ISO/FDIS 22266-1

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

General Information

• Status: Withdrawn
• ICS: 17.160 - Vibrations, shock and vibration measurements
• ICS: 29.160.40 - Generating sets
• Technical Committee: ISO/TC 108/SC 2 - Measurement and evaluation of mechanical vibration and shock as applied to machines, vehicles and structures
• Drafting Committee: ISO/TC 108/SC 2/WG 1 - Rotordynamics and vibration of machines
• Current Stage: 5098 - Project deleted
• Completion Date: 10-Feb-2017

Relations

• Revises: ISO 22266-1:2009 - Mechanical vibration — Torsional vibration of rotating machinery — Part 1: Land-based steam and gas turbine generator sets in excess of 50 MW
• Effective Date: 04-Nov-2015
• Consolidated By: ISO 11737-1:2018/Amd 1:2021 - Sterilization of health care products — Microbiological methods — Part 1: Determination of a population of microorganisms on products — Amendment 1
• Effective Date: 06-Jun-2022

Standards Content (Sample)

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
RECIPIENTS OF THIS DRAFT ARE INVITED TO
SUBMIT, WITH THEIR COMMENTS, NOTIFICATION
OF ANY RELEVANT PATENT RIGHTS OF WHICH
THEY ARE AWARE AND TO PROVIDE SUPPOR TING
DOCUMENTATION.
IN ADDITION TO THEIR EVALUATION AS
Reference number
BEING ACCEPTABLE FOR INDUSTRIAL, TECHNO-
ISO/FDIS 22266-1:2016(E)
LOGICAL, COMMERCIAL AND USER PURPOSES,
DRAFT INTERNATIONAL STANDARDS MAY ON
OCCASION HAVE TO BE CONSIDERED IN THE
LIGHT OF THEIR POTENTIAL TO BECOME STAN-
DARDS TO WHICH REFERENCE MAY BE MADE IN
©
NATIONAL REGULATIONS. ISO 2016

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ISO/FDIS 22266-1:2016(E)

COPYRIGHT PROTECTED DOCUMENT
© ISO 2016, Published in Switzerland
All rights reserved. Unless otherwise specified, no part of this publication may be reproduced or utilized otherwise in any form
or by any means, electronic or mechanical, including photocopying, or posting on the internet or an intranet, without prior
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ISO/FDIS 22266-1:2016(E)

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
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ISO/FDIS 22266-1:2016(E)

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.
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
any patent rights identified during the development of the document will be in the Introduction and/or
on the ISO list of patent declarations received (see www.iso.org/patents).
Any trade name used in this document is information given for the convenience of users and does not
constitute an endorsement.
For an explanation on the meaning of ISO specific terms and expressions related to conformity assessment,
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
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ISO/FDIS 22266-1:2016(E)

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).
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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
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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
2
ωω= 1−η
dn
where η is the damping ratio
1
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
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ISO/FDIS 22266-1:2016(E)

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
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ISO/FDIS 22266-1:2016(E)

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
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ISO/FDIS 22266-1:2016(E)

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.
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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
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ISO/FDIS 22266-1:2016(E)

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.
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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-
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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 th
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