ML17318A706

Last-Stage Blades of Large Steam Turbines

ML17318A706
Person / Time
Site:	Cook
Issue date:	03/17/1980
From:	Hohn A, Novacek P BBC BROWN BOVERI, INC. (FORMERLY BROWN BOVERI CORP.
To:
Shared Package
ML17318A689	List: ML17318A688 ML17318A690 ML17318A691 ML17318A693 ML17318A706 ML17318A707
References
CH-T-060040-E, CH-T-60040-E, NUDOCS 8004220034
Download: ML17318A706 (12)

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Ilute BROWN BOVERI Last-Ste Blades of Large Steam Turbines Publication No. CM-T060040 E A. Ifohn and P. ¹vacek The present article deals with the blades ih the last rotating row in large steam turbines, consfdering them as a machine element.

The static and dynamic stresses occurring in service are discussed and their effect on the design of the blades is demonstrated.

Some methods oftestfng which are used in the design ofprototypes are explained as they enable blades designed on pure theory to be tested under conditions comparable with those experienced"in

service, thereby enabling the behaviour of the blades in service to be pre-dicted. ¹wadays thfs performance ls from time to time checked ln servfce fn power stations; the article provides some information regarding test pr ocedures and the results obtained. In conclusion the authors discuss future develop-ments ln blade eonstructfon.

Introduction A feature ofthe last ten years in the construction ofsteam turbines was the marked risc in the unit ratings of ma-

chines. Atthe beginningof the 1960's the majority oflarge thermal machines in European installations were mainly in the 125 to 150 MW range. Today, both in Europe and in America, machines with an output of morc than 1000 MW are being installed. In them the volume of steam that has to bc handled on emerging from the blading (last-stage) may be ofthe order ofmagnitude of 10000m'/s (fossil fuelled plants) up to 25000 ms/s (in nuclear power plants with turbines employing saturated steam). In.order to handle such enormous quantities of steam in a reason-able number offlows, the cross-section offlowin the blade ducts has to be large and the last stage correspondingly long. On the other hand, in thc last row of blades of large steam turbines about 6% of the total heat drop of the steam flowing through the turbine is converted into me-chanical energy. Since these two factorshighoutputand quality of the energy conversion are also influenced by the last stage, particular attention has been paid to these blades during the past ten years. Here, developments in computer applications proved of great assistance to the engineers concerned with strength and flow problems. On the one hand, the blade angles and profiles had to be determined with the aid of a th~imensional flow cal-culation; on the other hand, the aptitude of thc twisted last-stage blades had to bc proved under service condi-tions.

The cost of producing thc last-stage blades for large turbines is high. Therefore, the turbine manufacturers endeavour to market a product which will perform its duties without any trbuble for many years.

Comparing new designs of last-stage blades with those of the past, it is strikingly evident that mechanically sound designs today dispcnsc with forms of "aids to survival".

Damping wires and in some cases cover strips are now a thing of the past for large turbines running at constant speed.

Apart from reducing costs and obtaining a better efficiency by this means, the machine is also made more reliable because thc unsupported blade is mounted under very definite conditions which makes calculation simpler.

Furthermore, with the methods of measuring now avail-able thc results obtained by calculation can easily be checked and calibrated in service.

Static Stresses of the Blade The cross-secuon of a blade varies considerably from bottom to top, the main axis of inertia of thc individual cross-sections being twisted from one to another as can be seen in Fig. 1 and 2.

The path taken by development can, be seen in Fig. 3 where a blade used forty years ago fora speed of 1200 rev/

min, which was slightly tapered and hardly twisted at all, is compared with a modern type ofblade from a 400 MW (3000 rev/min) turbine. What is most striking is thc differ-ence in the shape of the cross-sections along thc radius in the two designs.

Since the centrifugal stress az (see Table I) is responsible for the greater part of the total stress, even in twisted blades (Fig. 4), it may be adopted as a rough guide to the cross-section of the blade. Using the notation from Fig. 4 we then obtain:

Differential centrifugal force r

dK qrco~ F(r) dr

"I 4C NPy Islil I

~ 4 4

• I

,,4 0

V'4*

gE cc 4

g j,sf rle

'. P C ~t

+act l

+4t 44

~

sc$V,+4 Local blade stress Rf fNttta F(r) dr Rs F(R)

(2)

IfF(r) ~ F is constant, as was approximately the case in the older blades illustrated in Fig. 3, the relationship be-tween the tension due to centrifugal force in terms of the radius is given by ntoa az = (Rt' ra) 2 (3)

The stress in'he blade in this case increases quadratically from the tip to the base and attains its maximum value in the transition from blade to root (Fig.5). With this shaping thc designer has not made the best use of the materials and the attainable peripheral speeds therefore remain considerably below those of tapered blades. Ifon the other hand, an attempt is made to keep the tensile stress due to centrifugal force constant over thc grcatcr part of the length of the blade by differentiating equation (2) dF(r) ntns rdr F(r) az the following solution is obtained:

(R4s +

tte4 F(r) ~ Fte (5)

~

Modern last-stage blades have a

cross-section which roughly complies with equation (5) (see Fig. 4): az is al-most constant along the length of the blade. From equa-tion (5) it is apparent that the variation in cross-section of the blade is only dependent on the matertal chosen (n, az),

the speed (nt) and the geometry (R', r). This situation is illustrated in Fig.6 for the blade according'to Fig.2 made from three differen materials.

In service,

however, these blades are also subjected to other stresses besides az (Table I):

~Ieccl seve%

144%N Fig. I - Last-stage blades of a 600 MW turbine in the assetnbled state Due to inaccuracy in manufacture or duc to deliberate deviation of the linc connecting the centres of gravity of the various cross-sections from the radial, the blade is subjected to bending due to centrifugal force (as), which may be added to or subtracted from az according to its sign.

In the older designs (cylindrical blades) az and att are the sole stress components produced in the blade by rotation.

Consequently, calculation of the stresses duc to rotation is easy for such blades and can be readily analysed.

- In modern, long, last-stage blades, however, apart from the change in profile down the length of thc blade, the

Table I: Stresses in the t last-stage blades of turbines in service Type of stress Cause Static 1 Constant tension due to centrifugal force trz Centrifugal force produced by the blade mass situated above the given cross section 2 Flexural stress due to centrifugal force urn The departure: of the line joining the centre of gravity of the sections.rom the radial 3 Untwisting normal stress tra 4 Untwisting sheer stress rg 5 Flexural stress due to steam force trn The twisting of the blade due to centrifugal force The twisting of the blade due to centrifugal force Steam force acting on the blade Dynamic 6 Alternating flexural stresses aw Steam flow deviations from the preceding stationary blading, detachment, asymmetry (disturbances) in the design (at the horizontal joint), disturbing internals such as probes, critical speeds, short circuit at the generator Fig.2 - Shape of profile and velocity triangles of a last~go blade l000 mm long U>> Peripheral speed Ci >> Absolute velocity of the steam entering the blade Fig.3 - Comparison between lastwtage blades as made in l930 (n

l200 rev/min) and in l965 (n 3000 rev/min)

U I

U>> 605 m/s I.

I I

I

~465 I

I I

I I

322 C

a CI CI ta A

t r

IV'1 ia ai 1

I

>>c!

4*

EI'aowN cove%

isoess I

~MIll ~

individual cross-sections are subjected to successive twist in order to allow for thc change in peripheral speed over the height ofthe duct. Due to rotation ofthese blades, two additional stress components occur: the norinal stress trR and the sheer stress rlt due to the blade untwisting. A helpful model which shows how these stresses are pro-duced can be seen in Fig. 7. This shows that when the blade untwists under the effect ofa centrifugal force com-pression stresses arc produced in the outer sections while thc middle section is subjected to tension and torsion.

- Apart from these stress components produced by rota-tion the blade is also subjected to the forces produced by the fiowing medium. Here a distinction must be made be-tween the static component and a dynamic component of stress aw. Table I provides information about the causes leading to these stresses.

It is quite evident that for calculation ofthe stresses of the twisted tapered blade it is essential to use computers be-cause the stresses I to 5 in Table I at different points on the edges of the profile have to be calculated for differen cross-sections.

Fig. 4 shows thc result of such a stress cal-Fig. 6 - Distribution of the combined stress across the blade, showing the relationship bctwecn thc sum ev of all stress components according to Table l to the maximum value evws*

Ft ~ Reference a~ion (sce ot. Sl Rt ~ Hub radius Rs Radius at the tip R'

Radius ol'he reference cross-section r

~ Coordinates ev Rcfcrcnce stress-as ~ Tension due to centrifugal foNe Ft

~37 culation. Therein the rcfcrence stress trv ~ sum of the'tress components, was. determined according to the sheer stress hypothesis [I].

For practical applications it is extremely important to check the results of the stress calculations by random measurements, because when more is known about the stresses it is possible for optimum utilization of thc material to bc achieved.

Here the following checks are possible.

The stresses in the rotating blades are measured when, the bladed rotor is overspeed tested.

Groups of strain gauges are attached to the blade; the readings usually being transmitted to recorders by a system of sliprings.

It is, of course, possible that the actual measurement of the blade stress cannot be undertaken because no means of transmitting the measurement can be attached to the rotating rotor. Usually the employment of a slipring sys-tem to transmit thc measurement requires drillings in the rotor body for the leads, which in turn results in unde-sired stress concentration. The use of a telemetry system also imposes certain restrictions on the geometry of the rotor which have to be taken into account when it is designed. Therefore, when direct stress measurement is not possible, for the reasons given, the blade fitted with strain gauges may be run up to overspeed in stages and after each run examined at standstill to check for local exceed-ing of the yield point by measuring thc change in the electrical resistance ofthe strain gauge compared with the initial value. However, since this is only a means ofcali-brating the calculation, it is immaterial whether the blade consists oflow-alloy annealed material.

A further check is to measure thc angle of plastic un-twist. Whereas long last-stage blades untwist elastically. in service by between 5 to 8', under experimental conditions the blades can be brought to such peripheral speeds that plastic untwist occurs to an extent that can be measured.

By extrapolation to zero plastification the speed can bc determined at which the blade "stilljust" remains elastic.

Thc speed determined in this way at which plastification begins represents the upper limitat which the test blade may be used and from the relationship

+v/+Vtnss 0,7 0,76 trv is also applicable to blades having the same geometry but other strength values.

Attachment of the Blades to the Shaft 0.62 0,9

~ 0,2 0,

rg 0,

et g

oc 8

0 05 OY/dv ress I

0,55 I

>~~~anowN aovrte leeess t

The blades of the last stage of large turbines develop centrifugal forces ofsome hundreds oftons when running.

For this reason only very efficien methods ofattachment can bc considered. Among the systems in use at present, such as rhombus fixing in a peripheral slot, finger-shaped bolted fixing, straight or curved fir-tree roots, the last mentioned is an ideal means of attachment because it permits very close staggering ofthe blade cascade and the centrifugal force is produced in an optimum manner in thc shaft teeth. This design is illustrated in.Fig. l.

For reliability considerations it is essential to know the exact limits of the selected method of attachment; there-fore, in addition to calculations, photoelastic investiga-

0,5 az e<z 0,4 0.3 Ft 0,2 CI CI O,I 0.5 i

I I

I I

3 2

l 0

F/Ft I

aaowu eovsnt 100000

~ I Fig. 6-Variation ol'tions of a blade given by equation (5) l'or ditfcrent types of material A ~ Steel B ~ Titaniutn C

Fibre-reinforced plastic F'romcction Rs WKIWNCOVtN 100000 I Fig. 5 - Stress distribution in a hst-stage section 0

~ SpeciAc tnass of the blade material Us Peripheral speed at blade tip Other notation see Fig. 4.

blade with constant cross-Fig.y-Model to explain unwind-ing stress A

B ~

P a itf ~

Zl W

3 Blade tip Fixing Pressure Torsion Tension Leading edge Central clement ofthecross-scction Trailing ed ge 10000r I

tions and pullout tests on dummy blades are normally

'erformed.

Fig. 8 is a schematic illustration ofsuch a test bed on which Brown Boveri perform full-scale tests up to a force of 2000 t. In thc course of these tests not only the curve of deformation against force was plotted but also the notch stress at the bottom of the indentation in the fir-tree root. Using strain gauges with a grid of 04 mm local deformation was determined exactly to within a few per cent.

Effective Vibration 0 0 Oo tcoosa I Fig. Sa - Put&out test on a bent tir-tree root, with deformation diagram A ~ Working point (rated speed)

B Fracture P ~ Tension S ~ Clearance at thc bottom of thc groove, varying with tension The last-stage blade is subjected to forced vibration when running; the sources of disturbance are listed in Table L Thc magnitude ofthe forces acting on thc blade in scrvicc is, however, largely unknown. Consequently, the results of calculations of the alternating stresses caused by the forced vibrations are open to considerable doubt. For this reason it has become normal practice to judge the me-chanical quality of the blades according to the magnitude of the static stresses I to 5 (Table I) and the natural fre-quencies ofthc blade in relation to the exciting frequencies (multiples of the speed). With the aids to calculation that are availablc today the lower natural frequencies of the blade can bc calculated sufllcientiy accurately to avoid Fig. Sb - Rotor segment used for puuwut test Fig. Sc - Blade root indentations sheered otr in pWlwut test IttttISIitI8 sr "

4 "t

IIIIIII OIOWIIevce

resonance with possibl

'llating steam forces. Here, though, the following facts must be borne in mind.

A decisive factor for assessing the vibration behaviour of the blade is its frequencies at operating speed. As can be seen in Fig. 9, the centrifugal force has a stiffening effect on the blade, with the result that the natural fre-quency increases with the speed of rotation. This rise is different for the various orders and depends on the shape of the oscillation. The stifiening effect is greater with flexural than with torsional vibration of the blade (see nodal lines in Fig. 9).

- Differences ilgwu material quality and tolerable devia-tions in geomet~tof individual blades result in a scatter band at each order of natural frequency.

Simpliftcations and approximations which have" to be taken into account when setting up a model forcalculation result in discrepancies from reality.

For these reasons the calculation has to be recalibrated for the development of new blades whose shape differs from that of existing designs. The procedure adopted is roughly as follows:

Having calculated the first natural frequencies in terms of Fig.9-Natural l'requencies F of thc blade plotted against speed n

(rev/ min)

Nodal lines at the llrst four natural l'requencies.

A.B. C,D Permhted scatter bands of frequencies for zero speed measurement

a. b, c,...

Permincd scatter bands ol'blade frequencies at operating speed E ~ Spccd range in which the tbdng rigidityis inQuenced by centrifugal lorco 6

Usual scauer band (precision forged blade)

H Forbidden l'requency range for measurements at z<<ro and normal speed nI ~ Speed in rev/s

~

Values measured for zero speed vibration Values measured at dilfcrcnt speeds Fig. l0- Arrangement ofthe telemetry system for measurement ofvibra-tion on last stage blades in service D

Strain gauge on the blade lY ~ Shah

$ ~ Transmitter A

Pick-up ring 400

/

v/

/

I 9ns gus 300 7/ls

)

511s 200 Df3 H'

3 Its H l00 281 z

n l000 2000 3000 tMCQO ~ 1 eaoWN aovznl 100100 I

O s peed for blades with the desired dimensions and material qualities, each manufactured blade is checked at zero speed.

By measuring the natural frequencies of some blades during overspeed testing it is possible to check whether the rise in frequency as a function of thc speed was calculated correctly. Here the blade is made to vibrate by disturbance forces (excitation plates) many times larger than the disturbance forces actually experienced inservice.

On completion ofthis test which has to be carried out once forevery prototype last-stage blade, an economic selection ofthe manufactured blades can be performed by checking at zero speed (rr = 0) alone. In Fig. 9 the frequency ranges a, b, c... arc permissible at scrvicc speed, the range H is forbidden. Thus the permitted frequency ranges A, B, C at zero speed are fixetL To a limited extent completed blades whose routine zero speed measurement produces natural frequency values outside the ranges A, B, C can be brought inside these ranges by subsequent machining within the dimensional tolerance.

However, it must be remembered that this subsequent machining changes all the natural frequencies.

The series of tests is concluded by measurements in service. In 1968 Brown Boveri checked a 600 MWmachine by telemetry and was able to establish the dynamic behav-iour of last-stage blades 1000 mm long throughout the entire load range. The main obstacles were the develop-ment of a watertight and erosion proof means ofsticking and covering the strain gauges and the development of electronic equipment capable of withstanding centrifugal

'ccelerations up to 7000 g and temperatures up to 150 'C for long periods.

Fig. 10 shows the measuring set up.

These measurements were repeated successfully in 1969 and 1971 on a 300 MWmachine in which the steam had a high moisture content. Such measurements are nowadays desirable for various reasons and arc gaining in signifi-cance because:

Fig. 11 shows the arrangement ofthe strain gauges on one.

of the blades examined.. It was important to attach the strain gauges at points where there was a relatively high amplitude ofvibration in order that the result ofmeasure-ment at such points could be compared with the results of calculations and so that conclusions could bc drawn regarding the maximum stresses to which the blade was subjected.

Fig. II - Strain gauges No. 5l, 52. 53. 6I, 62 and 63 on the rear ofa test blade I'or measuring natural frequencies in service 52.62 63 5I,6I Little is known about the stimulation which causes the blade to vibrate in scrvicc.

- Thc length of the blades in thc last stages of steam turbines has been increased in recent years by all manu-facturers and willcontinue to increase as unit outputs are raised. This willmake the blades flexumllysofter and they will respond to external influences by more pronounced vibration.

- In many places today thc use of river water for cooling the condensers is no longer permitted. The usc of cooling

towers, however, results in warmer cooling water and consequently a higher pressure in the condenser compared with the fresh water cooling. Therefore, it is necessary to check whether the forces to which the final-stage blades are exposed as a result of the higher exhaust pressure do not represent an unreasonable strain.

During such measurements thc following operational conditions were examined:

- While the machine was running up, the resonance ofthe blades (damping) was tested. The vacuum was varied be-tween 30 and 250 mbar.

- The vibration of the final-stage blade was examined at dilferent loads and exhaust pressures up to 250 mbar.

By means ofshutdown tests with partial and fullvacuum breakage the behaviour of the finalwtage blades was also tested under these abnormal conditions.

te4141 I

The results of these mea.

ents can be summarized as follows:

The blades are sufficiently proof against vibration frac-ture but only when the natural frequencies are not mul-tiples of the speed of rotation.

The aerodynamic excitation forces are very small pro-vided steps are taken to avoid obvious sources of distur-bance when designing and manufacturing the turbine.

- The excitation forces resulting from errors in the pitch ofthe stationary blade segments or ofthe actual stationary blades themselves, occur at such high frequencies that they do not represent a direct hazard for the blades.

- Ifthe above requirements are taken into account, the maximum alternating stresses throughout the entire opera-tional range and at the exhaust pressures in use nowadays are only a fraction ofthe strength ofthe blade material.

- At high ex ressures the amplitudes of vibration may be expected to increase. They are highest at no load, because here the aerodynamic conditions are unfavour-able for the blades. For that point on the blade which is most severely

stressed, alternating stresses may occur which reach such a high level in relation to the fatigue strength of the blade material, even in good designs, that they can no longer be ignored. Fig. 12 gives some idea of the vibration of the blades during a shutdown with full vacuum breakage.

Therein the increase in the exhaust pressure p against time f can be seen, also the resultant drop in speed rt of the shaft due to increased ventilation, the curve of the temperature T in the exhaust area as well as the curve of the natural frequency F and the amplitude A of the blade vibration at point 63 in Fig. 11.

Fixing and Damping 20 Hx 3000 lg I60 I

I I

I I

I I

I iso I

I r

20 min 600 mm Hg 400 Fig. 12-Measuremcnt ofblade vibration during run-out ofthe machine followingfull vacuum breakage F

Natural I'requency of the test blade A

Amplitude measured by strain gauge No.63 in FIg. II at fre-quency F T ~ Temperature in exhaust region n

Speed of the turbine shaA'

~ Pressure in exhaust region r ~ Tltnc In the discussion of blade vibration the question often arises regarding the influence of the flexibilityof the blade fixing and of the damping on the vibration of the blade.

Owing to the enormous centrifugal forces acting on the blade in service, amounting to some hundreds oftons, the contact surfaces between the blade and the shaft are pressed against one another so strongly that not the slightest movement can occur at these points and therefore there is no variation in the natural frequencies ofthe blade which may be regarded as rigidly mounted. However, the indentations in the shaft and the foot of the blade have their own ehsticity values which diverge from the rigid, ideal case. As can be seen in Fig. 13, this influence on the natural frequency of the blade is negligibly small, because current designs possess rigidity values which come fairly close to the absolutely rigid fixing [2].

The damping is composed of components which depend on the method ofroot fixing, the ambient medium and the material from which the blade is made. In practice this damping is measured by recording the logarithmic decre-ment. As Fig. 14 shows, the damping changes with the amplitude of the alternating stress. At stress values which can in fact occur in turbines, it reaches an order of mag-nitude at which a distinct change in the resonant frequency can be detected.

I IOO Al 40 I IO imm/mm I000

[Qpq y lI lee les I 200 Erosion The last stages of large condensing turbines operate in the wet steam region, where the steam contains 5 to 12%

moisture.

Mainly responsible for erosion is the water which separates out in the outermost third of the last row of stationary blades. Drop by drop this water is tom off

895 Q

mVl Ch I

I t

I I

I

/

//

II IO 9

8 6'

D>>0 D<<0,02 D<<0,05 D <<O.l D>>O.I5 I500 Hz P

O,l 0,9 0,95 I,o I,05 Iooo D

500 0,05 0iO'O~'

iO'O~

IO' 10 I01 lenses I

O,OI 0

I00, 200 300 400 X

leo los I Fig. I3-Inliuence of the elasticity oi'he blade ibung on the natural fre.

qucncy A>> Fixing in the half plane hf RigidityofAxing Pt <<<< IO'kpcm Young's modulus E <<2 I. Io' picmc hf 8

Theoretical equivalent Axing model Ps rn <<Rigid equivalent Axing with aero mass n << Inclination ac point of Axing due to hf C <<Curve of frequencies P againsc Axing rigidity Pt, Pc, Ps <<Natural frequencies of Arst to third order D <<Rigidityofcommon kinds oi'blade Axings Fig. I4-Incremental function and damping of a vibrating blade y << Incremental function rr <<Frequency ratio: Exciting frequency to resonanc frequency D <<Lehr's coeAicient ofdamping 2nD Logarithmic dccremenc d (ID X

Amplitude ofalternating stress (kp/cmt)

A Displacement or thc peak frequency for y

[y (r/)1., due to damping 10-

thc trailing edge of the stary blades and accelerated by thc flow of steam. Thc Krops, which may be up to 0 2 mm across, reach velocities in the space between the stationary and moving blades which differ considerably from those of the steam flow [3). This implies that there is a relatively large difference between the peripheral speed of the bhde tip and the peripheral component of droplet

speed, as a result of which the droplets strike the leading edge of the blade with an abrasive efl'ect known as erosion.

Turbine manufacturers protect blades against erosion by armouring the leading edge. This can be done either by hardening the basic material (the method adopted by Brown Boveri) or by soldering on plates of Stellite.

The total amount of material lost by erosion is a loga-rithmic function of time; whereas the erosion rate is high during-the initial period of service, it almost ceases after about one year [4). An explanation for this is that the pores in the surface of the blade where erosion has taken place are partly fllled with water, so that the impact forces of the drops striking the blades are only transferred to the material in a damped form (Fig. 15).

Future Pros s

It is the size of modern power station turbines which is their most impressive feature. Machines with unit ratings ofover 1000 MWare being built; they have a total length ofabout 70 m, the diameter of the rotor measured across the tips of the final-stage blade varies between 45 and 5 5 m, depending on the manufacturer, while the casing surrounding the low-pressure rotor is almost as big as a private house. Since the trend toward further increases in unit capacities is continuing steps are already being taken to develop last-stage blades to even larger sizes in order to cope with the enormous steam volumesin a 1100 MW machine the amount'is about 30000 m'/s with a condenser vacuum of 0 05 barin a reasonable number of flows.

Here the stresses to which the rotor is subjected are of particular signiflcance. Since the rotor discs are made of material whose yield point cannot be extended much further, an attempt is made to enlarge the outlet area by reducing thc speed of the rotor and by employing a suit-able material for the last-stage blades.

Since, in current

designs, the centrifugal forces of the last-stage blades Fig. tS - Eroded leading edge ol'a blade MagniAeation 20x Fig. l6-Carbon Abres before being inserted In the plastic matrix MagniAeatton 4SOx

produce stresses in the section of the rotor, which may amount to 35% ofthe total rotor stress, the gcncral trend is towards blade materials with higher strengths but a lower specific weight. Foremost among such, materials is titanium.

In recent years, however, reinforced materials (plastics) have found new fields of application.

Among them, plastics reinforced with boron and carbon fibres exhibit properties which are quite equal to those of a high-alloy steel. The plastic blade therefore has a certain chance of being employed at thc cold end of power station turbines, provided the very severe problem of erosion can be over-come. Table II shows the properties of a material of this kind. Thc carbon fibres embedded in the matrix have a diameter ofabout 5 IO'm. Fig. 16 and 17 respectively show such fibres before being inserted in the plastic matrix and the surface of a fracture through such a composite bar.

The large steam turbines being built today are normally employed as base-load machines and they are expected to obtain a very high availability. As it is understandable that the indkvidual elements have to be carefully examined Table II: Properties of a plastic material reinforced with carbon fibres Matrix:

Fibre:

Density Tensile strength Young's modulus Heat transfer coclficient Coclficient of thermal expansion Epoxy resin Carbon (60% by volume) 16 Longitudinal 7500 Longitudinal 2 4

..10'ongitudinal 34 Transverse 2 9 Longitudinal Transverse

07. 10 e

28. IO'e g/cma kp/cm'p/ctn'/m

'C W/m 'C

/'C

/Oe even in the design stage, we make every effort to utilize.

modern computer and test facilitics as far as possible. The aim of all these efiorts is to ensure that future large ma-chines with ratings above 1000 MWwillbe just as reliable and compact. In the attainment of this aim size and quality ofthe last-stage blades plays an important part.

Fig. l7 - Fractured surface of a compound materhl employing carbon abte reinforcement Mattniacation lI00 tc Bibliography

[I] J. Montoya Garcfat Coupled bending and torsional vibrations in a twisted, rotating blade. Brown Boveri Rev.

1966 $3 (3) 216-230.

[2] F. Vagtt Obcr die Berechnung der Fundamcntdefor-mation. Published by Norske Videnskaps Aka'demi, 1925.

4.

1 V~"-

4I

[3] G. Gyarmathyt Grundlage einer Theoric der Nass-dampfturbine. Mitteilung Nr.6 published by the Insti-tute forThermal Turbo-Machines, Swiss Federal Institute ofTechnology, Zurich.

[4] J. Hossli: Problems in the construction ofturbines for nuclear power plants. Brown Boveri Publication 3340 E (1967).

[5] A. Hohn, P. ¹vacekt Dic Endschaufeln grosser Dampfturbinen aus mechanischer Sicht. Schweiz. Bauztg 1970 BB (30) 673-678.

BBC BROWN BOVERI BBC Brown, Bovefl g Company. Ltd., CH-5401 Baden/Switzerland printed In Swltrertand frtl2-7504I Ctasaittcauon No.01 01