ISO 22266-1:2009

INTERNATIONAL ISO
STANDARD 22266-1
First edition
2009-05-01
Mechanical vibration — Torsional
vibration of rotating machinery —
Part 1:
Land-based steam and gas turbine
generator sets in excess of 50 MW
Vibrations mécaniques — Vibration de torsion des machines
tournantes —
Partie 1: Groupes électrogènes à turbines à vapeur et à gaz situés
sur terre et excédant 50 MW
Reference number
©
ISO 2009
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©  ISO 2009
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ii © ISO 2009 – All rights reserved

Contents Page
Foreword. iv
Introduction . v
1 Scope . 1
2 Normative references . 1
3 Terms and definitions. 2
4 Fundamentals of torsional vibration. 7
4.1 General. 7
4.2 Influence of blades . 8
4.3 Influence of generator rotor windings. 9
5 Evaluation. 9
5.1 General. 9
5.2 Frequency margins. 9
5.3 Dynamic stress assessments. 12
6 Calculation of torsional vibration. 12
6.1 General. 12
6.2 Calculation data . 13
6.3 Calculation results. 13
6.4 Calculation report . 13
7 Measurement of torsional vibration. 13
7.1 General. 13
7.2 Method of measurement . 13
7.3 Measurement test report. 14
8 General requirements. 14
8.1 Set supplier responsibilities. 14
8.2 Guarantees . 14
8.3 Responsibilities . 14
Annex A (informative) Torsional vibration measurement techniques . 15
Annex B (informative) Examples of frequency margins relative to line and twice line frequencies
for shaft system modes that can be excited by torsional oscillations of the shaft. 21
Annex C (informative) Commonly experienced electrical faults . 23
Bibliography . 25

Foreword
ISO (the International Organization for Standardization) is a worldwide federation of national standards bodies
(ISO member bodies). The work of preparing International Standards is normally carried out through ISO
technical committees. Each member body interested in a subject for which a technical committee has been
established has the right to be represented on that committee. International organizations, governmental and
non-governmental, in liaison with ISO, also take part in the work. ISO collaborates closely with the
International Electrotechnical Commission (IEC) on all matters of electrotechnical standardization.
International Standards are drafted in accordance with the rules given in the ISO/IEC Directives, Part 2.
The main task of technical committees is to prepare International Standards. Draft International Standards
adopted by the technical committees are circulated to the member bodies for voting. Publication as an
International Standard requires approval by at least 75 % of the member bodies casting a vote.
Attention is drawn to the possibility that some of the elements of this document may be the subject of patent
rights. ISO shall not be held responsible for identifying any or all such patent rights.
ISO 22266-1 was prepared by Technical Committee ISO/TC 108, Mechanical vibration, shock and condition
monitoring, Subcommittee SC 2, Measurement and evaluation of mechanical vibration and shock as applied
to machines, vehicles and structures.
ISO 22266 consists of the following parts, under the general title Mechanical vibration — Torsional vibration of
rotating machinery:
⎯ Part 1: Land-based steam and gas turbine generator sets in excess of 50 MW
iv © ISO 2009 – All rights reserved

Introduction
During the 1970s, a number of major incidents occurred in power plants that were deemed to be caused by or
that were attributed to torsional vibration. In those incidents, generator rotors and some of the long turbine
blades of the low-pressure (LP) rotors were damaged. In general, they were due to modes of the coupled
shaft and blade system that were resonant with the grid excitation frequencies. Detailed investigations were
carried out and it became apparent that the mathematical models used at that time to predict the torsional
natural frequencies were not adequate. In particular, they did not take into account with sufficient accuracy the
coupling between long turbine blades and the shaft line. Therefore, advanced research work was carried out
to analyse the blade-to-discs-to-shaft coupling effects more accurately, and branch models were developed to
account properly for these effects in shaft system frequency calculations.
In the 1980s, dynamic torsional tests were also developed in the factory to verify the predicted dynamically
coupled blade-disc frequencies for the low-pressure rotors. These factory tests were very useful in identifying
any necessary corrective actions before the product went in service. However, it is not always possible to test
all the rotor elements that comprise the assembly. Hence, unless testing is carried out on the fully assembled
train on site, some discrepancy could still exist between the overall system models and the actual installed
machine.
There is inevitably some uncertainty regarding the accuracy of the calculated and measured torsional natural
frequencies. It is therefore necessary to design overall system torsional frequencies with sufficient margin
from the grid system frequencies to compensate for such inaccuracies. The acceptable margins will vary
depending on the extent to which any experimental validation of the calculated torsional frequencies is carried
out. The main objective of this part of ISO 22266 is to provide guidelines for the selection of frequency
margins in design and on the fully coupled machine on site.
In general, the presence of a natural frequency is only of concern if it coincides with an excitation frequency
within the margins defined in this part of ISO 22266 and has a modal distribution allowing energy to be fed into
the corresponding vibration mode. If either of these conditions is not satisfied, the presence of a natural
frequency is of no practical consequence, i.e. a particular mode of vibration is of no concern if it cannot be
excited. In the context of this part of ISO 22266, the excitation is due to variations in the electromechanical
torque, which is induced at the air gap of the generator. Any shaft torsional modes that are insensitive to these
induced excitation torques do not present a risk to the integrity of the turbine generator, regardless of the
value of the natural frequency of that mode (see 4.2 and 5.2).

INTERNATIONAL STANDARD ISO 22266-1:2009(E)

Mechanical vibration — Torsional vibration of rotating
machinery —
Part 1:
Land-based steam and gas turbine generator sets in excess
of 50 MW
1 Scope
This part of ISO 22266 provides guidelines for applying shaft torsional vibration criteria, under normal
operating conditions, for the coupled shaft system and long blades of a turbine generator set. In particular,
these apply to the torsional natural frequencies of the coupled shaft system at line and twice line frequencies
of the electrical network to which the turbine generator set is connected. In the event that torsional natural
frequencies do not conform with defined frequency margins, other possible actions available to vendors are
defined.
This part of ISO 22266 is applicable to
⎯ land-based steam turbine generator sets for power stations with power outputs greater than 50 MW and
normal operating speeds of 1 500 r/min, 1 800 r/min, 3 000 r/min and 3 600 r/min, and
⎯ land-based gas turbine generator sets for power stations with power outputs greater than 50 MW and
normal operating speeds of 3 000 r/min and 3 600 r/min.
Methods currently available for carrying out both analytical assessments and test validation of the shaft
system torsional natural frequencies are also described.
2 Normative references
The following referenced documents are indispensable for the application of this document. For dated
references, only the edition cited applies. For undated references, the latest edition of the referenced
document (including any amendments) applies.
1)
ISO 2041:— , Mechanical vibration, shock and condition monitoring — Vocabulary
ISO 2710-1, Reciprocating internal combustion engines — Vocabulary — Terms for engine design and
operation
ISO 2710-2, Reciprocating internal combustion engines — Vocabulary — Terms for engine maintenance

1) To be published. (Revision of ISO 2041:1990)
3 Terms and definitions
For the purposes of this document, the terms and definitions given in ISO 2041, ISO 2710-1 and ISO 2710-2
and the following apply.
3.1
set
assembly of one or more elements such as high-pressure, intermediate-pressure, low-pressure turbines and
generator and exciter elements
3.2
shaft system
fully connected assembly of all the rotating components of a set
NOTE 1 Figure 1 shows an example.
NOTE 2 When the torsional natural frequencies are calculated, it is the complete shaft system that is considered.
3.3
torsional vibration
oscillatory angular deformation (twist) of a rotating shaft system
3.4
torsional vibration magnitude
maximum oscillatory angular displacement measured in a cross section perpendicular to the axis of the shaft
system between the angular position considered and a given arbitrary reference position
3.5
natural frequency
frequency of free vibration of an undamped linear vibration system
NOTE 1 The same definition is given for natural frequency of a mechanical system in ISO 2041.
NOTE 2 It is usually not necessary to calculate the natural frequency for a damped system, which is
ωω= 1−η
dn
where η is the damping ratio.
3.6
modal vector
relative magnitude for the whole section, where the system is vibrating at its associated natural frequency and
an arbitrary cross section of the system is chosen as a reference and given a magnitude of unity
3.7
torsional mode shape
shape produced by connecting the modal vector magnitudes at each section
3.8
vibratory node
point on a mode shape where the relative modal vector magnitude is equal to zero
3.9
natural mode of torsional vibration
torsional mode shape which is produced when the shaft is vibrating at its natural frequency
EXAMPLE First mode of vibration or one-node mode of vibration, second mode of vibration or two-node mode of
vibration.
NOTE Figure 2 shows examples.
2 © ISO 2009 – All rights reserved

3.10
excitation torque
torsional torque produced by the generator, exciter or driven components that excites torsional vibration of the
shaft system
3.11
harmonic
each term of the Fourier series of the excitation or response signal
3.12
all-in-phase mode
mode of vibration in which all blades in a particular row vibrate in phase with one another
NOTE When the rotor disc and the blades couple under dynamic conditions, the combined system produces several
new “all-in-phase” frequencies that are different from the individual disc and blade frequencies (see Figure 3). These
modes are often referred to as zero-nodal diameter or “umbrella” modes.
3.13
resonant speed
characteristic speed at which resonances of the shaft system are excited
EXAMPLE The shaft speed at which the natural frequency of a torsional vibration mode equals the frequency of one
of the harmonics of the excitation torques.
NOTE The same definition is given for resonant speed/critical speed in ISO 2041.
3.14
additional torsional stress
stress due to the torsional vibrations of a given excitation harmonic superimposed on the torsional stress
corresponding to the mean torque transmitted in the given section of the shaft system being considered

Key
1 high-pressure (HP) rotor
2 low-pressure (LP) rotor 1
3 blades
4 LP rotor 2
5 LP rotor 3
6 generator rotor
7 excitation torque applied
8 exciter
Figure 1 — Six-rotor steam turbine generator system
a)  Second mode of vibration or two-node mode of vibration

b)  Sixth mode of vibration or six-node mode of vibration
Figure 2 — Typical torsional mode shapes of the shaft system

4 © ISO 2009 – All rights reserved

Frequencies in hertz
a)  Uncoupled frequencies of separated blade b)  Coupled frequencies of blade-disc
and disc assembly
a
Rotor central axis.
Figure 3 — Schematic illustration of blade-disc dynamic coupling
3.15
synthesized torsional stress
dynamic torsional stress generated at a section of the shaft system given by the vector sum of all the
harmonics of the excitation torques, taking into account both the magnitude and phase of the stress generated
by each harmonic
NOTE 1 See Figure 4, in which the six upper plots show, for a particular point on the shaft, the time history of the
additional torsional stress for each of the first six excitation harmonics. The lowest plot is the combined effect of vectorially
adding all of the individual harmonics.
NOTE 2 Mean torque is not used when elaborating the synthesized torsional stress.
Key
X time, s
Y torsional stress
Figure 4 — Typical dynamic torsional stress
3.16
prohibited frequency range
frequency range over which the stress caused by the torsional vibration exceeds the stress value permitted for
continuous operation
NOTE Although continuous operation in this frequency range is forbidden, passing through it in transient operation is
permissible, provided that it offers no danger of accumulated damage to the shaft system.
6 © ISO 2009 – All rights reserved

4 Fundamentals of torsional vibration
4.1 General
Torsional vibrations in turbine generator shaft systems are most commonly excited by variations in
electromechanical torque induced at the air gap of the generator. When operating under ideal
steady-state conditions involving balanced three-phase currents and voltages, the effects of higher harmonics
are negligible and the electromagnetic torque applied to the rotor in the generator air gap is essentially a
constant, non-varying torque that transfers the turbine mechanical power through the generator and
electrically to the power system. Under such ideal conditions, there will typically be little or no rotor torsional
vibrations. Torsional vibrations occur as a result of transient or unbalanced steady-state power system
disturbances which act to induce variations in the generator air gap magnetic field and, hence, the torque.
Table 1 summarizes the typical components of air gap torque variations for various types of system
disturbances. The magnitudes of these components depend upon the nature and severity of each disturbance.
These disturbances can be categorized as transient and steady-state. In general, transient disturbances are
cleared after a short time, but steady-state disturbances can persist for extended periods. Further details of
various electrical faults that could occur in power plants are provided in Annex C.
Table 1 — Types of disturbances
Excite at
Excite at
Excite at line (between 0,1
twice line
Types of disturbances Step change
frequency and 0,9) of
frequency
line frequency
Transient:
Three phase fault × ×
a
Unbalanced fault × × ×
Synchronization out-of-phase × ×
Open transmission line (three phases) ×
Close transmission line (three phases) × ×
Single pole switching × ×
Transient sub-synchronous resonance (SSR)  ×
Disturbances in the grid due to thyristor controlled
× ×
loads (e.g. variable speed electric motors)
Steady-state:
b
Line unbalance  ×
c
Load unbalance  ×
Steady-state sub-synchronous resonance (SSR)  ×
a
Unbalanced fault can be either line-to-line, line-to-ground or twice line to ground short circuits. Such faults can be seen either on
the transmission system or more severely at the generator terminals.
b
Line unbalance: Unbalance in transmission line or system, for example, untransposed transmission lines.
c
Load unbalance: Unbalance of the electrical load of the system.

In summary, torsional excitation of turbine generator shaft systems is induced at the generator terminals due
to the following reasons:
a) unbalanced short circuits that produce unidirectional, line and twice line frequency transient torques;
b) out-of-phase synchronization of the unit to the grid, which could produce very high levels of unidirectional
and line frequency transient torques;
c) excitations from other sources, including
⎯ three-phase short circuits,
⎯ transmission line switching, and
⎯ load variations induced and transmitted by heavy-duty operating equipment (such as electric arc
furnaces) in the vicinity;
d) sub-synchronous resonance, which can occur if the generator is connected to long transmission lines and
could excite the sub-synchronous torsional modes. Simple lump mass-spring systems are used in grid
system stability studies to model these sub-synchronous frequencies and their mode shapes;
e) line or load unbalance resulting in negative sequence currents that produce torques at twice the line
frequency.
In view of the possible excitation from the electrical grid, it is necessary to design the overall system torsional
natural frequencies with regard to both the line and twice line system frequencies. For those modes that can
be excited by torsional oscillations of the generator and are evaluated to be critical to the integrity of the unit,
there shall be sufficient margin from both the line and twice line system frequencies. This is the primary
consideration for avoiding any torsional vibration issues on large turbine generators. The following steps are
usually taken into account when defining the margin:
⎯ calculation uncertainty due to inaccuracies of the mathematical models,
⎯ experimental validation of the torsional natural frequencies,
⎯ desired margin between shaft system natural frequencies and the excitation frequency,
⎯ any specified/experienced grid frequency excursions, and
⎯ operating temperature effects.
Mechanical parts that are connected to the main rotor body could participate in torsional vibration if not
adequately designed for strength or tuned away from grid frequencies. These parts include shrunk-on
couplings, coupling bolts and long steam turbine blades. Among them, blade dynamic behaviour in torsional
vibration is complex and is discussed in more detail below.
4.2 Influence of blades
The mode shapes of zero-nodal diameter natural frequencies of blade rows are such that all blades in a row
vibrate in phase with one another. The tangential component of such modes can therefore be excited by
torsional oscillations of the shaft system. In addition, modal interaction takes place between the blades, discs
and shaft system such that the resulting natural frequencies of the combined blade-disc-shaft system are
different from those of the uncoupled components (see Figure 3). It is important to note that for other blade
modes with non-zero-nodal diameters, different sectors of the blade row vibrate in anti-phase to those of
adjacent sectors and are therefore not excited by torsional oscillations of the shaft system.
8 © ISO 2009 – All rights reserved

For short- and medium-height blade rows (e.g. high-pressure and medium-pressure turbines, and the first
several rows of low-pressure turbines), the frequencies of the lowest zero-nodal diameter modes are generally
far away from the frequencies of interest for torsional analysis. Therefore, when calculating the natural
frequencies of the coupled shaft system, such blades can be considered as rigid and only their torsional
inertias need be taken into account when calculating the shaft system torsional natural frequencies.
For longer blades (such as the last and penultimate rows of the LP turbine or the first compressor stage), the
frequencies of the zero-nodal diameter modes can be within the range of, or sufficiently close to, the line
and/or the twice line frequency in order to significantly affect the resulting system modes, which can then
become critical as far as torsion is concerned. These modes interact with those of the shaft system in such a
way that additional coupled shaft system modes are introduced with various combinations of blade vibration in
phase and anti-phase with the shaft system. Under adverse conditions, such modes could amplify rotor/blade
stresses due to external torques arising from grid disturbances. Consequently, when calculating the natural
frequencies of the coupled shaft system and blades, it is necessary to model the long blades as branched
systems that fully replicate the zero-nodal diameter (all-in phase) modes of these blades.
The criterion for assessing whether the blades can be represented by their torsional inertia only, or as
branched systems, is as follows. If the lowest zero-nodal diameter mode of the blade row and disc (or rotor
section for drum type rotors) is less than 2,5 times the nominal line frequency of the electrical grid system (i.e.
125 Hz in countries where the nominal grid frequency is 50 Hz and 150 Hz in countries where the nominal grid
frequency is 60 Hz), consideration should be given to modelling the blade row as a branched system.
Otherwise, it is only necessary to lump the total inertia of a blade row at the appropriate point in the shaft
system model. In general, it could be required that the last stage LP blades (and in some cases, penultimate
stage LP blades) be modelled as branched systems.
4.3 Influence of generator rotor windings
Special knowledge of the generator rotor structural design is needed for modelling the stiffness effects of the
rotor body section with its copper windings and wedges.
5 Evaluation
5.1 General
This part of ISO 22266 provides two methods for the evaluation of the torsional vibration characteristics of
coupled shaft systems including the blades:
a) the maximum frequency margin between the calculated natural frequencies and the relevant electrical
grid system frequencies (see 5.2);
b) dynamic stress analysis to ensure that the peak stresses induced by the transient fault conditions listed in
Table 1 are satisfactory (see 5.3).
Stress analysis for steady-state fault conditions is only required if the frequency margins defined in 5.2 are not
achieved.
Further information regarding the calculation of torsional vibration is given in Clause 6.
5.2 Frequency margins
The objective is to provide criteria that ensure that there are no shaft system modes that can be excited by
torsional oscillations of the shaft within close proximity of the line and twice line excitation frequencies of the
electrical grid system. It should be noted that the shaft system frequencies and associated modes, which are
insensitive to induced torsional forces, are permitted within the frequency exclusion zone (see 5.3).
Calculations by the equipment supplier would indicate whether a mode is responsive or non-responsive to grid
excitations. The allowable torsional frequency margins are shown in Figure 5 and given in Table 2, and are
described in a) to g) below.
If tests/calculations are carried out at room temperature, the relevant upper and lower frequency limits should be
increased by the temperature correction factor F.
NOTE See Table 2 for the definition of A to E.
Figure 5 — Definition of torsional frequency exclusion zone
Table 2 — Margins at line and twice line frequencies for both 50 Hz and 60 Hz machines
Description Frequency margin
Allowable upper grid frequency deviation A1
A
Allowable lower grid frequency deviation A2
B Margin between maximum/minimum allowable grid frequency and resonance peak B
C Calculation uncertainty C
Reduction in calculation uncertainty if a full-speed (dynamic) shop test carried out on
D D
generator rotor and static shop test (e.g. modal testing) carried out on LP rotor
Reduction in calculation uncertainty if full-speed shop tests carried out on the rotor(s) of
E concern, e.g. the generator, exciter, LP rotor, or if successful operating experience is E
available for a similar shaft system
Temperature effects
This compensates for the change in shaft stiffness in those cases where calculations or
F F
tests are carried out at room temperature instead of the normal operating temperature. F
is zero if temperature effects have been taken into account.

10 © ISO 2009 – All rights reserved

a) The allowable grid frequency deviation (electrical grid frequency oscillations leading to line and twice line
frequency excursions) limits, which apply for continuous full-load operation of the particular application, is
identified as A. This value, together with the additional margins given in Table 2, enables users to
evaluate the required frequency margins specific to their grids. This limit varies for different regions
throughout the world and should be agreed between the customer and the supplier. In many cases,
different upper and lower off-frequency variations (A1 and A2) are specified.
b) Margin B is required between the shaft system natural frequency and the maximum/minimum permitted
grid frequency to avoid any significant dynamic amplification near resonance.
c) The confidence in the accuracy of the assessment of the torsional natural frequency of the coupled shaft
system. For example, if the assessment is based on calculations alone, then a frequency margin
identified by C will apply to allow for possible calculation inaccuracies. This is the case if the calculations
are not validated by testing.
d) Confidence of the assessed frequency values increases if they are supported by experimental validation
permitting the calculation uncertainty margin C to be reduced. The extent to which the frequency margins
can be reduced will depend on the level of testing performed and the test configurations. For example, a
field test on the fully installed shaft system will give a greater level of confidence than that provided by
various levels of shop testing performed on individual components of the shaft system.
e) D is the reduction in margin C if a full-speed shop test on the generator and a static test on the LP rotor is
carried out. E is a larger reduction margin which applies if full-speed tests are carried out on generator,
LP or exciter rotors, or if successful operating experience is available for a similar shaft system.
f) Temperature influences the dynamic stiffness of rotors. Therefore, the actual operating temperature
should be included in the analysis. If the calculations are carried out at room temperature, compensation
F for temperature effect is required when the frequency margin is evaluated.
g) Different values of C, D and E can be applied for line and twice line frequencies, depending on customer
needs and special requirements of units.
The above frequency margin types can be dependent on a number of other factors, such as the location of the
power station, the integrity of the electrical network, accuracy of assessment and the operating history of the
supplied hardware. The specification of numerical values for factors A to F is therefore beyond the scope of
this part of ISO 22266. Examples of typical values together with the corresponding frequency margins are
given for information only in Annex B. However, it is emphasized that these may vary for different applications.
The actual values to be used are subject to agreement between the customer and the supplier of the specific
application.
The torsional vibration frequencies should be acceptable if one of the following criteria are satisfied (see
Figure 5).
Criterion 1
The calculated torsional natural frequencies of the coupled shaft and blade system without test verification
should be outside the range specified as +(A1 + B + C) and −(A2 + B + C) of the nominal line and twice line
system frequencies. This is the primary frequency exclusion zone (PFEZ), as shown in Figure 5. If this
criterion is met, no test whatsoever is needed.
If the calculation method has been confirmed by means of factory tests on individual rotors, the modelling
uncertainty is reduced. Hence, the required frequency margin can be reduced and the following alternative
criteria can be applied:
Criterion 2
If the calculation is validated by means of a full-speed shop dynamic test on the generator rotor and a static
test on an adjacent LP turbine rotor, the restricted frequency range would be +(A1 + B + C − D) and
−(A2 + B + C − D) of the nominal line and twice line system frequency. If this criterion is satisfied, there is no
requirement to carry out any further measurements to validate the calculations.
IMPORTANT — Caution should be exercised when interpreting results of static tests on bladed rotors
(see Clause A.4).
Criterion 3
If the calculation is validated by means of full speed shop dynamic tests on the generator rotor and an
associated LP turbine, the effect of blade coupling at full speed will be fully established. In this case, the
restricted frequency range could be further reduced to +(A1 + B + C − E) and −(A2 + B + C − E) of the nominal
line and twice line system frequency. If this criterion is satisfied, there is no requirement to carry out any
further measurements at the site.
Criterion 4
If a full-speed field test is carried out on the fully installed shaft system, the measured torsional natural
frequencies should primarily lie outside the range +(A1 + B) and −(A2 + B) of the nominal line and twice line
system frequency if the torsional natural frequencies are sensitive to grid frequencies. If the field test indicates
that the torsional natural frequencies are insensitive to grid system frequencies, the frequency exclusion zone
provided for by this criterion can be waived.
If the calculations or tests are carried out at room temperature conditions, the frequencies will be marginally
higher than those under service conditions due to the influence of temperature on the modulus of elasticity.
The effect of temperature on the modulus of elasticity for different materials is well established and the
appropriate correction factor can be readily calculated. In such cases, the value of F will be zero. Alternatively,
if such information is not available, a value provided by the equipment supplier should be applied for F, and
the calculated or measured frequencies at room temperature conditions should be reduced by this factor
before applying the frequency margins defined for Criteria 1 to 4.
If the above frequency margin criteria are not satisfied, action should be taken either to perform a more
detailed stress analysis to confirm that the dynamic stresses are satisfactory (see 5.3) or to modify the design
of critical components.
5.3 Dynamic stress assessments
Dynamic stress assessments shall be carried out to confirm the following:
a) the peak stresses induced by the transient fault conditions listed in Table 1 are satisfactory;
b) that the frequency margins specified in 5.2 are not achieved, in which case the shaft system could be
acceptable if the modes of concern are insensitive to the excitation and therefore do not pose any
problem to the system integrity.
In both of these cases, it is the responsibility of the supplier to demonstrate by calculation that the dynamic
stresses do not exceed acceptable values or to demonstrate that the same or similarly designed machines are
operating successfully in other units with comparable grid conditions. In particular, careful attention should be
paid to areas of potential high stress such as coupling bolts, blade roots and those regions of the shafts with
the smallest diameters.
The modelling techniques, calculation method and acceptance criteria are subject to agreement between the
customer and the supplier (see 6.1).
6 Calculation of torsional vibration
6.1 General
Provided that the details of the individual shaft system components are known, it is possible to calculate the
undamped, torsional natural frequencies and mode shapes of the shaft system, including blade-disc-shaft
coupled effects (free vibration). Then, if the frequency margins in 5.2 are not satisfied, the response of the
system to forced excitation mechanisms per Table 1 should be performed (forced vibration) and stress levels
should be in accordance with the supplier’s experience.
12 © ISO 2009 – All rights reserved

The supplier of the shaft system should be responsible for the calculation of torsional vibration using a
conventional method including, where appropriate, the excitation cases to be considered and any allowable
calculation simplification. The method should be agreed upon by the parties concerned.
6.2 Calculation data
The data to be taken into account for the torsional vibration calculation of the shaft system are the polar mass
moment of inertia and torsional stiffness characteristics of each constituent part of the complete shaft system,
its coupled blade-disc branched systems and the specific operating parameters. In addition, if it is necessary
to carry out a forced vibration calculation, knowledge of the torsional vibration damping and the relevant
excitation forces is required.
In some cases, the supplier might not be the original equipment manufacturer (OEM) of some of the shaft
system components (e.g. the turbine and generator may be manufactured by different suppliers).
6.3 Calculation results
The results obtained using the calculation methods described can determine
a) the natural frequencies and the corresponding mode shapes, and
b) the torsional stress margins or torque in the shaft system.
6.4 Calculation report
If the contract requires a torsional vibration calculation to be carried out, a suitable report should be provided
by the set supplier. The contents of the report should be decided between the customer and the supplier. In
general, the report should contain leading particulars of the unit, configuration of the shaft system (including a
summary of which blade rows have been modelled as branched systems) and calculation results. If the set
supplier has subcontracted the calculation then, it should be clearly stated in the report.
7 Measurement of torsional vibration
7.1 General
If the initial calculation shows that there are torsional natural frequencies within the PFEZ, it is necessary to
take further action. This involves either modifying the shaft system components or performing tests to validate
the calculation results and confirm that the application of the reduced frequency margins defined in 5.2 is
permissible. Depending on the particular circumstances, such measurements may be carried out on individual
components in the factory or on the fully installed unit on site. The requirement for, and extent of, any such
testing should be agreed between the set supplier and customer.
NOTE The requirement for testing can be waived if the supplier can demonstrate, to the satisfaction of the customer,
that the accuracy of the prediction method is such that a smaller PFEZ margin is satisfactory.
7.2 Method of measurement
A variety of different measurement techniques have been successfully employed in the past to measure the
torsional vibration characteristics. Annex A provides further background information. However, it is
emphasized that these are not the only available methods and others may be equally applicable. These
methods are subject to continuous improvement. Therefore, the one that is most appropriate for a specific
application will be dependent on a number of factors. Normally, the method adopted will be that which is
commonly used by the set supplier. Any variation from this is subject to agreement between the set supplier
and the customer. However, it should be recognized that testing of the fully coupled unit on site may be an
expensive and time-consuming process that should only be considered under exceptional circumstances. It is
for that reason that the preferred approach is for the set supplier to ensure that the frequency margins
specified in 5.2 are met at the design/manufacturing stage, thereby avoiding the necessity of site testing.
7.3 Measurement test report
If tests are carried out, a torsional vibration measurement test report should normally be provided by the set
supplier. The contents of the report should be decided between the customer and the supplier. It should
contain leading particulars of the unit, configuration of the shaft system, the measurement parameters and the
conditions at the test site. In addition, the type, accuracy and calibration method of the measuring equipment
and the positions of the measurement sensors should be recorded. If the set supplier has subcontracted the
torsional vibration measurements, then it should be clearly stated in the test report.
When the measurement conditions differ from the normal operating conditions, agreed correction factors
should be used to compensate for the effect of the different conditions.
8 General requirements
8.1 Set supplier responsibilities
The supplier of the turbine generator set shall be responsible for ensuring that the torsional vibration
characteristics of the shaft line are satisfactory. In those cases where different manufacturers supply the
turbine and generator, the turbine manufacturer will normally be responsible for the torsional vibration
assessment. It is the customer’s responsibility to ensure that all generator information required to enable the
torsional vibration calculation to be performed correctly is provided to the turbine manufacturer. Nevertheless,
in all cases, a clear responsibility agreement should be established between the respective suppliers and
customer.
8.2 Guarantees
Any guarantees that the set will operate satisfactorily with regard to torsional vibration are subject to
agreement between the customer and the set supplier.
8.3 Responsibilities
Where torsional vibration calculations of the complete shaft system are requested, the supplier of the set shall
be responsible for the calculations, even if he subcontracts them.
Where additional verification of the torsional vibration of the complete shaft system is required, the supplier of
the set shall be responsible for the measurements carried out even when he subcontracts them. In particular,
the supplier should select, in agreement with the customer or the inspection organization representing h
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