Steam Turbine Management in a Changing · PDF fileSteam Turbine Management OMMI ... of rotor stress and differential axial expansion is illustrated graphically in ... margin is afforded

Steam Turbine Management OMMI (Vo. 1, Issue 3) December 2003

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Steam Turbine Management in a Changing Market Eur Ing R.J. Martin CEng MI MechE Steam Turbine Consultant, Innogy, UK

Author Profile: Richard Martin is Corporate Steam Turbine Engineer for Innogy Plc, with 27 years power industry experience. He provides company focus for the resolution of steam turbine issues and is concerned with the prevention of major failures, trouble shooting / forensic investigation works and the provision of authoritative guidance on plant operation, life management, replacement and repair for the Innogy UK fleet and InnogyOne third party clients worldwide. Abstract Changes in the UK Electricity Supply Industry since transition from a government owned monopoly 12 years ago, include substantial and � under NETA � increasing commercial pressure for the flexible operation of large capacity utility steam turbine plant originally designed for base load operation. Flexibility enhancement comes at the price of increased risk: Plant integrity risk, invoked by degradation mechanisms which do not pertain under base load conditions, commercial risk associated with failure to supply to contract schedules and the potential for personnel risk if integrity issues are not managed appropriately. This paper describes the relationship between two key aspects of steam turbine plant operational integrity and flexible operation, and the InnogyOne methodology for safely optimising both.


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1. Introduction

For steam turbine plant, �flexible operating� capability is primarily defined by the following parameters: • Maximum number of start up / shut down cycles without incurring unacceptable

material degradation. • Minimum time from barring speed to full MW load for a range of initial HP and

IP casing temperatures. • Maximum rates at which MW load can be increased and reduced. Flexibility Enhancement: • Identifies design and operational characteristics / constraints which can

beneficially be changed. • Determines the limiting extent of such changes. • Establishes the alterations to plant design and operating practice necessary to

implement them. The application of this methodology to the principle flexibility limiting parameters of rotor stress and differential axial expansion is illustrated graphically in Appendices 1 & 2 and set out in greater detail in this paper.

2. Managing Risks Invoked By Rotor Thermal and Centrifugal Stress Steam turbine rotors are among the most critical and highly stressed components in modern power plants. The potential consequences of a rotor failure include blade loss due to disc head failure, spindle fracture from a thermally initiated circumferential crack and most significantly, fast fracture from a near bore defect causing a catastrophic burst. There have been only a few instances world-wide of catastrophic bursts of rotors, but the consequences are invariably severe as the fragmented shaft is unlikely to be contained within the casing. Fragments of up to 450 Kgs from the failed Gallatin IP/LP rotor which burst in 1974 (Figs 1 & 2), pierced the stations' concrete roof (Schemerling et al, 1976), whilst large fragments of the 71 tonne Irshing LP2 rotor which burst in 1987, were thrown 1300m from the station (Carlton et al, 1988). The risk of injury to personnel from flying debris and escaping high pressure steam is therefore significant. Manufacturers and electric power utilities quantify and limit the risk of such failures using the concept of "rotor life" which is the maximum number of service hours and / or hot and cold starts that the rotor can be safely subjected to. Reducing start up times to improve flexibility raises transient thermal stress levels at the rotor bore and surface, whilst the increase in annual start cycles substantially enhances rotor material degradation rate. Methodologies for assessing and managing the risks associated with these are set out below.


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Figure 1 - Gallatin Unit 2 IP/LP Rotor Fragmentation

Figure 2 - Gallatin Unit 2 IP/LP Rotor Fragmentation


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2.1 Bore Thermal Fatigue (T.F.) Cracking / Brittle Failure 2.1.1 Risk Management Methodology

The methodology set out in Table 1 is used to manage the risk of failure by propagation of an original forging defect in the near bore region to a critical size.

Table 1: Methodology for Assessment Bore High Strain Fatigue Life

Step Action Output Data 1

Estimate maximum size of defect, which would not be identified by original NDT inspection.

Initial Defect Size

2

Determine critical defect size at the most highly stressed area of the bore for worst operating condition (cold overspeed).

Critical defect size.

3

Determine rotor boundary conditions and temperature profile during typical hot and cold starts.

Rotor temperature profile during hot and cold starts.

4

Calculate thermal and centrifugal stresses at the bore during hot and cold starts.

Rotor bore stress levels during hot & cold starts.

5 Calculate crack propagation rate during hot and cold starts

Crack propagation rate due to cyclic stress (mm/start).

6

Calculate creep crack propagation rate at steady load.

Crack size increase per unit time due to creep crack propagation (mm/hr).

7

Calculate number of starts and operating hours required to propagate the smallest crack detectable by NDT to a safe proportion of the critical size.

Maximum permissible combination of operating hours and hot & cold starts.

2.1.2 Location of Critical Areas Critical defect size is determined by maximum total bore stress and bore temperature during the early part of a cold start when the material fracture toughness is relatively low. When thermal stresses are included, total bore stress is highest in the first or second stage disc, which is therefore the critical area.


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2.1.3 Typical Stress Strain Cycle at The Rotor Bore

Fig 3 (Carlton, 1978) illustrates bore stress strain conditions during a rapid cold start whereby the combination of high periphery and low bore temperature causes a tensile thermal hoop stress at the bore.

Figure 3 - Stress Strain Cycle at The Rotor Bore If the combined effect of thermal and centrifugal stresses during start-up is sufficient, yielding occurs at the bore. As the rotor warms through, thermal stresses decrease and the residual compressive stress (due to previous tensile yielding) reduces bore stress to less than the normal centrifugal stress. This reduction is compensated by increased stresses at larger radii in the rotor, which are redistributed by creep during operation. With sufficient operating duration, bore stress increases to the steady state value with attendant accumulation of bore creep strain. Subsequent starts severe enough to cause bore yielding, repeat the cycle, with each cycle increasing creep rate in the rotor body slightly until the equilibrium stress distribution is restored.


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2.1.4 Assessment of Assumed Initial Defect Size

NDT techniques can reliably identify defects down to 2mm Equivalent Flat Bottomed Hole EFBH), where EFBH provides a measure of defect size by comparing the amplitude of the ultrasonic echo to that of a standard flat bottomed hole. It is assumed that a defect of comparable size may have been missed during the last inspection. A safety margin is afforded by assuming the defect

area is eight times that of a 2mm EFBH, and that the shape is a semi elliptical crack of 3:1 aspect ratio (Carlton, 1995). This is approximately the equilibrium shape (i.e. remains constant during crack growth) and gives close to the maximum stress intensity for a given area. If:

a = Defect depth 2b = Defect length d = diameter of equivalent FBH

Then from geometry, for d = 2.0mm, a = 3.27mm, i.e. the depth of the smallest crack detectable by NDT is assumed to be 3.27 mm. Bore inspection using a system developed by Innogy, which combines MPI and Ultrasonic methods, has been shown to provide reliable detection of surface breaking defects down to 1.5mm deep (Anonymous 1981, Jan 2002 & Feb 2002). It is anticipated that ongoing NDT technique developments will enable reliable detection down to < 1mm depth (Murphy, 2000).

2.1.5 Assessment of Critical Defect Size It is assumed that the initial defect will be propagated by cyclic thermal and centrifugal stresses to a final or critical size, beyond which catastrophic brittle fracture would occur. The critical size depends on stress level, material FATT and temperature in the defect region. For bore defects the most arduous combination of these occurs during a cold overspeed test when thermal and centrifugal bore stresses are at their maximum value, whilst the rotor bore temperature (and hence material toughness and resistance to brittle fracture) is low. Based on FATT values for samples removed from service exposed rotors (Hirota et al, 1982; Nix, 1990) a conservative value of 170°C for the FATT is normally used to allow for any temper embrittlement which may have occurred in service. The Jones correlation (Jones, 1972) is then used to establish lower bound fracture toughness (KI c) as a function of temperature, where:

Fig 4

eTemperatur = T Where FATT+T 60

6600 = K Ic −


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For a rotor FATT of 170°C: The stress intensity factor K for a semi elliptical rotor bore defect of 3:1 aspect ratio and axial / radial orientation is given (Smith et al, 1967) as:

Where σ = Tangential stress at the rotor bore. Substituting for KIC the critical defect size ac at maximum stress σm is given by: 2.1.6 Assessment of Number of Cycles To Failure

Rotor bore T.F. life is expressed in terms of the number of hot and cold start cycles required to propagate a defect in the near bore region from the initial assumed size (a0) to the critical or final size (ac). The upper bound fatigue crack growth law for rotor steels is:

��.1

Where

(Haigh, 1971, & 1974; Mulvihill,et al 1988; Paris and Erdogan, 1963). For a defect of the assumed shape:

then

Combining 1 and 2, integrating from initial to final defect size and assuming σ is constant for each cycle, gives:

T2306600 = K IC −

a 1.42 = K σ

− σ m

2

c T) (230 1.426600 = a

K 10 x 1.0 = dnda 311 ∆−

rate growth Crack = dnda

a 1.42 = K σ ��.2

a 1.42 = K σ∆∆


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Where N = Number of cycles to failure. Additional �secondary� effect of creep crack propagation (CCP) at steady load, is summated with cyclic damage to give total �service invoked� bore crack depth. CCP assessment & summation methods exceed the scope of this paper. However analysis for a 500 MW IP rotor in �two shift� service (16 hrs per start), assuming upper bound CCP rate, shows a bore fatigue life of 1380 ECS if CCP is allowed for and 1150 ECS allowing for fatigue damage only (where ECS = cold starts + (hot starts /4)). Assumption of mean CCP rate would render CCP effect negligible (Carlton, 2002).

2.2 Rotor Surface T.F. Cracking / High Cycle Fatigue Failure

2.2.1 General Rotor surface T.F. damage describes the initiation and propagation of cracks at the rotor surface by low cycle high strain fatigue, due to cyclic thermal stressing under conditions of starting and load changing. Since start up and loading rates are constrained by these transient thermal stresses, effective optimisation of plant life / commercial performance requires assessment of the maximum temperature ramp rates that can be incurred without invoking excessive T.F. damage within planned life. Steam to metal temperature differentials are therefore monitored and controlled

during startup and shutdown to restrict surface stresses to values which will not cause

) ( 10 x 1.4a1

a1

= N 311-c0

σ∆

−

2

Fig 5 - Rotor Surface Thermal Fatigue - Critical Locations


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excessive compressive and tensile yielding, thereby limiting the plastic strain range during each cycle to acceptable values. Rotor surface thermal stresses are greatest in areas of high stress concentration and bore to periphery temperature differential. The most arduous combination of these is found at notches such as the heat relieve grooves of the glands at the inlet end of the rotor, fillet radii at the base of discs, balance holes in discs and blade grooving in reaction type rotors. Fig 5 refers. The life management policy addresses the risk of shaft fracture due to a circumferential crack originating at the periphery, and progressing by T.F. to a size large enough to allow failure by mechanical high cycle fatigue (due to rotor 'self weight' cyclic bending stresses).

2.2.2 Typical Stress Strain Cycle at the Rotor Surface Fig 6 shows conditions at a stress concentrating feature (or notch) at the rotor surface during startup and shutdown for severe thermal cycling conditions where yielding of the surface material occurs in tension as well as compression.

Fig 6 – Stress Strain Cycle at The Rotor Surface Referring to the notation: O � A: Rapid heating of rotor, during the very first start up induces compressive stress in surface material with compressive yielding as heating continues.


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A-B: After several hours at steady load, rotor body temperature rises towards surface temperature, surface compressive stress reduces and, due to previous compressive yielding (in stage O � A), becomes a tensile residual stress. B � C: During prolonged steady load operation (dwell period), with rotor surface temperature within the creep range (> 370°C), residual tensile surface stress relaxes by creep strain. C � D: Subsequent unloading or load reduction reduces inlet steam temperature by throttling, rapidly reducing surface temperature below body temperature. Surface material contracts faster than body material �below it�, inducing increased tensile surface stress with eventual plasticity in tension at D. D � E: Heat is conducted from the rotor body to the surface during shut down or part load period, reducing temperature gradient. Surface stress reduces so that the stress strain condition follows the line D - E. At E the rotor temperature is below the creep range, and there is therefore no creep relaxation of the compressive stress at this point. E � A: The next start up cycle will, if of the same degree of severity as the previous one, return the material to condition A, thus completing the cycle.

2.2.3 Calculation of Thermal Fatigue Life

Effect of Cycle Shape For a given material, the cycle shape is defined by two parameters: The total strain range (εr) and the start up (or loading) strain range (εc) as illustrated in Fig 7. Once calculated, these parameters are used to determine which thermal endurance test cycle results are applicable to the service cycle considered. Total Strain Range (εr) Total strain range (εr) is determined by: i) the total elastic stress range (σr) in the material adjacent to the notch. ii) a strain concentration factor (Ksr) based on the total elastic stress range (σr),

the stress concentration factor for the notch, and the material cyclic stress strain characteristics, where the relationship between these parameters is described by the formula:

EK . = srr

rσε


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Fig 7 - Start up (or Loading) Strain Range (εc) εc is determined solely by the start up or loading thermal stress (σc) which changes from compressive to tensile as the notch strain moves from A to B and is given by:

EK . = scc

cσε

Summation of Fatigue and Creep Damage

Overview Life expenditure per cycle comprises two components: i) Fatigue damage determined by the total strain range of the cycle. ii) Creep strain damage determined by the amount of stress relaxation during the

dwell period.


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The Innogy methodology employs the life-fraction method to summate creep and fatigue strain damage. Fatigue Damage Component Initial thermal stress relaxation during the period of steady state operation (the dwell period) is very rapid and is considered to produce a significant contribution to fatigue damage. The fatigue damage component is therefore derived from rotor material thermal fatigue test data, (endurance data) for a 0.5 hr dwell period, rather than from data for a zero dwell period (Batte et al 1984).

Creep Strain Damage Component

Figure 8 - Creep Life Expenditure During The Dwell Period

Creep strain damage per unit time reduces during the dwell period as thermal stress decays with time following a typical relaxation curve. This is modelled by dividing the relaxation curve into a large number of discreet time increments as shown in Fig 8, during each of which the stress is considered to be constant. Cumulative strain damage (i.e. creep life expenditure) is estimated for each time increment using creep rupture data and the life expended during the time increment (∆t) is derived using Robinsons' life fraction rule where:

tt = dr

c∆∆


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tr = Time to rupture at a constant stress representative of the stress during the time increment.

∆dc = Increment of creep damage during the time increment Creep life expenditure for the complete dwell period is the sum of the values for each time increment.

Summation

Thermal fatigue life expenditure for each cycle is derived by summating the creep and fatigue damage components. Since the number of cycles to crack initiation is the reciprocal of the life expenditure per cycle, then:

Where: N = Number of thermal stress cycles to crack initiation (Thermal fatigue life).

Nf = Number of fatigue cycles to crack initiation of a similar (test) cycle with 0.5 hr dwell (fatigue life).

dc = Creep damage in dwell period (i.e from 0.5hr to end of dwell period)

2.2.3 Thermal Fatigue Life Assessment Methodology The methodology summarised in Table 2 (Martin, 1996) is used to assess the number of hot and cold starts to crack initiation in areas of maximum stress concentration at the rotor surface (Fig 9 refers). T.F life is taken as the number of cycles to crack initiation and is conservative as no account is taken of the additional cycles required to propagate cracks to the critical size. The ratio of damage per cycle for hot and cold starts has been found to be typically 1:4. Therefore for guidance to operators, rotor T.F. life is expressed in terms of equivalent cold starts (E.C.S.) where E.C.S. = Cold Starts + Hot Starts/4. A Cold Start being defined as a start after > 30hrs shutdown period and a Hot Start as one following a shut down period of < 30hrs.

Figure 9 - Thermal Fatigue Cracking - 350 MW IP Rotor Stage 1 Disc.

d + N1 =

N1

cf

and

d.N + 1N = N

cf

f


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TABLE 2 - METHODOLOGY FOR ASSESSMENT OF ROTOR SURFACE T.F. LIFE STEP

ACTION

OUTPUT DATA

1

Determine turbine transient temperature conditions (start up rates and steam temperature ramp rates) from typical past or current startup procedures at the station.

Boundary conditions at rotor surface during typical hot and cold starts and duration and temperature of dwell periods.

2

Calculate rotor temperature profile and thermal stresses in areas of maximum stress concentration (Innogy rotor stress analysis program).

Maximum tensile (σt) and compressive (σc) thermal stress levels (and therefore total stress range (σr)), dwell periods and dwell temperatures for hot and cold starts.

3

Derive 0.2% proof stress (σ 0.2) and Youngs' Modulus (E) for rotor material at appropriate surface temperature.

0.2% proof stress (σ0.2) and Youngs' Modulus (E).

4

Derive strain concentration factor (Ksr) for notch. Use Ksr to derive total strain range (εr), where:

E . K = r

srrσε

Total strain range (εr).

5

Use εr in conjunction with high strain fatigue test data to derive fatigue damage per cycle for hot and cold starts. Fatigue damage per cycle (dc ).

6

Repeat step 4 to derive strain concentration factor (Ksc) for compressive strain range & derive compressive strain range (εc). Compressive strain range (εc).

7 Use εc, to derive Initial dwell stress. Initial Dwell Stress (σdi).

8

Use σdi in conjunction with damage factor for dwell period at appropriate temperature and dwell period duration to derive creep damage per cycle. Creep damage per cycle (dc).

9

Summate creep and T.F. damage per cycle to derive total damage per cycle.

Number of thermal cycles to crack initiation (N) for hot and cold starts.

3. Managing Rotor to Casing Differential Axial Expansion 3.1 Axial Clearances to Accommodate Rotor to Casing Relative Movement

Inherent in the design of multi stage labyrinth sealed turbines, is the incorporation of axial clearances between rotating and stationary components, adequate to accommodate all foreseeable service invoked changes in the shape or relative position of the rotor(s), casing(s) and supports. In general, increasing startup rate enhances differences in length change between rotor and casing, with attendant risk of axial fouling at speed between the two and significant consequential damage. This risk is


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managed by evaluating the potential for internal axial fouling under all operating conditions and modifying plant and operating conditions to optimise Relative axial movement between rotor and casing is controlled by the thrust bearing and constrained by the following intrinsic design features, illustrated in Figure 10. • Interleaved shaft end glands • Interstage glands • Blade tip seals • Rotor disc to diaphragm and fixed to moving blade axial clearances. • Journal bearing seals

The maintenance of adequate axial clearance in service at all these positions is essential, though some are more critical than others. Minor axial contact within a labyrinth gland for example may be incurred without obvious deleterious effect (save for local rubbing damage and a small loss of efficiency). However, heavy axial contact between a rotor disc and associated diaphragm will generate intense frictional heating with serious consequential damage / cracking to rotor and diaphragm.

3.2 Differential Axial Expansion

The risk of damage through loss of internal axial clearance is quantified and managed using the concept of rotor to casing differential axial expansion (DAE). DAE can be defined as the difference in service invoked change of axial length between rotor and casing, at a given axial position along the turbine shaft train and can be usefully illustrated graphically to aid operator and maintainer understanding of the concepts involved. The method used by Innogy One to construct a DAE graph is set out below.

3.2.1 DAE Graphs

DAE increases with distance from the turbine thrust collar as the influence of differential heating between rotor and casing on relative expansion, summates with affected axial length. The primary determinants of DAE are:

• Rotor bulk temperature / expansion: Rises faster than casing temperature as rotor mass is lower, rotor is entirely surrounded by steam (outer casing has steam on inside face only) and heat transfer co-efficient between spinning rotor and steam is higher than between stationary casing and steam.

• Casing temperature / expansion. • Rotor Poisson effect: Relative shortening of rotor train during run up when CF

induced radial dilation of the rotor(s) is compensated by axial shortening, to maintain constant material volume for the rotor material in accordance with Poisson�s ratio (typically 0.28 � 0.3).

• Casing support key wear / distortion allowing movement of a casing(s) with respect to (WRT) rotor by an amount equal to key clearance.


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Figure 10 - 500 MW HP Turbine – Positions of Axial Constraints


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• Diaphragm distortion � causing relative movement of diaphragm bore (and associated interstage gland) WRT periphery.

• Thrust bearing wear causing relative movement of rotor WRT casing. • Seizure of support system sliding surfaces causing compressive distortion of a

casing(s) with attendant �shortening� WRT rotor. The effect of each on DAE can be determined using a spreadsheet to: 1. Divide the rotor and casing into axial slices starting at the thrust collar. 2. Calculate thermal expansion of each slice (= length . ∆T. α). Where: ∆T = (slice service temp. � ambient temp) and α = coeff of expansion. 3. Summate expansion values for corresponding groups of rotor and casing slices to

derive the cumulative expansion of rotor and casing at a given position. 4. Calculate Poisson effect for each rotor slice using rotor stress analysis. 5. Add (or subtract) cumulative rotor expansion (including Poisson effect) from

that of the casing to derive DAE at given position. 6. Display results on a graph of DAE vs axial distance from:

1. Thrust collar (for cylinders adjacent to the thrust box - typically applies to HP & IP). By definition, DAE at thrust collar must be zero.

or 2. The end of the cylinder nearest to the thrust collar (typically applies to

LP turbines). Note: DAE at this position will equal the DAE limit for the adjacent cylinder (normally IP) plus any additional DAE incurred in the interconnecting shaft section.

7. Incorporate allowances for offsets such as key wear, diaphragm distortion etc. The process is illustrated by Figures 12 to 14 for the HP, IP, 2 x DFLP turbine shown in Fig 11.


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Figure 11 – HP, IP, 2 x DFLP Turbine Expansion Arrangements


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Fig 12– 2 x DFLP LP Rotor Thermal Expansion Graph

Fig 13– 2 x DFLP Rotor Poisson Effect Graph

Fig 14 2 x DFLP Turbine DAE Diagram

2 x DFLP Rotor Train Thermal Expansion

0.000

2.000

4.000

6.000

8.000

10.000

12.000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Distance from IP / LP1 Coupling (M)

Rot

or T

herm

al E

xpan

sion

(mm

)

2 x DFLP Rotor Train Poisson Effect

0.000

0.500

1.000

1.500

2.000

2.500

3.000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14


Rot

or T

herm

al E

xpan

sion

(mm

)

2 x DFLP Turbine DAE Lines

0.000

5.000

10.000

15.000

20.000

25.000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14


Rot

or T

herm

al E

xpan

sion

(mm

)

LP turbine DAE at 3000 RPM LP Turbine DAE after rundown


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Fig 15– DAE Diagram (after Rundown from limiting value) With Axial Clearances Shown DAE incurred at the position of the DAE measurement collar (14.256 m from IP/LP1 coupling in above examples) is the value displayed in the control room and is equal to the DAE limit for the cylinder(s) concerned. Graphical display of DAE enables optimisation of rotor axial positioning and the determination of optimum gland axial clearances. Transferring gland clearance data directly into the spreadsheet during overhauls enables the immediate evaluation and correction of critically low values (i.e. those falling below the DAE line). Fig 15 identifies two blade to diaphragm clearances in the front flow of LP2 turbine (at 10.758m & 10.951 m from IP/LP1 coupling) in this critical category. The methodology also identifies potential to enhance flexibility by modifying the rotor to increase critical gland axial clearances: The beneficial effect is illustrated graphically by Fig 16 which shows a 34% increase in DAE limit value afforded by machining 1.3mm from the castellation end faces of selected interstage and balance piston glands in an IP turbine.

2 x DFLP Turbine DAE / Axial Clearance Diagram

0.000

5.000

10.000

15.000

20.000

25.000

30.000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14


Rot

or T

herm

al E

xpan

sion

(mm

)

LP Turbine DAE after rundown Disc / Diaph Cnce Blade / Diaph Cnce


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Fig 16– Balance Piston Gland Modification to Increase DAE Limit

IP Rotor / Casing DAE Graph

0.000

1.000

2.000

3.000

4.000

5.000

6.000

7.000

8.000

9.000

0 1 2 3 4 5


Rot

or to

Cas

ing

DA

E -

(m

m)

DAE Line - Unmodified Gland Cnces Std Gland Cnce Modified DAE Line - Modif ied


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4. Conclusions

• Current electricity market conditions increasingly invoke the requirement for

highly flexible operation, on an ageing UK utility steam turbine fleet, originally designed for base load.

• Steam turbine integrity risk enhancements, invoked by flexible operation, can be

safely contained by the application of appropriate plant and operating limit modification methodologies and risk management procedures.

• Safe management of rotor periphery and bore fatigue risk issues, requires

evaluation of limiting temperature and speed change ramp rates and operation in accordance with attendant limits, to avoid unacceptable risks to personnel and plant with potential for massive damage.

• Increasing start-up rates enhances rotor to casing DAE, with attendant risk of

axial fouling at speed and significant consequential damage. This risk can be safely optimised by modelling internal axial clearances under operating conditions to identify necessary plant design modifications and operating limit changes.

5. References Anonymous, �Magnetic Particle Flow Detection�, British Standard 6072, 1981. Anonymous, �Procedure for Ultrasonic Inspection of Rotor Shaft Bores and Adjacent Bodies�, Innogy Procedure No NP/WI/NDT/706, January 2002 Anonymous, �Rotor Bore Magnetic Particle Inspection� Innogy Procedure No TS104/018, February 2002. Batte, A.D., Thomas, G., and Carlton, R.G., �Analysis of High Strain High Temperature Fatigue Data for a 1% CrMoV Rotor Forging Steel and a 0.5% Casing Casting Steel,� CEGB Report No GDCD/PE-T/MD/17/84, 1984. Carlton, R.G., � A procedure for the estimation of the Thermal Fatigue Lives of the High Temperature Components of Steam Turbines�. CEGB Report T/TUR/626.1, 1978. Carlton, R.G., Whitley, G.H., Wooldridge A., �The Implications of the Irsching LP Rotor Failure on the CEGB Policy for the Life Management of Turbine Generator Rotors�, CEGB Report PED/MAT/(88)6, 1988. Carlton, R.G., �HP & IP Rotors - Review of Number of Starts Allowable Before Bore Inspection� Report No CJS/RGC/011, Jan 1995. Carlton, R.G., Personal Communication, 2002


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(Hirota, Y., et al., �Changes of material properties and Life Management of Steam Turbine Components Under Long Term Service�, Mitsubishi Heavy Industries Review Vol 19 No 3, 1982. Haigh, J.R., �Fatigue Crack Propagation in a cast 1% CrMoV steel at 550°C�, C.E.R.L. Note RD/L/N 271/71, 1971. Haigh, J.R., �The Effects of Combined Fatigue and Creep on Crack growth in Tempered Cr-Mo-V steels�, C.E.R.L. Note RD/L/N 191/74, 1974. Jones, G.T., "The Relationship Between Fracture Toughness and Fracture Appearance Transition Temperature for Ferritic Materials", Proc. Institute of Mechanical Engineers Vol 186, p31 - 32, 1972. Martin, R.J., �Methods for Assessing and Extending the Service Life of High Temperature Steam Turbine Rotors and Casings Owned and Operated by National Power PLC�, NP Technology Report N0 TECH/TGG/001/96, National Power 1996. Mulvihill, P., Nix, K.J., Linley, T.C., & Mc Intyre, P., �Outline Procedures for Assessing the Influence of Fatigue and Stress Corrosion Crack Development on Turbogenerator Lifetime�, CEGB Report No TEC/L/0156/R88, 1988. Murphy, T.F., �Eddy Current of Steam Turbine / Generator Bores� GE Power Systems Schenectady NY, 2000. Nix, K.J., �Revalidation of Monobloc Rotors, Draft Strategy for Assessment of Critical Defect Sizes at Rotor Bores and Review of In Service Degradation of Fracture Toughness�, National Power Report No TEC/L/SIFD/0069/M90, 1990. Paris, P.C., and Erdogan, F., �A Critical Analysis of Crack Propagation Laws�, Journal of Basic Engineering, Transactions of the ASME, p528 � 534, 1963. Schemerling, J.M., and Hammon, J.C., �Investigation of the Tennessee Valley Authority Gallitin Unit No 2 Turbine Rotor Burst�, Proceedings of The American Power Conference, Chicago, Illinois, Apr 1976. Smith, et al, �Stress Intensity Factors for Semicircular Cracks�, Journal of Applied Mechanics, p.957, Dec 1967.


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Appendix 1 - Methodology for Managing Rotor Thermal & C.F Stress Invoked Risk

Speed Blade Load

Bore C.F. Hoop Stress

Speed & Load Change Rates

Steam temp ramp rates

Shutdown duration

Initial Rotor Temperature

Bore Thermal Stress

Tangential stress / strain range at rotor bore.

Material high strainfatigue resistance

Bore crackpropagation rate

Max hoop stressduring cold start

Bore Temp

Materialfracture toughness

NDT techniques

Assumed initialdefect size

Bore criticaldefect size

No of Starts to reach critical defect size @ rotor bore

Max startup rate without high brittle failure risk

Cyclic self weight bending stress range at rotor surface

Surface criticaldefect size

Material high strain fatigue resistance

Rotor Geometry

Stress concentration factorat rotor surface features

Thermal stress / strainrange at rotor surface

Surface crack propagation rate

Starts to reach critical defect size @ rotor surface

Max Startup Rate / Cyles Without Incurring High Rotor Brittle Failure Risk , or Surface / Bore Cracking


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Stage temperatures Heat Tx Coeffs Gland steam effect

Thermal Expansion

CF effect on each stage

Poisson effect

Rotor Expansion

Sliding Surfaces Register Keys

Pedestal Movement

Inner to Outer Casing DT Register Keys Casing supports

Inner / Outer Csg DAE Outer Csg / ground

Casing Expansion

Rebuild data

Rotor machining

Rotor Mods

Gland segment mods Diaphram moves

Stator Mods

Modification options

Shaft end glands Interstage Glands Blade Tip Seals Disc to Diaphram Cnce

Internal Axial Clearances

HP Turbine DAE Limit

IP Turbine DAE Limit

LP Turbine DAE Limit

Max Startup / Loading Rate Without Loss of Rotor Axial Clearance

Appendix 2 - Methodology for Managing Loss of Axial Clearance Risk

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Steam Turbine Management in a Changing · PDF fileSteam Turbine Management OMMI ... of rotor stress and differential axial expansion is illustrated graphically in ... margin is afforded