibration program saved the company
$640,000. This was a good return on their investment in one
engineer, two aides and their equipment.
General Guidelines for Setting Up your Program
Note that the following guidelines refer to software and hardware
developed by SKF, other monitoring equipment will follow the
same principals, but their displays and setups may be different.
Hopefully, you will be sent to software and hardware training. If
that is not possible, see if a trainer can be brought on site. Then,
read the hardware and software manuals. I’m serious! You will

13. 12
eventually figure it all out, but your results will have an influence
sooner if you are trained and understand the equipment.
Your greatest impact will be on those machines that, if they fail,
will cost the company money, often referred to as critical
machines. Survey the plant and list all of these machines;
machines that do not have spares or are one of a kind, machines
where any part of the production process passes through, and
machines that control some part of the process. Next, list all of the
machines that are spared but don’t automatically come on if there
is a failure of the primary equipment. The next list will be
machines that are spared and have automatic backup. Then,
finally there will be the rest of the plant equipment such as roof
fans, air conditioners, air compressors that are not associated with
process, along with other general rotating equipment.
Unless there are regulatory time requirements, most critical
machines are monitored once per month. Non-critical machines
are monitored every two months, and the balance of plant
equipment is monitored once per quarter. Again, there will be
deviations. In the power plant, we were out of business if the
reactor cooling pumps or the turbine/generator had problems.
Therefore, we monitored them once per week. You have to be
flexible and adapt to the situation where you are working. On the
balance of plant equipment, we checked them on a schedule that
started at the top of a list and worked to the bottom, then back to
the top and do it again. Depending on the workload, it might take
longer than a quarter to work the whole list, but there is not a
time constraint on balance of plant equipment.
Next, you have to build the route that you will walk as you collect
the data. Put it together so that each machine follows in a logical
order. Don’t do something in building A, then head over to building
B, then back to A, etc. For our example, we will use a typical
horizontal pump and motor set. In the original point setup, three
transducer location points will be set up for each bearing; vertical,
horizontal, and axial. Always take the data in the same order on

14. 13
each machine. For this example, start at the outboard motor
bearing, then the inboard motor bearing, go across the coupling to
the inboard pump bearing and then the outboard pump bearing.
Of course, in real life some of these positions may be covered by a
safety cage or insulation and you’ll have to modify as necessary,
and it may not be possible to collect data from every point. Label
and mark the points so that you can return to the same location
each time.
A Word about How Much Data to Collect
It is a true statement that most vibration programs collect too
much data. Part of this is the fear that something will happen and
it will not be documented, and the other part is because some
people just do what they are told by an instructor without
reasoning why.
For example, in our hypothetical pump and motor set, assume we
are concerned about a reoccurring unbalance problem. After the
initial data set is taken on all the bearings, the same twelve or
more points continue to be taken month after month. Balance is a
radial force and the horizontal position is usually more sensitive
than the vertical since the machine is bolted down and gravity is
assisting in the stability. Try it, you can shake a table back and
forth much easier than up and down! So, if you are concerned
about unbalance, after the initial data collection, use the software
to disable the vertical and axial readings and only take the
horizontal velocity reading on each bearing - and monitor the
frequency associated with rotation, (1X). This saves you the time
required to take vertical and axial readings on all four bearings.
Later, if you see an unbalance problem developing, or your trends
show an upward slope, you can enable the points and again take
vertical and axial readings and see if it is increasing in amplitude
as it is in the horizontal. (It should be increasing at a lesser rate.)
It is highly recommended to take the axial readings to monitor for
misalignment, but I assure you that, except for very rare cases, if
one end of a motor is moving in an axial direction, so is the other

15. 14
end, so it is not necessary to take axial readings on both ends
every time the route is collected. It’s interesting to note the one
occasion where this was not true. It was a case where the shaft,
which was enclosed in a collar, had broken inside the collar. The
two ends were held together and the shaft rotated, but axial forces
caused the readings to be different at each end. This was our first
clue that something was amiss.
In like manner, for monitoring bearing conditions on a normal
route, it is only necessary to take one enveloped acceleration
reading in the horizontal direction, per bearing, usually in the
same location as the velocity measurement. Additional readings on
the same bearing will not increase the likelihood that you will find
a bearing fault. Of course, if you are monitoring a thrust bearing,
take your enveloped acceleration readings in the axial direction.
Suggestions for Building the Points
Data is taken in two modes, the normal route data collection
and/or in an analyzer/trouble shooting mode. The point setups are
usually different for each mode. In the route mode, you are
primarily collecting data for trending. When an undesired trend
develops or you see other changes in the spectrum, then you
switch to trouble shooting to determine what has caused the
change. For example, the lines of resolution setting for route data
is usually set at 400 lines. This will provide you with enough
resolution to detect changes and monitor balance and alignment.
However, if you wish to verify sidebands, it may be necessary to
use 800 or 1600 lines. Remember that in the route mode, if you
use 800 lines, then it will take you twice as long to collect the data
as with 400 lines.
The same applies with averages, eight averages will double your
time to collect the data versus using four averages, and nobody
ever has enough time! In setting the Fmax, time and resolution
are again factors. If a high Fmax is set, then the data is collected
quickly but the information may be so compressed at the low end
of the spectrum that it is difficult to interpret. Most operators will

16. 15
set the initial Fmax between seven to ten times rotation speed.
This can always be modified as necessary. In velocity readings, the
Fmin should be set to 10 Hz to remove any ski slope that will
influence the overall amplitude, unless of course the rotation
speed is less that 600 RPM. If so, the Fmin must be adjusted. For
enveloped readings, the Fmin should be set to 0, as enveloped
signal processing has no integration and does not develop a ski
slope.
On each of these data collection points, the transducer can be
programmed to take as many types of readings as you wish.
Starting with the vertical position, you can take a velocity reading,
then a displacement reading. Then, in the horizontal position, you
may wish to take these same readings plus an enveloped
acceleration reading. The software will drive the data collector to
do what you tell it to do. Then, take your axial reading and move to
the next location, inboard motor bearing. Continue this until all the
readings are taken. There is more information on what type of
data to collect in the examples that follow of various problems and
their solutions.
Upon completion of the route, the data is uploaded to the
computer. At this point you need to change your thinking about
the computer. Now, think of it as a file cabinet that has the ability
to self examine all of its files. Once you have collected enough
data, at least six data sets, the software can develop trends of
each data type taken. In the point setups, the alarms are set so
that if a point exceeds the preset amplitude, the traffic light on the
software route data point will show yellow for 1st
alarm and red
for the 2nd
alarm. This means that you do not have to look at all
those hundreds of spectrums that you have collected, just those
that flash yellow or red alarms. There are a number of alarms that
you can activate; trend, amplitude, per cent change, etc. You can
use the default alarms in the beginning and change those that
need to be changed depending on machine requirements. For
example, a precision lath has much lower alarms than a woodshop
lath. Alarm amplitudes are set so that you are alerted before any

17. 16
vibration will influence the quality of the product or life of the
machine.
As you develop your skills, you will become an important part of
the maintenance team. They will depend on you to prevent those
midnight surprises that occur when a critical component fails. You
will contribute to the profitability of the company and become a
valuable employee.
With that in mind, provide your management with a monthly or
quarterly report on what you have reported and the results. It is
generally accepted that a machine failure will cost ten times as
much to repair as one that is shut down under controlled
conditions. Labor time cards are official documents. It is easy to
find how many hours were used for a repair. Multiply that number
by 10, subtract the hours reported used and you have a good
estimate of hours saved. Multiply that by the average wage and
you can logically claim to have saved that much money, to say
nothing of collateral damage when a failure occurs. Often people
don’t want to “ring their own bell” so to speak, but I assure you
that if you don’t, no one else will! Let the boss know what you are
doing and I speak from experience, he will pass the information to
his boss!

18. 17
Bearings
All rotating equipment has one thing in common, bearings. In
monitoring the health of our equipment, the condition of the
bearings is of prime importance. Consider that if the bearings are
good, even an out of balance, misaligned machine will operate.
However, if the bearings are damaged, the machine will soon fail,
even if properly assembled and balanced. Today’s technology has
developed new techniques for non-intrusive determination of the
bearing’s physical condition.
Early bearing fault detection usually consisted of an experienced
mechanic positioning a screwdriver on the bearing cap and the
other end of the screwdriver on the bone structure behind his ear.
Sometimes he was right and sometimes he was wrong, but he
was typically always too late. With the advent of portable vibration
measuring equipment, some operators noticed that the high
frequency energy generated by a failing bearing would excite the
natural frequency of the bearing. They would call a bad bearing
based on this information. Again, sometimes they were right and
sometimes they were wrong.
The next step in this evolution was to look for specific frequencies
generated by the bearing elements as they rotated, using velocity
measurements. With this improvement, the accuracy increased
but even the good technicians would often miss bearing flaws on
the very slow rotating machinery, considering anything below 100
RPM to be slow. In the last few years, with the inclusion of
enveloping algorithms, the accuracy has improved to nearly 100%.
A few bad bearings still get misdiagnosed, but they are rare.
The techniques explained in this section apply to all rolling element
bearings and both steel and hybrid ceramic bearings. These
techniques can also provide some information about the condition
of sleeve or journal bearings. Also, they apply to all bearing
manufacturer’s products. What is unique is that each vibration
data collector manufacturer uses different algorithms in
processing the electronic signal generated by the accelerometer.

19. 18
Therefore the results and reliability of other data gathering
equipment may not be equal to that used by the author.
The mathematical processing of an electrical signal, known as
enveloping, has been in existence for a number of years in the
laboratory. A simple explanation for the process is that by using
selective high frequency bypass filters, enveloping mathematically
enhances the repetitive signals generated as the rotating elements
pass over a damaged area, while at the same time degrades the
non-repetitive signals. This processed signal is demodulated and
presented to the user in the frequency range he desires.
Therefore, if you have a pump with a bad bearing, the bearing
signals, which are repetitive, will be enhanced while the non-
repetitive flow and possible cavitation noise will be degraded. If the
reader is interested in the mathematical functions required to
develop the enveloped signal, please refer to a signal processing
manual.
Journal or Plain Bearings
Bearings without rolling elements, where the shaft is rotating
inside a metal sleeve, are journal bearings. The bearing material is
poured into the housing and machined to tolerances determined
by the shaft size and other factors. The space between the shaft
and journal is filled with lubricant. As the shaft rotates, it moves
inside the bearing so that when in position, it is supported by a
wedge of oil and rotates in a stable condition. As a journal bearing
wears and metal is removed from the journal, the shaft loses this
support and will begin to move depending on the inherent
imbalance that is in most rotating elements. This movement will
be seen in the FFT as an increase in the rotational speed signal
(1X) and may be interpreted as just an out of balance condition.
The preferred way to monitor the condition of journal bearing is
with oil analysis. Based on this analysis report, in conjunction with
the vibration trends, a proper evaluation of the bearing’s condition
is possible.

20. 19
Rolling Element Bearings, Common Analysis
Since one of the analysis techniques involves trending of vibration
levels, it is important that the data collection location be marked,
and the same location be consistently used each time data is
collected. Common to most modern portable electronic data
collectors is the accelerometer. These sensors are generally
constructed with a man-made piezoelectric crystal that generates
an output voltage directly proportional to the amount of force
applied. The accelerometer is usually placed on the bearing cap, or
if not accessible, as near as possible to the bearing.
In those instances where it is not possible to safely position the
accelerometer by hand, the accelerometer may be permanently
stud mounted or glued to the machine and the signal wire
terminated in a safe location. Generally, the accelerometer will be
mounted using a magnet. Both methods are acceptable for
general vibration monitoring. In rare instances, a stinger (a steel
rod threaded onto the end of the accelerometer) may be attached
to the accelerometer to reach a bearing cap located in a tight
space, but stingers will alter the signal amplitude and frequency
and are not recommended for general usage.
For continuous monitoring of a machine, all of the points of
interest use a stud or epoxy mounted accelerometer. The signal
wires are then terminated at a common point where they are
multiplexed and routed to a permanently mounted data collector.
The signals from the data collector are then passed to a computer
controller which is programmed to store and process the data.
One accelerometer signal can be processed into four
presentations; acceleration, velocity, displacement, and enveloped
acceleration - and these may be processed for different frequency
ranges as needed. In other words, the velocity signal may be
presented in one spectrum from 0-30 Hz to check for balance and
alignment. A second spectrum may be generated with a range of
0-1000 Hz to disclose the rotor bar pass frequency or to check for
bearing degradation. In addition, other types of sensors can collect

21. 20
operational data such as shaft position, speed, temperature, flow,
pressure, etc. Generally any sensor that provides a voltage output
can be monitored, either with portable or permanently mounted
data collectors. The signal can then be collected and stored for
evaluation.
A rolling element bearing consists of four parts. The parts and
their fault abbreviations are; the outer ring, Ball Pass Frequency
Outer (BPFO); the inner ring, Ball Pass Frequency Inner (BPFI); the
cage, Fundamental Train Frequency (FTF); and the rolling
elements, Ball Spin Frequency (BSF). Each of these components
will generate a unique frequency which will vary in relation to the
shaft speed. Therefore, an accurate rotation speed is necessary to
perform the analysis. Over time, the bearing parts will wear and
the observed frequency may vary slightly from the calculated
frequency. As can be seen in the following frequency calculation
formulas, the frequency generated is based on the number of
rolling elements, the shaft rotation speed, ball diameter, pitch
diameter, and the contact angle. These formulas are provided,
although all of the data collection software on the market will do
the calculations if the user provides the bearing ID (i.e., an SKF
6309 or FAG 22222).
Bearing Frequency Formulas
(1) BPFO = (N/2) (RPM/60) (1-Bd/Pd) (cos )
(2) BPFI = (N/2) (RPM/60) (1+BD/Pd) (cos )
(3) BSF = (N/2) (RPM/60) (1-[BD/PD]2
) (cos)
(4) FTF = (1/2) (RPM/60) (1-Bd/Pd) (cos)
Where:
N = Number of balls or rollers
Bd = Ball diameter (in or mm)
Pd = Bearing Pitch diameter (in or mm)
 = Contact angle, ball to race

22. 21
These formulas are for the bearing mounted on the shaft and a
rotating inner ring. If the outer ring is rotating, reverse the (+) and
(–) in the formulas. Another handy rule of thumb to use when you
are in the field and the computer is back at the office and you need
a close approximation is:
BPFO = (RPM) (n) (0.4)
BPFI = (RPM) (n) (0.6)
The first four formulas will give the frequency results in Hertz (Hz).
Hz is cycles per second. If you desire them in cycles per minute
(CPM), multiply by 60.
Vibration amplitudes are measured in the following units:
Displacement (distance) is measured in Mils, one mil equals 0.001
inches. Metric measurements are in millimeters.
Velocity (speed) is measured in Inches Per Second, IPS. For
metrics, the units are mm/sec. For a quick approximation, 1
mm/sec equals 0.04 IPS. 25.4 mm/sec equals 1 IPS.
Acceleration (force) is measured in G’s, for both English and
Metric units.
An acceleration signal may be integrated to produce a velocity
signal, and double integrated to produce a displacement signal.
Most software programs will do the computations.
Enveloped Acceleration is a special measurement of acceleration
(gE), and there is no comparison or conversion to the standard
measurement methods.
Signal Processing
Although this manual is not about signal processing, it is necessary
to examine some of the characteristics of the process. All of the
major data collectors receive the accelerometer signal and either
store, or display it, as a time vs. amplitude signal. If you were
looking at an oscilloscope, the type of signal you would see would
show amplitude on the “Y” axis and time on the “X” axis. In order

23. 22
to see this same presentation in the frequency domain, a Fourier
transform must be applied to the energy shown in the time
display. The result is a display with the amplitude again in the “Y”
axis, but the “X” axis is now displayed as a frequency range which
the user can select to measure in either Hz or CPM.
For history buffs, Jean Baptiste Fourier was a famous French
mathematician who developed the basic theories for signal
analysis. One great help in using an enveloped Fourier transform
is that, not only is the very low signal enhanced by the enveloping
process, the display itself provides us with positive evidence of the
presence of bearing damage by overlaying the bearing fault
frequencies.
Although the only place in real life where a pure sine wave exists is
in the laboratory, a loaded rotating bearing will generate an
approximation. If there is no damage, and the bearing is heavily
loaded, the Fourier transform (FFT) may produce a single
frequency spike of energy at the bearing BPFO. The process is
sensitive enough to detect the minute outer ring movement that
takes place as three, then four, then three rolling elements pass
through the load zone. If the bearing is not heavily loaded, no
signal will be generated, so there will be no fault signals in the
spectrum. However, if there is damage, the sine wave is clipped or
truncated. An FFT of a clipped sine wave will result in the
fundamental frequency, BPFO for example, plus harmonics of the
BPFO frequency. The final results are that when the user
examines the FFT display, if there is no BPFO signal or if it is
present and there are no harmonics, then there is no damage in
the bearing. If harmonics of the bearing fault frequencies are
present, there is damage. Then the user has to evaluate these
damage indicators based on amplitude and shaft speed. For
general machine condition, if the FFT displays multiple harmonics
of the shaft rotation speed, this indicates looseness in the machine
parts and not damage in the bearing.

24. 23
A characteristic of all the bearing fault frequencies is that
approximately 90% of the fault frequencies, as a function of
rotation speed, are non-integers of shaft speed. For example, 1X
represents the shaft rotation speed and bearing fault frequencies
are seen as 4.733X or 6.382X. This is important because often the
nomenclatures of the bearings that are installed in older machines
are unknown. After the FFT is collected, if a series of harmonics
are seen and the first of the series is seen to be 4.733 orders of
rotation speed, then it is highly probable that it is generated by a
bearing fault. (This does not hold true for gear boxes with multiple
gears because internal speeds will also be non-integers in
relationship to the input and output shafts.) The other 10% of the
bearings will have frequencies that are very close to an integer of
1X, for example 4.020X or 5.003X and cannot be separated from
harmonics of rotation speed, which are indications of a different
problem.
Until the introduction of enveloped acceleration techniques, the
prime method of bearing analysis was the use of velocity
measurements. The difficulty with this technology is a function of
the mathematical calculations to develop velocity measurements.
Velocity is the product of dividing distance “D”, by time “T”,
resulting in IPS or mm/sec. The internal clearance in a new
bearing is in the order of 0.001,” and this space is filled with the
lubricant. The result is a very small “D” divided by a relatively large
time “T” and the resulting velocity is a very small number. When
the bearing becomes worn and more space is created (i.e., a larger
“D”), then the observer may begin to see velocity results, but by
that time the bearing is usually in trouble and may be nearing
failure. This is certainly the case with very slow speed bearings
where the “T” is even larger.
The introduction of Acceleration Enveloping (gE), with a standard
100 mv/eu accelerometer, provided a most excellent tool for
bearing analysis. The patented algorithm enhances the repetitive
signal when the rolling elements pass over a damaged area. The
same type of signal is produced when a damaged portion of the

25. 24
cage or the rolling element enters the load zone. In this case, the
mathematical calculations work in our favor. Force “F” equals
mass “M” times Acceleration “A”. In this case, we can arbitrarily set
“M” to equal 1, which provides us with F=1A. Using an
accelerometer, we can measure the acceleration forces generated
when any of the elements encounter a damaged area. Then, using
the unique algorithms and FFT processing, we can derive the
enveloped acceleration FFT for bearing vibration analysis.
The enveloping algorithm specifically captures impact signals. Out
of balance conditions do not usually generate this type of signal, so
unless the machine is making a once per revolution impact for
some reason, it is necessary to take a velocity FFT to detect the
rotation speed, 1X. With SKF software, using the “Set Speed”
function under the “Display” icon and manually entering the 1X
measurement, the rotation speed will be imposed on the FFT. This
will annotate the 1X signal and enable the use of the software’s
Frequency Analysis Module (FAM). The FAM markers will not be
calculated and overlaid on the FFT without the 1X information.
Although many harmonics of the fundamental fault frequency can
be selected for display in the FFT, three are enough to provide a
proper analysis.
Again, enveloped acceleration CAN NOT be compared to any other
type of vibration analysis. It is completely unique as a vibration
tool. It is provided in four frequency ranges but experience has
shown that the 3rd
filter is usually the primary band used in
bearing analysis, regardless of the rotation speed. However, the
shaft rotation speed is a direct factor on the severity analysis for
the bearing. In controlled conditions, a damaged bearing rotating
at 50 RPM generated a fault signal of 0.004 gE. Using a DC motor,
the speed was slowly increased to 3600 RPM where the amplitude
was 1.8 gE. (See Fig 1) Therefore, you must be aware of bearing
operating speeds when making your analysis.

26. 25
Figure 1. Relationship between Speed and Amplitude Using gE with an Induced
Fault in the Outer Ring.
Because the amplitude is dependent on the bearing shaft rotation
speed, a one size fits all damage vs. amplitude chart is not
possible. Using field experience, the following guidelines were
developed for bearings operating between 1500 and 3600 RPM:
Amplitude Action taken
0.0 to 0.2 gE No action required
0.2 to 0.5 gE Notify management, increase surveillance
0.5 to 0.8 gE Request bearing change when possible
0.8 to 1.0 gE Request bearing change ASAP
1.0 gE and up Request machine be taken out of service
These are only guidelines, experience and machine criticality will
have to be included as part of the decision process.
With bearings of any speed, the first clue of a problem is a
fundamental fault frequency with harmonics. If there is any
damage, there will be harmonics. The amplitude of the harmonics
is not a concern, only their presence is important. The second
warning sign is the existence of shaft rotation speed (1X) sideband
harmonics around the fault frequency. The presence of sidebands
is usually a sign that the bearing will need to be replaced in the

27. 26
near future regardless of the amplitude of the fundamental fault
frequency. Fault harmonics with sidebands and a raised noise floor
are an indication that the machine should be taken out of service
and the bearing replaced immediately. Note that when a bearing is
badly damaged the unique fault frequencies may not be present
because the entire load zone is engaged and no individual impacts
are occurring. In these cases the noise floor will slowly increase
with a failure to follow soon after.
With very slow rotations, less than 100 RPM, the time domain
should also be saved. Often the FFT signals will be covered by the
noise floor and the time domain will capture the impacts from
which the frequency can be calculated by the software. A standard
100 mv/eu accelerometer, 400 lines resolution and one average
are the usual setup. The range of the FFT should be set to 4-5
times BPFI in order to capture any harmonics of the fundamental
fault frequency. A default Fmax setting for unknown bearings is
10X.
The frequency response from a double row bearing will NOT
change even if both rings are involved. However, the amplitude
will increase as the two amplitudes are added together.
Field Examples
A major manufacturer of diesel engines complained that although
they had a good vibration program and took sufficient data, they
continued to have unexpected bearing failures of the bottom
bearing in the assembly line drive gear boxes. The complexity of
the gear box resulted in the loss of three days production when
these unscheduled outages occurred. The side effect was that they
had to keep at least three days production in the warehouse so
that they could continue to ship “just in time” to the end user.
There was not only the additional cost of this stored inventory,
when a gear box failed it was an around the clock repair effort
with increased use of overtime to return to production.

28. 27
An examination of their data and an interview with the technicians
did not reveal any major errors in their data collection methods or
their analysis techniques.
We then collected data on one of four gear boxes that had caused
them unscheduled outages. To show them the advantage of
enveloped acceleration, we took 3 separate readings from the
same location within seconds of each other, velocity, acceleration,
and enveloped acceleration. The assembly line was in full
operation and the speed of the drive shaft varied between 8.3 and
8.5 RPM.
Figure 2. Velocity, 8.4 RPM.

29. 28
Figure 3. Acceleration, 8.5 RPM.
Figure 4. Acceleration Enveloping, 8.4 RPM with BPFI Marker on Fundamental
Frequency.
The value of enveloped acceleration is immediately apparent. The
velocity FFT gives absolutely no indication of a problem developing
with this bearing. The acceleration FFT does have some
indications, but who is concerned when the amplitude is less than
0.00005 G’s! The enhanced presentation of gE clearly shows a
problem with the inner ring of this bearing. Based on the chart in

30. 29
Figure 1, the amplitude at 8.4 RPM of 0.006 gE would certainly
indicate this bearing required attention.
Another bearing on a transfer shaft was tested. According to their
records a different bearing was installed.
Figure 5. Velocity, 8.3 RPM.
Figure 6. Enveloped Acceleration, 8.3 RPM.
Again the enveloped acceleration technology shows its value. In
the velocity there is absolutely no indication of a problem. Also in
comparing the two gE spectrums, Figure 4 and Figure 6, note how

31. 30
much higher the second noise floor is as compared to the first.
This is a strong indication that this bearing is entering the final
stages of its life.
We were not able to remain on site during the repairs but a
follow-up call reported that the shutdown was scheduled and the
bearings were changed. Both bearings showed extensive spalling
in their respective races. The vibration group immediately added
enveloped acceleration into their routes. Within six months the
operations group was confident enough with their analysis that
they were able to close out the three day backup storage facility.
There have been no unplanned outages because of gear box
bearing failures and they reported a maintenance and operations
savings in excess of $1,000,000 for the year.
Cage Problems
It is all too common to observe at a new construction site, new
production equipment sitting at various locations covered with
plastic or a tarp. They have arrived before the building was
completed so are stored in the field. If this occurs over an
extended period of time, the end results will be damaged bearings.
No matter what time of the year, metal gets warmer in the
daytime and cooler at night, producing condensation. When this
condensation occurs inside the bearing, trouble begins in two
forms. First, the hydrogen molecule in the water has a strong
affinity to metal and attaches to the steel molecules resulting in
hydrogen embrittlement. Second, the oxygen oxidizes the metal
and you have rust. Then, several months later when the
equipment is installed and activated, loud grinding and scraping
noises emit from the bearings. This was the case at a new plant in
Richmond, Virginia. They were able to obtain seven of the needed
eight replacement bearings from the local bearing shop but could
not locate the eighth from any of the bearing shops in the area. In
desperation they obtained a bearing from a used parts shop and
proceeded with the installation. When this machine ran, it was

32. 31
much noisier than the other three and we were called in to
determine the cause.
We were told that the bearings were SKF 22222’s and that the fan
speed was about 1600 RPM. Figure 7 is the frequency spectrum
we collected on the suspect bearing. Using FAM, we can overlay on
the spectrum the frequency markers for each of the bearing
components. What is immediately seen is that the cage frequency
(FTF) lines up with an energy spike. For clarity, the other three
bearing frequency markers are not shown. The secret to
frequency analysis is identifying the sources for the energy seen in
the spectrum. In this case, the only thing in this machine that
would generate 675 CPM signal is a damaged cage in an SKF
22222 bearing. Note the Single Value cursor marker, cage
frequencies are always less than 1X.
Figure 7. FFT of Damaged Cage.
Based on this analysis, the bearing was removed and inspected.
Figure 8 is a photograph of the bearing showing the damaged
cage. Using the serial number on the bearing, it was determined
that it was over 21 years old! Sometime during its life, someone
had struck the brass cage and deformed it, either during an
installation or in removal. The deformed portion of the cage was
making point contact with the inner ring and generating the 675
CPM signal.

33. 32
Figure 8. Photograph of Damaged Cage.
This case illustrates how we find damaged components using
frequency analysis. It also points out the need to use care where
you purchase your bearings even if you are under pressure to get
the machine back in service. All major bearing manufacturers
provide customer training on the care and handling of rolling
element bearings. Somewhere in the past, someone had not
learned that you should not mount and dismount bearings with
hammers and drift pins.
Cracked Inner Race
There are very specific tolerances for bearing fits on the shaft and
in the housings. If they are followed, you can expect a long life
from the bearings. If, as in the next example, shaft fits are not
maintained, the results can be disastrous.
If a bearing is loose on the shaft, it will slowly rotate. The friction
will generate heat, which in turn will cause the shaft and inner ring
to expand. In this case, the shaft expanded more than the ring, to
the point where all the fit tolerances were exceeded and then the
ring cracked. This same condition can occur in situations where
the shaft is hollow and is heated with steam. If the heat is applied
too fast, the shaft will expand faster than the bearing inner ring,

34. 33
resulting in a cracked ring. Figure 9 is the enveloped spectrum
collected while the unit was in operation.
Figure 9. FFT of Cracked Inner Ring.
The owner said that the unit was operating at 1200 RPM and the
installed bearing was an SKF 2222. When we first looked at this
spectrum without the bearing frequency overlay, it appears that
we have multiple harmonics of the shaft speed, 1203 RPM which
would indicate looseness in the machine assembly. Figure 10
shows the value of further evaluation.

35. 34
Figure 10. Suspect Bearing with Bearing Frequency Fault Markers.
The bearing frequency overlay clearly shows us that we have a
problem with the inner ring. We can see the fundamental inner
ring frequency with sidebands of shaft speed. Inner ring defects
have a unique characteristic in that they will almost always
produce sidebands of the shaft speed. Using the software, we can
overlay sideband markers and see that they are the shaft speed.
These sidebands are created by the natural modulation caused by
the damaged area passing in and out of the load zone generating
a rising and falling signal.

36. 35
Figure 11. Suspect Bearing with Shaft Speed Sideband Markers.
With this evidence in hand we reported that the bearing had a
damaged inner ring and the overall amplitudes indicated a need
for immediate action.
Figure 13 is a photograph of the bearing. A piece of paper has
been inserted into the crack to make it more visible. Proof that the
bearing had been turning on the shaft is seen on the inside of the
ring, it is scratched, has black and blue heat marks and is coated
with fretting corrosion. Of course this is one of those “which came
first” problems, the crack or the looseness. Once the ring cracks it
will certainly turn on the shaft and if it were not scratched and
blued before, it soon will be. In any event, the bearing was
damaged when we first inspected it and it needed replacement.
Figure 12 is a spectrum taken on the same bearing at the same
location and at the same time as those above. The only difference,
besides the upper frequency limit, is that the acceleration signal is
processed to read out in velocity. Compare Figure 10 with Figure
12. The cursor is placed on the bearing frequency and the
amplitude reads 0.0004 IPS. No one would ever consider
changing a bearing with this low an amplitude, however we have

37. 36
enveloped acceleration readings that show a problem - and the
visual proof with the photo of the inner ring after it was removed.
This should convince anyone that enveloped acceleration is a
much more sensitive method of analyzing bearing conditions.
Figure 12. SKF 2222 Velocity Measurement, Cracked Ring.
Figure 13 is the photograph of the SKF 2222 with a cracked inner
ring. In black and white the fretting corrosion doesn’t show as well,
but the inner surface of the ring should be as clean and shiny as
the outer surfaces. A clean, straight-across crack is the
characteristic break resulting from excessive stress caused by an
expanded shaft.

38. 37
Figure 13. SKF 2222 with Cracked Inner Ring.
Very Slow Bearing Analysis
There are many bearings that rotate at less than one RPM, such
as slewing rings on cranes, radar antennas, and discharge vessels
in steel mills, all of which we have inspected over the years. The
following example is a bearing that was rotating at approximately
0.5 RPM. Operations will tell you that it is never convenient to
shut-down for maintenance, and in this case loss of production
plus replacement and labor would incur a cost of over $250,000.
Very slow bearings require the use of the time domain spectrums.
As can be seen in Figure 14, the FFT, where the BPFO is noted at
6.9 CPM, is not able to distinguish the unique, very low amplitude
frequencies and noise masks any other useful signals.

39. 38
Figure 14. FFT of Very Slow Bearing, Suspected Damage.
The solution then is to use the time domain spectrum.
Figure 15. Time Domain of Very Slow Bearing, Suspected Damage.
The BPFO of the bearing was calculated to be 6.9 CPM. The
harmonic marker is used to set the left and right cursors at 6.9
CPM. Then, by clicking on the bottom of the enclosed field, both
cursors can be dragged to a position where they appear to be in
the middle of the two pulses and will also line up with the next
pulse. This portion of the analysis is a judgment call on what lines

40. 39
up and what doesn’t. A number of spectrums had to be collected
to acquire this spectrum because it was found later that the
bearing housing was being distorted, so the rolling elements were
not always in contact with the ring.
A twin machine was available for comparison, so matching data
was collected there also.
Figure 16. FFT from 2nd Machine.
Figure 17. Time Domain from 2nd Machine.

41. 40
At the presentation for the manager, the two time spectrums were
overlaid as seen in Figure 18. With one shown in red and the other
in blue. Based on the overall amplitudes of the two spectrums, it
was deduced that there was nine times more energy being
generated inside the suspect bearing as in its twin. The decision
was made to change the bearing and severe damage was found in
the outer ring. It was also found that the mechanism designed to
allow for heat expansion had failed and was the source of the
housing distortion.
Figure 18. Composite Time Domain of Two Bearings.

42. 41
Figure 19. Composite FFT of Two Bearings.
Odds and Ends
Our experience has been that many times the owner of the
machine has no idea what bearings are installed. Usually the
machine has been in service many years with several overhauls by
several people and no one knows what bearings were installed. In
such cases, a helpful characteristic of the bearing fault frequency
calculations is that when the contact angle is greater than “0,” the
multiplier will result in a frequency that is a non-integer multiple
of the shaft speed. In Figure 20, the cursor is placed on an
unknown frequency spike and the Order information in the Single
Value box tells us it is 6.046X. We then place the harmonic marker
on this mystery frequency and see that we have harmonics. Based
on this information it would be prudent to do a physical inspection
of the bearing. The author has detected a number of damaged
bearings using this technique.
In Figure 20, a bearing had been deliberately damaged by the
owner before our arrival. He wanted to see if we could find it
among several others in the machine. We did, although the
damage was only several small scratches in the outer ring.

43. 42
Figure 20. Unknown Bearing @ 6.04X.
Remember that the computer bearing fault frequencies are
calculated based on new bearing dimensions. The bearings you
are inspecting are possibly worn and consequently the actual
frequencies generated may not fall exactly on the observed
frequency.
We have found through experience that approximately 90 percent
of inner ring failures are caused by poor installation techniques.
When the bearing is placed on the shaft by pushing on anything
but the inner ring, damage will occur. Force on the cage will
damage the cage and push the rolling elements against the lip of
the races causing damage to the rings. Even if the damage does
not influence the machine operation, it will result in noisy bearings.
Bearing ball and rollers are harder than the rings, 61-62 Rockwell
vs. 58-59 Rockwell. Therefore, damaged rollers and balls are rare
during normal conditions.
Care should be taken to prevent water from mixing with the
lubrication. One percent water in the lube system will reduce the
life of the bearing by 90 percent. Water dissolved in the lubricant
should not exceed 200 PPM

44. 43
And finally, over 60-70 percent of machine failures are caused by
the loss of the rolling element bearings. Why? Because of
misalignment! Other than thrust bearings, rolling element
bearings are designed to carry a radial load. When misalignment
occurs, an axial component is generated. When this becomes
excessive, the bearings begin to fail. Probably the one procedure
that would save the most money in any maintenance department
would be to improve the alignment methods using laser alignment
equipment.

46. 45
Balance Conditions
The general public usually equates any vibration as an out of
balance condition, no matter what the cause. Those in the
business of vibration analysis know, of course, that this is false as
there are many causes for vibration.
Technically, when the center of mass is not co-located with the
center of rotation, the rotating mass will present an out of balance
condition. On an FFT, the spectrum will be seen to have a spike of
energy at exactly the rotation speed of the shaft (i.e., at 1759 CPM
on a shaft rotating at 1759 RPM). When the vibration sensor is
placed on the shaft bearing housing, as the unbalance mass
passes the sensor it detects the centrifugal force that is generated.
Since this occurs once per revolution, it calculates how often this
will occur and presents the information in cycles per minute
(CPM). If desired, results can be shown in cycles per second (Hz).
The solution to an out of balance condition is to determine how
much mass, and the angular location of the mass that is not in the
center of rotation. There are numerous software and hardware
programs available to perform this function in the field or in the
maintenance shop. On shafts with multiple planes, as in a multiple
stage turbine, it may be necessary to perform a two plane balance
where the correction mass is placed on each of the stages so that
it influences both stages to arrive at a suitable balanced condition.
In the case of multiple planes attached to the same shaft, there
are two types of unbalance, coupled and un-coupled. In a coupled
condition, the unbalance forces are located in the same angular
location on both planes, for example both located at 134 degrees -
and if phase data is collected on both ends of the shaft, they will be
in phase. In an un-coupled condition, the balance forces are not in
the same angular location, one may be at 134 degrees and at 300
degrees on the other plane, and shaft end phase measurements
will be out of phase. What you will find in the field is usually a
combination of the two situations requiring the use balancing
hardware and software to arrive at a balanced condition.

47. 46
It is not the purpose of this publication to teach how to correct an
imbalance condition, but to recognize the condition when it is seen
on the FFT.
Field Data
Figure 1. Out of Balance Cooling Fan.
As seen by the amplitudes, this fan was in bad shape. It was
drawing air through a bag house and one bag had torn loose and
hung up on one of the fan blades. It was vibrating to the extent
that the fan blades were rubbing on the housing. This same FFT
will be seen in the section on Rubs to explain how this was
determined.
After the fan was cleaned and repaired, the following spectrum
was taken. Needless to say things were a lot quieter and
smoother.

48. 47
Figure 2. Cooling Fan after Balancing.
One question often asked is “What are those other spikes?” And
the answer is “I don’t know.” One is a harmonic of shaft speed and
the others are from random energy that could be coming from
other machines. You fix what you know is needed and if it isn’t
readily apparent where the other spikes are coming from, and they
don’t appear to be causing a problem, don’t worry about them.
In using an FFT for analysis, it is common to use “Auto Range” for
the “Y” axis. When this is the case, it is imperative that you pay
close attention to the amplitude scale when you compare FFTs.
The visual presentations may look the same, but Figure 1 is more
than 12 times the full scale amplitude of Figure 2. It is easy to fall
into this trap.

50. 49
Misalignment
Common Factors
Although the general public thinks most machinery vibration
problems are caused by something being out of balance, the
general consensus of the vibration community is that the root
cause of 70 to 75% of vibration problems is misalignment; and
misalignment is the root cause of 40-50% of bearing failures.
Machine failure because of misalignment usually proceeds in the
following manner:
Every rotating element has some amount of imbalance. All
manufacturers have a balance tolerance that they allow for during
assembly. To obtain a near perfect balance would be an extremely
expensive process that is used only on such applications as the
super quiet equipment for nuclear submarines. This acceptable,
small amount of inherent imbalance is dampened (absorbed) by
the rolling element bearings. Bearing clearances between the
rolling elements and races are approximately 0.0005," which is
filled with lubricant. For all practical purposes, there is no
clearance. As the unit is operated in a misaligned condition, the
push-pull of the rotating elements, through the coupling, causes
excessive un-designed loads on the bearings. These loads lead to
excessive clearances between the rolling elements and the races,
and the bearing is no longer providing the dampening needed to
restrain the inherent imbalance in the rotating element. The final
step occurs when some well meaning person notices the vibration
and writes the work order to repair or replace the unit, stating that
it is out of balance. So, without a correct diagnosis, unnecessary
work is performed. As a sidebar, there are statistics that show that
approximately 12% of any repair work performed will have to be
repeated for some reason, adding more to maintenance costs.

51. 50
Characteristics of an FFT Indicating Misalignment
Because misalignment induces a “push-pull” force for each shaft
rotation, two actions per rotation, there will be a display of energy
at twice running speed, (2X). As these forces are parallel to the
shaft, the best location to collect vibration data when checking for
misalignment will be on the ends of either the driver or driven unit
(i.e., in an axial direction). Since there are rarely perfectly aligned
units, it is not unusual to see some indications at 2X. However,
concern should be aroused if the 2X amplitude is more than 50%
of the 1X amplitude when the 1X is greater than 0.15 inches per
second. There is also a 1X factor, although it is masked by any
imbalance that is present and can only be documented by noting
the reduction in the 1X signal after the unit is aligned properly.
The phase relationship across the coupling will show that the
motor and driven unit are 180 degrees different, plus or minus
approximately 30 degrees. This difference is measured between
the inboard bearing of the motor and the inboard bearing of the
driven unit. This difference is caused by the push-pull action that
for one-half of the rotation is pulling the two pieces of equipment
together and then is pushing them apart on the second half of the
rotation, so that they are always moving in opposite directions, i.e.,
out of phase.
There are two types of misalignment, parallel and angular. In
parallel misalignment, the two shafts are parallel with each other
but not in the same horizontal plane. In angular misalignment, the
two shafts may be in the same plane but are at an angle to each
other. Although angular misalignment is the more common
problem, there is often a combination of both problems when the
machine base is not properly prepared. Parallel misalignment will
be more prominent in the radial measurement and angular in the
axial direction. Both will have the 180 degree out of phase
condition across the coupling.

52. 51
Field Examples
Figure 1. 1191 RPM Pump/Motor before Alignment.
Figure 2. 1191 RPM/Motor after Alignment.
This water injection pump was reported to be vibrating, the FFT is
displayed in Figure 1. Maintenance was requested to perform an
alignment. The face to face misalignment was found to be 0.010”.
Although the misalignment was corrected, the 1X amplitude is still
too high (Figure 2), indicating an out of balance condition that
needs to be corrected.
Another aspect of misalignment that is often overlooked is the
additional energy required to operate equipment in a misaligned
condition. Field testing has determined that the additional power
required to operate a misaligned pump/motor set is 2.5-3.0%

53. 52
higher than a properly set up unit. Using data from a large
manufacturing plant in the South, the savings for just one machine
was calculated in the following manner.
Three phase KW = (volts)(amperes)(P.F.)(1.732)
1000
(KW difference)($/KW)(7200 hrs/yr) = $ Annual Savings
7200 hrs/yr is assuming 6 day week - 50 wks/yr
EXAMPLE: Initial amperage = 27 after alignment = 25
Initial KW = (575)(27)(.8)(1.732)
1000 = 21.5 KW
After Align = (575)(25)(.8)(1.732)
1000 = 19.9 KW
21.5-19.9=1.6 KW Difference
(1.6KW)($.06/KWH)(7200 Hrs/yr) = $691.20/yr
As noted before, the major result of misalignment is damaged
bearings. Many coupling manufacturers advertise that their
couplings can absorb the stress of misalignment. This may be true,
but the bearings cannot. The energy used by misalignment is
converted to heat and vibration, energy that will eventually
damage the bearings no matter what coupling is installed.

54. 53
Gear Boxes
Common Factors
Gear boxes are used to transfer energy from one shaft to another.
Because there are always acceleration forces present and the
clearances between gears are very small during this transfer of
energy, an accelerometer is the preferred sensor to use in taking
measurements. For slow speed units, the signal may be processed
using velocity, but generally the best signal processing is
enveloped acceleration. At the higher frequencies that are often
encountered, other types of sensors are not suitable.
Every gear box is different. Even units that are made by the same
manufacturer and have the same data plates will produce different
vibration signals. Every manufacturer has tolerances for the gears
and other internal parts. Often, mating parts are made by different
suppliers. When all of these tolerances come together, the overall
result is within the OEM’s specifications, but each gear box will
produce a different FFT and time waveform.
The vibration signals generated inside the gear boxes are carried
through to the bearings. Therefore, mounting the sensors on the
bearing housing is the favored location. When the gears come into
mesh, a signal that is equal to the shaft speed times the number of
teeth on the gear mounted on that shaft will be generated. If the
gear box is ancient, it may be that the only way to know how many
teeth are on the gear is to physically count them, either in the field
or on the spare in the storeroom.
On gear boxes with multiple gear sets, the analyst must calculate
the individual gear mesh frequencies (GMF) in order to sort out the
signals seen on the FFT. In all cases, the gear mesh frequency
remains the number of teeth times the shaft speed. Sun gears
required additional calculations, which may be found on
engineering internet sites.

55. 54
In all cases, the time domain should be saved with the FFT. Use of
the time domain will be explained as we look at field data.
Field Examples
A national gold mining company asked for an analysis for a gear
box driving an ore crusher. This crusher was rotated for 6 months
in a clockwise direction and then reversed to operate 6 months in
a counter-clockwise direction to provide even wear on the gear
teeth. After the last reversal, it was noted that a low level buzz was
heard and the hand rails around the gear box were vibrating
enough that it could be felt when touched. A similar vibration was
felt on the gear box but not on the motor.
A 100 mv/eu accelerometer was mounted on the gear box inboard
bearing cap using a magnetic base. The following FFT and Time
spectrum were collected with 400 lines resolution and four
averages. A search of the maintenance files did not produce any
information about the number of gear teeth, and the shaft speeds
were determined with a strobe light and manually entered into
SKF’s software using “Set Speed.”
Figure 1. Slow Speed Gear Box, Gear Mesh @ 21,450 CPM.
By using the sideband marker and centering it on the gear mesh
signal, the sidebands can be seen to have been generated by the
modulation of the gear mesh and the output shaft which was

56. 55
measured to be rotating at 225 RPM. The drive motor RPM was
595. The driving gear is often referred to as the “bull gear” and
the driven gear is the “pinion gear”.
Figure 2. FFT with Sideband Markers Centered on GMF.
This tells us that the gear with the apparent damage is on the
output or pinion gear. Further confirmation of this analysis can be
seen in the time domain spectrum.
Figure 3. Time Waveform Slow Speed Gear Box.
Do not be concerned with slight differences in the calculated
speeds, a higher resolution measurement would bring them
closer, but higher resolution requires more time to collect the
data.

57. 56
The time spectrum displays the energy when a damaged tooth or
teeth come into mesh. The units are in “G’s.” The time span is set
by the Fmax of the FFT and is inversely proportional. In other
words, as the Fmax is increased, the time span is decreased.
Using the “Harmonic Marker” from the tool bar, two cursors are
generated. Using the mouse, each cursor is dragged until it
overlays a signal. Then the other cursor is aligned with the next
matching signal from the pattern. Once the two cursors are
aligned, both of them can be moved together by clicking and
holding the mouse on the line at the bottom of the pair.
Now we have two indications that the output shaft is the source of
the vibration. The maintenance department was informed that
there was a problem with the 36 tooth gear. A search of the store
room revealed that they did not have a spare. A rush order was
placed and a spare flown to the plant in 3 days.
Tear down of the gear box revealed a broken tooth on the output
shaft gear with some damage on the following teeth. If the
gearbox had been taken out of service and opened without the
knowledge of the source of the problem, three days of production
would have been lost without the output from the crusher.
The value of vibration analysis? At the time of this incident the
value of three days gold production at this mill was $1,500,000.
High speed gear boxes are analyzed using the same techniques as
with a slow speed box. The next example was collected at a large
chemical plant. Operators had reported a change in the sounds
emitting from the box and the platform where the steam turbine
driven gear box was mounted had begun to vibrate to a point
where it could be felt on the bottom of the feet. The bull gear had
recently been replaced with a new gear.

58. 57
Figure 4. FFT of Gear Box with 7500 CPM Sidebands.
Figure 5. Time Spectrum of Damaged 7559 CPM Drive Shaft.
Based on this FFT, with input shaft speed sidebands, the time
spectrum and the physical vibration that was present, it was
recommended that the unit be shut down and inspected.
When the unit was opened, 53 teeth on the bull gear were
damaged and broken. Because the gear had recently been
changed, the manufacturer was asked to examine the damaged
gear. It was found that the teeth had been cut improperly and

59. 58
were not meshing as designed. The old gear was returned to
service while another gear was produced.
If this unit had continued to operate, there was a high potential for
a catastrophic accident. The amount of kinetic energy in a rotating
mass weighing hundreds of pounds and rotating at 7500 RPM will
inflict tremendous damage to the surrounding area if it fails.
Occasionally, an analysis will be requested on equipment that is
old and has no maintenance records. Management has decided to
rebuild one of two gear boxes and asks which one should be
rebuilt. The solution is to collect data on the two units, and based
on the energy levels, select the unit with the highest levels.
Figure 6. FFT of Number One Gear Box.

60. 59
Figure 7. Time Spectrum of Number One Gear Box.
Then data was collected on the Number Two Gear Box.
Figure 8. FFT of Number Two Gear Box.

61. 60
Figure 9. Time Spectrum of Number Two Gear Box.
Based on the energy content of the two gear boxes and the time
spectrums, it was recommended that Number One gear box be
rebuilt with attention to the gear on the 1200 RPM output shaft. It
was clear that the Number Two gear box has some issues but the
amplitudes measured determined the decision.

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Electric Motor Analysis
The largest group of installed drivers in most plants is electric
motors, both three phase Alternating Current (AC) and Direct
Current (DC). In many plants, other than a vibration check on the
condition of the bearings, electric motors are installed and
forgotten except for an occasional shot of lubrication. The results
of this neglect are electrical failures and unplanned outages. With
the portable electronic data collectors on the market today, it is
possible to monitor both the physical and electrical condition of the
motor while it is installed and in operation.
Although the analysis technique is the same for both, it is
important to know that there are two types of AC motors,
continuous duty and intermittent duty. A continuous duty motor is
one that once started is allowed to operate for hours or days
before shutting down, such as a conveyor belt in a mining
operation. An intermittent duty motor is one that only operates
when required, such as on an air compressor that only operates
when the tank pressure falls below a set point. The construction of
the motors is different. The intermittent duty motor is made with
stronger rotor bars to withstand the torque applications that occur
each time the motor is started. The heavier construction is not
required of the continuous duty motors because they are only
torqued one time when first started. If a continuous duty motor is
installed in an intermittent application, there will soon be a failure
because of broken rotor bars.
Occasionally, it is difficult to determine whether the vibration is
mechanical, as in an unbalanced rotor or misalignment, or
electrically induced because of unbalanced power. The solution is
to set up in the analyzer mode the vibration measurement with
100 lines of resolution, one average and an Fmax above the 2X
signal. The power is then turned off. If the 1X and 2X signal slowly
decreases in frequency and amplitude as the motor slows down,
the fault is mechanical. However if the 1X or 2X signal instantly

63. 62
drops off the screen when power is removed, then the cause is an
electrical fault that ceased when the power was removed.
Historically, more motors are damaged by too much lubrication
rather than a lack of lubrication. The bearings only require a small
amount of grease but the general consensus is “ put in a little
more.” The bearing can only contain a limited amount of grease,
so the excess is pushed out into the motor housing where it will
often cause an electrical short and a premature failure of the
motor.
Analysis, AC Motors
Often the first indication of an electrical problem is the appearance
of a signal in the motor vibration FFT at exactly twice line
frequency. This signal can indicate several possible problems:
 Soft foot
 Poor construction, end bells not aligned
 Misalignment
 Insulation breakdown. Check by feeling the motor housing for
a hot spot.
The first three items all result in the rotor not being in the center
of the stator / magnetic field. The last item is a mechanical
problem that will get worse with time. In the evaluation of the FFT,
it may be necessary to zoom in on the signal for two pole motors
to distinguish between 2X, approximately 7100 CPM @ 60 Hz
power or an electrical problem at exactly twice line frequency,
7200 CPM @ 60 Hz. Note that a motor that is listed as operating
at 3600 RPM will in service and under load operate about 3550
RPM.
It is possible to generate a mechanical vibration if the voltage and
current is not balanced between the three leads. The voltage
should not vary between the three leads by more than 5%, and the
current should not vary by more than 3%. An industrial plant in the

64. 63
Chicago area was slowly surrounded by housing developments
over a period of 10 years. In examining the motor vibration trends
toward the end of that time period, it was noticed that the
vibration levels would peak in the morning and in the evening. An
investigation found that the voltages were unbalanced during
those periods of time. Further investigation by the power company
found that the housing developments were primarily being fed by
just one of the three leads from the main transformer. The result
was that in the morning, when the homeowners were starting
their day, and in the evening, when everyone was home, the
voltage in that lead would drop more than 5% with reference to
the other two. In theory, this could be fixed, but in practice it was
too expensive and the plant had to live with the condition.
AC motor current analysis uses an unpowered clamp-on ammeter
to evaluate the current flow in the power circuits of a three phase
motor. The data collector is set to collect volts RMS, in log scale.
The motor must be loaded to at least 75% of its rated full load. The
sine wave of the AC current will induce a voltage in the ammeter
and, in effect, turn the motor into a sensor, where the detected
motor current is the information carrier on the motor’s mechanical
condition. Mechanical faults in the motor will alter the motor
current and change the response seen in the FFT.
Never place a clamp on a high voltage line, the maximum safe
voltage is 240 volts. Most ammeters are rated for 600 volts but
this can be very dangerous under some conditions and is not
recommended in any case. Instead, place the clamp on the power
line coming from the step down transformer to the installed amp
meter, usually on the front panel of the motor control center. The
voltage is commonly stepped down 100:1, but the current signal
that we are interested in observing is not altered. If continuous
monitoring is desired, a ring magnet can be permanently installed
on the lead. If one of the three leads is used, it does not matter
which of the leads is used for the analysis. In the past, it was
required for the analyst to measure the responses seen in the FFT
and calculate the amount of damage based on that response.

65. 64
Today’s data collectors, with installed software, perform the
calculations and present to the analyst, in chart form, the damage
report. Figure 1 is a typical FFT taken of a damaged AC motor.
Figure 1. Damaged Motor.
Figure 2. Good Motor.
Using a logarithm scale because of the extremely low amplitudes,
the data collector evaluates the amplitudes of the sidebands with
reference to the 60 Hz line voltage. Note that the damaged motor
sideband has an amplitude of 9.084E-2 volts and the good motor
has an amplitude of 3.0793E-3. The results are presented to the
analyst as a measurement of increased resistance, which is usually
caused by broken rotor bars. The higher the amplitude of the
sideband the greater the resistance and the probability of more

66. 65
broken rotor bars or other damage. Because this is a measure of
resistance, the additional resistance may be somewhere else in the
circuit, such as a bad connection in the motor control center or a
loose splice at a repair point, but the most common fault is broken
rotor bars. As the number of broken bars increases, the motor will
begin to fail and start slower and slower until it overheats and
fails. Usually after four bars are broken the motor will completely
fail.
A type of motor application that may cause trouble are the high
efficiency motors that are purchased with the intent of reducing
electrical costs. These motors are more efficient because of
thinner insulation and smaller air gaps. What has happened in a
number of plants that also have variable frequency drives (VFD) is
that the electronics of the VFDs and other electronic equipment
will feed back high voltage spikes into the plant electrical grid.
These very high frequency spikes have been measured in the
thousands of volts. The thinner insulation of the high efficiency
motors will develop pin holes from the voltage spikes, and these
internal shorts lead to reduced motor life.
In the past few years, VFDs have been sold to a number of plants
as a tool that will allow them to set the speed of their equipment
to whatever speed they desire for the best production rates. A
problem is created when the speed desired is within a resonance
range of the motor or other components. A number of failure
cases have been documented where the speed was lowered from
the design speed, which resulted in the rotor operating in a
resonant condition.
A plant in Texas was having bearing failures on a monthly basis in
the main conveyor motor after installing a VFD. It was simple to
place the accelerometer on the motor and watch the amplitudes
rise and fall as the resonant speed range was passed through as
the conveyor speed was adjusted. In that case, the solution was to
post a notice not to operate at those specific resonant speeds.
After that, the bearing failures ceased. If the rotor is resonant, the

67. 66
increased vibration causes chaffing of the internal insulation with
resulting breakdown and an electrical failure due to shorting. A
VFD will not solve all speed related production problems, and they
must be evaluated with care as to their influence on the driven
equipment and the electrical grid.
Rotor Bar Pass Frequency
If there is damage in the stator of the motor, that area of the
generated magnetic field will be altered and reduced.
Consequently, the field 180 degrees on the other side of the stator
will have a higher amplitude. When the motor is in operation, as
each rotor bar passes through this area of higher magnetic force,
the rotor is pulled in that direction. In Figure 3 this mechanical
effect can be seen at 51X, the number of rotor bars in this AC
motor.
Figure 3. Rotor Bar Pass Frequency at 51X.
The sidebands around the rotor bar pass frequency are equal to
1X, although with this lower resolution of 400 lines the markers
do not all line up perfectly. It is recommended that one high Fmax
of this type, in G or gE, be collected on each motor.
The number of rotor bars in any motor is usually very difficult to
determine. Motor manufacturers are reluctant to divulge this
information. It is recommended that any time a motor is sent out

68. 67
for repair that a request for a rotor bar count be included in the
purchase order.
Analysis, DC Motors
Because there is no sine wave with DC current, a battery powered
ammeter is required to detect DC current. This type of ammeter is
referred to as a Hall Effect transducer. DC motors are used where
it is necessary to change the speed of the motor during operations.
They can be driven by a direct current, or the most common way
is AC current converted to a direct current by the use of silicon
controlled rectifiers (SCR). In simple terms, the SCRs take the AC
sine wave and chop it into thousands of pieces and then put it back
together as a straight line (i.e., a DC voltage). The SCRs can fail.
The response to this failure will be a strong signal at 21,600 CPM
where the line current is 60 Hz (6 X 3600 CPM = 21,600 CPM).
With DC motors, motor speed is proportional to the amount of
voltage applied to the motor. A danger here is that the motor can
be over sped by excessive voltage, a problem that the railroads
have with their DC drive locomotives.
The most common failure mode for DC motors is a break down in
the internal insulation. It is suspected that, as the speed is varied,
resonant speeds are often maintained, increasing the vibration
levels which causes the damage. Of course, the environment is
often a cause. It has been noted that many motors in the
basements of paper mills are often covered with trash and pulp
waste. This causes overheating insulation failure and motor
failure. Railroads have this environmental problem with trash and
dirt thrown up from the rail bed.
Experimental Testing, DC and AC Motors
It is possible to monitor the condition of the insulation in a DC
motor using a portable data collector. With a Hall Effect
transducer, a spectrum can be taken on the armature lead and
then processed with enveloping technology. Lab tests using this

69. 68
procedure on an unloaded DC motor found that the trend of the
overall signal amplitude was fairly linear as the damage was
mechanically increased. Then, when the damage was repaired, the
amplitude returned to the levels seen prior to the damage being
introduced. The next figures show the difference between an
undamaged DC motor and the same motor with a small portion of
insulation removed from the rotor windings.
Figure 4. Undamaged DC Motor.
Figure 5. DC Motor with Damaged Insulation.
This was an experimental test, but it can clearly be seen that a
small amount of damage results in a 1000 percent increase in
amplitudes.

70. 69
Another experiment using enveloped AC motor current has been
conducted. With the enhanced sensitivity available with enveloping,
very small defects in rotor bars can be disclosed.
A major supplier of quiet AC motors for the U.S. Navy submarine
fleet requested assistance in the detection of occlusions in cast
rotor bars. An occlusion, or bubble, cast into the metal has a
similar influence on the motor current field as a cracked or broken
rotor bar, except the amplitudes are not detectable in standard
motor current analysis FFT’s. To simulate an occlusion, a hole was
drilled 25% of the way through one rotor bar of a 250 HP AC
motor installed on a test stand. The data collector was in the
enveloped acceleration mode, 1st
filter, RMS, 400 lines and 1
average, Fmax = 12 Hz. The motor was loaded to 75% of full load.
Figure 6. Undamaged AC Motor, Enveloped Current.

71. 70
Figure 7. Rotor Bar Drilled 25% Through.
The analysis requires dividing the Pole Pass Frequency @ 59.4
CPM by the overall amplitude, 0.016 for the good motor in Figure
5. The resultant was 0.25 or 25%, which is the mid range for a
good motor.
For the damaged motor in Figure 6, the Pole Pass Frequency @
116.1 has an amplitude of 0.019 divided by the overall amplitude
of 0.053 equals 0.37 or 37%. The 12% increase is due to the
induced damage.
Additional tests drilled the bar 70% through, which resulted in a
reading of 59% - or a 34% increase over a good motor.
The results of the testing proved that an enveloped motor current
would provide the sensitivity to evaluate rotor bars that had an
occlusion or other minor or other damage in the rotor circuit.

72. 71
Resonance
The root of the word resonance comes from a musical word,
resound. If a musical reed that is tuned to a “C” note, for example,
is placed next to another “C” note reed and the first reed is made
to vibrate, it will cause the second reed to also “resound” and
vibrate. In the mechanical world, this same influence takes place
when an out of balance rotor rotating at 1800 RPM, which would
be vibrating at 1800 CPM, is mounted on a frame that has a
natural frequency of 1800 CPM, the frame will also begin to
vibrate. The motor and frame would then be said to be “in
resonance.” This example shows that resonance is not a vibration
source, but the response to another vibration.
Early tachometers used this principal in their construction. A rod
would have a series of reeds that were tuned to the RPM range of
interest. The rod would be placed on the machine and the reed
that was vibrating at the highest amplitude would be recorded as
the rotation speed of the machine. The smaller the difference in
RPM tuning between each reed, the more accurate the results.
The frequency of a structure is influenced by its stiffness and
mass. If these values are known, it is possible to calculate the
natural or resonance frequency of a machine or structure. Another
way is to perform a “bump” test and “ring” the structure by
striking it and measuring the response with a frequency spectrum.
A large bell is not rung by striking it with a straw. When a bump
test is performed, the mass that is used as a hammer must be
large enough to excite the structure, just as a heavy clapper rings
the bell. Structures will have a different response when struck in a
horizontal plane vs. a strike in the vertical plane because the
stiffness will be different in each plane.
When structures are in a resonance condition, a small vibration
input such as imbalance or misalignment is amplified and in time
may damage the structure. The natural frequency of the structure
should not be within 20% of the rotation speed of the rotating
element. So if 1000 RPM is the rotation speed, the structure

73. 72
should be designed to have no natural frequencies between 800
and 1200 CPM. There are only two ways to change the natural
frequency of a structure, increase the mass to lower the response
or stiffen the structure to raise it. Mass can be supplied by addition
of bags of lead shot, and added bracing will provide addition
stiffness, both have been used successfully.
Although structures are typically designed properly, the owner
often gets into trouble by operating the machine at a speed not
included in the original design. This has been the experience of
many paper making machines, which over the years have been
slowly increased in speed to increase production. It is not unusual
to find placards on operating panels with strict instructions on
speed ranges that must not be engaged during operation when
equipment is operated with variable speed drives. Also, it is
possible with variable speed drives to operate at a speed which
excites an internal component, such as bearings. If a variable
speed drive is involved and the bearing life is abnormally short,
there are cases where the bearings have been damaged by
resonance. Since the bearing’s mass and stiffness cannot be
changed, it is again necessary to block out a specific operating
speed range.
The terms resonance and natural frequency mean the same thing,
it is the frequency determined by calculations or by performing a
bump test. A critical frequency is the frequency that, if maintained,
will result in damage if the generation of that frequency is
continued.
Field Experience
Modern data collectors now have built into the software the
parameters to collect and process the frequency spectrum to
display the response frequency when a bump test is performed.
Every machine is different, and it may be necessary to perform the
test several times as the parameters are modified to improve the
response. The time spectrum provides the most important

74. 73
information, so both the frequency and time spectrums should be
displayed together. Note in Figure 1 that until approximately 0.15
sec there is no input other than background noise. At 0.16
seconds, an impact occurs and the structure rings with the signal
decaying down to the background noise at 0.39 seconds. It is
mandatory that the signal returns to the background level in order
to obtain an accurate frequency response, because the energy
shown in the time spectrum is the energy used to develop the FFT.
If the signal is still ringing as it reaches the max time of the
spectrum, energy will be lost and the frequency will be in error. An
error will also occur if two or more bumps are captured in the time
spectrum.
Figure 1. Ring Response in Time Spectrum.

75. 74
Figure 2. Frequency Response from Bump Test.
Although this is an actual response, this is not the type of response
one would see in a normal structure. This test was performed on a
large bell, which after several attempts provided the proper
response for a bell that was designed to an “E” tuning, 41,200
CPM. The small harmonics that are seen are what give bells their
pleasing overtones.
A cold roll steel mill was experiencing premature failures in the
rolling stand, and an examination was requested to determine if a
natural frequency was the culprit. As previously stated, the
hammer used to excite the structure during a bump test has to
have enough mass to ring the structure. In this case, the cold roll
mill stand, which weights multiple tons, was bumped by the heavy
coil of steel as it rolled into the mill stand. Numerous attempts
were required to capture the signal, as it rapidly scrolled across
the screen.

76. 75
Figure 3. Time Response from Mill Stand.
Figure 4. Frequency Response of Mill Stand.
The frequency response shows the usual multiple responses to a
bump test of a large structure. The two responses at 976 CPM
and 9450 CPM are prominent, but do not coincide with any
internal rotating speeds, nor do any of the other spikes. With this
information, the owners were informed that it did not appear that

77. 76
a resonance condition was the cause of the premature bearing
failures. Later, testing of the bearing lubrication indicated high
levels of water that was used in the rolling process. It has been
shown that 1% water in the lubrication will reduce the life of a
bearing by 90% and that the maximum water content should not
exceed 200 PPM.
The FFT and time spectrum will give visual clues that a resonance
condition exists. Often, it is of no consequence, but if the frequency
response overlays an operating speed, trouble may result. The
characteristic of an FFT showing a resonance condition is seen in
Figure 5, where the base of the energy spikes are very broad.
Depending on the Fmax and the resolution, this broad base may
be made up of many small energy spikes. The term for such a
display is that it is a “haystack,” and they are often in harmonic
order.
Figure 5. Harmonic Haystacks, Possible Resonance.
In many cases the time spectrum gives a better clue. A series of
pulses can be seen across the full span of the spectrum, pulses
result in the vibration amplitudes rise and fall. The term for this
type of display is that you see “angelfish” in the spectrum as this is
the general shape of that tropical fish.

78. 77
Figure 6. Resonant Angelfish Seen in the Time Spectrum.
A question that is often asked when a resonant condition is
detected is, “how long will it run like this”? There is not an answer
for this question. It is best answered with another question, “How
many times can you bend a paper clip before it breaks”? As the
structure flexes, the material will become fatigued and at some
point will crack or break apart. When it will happen is anyone’s
guess. There have been several incidents involving wind turbines
where the blades have broken off because of a resonance
condition excited by the vibration generated when the blades pass
by the support structure. No matter what material the blades are
constructed of, sooner or later, when excited by a resonance, the
bending moments and amplitude will exceed the design limits and
the blade will break.
Another way to determine a critical frequency with a variable
speed setup is to slowly increase the speed through the critical
frequency and then continue to increase the speed until the signal
amplitude has returned to background levels.
The manufacturer of large steel shafts asked if we could verify the
critical speed that had been calculated for a new type shaft. The
shaft was in a variable speed lath where it had been turned. The
data collector was placed in “Peak Hold” and velocity, 400 lines;
the sensor was placed on the head stock of the lath. As the speed
was slowly increased, two low frequency low amplitude signals

79. 78
were noted, and as the speed was further increased, the prime
critical frequency was reached at 1020 CPM. Since we were using
Peak Hold, the signal remained on the screen to be saved for later
analysis. See Figure 7.
Figure 7. Shaft Critical Speed at 1020 CPM.
The company was very pleased since their calculated critical speed
was 1000 CPM. This spectrum was taken on the speed run up.
The frequency would have been different if it was collected as the
shaft coasted down in that the lack of torque that was present on
run up would have reduced the stiffness.
It is common to see in pictures after an earthquake that some
buildings have remained standing while those on either side will
have collapsed. Contractors are sometimes blamed for shoddy
work when all three buildings may have been built by the same
company and met all building inspection requirements. The
difference in the buildings is not their construction but in the way
the materials are used in the construction. Different architects call
for different construction methods and materials in their buildings
and they will have a different stiffness and a different natural
frequency. Also, the height of the building has a direct relationship
to the natural frequency of the building. The seismic wave
generated by an earthquake will have a very low frequency, in the
0 to 5 Hz range. If that is also the natural frequency range of the
building and the seismic event is long enough or hard enough, that

80. 79
building will come down while the building with the higher natural
frequency will sway and shake, but remain standing.
Historical Note
The Tacoma Narrows Bridge near Seattle, WA is often mentioned
as an example of how resonance can destroy a structure. In the
case of the bridge, it is often stated that the wind blowing through
the support cables caused the bridge decking to vibrate and
eventually break and fall into the water. Actually, the failure mode
was from aero elastic flutter caused by the wind blowing on the
bridge deck that lacked the stiffness necessary to remain stable.
When the wind blew from the right direction and velocity (40
MPH) the 2800 ft span twisted and buckled like a flag in the wind
until it was destroyed.

82. 81
Enveloped Acceleration (gE)
The initial development of enveloped acceleration, or
demodulation, as a signal processing technique occurred many
years ago as a laboratory analysis tool. It remained there until the
early 1990’s when advancements in electronic signal processing
enabled the development of a truly portable analyzer for field
work. Although the original intent was to provide a tool for bearing
analysis, further development has resulted in a multiuse tool for
gears, bearings, detection of looseness, motor faults and in
general the analysis of any system that produces a once per
revolution impact or electrical signal. For example, in bearings as
each rolling element makes contact with a fault, as gears come
into mesh, or when a loose component cycles between its two
extremes, an impulsive signal is produced. Imbalance signals and
other low level non-impact signals do not respond to the
enveloping process which requires the use of a velocity FFT to
determine a precise 1X signal that can be transferred to the
enveloped FFT. The use of enveloped acceleration will allow you to
at least double the warning time before a bearing enters the end
of its useful life.
Acceleration envelope signals are band passed signals that remove
the lower and higher frequencies around the frequency of interest.
The equipment is designed so that the user can choose the
frequency range of interest, for example 500 to 10K Hz. This
would remove all signals below 500 Hz and above 10K Hz. In the
SKF data collectors, there are four filter ranges available for
selection. Experience has found that the third filter, 500 to 10K Hz
is the most useful for bearing analysis in speed ranges from below
1 RPM to over 5,000 RPM.
The enveloping process collects and converts the higher frequency
harmonics of the fault frequency and demodulates them for
presentation in the FFT. There is NO relationship between the
frequency range of the FFT and the selected frequency range of
the filter band. Because there is no integration of the

83. 82
accelerometer signal, there is not a “ski slope” problem in the FFT,
as with velocity when internal electronic noise is integrated.
Therefore, the minimum frequency for the FFT may be set to “0”
for slow speed equipment. The maximum frequency of the FFT is
initially set to 10X unless the bearing frequencies are known, in
which case the Fmax may be set between 4 and 5 times the
bearing inner race frequency, (BPFI). Although there will be many
harmonics generated by the fault, the presence of the
fundamental fault frequency plus three harmonics are sufficient to
analyze the bearing’s condition. The BPFI frequency is used as an
Fmax because that range will also include any outer ring faults.
The overall process may be stated thus. By filtering and removing
low level signals, the harmonic content of a specific filter range are
collected. These repetitive signals are then mathematically
manipulated to further enhance the repetitive signals and degrade
the random noise and low level machine signals. The signals are
then demodulated from the higher ranges and presented in an
FFT setup to display the fundamental fault frequency plus
harmonics.
The presence of harmonics is the key to successfully using
enveloped acceleration. Even if the fundamental fault frequency is
present but there are no harmonics, then there is no damage in
the bearing. If the fundamental fault frequency is seen and there
are harmonics of that frequency, then there is damage in the
bearing. For gear analysis, there will always be harmonics because
of the rising and falling forces as the gears go in and out of mesh.
Looseness, by definition, generates an FFT with multiple
harmonics of 1X.
Bearing analysis using enveloped acceleration requires knowing
the rotation speed of the bearing shaft. As noted in the section on
bearing analysis, the determination of the severity of damage is
based on the presence of harmonics and the amplitude of the
fundamental frequency. Repeating that information for machines
operating between 1500 and 3600 RPM:

84. 83
“Because the amplitude is dependent on the bearing shaft rotation
speed, a one size fits all damage vs. amplitude chart is not
possible. Using field experience, the following guidelines were
developed for bearings operating between 1500 and 3600 RPM.
Amplitude Action taken
0.0 to 0.2 gE No action required
0.2 to 0.5 gE Notify management, increase surveillance
0.5 to 0.8 gE Request bearing change when possible
0.8 to 1.0 gE Request bearing change ASAP
1.0 gE and up Request machine be taken
out of service
These are only guidelines, experience and machine criticality will
have to be included in the decisions
The dependence on the speed of the bearing shaft is seen in the
following chart:
Figure 1. Relationship Between Speed and Amplitude using gE with an Induced
Fault in the Outer Ring.
Always remember that enveloped acceleration is a unique,
mathematically derived analysis tool. It cannot be compared with
any other method of analysis nor is there any “conversion factor”

85. 84
to match it to any of the many alert and warning charts that have
been published.

86. 85
Spectral Emitted Energy (SEE) Technology
SEE Technology is a machine condition monitoring technology
developed by SKF. The two unique components of the technology
are the use of a high frequency acoustical accelerometer to collect
the data, and the enveloped processing of the signal from the high
frequency range to the low frequency range that is observable in
an FFT. It is useful in evaluating the condition of rolling element
bearings, lubrication problems and gear meshing faults.
How it Works
A spall in a bearing ring will undergo a metallurgical change of the
metal crystals before the spall breaks through the surface of the
ring. When a rolling element of a bearing rolls over this area of the
ring, the metal crystals are moved against one another by the
loads in the bearing. This movement of crystals grating against
each other generates high frequency acoustical signals in the
range of 250,000 to 350,000 Hz. As each roller passes over the
area, the signals are generated in a repeating pattern with
intervals that can be measured, just as with enveloped
acceleration measurements. This rate of acoustical pulses is then
enveloped and demodulated to be presented in an FFT as a
developing bearing fault. With this early warning, the bearing
monitoring can continue until the spalled area breaks loose. This
will be seen as an increase in signal amplitude which will continue
to increase as the damaged area grows. Using this high frequency
process removes the normal machine noises and signals induced
by looseness, imbalance and other random noises.
If there is a lack of lubrication or a contaminated lubricant
situation, the rolling elements will make metal to metal contact
and again the acoustical signals are generated. However with the
lack of lubrication or contamination, there are no individual pulsed
signals. The sound is a continuous signal that, when displayed in
the FFT, is seen as a raised noise floor which over time will
continue to rise in amplitude. Once a “normal” plot is determined

87. 86
by several observations, an increase can be noticed by overlaying a
new plot with an older plot. Confirmation of lubrication problems
can be confirmed by a reduction in amplitudes after the application
of fresh lubrication.
One characteristic of high frequency acoustical signals is that they
attenuate (fade out) very rapidly in the air. To overcome this, the
acoustical accelerometer must be attached to the bearing housing
not only with a flat rare earth magnet but a couplant must be used
between the housing and the accelerometer. The best couplant is
a light grease or heavy oil to fill any air gaps. It is also
recommended that the pickup contact point on the bearing
housing be milled so that it is flat and has no paint on the surface.
Even if it is not possible to mill the point, the paint must be
removed.
Based on case histories and field experience, the following
guidelines have been developed to rate the damage severity:
Amount Action
0 to 3 SEE No action required
3 to 20 SEE Possible lubrication contamination /
small bearing defect
20 to 100 SEE Bearing defect or contamination
100 SEE and above Severe bearing problem
Field Examples
This example was a bearing check on a variable frequency drive
(VFD) motor that was operating between 700 and 715 RPM. SEE
readings show the harmonic activity just as the gE readings,
although the amplitudes indicate there are no problems with the
outboard bearing.

88. 87
Figure 1. SEE Spectrum, Outboard Motor Bearing, 30K CPM.
Figure 2. SEE Spectrum, Inboard Motor Bearing, 30K CPM.
Although the SEE data indicates very minor looseness as
evidenced by the multiple harmonics of shaft rotation speed, the

89. 88
amplitudes did not indicate a bearing problem. Additional data was
taken using gE.
Figure 3. gE Spectrum, Inboard Motor Bearing, 15K CPM.
To compare the readings, another spectrum was collected in
velocity, (IPS)
Figure 4. Velocity Spectrum, Inboard Motor Bearing, 7K CPM.

90. 89
The conclusion reached on this case based on the SEE reading
was there was not a problem with the bearing. However based on
the multiple harmonics in the enveloped acceleration and velocity
readings, there was excessive looseness in some portion of the
machine and further investigation was warranted.
This case is a good example why multiple methods of investigation
are very important. While one method may be sufficient in many
cases, here is a case where one technology cleared the bearings
and another found the apparent source of the vibration.
Another area for use of SEE technology is in detecting rubbing
seals in steam turbines and other types of equipment where it is
necessary to contain the driving steam or products. A seal that
rubs because it is not evenly loaded will generate a mechanical
vibration that can be very destructive. Once a “normal” acoustic
signal is determined for the seal, any changes due to wear or
housing distortion or shaft unbalance can be determined by
comparison of old and new spectrums.

92. 91
Mechanical Looseness
Even the most carefully assembled machines can develop
looseness over time. Nuts work themselves loose, pillow block
bolts break, shafts become worn and undersized, and occasionally
a wrong sized component is installed, allowing excessive
clearances.
Common Factors of Looseness
The predominant FFT feature of equipment that has loose
components is a display of multiple harmonics of running speed.
The rotating speed of the equipment is referred to as 1X, and in
this example is 1200 RPM. The second harmonic (2X) would then
be at 2400 CPM, the third (3X) at 3600 CPM, etc. The spacing
between each harmonic is equal to the shaft rotation speed,
indicating mechanical looseness. It is not unusual to see harmonics
extending to 10X and higher. Some texts will refer to 1X, 2X, etc.
as "orders," where the nomenclature 5X would be referred to as 5
orders of running speed. Although the amplitudes of the
harmonics vary because of the phase relationship between the
harmonics, there is no particular interpretation that can be
attached to the amplitude of an individual or group of harmonics in
any analysis. However, it is possible to determine, to some extent,
the source of the looseness by taking amplitude measurements at
several locations on the machine, the amplitude will be higher the
closer you are to the source of looseness.

93. 92
Figure 1. Looseness from Broken Base Bolt.
When dealing with plain bearings, it is possible to generate
multiple harmonics when the bearing has become worn and is not
properly supporting the rotating shaft. When this occurs the shaft
will be loose in the bearing and will move in response to outside
forces such as unbalance and operational processes.
Figure 2. Shaft Loose in Worn Plain Bearing.
An exception to ignoring the amplitudes of the harmonics applies
to loose installations of bearings. It has been found that when the
third harmonic is higher than the second and fourth, the bearing is
loose on the shaft. When the fourth harmonic is higher than the

94. 93
third and fifth, the bearing is loose in the housing. This works in
both velocity and enveloped acceleration as seen in Figures 3 & 4.
Figure 3. Bearing Loose on the Shaft.
Note by the amplitudes in the Harmonic Box that the amplitude of
the third harmonic is nearly eight times the amplitude of the
second and fourth.
Figure 4. Bearing Loose in the Housing.

95. 94
Note, by the amplitudes in the Harmonic box, that the amplitude
of the fourth harmonic is more than four times the amplitude of
the third and fifth.
The looseness of bearings can be confirmed when they are
removed. If the bearing has been loose on the shaft, the inner
surface of the inner ring will be covered with a brown material
composed of water, rust and grease. The term applied to this
condition is corrosion fretting. If the bearing is loose in the
housing, this condition will be found on the outside surface of the
outer ring.

96. 95
Leak Detection
There are a number of acoustical instruments on the market that
provide a good tool to be used for leak detection. Although it has
not been developed as such a tool, SEE technology, with its
acoustic accelerometer, could be developed for this purpose. A
portable vibration data collector, such as the Microlog, can be
utilized as a valve and pipe leak detector using the acceleration
mode (G’s) for data collection.
In theory whenever any liquid or gas is forced through an orifice,
an acoustical signal is generated. Consider the sound of a toy
whistle or the sound of a liquid discharging from a nozzle. This
signal also transfers to the area surrounding the orifice and
imparts a mechanical vibration to the area. Empirical data
collected in the field has shown this theory to be true and logical,
sound is the vibration of air molecules that impact on the ear
drums and is interpreted as a multiple frequency sound. These
same vibrations impact the pipe or valve and cause a high
frequency vibration which can be detected by an accelerometer.
Field Experience
Nuclear power plants have many unique operating parameters.
When the plant is operating, the containment area where the
reactors are located cannot be occupied at any time. However,
there are many pieces of equipment that must be monitored,
specifically valves involved with the radioactive water flow inside
containment. The plant had 26 valves with drain lines installed in
the housing to collect any leakage, and this was collected and
measured at a manifold for a measure of total leakage.
Under Nuclear Regulatory Commission (NRC) rules at the time, if
the valve or valves that were leaking could be identified, then the
total leakage was allowed to be 2.5 gallons per minute (GPM) but
if the leaking valve could not be identified, then the allowable
leakage was only 1.5 GPM. Because all the valve leak collection

97. 96
lines lead to one manifold where the leakage was measured, it was
not possible to tell which specific valve or valves of the 26 were
leaking.
If the product, in this case water, remains at a constant pressure
and temperature, and a valve in the system begins to leak, it was
proven in the laboratory that there is a linear response between
the volume of leakage and the amplitude of the signal. Multiple
tests were performed using containment conditions and the data
plotted to determine the slope of this response curve which was
accepted by the NRC. At the next outage, accelerometers were
stud mounted on each valve body with the signal wires routed to a
safe area outside containment.
Thereafter, each valve could be monitored during operation, and if
it began to leak, using the developed leak vs. amplitude chart
(Figure 1), the volume of leakage could be calculated. This allowed
the plant to continue operations until the total leakage reached the
limit of 2.5 GPM. Although this seems to be a small matter, it
allowed the plant to operate much longer between shutdowns
which greatly increased its efficiency. When a down day costs in
excess of one million dollars, efficiency is very important.
Figure 1. Amplitude in G's vs. Flow Chart 0.34 G’s Equals 0.91 GPM.

98. 97
The U.S. Navy Laboratory needed to know if an off the shelf data
collector could detect a leak through a valve of 1/8th
of a GPM.
Although the techniques developed in the nuclear plant could have
been used, certain Navy applications required that they didn’t need
to know how much the valve was leaking, but just the fact that
there was a leak as small as 1/8th
of a GPM.
In the laboratory, a valve was connected to a water source, a
standard accelerometer was attached to the top of the valve stem
with a two bar magnet and with no fluid emitting from the
discharge, and a spectrum was taken, Figure 2.
Figure 2. Navy Valve with No Leak.
Not knowing what frequencies would be generated, the data
collector was ranged to its maximum, 1.2M CPM. There is a bit of
energy in the very low range which is attributed to background
noise in the system. Next, the valve was opened sufficiently to
allow a measured 1/8th of a GPM flow. With the same setup,
another spectrum was taken, Figure 3.

99. 98
Figure 3. Navy Valve with 1/8th GPM Leak.
When the two spectrums were overlaid in color, it was readily
apparent that the 1/8th of a GPM leak was detectable. The Navy
was satisfied that the technology worked and they were never
heard from again. Note that if an overlay presentation is prepared,
make sure that the amplitude and frequency scales are the same
so that the comparison is valid.
If it is desired to know if a valve or a pipe joint such as an air line is
leaking, another technique may be used. As noted previously,
when a gas or liquid passes through an orifice, a vibration is
induced into the pipe. As seen in Figure 3, this vibration covers a
wide frequency spectrum. However, because it is in the higher
frequencies, it will attenuate as it moves down the pipe. Testing
has determined that a leak generated signal will attenuate in
approximately 10 pipe diameters. For example, on a two inch line,
the vibration would attenuate within 20 inches.
The procedure is to mount with a magnet an accelerometer on the
valve stem of the valve. This is the ideal site because the valve
stem is connected directly to the valve plate. If it is not possible to

100. 99
reach the valve stem, mount the sensor on the valve body.
Caution: The temperature of a steam line will often exceed the
temperature limits for the accelerometer. In these cases, a steel
stand off should be used to prevent the destruction of the sensor.
A reading is taken using the acceleration or acceleration
enveloping mode and the amplitude noted. Next a reading is taken
at least 10 pipe diameters down stream of the valve with the same
setup. This second reading provides a background noise reading.
Compare the two readings and if the reading from the valve is 20
to 25% higher than the background reading, the valve is probably
leaking.
This procedure was used in a plant with a large number of valves
which were required by regulation to be examined for leakage.
These valves dumped into a large manifold and returned to a
storage tank. As with the nuclear plant, if a valve was leaking they
could tell by the level controls on the tank but they had no way to
determine which valves were leaking. The prior procedure was to
rebuild a portion of the valves at each outage, if they were leaking
or not, which of course they had no way of knowing. They would
start at the beginning of the outage and work on those valves until
it was time to resume operations.
Using leak detection, while the plant was in operation, a
measurement on each valve was taken and then ranked high to
low on a chart. Maintenance was instructed to start at the top of
the list and work their way down until valves were found to have
no leak indicators and then stop. After several cycles, the
amplitudes that were determined to be background noise were
documented and no more valves were opened that were below or
equal to that amplitude. This saved many man hours of
unnecessary work and allowed the crew to go to other jobs. As a
side note, historical data shows that overall, maintenance will have
to return to 11% of their overhauls and repairs and redo the work
because of a pinched gasket, missing bolt or some other mistake.
A wise man said, “If it ain’t broke, don’t try to fix it”!

102. 101
Detecting Rubs
On initial startup of a new machine, rubs may occur due to
misaligned bearing housings, an unbalanced rotor, misalignment
with the driver, a distorted base, the list goes on. In any event, the
prime FFT indicator of a rub between a rotating element and some
stationary portion of the machine is a signal at 0.5X. There may
be, or may not be, harmonics associated with this signal and they
may also be seen at 1/5, 1/4, or 1/3 of 1X signal. Because so
many submultiples of 1X are possible, the possibility of exciting a
resonance is increased. A usual indication of a resonance is if any
of the energy spikes in the FFT have a broad base and are
surrounded by a cluster of small spikes, as covered in the section
on resonance.
When data is collected at a lower resolution, in this case 400 lines,
the number on the value chart may not be exactly 0.5X as in the
first example where the chart indicates 0.488 for the signal. This is
an acceptable tolerance versus taking the additional time to take a
higher resolution spectrum. The examples chosen show both gE
and velocity plots of a rubbing rotating element.
Field Examples
The first example is an outboard motor bearing on a new
inexpensive motor. As often stated, you get what you pay for- this
motor was unacceptable. When the motor was started, the
vibration was excessive and it was shut down. When the FFT was
taken, the rub signal was quickly apparent and the motor was
removed to the shop.
On teardown it was noted that the outboard and inboard bearing
housings were offset from the centerline so that the rotor was
cocked to one side and there was a bright shiny area on the
housing where the rotor hub had rubbed. The motor was
discarded and the purchasing department told not to buy any
more of this brand.

103. 102
Figure 1. Velocity FFT of Outboard Motor Bearing.
When maintenance is deferred for financial, operational or other
reasons, machines will degrade until they break down or begin to
make enough noise that someone will pay attention to their
condition. In this case, the plain bearings in the centrifugal pump
had deteriorated to the extent that the impeller was rubbing on
the pump housing. An inspection disclosed the typical shiny metal
from the rub and a measurement of the bearing revealed the
excessive clearance due to wear. The bearing was replaced and
the pump returned to service.

104. 103
Figure 2. gE FFT, Impeller Rub.
As seen in this example, Figure 2, the speed of rotation does not
change the resulting FFT of a rub. The defining signal will appear
at 0.5X. Using enveloped acceleration will more likely show
harmonics as the enveloping process is more sensitive than
velocity. As with any vibration, the higher the amplitude, the more
urgent the need for repair.
The last example is an FFT displaying the results of a rub, and a
time spectrum is also included. The time spectrum does not add
anything to the diagnosis but verifies that the signal at 750 CPM is
prominent. Based on the observed amplitudes, it was decided not
to take the fan out of service at that time but to increase the
observations and trend the specific 750 CPM signal. This is
acceptable as many minor machine faults will operate for
extended periods without any problem, as long as they are
monitored on a cycle in accordance with their criticality. The more
critical the machine, the more often it should be monitored.

105. 104
Figure 3. FFT, Outboard Fan Bearing, Belt Drive.
Figure 4. Time Spectrum, Outboard Fan Bearing, Belt Drive.

106. 105
Blade Pass
Blade pass frequency has been in the news lately as more people
become annoyed at the noise made by electrical generating wind
turbines. As each blade passes by the support structure, the air
flow is disturbed and a pulsing noise/vibration is generated. The
frequency of the noise is determined by the rotating speed of the
blades and can be uncomfortable to some people. In a rotating
pump or compressor, a similar noise or vibration will be generated
as the compressor blade passes the discharge and the pumped
material is passed from the higher pressure housing to the
discharge pipe. This vibration will remain constant under constant
operating conditions, and will begin to change as wear takes place
and clearances between the rotating element and the housing
change. When it becomes severe, backflow may occur and the
noise and vibration levels will increase in the FFT. It is helpful to
trend the overall vibration levels to be alerted to this developing
condition. Until the wear conditions become excessive, blade pass
is not usually a problem unless the generated frequencies excite a
resonance in the intake or discharge piping. If this occurs,
additional piping supports should be installed at the maximum
vibration points to stiffen the pipe and change the response
frequency. If there is damage to just one of the vanes, the signal
will be seen at 1X because that vane only passes the discharge
point once per revolution.
Field Experience
Propeller driven aircraft will generate a blade pass frequency as
the propellers pass the leading edge of the wing, both on the down
sweep and the up sweep. For each rotation of the drive shaft, four
blades will pass the leading edge resulting in a spike of energy at 4
times shaft speed. Figure 1 is an example of blade pass taken in a
small twin engine turbo prop aircraft. The wing spar passed
through on the floor of the passenger compartment and was

107. 106
accessible in flight. The engine was rotating at 1550 RPM and the
blade pass frequency is seen at 6200 CPM with some harmonics.
Figure 1. Four Blade Propeller, Engine Speed 1550 RPM.
In the case of an aircraft, the blade pass frequency is important. A
major manufacturer of large aircraft had 3 aircraft crash in a short
time span several years after they were put in service. The aircraft
fleet was grounded and the investigation found that a harmonic of
the blade pass frequency was exciting the natural frequency of the
wing. For example, in Figure 1 it could have been the harmonic at
12,400 CPM. The result was an undetected flexing of the wing
during flight. After millions of cycles, metal fatigue crystallized the
metal and the wing broke off in flight. The solution was to add
additional structure and stiffen the wing and then to verify that the
resonant condition no longer existed.
Although the blade pass frequency is rarely a problem, in some
cases the pressure differential between the housing and the
discharge line can generate a vibration signal that will hide other
important signals when data is collected in amplitude auto range.
For example, Figure 2 is a spectrum from a screw compressor.
The vibration from the discharge, which is nearly 11 gE, drives the
signals from a bearing fault to the bottom of the spectrum where
it appears there is not a problem. It is necessary to always look at
the amplitude axis to verify the scale. In this case, the bearing
signal is at 2.6 gE, an indication of a serious problem. The unit was

108. 107
shut down for an inspection and the bearing was spalled around
the entire outer ring and all the rollers were damaged.
Figure 2. High Amplitude Discharge Signal.
Again, be alert when using auto range, it can give a visually
misleading impression when one signal is much higher than the
others, driving them down to a height that at a glance doesn’t
appear to be a problem.

110. 109
Unusual Situations
The Mystery Tape
The super secret B-2 stealth bomber was in production and a few
were flying. As with any aircraft, testing continues throughout the
life of the airplane as various modifications and upgrades are
installed. We were requested to come to the company test labs
and give our opinion about a situation they were working on.
When we arrived and were finally cleared through all the security,
we were taken to an electronics laboratory and shown a tape
recorder. We were told that the tape had signals on it and they
wanted to know what we thought the signals indicated. We asked
where the tape came from, couldn’t tell us, that was classified. We
asked what kind of equipment was being monitored, pump, fan, or
what. Couldn’t tell us, that’s classified. O.K., how fast is this
equipment rotating? Couldn’t tell us, that’s classified.
So we plugged the tape recorder into the data collector and
collected the spectrum in Figure 1.
Figure 1. FFT from Aircraft Mystery Tape.

111. 110
Figure 2. Time Spectrum from Mystery Tape.
Based on the harmonic content of the signal and the location of
the first peak, we told the group that what ever this thing was, it
was rotating at 2,325 RPM and it appeared to be loose in its
mounting or had a loose component.
With that, the lead technician slapped his hand down on the table
and yelled, “I told them that #**% thing was loose!” He then
turned to us said “Thank you” and showed us the door. After all
that, all we know is that it was some part of a B-2 bomber.
The B-1 Bomber
The B-1 bomber is a swing wing aircraft that has been in several
war time actions during its life span. During post flight inspections
it was noticed that pieces of Teflon were around the base of the
wing pivots. These pivots are constructed of a hollow titanium post
approximately two feet in diameter with walls 2 inches thick and a
solid bottom. During assembly it was filled with liquid nitrogen.
The next part installed is a Teflon shell around the post, followed
by a steel shell over the Teflon to which are fastened the wing
spars and support equipment. When the titanium returns to room
temperature, the whole assembly is completely locked together.
This results in a wing pivot post that uses the Teflon shell as a
lubricating surface between the moving metal parts. Finding pieces

112. 111
of Teflon on the deck indicated unexpected wear on this critical
component. We were asked if we could detect the metal to metal
contact that might take place if areas of the Teflon were displaced.
In our lab we laid out a strip of Teflon on a steel lath bed. We
placed the accelerometer on the end of the lath. A nail was then
dragged along the Teflon strip, then off the strip while a long
duration time spectrum was collected. The jump in the noise level
showed that we could tell the difference in the metal on Teflon vs.
metal on metal.
We went to Dyess AFB, TX where the B-1’s were based. We were
taken to two aircraft. They thought the first had a problem - the
second was not suspect. They didn’t tell us which aircraft was good
and which one was bad, they wanted us to tell them. Their
concern was about noises heard in the cockpit when the wings
were sweeping. Did it have a serious missing Teflon problem?
We instrumented both aircraft, one at a time, with four data
collectors, two on each wing. Before we went to the base we were
told that the wings would sweep from 15 to 67.5 degrees and we
could collect data while they were sweeping. After we arrived, they
told us that they could only move the wings 5 degrees at a time
because, on the ground, the wings drooped and had to be
supported by rolling stands while moving. The stands wouldn’t
move with the wing unless someone pulled them along with the
wing movement. That put a monkey wrench in our original plan so
we went to plan B and measured each wing as it moved 5
degrees, waited while they repositioned the stands and took data
for the next 5 degrees until the wings were fully swept to the aft
position.
There are times when you find things you are not expecting. In
reviewing the data it was noticed that something was generating
multiple harmonics. We asked what was rotating at 141 RPM and
were told it was the jack screw that moves the wings forward and
aft. We suggested that they check the mountings because it
appeared there was some minor looseness. The Colonel said that

113. 112
wasn’t possible, but we insisted there was looseness. Turned out
the mounting plate for the jack shaft was loose where it attached
to the aircraft bulkhead. Figures 3 and 4 are the time and
frequency spectrums of the loose jack shaft. We didn’t hear any
more from the Colonel.
Figure 3. Time Spectrum of Loose Jack Shaft.
Figure 4. Frequency Spectrum of Loose Jack Shaft.

114. 113
Based on what we had done in the laboratory we knew the
amplitudes would be higher for metal to metal vs. metal to Teflon
contact. Using the overall amplitudes we plotted the amplitude for
each 5 degree sweep. Figure 5 is a plot of those amplitudes.
0
0.01
0.02
0.03
0.04
0.05
0.06
0.07
0.08
0.09
0 10 20 30 40 50
Figure 5. Plot of Overall Amplitudes, Aircraft #1.
Figure 6 is the plot of the overall amplitudes for the second
aircraft.
0
0.01
0.02
0.03
0.04
0.05
0.06
0.07
0.08
0.09
0.1
0 10 20 30 40 50
Figure 6. Plot of Overall Amplitudes, Aircraft #2.
It is apparent that there is much more scatter in the plot for
aircraft #2. Based on this comparison we reported that the #2
aircraft pivot pin appeared to have metal to metal contacts. The

115. 114
reply was that was what they had found also, they just wanted us
to confirm that their method using capacitance measurements
worked. So we didn’t get to test all of the B-1’s!
Water Hammer
A major manufacturer opened a new plant, and within a short
time began to have failures on the metal cutting positive
displacement coolant pumps. After several months it was decided
that the failures were caused by a vibration problem, and we were
asked to find and solve the problem.
During the cutting cycle, the pumps would operate at an initial
high pressure of 100 PSI. After the initial cuts were made, a valve
would automatically close and the pressure would be reduced to
75 PSI for the remainder of the cut. The failures were primarily
with the pistons they were cracking and breaking apart. We
examined the pieces and noticed that the steel screws securing the
cylinder heads to the aluminum housing were literally yanked out
of the aluminum and contained shreds of the aluminum in the
threads of the screws. We then collect the time domain spectrums
as seen in Figure 7.
Figure 7. Time Domain of Water Hammer.
The time of this pulse is 0.07 seconds and represents what
happens when you slam a valve closed when the pressure is 100
PSI. It is called a water hammer and is very destructive. Water

116. 115
moving at a high velocity has a large amount of kinetic energy.
When you slam that mass into a fast closing valve, the mass
impacts the entire line because a liquid does not compress.
In seeking a solution, we recommended a “T” bar pipe be installed
on the coolant line so that the fluid could surge into the “T” bar
and the trapped air would cushion the hammer. They then
contacted the cutting machine manufacturer and asked their
opinion. Their answer was, “Oh, don’t you have the new
modification, we’ve added a surge chamber on that line.” No, they
didn’t have the modification, but they do now.
Trending
The usual sequence of events in the slow degradation of a
machine’s condition can usually be seen in the overall vibration
amplitudes. If there are hundreds of machines being monitored, it
becomes physically impossible to look at all the data that is
collected. Therefore, the analyst must rely on the software to
evaluate changes that take place over time.
The computer is nothing more than a large data storage box. The
software will organize and monitor the data within the parameters
that you establish. Therefore it is not necessary to look at all data
from each machine but to monitor the results provided by the
software in a programmed alarm report. This report, for example,
will alert the analyst as to which machines have increased
vibration levels that exceed the programmed alert level. When the
amplitude continues to increase and the programmed alarm level
is exceeded, another report is issued.
When these reports are received, the analyst knows to look at that
specific machine and begin a detailed analysis to determine the
cause of the increasing vibration. Often this analysis will find that
nothing is amiss but that the alert/alarm setpoints are set too low.
Where a machine failure was unexpected, it may be found that the
alert/alarm was set too high. In both cases, the setpoints can be
adjusted to reflect the reality of the actual conditions.

117. 116
This ability to adjust the setpoint of any of the multiple monitoring
parameters is a very strong point for the software. Do not rely of
default setpoints, adjust them to work for you.

118. 117
Conclusion
On a personal note, during my working years I've had the
opportunity to work in three separate career fields. First, as a pilot
in the Air Force, second as a plant engineer in four different
industries and finally as an applications engineer for SKF Reliability
Systems. I enjoyed each "career," but the applications job was the
most rewarding. Each time a problem was presented, in any of the
35 countries in which I worked, an opportunity existed for me to
make a difference in some aspect of a production facility. Whether
it was to improve efficiency, save money, or make the machines
safer to operate, every job was important.
It should be said that there is a lot of satisfaction in knowing that,
at the end of the day, your actions have made a positive impact -
and it’s even better if you’re able to have fun while you were doing
it. It's true when they say, "if you enjoy your job, you’ll never have
to work a day in your life."
And when tomorrow comes, you get to do it all over again
because; even if the machines are the same, there will be
differences in their responses to the forces in effect.
There's always plenty to do, and YOU too can make a difference.
Good hunting,
Dr. Bob