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Safet y Engi neer ing for t he 423 MW-Pelton-Runn ersat Bieudron

VA TECH HYDRO

Pa per to be presented a t the20th IAHR Sympos ium Augus t 6 9, 2000 Charlot te, N.C. USA


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Safet y Engi neer ing for t he 423 MW-Pelton -Runn ers at B ieudr on

R. Angehrn1

VATEC H ES CHER WYS S , Zurich, S witzerla nd

ABSTRACT

The e xtreme output and hea d co nditions a t Bieudron (3 x 423 MW, H = 1869 m) and the

unusua l runner size (outer dia meter 4.63 m) de ma nde d a bo ve-a vera g e eng ineering effort

for runner development and design. Apart from developing profiles for best hydraulic effi-

ciency the mos t dema nding tasks a t this stag e w ere to me et the mec hanica l req uirements of

the integrally cast runners mainly with regard to corrosion fatigue and cavitation behaviour.

Life c ycle a ss es sme nt and optimisa tion o f the inspe ction pe riod s rely on the s tres s res ults of

theoretica l a na lysis b y FEA. As a wo rld no velty, the runner stres s a na lysis is ba se d on b ucket

pressure distributions from laboratory measurements, carried out for the first time on a rotat-

ing runner. Initial experience after about 2000 hours of operation with each of the 3 turbinescan be summarized as successful: the measured hydraulic efficiency complies with the

guarantee figures. No signs of fatigue damage have been encountered up to now on the

exceptionally high-loaded runners.

INTRODUCTIONAs scheduled, the new Swiss power station Bieudron2 (3 x 423 MW, H = 1869 m) was suc-

cessfully commissioned by the end of 1998 [1]. During the 1s t year of operation until the end

of Octob er 1999, the 3 units o perated be tween 1500 a nd 2200 hours a nd the runner buckets

experienc ed up to 200 million loa d cy cles by jet impinge men t. The e xtreme c ond itions in the

pla nt ca n b e d es cribe d, for example, by the ma ximum jet force which is only a little les s than

1 Meg a -Newton (1 MN). This va lue ca n bes t be envisa ge d a s the d ea dwe ight of a mode rn

hig h s peed loc omo tive which, how ever, wo uld be hopping w ith hig h freq uency from buc ket

to buc ket a nd no t just rolling or res ting on its ra ils. The q uestion o f how to d es ig n a nd m a nu-

fa cture runners for suc h de ma nding co nditions is the ma in topic of this pa per.

RUNNER DESIGNDevelopment of the turbines for Bieudron power plant meant a real step forward in the de-

sign o f Pelton turbines [2]. The d es ig n req uireme nts de ma nde d ne w e ng ineering a nd d e-

sign tools, as well as a major extension of existing physical know-how on Pelton turbines.Among a se ries of other spe cial R&D ta sks for Bieudron, three main studies may be men-

tioned:

1. On a single jet mod el turbine, a de tailed experimenta l flow study w a s pe rformed to d e-

velop the necessary knowledge for final selection of the basic turbine concept (three 5-jet

turbines instead of four 4-jet units) as well as for determination of the acceptable amount

of c a vita tion eros ion [3, 4].

2. Optimisa tion o f the co mplete turbine profile on a hom olog ous vertica l 5-jet mod el turbine;

determination of the prototype efficiency guarantees.

3. Deta iled d es ig n study for all ma in turbine parts, ma inly foc used on the mecha nica l de sign

of the runners which is the main issue of this paper.

1 Richard Angehrn, Senior Engineer, VATECH ESCHER WYSS Ltd., Hardstr. 319/P.O. Box, CH-8023 Zurich

Tel. (+ 41)1 278 2406, Fax (+ 41) 1 278 2819, e-mail: richa rd.a ng ehrn@ vate w.ch2 Pelton turbine engineering and supply was by the Groupement Cleuson-Dixence, GCD, comprising Hydro Vevey

and the former Sulzer Hydro, now VATECH ES CHER WYSS , as cons ortium lead er


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S ome cha ra cteristic d a ta a re s hown in Ta bles 1 a nd 2 a s follow s:

Ta ble 1. Turb ine da ta

Number of turb ines 3Ra ted pla nt output 1200 MWRa te d turb ine o utput 423 MW

Ra ted hea d 1869 mS ync hro no us s pe ed 428.6 min-1

Runa wa y speed 756 min-1

Ra ted flow 25 m3/s

J et pitch dia meter 3993 mmNumber of nozzles 5

Ta ble 2. Runn er da ta

Outer d ia meter 4630 mmWidth o f buckets 620 mmNumber of buckets 26

Weig ht of runner 29000 kgWeig ht of a bucket 380 kgMa x. jet force 944 kNCe ntrifuga l bucket force 1470 kNat synchronous speedType o f coupling Fric tion typeMaterial3 stainless ca st steel

G -X5 CrNi 13 4

MEASUREMENT OF PRESSURE DISTRIBUTIONApproximate pressure data can be gained from stationary measurements on individual

buckets [5, 6]. However, the pressure distribution on a rotating runner can only be deter-mined with cons ide ra ble outla y in instrumentation a nd d a ta proc es sing, s ince CFD is no t yet

available. Measurements of unsteady pressure distribution were carried out on a Bieudron

mod el runner under homologo us co nditions in a tes t rig [7]. Fig . 1 show s a sna psho t of

pressure distribution by jet impingement, as an example.

Fig . 2 show s the rela tive bend ing mome nt at the bucket roo t as a function o f rota tion a ng le

or time. The curve is d erived from b end ing mome nt mea surements by s tra in ga uge s a t a

bucket roo t, and the slig htly de via ting points a re b a se d o n integra tion o f the press ure s ig na ls

over the buc ket. The c oincidence a t full imping eme nt is nea rly pe rfec t. S ome underestima -

tion a t jet entra nce is d ue to la ck of sens ors nea r the inlet ed g es be ca use of their thinnes s.

The d a ta ob tained o n hydrody na mic loa d distribution is no t projec t-rela ted a nd forms a n

important basis for design optimisation, also in connection with new manufacturing meth-

ods such as MicroGuss and FiberGuss [8, 9, 10]. As a world novelty, the runner design

stress analysis was based on bucket pressure distributions from laboratory measurements,

carried out for the first time on a rotating runner.

3 Cast in 1991 at Georg Fischer, Schaffhausen, Switzerland, shortly before closing of their steel foundry

runner rotation angle []

relativemomentM[

-]based onstrain gagemeasurementat bucket root

based onintegratedpressuremeasurement

0 8 16 24 32 40 48 56 64 720

0.1

0.2

0.3

0.40.5

0.6

0.7

0.8

0.9

1

Figure 1. Pressure distribution at theinstant of maximum load ing; p

max

co rresponds to 21% of the net head

pressure.

Figure 2. Characteristic of bend ing moment atthe bucket root during jet impingement.Comparison of the measured moment and the

moment based on p-measurement.


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FINITE ELEMENT STRESS ANALYSIS

At the time o f ca rrying out this a na lysis, so lid mo de lling ba se d o n 3D-CAD wa s not a de q ua -

tely d evelope d. The g eo metry wa s de fined ba se d on B-spline s urfa ce de finition [11], and cre-

a tion o f the 3D-mes h need ed co nsiderab le ma nual effort. Due to the co mplex ge ome try a nd

large differences in wall thickness, modelling was carried out exclusively with tetrahedral

elements. The 10-nod e element use d here is spe cially reco mmend ed by the FE so ftwa re sup-

plier for such applications, and gives better accuracy than 4-node elements, although withsubstantially more computing outlay (8500 degrees of freedom or unknowns in the simul-

taneous eq uation system). For symmetry rea so ns, only one bucket ha lf wa s mode lled.

Finite element stres s a na lysis wa s then ca rried out for ce ntrifuga l force loa ding a nd s elec ted

pressure loading according to project conditions [1, 4]. Various geometrical alternatives were

co mputed. In the fina l de sign version the wa ll thickness wa s prog res sively increa se d (up to

40 %) along the b ucket rim tow a rds the roo t (b). As sh ow n in Fig . 3b, this led to s tres s reliev-

ing a t this po int (b) and eq ua lising of the stres se s a t the three m os t hig hly loa de d loca tions

middle ridge fillet (a), bucket rim (b) and inlet edge (c). An interesting point is that originally

(Fig . 3a ) the highes t stres s a t (b) did not result from pea k jet force a nd be nding mome nt, but

occurred under the conditions shown here after the beginning of the unloading phase, with

bucket emptying obliquely outwards, in arrow direction. Due to the force exerted on thebucket shells, extra loading is applied not only to the bucket rim (b) but also to the inlet edge

(c). Unde r centrifuga l loa ding the w a ll thicknes s modifica tion led to a 45% reduc tion in pea k

stress at the bucket rim (b) of the final design, Fig. 3b.

FATIGUE EVALUATION AND LIFE TIMEIn Fig. 4 the FEA stress results are compared with the results of a traditional analysis method

ca lled P ELTBE4. The d ifferences , a lthough no t too la rg e, ne ed so me c la rifica tion. The s tatic

stres s pa rt (= mea n stres s) res ulting from FEA is ne g lig ible a t the mo uth, a vera g e a t the hot

sp ot (= root fillet o f the m idd le ridg e) a nd hig hes t a t the b ucket rim. The P ELTBE va lue lies

4 P ELTBE is a c omp uter prog ramm for Pelton stress analysis based on be am theory

3a) original design 3b) final designFigure

Figure 3. Maximum p rincipal stress due to highest pressure loading at a net head of 1869 mand a turbine power of 423 MW; the jet position angle with reference to the bucket rim p lane isnearly perp endicular (93).

c

a

b

ca

b

ma x

= 40 MPa ma x

= 34 MPa


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midw a y be tween the la tter two . Keep ing the expla na tions a bo ve in mind, this is understa nda -

ble. The d yna mic s tres s pa rts o f FEA a t the ho t spo t, mouth and rim a re w ell eq ua lise d, this

be ing a q ua lity la be l of the de sign . The P ELTBE va lue, h ow ever, is a t a 60% hig her level. For

a b etter understa nding , one ha s to be a wa re of the co nse rvative PELTBE a ss umptions :

At the b ucket roo t the s ec tion a rea a nd the a rea mome nt of inertia a re c a lcula ted w ith a redu-

ced profile width, namely with a width that corresponds to the thickness of the adjoining

disk. However, since the effective root stiffness at Bieudron is higher than usual the support-

ing width is larger than assumed.

Based on the stress analysis a runner lifetime of 80000 hours at full power, equivalent to

a bout 1010loa d c ycles, wa s indica ted in the contra ct. Pa rt-loa d hours a re e valuated a cc ord-

ing to their red uced da ma g ing effec t co mpa red with full pow er opera tion (Cha pter 10).The exa mple s how s a de sign optimisa tion s tep which wa s e na bled by a pplica tion o f the

a bo ve-mentioned too ls. The tra ditiona l metho d w ould not ha ve a llow ed s ufficiently de ep in-

sight into the problem to establish this improvement.

FINITE ELEMENT MODAL ANALYSIS VS. EXPERIMENTThe theo retica l ana lysis (Fig . 5) a ss umes ide ntica l ma ss a nd stiffnes s of a ll buckets a lthough

the reality is different. However, valuable information is obtained on the natural frequency

amplitudes of the basic bending vibration modes. Since with integrally cast runners the

buckets are not fully machined at rear, their geometry can deviate from the precise design

profile. This c a use s s om e sca tter of the na tura l freq uenc ies , and d evia tions from the theo reti-ca l a na lysis. This m ethod s hows up a ll mod es in the freq uency rang e o f interes t, whereas

measurement by hammer excitation normally does not. Namely the two lowest modes of

Fig. 5 were not contained in the response spectra.

HYDRAULIC EXCITATION SPECTRUMAss uming that the b ending mome nt cha ra cteristic (Fig . 2) determined on the mo de l runner

is tra ns fera ble to the full s ize prototype , the d isc rete Fourier spe ctrum of the buc ket excitation

forces can be established (Fig. 6). Compared with previously published examples [13, 14], it

is advantageous for the Bieudron runners that the harmonics in the vicinity of the lowest

16.516.5 1616

26.5

17

Mean Stress [MPa]

StressAmplitude[MPa]

0 50 100 150 200 250 3000

10

20

30

40

50

60

70

80

90

100R=0

contractuel

limit of stress

amplitude

a= 27MPa

PELTBE hot spot

(middle ridge illet)

FEA hot spot (a)

F EA mouth (b)

FEA rim (c)107

108

109

1010

1011

loadcycles

Figure 4. Buc ketdesign stress resultsat 1869 m net headand 423 MW.PELTBE vs. FEA-results. The Haigh-diagram definesfatigue limits for caststeel G-X5CrN i13 4based on tests of theDarmstadt Betriebs-festigkeit Laboratory,LBF [12] . Samples

are taken from nearsurface zones. Loadcycle numbers > 1E8are extrapolated .


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Bieudron 1st natural mode, f = 508 Hz

i = 1 a = 0

Bieudron 4th natural mode, f = 706 Hz

3 2

Bieudron 2nd natural mode, f = 576 Hz

2 1nd


4 3


5 4


6 5

bucket excitation spectra

frequency [Hz]

relativeam

plitude

1 3 5 7 9 11 13 15 17 19 21 232 4 6 8 10 12 14 16 18 20 22

0 100 200 300 400 500 600 700 800 900

integers i of harmonicsof basic frequency

f1

= n/60*Z0

f1

= i*f1

5-jets

1-jet

0.0001

0.001

0.01

0.1

1

mea sured na tura l buc ket freq uencies (a pprox. 760 775 Hz, Fig . 9) ha ve a 50 % hig her orde r

number, namely i = 21 a t 750 Hz a nd i = 22 a t 785.8 Hz. Co nseq uently, the rela tive a mplitude

of the b end ing mome nt or stres s a t the b ucket roo t is sma ller (< 0.0003, Fig . 6). For this

rea so n, the dyna mic s tres s c ha ra cteristic (Fig . 7) co ntains less supe rimpos ed d yna mic s tres s

than originally expected, thus increasing the runner safety. Nevertheless, careful detuning

wa s ca rried out, a s expla ined in the follow ing .

Figure 5. Calculated frequencies of the first 6 natural modes of bucket bending vibration:f1= 508 Hz, f

2= 576 Hz, f

3= 706 Hz, f

4= 761 Hz, f

5= 761 Hz, f

6= 762 Hz. i = number of

mode, n = i 1 = number of vibration node diameters nd.

Figure 6. Discrete Fourier spec trum of jet impingement force on p rototype for 1 and 5-jetoperation (Z

0= 1 and 5), scaled up from model test. The frequency step for 1 and 5-jet

operation is 7.14 Hz and 35.72 Hz respec tively.

0

1

0 72

time

momentM

0

1

0 72

time

momentM


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DETUNINGThe g oa l of detuning is to a void reso na nce, in other words: to prevent a co incide nce of the

na tura l freq uenc ies of the b ucket b end ing vibra tions (Fig . 5) with multiples (or ha rmonics ) of

the basic excitation frequency f1. Resonance would increase the superimposed stress

B

(Fig . 7) since, due to low d a mping , a

high dynamic magnification factor upto 1000 is to be expected .

A frequently asked question concerns

the added mass influence of the water.

This wa s investig a ted a nd a nsw ered

during the course of earlier measure-

ments [13]. It wa s found tha t the a dd ed

mass effect, lowering the natural fre-

q uencies, wa s co mpensa ted by the stif-

fening effect due to the centrifugal

forces . Therefore, thes e effec ts a re ne-

glected here.The a ctua l dyna mic s ystem, c ons isting

of 26 slightly different buckets coupled

by the runner disk, is complex and dif-

ferent from the theoretically investiga-

ted system in Chapter 6. In the past,

each bucket was regarded as a single-

degree-of-freedom system, separated

from the others. Corrections of the

natural frequencies, if necessary, were

executed by grinding the rear of thebucket co ncerned. Subs eq uent chec ks,

how ever, so metimes revea led unexpect-

ed de via tions . The reas on is that the s ys-

tem cannot be separated as described

above. It must be analysed as a multi-

de g ree-of-freed om s ystem (Fig . 8).

Therefore, a ne w de tuning p roc ed ure

[15] wa s d evelope d, initia ted a nd s pon-

so red by the former S ulzer Hydro. The a im w a s to o ptimise the proce ss of fina l bucket grind-

ing . Not only the na tura l freq uencies s hould be reg a rde d b ut also the fina l runner ba la ncing

which ha s to me et the q ua lity req uirements of IS O 1940/1.The ne w proc ed ure is b a se d o n the principles of experimen tal mod a l a na lys is. Mod el ide nti-

fica tion is ca rried out b y me a sureme nt of the d yna mic flexibility ma trix H. The elements hiko f

the matrix H contain the response of the displacement coordinate i due to a unit excitation

(excitation with a unit force as amplitude) at the coordinate k with freq uenc ies in the releva nt

range. In practice, the matrix H is determined by

hammering on a bucket in circumferential direction

and measuring the acceleration of the buckets in

the direc tion o f the displac eme nt coordinate. The

force of the ha mmer is a lso mea sured by a n a cc el-

eration se nso r inside the ha mmer.

h ()11

symm.

h ()21

h ()22

H = H ()31

h ()32

h()33

mass

m1

stiffness

direction of the displacement cooridnate

kc1

kc2kc3 kc4

kc5

k5

k4k3

k1

coupling

stiffness kci

k2

m2 m3 m4

m5

Figure 7. Typical stress trace at a bucket root.

B

= s upe rim po s eddyna mic s tress pa rtdue to harmonicresponse

J

= d ynam ic s tre s s pa rtdue to jets

c

= centrifugal s tress

a

= s tress amplitude

m

= mea n s tres s

Figure 8. The model for d etuning is a multi-deg ree-of freedom system


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Fig. 9 shows the response curves derived from bucket dynamic flexibility measurement.

Multiple close resonances can be recognised which are caused by coupling of the slightly

different buckets. The mo de l is ide ntified by a da pting the unknown pa ra meters (mas se s a nd

stiffnesses) in a way which minimises the differences between measured and reconstructed

excitation. The difference (Fig . 9, 10) be tween the mea sured a nd ca lcula ted dyna mic

flexibilities indicates the quality of the identified model. When the model is found to reflecta ctua l conditions sufficiently we ll, it ca n be used for pred icting wha t hap pens if sma ll ma ss

corrections are necessary by grinding, either due to imbalance of the runner or to correct

bad locations of the natural frequencies.

Detuning is mainly required on runners rotating at speeds higher than about 400 rpm in

turbines w ith 4 or more no zzles , whe re the b ucket freq uenc ies a re relatively low (< 700 Hz).

NON-DESTRUCTIVE TESTINGThe d ec ision limits for NDT a re strictly ba se d o n a ctua l stres s levels a nd o n a dm iss ible fla w

size a s d etermined by fra cture mec ha nics . Qua lity control wa s c a rried out using UT, MT a nd

P T, whe rea s RT wa s a pplied only a t the firs t qua lity co ntrol sta g e in the found ry.

Nine inspe ctions b y MT, three on e a ch runner, we re ca rrried out b etwe en c om miss ioning in1998 and the end of Octob er 1999 without finding a ny ina dmissible fla ws or crac ks.

SAFETY ASPECTS DURING OPERATIONAND MAINTENANCEAs typical with power stations for peak energy production, the Bieudronunits o perate o ver a

wide power range and go through several start-stop-cycles per day. Figure 11 shows the

distribution of service hours within five power ranges, averaged over all units.

In the following, some points are discussed concerning the possibilities of part load opera-

tion a nd the eva lua tion o f service hours with rega rd to fatig ue.

Pelton

runner

math.

model

excitation response

deviation

+ -

Figure 9. Frequency response curves ofmeasured and simu lated flexibility of abucket. Multiple close resonances can berecognized which are caused by couplingof the slightly different buc kets.

Figure 10. The identification of the math-ematical model is based on the adaption ofthe unknown parameters (bucket mass,stiffness) in a manner wh ich m inimizes thedifferences between the measured and thereconstructed excitation.


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Modes of Part Load Operation and Consequences

With smooth starting and stopping procedures, the start-stop-cycles of the turbines nor-

mally have much less influence on material fatigue than the dynamic excitation due to jet

pulsation.

A multi-jet-turbine can be operated at part load in two different ways, each having differentinfluenc es (Ta b. 3). Althoug h efficiency loss oc curs, there a re pla nts w here a ll noz zles a re

a lwa ys in ope ra tion, even a t part loa d. How ever, a t Bieudrona n operating mode w as chosen

where the number of jets is strictly adapted to the actual load level. Figure 12 shows the

resultant difference in stress amplitude at the bucket root.

Figure 11. Typic al B ieudron

turbine load spec trum afterthe first year of operationsince c ommissioning. Effectivehours and correspondingPmax hours.

Table 3. Consequences ofdifferent part loadoperation modes:a) adapting and b ) non-adapting the number of

jets.

Figure 12. Stressamplitude at the buc ketroot in function of turbinepower at different partload operation modes:a) adapting as atBieud ron andb) non-adapting thenumber of jets. Theeffective switch pointsdeviate slightly from theschematic po ints shown

in the diagram.

1.0

0.8

0.6

0.4

0.2

0.0

effective hours

corresponding Pmax-hours

total

noload

0-100M

W

101-200

MW

301-420

MW

201-300

MW

0

0.2

0.4

0.6

0.8

1

0 0.2 0.4 0.6 0.8 1

relative power, P/Pmax

relativestressamplitude

1 2 3 4 5 jets

active

a) b)

influe nc e pa ra m ete rs a ) a d a p ting the b ) no n-a d a p ting

num ber o f je ts the num ber o f je ts

s tres s a mplitude hig her lower

at bucket root

inspection periods s horter long er

life time s horter long er

la tera l sha ft force > 0 0

bend ing moment > 0 0

of turbine shaft

turb ine bea ring loa d > 0 0

hydra ulic effic iency better lower


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Evaluation of Load Spectrum and Ins pection Periods

Inspection periods in a pla nt should be defined ba sed on po ss ible c ra ck propa ga tion s ce-

narios using linear elastic fracture mechanics. Since they should also take into account the

actual stress level, the load spectrum has to be evaluated when calculating inspection peri-

ods. For the evaluation shown here, the Paris equation (1) and the basic equation (2) of

fracture mechanics areused, leading to the follow-

ing equations 36:

)1(/ m

KCdNda

)3()/( 2/mmm aCYdadN

Integrating equation (3):

If m = 2 applies the inte-

g ra ls a re:

)5(ln)/(11

222

12

a

aCYNNN

If a1 and a

2 are assumed to

be s et at the start a nd end of

an inspection period, and C

and Y are constants, the fol-lowing equation is valid and

d e f i n es t h e f a t i gu e c r ack

damage potential FCD, as-

suming it constant:

)6(.2 constFCD

From equation (6) and

Fig . 12 it can be see n that

at a service load of e.g.

60% (or a sma ll po rtion b e-low) the following evaluation factors for service hours at part load apply:

in op erating mod e a ) the e valua tion fa ctor co rres pond s to 0.6. This is due to the reduce d

number of loa d c ycles if 2 nozzles a re c los ed . The s tres s a mplitude, ho wever, is a t the

maximum;

in operating mode (b) the number of load cycles is not changed since all nozzles stay

ope n, how ever, due to the reduced stress a mplitude the eva lua tion fac tor is 0.62= 0.36.

The eva lua tion fac tors w ould b e slig htly d ifferent if the exponent mof the Paris equation is at

m = 2.25 instead of 2, as indicated in [16] for martensitic steels. In [17] the corresponding

pa ra meter is 2.55. Thes e de via tions ha ve b een neg lec ted a t the eva lua tion fa ctor de finition.

0

0.2

0.4

0.6

0.8

1

0 0.2 0.4 0.6 0.8 1

relative power, P/Pmax

evaluation

factor

a) b)

Figure 13. Evaluation fac tors for service hours at the twopossible part load operation modes: a) adapting as atBieud ron and b ) non-adap ting the number of jets.

N0 N N number of cycles, N1 2

inspection

interval

P < Pmax

Pmax

dN

da

crack

length

a

a

aa

2

1

0

Figure 14. Crack p ropagation behaviour in c onstant amp li-tude fatigue load ing.


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11

The insp ec tion pe riod s ha ve been de fined 5 as follows:

1s t inspe ction 200 h after commiss ioning (service hours are based on

2nd inspe ction 400 h after the 1st inspection operation at Pmax)

3rd inspe ction 400 h after the 2nd inspection

4th inspe ction 600 h after the 3rd inspection

5th

inspe ction 600 h a fter the 4th inspection, etc.

As a conclusion it can be said that the Bieudron runners have been designed on top-state of

the art engineering practice, and represent a milestone in Pelton technology.

ACKNOWLEDGEMENTSThe a uthor would like to tha nk a ll his c ollea g ues who co ntributed to the de velop ment o f

these new tools a nd their suc ce ss ful a pplica tion in the pres ented projec t.

REFERENCES[1] Loth, P., A reco rd brea ker. Cleuso n-Dixence w ill go on line ..., Pa rt Hydromec ha nics , Internationa l Wa ter

Power &Dam Construction, June 1998, Pages 22-24.[2] Keck, H., Sc hrer, Ch., Cunod , R., Cateni, A., Pelton technolog y for new plants and modernization s chemes,

International journal on Hydropower &Dams, Vol. 4, Iss. 2, 1997, pages 104-108.[3] Ba chma nn, P., S ch rer, C h., S taubli, T., Vullioud, G ., Experimenta l flow studies o n one 1-jet mod el Pelton

Turbine , 14th IAHR S ympo sium, B elgra de , Yug os lavia , 1990.[4] Be zinge , A., Ba chm a nn, P., Vullioud , G ., Da s P rojekt Cleus on - Dixenc e , VE/S EV/VDE/-Fac hta g ung

Wasserkraft - Regenerative Energie fr heute und morgen, Vienna, May 1992.[5] Obretenov, V.S., Ana lysis o f the Flow o n a Pelton Turbine Buc ket, P roc . of the Lenin Highe r Inst. of Mecha nica l

and Electrical Engineering, Vol.41, Book 3, Sofia, 1987.[6] G rozev, G., Ob retenov, V., Trifonov, T., Investiga tion o f the Distribution o f P ress ure over the B uckets o f a Pe lton

Turbine, P roc. o f the C onferenc e on Hyd raulic Mach inery, Turboinstitut, Ljublja na , S ept. 1988, P a g es 119-125.

[7] Angehrn, R., Rettich, J ., Sch rer, Ch., Pelton runner des ign b as ed o n mea sured unstead y press ure distributionsin the b ucket, Hydropowe r &Dams , Issue S ix, 1999.

[8] S chnee beli, F., Ba ltis, E., Keck, H., New Techno log y Ea rns Acc eptanc e, S ulzer Technica l Review, No. 1, 1996.[9] Kalberer, A., Krause , M., A Review o f Experienc e with MicroCa st Pe lton Wheels, Hydropo wer & Dam s,

Issue 1, 1996.[10] Kra use , M., Ried el, A., Innovationen b ei der Fertig ung und Re pa ratur von Wa ss erkra fta nlag en MicroG uss TM

und B es chichtung en, 9th Interna t. Se mina r on Hydro Po wer P la nts, Wien, 1996.[11] G rein, H., S chne eb eli, F., Ba ntli, H., 3-Dimen siona l S urfa ce Mode lling - a Des ig n and Ma nufac turing Too l for

Hyd raulic Ma ch inery, S ulzer Tec hnica l Review 2/1989.[12] Ostermann, H., Rckert, H., Ausfallsichere Be mes sung von Laufrde rn fr Wa ss erkraftma sc hinen a us rost-

freiem Stahlguss unter Bercksichtigung von Korrosion und Gefgezustand , Abschlussbericht zumG eme ins cha ftsp rogra mm S tah lgus s , B MFT-Industrie-LBF, Fraunho fer-Institut fr Be trieb sfes tig keit, Da rm-stadt, Sept. 1983.

[13] Ang ehrn R., Duba s M., Experimenta l S tres s Ana lys is o n a 260 MW Pe lton Runner, P roc. o f the 11th IAHR-S ympos ium, Amsterda m, 1982.

[14] G rein, H., Ang ehrn, R., Lorenz, M., Bez ing e, A., Inspe ction Pe riod s o f P elton Runners , Proc . of the 12th IAHR-Symposium, Stirling, 1984.

[15] Sc hmied, J . , Von der Sc hwingung smes sung zum S imulationsmode ll , Technische Runds cha u Nr. 17,Bern, 1998.

[16] Fuchs , H.O., S tephe ns, R.I., Metal Fa tig ue in Eng inee ring , Wiley &S ons , NewYork, 1980.[17] G rein, H.L., Ange hrn, R., S ervice Life of P elton Runners under C orros ion Fatigue , Interna tiona l Sympo si-

um o n Fluid Mac hinery Troubles hoo ting , ASME Winter Annua l Mee ting , Anahe im, Ca lifornia , De c. 1986,FED-Vol. 46/P WR-Vol.2.

5 The inspec tion periods d efiniton is cohe rent with the results of a cra ck propag ation ca se s tudy which wa s exec uted on a

260 MW-runner after a damage occured in 1981 [17]. As a consequence, stringent inspection periods were introducedthere a nd it can b e reported ab out succes sful experience up to now.


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