Raabe J. Hydro power - the design, use, and function of hydromechanical, hydraulic, and electrical еquipment

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10,5.2.

Influence

bearings on shaft vibration

10.5.2.1.

General remarks

The rotor of the turbine and generator and the shaft are excited usually

rotating forces

due

to mechanical, hydromechanical and electrical unbalance of the rotor masses. in

consequence of misalignment during assembly, imperfect balancing or short circuit. The

system of shaft and rotor masses, which is

elnstically supported and damped by

the bearings, responds to these excitations by shaft oscillations with finite amplitude of

different kinds

[lo.

11.

flexural vibrations as the most obvious type of vibrations are considered now separately. 1.e.. for

the generator to work properly, vibration must not accelerate to levels at which it misht damage the

components, or reach such an amplitude as to cause wear or wiping off the bearing's white metal.

fie rotary system must be properly insulated from the stationary one, so as to avoid subjecting

to excessive stresses. These differing requirements tend to be in contrast, as a bearing system, which

limits shaft vibration tends to put greater forces on the foundations and vice versa.

The basic element of damping and elasticity between rotor and the foundation is the

oil film in the

bearings. As already stated, the general way to limit vibration amplitude, is to design the machine

so, that its lowest critical speed is greater than any possible

operatirlg speed (runaway). Hydraulic

units have considerable shaft stiffness and the masses are limited by electrical, mechanical and

governing requirements.

So any method of increasing the critical speed depends on the possibility of reducing the elasticity

of the guide bearings. In a very stiff system small values of

the amplitude

shnit vibrations can

involve significant unbalanced stresses which can exert high forces on the foundations.

For these reasons the traditional criterion of creating a large difference between critical speed due

to shaft stiffness and the rotor masses and the working speed cannot be considered

an abscjlu~e

guarantee of safety.

is necessary to make

thourough analysis of the dynamic response of the rotor which takes into

account three parameters: unbalance, vibration amplitude and

the feature of the bcariugs. Sincc the

bearing box and its support are usually rather

stiff, as a minimum requirement of a good bearin:

design, the features of the bearings are reduced to that of the lubricating film in the gap bctwecn seat

and shaft.

10.5.2.2.

The spring rate

the film

lubricant

As a model consider two plane faces parallel to each other, at distance

with an area

The clearance is filled with a lubricant of bulk modulus

E,.,

the boundaries might be

completely leak proof. Expressing the spring-induced force

dy,

where

is the

spring rate and

the compression of clearance, and using the relation

(8.3-5)

as the

equation of state

the lubricant, the spring rate reads

Hence the stiffness of an oil film increases with decreasing

cleararlce

Thus an oil

fila

loaded by an unbalance, by which the oil is squcezed out of the

gap,

becomes suddenly

more stiff than an unloaded film existing in a well balanced system.

10.5.2.3.

The

damping coefficient

the

lubricant's film

Imagine an oil-filled gap, length

in the flow direction, width

clearance

J*.

Both the walls

this gap might be squeezed by a damping force

with the clearance rate

I;..

Hence by

definition the damping coefficient

&/j.

-This

sclrlcezing ;~ctui~res

prcssurc

which

presses the oil out of thc grip in the

stre;~mwise dirccr~on. It.ngtli

kvitll

thc avcrnsc spcctl

To bal:il~cc ti~c will1 shear stress

~l/y.

Nz)\l~~r's

law rcquires

tlc/y.

Continuity rcquires

j//y.

Wit11 Ap, the

dninping force becomes

Aplb

z12b/y

rlc12b/y2

t~yI'h/y'-

dy.

Reducing

yields the damping cocflicient of the film

might' be the axial length of a journal bearing and

the

circumferential

length of its

loadcd face,

the clearance of its gap. Hencc the damping coefficient of

bearing with

an unbalanced load, squeezing the oil out of the

gap, increases with decreasing clearance,

whereas the damping coefficient of a balanced system is nearly zero.

10.5.2.4.

Control

bearings

Regular working of a large hydroelectric unit

reql~ires

continuous measurement of

amplitudes due to shaft vibration to be installed,

e.g., at the bearings of the rotor. Such

measuring set consists of vibration transducers, signal amplifier and an indicator

and

rocording device with an appropriate alarm and locking signal. In this way

is possible

to intervene

quicklj in the case of alteratioli in the dynamic balance of the unit, caused

~nomalies (short circuit, bearing wear) and also to forecast the need for any normal

or emergency

maintainance,

[LO.

1211.

10.5.3.

nearing

design

and arrangement

10.5.3.1.

General remarks

In consequence of their size and load, particularly in the axial direction the bearings of

larger sets are only of the sliding type. Moreover the

12/licllell

tilting pad bearing type is

used exclusively for the axial

load, composed of hydraulic axial thrust and the weight of

rotating parts

the case of vertical shafts. For shaft diameters above about

1,5

m this

bearing

type is also preferred for the radial (guide) bearing.

Any bearing has to transmit a certain heat flow through its casing into the supporting

structure towards the foundations. This originates partly from

tlie waste heat of the

alternator

(in

the case of a bearing due

the latter) and partly from the bearing itself.

Fig.

10.5.4.

Principal bearing

arrangements. a)

Radial bear-

ings, both

the rotors of

overhung design, thrust bearing

on the head cover.

Radial

bearings, thrust bearing on the

pit. Variants: middle or

upper

radial bearing omitted.

c)3

Radial bearings, thrust bearing

or? the generator stator.

several designs have been proposed to avoid additional stresses in this structure due to

thermal expansion. One solution consists in radial arms transmittins any radial expan-

sion via strongly prestressed springs into the concrete of the pit wall.

Another solution uses arms which are slightly inclined to the radius, so-called "skew

spokes", [lo. 173). Thus heat expansion of these arms is converted into a central rotation

the bearing casing about its axis. Thereby the skew spokes are subjected

bcnding.

With the largest heat expansion of the bearing's support, the clearance of a radial bearing

must not fall short of the prescribed threshold.

Fig. 10.5.4 shows typical bearing arrangements of hydro turbo generators with vertical

shaft. To this typical bearing arrangements

horizontal sets have to be added. See

Figs.

3.4.1; 3.4.19; 3.4.20; 3.4.22; 3.4.23; 3.4.28; 3.4.33; 10.2.10 to 10.2.18.

10.5.3.2.

Guide bearings

Usually two guide bearings are used, as their load is statically determined. Overhung

design is preferred at the turbine runner especially in reaction turbines. This keeps the

draft tube bend free of any obstruction by the shaft. In some ternary pumped-storage sets

the runner can be disconnected either from the alternator-motor (Fig.

3.4.38) or from the

pump (Fig. 3.4.39). This requires two guide bearings on both sides of the runner. This

holds also for rarely found double flow

FTs with a twin runner supplied by one spiral

casing.

The relatively big alternator rotor has an overhung design only up to heads

about

250m

and for vertical sets. In this head range the umbrella or semi-umbrella type of

alternator rotor prevails (Fig.

10.5.4a). Above this head the alternator rotor is supported

both ends (Fig. 10.5.4b).

the high-head range the two and three bearing arrangement is used. In smaller low

head sets with submerged turbine also three or more bearings are found, to make the

generator easily accessible and to place it above the tail race level.

The lubrication of the bearing requires the existence of

wedge-fonned clearances between

shaft and bearing sleeve in the order of

0,5

times the shaft diameter [10.174]. To

take out-of-balance loads like the magnetic pull the radial support of the bearing must

of solid design in all radial directions.

10.5.3.3.

Tllrust bearing

design,

brake

Vertical sets, especially of higher head, with axial thrust downward oriented, need only

one thrust bearing

(Fig.

10.5.2). Horizontal sets need two (Figs. 10.2.10; 10.5.5). For the

brief

and small reversed axial thrust in vertical sets during transients in KTs,

thrust

collar is provided. This is usually on the lower extremity of the bell-shaped part

the

shaft that is immersed in an annular oil tank, to facilitate lubrication of guide bearing

during starting (see Fig. 10.5.6).

The clearance of the surge face pad on the lower front of the guide bearing casing must

in thc order of the axial deflection which may be experienced under axial thrust.

For heads up to 320 m (Fig. 10.5.4) the thrust bearing is underneath the alternator and

then

supported either by a spider oil the pit wall or a supporting truncated cone on the

kid

cover (Figs. 10.5.4a; 10.5.4 b). In small sets and under highest heads, e-g., in PTs the

firust bearing rests on the generator stator by means of a spider (Fig. 10.5.4~).

Ie'ig.

10.5.5.

Guide bearin

two

;lxi;11

(thrust) bearin

c.g.,

T;T

with liorizonta

Oil circulation of thc t

bcaring

pulnp

feed

ring pipc 12, with radial

chcs, blowinl: out

the

oil agai

thc rotating thrust collar

starting

lift

by prcssurizc

from the bore 25.

thrust collar,

tilted

prop with calotte to sup

pad,

callote pressed

iron ring,

spaccr

cage

pads,

pcdcstal of bcaring, cast iron,

drivcr for

1,s

oil level,

shah

10 split support ring,

fast-

ening ring for

10,

12 pipe con-

nection for fresh oil,

injection

pipe,

oil

pan, 16 ovcflow

edge, 20 bearing sleeve, 25 bore

for prcssurizcd oil during Start.

For an overhung alternator rotor the accessibility of the thrust bearing

may

be facilitated

by a removable ring, connecting the upper

end of the shaft with the upper front face of

thc rotor (Fig. 10.2.1). Tliis ring then serves as a thrust collar, which transmits the axial

thrust from the shaft tc the tilted segments.

For a revision, both the rotors must be supported by other devices,

e.g., the generator

brake, then used as jack. Then after removing the connecting bolts and the ring the tilted

pads become easily-accessible.

Each pad may be supported by springs or it rests on a radial edge or a spherical calotte

which presses

a ring of mild steel (Fig. 10.5.7). The pad may be supported also by a

pressurized plunger piston in an elastic

toroidal shell, supplied from a common oil

pressure vessel, under the controlled pressure of all the pads together, Fig.

10.5.8 [10.134].

At heads above about

1000

In,

a water-lubricated hydrostatic thrust bearing, at the lower

end of the shaft, reduces the tip diameter of this

besring and the corresponding disk

friction loss

[10.135]. This follows from the head-conditioned pressure, high

compared

with the mean

value of about 35 bar, which is attainable in the gap of a tilted pad, (see

Fig.

10.5.9).

In large units the elastic and thermal deformation of thc pad face from its desired plane

form is sometimes automatically compensated

meails of

special construction

[10.133]. Thus the centre of pressure distribution at the shoe is always retained.

Fig.

10.5.6. Oil lubricated lower bell-type guide bearing

of a

KT.

Small collar for the transitory uplift during a

o top page.

Cooling by the working fluid (water) on the

water deflection shield.

lnlct of the suction pipe near to

the cooling surface.

Oil circulation by the combined

&on of viscosity, shaft rotation and inclined groves

(after Sulzer Escher Wyss).

Before the start? each pad is fed through a flexible hose by pressurized oil, to ensure fluid friction from

the beginning of operation. To limit the period of mixed friction during

runnins down, each set has

brake device. This is of particular importance for peak load or pumped storage sets with their

frequent load change or their cyclic service of generating and pumping.

The present tendency is to limit the proper mechanical braking to the final phase of deceleration,

below 10

of the rated speed. The brake is on the spider below the alternator rotor and presses its

oil operated piston against the lower front face of this rotor. This mechanical braking is preceded

either by hydraulic or electric short circuit braking. In this way problems of dust, wear,

overheatins

ofsliding surface and vibrations are minimized

[11.69].

The brake is designed to cope with emergency

situations which would arise up to

50%

of rated speed. In Fig. 10.5.10 brake and jacking device are

arranged separately.

To eliminate any dust and wear during braking at the

pun-lp-turbine set of Kiihtai. Austria, any

mechanical braking has

been cancelled in favour of an electrodynamical stoppase. Usually then the

oil-operated mechanical brake is left for the emergency case, and as jacking device during revision

for the rotating parts

[10.175].

Under rated speed the oil is circulated either by flinger bore holes in the thrust collar or by a separate

pump. The latter feeds a ring

pipe with radial passages, from which the oil spouts against the !o\ser

front face of the rotating thrust collar. Then it is drawn into the oil wedge on the pad. After the pad

the heated oil is wiped off by a stripper.

Fig.

10.5.7.

Elevation of an axial thrust bearing

with tilted segments, oil lubricated. Cooler in a

non

pressurized oil tank. Circulation by axial

booster pump

radial bore

and bamc

past

outlet from

Oil flow through the bearing and

the lubricating gap

in shunt to the oil flow

Passing the cooler

Air separation from the oil

after the cooler on the free level of the oil tank.

(After Sulzer Escher Wyss.)

Cool

Cooling

Water

Ill

Fig. 10.5.8.

Elevation

of the thrust bearing of large FT. Plant Tucurui, Tocantin, Brazil (owner

Elcctrcnorte) built by Neyrpic. See Fig. 10.3.4. External diamster of casing

5,1,

tilted pad cxter~al

dinmetcr 4,205 m, internal diameter 2,75. Shaft diameter 1,75

Axial thrust 2410 tons. Oil: 8,5"

Er.gler at 50 "C. Cooling water 90 m3/h. Oil volume 5,3

m3,

auxiliary oil volume 4,3

m3.

Tilted pad

supported

a hyd;ostatically loaded piston 7 usually carried

pressurized oil in a toroidal axially

flexible vessel

communicating with each other. In emergency the pad carried by the calotte 8. The

toroidal vessel enables also

slight radial deflection,

the pad floats. Oil from cooler enters at the

in!ler edge 11

the bottom ring, is then centrifuged by the thrust collar

either between the pads

directly, or through the lubricating gap into the outer part of the case.

Fastening of the thrust collar

3,ti!tcd pad,

joint.

oil injection for jacking at start up,

toroidal shell, 7 upper support disk

of the pad,

8 calotte,

bottom support disk of the pad,

oil exit to the cooler, 11 oil inlet from the

cooler,

oil level.

radial splint,

guide cage for the pad,

bottom plate of the pad with

pipe

for pressurized oil from a balancing chamber. (Drawing courtesy Neyrpic, Grenoble, France).

Fig. 10.5.9. Thrust bearing with hydrostatic water lubrication in one of the 6-stage storage pumps

of San Fiorano pumped storage plant, (owner

ENEL,

Italy) with ternary sets built by Sulzer Escher

Wyss de Pretto.

1483

105,8

(see also Fig. 1.6.8). 1 thrust collar, clamped on the

shaft, with discharge orifice,

rotary thrust disk attached to 1,

stationary support of

radial

gap to relieve pressure chamber

14,s pressurized, sealed, and water filled calotte, 6 bore

balance

the pressure in

and 14,7 valve for the admission of pressurized water from the penstock,

filter,

flow control, 10 water head,

admission pipe for the pressurized water, 12 strain gauge to

indicate water load,

throttle,

pressure chamber, 15 gap between stationary parts

and 5,

16 leakage pan, 17 casing, 18 comb labyrinth, 19 transducer, 20 clearance indicator.

10.5.4.

Lubricants and

their

cooling

general oil is used as

lubricant, in some cases also grease water emulsion or (in the USSR)

Purified water. The oil lubrication of the bearing next

the water side needs protection against

Water pollution in the form of a shaft seal

(Fig.

10.5.1

I).

This requires

leakage sump between the

oil collector and the shaft seal, whose level must be controlled by

float and a pump.

Fig.

10.5.10.

:\rr:lngcnlcnt

thc braking

jacking cyli~~dcr-s on

tI1~

lowcr bcnring brae

braking cylinclcr,

jacking

cylinder,

ing ring. (Ur;twing fronl Hrown Bovcri

Cic: Spcci;ll issue: Synchronous lnachines

hydro clcctric power plilnts. Publ., no.

CH-T

130082

39,

Fig. 82). 1,

oftcn combined

Cooling of the lubricant oil: For the guidc bearings next to the runner, c.g., in

KT, the oil

may

be coolcd by the working fluid, i.e. along the water deflection shield (Fig. 10.5.6). Here the cooling

is the more effective the closcr the oil circuit intake is to the cooled wall. In othcr bearings more

distant from thc discharge

special cooler must be providcd. This may be inside or outside the oil

tailk. The oil flow and its temperature at inlet and outlet of the cooler must bc

controlled.

In the case

of excessive temperature

the sct is stopped automatically. Through the coolcr passes a flow which

is a multiple of that,

drawn into the wedge-formed gap of the bearing. This lubricating oil flow

circulates

shunt to the cooled oil flow so as to mix intcnsivcly at thc joints with the latter. The

circulrition through the coolcr may bc via special pumps. Usually

is donc by flinger bores in the

thrust collar (Fig.

10.5.7)

flinger shell (Fig. 10.3.10). The conversion of the high oil speed into

pressure,

needed to overcome the friction in the cooler, may be effected by baffles or a

Pitot

tube

(Fig. 10.3.10). The circulation of the cooling flow in the guide bearings also may be effected by viscous

shear stress

the

rotating shaft in connection

with

hclical groove within the bearing sleeve

(Fig.

10.5.6).

10.5.5.

Runaway and

its

problems

10.5.5.1.

General

remarks

In the case of runaway the specific head gH is converted through dissipation into waste

heat. I-Ience

(29HlcD)1i2,

(10.5-

12)

where

is a characteristic absolute velocity of the turbine and

a coefficient. At

runaway the torque becomes zero and hence by

Euler's

equation also the difference of

angular momentum between the rotor inlet and outlet

c,,

r,. In a

this

condition reads

Ac,

Thus both the velocity triangles coincide here under runaway.

Let

be the absolute velocity of this runaway triangle. Hence at a certain

c,,

the

peripheral blade speed is the larger the smaller the angle

the blade makes with the

circumference. Thus the highest possible runaway speed in a

occurs at overgate and

s~nall blade angle at cam-off operation. This may be caused by the longer opening time

the runner servomotor has compared with that of the gates (Cap. 11.2). Such a situation

may occur, when the

co~ltrol loops of gate and runner are connected in series in which

~ig.

10.5.1

Shaft seal with graphite ring pressed against

a chromed sleeve by a ring spring. Conical joint between

and inner ring causes slippage of both the rings against

@ch other and thus an axial tightness of the ring chamber.

Lubrication and cooling by the admission of purified block-

age water

between the upper and the lower seal ring

&amber. Left: standstill seal pressed down before a revision

the shaft seal.

case the tachometer first actuates the gate servomotor. Therefore a connection of both

the loops in parallel is preferred, e.g.

the

French

firm Neyrpic (Cap. 11.2).

10.5.5.2.

Protection

the set against runaway

Contrary to practice with thermal turbo generators, the hydro turbo set in general has

withstand the rimaway speed

n,,

of its turbine. This is

multiple (1,4 to 3,3) of the rated

speed, depending on the design (see Table

9.2.1). The first safety precaution is the speed

governor of the set (if any at all). At least each set has an emergency shut down device

for the case a certain overspeed (about 1,3 the rated one)

surpassed {41.21],

[IS].

Since the closing time of gates is limited

water hammer-induced .pressure surge in the

penstock, the set may reach rather high overspeed du'ring emergency shut down. When

this device fails,

e.g., by jamming, the then unloaded set may attain its runaway speed.

As a general rule,

may be stated, that sophisticated devices for avoiding runaway (such

jet deflectors in impl~lse Ts or braking runner blades swinging out in axial Ts) are

impracticable and not reliable in larger sets

[10.68; 8.1 321.

micro power stations runaway may be the state in which the set passes the time between its

working periods. In such plants the peripheral blade speed may be small, so as to have more an idling

period instead of runaway, useful to ensure the lubrication of bearings.

Sometimes in double regulated

KTs the runner servo motor may be short-circuited by an overspeed-

actuated valve. Thereafter the blades can follow their

inbuilt opening tendency to ensure the lowest

possible runaway speed (see above). This implies that the outermost components of the alternator

rotor like the poles are

ovzrstrained so as to need rewinding. But the core of the set has to withstand

runaway until the bulkhcnds are

insertcd

[10.176].

smaller units with step up gear for the alternator the latter may be protected against runaway by

loosening the clutch then used to attach the gear casing to the ground. Remember that any gear

needs a connection to the foundations so as to lead the difference in torque between the input and

output

shafr into the grobnd. When this connection is interrupted then the gear functions as a

coupling. Thereby thc alternator rotor is protected against runaway speed. Such a device has been

Proved successfully in thc bulb turbine at Ossbergshausen on Aggcr, West Germany, with a plane-

tary gear

Knrp13-Stoeckic.hr,

[10.177].

Thc

high head PT sets at Silz in Tyrol, Austria, of thc Sellrain scheme arc an exceptional example

ofthe sit~~ation, that the lowest critical speed falls short

the runaway speed. Here a thickening of

the

shaft suflicient to raise the critical speed above the runaway specd would have required the use

of rather expensive material in the alternator rotor because of the resulting increased diameter.

sucll

caw, carc must be taken, th,it undcr all imnglnnblc opcr;~tions thc fundway spccd ncvcr

co~ncidcq nit11 the critical SPI-C~.

Tlic

I:~ttcr

only p:lsscd dur~ng running tl~\!~n or 5pccdrng

up.

I11

two rcccrltly buiit scts

with

Isogjrc punlpturbincs, Grimscl. Swirzcrl;~nd, Fig.

3.4.3,

and lZ/laltn,

upper stagc, Austrin, Fig.

1.6.7,

1.70;

10.1

161,

thc lowcst critical spccd f;~lls short of runaway -\pc~d.

This may

a rcasonnble practice,

the runawily spccd dcviates from thc lowest critical speed ulldcr

all

possiblc modcs ol

operation.

Howcver, for the Malta set this is iiot

assured,

as the scts opcratc with a largc variation in

head

:l,,d

runaway

speed

to tlccd, by thc way, also a cl~unge polc

alternator.

Hcncc thc critical speed

occasionnl!y may coincide with thc runaway spccd. In this case all thc cmcrgcncy shut down devices,

govciiior, ovcrspced governor, valves and gates must opcratc safely.

10.6.

The

comyrltation of

flow

Francis

runner

10.6.1.

Flaw

prediction for

given

runner and working data

10.6.1.1.

Introduction

This section is considered to be supplementary to the simplified runner design of a

ii~ Cap. 10.3.6, which may be used as a preparation. In the following the problem is

treated of the distribution of relative velocity in

given runner vane channel under given

flow

rate with respect to face to face and shroud to hub distribution even with regard to

loss

and a slight twist of the streamface within the runner channel.

10.6.1.2. Assumptions

Consider the runner of a FT (Fig. 10.6.1) with given geometry under known working data

II,

and

(o)

around the bep. The flow does not stall and is steady. The absolute

flow

is irrotational. Thus the axisymmetric flow pattern

the mcridian gained by the net work

mcthod due to potential flow, Cap.

5.2,

is taken

a first approximation. From this

stepwise approximation may be made to the real streamline pattern.

The

pheliome~ia at the ends of a vane'are neglected. Hence the flow is along the vane

surfzce. The flow is treated on a surface with coordinates

and

(Fig.

10.6.2). The

azimuthal direction

is peripherally orientated. It coincides with the sense of the angular

velocity

of the runner. The a-lines are oriented liom the shroud towards the hub and

follow at leas! on the runner vane face the direction of the latter. They make an angle

n/2

with the periphery (Fig. 10.6.2a). They coincide (Fig. 10.6.1.2) on the runner inlet

and outlet with the axisymmetric surfaces

and 1, touching the inlet and outlet edge

and 1 of the runner vane. The other axisymmetric surfaces that contain the a-lines vary

conti~iuously in streamwise direction between these faces

and

at the extremity of the

rotor vane passage.

The projections of the a-lines in the local meridian, denoted by

a',

make an angle

with

the local

rzdius

(Fig. 10.6.1.1). The a, cp-face has the advantage, that it neither intersects

the

axisym~netric faces

and

touching the extremities of the runner vanes, nor the

runner vane itself, when proceeding along a direction from shroud to hub or when

proceeding in the cp-direction from the suction face to the pressure face of the vane. With

this system the blockage

the flow passage by the vane and its boundary layers must

be accounted for by an adequate contraction coefticient.