Unit Upgrades Fatigue and Fracture Mechanics analysis form a Utilities

Standpoint.

I know this is a dry topic

BUT

Why is this important?

Picture of U55 with yielding rim ledge - 2013

Picture of U55 with yielding rim ledge. 2013

Picture of U56 with cracked rim ledge. 2015

Picture of U56 with cracked rim ledge. 2015

Picture of Bonneville with broken rim ledge.

Side Note – what we had to do – Seasoned Rim

What the ledge looked like

I’m I an expert – NO But I know that this is important – don’t want anyone to get hurt, don’t want hydro to get a bad reputation. Our equipment is getting older, upgrades and life extensions are common.

What happen here? Multiple things but the fracture of the headcover bolts was due to fatigue. There are a lot of major components that if they failed could be catastrophic.

Questions to ask yourselves –

1. How long will the equipment last, if I continue to operate it as I have in the past – Forever is a long time!

2. How Long will the equipment last if I change how I operate the equipment in the future

Important things for a Utility to consider for unit upgrades and life extensions.

1. How much power upgrade is obtainable, but also how much power upgrade is acceptable?

2. What were the OEM’s design criteria – all those years ago? a. How many starts is it designed for? b. How many load swings?

3. How was the unit operated in the past? a. Do you know? b. Can you make accurate assumptions?

4. How do you expect the unit to be operated in the future? 5. What other potential modifications has your utility made to the unit

over the years? a. Replacement or modifications to the packing sleeve? b. New runner with higher hydraulic thrust? c. More or less load rejections, load rejections from higher loads =

higher maximum overspeed’s?

Analysis indicated 37 more years until earliest crack initiation – inspection found it clean

What else has changed since that unit was designed all those years ago? • Materials – especially in the spiders

o Not high-grade steels. o Welding practices changed o Inspection techniques changed

• FEA and fracture Mechanics has evolved since the 1940,50, 60’s etc. o The OEM’s did not have the tools or expectation of how

unit operations and life extensions would evolve. o Fatigue considerations omitted from non-pump storage

unit design o Designed to OEM endurance limit criteria – if stresses <

criteria design acceptable. o The longer the machines operate and the more starts they

accumulate the less valid these initial assumption become.

What to do / How to start?

Power upgrade – new runner/Generator higher outputs effects: • Increased hydraulic thrust – Thrust bearing • Turbine/Gen shaft and couplings • Rotor Spider • Spider to rim iron • Rim iron to poles • Main bracket • Multi-turn windings replaced by Robel Bars – less parallel

circuits – higher unbalanced magnetic pull.

What data should be provided in your SOW for the FEA analysis?

The steady state operation at full power: • Constant torque transmitted through the turbine at the rating • Year of operation • Hydraulic thrust load • Weight of the runner/shafting (connected components) The alternating loads: • Number of Starts and Stops (Shafts, poles, spider, rim) • Load rejections per year (Peak RPM data from 100% load and

Run-a-way speed, may be needed) (Poles, Rim, Spider) • Periodic variations in the torque from the nominal level (torque

pulsations) (Shafts, Spider) • Side loading due to hydraulic thrust imbalance in the runner

(Shaft)

Example

The following operating criteria should be used for the analysis Loading event Original

Runner (Nohab)

Original Runner (Nohab) Plant Expansion

Existing Runner (Noell)

Future Operation

Years operated 1967-1986 1986-2000 2000-2020 2020-Future Normal Start Stops (per Month)

12 16 16 30?

Load Variations 144/year 192/year 192/year ??????? Load Rejections (per year)

5 5 5 5

Draft Tube Pulsations₁ (pulsations/ min)

40 40 40 40

Runner Side Load2 (per minute)

120 120 120 120

Static loading for all fatigue cases. Applied Loads (at 250 ft. Net Head)

Original Runner (Nohab – 155MW)3

Existing Runner (Noell – 165MW)

Existing Runner (Noell new Gen. 180MW-future ops. 2020+)

Runner Torque (ft/lbs.)

9,097,300 9,685,000 10,565,000

Hydraulic Thrust (lbs.)

651,000 (note #4)

1,050,000 (note #5)

1,050,000

Bolt Coupling preload

950,000 lbf per bolt

950,000 lbf per bolt

950,000 lbf per bolt

Runner Weight (lbs.)

140,630 141,940 141,940

Discuss ways to determine hydraulic thrust

Alternating Loads and Cycles for fatigue cases Load Type Value Original

Runner (Nohab – 155MW)3

Existing Runner (Noell – 165MW)

Existing Runner (Noell new Gen. 180MW-future ops. 2020+)

Start/Stops 0-120 RPM 5424 3840 360/year Load Rejections

120-210 RPM

165 100 5/year

Load Variation

0-100% nominal torque

3762 2280 114/year to 10/day

Torque Pulsations

±316,950 ft-lbf (±3% Max Torque)

6.943 x 10^8

4.208 x 10^8 2.104 x 10^7 /year

Runner Side Loading

0-775,686 ft-lbs

2.08 x 10^9 1.262 x 10^9 6.311 x 10^7 / year

Example (cont.)

Relevant drawings and notes of know modifications/ Examples: An Alternating runner side load, due to hydraulic imbalance should also be applied. It is calculated based on the average seal diameter times the height of the seals times 10% of the normal hydraulic head. (Each OEM has a somewhat different criterion – you can provide or ask for then to provide and given an explanation) The load will be applied once per revolution. The lever arm shall be from the mid-plane of the runner distributor to the mid-plane of the Turbine Guide Bearing (Aver seal diameter =202.3”, Height = 58.8”, Normal Head 250ft. with a lever arm of 72.2” (Distributor centerline to center of TGB journal)). Low load, high cycle condition due to draft tube pressure pulsations. The load will be applied at 1/3 operating speed and be based on ±3% power swings at the maximum turbine power. (Each OEM has a somewhat different criterion – you can provide or ask for then to provide and given an explanation) The Turbine shaft is made from SIS 14 1650 (Swedish industrial standard). A carbon seal ring carrier (CSRC) sits on a ledge on the shaft (SMD-27) with a 1/32” radius just above the shaft vent holes. The diameter that the CSRC mates to was originally 49.213” ±.000/.008”. This surface was cleaned up due to significant corrosion per the notes on the drawing, smallest diameter is now 49.180”. There are also 4 drive pins for the CSRC that fit into 4 holes in the shaft. The original holes were 1 3/16” diameter per the shaft drawings. The pin and hole were designed to be body fit. No actual documentation exists but it is known that these holes were in such bad shape that they were modified by oversizing the holes. For this reason, it should assumed that the final hole diameters are 1 3/8” diameter with an edge radius of 1/32” for analysis purposes. These areas will be accessible during the rewind for inspection and NDE.

What do you want your FEA consultant / OEM to do?

What do you want to know from the report?

1. Can I upgrade? 2. How long can the component last? 3. When should I do an inspection? 4. Where should I look (U53) 5. What should I look for?

Some Background Fatigue Analysis Types • Strain Life • Stress Life • Facture Mechanics

When you are reviewing these analysis, you will see some very high stress – which are acceptable. WHY – elastic plastic deformation are redistributing the loads. Discuss – shaft stress typical design load <6000 psi stress – but

Strain Life (EN) • Used to determine crack initiation • Best for dealing with Low Cycle Fatigue (LCF) < 100,00

cycles • Also works for High Cycle Fatigue (HCF) • You will hear the terms Elastic/Plastic and non-linear

when discussing Strain Life analysis with your consultants • Strain Life analysis requires the total strain as an input

(elastic & plastic) • Using cyclic strain, the consultant will calculate local

stress/stains (sum of elastic and plastic strains). • Strain Life (typical methods) will always identify a material

cracking even if below the materials endurance limit, which we would not expect. This is an inherit conservative approach, due to the calculation method.

Stress Life (Typical OEM Process) • Uses S-N Curves for specific materials. (S-N stress-life

diagrams) • S-N Curves not available for all materials (limitation) • ASME Method used for HCF and LCF • FKM Standard HCF >10,000 cycles only • These Codes provide design lives and are not meant to

predict actual lives – What does this MEAN?

Safety factors – factor of safety of 2 on stress and 20 on fatigue

Other Considerations

Material Issues – if someone other than the OEM is doing your upgrade analysis you often don’t have material properties – Our solution has been to take shaving and hardness measurements in low stress areas. Not having the right material properties can have a significant impact on results.

Material surface condition, roughness, corrosion in fresh water, etc. – Surface correction factors can have a large effect on HCF.

An Example:

Stress Amplitude are calculated out of the FEA analysis – they are used in the S-N curve to determine the allowable number of cycles.

Miner’s rule for multiple loading scenarios

Instead of stating the number of starts/stops in your scope – maybe you ask for the maximum number of start/stops still available for the component. Then you can use this information to determine your inspection intervals and how to limit potential operations that will compromise future life.

Fracture Mechanics / LEFM (Linear Elastic Fracture Mechanics) • Starts with an assumed flaw of known size and

determines the crack’s growth rate until rapid fracture. • Fracture Mechanics is also referred to as Crack Life. • Based on the assumption, the stress field near the crack

tip is calculated using the theory of elasticity. When the stresses near the crack tip exceed the material fracture toughness, the crack will grow.

• Used to determine Inspection intervals o What is the smallest crack you could find (1/16” or

1/8” by PT?) o From your minimum detectable crack size – Fracture

Mechanics will calculate the time it will take for the crack to grow to a critical (failure) size.

o Then you could set up an inspection interval at ½ or ⅓ the crack growth rate to ensure serviceability of the component.

Take away – Strain life methods can be used to determine crack initiation with Fracture Mechanics used to determine the crack life.

CRACK INITIATION + CRACK LIFE = TOTAL LIFE

Said another way - Strain based analysis is used to determine crack initiation then once the years to an initial crack are determined - use a stress based LEFM analysis to determine time to critical crack size. LEFM uses stress intensity at the crack tip to determine crack growth.

OEM will typically perform a stress based (LINEAR) FEA analysis and then compare to the ASME BPV S-N Curves. Advantage - Defined method with pedigree. Conservative. If method says life is OK, very likely this is true.

Disadvantage - Conservative. Often too much so. For instance. Many units have operated with pole attachments exceeding the design life predicted by the ASME code. May drive unnecessary replacement work.

Why so much talking about Stress based FEA and ASME / FKM codes vs. Strain based and LEFM analysis. If you’re having an OEM preform the analysis for a unit upgrade and have asked them to do all the upgrade analysis – this gives you some background on what and how they are doing it. If you have upgraded a unit in the past (new runner), but did little analysis to determine the acceptability of the upgrade, you may hire a consultant to perform the analysis. Same for having a significant change in how you operate your units.

If so – you may want to perform a stress-based analysis (to determine life to crack initiation). And the LEFM from crack initial to failure. Then with this information you can determine inspections intervals. Every time you inspect and do not find any cracks, you can reset your “crack clock” and the components life is extended. This information also tells you where to look.

Conclusion:

1. Provide clear past history data 2. Provide expected future operations or 3. Request an analysis to be performed parametrically

4. OEM like to sell equipment – Ask can the component be saved by inspection and geometry changes – larger radius – weld reinforcement – gusset plate size changes.

A.M. Goodison – Design Engineer on 1930 era Westinghouse drawings

Governing Codes:

1. ASME BPVC-section VIII- division 2 2. API579-1 / ASME FFS-1 (Fit for service) 3. FKM-Guideline (German standard) – Not applicable to LCF see note

CEATI Publication #0357 – Turbine Generator Shaft Stress Analysis methods & Limitations – Provides guidance in support of developing stress analysis / life assessment methods and up-rating guidance for hydroelectric turbine-generator shafts.

Questions?