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: ML17318A690 ML17318A691 ML17318A693 ML17318A706 ML17318A707
References
CH-T-060040-E, CH-T-60040-E, NUDOCS 8004220034
Download: ML17318A706 (12)

Text

Last-Ste Blades Ilute BROWN BOVERI of Large Steam Turbines Publication No. CM-T 060040 E A. Ifohn and P. ¹vacek determined with the aid of a th~imensional flow cal-culation; on the other hand, the aptitude of thc twisted The present article deals with the blades ih the last rotating last-stage blades had to bc proved under service condi-row in large steam turbines, consfdering them as a machine tions.

element. The static and dynamic stresses occurring in The cost of producing thc last-stage blades for large service are discussed and their effect on the design of the turbines is high. Therefore, the turbine manufacturers blades is demonstrated. Some methods of testfng which are endeavour to market a product which will perform its used in the design ofprototypes are explained as they enable duties without any trbuble for many years.

blades designed on pure theory to be tested under conditions Comparing new designs of last-stage blades with those of comparable with those experienced"in service, thereby the past, it is strikingly evident that mechanically sound enabling the behaviour of the blades in service to be pre- designs today dispcnsc with forms of "aids to survival".

dicted. ¹wadays thfs performance ls from time to time Damping wires and in some cases cover strips are now a checked ln servfce fn power stations; the article provides thing of the past for large turbines running at constant some information regarding test pr ocedures and the results speed.

obtained. In conclusion the authors discuss future develop- Apart from reducing costs and obtaining a better efficiency ments ln blade eonstructfon. by this means, the machine is also made more reliable because thc unsupported blade is mounted under very definite conditions which makes calculation simpler.

Introduction Furthermore, with the methods of measuring now avail-able thc results obtained by calculation can easily be A feature of the last ten years in the construction of steam checked and calibrated in service.

turbines was the marked risc in the unit ratings of ma-

chines. At the beginningof the 1960's the majority of large thermal machines in European installations were mainly Static Stresses of the Blade in the 125 to 150 MW range. Today, both in Europe and in America, machines with an output of morc than The cross-secuon of a blade varies considerably from 1000 MW are being installed. In them the volume of bottom to top, the main axis of inertia of thc individual steam that has to bc handled on emerging from the blading cross-sections being twisted from one to another as can (last-stage) may be of the order of magnitude of 10000m'/s be seen in Fig. 1 and 2.

(fossil fuelled plants) up to 25000 ms/s (in nuclear power The path taken by development can, be seen in Fig. 3 plants with turbines employing saturated steam). In. order where a blade used forty years ago for a speed of 1200 rev/

to handle such enormous quantities of steam in a reason- min, which was slightly tapered and hardly twisted at all, able number of flows, the cross-section of flow in the blade is compared with a modern type of blade from a 400 MW ducts has to be large and the last stage correspondingly (3000 rev/min) turbine. What is most striking is thc differ-long. On the other hand, in thc last row of blades of large ence in the shape of the cross-sections along thc radius in steam turbines about 6% of the total heat drop of the the two designs.

steam flowing through the turbine is converted into me- Since the centrifugal stress az (see Table I) is responsible chanical energy. Since these two factors highoutputand for the greater part of the total stress, even in twisted quality of the energy conversion are also influenced by blades (Fig. 4), it may be adopted as a rough guide to the the last stage, particular attention has been paid to these cross-section of the blade. Using the notation from Fig. 4 blades during the past ten years. Here, developments in we then obtain:

computer applications proved of great assistance to the Differential centrifugal force r engineers concerned with strength and flow problems. On the one hand, the blade angles and profiles had to be dK qrco~ F(r) dr

0 Local blade stress R

f fNttta F(r) dr

'. P Rs (2)

F(R)

IfF(r) ~ F is constant, as was approximately the case in "I the older blades illustrated in Fig. 3, the relationship be-4C C ~t tween the tension due to centrifugal force in terms of the radius is given by az =

ntoa 2

(Rt' ra) (3)

NPy 4 cc +act l The stress in'he blade in this case increases quadratically from the tip to the base and attains its maximum value in

+4 t 44 ~

the transition from blade to root (Fig.5). With this g 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. If on the other hand, an attempt is made to keep the tensile sc$V,+ 4 stress due to centrifugal force constant over thc grcatcr j,sf part of the length of the blade by differentiating equation rle V '4 (2)

Isl

• il dF(r) ntns rdr I

gE F(r) az

~ 4 the following solution is obtained:

F(r) ~ Fte

+

tte4 (R4s (5) ~

4 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-I 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

,,4 made from three differen materials.

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

Due to inaccuracy in manufacture or duc to deliberate

~ Ieccl seve% 144%N deviation of the linc connecting the centres of gravity of Fig. I - Last-stage blades of a 600 MW turbine in the assetnbled state 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 produced by the blade mass situated centrifugal force trz above the given cross section 2 Flexural stress due to The departure: of the line joining the centre of gravity of centrifugal force urn the sections .rom the radial 3 Untwisting normal stress tra The twisting of the blade due to centrifugal force 4 Untwisting sheer stress rg The twisting of the blade due to centrifugal force 5 Flexural stress due to steam Steam force acting on the blade force trn 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.3 - Comparison between lastwtage blades as made in l930 (n l 200 rev/min) and in l 965 (n 3000 rev/min)

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 U

I U>> 605 m/s IV'1 I . ta ia ai I

C

~465 a CI 1 CI I

I A

I t >>c!

r I 4*

I I I I 322 EI'aowN cove% i soess I

~ M Ill ~

individual cross-sections are subjected to successive twist culation. Therein the rcfcrence stress trv ~ sum of in order to allow for thc change in peripheral speed over components, was. determined according to the the'tress the height of the duct. Due to rotation of these blades, two sheer stress hypothesis [I ].

additional stress components occur: the norinal stress trR For practical applications it is extremely important to and the sheer stress rlt due to the blade untwisting. A check the results of the stress calculations by random helpful model which shows how these stresses are pro- measurements, because when more is known about the duced can be seen in Fig. 7. This shows that when the stresses it is possible for optimum utilization of thc blade untwists under the effect of a centrifugal force com- material to bc achieved. Here the following checks are pression stresses arc produced in the outer sections while possible.

thc middle section is subjected to tension and torsion.

- Apart from these stress components produced by rota- The stresses in the rotating blades are measured when, tion the blade is also subjected to the forces produced by the bladed rotor is overspeed tested. Groups of strain the fiowing medium. Here a distinction must be made be- gauges are attached to the blade; the readings usually tween the static component and a dynamic component of being transmitted to recorders by a system of sliprings.

stress aw. Table I provides information about the causes It is, of course, possible that the actual measurement of leading to these stresses. the blade stress cannot be undertaken because no means of transmitting the measurement can be attached to the It is quite evident that for calculation of the stresses of the rotating rotor. Usually the employment of a slipring sys-twisted tapered blade it is essential to use computers be- tem to transmit thc measurement requires drillings in the cause the stresses I to 5 in Table I at different points on rotor body for the leads, which in turn results in unde-the edges of the profile have to be calculated for differen sired stress concentration. The use of a telemetry system cross-sections. Fig. 4 shows thc result of such a stress cal- also imposes certain restrictions on the geometry of the rotor which have to be taken into account when it is d esigned. 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 Fig. 6 Distribution of the combined stress across the blade, showing electrical resistance of the strain gauge compared with the the relationship bctwecn thc sum ev of all stress components according to Table l to the maximum value evws* initial value. However, since this is only a means of cali-Ft ~ Reference a~ion (sce ot. Sl brating the calculation, it is immaterial whether the blade Rt ~ Hub radius consists of low-alloy annealed material.

Rs Radius at the tip A further check is to measure thc angle of plastic un-R' Radius ol'he reference cross-section r ~ Coordinates twist. Whereas long last-stage blades untwist elastically. in ev Rcfcrcnce stress- service by between 5 to 8', under experimental conditions as ~ Tension due to centrifugal foNe 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 Ft begins represents the upper limit at which the test blade

~ 37 may be used and from the relationship trv is also applicable to blades having the same geometry but other strength values.

/

+v +Vtnss 0,7 Attachment of the Blades to the Shaft 0,76 0.62 The blades of the last stage of large turbines develop centrifugal forces of some hundreds of tons when running.

0,9 For this reason only very efficien methods of attachment 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 of the blade cascade and the 0 05 centrifugal force is produced in an optimum manner in thc OY /dv ress shaft teeth. This design is illustrated in.Fig. l.

I For reliability considerations it is essential to know the et oc ~ 0,2 0, rg 0, 8 0,55 exact limits of the selected method of attachment; there-g I

>~~~anowN aovrte leeess t fore, in addition to calculations, photoelastic investiga-

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

I I

I 3 2 l 0 F/Ft I aaowu eovsnt 100000 ~ I 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 WKIWN COVtN 100000 I Fig. 5 - Stress distribution in a hst-stage blade with constant cross-section 0 ~ SpeciAc tnass of the blade material Us Peripheral speed at blade tip Other notation see Fig. 4.

Fig.y- Model to explain unwind-ing stress A Blade tip B ~ Fixing P a Pressure itf ~ Torsion Z Tension l Leading edge W Central clement of thecross-scction 3 Trailing ed ge 10000r I

tions and pullout tests on dummy blades are normally Fig. 8 is a schematic illustration of such a test 'erformed.

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 The last-stage blade is subjected to forced vibration when 0 0 running; the sources of disturbance are listed in Table L Thc magnitude of the forces acting on thc blade in scrvicc is, however, largely unknown. Consequently, the results of calculations of the alternating stresses caused by the Oo forced vibrations are open to considerable doubt. For this reason it has become normal practice to judge the me-tcoosa I chanical quality of the blades according to the magnitude

- Put&out test on a bent tir-tree root, with deformation of the static stresses I to 5 (Table I) and the natural fre-Fig. Sa diagram quencies of thc blade in relation to the exciting frequencies A ~ Working point (rated speed) (multiples of the speed). With the aids to calculation that B Fracture P ~ Tension are availablc today the lower natural frequencies of the S ~ Clearance at thc bottom of thc groove, varying with tension 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 IttttISI itI8 sr "

4 "t

IIIIIII OIOWI Ievce

resonance with possibl 'llating steam forces. Here, - Differences ilgwu material quality and tolerable devia-though, the following facts must be borne in mind. tions in geomet~tof individual blades result in a scatter band at each order of natural frequency.

A decisive factor for assessing the vibration behaviour Simpliftcations and approximations which have" to be of the blade is its frequencies at operating speed. As can taken into account when setting up a model for calculation be seen in Fig. 9, the centrifugal force has a stiffening result in discrepancies from reality.

effect on the blade, with the result that the natural fre-quency increases with the speed of rotation. This rise is For these reasons the calculation has to be recalibrated different for the various orders and depends on the shape for the development of new blades whose shape differs of the oscillation. The stifiening effect is greater with from that of existing designs. The procedure adopted is flexural than with torsional vibration of the blade (see roughly as follows:

nodal lines in Fig. 9). 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 Fig. l0- Arrangement of the telemetry system for measurement of vibra-measurement tion on last stage blades in service

a. b, c, ... Permincd scatter bands ol'blade frequencies at operating speed D Strain gauge on the blade E ~ Spccd range in which the tbdng rigidity is inQuenced by centrifugal lY ~ Shah lorco $ ~ Transmitter 6 Usual scauer band (precision forged blade) A Pick-up ring 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

/

v/

9ns

/

400 I gus 7/ls 300 )

511s Df3 200 H' 3 Its H

l00 281 z n

l 000 2000 3000 tMCQO ~ 1 eaoWN aovznl 100100 I

s peed O

for blades with the desired dimensions and material Fig. 11 shows the arrangement of the strain gauges on one.

qualities, each manufactured blade is checked at zero of the blades examined.. It was important to attach the speed. By measuring the natural frequencies of some strain gauges at points where there was a relatively high blades during overspeed testing it is possible to check amplitude of vibration in order that the result of measure-whether the rise in frequency as a function of thc speed ment at such points could be compared with the results of was calculated correctly. Here the blade is made to vibrate calculations and so that conclusions could bc drawn by disturbance forces (excitation plates) many times larger regarding the maximum stresses to which the blade was than the disturbance forces actually experienced in service. subjected.

On completion of this test which has to be carried out once for every prototype last-stage blade, an economic selection of the 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 -

Fig. I I Strain gauges No. 5l, 52. 53. 6I, 62 and 63 on the rear of a test service. In 1968 Brown Boveri checked a 600 MW machine blade I'or measuring natural frequencies in service 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 of sticking 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 52.62 and 1971 on a 300 MW machine in which the steam had a high moisture content. Such measurements are nowadays desirable for various reasons and arc gaining in signifi- 63 cance because: 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 will continue to increase as unit outputs are raised. This will make the blades flexumlly softer 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 of the 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 of shutdown tests with partial and full vacuum 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 - At high ex ressures the amplitudes of vibration follows: may be expected to increase. They are highest at no load, because here the aerodynamic conditions are unfavour-The blades are sufficiently proof against vibration frac- able for the blades. For that point on the blade which is ture but only when the natural frequencies are not mul- most severely stressed, alternating stresses may occur tiples of the speed of rotation. which reach such a high level in relation to the fatigue The aerodynamic excitation forces are very small pro- strength of the blade material, even in good designs, that vided steps are taken to avoid obvious sources of distur- they can no longer be ignored. Fig. 12 gives some idea of bance when designing and manufacturing the turbine. the vibration of the blades during a shutdown with full

- The excitation forces resulting from errors in the pitch vacuum breakage. Therein the increase in the exhaust of the stationary blade segments or of the actual stationary pressure p against time f can be seen, also the resultant blades themselves, occur at such high frequencies that they drop in speed rt of the shaft due to increased ventilation, do not represent a direct hazard for the blades. the curve of the temperature T in the exhaust area as well

- If the above requirements are taken into account, the as the curve of the natural frequency F and the amplitude maximum alternating stresses throughout the entire opera- A of the blade vibration at point 63 in Fig. 11.

tional range and at the exhaust pressures in use nowadays are only a fraction of the strength of the blade material.

Fixing and Damping 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 of tons, 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 of the blade which may be regarded as rigidly mounted. However, the Fig. 12- Measuremcnt ofblade vibration during run-out of the machine indentations in the shaft and the foot of the blade have following full vacuum breakage their own ehsticity values which diverge from the rigid, F Natural I'requency of the test blade ideal case. As can be seen in Fig. 13, this influence on the A Amplitude measured by strain gauge No.63 in FIg. II at fre- natural frequency of the blade is negligibly small, because quency F current designs possess rigidity values which come fairly T~ Temperature in exhaust region n Speed of the turbine shaA' close to the absolutely rigid fixing [2].

~ Pressure in exhaust region The damping is composed of components which depend r ~ Tltnc on the method of root fixing, the ambient medium and the 3000 material from which the blade is made. In practice this 20 damping is measured by recording the logarithmic decre-Hx ment. As Fig. 14 shows, the damping changes with the amplitude of the alternating stress. At stress values which 600 can in fact occur in turbines, it reaches an order of mag-mm Hg nitude at which a distinct change in the resonant frequency I

can be detected.

lg I I 20 I

min I

I 400 I

I iso I

I60 I

r I IOO I 000 Al 200 Erosion 40 I IO imm/mm [Qpq y The last stages of large condensing turbines operate in the wet steam region, where the steam contains 5 to 12%

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

Q II IO D>>0 m I Vl 9 Ch I D<<0,02 895 t 8 I

I I 6'

/ D<<0,05

/ D <<O.l

/

D>>O.I5 0,9 0,95 I,o I,05 I 500 Hz P

O,l Iooo D 0,05 500 O,OI 0

iO'O~' iO'O~ IO' 10 I01 0 I 00, 200 300 400 lenses I X leo los I Fig. I3 - Inliuence of the elasticity oi'he blade ibung on the natural fre. Fig. I4- Incremental function and damping of a vibrating blade qucncy y << Incremental function A>> Fixing in the half plane rr <<Frequency ratio: Exciting frequency to resonanc frequency RigidityofAxing Pt <<

hf << IO'kpcm D <<Lehr's coeAicient of damping 2nD (I D Young's modulus E <<2 I. Io' picmc Logarithmic dccremenc d 8 Theoretical equivalent Axing model Ps hf X Amplitude of alternating stress (kp/cmt)

A Displacement or thc peak frequency for y [y (r/)1 ., due to rn <<Rigid equivalent Axing with aero mass damping 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 <<Rigidity of common kinds oi'blade Axings 10-

thc trailing edge of the stary blades and accelerated Future Pros s by thc flow of steam. Thc Krops, which may be up to 0 2 mm across, reach velocities in the space between the It is the size of modern power station turbines which is stationary and moving blades which differ considerably their most impressive feature. Machines with unit ratings from those of the steam flow [3). This implies that there of over 1000 MW are being built; they have a total length is a relatively large difference between the peripheral speed of about 70 m, the diameter of the rotor measured across of the bhde tip and the peripheral component of droplet the tips of the final-stage blade varies between 45 and speed, as a result of which the droplets strike the leading 5 5 m, depending on the manufacturer, while the casing edge of the blade with an abrasive efl'ect known as surrounding the low-pressure rotor is almost as big as a erosion. private house. Since the trend toward further increases in Turbine manufacturers protect blades against erosion by unit capacities is continuing steps are already being taken armouring the leading edge. This can be done either by to develop last-stage blades to even larger sizes in order to hardening the basic material (the method adopted by cope with the enormous steam volumes in a 1100 MW Brown Boveri) or by soldering on plates of Stellite. machine the amount'is about 30000 m'/s with a condenser The total amount of material lost by erosion is a loga-vacuum of 0 05 bar in a reasonable number of flows.

rithmic function of time; whereas the erosion rate is high Here the stresses to which the rotor is subjected are of during-the initial period of service, it almost ceases after particular signiflcance. Since the rotor discs are made of about one year [4). An explanation for this is that the material whose yield point cannot be extended much pores in the surface of the blade where erosion has taken further, an attempt is made to enlarge the outlet area by place are partly fllled with water, so that the impact forces reducing thc speed of the rotor and by employing a suit-of the drops striking the blades are only transferred to the able material for the last-stage blades. Since, in current material in a damped form (Fig. 15). designs, the centrifugal forces of the last-stage blades Fig. tS - Eroded leading edge ol'a blade Fig. l6- Carbon Abres before being inserted In the plastic matrix MagniAeation 20x MagniAeatton 4SOx

produce stresses in the section of the rotor, which may even in the design stage, we make every effort to utilize.

amount to 35% of the total rotor stress, the gcncral trend modern computer and test facilitics as far as possible. The is towards blade materials with higher strengths but a aim of all these efiorts is to ensure that future large ma-lower specific weight. Foremost among such, materials is chines with ratings above 1000 MW will be just as reliable titanium. and compact. In the attainment of this aim size and In recent years, however, reinforced materials (plastics) quality of the last-stage blades plays an important part.

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 Table II: Properties of a plastic material reinforced with steel. The plastic blade therefore has a certain chance of carbon fibres being employed at thc cold end of power station turbines, provided the very severe problem of erosion can be over- Matrix: Epoxy resin come. Table II shows the properties of a material of this Fibre: Carbon (60% by volume) kind. Thc carbon fibres embedded in the matrix have a diameter of about 5 IO'm. Fig. 16 and 17 respectively Density 16 g/cma show such fibres before being inserted in the plastic matrix Tensile strength Longitudinal 7500 and the surface of a fracture through such a composite Young's modulus Longitudinal 24 kp/cm'p/ctn'/m bar. Heat transfer 34 ..10'ongitudinal

'C The large steam turbines being built today are normally coclficient Transverse 29 W/m 'C employed as base-load machines and they are expected to Coclficient of 07.

obtain a very high availability. As it is understandable thermal Longitudinal 10 e /'C that the indkvidual elements have to be carefully examined expansion Transverse 28. IO'e /Oe Fig. l7 Fractured surface of a compound materhl employing carbon Bibliography abte reinforcement Mattniacation l I 00 tc [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.

[3] G. Gyarmathyt Grundlage einer Theoric der Nass-dampfturbine. Mitteilung Nr.6 published by the Insti-

4. V~"- 4I 1 tute for Thermal Turbo-Machines, Swiss Federal Institute of Technology, Zurich.

[4] J. Hossli: Problems in the construction of turbines 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