ISO/FDIS 22266-1

FINAL
INTERNATIONAL ISO/FDIS
DRAFT
STANDARD 22266-1
ISO/TC 108/SC 2
Mechanical vibration — Torsional
Secretariat: DIN
vibration of rotating machinery —
Voting begins on:
2016-06-08
Part 1:
Voting terminates on:
Land-based steam and gas turbine
2016-08-03
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
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©
NATIONAL REGULATIONS. ISO 2016

© ISO 2016, Published in Switzerland
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ii © ISO 2016 – All rights reserved

Contents Page
Foreword .iv
Introduction .v
1 Scope . 1
2 Normative references . 1
3 Terms and definitions . 1
4 Fundamentals of torsional vibration . 7
4.1 General . 7
4.2 Influence of blades . 9
4.3 Influence of generator rotor windings . 9
5 Evaluation . 9
5.1 General . 9
5.2 Frequency margins .10
5.3 Dynamic stress assessments .12
6 Calculation of torsional vibration .13
6.1 General .13
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 .14
7.3 Measurement 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 .16
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
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electrotechnical standardization.
The procedures used to develop this document and those intended for its further maintenance are
described in the ISO/IEC Directives, Part 1. In particular the different approval criteria needed for the
different types of ISO documents should be noted. This document was drafted in accordance with the
editorial rules of the ISO/IEC Directives, Part 2 (see www.iso.org/directives).
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. Details of
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on the ISO list of patent declarations received (see www.iso.org/patents).
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as well as information about ISO’s adherence to the World Trade Organization (WTO) principles in the
Technical Barriers to Trade (TBT) see the following URL: www.iso.org/iso/foreword.html.
The committee responsible for this document is 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.
This second edition cancels and replaces the first edition (ISO 22266:2009), of which it constitutes a
minor revision.
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 2016 – 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).
FINAL DRAFT INTERNATIONAL STANDARD ISO/FDIS 22266-1:2016(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 documents, in whole or in part, are normatively referenced in this document and are
indispensable for its application. For dated references, only the edition cited applies. For undated
references, the latest edition of the referenced document (including any amendments) applies.
ISO 2041, Mechanical vibration, shock and condition monitoring — Vocabulary
ISO 2710-1, Reciprocating internal combustion engines — Vocabulary — Part 1: Terms for engine design
and operation
ISO 2710-2, Reciprocating internal combustion engines — Vocabulary — Part 2: Terms for engine
maintenance
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 (3.1)
Note 1 to entry: Figure 1 shows an example.
Note 2 to entry: 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 (3.2) 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 to entry: 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
2 6 8
3 4 5
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
3.6
modal vector
relative magnitude for the whole section, where the system is vibrating at its associated natural
frequency (3.5) 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
2 © ISO 2016 – All rights reserved

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 (3.7) which is produced when the shaft is vibrating at its natural frequency (3.5)
EXAMPLE First mode of vibration or one-node mode of vibration, second mode of vibration or two-node
mode of vibration.
Note 1 to entry: Figure 2 shows examples.
3.10
excitation torque
torsional torque produced by the generator, exciter or driven components that excites torsional vibration
(3.3) of the shaft system (3.2)
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 1 to entry: 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
resonance speed
characteristic speed at which resonances of the shaft system (3.2) are excited
EXAMPLE The shaft speed at which the natural frequency (3.5) of a torsional vibration mode equals the
frequency of one of the harmonics (3.11) of the excitation torques (3.10).
Note 1 to entry: The same definition is given in ISO 2041 in a more general way.
3.14
additional torsional stress
stress due to the torsional vibrations (3.3) of a given excitation harmonic superimposed on the torsional
stress corresponding to the mean torque transmitted in the given section of the shaft system (3.2) being
considered
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 2016 – All rights reserved

a)  Uncoupled frequencies of separated b)  Coupled frequencies of
blade and disc blade-disc assembly
Key
1 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 (3.2) given by the vector sum of all
the harmonics (3.11) of the excitation torques (3.10), taking into account both the magnitude and phase
of the stress generated by each harmonic
Note 1 to entry: A typical short circuit fault is shown in Figure 4 a) which indicates that the fault generates large
torque instantaneously and it clears within a few seconds. The frequency and amplitude content of the fault is
shown in Figure 4 b) for a 60 Hz machine. This indicates the energy is concentered mainly in the line and the
twice line frequencies. Resulting stress responses due to the short circuit fault are shown in Figure 4 c) at two
different rotor locations. The torsional stress responses are seen to follow the behaviour of the fault; ultimately,
they die down over time. Multiple short circuit faults over the life span of rotating machinery could accumulate
stresses in rotor shafts that could eventually lead to damaging shafts severely. Therefore, it is a good practice to
avoid line and twice line frequencies in the design of rotating machines.
Note 2 to entry: Mean torque is not used when elaborating the synthesized torsional stress.
a) Short circuit fault — Amplitude vs time
b) Short circuit fault — Amplitude vs frequency
c) Stress plots for the short circuit fault — Stress vs time
Key
X1 time, s Y1 normalized torque amplitude (dimensionless)
X2 frequency, Hz Y2 torsional stress, MPa
Figure 4 — Example representation of a short circuit fault
6 © ISO 2016 – All rights reserved

3.16
prohibited frequency range
frequency range over which the stress caused by the torsional vibration (3.3) exceeds the stress value
permitted for continuous operation
Note 1 to entry: 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 (3.2).
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 (between
Excite at line
Types of disturbances Step change twice line 0,1 and 0,9)
frequency
frequency of 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)
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.
Table 1 (continued)
Excite at
Excite at (between
Excite at line
Types of disturbances Step change twice line 0,1 and 0,9)
frequency
frequency of line
frequency
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;
— 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-
8 © ISO 2016 – All rights reserved

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.
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 the following 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 5.2 a) to g).
Figure 5 — Definition of torsional frequency exclusion zone
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.
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
Margin between maximum/minimum allowable grid frequency and resonance
B B
peak
C Calculation uncertainty C
10 © ISO 2016 – All rights reserved

Table 2 (continued)
Description Frequency margin
Reduction in calculation uncertainty if a full-speed (dynamic) shop test carried
D out on generator rotor and static shop test (e.g. modal testing) carried out on LP D
rotor
Reduction in calculation uncertainty if full-speed shop tests carried out on the
E rotor(s) of concern, e.g. the generator, exciter, LP rotor or if successful operating E
experience is available for a similar shaft system
Temperature effects
This compensates for the change in shaft stiffness in those cases where
F F
calculations or tests are carried out at room temperature instead of the normal
operating temperature. F is zero if temperature effects have been taken into
account.
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 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.
12 © ISO 2016 – All rights reserved

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.
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
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