Vibration Analysis calculation ans methode.pdf

© GermanischerLloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration analysis and calculation, shock
and noise
SaptarshiBasu–GL trainer

© GermanischerLloyd 2010 Propulsion system Integration 13~25/02/2012
ILL EFFECTS OF NOISE AND VIBRATION
•Excessive ship vibration is to be avoided for
passenger comfort and crew habitability.
•In addition to undesired effects on humans, excessive
ship vibration may result in the fatigue failure of local
structural members or malfunction of machinery and
equipment.
•For naval application noise and vibration means
decrease in stealth capability, increase in detectability
and susceptibility.
No. 2

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VIBRATION MONITORING
•VIBRATION THOUGH HAS SEVERAL UNDESIRABLE
EFFECTS ITS CHARECTERISTICS CAN BE
MONITORED TO GET INFORMATION ABOUT THE
HEALTH OF THE MACHINERY AND ANY IMPENDING
DISASTER OR HIDDEN DEFECT.
•ADVANCED ANALYSIS TECHNIQUES SUCH AS THE
FOURIER FAST TRANSFORM and INSTRUMENTS
LIKE SHOCK PULSE METER.
No. 3

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VIBRATION MONITORING
•Real-Time Vibration Monitoring System (RTVMS)
produces real-time vibration spectral data from critical
components discrete spectral signatures, which are
prime indicators of machinery health, can be assessed
at high speeds and utilized to mitigate potential
catastrophic engine failures.
•The vibration source data (accelerometers) from the
must be acquired at high sample rates in order to
provide the best time and frequency resolutions in the
frequency domain for performing enhanced engine
health monitoring.
No. 4

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Real-Time Vibration Monitoring System
(RTVMS)
•After the high-speed acquisition task is performed, the
data must be transferred rapidly to Digital Signal
Processing (DSP) modules.
• The DSP modules take the digital data, perform a Fast
Fourier Transform (FFT) to produce frequency
spectral data, and summarily run the pertinent health
algorithms on-board the DSP chips.
•The complexity of the DSP operations is dependent on
the health algorithms being utilized
No. 5

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Real-Time Vibration Monitoring System
(RTVMS)
•The DSPs can act independently if the processing
load is light or may be utilized as multiple parallel
processors (MPP) if the processing tasks are heavy.
•The use of MPP operations allows a system to utilize
multiple DSP’s, which communicate through the DSP
communication ports, and act in parallel to perform
immense processing tasks without any additional and
harmful processing latency. The system is a real-time
distributed processing system that performs Multiple
Instructions on Multiple Data (MIMD).
No. 6

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Health Management Computer Block
Diagram
No. 7

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
HUMAN ERROR related human fatigue
•Human error resulting from fatigue-impaired performance
has been identified as the cause of numerous
transportation mishaps.
•Incorporating human factors into a ship’s design can help
combat fatigue, increase alertness, and decrease human
error.
•Human Error – Any deviation from a system performance
standard which is caused indirectly or directly by an
operator and which has significant consequences to the
system operation in which it was made
No. 8

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
OSVs: Crew and Safe Operations
▪Late 90’s – Considerable increase of accidents, mostly
in the North Sea:
-Supply vessels
-Anchor handling vessels
-Rescue vessels
▪Last days/hours of rotation/watch0
2
4
6
8
10
12
14
1995 1996 1997 1998 1999 2000

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
What are the causes of your sleep problems at sea?
3
4
6
8
11
12
15
18
22
28
29
32
51
0 10 20 30 40 50 60
Snoring
Physical problems (write):
Feel that something on board is not safe
Conflicts/worries at work
Other (write):
Conflicts/worries at home
Bothered by shift-working
Need to get up to urinate
Bad mattress/pillow/duvet
Wakened by alarms or other loud noises
Cabin temperature/humidity
Vibration in cabin
Noise
Percentage
Reprinted with permission from Sintef
Sleep problems for OSV crews

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Effects of human fatigue
•Fatigue drastically reduces human alertness levels and negatively affects job
performance.
•Although fatigue is difficult to define, it can be generalized as "impaired
alertness“.
•Fatigue affects humans in different ways, although most people suffer from:
•decreased problem solving ability
•increased risk taking
•delayed reaction time
•moodiness
•inability to concentrate, and
•inattentiveness.
•Logical reasoning and decision-making are affected by fatigue and it impairs
human physical abilities such as strength, speed, coordination, and balance
No. 11

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
CAUSES OF FATIGUE
•Fatigue may result from:
•poor sleep quality
•sleep deprivation
•physical/mental exertion
•emotional stress
•disruption of circadian rhythms
•poor physical condition,
•drug/alcohol use.
•Everyone has felt the effects of fatigue. It is usually described as an
uncontrollable urge to sleep or rest. It has also been described as a "fog"
that comes over the brain at certain times of the day. Of importance is that
fatigue lowers alertness levels and impairs performance.
No. 12

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
FACTORS AFFECTING SLEEP
No. 13

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
RESTORATIVE SLEEP ENVIRONMENT
•The sleeping environment determines a person’s
ability to get three of the four factors of restorative
sleep, namely quality, continuity, and quantity.
Quantity and time of day are impacted by work
schedules and operational commitments.
•The design factors that create a good sleeping
environment also impact the watch station, working,
and recreational environments.
No. 14

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
SLEEP DEPRIVATION & ALTERNESS
Effects of Reduced Sleep on
Alertness Levels
Effect of Successive Days of
Reduced Sleep on Alertness
Levels
No. 15

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
DESIGN FACTORS AFFECTING SLEEPING
ENVIRONMENT
•The foundation of the diagram is the six design factors, all of which naval
architects and marine engineer’s can directly control.
•These factors determine how comfortable the working and sleeping
environments are aboard a ship:
•lighting
•noise
•vibrations
•ventilation
•temperature, and
•ship motions
•These all can be incorporated through the preliminary design of a vessel
and maintained throughout ships post-production life
No. 16

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
NOISE AND SLEEP
•Noise affects sleep patterns, which greatly contributes to
fatigue. It makes it difficult to fall asleep, can wake a
person throughout the night, and pulls a person from
deeper to lighter sleep stages.
•Nightly interruptions can get so frequent that a person
may begin to forget that they were awoken and return to
sleep very quickly.
•This pattern is particularly dangerous because the person
is not getting enough deep sleep and will be drowsy the
next day.
No. 17

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
NOISE LEVEL AND SLEEP DISTURBANCE
•The noise levels at which sleep disturbances occur are low. Levels of 40 to 50 dBA
(lower than a casual conversation) have caused difficulty in falling asleep and has
extended the time of falling asleep to one hour.
•As the sound levels increase it becomes more difficult to fall asleep.
•Three other important findings are listed:
•70 dBA is enough to significantly change the sleep patterns of most subjects.
•Long-term exposure to noise affects sleep.
•Short sound duration awakens more than long and steady noise.
The effect that noise has on sleep challenges designers of shipboard general
arrangements. Finding the optimal location for sleeping quarters and crew
recreation compartments is critical. Noise is an important factor and is not usually
considered with fatigue in mind
No. 18

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
HEALTH EFFECTS OF NOISE
•Noise can be defined as unwanted or undesirable
sound. It is present in most compartments of a ship
and it is difficult to avoid.
•Noise comes from numerous sources including
engines, generators, pumps, and air conditioners.
There are many human physiological and physical
impacts of noise in the workplace that cause fatigue
and negatively impair human performance.
•Noise also affects sleep patterns and decreases the
restorative quality of rest.
No. 19

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
NOISE AND HEARING LOSS
•Long-term exposure to excessive noise can result in
permanent hearing loss.
•The extent of the hearing damage is dependent upon
noise intensity and frequency.
•Temporary loss of hearing is the result of short-term
exposure to noise and can lead to permanent hearing
loss.
No. 20

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
PHYSIOLOGICAL IMPACTS OF NOISE
•There are also physiological impacts of noise and these have
not been adequately addressed in the guidelines used to
establish acceptable noise levels. Although the physiological
effects are less perceptible, they have a considerable impact on
human performance and this makes them the most dangerous
cause of noise induced fatigue.
•The physiological changes that occur due to noise are the result
of the natural "fight or flight" response of the human body. The
body perceives all noise as a threat or warning of danger and
continuously responds to it accordingly, even at low noise
levels and while a person is asleep
No. 21

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Mariners working in a noisy environment tend to be
moody, irritable, and unable to effectively deal with
minor frustrations.
•Noise causes blood pressure to go up, increases heart
and breathing rates, accelerates the metabolism, and
a low-level muscular tension takes over the body
("fight or flight" effects).
•The physiological changes described above also
occur when a person is asleep, affecting their ability
to get restorative sleep and leading to fatigue.
No. 22
PHYSIOLOGICAL IMPACTS OF NOISE

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Problems caused by this type of stress are listed below:
•(a) Neuropsychological disturbances: headaches, fatigue,
insomnia, irritability, neuroticism
•(b) Cardiovascular system disturbances: hypertension,
hypotension, cardiac disease
•(c) Digestive disorders: ulcers, colitis
•(d) Endocrine and biochemical disorders
•(e) Sleep disturbance
No. 23
PHYSIOLOGICAL IMPACTS OF NOISE

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
NOISE EXPOSURE
•Noise levels that cause the human body to respond in
these ways vary with individuals.
•If the noise continues for long periods, the factors
compound and it becomes harder to relax. The factors
increase as the noise levels increase.
•The Occupational Safety and Health Administration
(OSHA) and numerous human factor design guidelines
have prescribed values for intensities and exposure
duration at which operators can safely be subjected to
noise.
No. 24

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Definitions
•Peak Sound Pressure (P
peak)
•This is measured in units of pressure (pascals) and not decibels. It is the maximum
instantaneous noise pressure measured on a 'C' weighted frequency scale. 'C' weighting is
used as it is almost a linear weighting which is incorporated in many commercial sound level
meters.
•Daily Noise Exposure Level (L
EX,8h)
•Because noise exposure depends on both the amplitude of noise and the duration of
exposure, the daily noise exposure level is the time weighted average of the noise level
experienced. It is normalized for an 8 hour working day so that if for example the exposure
time per day is more than 8 hours, the noise level to which an employee is exposed must be
reduced.
•The L
EX,8h is a direct replacement of the old L
EP,d and is measured in dBA.
•Weekly Noise Exposure Level
•This is simply the time weighted average of daily noise exposure levels for a standard 40 hour
working week.
No. 25

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Noise exposure control
•There are two action values and a limit value:
•Action Value
•The two action values are:
•1. The Lower Exposure Action Value
•2. The Upper Exposure Action Value
•(1) The Lower Exposure Action Value is 80dBA L
EX,8h and peak
pressure P
peak of 112 pascals.
•(2) The Upper Exposure Action Value is 85dBA L
EX,8h and peak
pressure, P
peak of 140 pascals.
•These 'Action Values' do not take into account the attenuating
effect of ear protectors that employees would be wearing.
No. 26

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Limit Value
•The exposure limit value is similar to the action value
except that the attenuation provided by ear protection is
taken into account. The exposure limit value is 87dBA
L
ex,8L and peak pressure P
peak of 200 pascals.
•Variable Daily Exposure
•In some industries particularly where production is of a
batch rather than continuous nature, noise exposure
varies greatly from day to day. When this the case, the
Directive suggests the use of a weekly noise exposure
level.
No. 27
Noise exposure control

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Reducing Noise Exposure
•Wherever possible, reduction of noise at source is
preferred. In particular the employer should use quiet
working methods and equipment. Work places and
work stations should be designed to minimize the
noise exposure.
•The employer should reduce noise using suitable
techniques depending on whether noise is
predominantly air borne or structure borne. Also
equipment should be well maintained as it is know
that poorly maintained equipment tends to be noisier.
No. 28

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Work should be planned so that times spent in noisy situations is limited
and that rest from noise is provided.
•Upper Action Value
•In particular, if the upper action value is exceeded the employer is to reduce
noise by either technical means and or organizational means. Warning signs
must be displayed and if possible access to these noisy areas restricted.
•Ear Protection
•Once the possibility of noise reduction by technical or organisational means
has been eliminated, efficient ear protection must be made available and
must be used. The employer must see that this is done.
•There are no circumstances in which employees can be allowed to exceed
the Exposure Limit Value.
No. 29
Reducing Noise Exposure

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
TIME EXPOSURE OF NOISE
No. 30
The sound levels and permissible duration exposure time per day given by OSHA
are listed below and are a good generalization of standards used by the military:
Sound Level and Duration per day

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
NOISE LIMITS AS PER SPACE
No. 31

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Sound pressure
SOUND PRESSURE LEVEL
SOUND PRESSURE
Woods
Bedroom
Living roomLibrary
Business office
Conversation
Street traffic
Heavy truck
Pneumatic
chipper
Pop group
Jet take-off
at 100 meters distance
Jet engine
at 25 meters distance
140 dB(A) Threshold of Pain
130 dB
120 dB
110 dB
100 dB
90 dB
80 dB
70 dB
60 dB
50 dB
40 dB
30 dB
20 dB
10 dB
0 dB (A) Threshold of Hearing20 μPa
100 μPa
1.000 μPa
10.000 μPa
100.000 μPa
1.000.000 μPa
10.000.000 μPa
100.000.000 μPa

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Sound pressure
Airborne + 62 dB = Waterborne

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
IMO on underwater noise
◼The Marine Environmental Protection
Committee (MEPC) of the International
Maritime Organization (IMO) July 2009:
“The committee urged governments to
review their commercial fleets to
identify the ships that contribute most
to underwater noise pollution”
◼IFAW estimates that the noisiest 10%
of ships contribute the majority of the
noise problem

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Noise sources
Propellers,
thrusters
Diesel engines, generators, electric
motors, gears
Water
flow

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Underwater Sound Levels
Ships
underway
Broadband source
level
(underwater dB at 1 m)
Tug and barge 171 dB
Supply ship 181 dB
Large tanker 186 dB
Source Broadband
source level
(underwater dB at 1 m)
Grey whale moans 142 - 185 dB
Bowhead whale tonals,
moans, and song
128 - 189 dB
Humpback whale song 144 - 174 dB
Photo source: www.iucn.org

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Noise transmission
•Audible noise can be broken down into two
categories: (1) airborne and (2) structure-borne.
•Airborne noise is what causes stress and hearing
loss.
•Structure-borne noise induces vibrations that can
damage machinery and marine structures. Both of
these noise types vary in frequency and intensity.
No. 37

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Noise Control
•A ship designer must consider a number of different issues when designing to reduce and
control noise. There are three locations where noise can be minimized and four basic
methods of controlling it.
•Isolation: Minimizes noise by reducing vibrations caused by machinery or equipment.
Vibration-absorbing materials are used such as rubber mounts, pads, or springs. The type of
material for an application is based on weight, vibration frequency, and desired degree of
isolation.
•Barriers: Minimizes noise by blocking sound transmission through the use of high mass,
resilient, or limp mass materials. Using more mass increases the effect and barriers work
better at higher frequencies.
•Damping: Minimizes noise by adding mass to the vibrating structure or by connecting it to a
surface that does not want to vibrate. Damping materials are selected by considering the
thickness of the vibrating surface, the desired reduction, and the environment.
•Absorption: Minimizes noise with resonators and open-celled porous material, which
converts sound energy to heat. Materials used are based on the noise frequency, desired
reduction, and environment.
No. 38

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Locations of barriers for noise transmission
•In most applications, the use of barriers is the most effective
means of reducing airborne noise. In order for barriers to be
effective designers must use the proper absorption materials.
These materials can be heavy, expensive, and take up critical
space. Despite this, specially enclosed workspaces can have
as much as an eight to nine dBA reduction.
No. 39

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
HEALTH EFFECT OF VIBRATION
•Vibrations resonate throughout the hull structure and
the entire crew can be affected.
•The propagation of these vibrations along the decks
and bulkheads subject the crew to whole body
vibration and noise.
•The effects of whole body vibration are well studied
and documented.
•There are two types of effects short term and long
term
No. 40

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
PHYSIOLOGICAL IMPACTS OF VIBRATION
•short term effects:-
•headaches,
•Stress, and
•Fatigue
•Long term effects:-
•hearing loss
•constant body agitation
•musculoskeletal injuries,
•back disorders, and
•bone degeneration.
No. 41

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Physiological:
•Cardiac rhythm increases
•Respiration rhythm increases
•Blood circulation increases
•Vasoconstriction
•Endocrine secretions
•Central nervous system affected
•Comfort and Performance:
•Pain
•Nausea
•Vision problems
•Posture
•Movement and coordination decline
•Force
•Perceptions altered
No. 42
HEALTH EFFECT OF VIBRATION

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration Exposure Criteria
No. 43
Vibration Exposure Criteria for the
Longitudinal Directions with Respect to
Body Axis
Vibration Exposure Criteria for the
Transverse Directions with Respect to
Body Axis

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration Control
•Listed below are the three effective ways:
•Source Control
•- Reduce vibration intensity
•- Avoid resonance
•Path Control
•- Limit exposure time
•- Reduce vibration transmission (structural dampening)
•- Use vibration isolators
•Receiver Control
•- use vibration isolators
•- adapt posture
•- reduce contact area
No. 44

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
45
MIL-STD-167-1A changes from previous version
•Definition of response prominence
•A resonance with transmissibility (output/input) greater than 1.5:
•“Response prominence is a general term denoting a resonance or other distinct
maximum, regardless of magnitude, in a transmissibility function, including local
maxima which may exist at the frequency endpoints of the transmissibility function.
Typically, a response prominence is identified by the frequency of its maximum
response, which is the response prominence frequency. A response prominence of a
system in forced oscillation exists when any change, for both plus and minus
increments however small, in the frequency of excitation results in a decrease of the
system response at the observing sensor registering the maximum. A response
prominence may occur in an internal part of the equipment, with little or not outward
manifestation at the vibration measurement point, and in some cases, the response
may be detected by observing some other type of output function of the equipment,
such as voltage, current, or any other measurable physical parameter. Instructions on
how to identify response prominences is provided in Appendix A”.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
46
Balance Definitions
•Rotor, flexible – A flexible rotor is one that does not meet the criteria for a
rigid rotor and operates above its first resonance. The unbalance of a
flexible rotor changes with speed. Any value of unbalance assigned to a
flexible rotor must be at a particular speed. The balancing of flexible rotors
requires correction in more than two planes. A rotor which operates above
n resonances requires n+2 balance planes of correction. A rotor which
operates between the second and third resonances, for example, requires 2
+ 2 balance planes of correction.
•Rotor, rigid – A rotor is considered to be rigid when its unbalance can be
corrected in any two arbitrary selected planes and it operates below its first
resonance. After correction, its residual unbalance does not exceed the
allowed tolerance, relative to the shaft axis, at any speed up to the maximum
service speed and when running under conditions which approximate
closely to those of the final supporting system.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
47
MIL-STD-167-1A
•15 SECOND TEST/ ID RESPONSE PROMINANCES
AND/OR RESONANCES
•5 MIN SCREENING TEST/ 40% OF CYCLES
UNOFFICIAL ENDURANCE TEST
•2 HR ENDURANCE TEST
•FOR ONE CLASS INSTALLATION, TEST UP TO AND
INCLUDING: F = (DESIGN RPM/60) x (NUMBER OF
PROPELLER BLADES) x 1.15 NOTE: ROUND ANSWER UP TO
NEXT HIGHER INTEGRAL FREQUENCY

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
48
DRAFT MIL-STD-810G METHOD 528
•PRETEST
•CHECK BOLTS AND WASHERS BEFORE, DURING (WHEN
CHANGING DIRECTION OF VIBRATION), AND AFTER TEST.
ENSURE ALL BOLTS ARE PROPER GRIP LENGTH AND THAT
THE WASHERS ARE NOT ROTATING.
•IDENTIFY POTENTIAL SIGNS OF HIGH STRESS
CONCENTRATION. CONSIDER COMPOSITE AND CAST
MATERIALS.
•TAILORING : TAILORING IS A FUNCTION OF THE QUALITY OF
THE COMBATANT’S PROPELLERS. THE ALTERNATING
THRUST IS HIGH FOR FLAT PROPELLERS AND DECREASES
SIGNIFICANTLY FOR MODERN HIGHLY SKEWED
PROPELLERS.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
49
Draft MIL-STD-810G, METHOD 528
•Definitions – Random Vibration
•Power Spectral Density – mean square value of the signal in the
frequency interval ∆f, at the centered frequency in g
2
/Hz. In
most cases, the random vibration test would be less severe than
the sinusoidal test of MIL-STD-167-1A; however, it may be
more representative of the actual environment. However, a
waiver is required from SEA 05P1.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
50
Draft MIL-STD-810G, METHOD 528
•BASIS OF USING RANDOM VIBRATION TESTING
•Actual ship environmental data is available and indicates
random vibration is the best representation of the environment.
•The Alternating Thrust of the propeller is below 1.5 % of mean
thrust.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
51
Draft MIL-STD-810G, METHOD 528
•Vibration Environment
•Validate the accelerometer's sensitivity before and after testing.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
52
ACCELEROMETER PROBLEMS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
53
Vibration Standards
(.557/.45 =1.2 test levels exceed ship environment)
(.297/.075 = 4.0)
(.28/.05 = 5.6)
For stern of combatant test levels are 7 times higher than shipboard measurements (⅓ (1.2 + 4.0 +
5.6) = 3.6); (Assuming Mag. Factor of 2: 3.6 X 2 ≈ 7)

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
54
Isolation Mountings
•For Type I testing of material to be installed shipboard on isolation mounts, testing
shall be performed on isolation mounts or hard mounted to the testing machine, or
as specified. Type I testing of a particular test item on isolation mounts is valid only
for the isolation mount type and configuration used during testing. Ensure the
transmissibility across the mounts does not exceed 1.5 within the blade frequency
range of 80% to 115% of design RPM. If material is tested for Type I vibrations hard
mounted to the test fixture throughout the duration of the test, the test is valid for
either hard mounted or isolation mounted shipboard installations, provided the
isolation mounts are Navy standard mounts contained in MIL-M-17191, MIL-M-17508,
MIL-M-19379, MIL-M-19863, MIL-M-21649, MIL-M-24476, or distributed isolation
material (DIM).
•The endurance test is for a total period of two hours at the frequency moist seriously
affecting the equipment. One of the frequencies selected should be the isolation
mount frequency if the test is to be performed on isolation mounts.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
55
Design Power Calculation
•IAI = ω
2
X
•F = ma = ω
2
Xm
•T(torque) = ω
2
X
2
m
•P(power) = T ω = ω
3
X
3
m
•P = (2Πf)
3
X
2
m; f= RPM/60
•P ≈ (RPM)
3
•½ Power ≈ (1/2)
⅓
≈ 0.8 Design RPM
•½ Power ≈ 80% Design RPM
Record Response Prominences
•Output/input = 1.5 and greater

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
56
Balancing
•A process of minimizing the distance, е, between the mass centerline and the
geometric centerline of a rotor.
•G: balance quality grade in mm/sec.
•G = ω е (eccentricity from balance in mm)
•е = G/ ω (units of ω are rad/sec)
•Note “е” is the distance between the shaft axis and rotor center of gravity.
•Mass unbalance occurs when the center of mass of a rotor does not coincide with
the rotor’s geometric center.
Mass
center
Geometric center е

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
57
Balance Limits for Rigid Rotors
•When balanced as specified, the maximum allowable residual unbalance is given by
the following formula:
•Given:
•U = Wе and G = ω е = 2 Πf е
•U is the maximum allowable residual unbalance
•G is the total balance quality grade (mm/sec) as specified
•W is weight of the rotor (lbs)
•N is the maximum rotor rpm
•е is the eccentricity limit (mm)
•It can be shown that
• U = 60GW / 2 Π N (lbs – mm)
•or U = 6GW/N (oz – in)
•For rigid rotors that operate below 1000 rpm, the total balance quality grade shall not
exceed G = 2.5 mm/s. For rigid rotors that operate at 1000 rpm and above, the total
balance quality grade shall not exceed G = 1.0 mm/s. For rigid rotors that require
low noise, a balance quality grade of G = 1.0 mm/s can be specified for all speeds.
For guidance on balance quality grades of rigid rotors, see ANSI S2.19.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
58
Balance Quality Grades
•Application of U = 6GW/N (U in oz/sec, G in mm/sec, W in lbs, N in RPM)
•In MIL-STD-167-1, U(per plane) = 4W/N
•What is the balance quality grade, G, for this requirement ?
•Set 6GW/N = 4W/N ; 6G = 4; G = 2/3 per plane
•G (total) = 2(2/3) = 4/3 mm/sec for two planes
•In MIL-STD-167-1A, G(total) has been reduced to 1 mm/sec for noise
sensitive rotors.
•This is a more severe requirement. Since:
•G(total) = ω е (e is ecentricity from balance in mm)
•The distance “e” between the shaft axis and the rotor center of gravity has been
reduced in MIL-STD-167-1A

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
59
Balance Quality Grades
•Allowable Unbalance Example
•What is the maximum allowable residual unbalance, U in oz-in, For a 500 lb
Rotor which is to be balanced to balance quality grades of 2.5, 4/3 and 1.0
mm/sec at 6000 RPM?
•For G = 2.5 mm/sec : U = 6G(W/N) = 6(2.5) 500/6000 = 1.25 oz-in total and G =
ω е ; е = G/ ω = 2.5/(2 Π 6000/60) = 2.5/200 Π = 0.004 mm allowed eccentricity
(ω = 2 Π f = 2 Π RPM/60)
•For G = 4/3 mm/sec (Old 4W/N requirement) : U = 6GW/N = 6(4/3) 500/6000 =
2/3 oz-in total ; G = ω е ; е = G/ ω = 4/3/(2 Π6000/60) = (4/3) 1/628 = 0.0021 mm
allowed eccentricity
•For G = 1.0 mm/sec ; U = 6GW/N = 6(1) 500/6000 = ½ oz-in total; For low noise
rigid rotors G= ω е ; е = G/ ω = 1.0/(2 Π6000/60) = 0.0016 mm allowed
eccentricity.
•The lower the balance quality grade, the quieter the rigid rotor because of the
small eccentricity, е.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
60
Approval Process for Vibration
Testing Data Reports
Certified Government
Vibration
Testing Approver
Test from
Certified Manufacturer
Tester
Test from
Certified Contractor
Tester
Test from
Certified Government
Tester
Disapprove
Approve
NAVSEA

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VIBRATION DEFINATIONS
61
•Vibration may be in any linear direction, and it may be rotational (torsional).
Vibration may be resonant, at one of its natural frequencies or forced.
•It may affect any group of components, or any one. It can occur at any
frequency up to those which are more commonly called noise.
•As ship design advances, particularly with regard to structural optimization
and high speeds to meet market demands, there is a tendency for noise and
vibration problems to become more pronounced.
•Design practice should include elements of model testing, calculation and
heuristic deduction from previous experience

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The total vertical movement is called the Peak to Peak Displacement of the vibration.
This is an indication of the amount of the amount of lateral movement of the
machine and is good indication of the amount of out of balance in a machine when
the value is compared to a standard for that machine. This parameter is often used
when balancing.
•The Vibration Velocity is the speed of movement of this point ,being highest as the
point passes through its at rest position. It gives good guide to the amount of
energy being generated by the vibrating object. This energy usually results in wear
and eventual failure.
•The amount of energy is proportional to the square of the velocity of vibration.
Velocity being a good indication of the amount of wear taking place in a machine is
used exclusively in monitoring systems.
62
VIBRATION DEFINATIONS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•For analysis purposes the r.m.s value is used.
•For very low speed machines where the velocity is low
the displacement may be used instead.
•The Vibration Frequency is the time taken to complete
one cycle .The shaft below is said to have a
fundamental frequency equal to the shaft rotational
velocity.
63
VIBRATION DEFINATIONS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
SHAFT VIBRATION
•Any elastically coupled shaft or other system will have one or
more natural frequencies which, if excited, can build up to an
amplitude which is perfectly capable of breaking crankshafts.
•‘Elastic’ in this sense means that a displacement or a twist
from rest creates a force or torque tending to return the
system to its position of rest, and which is proportional to the
displacement.
•An elastic system, once set in motion in this way, will go on
swinging, or vibrating, about its equilibrium position forever,
in the theoretical absence of any damping influence
64

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
FUNDAMENTAL OR NATURAL FREQUENCY
•The frequency of torsional vibration of a single mass
will be
Where,
• q is the stiffness in newton-metres per radian, and
•I is the moment of inertia of the attached mass in kg
metres
2
.
65

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•For a transverse or axial vibration
Where,
• s is the stiffness in newtons per metre of deflection and
m is the mass attached in kg.
•The essence of control is to adjust these two
parameters, q and I (ors and m), to achieve a frequency
which does not coincide with any of the forcing
frequencies.
66
FUNDAMENTAL OR NATURAL FREQUENCY

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration frequency
•The equation is true only when the majority of the vibration
occurs at one frequency.
•In reality machines vibrate in a much more complex way with
vibration occurring at several frequencies.
•By analysis of the frequency at which each of the vibrations are
occurring it is possible to ascertain whether they are being
generated from within the system or externally.
•By further analysis it is possible to locate the source of
vibration within complex machinery.
67

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration spectrum
68
Vibration phase can be defined as the angular relationship between the
positions of maximum vibrations and some fixed point on a rotating shaft at
any instant.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration effects and types
69

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VIBRATION PARAMETERS
•It is important to understand that with sinusoidal vibration, the
relationship between acceleration, velocity and displacement is
fixed and frequency dependent.
•It is not possible to vary any one of these three parameters without
affecting another, and for this reason, one must consider all of them
simultaneously when specifying or observing sine vibration.
•The three parameters of acceleration, velocity and displacement are
all linear scalar quantities and in that respect, at any given
frequency, each has a constant, proportional relationship with the
other
70

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•In other words, if the frequency is held constant,
increasing or decreasing the amplitude of any one of
the three parameters results in a corresponding
proportional increase or decrease in both of the other
two parameters.
•However, the constant of proportionality between the
three parameters is frequency dependent and
therefore not the same at different frequencies.
71
VIBRATION PARAMETERS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Acceleration, Velocity and Displacement
•In general, sinusoidal vibration testing uses the following
conventions for measurement of vibration levels.
•Acceleration is normally specified and measured in its peak
sinusoidal value and is normally expressed in standardized
and normalized dimensionless units of g’s peak. In fact, a g
is numerically equal to the acceleration of gravity under
standard conditions, however, most engineering
calculations utilize the dimensionless unit of g’s and
convert to normal dimensioned units only when required.
•
72

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Velocity is specified in peak amplitude as well. Although not often used in vibration
testing applications, velocity is of primary concern to those interested in machinery
condition monitoring. The normal units of velocity are inches per second in the
English system or millimeters per second in the metric system of units.
•Displacement is usually expressed in normal linear dimensions, however, it is
measured over the total vibration excursion or peak to peak amplitude. The normal
units of displacement are inches for English or millimeters for the metric system of
units.
•The second is that velocity has a proportionally increasing (or decreasing)
relationship with either displacement or acceleration. In other words, the velocity
will increase (or decrease) in direct proportion to the frequency if either of the other
parameters are held constant. Velocity is of interest when damping components or
back EMF issues are important to the testing.
73
Acceleration, Velocity and Displacement

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•As mentioned previously, these quantities are not independent and
are related to each other by the frequency of the vibration. Knowing
any one of the three parameter levels, along with the frequency of
operation, is enough to completely predict the other two levels. The
sinusoidal equations of motion stated in normal vibration testing units
are as follows:
•where:
g= acceleration, g’s peak
D= displacement, inches, peak to peak
V= velocity, inches per second, peak
f = frequency, Hz
74
Acceleration, Velocity and Displacement

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration characteristics
75

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
ACCELERATION MEASUREMENT
76

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration measurements units
•There are three different ways of expressing vibration
measurements
•1. Peak to peak
•2. Half peak
•3. Root mean Square
•They are related as follows.
•R.M.S. = Peak to Peak / 2.83
•Half peak value = Peak to peak / 2 .
77

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VIBRATION MEASUREMENT
•Standard measurement groups include:-
• FFT,
•order tracking,
•octave,
•swept-sine,
•correlation,
•time capture, and
•time/histogram
78

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Averaging
•A wide selection of averaging techniques to improve your signal-to-noise
ratio. RMS averaging reduces signal fluctuations, while vector averaging
minimizes noise from synchronous signals. Peak hold averaging is also
available. Both linear and exponential averaging are provided for each mode
•Order Tracking
•Order tracking is used to evaluate the behavior of rotating machinery.
Measurement data is displayed as a function of multiples of the shaft
frequency (orders), rather than absolute frequency. Combined with a waterfall
plot, or "order map" of your data as a function of time or rpm. Using the slice
feature, the amplitude profile of specific orders in the map can be analyzed. In
tracked order mode, the intensity of individual orders vs. rpm is measured
79
VIBRATION MEASUREMENT

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Octave Analysis
•Real-time 1/1, 1/3 and 1/12 octave analysis, at frequencies up to
40 kHz (single channel) or 20 kHz (two channel). Octave
analysis is fully compliant with ANSI S1.11-1986 (Order 3, type
1-D) and IEC 225-1966. Switchable analog A-weighting filters,
as well as A, B and C weighting math functions, are included.
•Averaging choices include exponential time averaging, linear
time averaging, peak hold, and equal confidence averaging.
Broadband sound level is measured and displayed as the last
band in the octave graph.
80
VIBRATION MEASUREMENT

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Swept-Sine Measurements
•Swept-sine mode is ideal for signal analysis that involves high dynamic range or
wide frequency spans. Gain is optimized at each point in the measurement,
producing up to 145 dB of dynamic range. A frequency resolution of up to 2000
points is also provided. Auto-ranging can be used with source auto-leveling to
maintain a constant input or output level at the device under test (to test response
at a specific amplitude, for instance).
•Time/Histogram
•The time/histogram measurement group is used to analyze time-domain data. A
histogram of the time data vs. signal amplitude is provided for accurate time
domain signal characterization. Statistical analysis capabilities include both
probability density function (PDF) and cumulative density function (CDF). The
sample rate, number of samples, and number of bins can all be adjusted
81
VIBRATION MEASUREMENT

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Time Capture
•Analog waveforms can be captured at sampling rates of 262
kHz or any binary sub-multiple, allowing to optimize sampling
rate and storage for any application. For example, 8 Msamples
of memory will capture 32 seconds of time domain data at the
maximum 262 kHz sample rate, or about 9 hours of data at a
256 Hz sample rate. Once captured, any portion of the signal
can be played back. The convenient Auto-Pan feature lets you
display measurement results synchronously with the
corresponding portion of the capture buffer to identify important
features
82
VIBRATION MEASUREMENT

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Waterfall
•Waterfall plots are a convenient way of viewing a time history of data. Each
successive measurement record is plotted along the z-axis making it easy to
see trends in the data. All FFT, octave and order tracking measurements can
be stored in waterfall buffer memory.
• Waterfall traces can be stored every n time records for FFT and order
tracking measurements. For order tracking measurements, new records can
be acquired at a specific time interval or change in rpm. In octave
measurements, the storage interval is in seconds (as fast as every 4 ms).
While displaying waterfall plots, you can adjust the skew angle to reveal
important features, or change the baseline threshold to eliminate low-level
clutter. Any z-axis slice or x-axis record can be saved to disk or displayed
separately for analysis.
83
VIBRATION MEASUREMENT

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Order Tracking Waterfall
84
click to enlarge click to enlarge
VIBRATION MEASUREMENT PLOTS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
FACTORS AFFECTING VIBRATION PARAMETERS
•The four elements of importance in ship vibration are:
•• Excitation,
•• Stiffness,
•• Frequency Ratio, and
•• Damping
85

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
REDUCING VIBRATION PARAMETERS
•i) Reduce exciting force amplitude, F. In propeller-induced ship
vibration, the excitation may be reduced by changing the propeller
unsteady hydrodynamics. This may involve lines or clearance
changes to reduce the non-uniformity of the wake inflow or may
involve geometric changes to the propeller itself.
•ii) Increase stiffness, K. Stiffness is defined as spring force per
unit deflection. In general, stiffness is to be increased rather than
decreased when variations in natural frequency are to be
accomplished by variations in stiffness. It is not a recommended
practice to reduce system stiffness in attempts to reduce vibration.
No. 86

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•iii) Avoid values of frequency ratio near unity; ω/ωn = 1 is
the resonant condition. At resonance, the excitation is
opposed only by damping. Note that ω/ω
n can be varied
by varying either excitation frequency ω or natural
frequency ω
n. The spectrum of ω can be changed by
changing the RPM of a relevant rotating machinery source,
or, in the case of propeller-induced vibration, by changing
the propeller RPM or its number of blades, ω
n is changed
by changes in system mass and/or stiffness; increasing
stiffness is the usual and preferred approach.
No. 87
REDUCING VIBRATION PARAMETERS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•iv) Increase damping, ζ. Damping of structural
systems in general, and of ships in particular, is small;
ζ << 1. Therefore, except very near resonance, the
vibratory amplitude is approximately damping
independent. Furthermore, damping is difficult to
increase significantly in systems such as ships; ζ is,
in general, the least effective of the four parameters
available to the designer for implementing changes in
ship vibratory characteristics.
No. 88
REDUCING VIBRATION PARAMETERS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration mitigation on a naval vessel
•It is important to prevent or counter vibrations as it ultimately becomes destructive. ‘
Important’ becomes ‘imperative’ on naval and other vessels that, for operational
reasons ,must operate silently.
•Much vibration can be avoided by careful design and manufacture ,for example
ensuring that rotating masses on machines are balanced.
•Otherwise ,it can be dealt with by isolating the responsible machinery from its base
support structure ,interrupting the path which vibration is transmitted from its
source,
•Impressing a counter vibration on the source such that the unwanted vibration is
cancelled , or by controlling the response of ship’s structure by imposing a counter
vibration on it.
•Instead of attempting to isolate structure from source vibration ,the response of the
structure to that vibration could be sensed and actively controlled using strategically
placed transducers and cancellation techniques
89

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Strategies of naval vibration minimization
•Naval designers use finite element and impedance
modeling to help designers achieve structures whose
natural vibration frequencies are well separated from
the excitation frequencies likely to be generated by
the ship’s propulsion and other machinery.
•Failure to ensure this can lead to resonances in which
structural oscillations ,fed with energy from in phase
source of vibrations, may grow to damaging
proportions.
90

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•If , despite these precautions ,vibration still develops
,it may be necessary to address the problem by
judiciously adding balance weights, and a number of
firms specialize in static and dynamic balancing.
•Classic ‘fix' now is to mount them to their base via
anti vibration mounts incorporating resilient material.
Natural rubber ,once the material of choice for
absorbing vibration energy ,has since been joined by
a range of elastomer and other materials. This is
called a passive machinery mount/vibration isolator
91
Strategies of naval vibration minimization

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Passive vibration isolator
•A passive machinery mount/vibration isolator has two roles.;
•to support and secure the machine in question ( propulsion engine, generator ,pump
etc.) and
•to isolate vibration source (machine) from the receiver ( base structure)
•The former requires the mount to be stiff as possible , while for the latter it needs to
be highly resilient.– indeed, for maximum effect, as soft as possible.
•This is a contradictory requirement, so passive dampers tend to be a compromise
and are least effective at low frequencies.
•Nevertheless ,passive mounts can substantially reduce medium or high frequency
vibrations.-typically by some 10dB. Most solid mounts comprise metal/elastomer
combinations in which metal most of the stiffness required to support and locate the
subject machine while the elastomer absorbs the vibration energy.
92

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Type of passive vibration isolators
No. 93

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Tuned vibration absorbers (TVAs)
•Whilst passive dampers based on resilient materials
are often cost effective ,their performance can
sometimes be bettered by alternative source of
resilience.
•Thus, engineers have devised fluidic/hydraulic,
pneumatic and electro-magnetic systems. Also
available are tuned vibration absorbers (TVAs) ,the
performance of which is optimized for particular
frequencies.
94

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•For low speed engines ,electro-mechanical vibration
compensators can be more effective. These utilize
rotating masses electrically driven in synchronism
with mechanical out of balance forces, but anti phase
to produce a canceling effect.
•A signal representing engine rev/min sensed by a
tacho-graph, is fed to a synchronizer module which
ensures correct drive speed and phasing.
95
Tuned vibration absorbers (TVAs)

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Electrically operated tuned vibration
absorber
96

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Location of the TVA
•Units can be selected to counter vibration propagated
transversely, longitudinally or vertically.
•Typically a pair would be mounted at specific points
near main engine ,to feed counter vibrations into the
base structure.
•The system can also be beneficial for certain propeller
vibration modes ,and in those cases it is usually
mounted at the thrust block.
97

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Active vibration isolator
•Another approach shown to be effective with slow-turning
engines is to supplement passive isolation with the
element of active cancellation.
•In essence ,the active means which is somewhat
analogous to the now well-known ‘anti noise’ cancellation
technique. -- impose vibration on a machinery mount in
such away that it cancels the unwanted source vibration .
As a result ,vibration passing to the base structure is
minimized.
98

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•System requires an electro-mechanical actuator to impose the
anti-vibration oscillations, accelerometer sensors to sense the
source of vibration ,a central data acquisition
/processor/control unit and a power supply.
•Sensor signals are digitized and analyzed by the processor
,which then generates an appropriate digital cancellation signal.
This is converted back to analogue and passed via a power
amplifier to the actuator.
•The control loop can be closed by a feed back or feed forward (
anticipatory) term depending on the situation. A tachograph
fitted to the subject machine provides the necessary
revolutions rate reference signal.
99
Active vibration isolator system

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•An active control element can be included as a part of the isolator itself ,in
which case it is in series (in line) with the passive resilient element, or it can
be mounted alongside it so that the active and passive elements are in
parallel.
•The latter arrangement has the advantage that the active element does not
have to bear the source machine’s weight ,but the inertial actuator has to be
powerful enough to overcome the stiffness of the passive mount.
•State-of –the-art active control systems can provide as much as 20-30 dB
attenuation, including at the low frequencies that defeat passive isolators.
•Smart Spring mount fail to a safe condition since, should the active
component or power fail ,they can still operate passively.
100
Active vibration isolator control element

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
101
Active vibration isolator (parallel config)

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Concept Design Approach
•Concept design is where the vibration avoidance process
must begin. It is clear that if the vibration problems,
repeatedly identified by experience as the most important, are
addressed at the earliest design stage, ultimately serious
problems, involving great cost in correction efforts, may be
avoided.
•The focus is on planning for vibration early at the Concept
Design stage, where there has been no development of details.
If as much as possible can be done in concept design with the
simple tools and rules of thumb available at that level, it will
help to avoid major vibration problems.
102

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•While quantification of all four elements is required in
calculating the vibration response level, acceptable results may
consistently be achieved with reasonable effort by focusing
attention in concept design on two of the four elements. The
two of the four elements of importance are excitation and
frequency ratio.
•While quantification of all four elements is required in
calculating the vibration response level, acceptable results may
consistently be achieved with reasonable effort by focusing
attention in concept design on two of the four elements.
No. 103
Concept Design Approach

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The two of the four elements (excitation, stiffness,
frequency ratio, damping) of importance are excitation
and frequency ratio. The achievement in design of
two objectives with regard to these elements has
resulted in many successful ships:
•• Minimize dominant vibratory excitations, within the normal
constraints imposed by other design variables, and
•• Avoid resonances involving active participation of major
subsystems in frequency ranges where the dominant excitations
are strongest.
No. 104
Concept Design Approach

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Attention to vibration design of ships
•Experience has shown that attention to vibration in concept design of large
ships can usually be paid to the following items:
•i) Hull girder vertical vibration excited by the main engine.
•ii) Main machinery/shafting system longitudinal vibration excited by the
propeller.
•iii) Superstructure fore-and-aft vibration excited by hull girder vertical
vibration and/or main propulsion machinery/shafting system longitudinal
vibration.
•A myriad of local vibrations, such as hand-rails, antennas, plating panels,
etc., may be encountered on new vessel trials in addition to these three.
•But local problems usually involve local structural resonances and often
considered as minor problems, as the correction approach by local
stiffening may be easily achievable
No. 105

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
DESIGNING OUT VIBRATION
No. 106

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Effect of engine vibrations
No. 107

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Sources of vibration excitation
•There are a number of sources of vibration and noise present
in a ship or marine vehicle. Typically these may include:
1.• The prime movers - typically diesel engines.
2.• Shaft-line dynamics
3.• Propeller radiated pressures and bearing forces.
4.• Air conditioning systems.
5.• Maneuvering devices such as transverse propulsion units
6.• Cargo handling and mooring machinery.
7.• Vortex shedding mechanisms
8.• Intakes and exhausts.
9.• Slamming phenomena.
108
The major sources are the low-speed diesel
main engine and the propeller. Gas turbines are
generally considered to give less excitation
than diesel engines.

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Other causes of vibration form prime movers
and auxiliaries
•Typical causes could be
•ŠUnbalance,
•Misalignment,
•Damaged or worn bearings,
•Damaged or worn teeth
•Resonance,
•loose components
•Bending or eccentricity of shafts,
•Electromagnetic effects,
•Unequal thermal effects
•Aerodynamic forces (turbocharger)
•Hydraulic forces
•Bad belt drives
•Oil whirl
•Reciprocating forces.
•The great majority of the above create a vibration at a multiple of the fundamental
No. 109

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Excitation due to slow speed diesel engine
•Excitations generated by the engine can be divided into two
categories:
•1) Primary excitations, which are forces and moments originating
from the combustion pressure and the inertia forces of the rotating
and reciprocating masses. These are characteristics of the engine
as such, and they can be calculated in advance and be stated as
part of the engine specification, with reference to a certain speed
and power
•2) Secondary excitations, stemming from a forced vibratory
response in a sub-structure. The vibration characteristics of sub-
structures are almost independent of the remaining ship structure
No. 110

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vibration Aspects of Two-stroke Diesel Engines
•The vibration characteristics of the two-
stroke low speed diesel engines can for
practical purposes be, split up into four
categories:-
•• External unbalanced moments :These can
be classified as unbalanced 1
st
, 2
nd
and may
be 4
th
order external moments, which need to
be considered only for certain cylinder
numbers
•• Guide force moments
•• Axial vibrations in the shaft system
•• Torsional vibrations in the shaft system.
No. 111

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
External unbalanced moments and guide
forces
No. 112

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Diesel engine forces
•Diesel engine force components are
comprised of static loads (I.e. loads
arising from bolted assembly) and
dynamic loads.
•The dynamic loads are due to forces
arising from two sources.
1.The fluctuating gas pressure in the
cylinder.
2.The inertia forces.
3.Rotating masses
No. 113

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
External unbalanced moments
•The inertia forces originating from the unbalanced rotating and reciprocating
masses of the engine create unbalanced external moments although the external
forces are zero.
•This can be Mathematically, expressed as follows:-
•Of these moments, only the 1st order (one cycle per revolution) and the 2nd order
(two cycles per revolution) need to be considered, and then only for engines with a
low number of cylinders.
•On some large bore engines the 4th external order moment may also have to be
examined.
•The inertia forces on engines with more than 6 cylinders tend, more or less, to
neutralize themselves.
No. 114

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
1
st
and 2
nd
order external moments
•The external moments are known as the 1
st
,order moments (acting
in both the vertical and horizontal directions) and 2
nd
order moments
(acting in the vertical direction only, because they originate solely
in the inertia forces on the reciprocating masses.
•The 1
st
order moments acts with a frequency corresponding to the
engine speed x 1.
•Generally speaking, the 1
st
order moment causes no vibration
problems. For 4-cylinder engines, however, it is recommendable to
evaluate the risk because in rare cases this cylinder configuration
may cause vibration levels of a disturbing magnitude
No. 115

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
1
st
order resonance and solutions
•Resonance with a 1st order moment may occur for hull vibrations with 2 and/or 3
nodes.
•This resonance can be calculated with reasonable accuracy, and the calculation for
the specific plant will show whether or not a compensator is necessary.
•In rare cases, where the 1st order moment may cause resonance with both the
vertical and the horizontal hull vibration mode in the normal speed range of the
engine,
•The adjustable counter-weights should be positioned so as to make the vertical
moment harmless, and a 1st order compensator fitted in the chain tightener wheel
in order to neutralize the horizontal moment.
•With a 1st order moment compensator fitted aft, the horizontal moment will
decrease to between 0 and 30% of the value, depending on the position of the node.
The 1st order vertical moment will decrease to about 30% of the value.
No. 116

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
1
st
order vertical & horizontal moment
compensator
No. 117

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
2
nd
order resonance
•The 2nd order moment acts with a frequency corresponding
to twice the engine speed. The 2nd order moment acts in the
vertical direction only.
•Owing to the magnitude of the 2
nd
order moment, it is only
relevant to compensate this moment on 4, 5 and 8-cylinder
engines,
•Resonance with 4 and 5 node vertical hull girder vibration
modes can occur in the normal engine speed range.
•In order to control the resulting vibratory responses, a 2nd
order compensator can be installed
No. 118

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Solutions for 2
nd
order compensator
•Several solutions, from which the most cost-efficient one can be chosen, are available to cope
with the 2
nd
order vertical moment:
•a) No compensators, if considered unnecessary on the basis of the natural frequency, nodal
point and size of the 2nd order moment
•b) A compensator mounted on the aft end of the engine driven by the main chain drive,
•c) A compensator mounted on the front end, driven from the crank shaft through a separate
chain drive
•d) Compensators on both the aft and fore ends of the engine, completely eliminating the
external 2nd order moments,
•e) An electrically driven compensator, synchronized to the correct phase relative to the free
moment.
• This type of compensator needs preparations in the form of an extra seating, prefer-able in
the steering gear room, where deflections are largest and the compensator, therefore, will
have the greatest effect,
No. 119

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
2
nd
order moment compensators
No. 120

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
No. 121
2
nd
order moment compensators

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Action of 2
nd
order compensator
•Compensation of an external moment by means of a compensating
force is possible if there is an adequate distance from the position
where the force is acting to the node of the vibration (i.e. an
excitation force is inefficient when acting in a node).
•The counterweights on the chain wheel produce a centrifugal force
which creates a moment, the size of which is found by multiplying
the force by the distance to the node.
•Due to the positioning of these counter-weights, the direction of the
compensating moment will always be opposite to the direction of the
external moment
No. 122

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Guide force moments
•The so-called guide force moments are caused by the gas
force on the piston, and by inertia forces.
•When the piston is not exactly in its top or bottom
position, the gas force, transferred through the connecting
rod, will have a component acting on the crank-shaft
perpendicular to the axis of the cylinder. Its resultant is
acting on the guide shoe and, together, they form a guide
force moment,
•In a multi-cylinder engine, gas and inertia forces and their
resultants form a system of guide force moments
containing all orders.
No. 123

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
TYPES OF GUIDE FORCES.
•Two kinds of guide force moments exist:
•The so-called H and X-moments.
•The H-type guide force moment, which is dominating on
engines with less than seven cylinders, tends to rock the
engine top in the transverse direction. The main order of
the H-moment is equal to the cylinder number.
•The X-type guide force moment is the dominating for
engines with more than six cylinders, The X-moment tends
to twist the engine in an X-like shape, and the main order
is equal to half the number.
No. 124

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Guide forces and top bracings
•For engines with odd numbers of cylinders, the main
orders are mostly the two orders closest to half the
number of cylinders.
•In order to counteract the possible impact on the hull from
guide force moments, it is recommend the installation of a
set of top bracings between the upper gallery of the main
engine and the hull structure (casing side).
•The top bracing can either be mechanical with frictional
connection or hydraulically adjustable
No. 125

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
H-type Guide Force Moment (M
H)
•Each cylinder unit produces a force couple consisting of:
•1: A force at level of crankshaft centre-line.
•2: Another force at level of the guide plane.
•The position of the force changes over one revolution, as
the guide shoe reciprocates on the guide plane. As the
deflection shape for the H-type is equal for each cylinder
the Nth order
•H-type guide force moment for an N-cylinder engine with
regular firing order is: N • MH(one cylinder).
No. 126

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The size of the forces in the force couple is:
•Force = MH /L kN
•where L is the distance between crankshaft level and the middle
position of the guide plane (i.e. the length of the connecting rod).
•As the interaction between engine and hull is at the engine seating
and the top bracing positions, this force couple may alternatively be
applied in those positions with a vertical distance of (LZ).
•Then the force can be calculated as:
•ForceZ =MH /LZ kN
No. 127
H-type Guide Force Moment (M
H)

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
X-type Guide Force Moment (M
X)
•The X-type guide force moment is calculated based on the same
force couple as described, however, as the deflection shape is
twisting the engine each cylinder unit does not contribute with equal
amount.
•The centre units do not contribute very much whereas the units at
each end contributes much.
•A so-called ”Bi-moment” can be calculated
•The X-type guide force moment is then defined as:
No. 128

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Alignment of guide forces
No. 129

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Action of the bracings
•These bracings act as detunes of the system double
bottom and main engine, which means that the natural
frequency of the vibration system will be increased to
such an extent that resonance occurs above the running
range of engine speed and the guide force moments will,
therefore, be harmless.
•The mechanical top bracing comprises stiff connections
(links) with friction plates and alternatively a hydraulic top
bracing, which allow adjustment to the loading conditions
of the ship.
No. 130

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Hydraulic and mechanical top bracings
No. 131

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Internal forces and moments
•It is the responsibility of the engine de-signer to provide the
engine frame with sufficient stiffness to cope with the
internal forces and moments so that de-flections and
corresponding stresses can be kept within acceptable limits.
•If the engine frame could be assumed to be infinitely stiff,
internal moments and forces would not be able to give
excitations to the ship’s structure. How-ever, it is obvious
that an infinitely stiff engine frame cannot be obtained and,
•therefore, it is the relative stiffness between the engine
frame and the connected hull structure which has to be
considered
No. 132

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
No. 133
Internal & external forces and moments

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Secondary forces
•These excitation forces are not generated within the
engine, but are resultant of the interaction effect of the
prime mover the propeller and the shafting system.
•Performance of the system also depends on the response
of the system to the oscillations imposed on to the system.
•The vibration characteristics may be modified by the
impressed effect on it by the remaining system.
•There are two main types of these vibrations:-
•Axial vibrations and
•torsional vibrations.
No. 134

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Torsional vibrations
•The varying gas pressure in the cylinders
during the working cycle and the
crankshaft/connecting rod mechanism
create a varying torque in the crank-shaft. It
is these variations that cause the excitation
of torsional vibration of the shaft system.
•Torsional vibration causes extra stresses,
which may be detrimental to the shaft
system.
•The stresses will show peak values at
resonances, i.e. where the number of
revolutions multiplied by the order of
excitation corresponds to the natural
frequency
No. 135

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Power Related Unbalance (PRU)
•To evaluate if there is a risk that 1
st
and 2
nd
order external moments will excite
disturbing hull vibrations, the concept Power Related Unbalance can be used as a
guidance.
•With the PRU-value, stating the external moment relative to the engine power, it is
possible to give an estimate of the risk of hull vibrations for a specific engine.
Based on service experience from a greater number of large ships with engines of
different types and cylinder numbers, the PRU-values have been classified in four
groups as follows:-
•from 0 to 60 . . . . . . . . . . . . . . . . . . . . . not relevant
•from 60 to 120 . . . . . . . . . . . . . . . . . . . . . . unlikely
•from 120 to 220 . . . . . . . . . . . . . . . . . . . . . . . likely
•above 220 . . . . . . . . . . . . . . . . . . . . . . . most likely
No. 137

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Power unbalance and torsional vibration
•Designers tend to rely on reasonably correct balance among cylinders.
•It is important to realize that an engine with one cylinder cut out for any
reason, or one with a serious imbalance between cylinder loads or timings,
may inadvertently be aggravating a summation of vectors which the designer,
expecting it to be small, had allowed to remain near the running speed range.
•In general any kind of irregularity in the cylinder firings produces and enlarged
vibratory stresses in the components of the propulsion plant.
•The absence of firing of one unit significantly changes the whole picture of
the propulsion plant vibration behavior.
•Misfiring in any one cylinder gives rise to resonances that are small or even
negligible during the normal operation of plant
No. 138

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Coupling of variable torque, angular
momentum and variable propeller thrust
•Torsional vibrations of the entire shaft system are mainly excited by
the tangential force T.
•Torsional vibration can, excite vibration in the hull through the
coupling phenomena present in the connecting rod mechanism and in
the propeller.
•Torsional vibration induced moments and forces due to connecting
rod mechanism . If a harmonic angular velocity is superimposed upon
the normal uniform rotation of the crank-throw, as in the case of
torsional vibrations, this will cause harmonic forces and moments to
occur.
•However, due to the connecting rod mechanism, the reaction forces
will not solely be of the same order as the super-imposed torsional
vibration, but significant orders of n-2, n-l, n+l and n+2 will also
appear.
No. 139

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Torsional vibration analysis
•The reciprocating and rotating masses of the engine
including the crankshaft, the thrust shaft, the intermediate
shaft(s), the propeller shaft and the propeller are for
calculation purposes considered as a system of rotating
masses (inertias) interconnected by torsional springs.
•The gas pressure of the engine acts through the
connecting rod mechanism with a varying torque on each
crank throw, exciting torsional vibration in the system with
different frequencies
No. 140

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Potentially the most damaging form of vibration is the
torsional mode, affecting the crankshaft and propeller
shafting (or generator shafting).
•There will be as many ‘modes’ in which the shaft can be
induced to vibrate naturally as there are shaft elements.
•A node is found where the deflection is zero and the
amplitude changes sign. The more nodes that are
present, the higher the corresponding natural frequency.
No. 141
Torsional vibration analysis

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•In general, only torsional vibrations with one and two
nodes need to be considered.
•The main critical order, causing the largest extra stresses
in the shaft line, is normally the vibration with order equal
to the number of cylinders.
•This resonance is positioned at the engine speed
corresponding to the natural torsional frequency divided by
the number of cylinders.
•The torsional vibration conditions may, for certain
installations require a torsional vibration damper.
No. 142
Torsional vibration solutions

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
MULTINODE SHAFTING OSCILLATION
•In the one-node case, when the masses forward of the
node swing clockwise, those aft of it swing anti-
clockwise and vice versa.
•In the two-node case, when those masses forward of the
first node swing clockwise, so do those aft of the second
node, while those between the two nodes swing anti-
clockwise, and vice versa.
•The problem arises when the forcing frequencies of the
externally applied, or input, vibration coincide with, or
approach closely, one of these natural frequencies.
No. 143

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Based on statistics, this need may arise for the following types of
installation:
•• Plants with controllable pitch propeller
•• Plants with unusual shafting layout and for special owner/yard requirements
•• Plants with 8, 11 or 12-cylinder engines
•The so-called QPT (Quick Passage of a barred speed range Technique), is
an alternative option to a torsional vibration damper, on a plant equipped
with a controllable pitch propeller. The QPT could be implemented in the
governor in order to limit the vibratory stresses during the passage of the
barred speed range.
•The application of the QPT has to be decided by the engine maker
No. 145
Torsional vibration solutions

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Factors affecting amplitude and frequency
•The most significant masses in any mode of
vibration are those with the greatest amplitude
on the corresponding elastic curve. Changing
them would have the greatest effect on
frequency.
•The most vulnerable shaft sections are those
whose combination of torque and diameter
induce in them the greatest stress.
•The most significant shaft sections are those
with the steepest change of amplitude on the
elastic curve and therefore the highest torque.
•These are usually near the nodes but this
depends on the relative shaft diameter.
•Changing the diameter of such a section of shaft
will also have a greater effect on the frequency.
No. 146

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Effect of running the engine at critical RPM
•If an engine were run at or near a major critical speed it would sound rough
because, at mid-stroke, the torsional oscillation of the cranks with the biggest
amplitude would cause a longitudinal vibration of the connecting rod.
•This would set up in turn a lateral vibration of the piston and hence of the
entablature.
•Gearing, if on a shaft section with a high amplitude, would also probably be
distinctly noisy.
•It is usually difficult, and sometimes impossible, to control all the possible critical
speed, so that in a variable speed propulsion engine it is sometimes necessary to
‘bar’ a range of speeds where vibration is considered too dangerous for
continuous operation.
•Torsional vibrations can sometimes affect camshafts also
No. 148

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Control of stresses during the resonance
•Designers can nowadays adjust the frequency of
resonance, the forcing impulses and the resultant
stresses by adjusting shaft sizes, number of propeller
blades, crankshaft balance weights and firing orders,
•By using viscous or other dampers, detuning couplings
and so on.
•Gearing, of course, creates further complications—and
possibilities. Branched systems, involving twin input or
multiple output gearboxes, introduce complications in
solving them; but the principles remain the same.
No. 149

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The classification societies prescribe two limits, Ʈ
1, and Ʈ
2, for the
torsional stress in the speed range up to 80 per cent of MCR :
•The lower Ʈ
1,:
•Determines a stress level which may only be exceeded for a short time, i.e.
not during continuous running, which means that the propulsion plant
requires a barred speed range of revolutions.
•The upper limit Ʈ
2 :
•May not be exceeded at all,
•At engine speeds where the lower limit Ʈ
1, is exceeded, it will be necessary
to introduce a “barred speed range” in which continuous operation is
prohibited.
No. 151
Stress limits and barred speed range

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Four, five and six-cylinder engines, require special attention. On
account of the heavy excitation, the natural frequency of the system
with one-node vibration should be situated away from the normal
operating speed range, to avoid its effect.
•This can be achieved by changing the masses and/or the stiffness of
the system so as to give a much higher, or much lower, natural
frequency, called under-critical or overcritical running, respectively.
•Owing to the very large variety of possible shafting arrangements that
may be used in combination with a specific engine, only detailed
torsional vibration calculations of the specific plant can determine
whether or not a torsional vibration damper is necessary.
No. 152
Torsional vibration calculations

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
TORSIONAL VIBRATION DAMPERS
•There two type of dampers for torsional vibrations:-
•The spring mass type
•The viscous type.
•Torsional dampers are placed behind the engine as
vibrational dampers when the powertrain does not include a
separating and starting clutch.
•The purpose of using a torsional damper is to keep engine
torque peaks as well as operational irregularities away from
the powertrain and connected units.
No. 153

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•A torsional damper ensures “peace and quiet”
operation.
•If the forces operating in the powertrain area were not
countered, the powertrain components would also
show considerably higher levels of wear.
•A standard solution today for decoupling torsional
vibrations in powertrains is to use a bolt-on torsional
damper that builds on the technology in clutch discs
with torsional damping.
No. 154
TORSIONAL VIBRATION DAMPERS

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
The spring mass type
•The torsional damper consists of a set of coil springs
positioned in windows that allow a limited amount of rotary
movement between the crankshaft and the transmission
input shaft and a friction device.
•By selecting the right torsional damper size and spring set,
characteristic curves can be adjusted to meet the individual
needs of specific applications.
•Vibrational decoupling can therefore be adapted in
optimum fashion, and ignition-related rotational
irregularities can be reduced.
No. 156

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VISCOUS TYPE DETUNERS
No. 157
torsional damper Dampers De tuners: Reducing Vibration of Marine Engines

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
VISCOUS TYPE DETUNERS
•The torsional damper is integrated into the respective
installation space by a simple adjustment of the external
bolt-on area and by selecting the corresponding spline
profile to match the drive shaft.
•The most famous type of torsional damper used on marine
engine of a ship is Viscous type dampers, which consist of
an inertia ring added to the crankshaft enclosed in a thin
layer of highly viscous fluid like silicon.
•The inertia ring is free to rotate and applies a lagging torque
on the crankshaft due to its lagging torsional motion.
No. 158

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Lumped mass system
•The continuous shafting system needs to be divided in the
so-called lumped mass system where, after applying
equations of motion, one evaluates natural frequencies,
accompanied mode shapes and, in the case of forced
torsional vibrations, angular displacements of all masses.
•After that, it is straightforward to determine vibration
torques and stresses.
•Equations of motion of the lumped mass system could
be gathered in a common matrix equation:
No. 160

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Lumped mass equation
•where J is the diagonal inertia matrix, C is the symmetric
damping matrix, K is the symmetric stiffness matrix, and
θ, θ and θ are the angular acceleration, velocity and
displacement vectors, respectively. On the right hand
side, f denotes the applied load, expressed with vibration
excitation vector.
•Forced damped torsional vibration response could be
obtained in various ways. By assuming harmonic
excitation:
No. 161

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•and harmonic response in the form:
•where F is the complex excitation torque amplitude, Ω is
the excitation frequency, t is the time, and Θ is the
complex angular displacement amplitude, the system of
equations readily transforms into a system of linear
algebraic equations with complex coefficients:
•Vibration torque amplitudes between the adjacent
masses could then be obtained from:
No. 162
Lumped mass equation

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Estimation of vibration stresses
•where k
t is the shaft stiffness, (θi+1– θi) is the
amplitude of the shaft element twist, and m is the
number of shaft elements. Afterwards, the vibration
stresses could be easily determined from
•where d is the shaft element diameter.
No. 163

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Axial vibrations
•When the crank throw is loaded by the gas pressure
through the connecting rod mechanism, the arms of the
crank throw deflect in the axial direction of the
crankshaft, exciting axial vibrations.
•Through the thrust bearing, the system is connected to
the ship`s hull. Generally, only zero-node axial vibrations
are of interest. Thus the effect of the additional bending
stresses in the crankshaft and possible vibrations of the
ship`s structure due to the reaction force in the thrust
bearing are to be considered.
No. 164

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•An axial damper is fitted as standard to all MC engines minimizing
the effects of the axial vibrations.
•For an extremely long shaft line in certain large size container
vessels, a second axial vibration damper positioned on the
intermediate shaft, designed to control the on-node axial vibrations
can be applied.
•Alternating thrust, the excitation for longitudinal vibration of the
shafting/main machinery system, occurs at blade rate frequency
(Propeller RPM × Blade number N) and its multiples.
•The fundamental is usually much larger than any of its harmonics,
however. Alternating thrust is produced by the blade number
circumferential harmonic of the hull wake.
No. 165
Axial vibrations

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Axial vibration frequency and mode
•When the crank-throw is loaded by the gas
force through the connecting rod mechanism,
the arms of the crank throw deflect in the axial
direction of the crank-shaft, exciting axial
vibrations which, through the thrust bearing,
may be transferred to the ship’s hull.
•The dominating order of the axial vibration is
equivalent to the number of cylinders for
engines with less than seven cylinders. For
engines with more than six cylinders, the
dominating order is equal to half the numbers
of cylinders.
•For engines with odd numbers of cylinders, the
dominating orders are mostly the two orders
closest to half the cylinder number.
No. 167

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Axial vibration damper
•These influenced the
vibration behavior of the
crankshaft, the engine
frame, and the
superstructure.
•The axial vibration damper
alone actually eliminates the
problems, and reduces the
vibration level in the deck
house to below the IS0
recommended values.
No. 168

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The Axial damper is fitted on the crankshaft of the engine to dampen the shaft
generated axial vibration i.e. oscillation of the shaft in forward and aft directions,
parallel to the shaft horizontal line.
•It consists of a damping flange integrated to the crankshaft and placed near the
last main bearing girder, inside a cylindrical casing. The casing is filled with
system oil on both side of flanges supplied via small orifice. This oil provides the
damping effect.
•When the crankshaft vibrates axially, the oil in the sides of damping flange
circulates inside the casing through a throttling valve provided from one side of
the flange to the other, which gives a damping effect.
•The casing is provided with high temperature alarm and pressure monitoring
alarms located on both sides of damping flanges. They give alarm if one side oil
pressure drops more than the set value as a result of low LO supply, sealing ring
failure etc.
No. 169
Axial vibration damper

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
AXIAL VIBRATIONS AND DAMPERS
•The torsional deformation causes
changes in the length of the crankshaft
which is seen as axial vibration at the
free end of the crankshaft.
•The torsional vibration also causes the
propeller to rotate with varying speed,
which in turn gives a varying thrust.
•The varying thrust excites the
propulsion shafting axially, which also
causes axial vibration to be seen at the
free end of the crankshaft.
No. 170

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Propeller excitations due to non-uniform
wake field
•Excitations due to the propeller working in the non-uniform
wake field will be transmitted to the hull either through the
shaft system as forces and moments or through the water
as pressure fluctuations acting on the hull surface,
•The forces and moments should also be considered when
calculating the torsional, axial, and lateral vibrations of the
shaft system.
•The excitation can be reduced by modifying wake field and
propeller design,
No. 171

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Axial vibration nodes
•Axial vibrations are longitudinal
shafting vibrations. The mass-
elastic system is used for axial
vibration calculations and the
mode shapes of the two lowest
modes which are of relevance.
• For engines more than 6
cylinders main critical resonance
with O-node vibration mode
below MCR speed.
No. 172
The 1 -node vibration mode is normally
of less importance. Its natural frequency is
determined by the mass and stiffness of the entire
shafting system. Especially the stiffness of the
thrust bearing and its support is very decisive.
Normally, the natural frequency is so
high that no dynamic amplification of
this mode will occur

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Hull Wake
•Hull wake is one of the most critical aspects in avoidance
of unacceptable ship vibration.
•Propeller-induced vibration problems in general start with
unfavorable hull lines in the stern aperture region, as
manifest in the non-uniform wake in which the propeller
must operate. Unfortunately, propeller excitation is far
more difficult to quantify than the excitation from internal
machinery sources.
•This is because of the complexity of the unsteady
hydrodynamics of the propeller operating in the non-
uniform hull wake
No. 173

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Non-uniform hull wake
•In fact, the non-uniform hull wake is the most
complicated part; it is unfortunate that it is also the
most important part.
•Propeller-induced vibration would not be a
consideration in ship design if the propeller disk
inflow were circumferentially uniform. Any treatment
of propeller excitation must begin with a consideration
of the hull wake.
No. 174

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Wake field analysis.
•For engineering simplification, the basic concepts allow for the
circumferential non-uniformity of hull wakes, but assume, for steady
operation, that wake is time invariant in a ship-fixed coordinate
system.
•Nominal wake data from model scale measurements in towing tanks
are presented either as contour plots or as curves of velocity versus
angular position at different radii in the propeller disc.
•The position angle, θ, is taken as positive counterclockwise, looking
forward, and x is positive aft. The axial wake velocity v
X and
tangential wake velocity v
T are dimensionless on ship forward
speed, U.
No. 175

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Nominal wake distribution
•the axial velocity is symmetric in
θ about top-dead-center (even
function) and the tangential
velocity is asymmetric (odd
function).
•This is a characteristic of single
screw ships due to the
transverse symmetry of the hull
relative to the propeller disk;
such symmetry in the wake does
not, of course, exist with twin-
screw ships
No. 176

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Wake asymmetry
•The streamlines are more or less horizontal along the skeg
and into the propeller disk. The flow components along the
steep buttock lines forward of the propeller disk are small.
•The dominant axial velocity field of the resultant wake has a
substantial defect running vertically through the disk along
its vertical centerline, at all radii.
•This defect is the shadow of the skeg immediately forward.
The tangential flow in the propeller disk, being the
combination of the component of the upward flow toward
the free surface and any disk inclination relative to the
baseline, is much smaller.
No. 177

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Propeller cavitation and vibration
•Each of the blades will be lightly loaded in position
90 – 135 deg ( high axial velocity ), while in position
0 degree it will be heavily loaded.
•The area around 0 deg in Fig 1 is called the wake
peak.
•In such a wake peak ,the blade loading will increase
as the blade continues through it and cavitation will
occur at the back of the blade( suction side
cavitation) .
•When the blade moves out of the wake peak ,the
loading will decrease and the cavitation gradually
disappears.
•This variation in cavity volume per unit time makes
the largest contribution to propeller-induced
vibration of the hull.
No. 178

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Propeller Radiated Signatures
•The basis of the development of the propeller induced hull
pressure signature is the acceleration of the cavity
volumes with respect to time on the propeller blades,
modified by the self induced component of pressure
generation arising from the vibration of the ship structure
at the point of interest.
•As such, the hydrodynamic excitation process is a time
domain event whose physical processes can better
understood through the pressure time series.
No. 179

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•There are blade surface pressure changes which vary
from blade to blade in a single revolution and changes
from one revolution to the next.
•These changes are random in nature and result from
the interaction of the temporal changes in the flow;
the flow field, this being the sum of the steady inflow
field and the seaway induced velocities; and the blade
to blade geometric variations due to the
manufacturing tolerances of the propeller blades.
•
No. 180
Propeller Radiated Signatures

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•These changes influence both the general form of the
cavity volume variation and the higher frequencies
and noise generated from the random perturbations of
the topological form of the underlying cavity structure.
•A number of candidate approaches offer themselves
and among these are Short Form Fourier Transforms,
Joint Time-frequency analysis, wavelet techniques
and a double integral analysis of the underlying
pressure signature.
No. 181
Propeller Radiated Signatures

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Highly Skewed Propellers
•According to one manufacturer ,the highly –skewed blade
design offers perhaps the most beneficial type of propeller
by reducing hull vibration and improving fuel consumption
on most types of vessel.
•Highly skewed blades have been used for decades on
fixed pitch propellers, but their application on controllable
pitch propellers has been fairly limited. The highly skewed
design is characterized by a remarkable backward sweep
of the edge in relation to the direction of rotation.
No. 182

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•On a propeller with sufficient skew on the blades ,the
duration of the cavitation will be lengthened compared to
conventional propeller .This reduces the rate of variation
of cavitation with time and therefore vibration.
•A skewed propeller will also reduce the dynamic forces
absorbed through the propeller shaft.
•There has been no full scale measurements but it has been
experienced that a reduction in propeller induced vibration
level of about 50 per cent, where it has been possible to
compare conventional with highly skewed blades.
No. 183
Highly Skewed Propellers

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
Vortex Shedding Mechanisms
•Vibration induced from the flow over structural
discontinuities such as sea chest openings has been a
troublesome feature in some ships and has prevented
the meeting of localized comfort criteria.
•Such vibrations, which commonly manifest themselves
in local structural resonant behavior, are clearly not
directly related to machinery rotational speeds.
•Rather, they are related to vortex shedding over the sea
chest hull opening grills and, therefore, are Strouhal and
Froude number dependent based on ship speed.
No. 184

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•These have included A-brackets, extended centre-line
skegs and fin appendages fitted to ships to improve course
keeping stability.
•The characteristics of these problems were high vibration
levels in the ship structure or failure of the structural
elements.
•Vortex shedding occurs when the fluid flow around the
after part of an appendage is separated from the structure
at a given Reynolds number and the oscillating pressures
cause the elastic structure to vibrate.
No. 185
Vortex Shedding Mechanisms

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The shedding frequency is given in terms of Strouhal
number and for bodies with rough surfaces at ship scale
it is frequently acceptable for estimation purposes to use
a value for the Strouhal number of 0.2.
•When structures vibrate in the transverse direction with a
frequency at or near the vortex shedding frequency they
tend to increase the strength of the shed vorticity which,
in turn, may increase the structural excitation.
No. 186
Vortex Shedding Mechanisms

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Furthermore, if the vortex shedding frequency is close to the
natural frequency of the structure it will move to the frequency of
the structure. Then once the vortex shedding frequency is
synchronized with the frequency of the structure it will often tend
to remain at that frequency even when the flow speed changes
over a limited range.
•The dynamic behavior of structures subjected to vortex shedding
excitation depends upon the ship speed, the structural profile
and its trailing edge shape, the structural natural frequencies and
damping and the interaction between the fluid flow and structural
vibrations.
No. 187
Vortex Shedding Mechanisms

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Reduction of the vibration amplitudes of the structure
caused by vortex shedding may be achieved by:
•• Avoidance of resonance between the vortex-induced excitations
and the structural natural frequency.
•• Lowering the vortex excitation levels.
•• Reducing response of the structure.
•Resonance can be avoided by modifying either the
vortex excitation frequency or the structural natural
frequency.
No. 188
Vortex Shedding Avoidance

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•Ordinarily the structural natural frequency should be
increased sufficiently to avoid resonances with vortex
shedding mechanisms. That may be achieved by increasing
the structure’s stiffness or changing the aspect ratio.
•Other solutions can be to increase the vortex shedding
excitation frequency by changing the structure’s trailing edge
shape. In all cases it is necessary to evaluate the structural
natural frequencies and ensure that they are not coincident
with the vortex shedding and propeller excitations
No. 189
Vortex Shedding Avoidance

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
After-body Slamming
•Shock impacts such as slamming also need consideration since
as well as generating structural tertiary stresses in the ship
structure, these events can be disturbing to passengers.
•In particular after-body slamming can excite resonant conditions
in the ship structure ; most typically the 2-node vertical mode.
•The incidence of after-body slamming, in contrast to fore-body
slamming, frequently reduces with increasing ship speed. This
is because the ship’s entrained wave system increases at higher
speed and gives a measure of protection to the hull after-body
from the otherwise uninterrupted incidence of the environmental
wave system.
No. 190

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•The incidence of after-body slamming, in contrast to fore-
body slamming, frequently reduces with increasing ship
speed.
•This is because the ship’s entrained wave system
increases at higher speed and gives a measure of
protection to the hull after-body from the otherwise
uninterrupted incidence of the environmental wave system.
No. 191
After-body Slamming

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•In addition to being a function of reducing ship speed,
the slamming threshold speed is also dependent on the
sea state, recognizing that the resultant sea state
comprises both underlying swell and wind induced
wave components which strongly influence the
directional slamming threshold.
•Furthermore, a common characteristic possessed by
ships that suffer from after-body slamming is a
relatively flat after-body design coupled with relatively
small immersion.
No. 192
After-body Slamming

© Germanischer Lloyd 2010 Propulsion system Integration 13~25/02/2012
•after-body slamming has been known to occur in sea
conditions with wave heights less than 1m.
•Consequently, the exploration at an early design stage
of hull forms that avoid this problem in association
with the predicted sea and ship motions is of
particular importance
No. 193
After-body Slamming

Vibration Analysis calculation ans methode.pdf

About This Presentation

Slide Content

Tags

Categories

Download

Quick Actions

Statistics

Related Slideshows

Vibration Analysis calculation ans methode.pdf

About This Presentation

Slide Content

Slide 1

Slide 2

Slide 3

Slide 4

Slide 5

Slide 6

Slide 7

Slide 8

Slide 9

Slide 10

Slide 11

Slide 12

Slide 13

Slide 14

Slide 15

Slide 16

Slide 17

Slide 18

Slide 19

Slide 20

Slide 21

Slide 22

Slide 23

Slide 24

Slide 25

Slide 26

Slide 27

Slide 28

Slide 29

Slide 30

Slide 31

Slide 32

Slide 33

Slide 34

Slide 35

Slide 36

Slide 37

Slide 38

Slide 39

Slide 40

Slide 41

Slide 42

Slide 43

Slide 44

Slide 45

Slide 46

Slide 47

Slide 48

Slide 49

Slide 50

Slide 51

Slide 52

Slide 53

Slide 54

Slide 55

Slide 56

Slide 57

Slide 58

Slide 59

Slide 60

Slide 61

Slide 62

Slide 63

Slide 64

Slide 65

Slide 66

Slide 67

Slide 68

Slide 69

Slide 70

Slide 71

Slide 72

Slide 73

Slide 74

Slide 75

Slide 76

Slide 77