Steam Turbine Theory, Design, & Maintenance.pdf

Steam Turbine Theory, Design, & Maintenance

1. 6 July, 2007 Page 1 of 58
STEAM TURBINE THEORY
(DESIGN AND MAINTENANCE)
PREPARED BY
SYED SALEEM AHMED
MECHANICAL DEPARTMENT
HUB POWER PLANT
INTERNATIONAL POWER GLOBAL DEVELOPMENT Ltd.

2. 6 July, 2007 Page 2 of 58 Syed Saleem Ahmed
CONTENTS
‰CODE (HUB POWER PLANT) ………………………………………………………………….. 4
‰CODES …………………………………………………………………………………………….. 5 - 8
‰SECTIONAL DRAWING OF HP/IP AND LP TURBINE ………………………………………. 9
‰CHEST ……………………………………………………………………………………………… 10
‰RING-TYPE EXPANSION JOINT (MAINTENANCE) …………………………………………. 11
‰EXPANSION JOINT AND STEAM SEAL RING (BLOW-DOWN) …………………………… 12
‰THERMOCOUPLE (SHELL) ……………………………………………………………………… 13
‰ATMOSPHERIC RELIEF DIAPHRAGM ………………………………………………………… 14
‰CROSSOVER PIPE ……………………………………………………………………………... .. 15
ASSEMBLY. DISASSEMBLY. NUT TIGHTENING INSTRUCTIONS
‰MAIN JOURNAL BEARINGS ……………………………………………………………………. 16 - 23
O ELLIPTICAL TYPE WITH RING
O ALIGNMENTS OF BEARING TO JOURNAL (TILT & TWIST)
O DOUBLE TILTING PAD TYPE
O BEARING POSITION CHANGE
O FITTING INSTRUCTIONS – BEARING RING TO PEDESTAL
‰THRUST BEARING (TAPER LAND) …………………………………………………………….. 24 - 25
GENERAL DESIGN. THRUST BEARING WEAR DETECTOR.
‰MAIN OIL PUMP (DOUBLE SUCTION CENTRIFUGAL) …………………………………… .. 26 - 27
‰OVER-SPEED TRIP DEVICE (HIGH SPEED, OIL TRIPPED TYPE) ……………………….. 28 - 31
DESIGN. OPERATION. ADJUSTMENTS.
‰BUCKET TENON PEENING ……………………………………………………………………… 32 - 33
TOOLING. PRE-PEENING INSPECTION . PEENING PROCEDURES. FOX-HOLED COVERS (RECESSED).
FIELD PEENING AND RE-PEENING OF COLD BUCKET MATERIALS.
‰DOVETAIL BUCKET (TANGENTIAL ENTRY) ………………………………………………. 34
‰INSTALLATION OF BALANCE WEIGHT ……………………………………………………. 35 - 36
WEIGHTS AND WEIGHT GROOVES. BALANCE PLUGS. BALANCE WEIGHT MATERIAL.
‰SHELL ARM KEYS ( ALIGNMENT AND INSTALLATION ) ……………………………….. 37 - 38
BLOCK OR SQUARE KEY. “T” KEY. “L” KEY. SHOE KEY. 10
0
ANGLE KEY. CENTERLINE SUPPORTED
DESIGN
‰SHELL ARM HOLD-DOWN BOLTS …………………………………………………………. 39
‰ARRANGEMENT OF HP/IP SHELL BOLTING …………………………………………….. 40 - 41

3. 6 July, 2007 Page 3 of 58 Syed Saleem Ahmed


CONTENTS –
Continued



‰ LP TURBINE INNER CASING GIB AND KEY CLEARANCE …………………………….. 42 - 44

‰ DIAPHRAGM ALIGNMENT (INSTALLATION AND MAINTENANCE) ……………………. 45 - 47
AXIAL POSITIONING. TRANSVER SE POSITIONING. CENTERING PIN TYPE. MAINTENANCE
INSPECTIONS

‰ ALIGNMENT OF TURBINE INTERNAL PARTS ……………………………………………. 48 - 50

‰ SHAFT SEAL PACKING ……………………………………………………………………….. 51 - 56
O RETAINER DESIGN ( PIN DESIGN, RIVET DESIGN, KEY DESIGN)
O METALLIC LABYRINTH TYPE
O CLEARANCES ( BUTT, RADIAL, AXIAL)

‰ DIFFERENTIAL EXPANSION OF STEAM TURBINE GENERATORS ………………….. 57 - 58
DIFFERENTIAL EXPANSION. THE ROTOR LONG CO NDITION. THE ROTOR SHORT CONDITION.
ROTOR EXPANSION.

4. 6 July, 2007 Page 9 of 58 Syed Saleem Ahmed

5. 6 July, 2007 Page 10 of 58 Syed Saleem Ahmed
Chest
Looking from Turbine End
NOZZLE BOX (HALF)

6. 6 July, 2007 Page 11 of 58 Syed Saleem Ahmed
Ring-Type Expansion Joint
Maintenance

The recommendations for maintenance and inspection are as follows:
‰ Inspect the hardened surfaces of the sleeves for scoring which may have occurred during operation
or disassembly. Scoring of the sleeves would indicate that the bores of the rings are also scored.
Light scoring on the sleeves should be polished out to remove burrs and roughness.
Serious scoring or galling would require repair or replacement of the hardened surfaces.

‰ Inspect the sealing ring assemblies for:
ƒ Scoring of the ring inside diameters. (Light scoring should be polished out. Rings, which are
seriously scored, resulting in excessive steam leakage, should be replaced).
ƒ Condition of the locking rings and/or welding for the retaining rings.
ƒ Freedom of the rings to move and the amount of oxide buildup in the assemblies. (A light, tight
oxide scale is not harmful. A heavy oxide scale, which appears to prevent the rings from moving
properly or is flaking off, should be cleaned off by disassembly of the rings).

‰ Rings, which are not free to move, may be soaked with penetrating oil as a means of loosening
them.

‰ The radial clearances between the outer ring outside diameters and the shell bores and between
the inner ring inside diameters and the sleeves are important as far as steam leakage through the
assembly is concerned. The rings must be flat. That is the sides should be parallel.

‰ The axial clearance in the assembly is not of critical importance although it may not be excessive
since the steam pressure holds the rings tightly against one end. Some axial clearance is
necessary since the rings heat and expand more rapidly than the shell in which they are held.

7. 6 July, 2007 Page 12 of 58 Syed Saleem Ahmed
Expansion Joint and Steam seal rings
Blow-Down

8. 6 July, 2007 Page 13 of 58 Syed Saleem Ahmed
Thermocouple
Shell

9. 6 July, 2007 Page 14 of 58 Syed Saleem Ahmed
Atmospheric relief diaphragm


The atmospheric relief diaphragm is a safety feature which protects the exhaust hood and condenser
against excessive steam pressure in case the condenser water for any reason is lost.
The device consists of a hard-rolled silver-bearing copper sheet. In normal operation of the turbine
with proper vacuum conditions, the diaphragm is dished inward against the supporting grid by
atmospheric pressure. Should the vacuum conditions fail for any reason and the internal exhaust-
hood pressure rise to approximately 0.034 MPa, it would force the diaphragm outward against the
cutting knife. The diaphragm would be cut free as a disk relieving the exhaust pressure to
atmosphere.

10. 6 July, 2007 Page 15 of 58 Syed Saleem Ahmed
Crossover Pipe

Assembly:
Place crossover pipe in place and install washers & nuts “P” and “K” to studs on flange connections.
Remove locking bolts “R” (8 total) from crossover at this time. This is very important other wise
interference and binding will occur during expansion movement.

Disassembly:
Replace locking bolts “R” (8 total). Do not attempt to disconnect crossover until this has been done.
At room temperature the locking bolts should line up with the holes in the spacer blocks. If not,
jacking bolts “V” must be installed and tightened before loosening studs and nuts on flanges.
Position crossover with the jacking bolts can be installed.

Nut tightening instructions:

Tightening nuts up firmly without force until gasket is ready to be compressed and then tighten to
the amount of 3 flats on hexagon.
The above number of flats will compress the gaskets and stress the studs the required amount.

11. 6 July, 2007 Page 16 of 58 Syed Saleem Ahmed
Main Journal Bearings
Elliptical type with ring

It is basically consist of a bearing body with the tin-based babbit directly applied to it, the babbit is
retained in dovetail grooves and a bearing ring, both splitted horizontally in two halves.
The bearing body has a outer spherical surface (ball seat) mating with the internal spherical surface of
the bearing ring.

The bearing bore has an elliptical shape, with a total vertical diametrical clearance of 1.33
0
/00 and a
total horizontal diametrical clearance of 2.66
0
/00.
Ex : Total vertical diametrical clearance = (1.33×Journal Ø mm) /1000
= (1.33×535) /1000
= 0.71 mm
The proper operation of the bearing depends, on a proper clearance: therefore when the vertical bore
clearance is close to twice the nominal clearance.

























The approximate ellipse is obtained by machining the bore to the larger horizontal diameter with shims
inserted in the joints of the bearing; the shims are then removed for final assembly. The bore has one
or two (depending on bearing size) overshot oil grooves extending over the top half of the babbit lining:
this single or double overshot oil groove has been provided mainly for cooling purposes. The depth of
the single or double groove varies from 1.5 to 5.0 mm depending on bearing size.
To facilitate the entrance and discharge of oil the bearing has the babbit cut away at the horizontal joint.
In order to minimize oil leakage into the oil deflectors, a circumferential end–leakage groove drains oil
from the out-board end of the bearing into the standard. That portion of the bearing bore adjacent to the
end-leakage groove is made cylindrical so as to reduce the oil leakage clearance.
The two halves of the bearing rings holding the bearing body are bolted together tightly when the
bearing is assembled.
The ring is secured to the standard or turbine casing, which is mounted on, by bolts and is properly
located by a circumferential tongue fitting into the a groove in the standard or turbine casing.
An anti-rotation pin is located in the ring and extends in to the bearing shell to prevent bearing rotation.
The bearing ring is provided with shims to adjust the proper position of the bearing.

12. 6 July, 2007 Page 17 of 58 Syed Saleem Ahmed
Main Journal Bearings
Alignments of bearing to journal (tilt & twist)


For reliable operation, the bearing must be installed parallel to the journal. The spherical ball seat
feature allows the bearing to be adjusted in two directions to obtain the proper alignment. Thickness
gauge readings should be taken between each side of the bearing and the journal (4 places) to be sure
that no “twist” exists. Similar clearance measurements at each end and at the top of the bearing must
be made to check that no “tilt” is present.



SIDE MISALIGNMENT (TWIST)


Twist = (GS – IS + ID – GD) / 2



Twist alignment limits = ± 0.3 × brg. Ø (mm)
1000


Example = 535 mm Dia. Brg.


Maximum twist allowable = ± 0.3 × 535
1000



= ± 0.16 mm




VERTICAL MISALIGNMENT (TILT)


Tilt = (DG – DI) – (LG – LI)



Tilt alignment limits = ± 0.1 × brg. Ø (mm)
1000


Example = 535 mm Dia. Brg.


Maximum tilt allowable = ± 0.1 × 535
1000



= ± 0.05 mm

13. 6 July, 2007 Page 18 of 58 Syed Saleem Ahmed
Main Journal Bearings
Double tilting pad type

These bearings provide maximum stability and freedom from shaft vibration. They are used in this unit
for #1, #2 and #3 journal bearing.

The tilting-pad design consists of a housing containing six steel pads with babbit centrifugally cast on
the bearing surface. The pads are supported on a cylindrical bore in the bearing casing, three in each
half. The pads have a smaller OD than the casing bore so as to be free to pivot in the direction of shaft
rotation and adapt themselves to the optimum oil film wedge. A large radius is also machined on the
pad backs in the axial direction to allow the pads to align themselves to the journal. Hook fits on the
shell retain the pads, and they are prevented from rotating by means of induction hardened locking
pins.

















Oil enters a feed annulus which directs flow through various drilled passageways. These passageways
are located so that each pad receives its own supply of lubricating oil. Drain holes located on the
downcoming side of the journal in each drain annulus are sized to restrict flow sufficiently so as to build
up a slight discharge pressure. The rest of the oil discharges through the ends of the bearing. To
minimize this end leakage, seal teeth are provided.

A standard NPT pipe tap is provided in the feed pocket at the horizontal joint on the upcoming side of
the journal to allow installation of temporary piping for flushing during installation and maintenance
outages.

The slide-in cover plate is provided mainly to allow access to the internal oil feed passageways.

The four strap-down bolts which pass through casing ears into the standard should be snuggly
tightened but care should be exercised so that ears deflection does not exceed 0.05 mm at the
outermost edge. Each of the four strap-down bolts should be then secured after tightening through the
use of lock plates provided.
Circular wear plates are installed in the casing bore under each lower half pad while hardened bushing
inserts are installed in the top pads. Periodically the back of the tilting pads, wear plates, bushings and
lock pins should be inspected for wear and replaced as necessary.

These bearings are machined at the factory with a cylindrical bore providing a clearance of 0.033
mm/25.4 mm for sizes up to and including 381 mm and 0.038 mm/25.4 for sizes 406 mm and above.
The stability characteristics of this bearing depend to a great extent on proper pad curvatures and a
close clearance, the maximum allowable clearance being 0.050 mm/25.4 mm of bore diameter. When
the clearance approaches this limit, or if shaft vibration has increased, consideration must be given to
re-establishing the proper bore geometry.

14. 6 July, 2007 Page 19 of 58 Syed Saleem Ahmed
Main Journal Bearings
Double tilting pad type

To prevent locking pins coming loose during operation, pins must be staked in four places with a tool
having a 1.58 mm radius.
It is important to periodically check the relative radius of curvature of the pads with respect to the
journal, especially if operating temperature is approaching the alarm point. During service, there is a
tendency for the pads to take up the radius of journal (especially the bottom pad or pads). If these pads
do assume this radius, then their load carrying capacity is seriously reduced and a failure may result
(Fig.1). Pads may be checked for proper curvature by bluing them against a mandrel which is equal in
size to the journal plus the diametrical clearance.

NOTE: UNDER NO CIRCUMSTANCES SHOULD THE PADS BE BLUED AGAINST THE TURBINE
SHAFT.

If this check indicates that the pads blue only at the leading and trailing edge, i.e., the radius of
curvature is less than design, they should be carefully hand scraped or shimmed in and rebored to
establish proper curvature. Leading and trailing edge radius should be restored to 9.52 mm.
Since the pads are normally loose and free to pivot in their casing. It is impossible to measure the bore
in the conventional manner. To properly measure the bore, the locking pins may be removed and
temporary holding bolts inserted through the shell and screwed into the backs of the pads. In this
manner the pads can be equalized by temporarily inserting four 6.00 mm set screws in the corner of
each pad and pulling the pads tightly against the housing (Fig.2). The bore can then be carefully
measured across the center of each pad. During reassembly, the clearance to the shaft can be
reconfirmed by using Plastigage or lead fuse wire.


























Reduction in clearance can be made by addition of suitable shims under each wear pad and adjusting
plate. The pads should be assembled in their casing as described previously with temporary holding
bolts and set screws, and finish bore to suit the journal diameter. After machining and installation lock
pins, the pads should be checked to be sure they are free to pivot in both the circumferential as well as
the axial direction.

15. 6 July, 2007 Page 20 of 58 Syed Saleem Ahmed
Main Journal Bearings
Double tilting pad type


A tilting pad bearing tends to operate with a slightly higher metal temperature than a corresponding
sleeve type elliptical bearing of the same size.


























Bolts identified with a require tightening torque at final assembling according to values shown in table.

TIGHTENING TORQUE
IDENTIFICATION SCREW DIMENSION KGM
Alignment pad M12×25 4 - 4.5
Half joint M20×140 15 – 16.5
Half joint M24×140 34.5 – 38
Tilting pad Locking pin M20 15 – 16.5


OPERATIONAL RECOMMENDATION DURING RUNNING

Bearing header oil pressure should be 1.8 bar at turbine centerline. The inlet oil should be 43
0
C to 49
0
C
and the oil temperature rise should not exceed 28
0
C.

Normal operating metal temperatures for various bearing designs are:
Tilting Pad 82 - 104
0
C
Elliptical 77 - 88
0
C
The recording instrument should be set usually to alarm at 107
0
C. A journal bearing should not be
operated with metal temperatures above 121
0
C.
Bearing metal temperatures should be fairly constant under full-load conditions. Any sudden increase
in temperature of more than 5 – 6
0
C should be considered abnormal.

16. 6 July, 2007 Page 21 of 58 Syed Saleem Ahmed
Main Journal Bearings
Bearing Position change

¾ To shift bearing vertically, with reference to fig. draw a straight line from desired rise on shift
scale through pad angle on vertical scale and read shim change for each pad on shim scale.
Example: To raise 0.20 mm with 30
0
pads, add 0.173mm shim to each pad. For pads on the
vertical center line (0
0
), the shim change equals the vertical shift.

¾ To shift bearing horizontally, draw a straight line from desired movement on shift scale through
pad angle on horizontal scale and read shim change on shim scale. Remove that shim from side
direction of motion and add it to the corresponding pad on the other side.
Example: To shift right 0.15 mm with 30
0
pads, change 0.075mm shim from right pad to left
pad.















Shim change = vertical shift × COS θ
Shim change = horizontal shift × SIN θ














To raise 0.20 mm – add 0.173 mm shim. To raise 0.20 mm – add 0.173 mm shim.
To shift right 0.15 mm – add 0.075mm shim To shift right 0.15 mm – remove 0.075mm shim
Total of above – add 0.248 mm shim. Total of above – add 0.098 mm shim.

17. 6 July, 2007 Page 22 of 58 Syed Saleem Ahmed
Main Journal Bearings
Fitting Instructions – Bearing ring to Pedestal

Coupling checks and rotor position checks should be obtained during disassembly of the unit. These
checks provide the benchmark for unit realignment.

RING FITTING

The horizontal joint of the ring parallel with the horizontal joint of the bearing pedestal within 0.127 mm.
The bearing ring tongue fit to pedestal should be checked (Fig.1). There should not be more than 0.25
mm total axial clearance.

18. 6 July, 2007 Page 23 of 58 Syed Saleem Ahmed
Main Journal Bearings
Fitting Instructions – Bearing ring to Pedestal

After making a pad shim change, reassemble the complete bearing ring. Apply a light coating of bluing
to the pedestal bore and rotate the bearing and ring assembly slightly. (Not more than 3.00 mm in each
direction). To rotate the assembly, it is preferable to push down on the ring ears rather than lift up. The
assembly can most easily be rotated by drawing up on the ring ear bolts on one side of the ring with the
opposite side bolts relaxed and make a feeler check between each pad and pedestal bore. It should not
be possible to insert a 0.04 mm feeler at any pad location.

Note: Shimming with bundles of thin shims must be avoided in order to prevent a possible loss in
bearing support stiffness. No more than 6 or 7 thin shims should remain under the ring pads after
shimming. A thin shim can be considered to be anything less than 3.00 mm shim found under each
bearing ring plate. Shims less than 0.05 mm should not be used.


Remove the bearing /ring and examine the contact of the pads with the bore of the supporting standard.
If pad contact, is not sufficient. Scrape the out side diameter of the pads until at least 80% contact is
obtained. Pads containing oil feed and drain holes should have contact completely around the hole
circumference.

Note: After the rotors are installed and it is found necessary to make a final bearing alignment change,
it will not be necessary to re-check the contact on the pads if the rotor is moved 0.254 mm or less.

19. 6 July, 2007 Page 24 of 58 Syed Saleem Ahmed
Thrust Bearing
Taper-land, straddle ball-seat type

Thrust bearing independently mounted inside the front standard, it absorbs the axial thrust of the
turbine and generator rotors, which are connected by a solid coupling.
GENERAL DESIGN
This taper-land thrust bearing consists of two stationary thrust plates and two rotating thrust collars on
the turbine shaft which will provide the front and back faces to the bearing. These plates are supported
in a casing so that they may be positioned against the rotating faces of the collars.
The surfaces of the two thrust plates are babbitted, and have taper lands of fixed converging surfaces,
permitting a wedge of oil to exist between the rotating thrust collars and the thrust plates. The thrust
plates are constructed as split copper rings, with the babbitted surfaces divided into lands by radial, oil
feed grooves. The surface of each land is tapered, so that it slopes toward the rotating collar, both in
the direction of rotation. The radial grooves are dammed at the outer ends, maintaining an oil pressure
in the groove.
There are two methods of generating the pressure within the bearing it self. The first results from the
geometry of the bearing & the second from the load application.

























Outside diameter of the casing is machined to a spherical surface, seating inside the bearing ring. This
will permit the thrust plates to align, relative to the thrust collars.
Thrust on the bearing is transmitted to the casing and the ring assembly, which is held in its axial
position by tongue-and-groove fit in the standard.
An adjusting shim is provided between each thrust plate and the casing. These shims are use to adjust
the axial position of the rotor, or change the thrust bearing clearance, if so required.
Metal temperatures on loaded thrust plate may normally be 60-80
0
C, and 50-65
0
C on the unloaded
plate. The high-temperature alarm on the thermocouple recorder should be set for 82
0
C. Under no
condition should the thrust bearing be operated above 88
0
C.

THRUST
BEARING WEAR DETECTOR For the protection of the turbine in the event of a
thrust bearing failure, a thrust bearing wear detector is provided.

20. 6 July, 2007 Page 25 of 58 Syed Saleem Ahmed
Thrust Bearing
Taper-land
























































ROTOR ROTATION
MAX. SLOPE AREA 0.23 MM ± 0.025
SLOPE 0.15 MM ± 0.025
CONTACT AREA
THERMOCOUPLE HOLE
DIRECTIONS OF SLOPE
RADIAL GROOVES

21. 6 July, 2007 Page 26 of 58 Syed Saleem Ahmed
Main oil Pump

Double-suction, Centrifugal

The main oil pump provided on this unit is a double-suction, single-stage, centrifugal-type pump, capable of
providing both the hydraulic and bearing oil requirements of the machine.
A spring-based, steady bearing is provided on the front end of the pump to stabilize the end of the shaft.
The bearing support is held against the pump casing by a series of compression springs which allow the
bearing support to center itself.
Split bronze sealing rings are provided at the eyes of the impeller to seal the suction pressure from the
discharge oil. A similar sealing ring is also provided on the suction side of the pump on the end opposite
the steady bearing. The sealing rings around the eyes of the impeller are free to rotate and to center
themselves when starting up. As pressure is built up at discharge of the impeller, the pressure drop across
the rings lock them into the centered position and the rings remain stationary. The sealing ring on the
inboard end of the pump will not have sufficient pressure drop across it to keep from rotating. A pin is
provided in the ring which rests in a notch in the pump casing to keep it from turning. Each individual
sealing ring runs on a replaceable steel wear ring or collar.

MAINTENANCE

The sealing rings are to be assembled so the milled oil pockets in the inner diameter of the ring face toward
the highest oil pressure chamber. The small drain slot cut in each pocket should be assembled so the
discharge edge points in the direction of shaft rotation.
The lower half of the bearing should be assembled to have 0.04 mm to 0.05 mm clearance with the shaft.
From this position of the steady bearing support, the support should have 2.3 mm in radial clearance
around the bolts for the compression spring assemblies.

22. 6 July, 2007 Page 27 of 58 Syed Saleem Ahmed
Main oil Pump

Double-suction, Centrifugal




















































COLLAR SCREW

23. 6 July, 2007 Page 28 of 58 Syed Saleem Ahmed
Overspeed trip device

High-Speed, Oil-Tripped Type

The over-speed trip device consists of unbalanced ring which is actuated by centrifugal force against the
force of a spring when the turbine over-speeds. This movement puts the ring in an eccentric position so
that it strikes the trip finger of the mechanical trip linkage which operates the mechanical trip valve and
closes all turbine steam admission valves.

DESIGN
The supporting sleeve is screwed on the end of the control rotor stub shaft and is encircled by the ring
assembly.
The helical spring is mounted in the supporting sleeve and is held in place by an adjustable spindle which
is used for varying the compression of the spring. The spindle of the ring assembly is guided by two Teflon
bushings retained by snap rings.

OPERATION
The overspeed trip device may be tested by tripping it at normal speed by the application of oil through the
oil trip valve. The oil is directed into a circumferential annulus and through the supporting sleeve, into two
pockets formed by the closed side of the ring. The weight of the oil unbalances the ring, causing the
overspeed trip device to trip. After the overspeed trip device trips and the oil trip valve is closed, the oil
pockets are emptied by centrifugal force through two small holes in the ring. The ring will then reset.

ADJUSTMENTS
The device is normally adjusted to trip at 110 - 111 percent.

Adjustment of the spring compression may be made by turning the spindle with a suitable screw driver.
Before attempting to turn the spindle, remove the cotter pin which locks it in place. After the adjustment is
made, replace this cotter pin. If the cotter pin needs replacing, make sure the same size and length are
used as those of the one removed. A longer cotter pin can change the setting considerably.

The tripping speed of the overspeed trip device may be adjusted by the following methods:

o By turning the spindle clockwise in half-turn increments, the tripping speed is increased. If turning it
counterclockwise, the tripping speed reduced. One-half turn spindle change will also the tripping
speed by approximately 150-180 rpm.

o By changing the length of the adjusting plug in the spindle.
a) Decreasing the overall length by 3 mm, decrease the tripping speed 30- 40 rpm.
b) Increasing the overall length by 3 mm, increase the tripping speed 30 – 40 rpm.

o By changing the position of the adjusting plug in the spindle.
a) Backing the plug out one turn increases the tripping speed by approximately 15 rpm.
b) Turning the plug in one turn decreases the tripping speed by approximately 15 rpm.

Adjustments to the plug or spindle must be such that the plug or spindle never extends beyond the outer
diameter of the ring.

24. 6 July, 2007 Page 29 of 58 Syed Saleem Ahmed
Overspeed Governor

25. 6 July, 2007 Page 30 of 58 Syed Saleem Ahmed
Overspeed Governor





















































RING OF TRIPPING GOVERNOR
OIL SUPPLY TUBE
TRIP
FINGER

26. 6 July, 2007 Page 31 of 58 Syed Saleem Ahmed
Overspeed Governor























































THRUST BRG WEAR DETECTOR
MECHANICAL TRIP VALVE
LOCKOUT SOLENOID VALVE
MECHANICAL TRIP SOLENOID VALVE

27. 6 July, 2007 Page 32 of 58 Syed Saleem Ahmed
Bucket Tenon Peening


Tenon peening is the method of banding buckets together in groups. The primary reasons for bucket
groups or covered buckets are vibration dampening and a method of keeping the steam within the
steam path.

Tooling
The peening tool to be used is riveting hammer (gun). These hexagonal-shaped replacement parts will
prevent the anvil from rotating and should give the worker better control while the gun is in
operation(see Fig. 1 and 2).












PRE-PEENING INSPECTION
The radius at the tip of the tenon before peening should be smooth and full size. The cover-to-tenon fit
must not be any larger than 0.38 mm. This is to be measured on one side of the tenon with opposite
side of the tenon against the cover as shown on Fig. 3.











PEENING PROCEDURES

General
All peening of tenons on buckets with tangential entry dovetails is to start 180
0
from the notch opening.
Axial entry first stages may start with any group.

Single Covers
For single covers, peening should start at or adjacent to the middle of the cover group and then
alternate either side of the first peened tenon (see Fig. 4).

28. 6 July, 2007 Page 33 of 58 Syed Saleem Ahmed
Bucket Tenon Peening


Fox-holed Covers (Recessed)
In a great many instances when covers are replaced in the field, the bucket tenons are reworked and a
“fox-holed” cover is needed due to the reduced tenon height. In Fig. 7 it is shown that the minimum
tenon stick-thru is 2.3 mm (the length of tenon above the bottom cover recess). The peened tenon
should remain 0.40 mm away from the side of the fox-hole.






FIELD PEENING AND REPEENING OF COLD BUCKET MATERIALS
Original Peening of B50A125, B50A365, B50A332, B50A490 and B50A240
Starting at the center of the tenon, Fig. 8, the material is worked out-ward to swell the tenon to fill the hole
in the bucket cover as shown in Figs. 9 and 10. During all phases of the peening procedure, the operator
should make continuous checks to be sure that the cover is down tight against the bucket tips.
A sight overlap of the tenon head lip beyond the top edge of the cover tenon hole top chamfer (0.25 mm
minimum overlap beyond chamfer).
















As shown in Fig. 11, the top of the cover chamfer “A” must be completely filled all around during final
shaping. The height “B” of the edge of the peened tenon head should be at least 0.80 mm with no feather
edges at any point around the periphery. All tenons are to be nondestructive tested after peening.

29. 6 July, 2007 Page 34 of 58 Syed Saleem Ahmed
Dovetail Buckets
Tangential Entry














































Notes:

Whenever buckets are disassembled from a
wheel, the wheel dovetail must be thoroughly
magnetic particle tested and if the buckets
are to be reused, they must be thoroughly
magnetic particle tested including the bucket
dovetails.

30. 6 July, 2007 Page 35 of 58 Syed Saleem Ahmed
Installation of balance weights


WEIGHTS AND WEIGHT GROOVES
Once the weight is positioned in the weight groove, the set screw is tightened against the inner surface
of the weight groove. This force the weight out until the inclined surface of the weight is tight against a
parallel inclined surface of the weight groove (Fig.1). For most units, an access opening exists at every
90
0
of balance groove arc in order to permit weight installation. After the weight is inserted in the
access opening, it should easily slide to the desired position.


















When balancing couplings, a T-shaped weight is used (Fig.2). The lips of the weight groove should be
staked at each end of the weight group so that they won’t loosen and slide out , and final locations and
amounts of installed weights should be recorded for future reference.

31. 6 July, 2007 Page 36 of 58 Syed Saleem Ahmed
Installation of balance weights


BALANCE PLUGS
High temperature rotors and nuclear low pressure turbines use balance plugs. When balance plugs are
installed, the weight is guided through a turbine access opening with the aid of an insertion tool (Fig.3).
Extreme caution must be exercised whenever balance plugs are installed. They must not be dropped
into the shell.





















There is usually a thermal expansion problem that is encountered when balance plugs are installed on
high temperature rotors. The rotor temperature can possibly approach 482
0
C, while the plug is at
ambient temperature. If the weight is installed under these conditions, it can become stuck where it
won’t go in or come back out. It order to prevent this, use the following procedure whenever balance
plugs are installed. The same precautions must be taken when cleaning a plug hole with a tap. The tap
may be heated by inserting it into the access hole and waiting a few minutes.
1) Tap has been heated to approximately rotor temperature.
2) Install the weight onto the insertion tool.
3) Lubricate the balance weight threads with high temperature anti-seize compound.
4) Heat the weight to approximately rotor temperature.
5) Thread the weight into the rotor balance hole.
6) Remove the insertion tool from the weight.
7) Tighten the weight with the tightening tool.


BALANCE WEIGHT MATERIAL

If weights are not available and must be made, it is very important that a suitable material is used. The
material depends on the weight location, rotor material, and the type weight.
Typical weight material is B50A332A (a 12 chrome material equivalent to ASI type 410) for those
weights fitting in grooves and B5F5B3 for most plug type weights. Plugs for the mid spans of fossil HP
rotors must be copper plated.
“Heaviment” weight (made from tungsten which has a high density) may be used in order to install
more weight in a given place. It is approximately 2.2 times heavier than normal balance weight material.

32. 6 July, 2007 Page 37 of 58 Syed Saleem Ahmed
Shell Arm keys
Alignment and Installation

The various types of keys shown in the following figures are elevation views.
KEY TYPES
Block or Square Key: The oldest and simplest type of key called the “block” or “square” key (Fig. 1).
The block key is the least expensive and easiest design to manufacture, fit and install. It is adequate
for the small units.
“T” Key: The “T” key (Fig. 2) is refinement of the block key. It is somewhat more costly and difficult to
fit. It affords a larger bearing area and the problem of obtaining adequate strength in the standard is
minimized because the key way can be smaller.




“L” Key: The “L” key (Fig.3) is a refinement of the “T” key, which can be machined and fitted more
easily. The filler piece takes the axial thrust.
Shoe Key: This key is made of cast iron. It is used where the standard is fixed and the shell arms
slides on the key. Alignment is maintained only by the gibs at the vertical centerline. There is no fit of
the key in the shell arm, only in the standard.








10
0
Angle Key: It is the only key that does not lie in a horizontal plane. The lower piece is hardened at
the sliding surface. The center piece is a chrome-moly-vanadium forging. This key is employed on
some large turbines which do not have centerline supported shells. As the shell arm expands
with heat, it tends to lift the lower shell upward, decreasing the clearances at the bottom of the turbine
rotor. The 10
0
angle centerline key compensates for this tendency.

33. 6 July, 2007 Page 38 of 58 Syed Saleem Ahmed
Shell Arm keys
Alignment and Installation

Centerline Supported Design: Figures 7 and 8 show two types of the centerline supported or
suspended shell arm construction. With this construction, the shells are supported exactly at the
horizontal joint so that clearances are not affected by the shell arm expansion.
The centerline support design is simpler to fit, less costly to manufacture, and provides more
dependable alignment accuracy under all conditions. Alignment adjustments for shell height and for
axial position can be made completely independent of each other, since different keys are involved.

Note there are several keys associated with the centerline supported shell arm construction.

Part A: is the “tops-off” key or support plate. It is used during alignment of the turbine or whenever the
upper half shell is to be removed. It must be put in place (using the jacks) before the horizontal joint
bolts are loosened preparatory to removing the upper half shell; otherwise there is nothing to support
the lower half shell and it may drop down as the bolts are loosened. The “tops-off” key (A) must
never be left in place when the unit is running, as this will nullify the centerline support design. The
lower shell will expand upward from the top-off key and decrease the clearance at the bottom of all the
packings. Serious rubbing may result. The top-off key is bolted to the standard as shown by the
phantom lines when not in use, for safe keeping.

Part B: The key at ‘B” is called the “safety” key. It is left in place at all times, except it may be
temporarily removed during alignment work. Its purpose is to prevent the lower shell from dropping
down too far, in case the tops-off key is accidentally left out when loosening the horizontal joint bolts. It
will never support any weight unless the top half shell is mistakenly removed without first assembly the
tops-off key. Th safety key has a clearance of 1.50 mm ±0.12 mm and this clearance should be
maintained if alignment changes are made.

Part C: is the “tops-on” key or support plate and is made of cast iron. It carries the weight of the upper
and lower shells at all times except when the upper half casing is to be removed.

Part D: The remaining two keys (D) locate the front standard axially, and transmit the forces of
expansion and contraction from the shells to the sliding standard.














The tops-on key will always be thicker than the tops-on key way. This is normal and necessary in order
to maintain proper alignment in the outer shell bore that was established at tops-off condition. The
deflection of the entire mass (upper and lower shells) requires the key to be thicker than the key way.
The difference between the key and key-way can vary from 0.38 mm for light fossil shells to as much
as 0.76 mm nuclear shells.

All keys except the shoe key, the tops-on key, “C”, in the centerline supported type of construction and
the 10
0
angle key are made of low carbon steel (AISI-C-1020) and such material for making
replacements can usually be obtained fairly easily.

34. 6 July, 2007 Page 39 of 58 Syed Saleem Ahmed
Shell Arm Hold-down bolts

Holds down bolts are employed near the keys on many units. The purposes of hold down bolts are to
prevent the shell from being lifted by expansion forces from the connected piping and to hold the lower
shells down while upper halves are being lifted off.
Since the turbine shell is keyed to the standard on the vertical center-line, the shell must be able to
expand outward in a direction perpendicular to the shaft center-line as it heats. Some hold down bolts
are designed long enough to be sufficiently flexible to deflect as the shell expands and others are
designed with clearance under the nut s to allow the shell to move in relation to the bolt.














Hold down bolts should be assembled with 0.076 mm to 0.127 mm clearance under the head unless
instructions for a particular unit state other wise. Excessive tightening is not necessary and may
stretch the bolts or restrain the desired free sliding action of the adjacent keys.
Older designs often employed hold down bolts passing through the key. The hold down bolt does
not pass through the key on modern suspended shell arm designs, but is located to the side. Where
the bolt passes through the key, it serves as a keeper to prevent the key gradually working out on the
key way. A small keeper plate is employed where the bolt will not serve as a keeper.
If it is necessary to remove the hold-down bolts extreme care should be taken that, as the bolts are
removed, the shell does not raise up into the rotor and damage the packings.

35. 6 July, 2007 Page 40 of 58 Syed Saleem Ahmed
Arrangement of HP/IP shell bolting






































A Main steam inlet M52 х3 0.17 - 0.22
B Crossover M52 х3 340 kgm
C 1
st
stage press. tap conn. M22 х2.5 21 kgm
E Bal. wt. (Mid span) M16 07 kgm
F Blow down line M39 х3 105 kgm
G 1
st
stage tcpl. Holder M20 х2.5 15 kgm
H Packing casing T.E M24 х3 26 kgm
J Nozzle-box (hold-down) M30 27 kgm
K Thermo holder Asm. M22 21 kgm
M Inspection Flange M20 15 kgm
N Control valve M64 х3 0.15 - 0.19

36. 6 July, 2007 Page 41 of 58 Syed Saleem Ahmed
Arrangement of HP/IP shell bolting

37. 6 July, 2007 Page 42 of 58 Syed Saleem Ahmed
LP Casing Gib and Key Clearance

38. 6 July, 2007 Page 43 of 58 Syed Saleem Ahmed

LP Turbine Inner Casing

39. 6 July, 2007 Page 44 of 58 Syed Saleem Ahmed
LP Casing Gib and Key

40. 6 July, 2007 Page 45 of 58 Syed Saleem Ahmed
Diaphragm Alignment
Installation and Maintenance
During turbine operation, the diaphragms are subjected to three principle forces as they fulfill their
function of controlling steam conditions from stage to stage:
‰ Their own weight.
‰ The force of the steam tending to push the diaphragm axially downstream.
‰ A rotating force tending to cause the diaphragm to rotate opposite to shaft rotation, caused by the
reaction of steam against the partitions.

The methods of assembly diaphragms into the turbine must provide for these forces, and must also
locate the diaphragms vertically and transversely. Axial positioning is fixed, and is determined by the
shell ledge (steam face).
AXIAL POSITIONING
The steam forces the diaphragm firmly against the face when there is significant steam flow, thus
providing a seal as well as locating the diaphragm axially. When the steam flow is low, there may not
be sufficient force to hold the diaphragm firmly against the steam face, and it may tend to rattle, and
also may permit leakage that would erode the steam face. This is minimized by providing axial crush
pins on the upstream side of the diaphragm, which serve as stops against upstream movement. There
will be a varying number of axial crush pins on lower half of diaphragm depending on diameter. The
axial clearance between crush pin and shell should be 0.07 – 0.38 mm and clearance is necessary for
expansion.

41. 6 July, 2007 Page 46 of 58 Syed Saleem Ahmed
Diaphragm Alignment
Installation and Maintenance
TRANSVERSE POSITIONING
Two principle methods, with variations of each, are used to locate diaphragms transversely, and to
prevent rotation. These are by means of centering pins at or near the bottom centerline of the
diaphragm (see Fig. 3) or by arch spring supports at the horizontal joint behind the support bars (see
Fig. 2). In both methods, vertical positioning is determined by shims beneath the support bars.


















Centering pin type
Centering pins are used on all diaphragms whose average operating temperature exceeds 232
0
C.
Since 1968, however, all diaphragms for new units have been centering pin construction entirely.
‰ Vertical Adjustments: During installation, vertical position of the diaphragm is adjusted, if
necessary, by means of the support bar shim and thin stainless steel shims. No more than two thin
shims may be used under the main shim with a total thickness not to exceed 1.5 mm. Minimum
thickness of any single shim is 0.30 mm.
‰ Transverse Adjustment: A diaphragm can be adjusted side to side any desired amount S
by
“rolling” it. This is done by decreasing the support shim thickness by S on one side and by
increasing the shim thickness by S on the other side.
‰ Levelness Requirement: The diaphragms must not be “rolled” so much that the diaphragm is out-
of-level, in relation to the horizontal joint of the shell, more than 0.025 mm/25.4 mm of diameter. If
such an out-of-level condition exists, the diaphragm should be releveled by replacing the centering
pin with a new pin that has an offset tongue.
‰ Centering Pin Clearnace: When diaphragm alignment is complete, the total “side slip” must be
0.025 – 0.075 mm (i.e., if centered there would be 0.01 – 0.04 mm clearance on each side of the
centering pin) as shown at A in Fig. 3. The clearance at B should be the same as between the
casing and diaphragm.
‰ Clearance Spacer: The clearance spacer at the top of the support bar is provided to permit
adjustment of the space between the top of the support bar and the horizontal joint of the upper
half section. This clearance should be 0.075 – 0.20 mm, and must be checked at installation, and
any other time that an alignment change is made to the diaphragm. This clearance is necessary to
limit the amount the diaphragm can move vertically upward during turbine operation. The inboard
side of this spacer bar has 0.80 mm relief, so that the top half of the diaphragm will not be
prevented from closing at the horizontal joint should the shell horizontal joint be located above the
diaphragm horizontal joint.

42. 6 July, 2007 Page 47 of 58 Syed Saleem Ahmed
Diaphragm Alignment
Installation and Maintenance



























MAINTENANCE INSPECTIONS
‰ Lower axial crush pin clearances may increase from the normal after a turbine has been in service.
This increase in clearance is usually due either to an indentation in the ledge, or mushrooming of
the pin, or both. A clearance of 0.38 – 0.50 mm at the crush pins is reasonable, and should not
require corrections unless there is sufficient indentation in the shell ledge to prevent the necessary
radial expansion of the diaphragm. Clearances greater than 0.50 mm should be replaced.
Replacement pin material is B7B5B (12CR) material. Diaphragm crush pins can be built up by
welding, using B21B88A1 (AWS A5.11 Class ENicrFe3) commercial equivalent welding electrode.
‰ “Dished” diaphragms are a result of creep in the web/partition weld material. Opening turbine rotor-
to-diaphragm clearances show this type of defect quite clearly when compared to the installation
and/or previous outage closing clearances.
VISUAL INSPECTION can also indicate diaphragm dishing. Anytime the downstream side of a
diaphragm has rubbed on the adjacent stage wheel or buckets, diaphragm dishing should be
suspected. Drop
checks with a straight edge are recommended method of inspecting for a “dished” condition. All
axial drawing dimensions are referenced from the discharge side appendage of the diaphragm. To
make a drop check, a straight edge should be laid diametrically across the appendage, and all axial
dimensions should be checked at a minimum of six points around the diaphragm. If the downstream
surface of the appendage has been damaged or eroded, the straight edge should be set up on
blocks resting on the sealing face of the diaphragm. The six points should include measurements
on both sides of the horizontal joint and on the vertical center-line. At each of these six locations,
the following drop checks should be made:
ƒ Diaphragm steam face to packing steam face.
ƒ Diaphragm steam face to partition trailing edge – root and tip.
ƒ Diaphragm steam face to inner and outer set back faces.
Any drop checks in above which vary from the drawing by more than 1.14 mm should be reported.

43. 6 July, 2007 Page 48 of 58 Syed Saleem Ahmed
The generator stator frame feet are machined in
the factory at an angle so that the generator
coupling face is vertical. Bearing no. 5 will be
about 0.22mm and bearing no. 6 will be about
2.62mm above the level line.
Alignment of turbine internal parts

against common base line through piano wire
It is not as simple as lining up all of the diaphragm bottom halves against a common reference line. If
you plan properly, use the correct techniques and technology; you will save time on the diaphragm
alignment process and have a nearly rub free startup.
First you have to determine what the alignment reference line will be. First there is the catenary curve
of the turbine rotor. This is the natural sag that occurs in the turbine rotor that will not spin out of it when
the rotor is at running speed.
Figure shows how this fact influences the alignment of the bearings. Figure (a) is incorrect; in practice
such shafts would be lined out as in (b). Two bearings near the centre of the set are aligned horizontally
and after the shafts are in position, the outer bearings are raised to bring the coupling into alignment.
On a large turbine the outer bearings may be about ½ inch above the level of the central bearings.










The turbine (D5 unit’s) alignment is to be made with the bearings no 3 and 4 are set to a level line. The
no 1 bearing is set 1mm above the level line through the low pressure section bearings. The no 2
bearing is set 0.25mm low to compensate for vertical thermal expansion. This alignment will make the
HP rotor coupling half low to the LP rotor half by 0.47mm (the periphery check will show a 0.94mm
difference). The faces will be with an expected 0.01mm opening at the bottom.












Next there is the thermal growth that occurs at turbine components that will effect where the rotor sits.
Thermal growth will cause the rotor to rise up or come down at a bearing which will move the entire
alignment line. The catenary curve and the thermal growth have to be considered with the turbine
completely assembled.

44. 6 July, 2007 Page 49 of 58 Syed Saleem Ahmed
Alignment of turbine internal parts

against common base line through piano wire

Rotor disassembly readings should always be performed before the rotor is removed from the turbine.
These readings are taken at two points that are immovable, this way the rotor can be installed back into
the turbine the same way it was taken out. The oil deflectors are ideal positions to take the rotor
disassembly readings.

o Calculate wire sag of every location, where adjustment to be needed.
F = W × X × (2L – X)
2P
w = weight per unit length of piano wire in kg/mm.
P = hanging weight (Example: 13.608kg)
L = Length from 1st pulley to center of piano wire.
X = Distance from measuring point to nearest pull.
F = Wire sag.

o Align piano wire on bearing # 1 and 4 supporting standard with respect to catenary value + half of
bearing clearance + wire sag. Readings on bearing # 1 and 4 oil deflector housing should be noted
(During the diaphragm alignment the line can always be referenced to the oil deflectors)
Example:
Readings Point of
Measurement
Wire
Sag LHS BOT RHS
Difference After sag correction
4.98 4.93 0.05 0.00 0.05 0.00 Bearing # 1 standard
0.99
2.74 -2.19 -1.20

o Align all 4 bearing rings (without tilting pads). After alignment, examine the contact of the pads with
the bore of the supporting standard. If pad contact, is not sufficient. Scrape the pads until at least
80% contact is obtained.

o Take all long wire measurements from alternate diaphragms, gland housings and casings as tops-
off and tops-on state. The diaphragms should all be moved to left side of the shell. Generally the
tops-off data will show the diaphragm sitting higher than the tops-on data.

45. 6 July, 2007 Page 50 of 58 Syed Saleem Ahmed
Alignment of turbine internal parts

against common base line through piano wire


‰ Base line (HP/IP) = Difference with & without upper parts + Correction of rotor up thrust
+ Rotor sag + Diaphragm ovality/2.

Examples: (Note: Algebric values should be used)
Point of
Measurement
Difference without &
with upper parts
Correction of
rotor up-thrust
Rotor
sag
Diaphragm
ovality
Base
line
Stage no. 09
Diaphragm
(-1.63) - (-1.34) = -0.29 - 0.15 0.11 N/A - 0.33
Gland no. 02 0 -0.15 0.20 N/A + 0.05
Stage no. 13
Diaphragm
(-1.07) - (-0.69) = -0.38 - 0.15 0.15 N/A - 0.38
Stage no. 15
Diaphragm
(-1.07) - (-0.69) = -0.38 - 0.15 0.13 -0.13 - 0.53


‰ Base line (LP) = Rotor sag + Effect of vacuum on rotor + Correction of rotor up thrust
+ Diaphragm ovality/2

Example: (Note: Algebric values should be used)
Point of
Measurement
Rotor
sag
Effect of vacuum
on rotor
Correction of
rotor up-thrust
Diaphragm
ovality
Base
line
Stage no 18
Diaph. (L.A)
0.20 + 0.20 - 0.05 N/A + 0.35
Stage no 20
Diaph. (L.A)
0.16 + 0.20 - 0.05 + 0.12 mm + 0.43

‰ Align short wire for HP/IP on bearing # 1 and 2 and for LP on bearing # 3 and 4 rings with respect to
half of bearing clearance + wire sag.

‰ Align all HP/IP/LP internal’s according to base line.

‰ A final check of the diaphragm to rotor clearance will be performed when the rotor is installed. All of
the diaphragm centers have to sit at a certain horizontal and vertical deviation from a reference line
to within 0.12 mm when the rotors are in and the turbine is closed up. Remember that the gap on
the bottom may be smaller than what is acceptable, but this gap will increase in the tops-on state.

46. 6 July, 2007 Page 51 of 58 Syed Saleem Ahmed
Shaft Seal Packings
Retainer design

Three different designs to serve two important functions.
The three designs are:
1) Pin design 2) Rivet design 3) Key design

The two important functions provided are:
1) To prevent rotation of the packing ring segments during turbine operation.
2) To hold the ring segments in the upper half of the diaphragm as it is lifted from the turbine.

Pin Design
This design is used in both cast steel or fabricated steel
diaphragms. The pin is secured by peening so that the
diaphragm material flows over the large end of the
12mm diameter pin hole on the diaphragm’s admission
side. The 6mm diameter hole on the discharge side.













Rivet Design
This design is used entirely in cast iron diaphragms,
Since the cast iron is not very ductile and will not
adequately flow when peened.
Both ends of the mating hole in the diaphragm should
be chamfered. The rivet always installed so that the
headed end is on the diaphragm’s admission side, and
the flared end is on its discharge side.










Key Design
This design is employed in diaphragms which have
more than one ring packing segments and/or where
dimensional limitations will not permit the use of the
pin or rivet.

47. 6 July, 2007 Page 52 of 58 Syed Saleem Ahmed
Shaft Seal Packings
Metallic labyrinth type

The packing is a steam throttling device consist-
ing of stationary and rotating teeth arranged
concentrically with small radial clearance.

The rotating element consists of step recesses
machined directly in the turbine rotor, while the
stationary elements are segment rings provided
with teeth. The packing rings are basically of
two types: (1) For low temperature service
(below 400
O
C) the ring segments are made of
BHT leaded bronze with teeth machined integral
in the material; (2) For high temperature service
(above 400
O
C) the ring segments are steel with
moly Ascoloy teeth inserted.


















The packing segments are spring backed. These springs are made of Inconel-X material and are of
sufficient strength to hold the segments in place. The segmental spring-back construction is to
provide additional clearance in case the rotor should become distorted due to some operating
condition.

Packing springs should be inspected for cracks and straightness and any springs that are covered with
scale should be cleaned with 220 grit aluminum oxide prior to inspecting. Packing springs should
be red-dye tested. Bent springs should be replaced. The hook fit between the segment and the
spring should be checked during assembly, insuring there is at least 1.50mm of engagement. If there
is no engagement, the springs should then be replaced with a design having a longer hook.

A typical packing spring is shown in Fig. and its
dimensions are designated by letters as follows
A- hook length
B- length
C- active length
D- width
E- thickness

48. 6 July, 2007 Page 53 of 58 Syed Saleem Ahmed
Shaft Seal Packings
Clearances

Butt clearance:
The ring segments are machined with proper joint clearance so as to provide complete circumferential
protection after the segments have expanded due to operating temperatures.

Diaphragm Packing Steam Seal Packing
Packing
Ring
Dwg.
Clear.
Max.
Clear.
Hook
Dia
Material Packing Ring
Dwg.
Clear.
Max.
Clear.
Hook
Dia.
Material
St-9 5.08 6.60 728 Brass Man. 1 Anc.1 4.24 5.76 550 Brass
St-8 5.40 6.92 728 Brass Man. 1 Anc.2 4.24 5.76 550 Brass
St-7 5.64 7.16 728 Brass Man. 1 Anc.3 5.32 6.84 697 Brass
St-6 0.32 1.84 728 S. Steel M an.1 Anc.4 5.32 6.84 697 Brass
St-5 0.32 1.84 728 S. Steel M an.1 Anc.5 5.32 6.84 697 Brass
St-4 0.32 1.84 728 S. Steel M an.1 Anc.6 5.32 6.84 697 Brass
St-3 0.32 1.84 728 S. Steel Man. 2 Anc.1 0.32 1.84 728 S. Steel
St-2 0.32 1.84 728 S. Steel Man. 2 Anc.2 0.32 1.84 728 S. Steel
St-10 0.32 1.84 728 S. Steel Man.2 Anc.3 0. 32 1.84 728 S. Steel
St-11 0.32 1.84 728 S. Steel Man.2 Anc.4 0. 32 1.84 728 S. Steel
St-12 0.32 1.84 728 S. Steel M an.3 Anc.1 5.32 6.84 688 Brass
St-13 0.32 1.84 728 S. Steel M an.3 Anc.2 5.32 6.84 688 Brass
St-14 0.32 1.84 728 S. Steel M an.3 Anc.3 5.32 6.84 688 Brass
St-15 5.16 6.68 728 Brass Man. 3 Anc.4 4.48 6.00 580 Brass
St-16 4.72 6.24 728 Brass Man. 3 Anc.5 4.48 6.00 580 Brass
St-17 - - - - Man.4 Fila 1 5.16 6.53 Brass
St-18 4.96 Brass Man.4 Fila 2 5.16 6.53 Brass
St-19 4.00 Brass Man.4 Fila 3 6.28 7.65 Brass
St-20 4.00 Brass Man.5 Fila 1 6.28 7.65 Brass
St-21 4.00 Brass Man.5 Fila 2 5.16 6.53 Brass


Man.5 Fila 3 5.16 6.53 Brass

Notes:
• “Drawing” clearance and “Maximum” clearance are total clearance for the entire 360
0
.
• Fossil Turbines – Maximum clearance = Drawing clearance + 2πD
3000
Nuclear Turbines – Maximum clearance = Drawing clearance + πD
3000
D = Hook Diameter (mm)

49. 6 July, 2007 Page 54 of 58 Syed Saleem Ahmed
Steam Seal Packings
Clearances

Radial clearance:
Prior to packing segment removal measure and record the radial clearance,
top, bottom, left and right sides, between packing ring and rotor at points
1, 2, 3 and 4. The radial clearance should be taken on a short tooth.



The radial clearance is corrected by “dropping” the hook diameter. The teeth of the steel packing ring
can be sharpened as many times as necessary until the radial height of the tooth has been reduced by
0.75mm. After 0.75mm has been removed from the teeth due to sharpening, the teeth should be
replaced.

Spill strip and
Diaphragm packing
Gland packing
Stage
no’s
R Z W Packing Ring R
St-9 0.35 1 1.3 Man.1 Anc.1 0.65
St-8 0.35 1 1.3 Man.1 Anc.2 0.65
St-7 0.35 1 1.3 Man.1 Anc.3 0.35
St-6 0.35 1 1.3 Man.1 Anc.4 0.35
St-5 0.35 1 1.3 Man.1 Anc.5 0.35
St-4 0.35 1 1.3 Man.1 Anc.6 0.35
St-3 0.35 1 1.3 Man.2 Anc.1 0.50
St-2 0.35 1 1.3 Man.2 Anc.2 0.50
St-1 - 1 1.3 Man.2 Anc.3 0.50
St-10 0.75 1 1.3 Man.2 Anc.4 0.50
St-11 0.35 1 1.3 Man.3 Anc.1 0.35
St-12 0.35 1.3 1.6 Man.3 Anc.2 0.35
St-13 0.35 1.3 1.6 Man.3 Anc.3 0.35
St-14 0.35 1.3 1.6 Man.3 Anc.4 0.65
St-15 0.35 1.3 1.6 Man.3 Anc.5 0.65
St-16 0.35 1.3 1.6 Man.4 Fila 1 0.65
St-17 N/A 1.5 1.8 Man.4 Fila 2 0.65
St-18 0.90 1.5 N/A Man.4 Fila 3 0.65
St-19 0.90 1.5 N/A Man.5 Fila 1 0.65
St-20 0.90 2.0 N/A Man.5 Fila 2 0.65
St-21 0.90 5.8 N/A

Man.5 Fila 3 0.65

Field Tolerances: Z = ±0.40, W = ±0.25, R = +0.25, -0.15


Notes
• When standing at the front standard and looking toward the
generator. The side to the left is always the left side of the turbine.
• The packing assemblies are numbered starting from the turbine
end of the machine and working toward the generator. The
packing rings in one assembly are also numbered from the turbine
end of the assembly working toward the generator.

50. 6 July, 2007 Page 55 of 58 Syed Saleem Ahmed
Steam Seal Packings
Clearances

Axial clearance: Occasionally, adjustment of the axial location of a packing ring is required to
accommodate rotor movements during operation and eliminate axial rubs between the rotor and the
packing ring. The maximum amount of axial correction, due to design stress consideration, is 1.50mm
and this is accomplished by machining stock off the seal face and the downstream side of the hook.
After the material is removed from the downstream side of the ring, crush pins are needed on the
upstream side of the ring to avoid the possibility of inserting the ring backwards. The crush pins should
be machined so their projection is equal to the amount of material removed from the downstream side.
The crush pins should be B5F5B machine screws. There should be four pins for each segment, each
41mm in from the end. Two pins should be on the hook and two more on the neck in the same position.

51. 6 July, 2007 Page 56 of 58 Syed Saleem Ahmed
Steam Seal Packings
Clearances

Axial clearance:

Stage
no’s
X Y Z-3 Z-4
Fig.
no.
N L
Packing
Ring
X Y
Fig.
no.
St-9 7.5 4.5 - - 3 2.6 3. 4 Man.1 Anc.1 4.75 2.75 2
St-8 7.5 4.5 3.0 5.2 6 3.6 3.6 Man. 1 Anc.2 4.75 2.75 1
St-7 7.5 4.5 2.8 5.0 6 3.4 4.0 Man. 1 Anc.3 5.75 3.75 1
St-6 7.5 4.5 3.8 5.0 6 3.4 4.2 Man. 1 Anc.4 5.75 3.75 1
St-5 7.5 4.5 4.0 4.7 6 3.6 4.4 Man. 1 Anc.5 5.75 3.75 1
St-4 7.5 4.5 4.0 2.9 - 4.6 4.6 Man. 1 Anc.6 5.75 3.75 1
St-3 7.5 4.5 4.2 3.0 - 4.4 4.8 Man. 2 Anc.1 7.05 4.05 2
St-2 7.5 4.5 4.4 3.5 - 5. 0 5 Man.2 Anc.2 7.05 4.05 2
St-1 N/A N/A - - - 5.0 5 Man.2 Anc.3 7.05 4.05 2
St-10 8.25 10.85 4.9 8.0 - 4.8 6.2 Man. 2 Anc.4 7.05 4.05 2
St-11 8.25 10.85 5.2 8.7 - 5.0 6.4 Man. 3 Anc.1 11.05 8.05 2
St-12 8.25 10.85 5.4 9.0 - 5.2 6.6 Man. 3 Anc.2 11.05 8.05 2
St-13 8.25 10.85 5.7 9.2 - 5.5 6.8 Man. 3 Anc.3 11.05 8.05 2
St-14 8.25 10.85 5.9 9.5 - 5.8 7.2 Man. 3 Anc.4 11.05 8.05 2
St-15 8.25 10.85 - - 3 6.2 7. 4 Man.3 Anc.5 11.05 8.05 1
St-16 8.25 10.85 - - 3 6.4 7. 6 Man.4 Fila 1 15.20 10.20 2
St-21 LT N/A N/A - - - 57.4 23.4 Man.4 Fila 2 15.20 10.20 1
St-20 LT N/A N/A - - 4 20.0 20.23 Man.4 Fila 3 15.20 10.20 2
St-19 LT N/A N/A - - 4 20.6 19.83 Man.5 Fila 1 17.7 7.70 1
St-18 LT N/A N/A - - 1 13.9 20.41 Man.5 Fila 2 17.7 7.70 2
St-17 LT N/A N/A - - 1 13.9 17.55

Man.5 Fila 3 17.7 7.70 1
St-17 LA N/A N/A - - 1 7.1 10.75
St-18 LA N/A N/A - - 1 7.1 13.31
St-19 LA N/A N/A - - 4 13.6 13.63
St-20 LA N/A N/A - - 4 13.8 14.13
St-21 LA N/A N/A - - - 44.7 17.3

Field Tolerances: X, Y = ±0.75, Z-3, Z-4 = ±0.50, N, L = ±0.25














Notes

• Rotor must be blocked away from the generator as far as thrust allows when checking clearances.
Using a force of 5 Tons for one rotor or 10 Tons for all rotors and generator field.
• Wedge all packings in direction of steam flow. The “X” clearance is always on the turbine side of the
long tooth and the “Y” clearance on the generator side.
• Refer to the clearance diagram drawing for detailed clearance locations.
• All clearances must be taken with the HP rotor positioned on stage 2 and LP rotor positioned on
stage 17 LT with an axial “N” clearance.
• Clearances at “N” on each stage wheel are taken at the left horizontal shell joint.

52. 6 July, 2007 Page 57 of 58 Syed Saleem Ahmed

Differential Expansion of Steam Turbine Generators

Differential Expansion
Differential Expansion monitoring measures the change in axial clearances between the machine rotor and
casing caused by thermal changes natural in most machines. Excessive differential expansion can produce
one of the most catastrophic failures of rotary machinery (resulting in very expensive repairs and possibly
machine replacement).

There are two basic machine conditions that can occur during machine operation:

The Rotor Long condition
When high temperature steam is applied to the
machine (as during a machine startup), the rotor
thermally expands at a faster rate than the
machine casing because of its smaller mass
and different metallurgy.
If the rate of rotor growth in relation to the rate
of casing growth is not controlled, the moving
and stationary parts of
the machine come into contact. To control the
system’s rate of expansion, steam temperatures
and flows are controlled. Turbine soak periods
further control the expansion process.
“ Soak period” is a term that describes the
amount of time, and the machine speed, at
which the machine is held to allow the machine’s casing
casing to thermally expand and catch up to
the rotor assembly. During these periods, steam temperatures and flows are strictly maintained to prevent
diminishing the machine’s internal clearances.
Care should be given to seal steam temperatures applied to the unit, as even excessive seal steam
temperatures have been known to cause excessive differential expansion values during machine startup
procedures.

The Rotor Short condition
The rotor short condition typically occurs during
steady state operation. “Rotor short” describes
the condition where the rotor shirks with respect
to the machine case.
Cooling effect is typically caused by a severe
reduction in steam temperature or flow within
the machine and is normally referred to as
“quenching the rotor”. If this quenching action
exceeds acceptable limits, the rotor short
condtion can occur very quickly. Acceptable
limits in the rotor short direction are typically
small when compered to those for the rotor
long dircetion. The rotor short condition how-
ever is designed to have smaller clearances to
increase unit efficiency, as this condition typically
should not occur during normal opertion.

53. 6 July, 2007 Page 58 of 58 Syed Saleem Ahmed
Differential Expansion of Steam Turbine Generators


Rotor Expansion
To allow unrestricted growth of the turbine rotor assembly, the machine rotor is restricted from axial
movement within the machine case at one fixed point and allowed to move from that point. The point at
which this attachment is made is at the machine’s thrust bearing. Figures 3 and 4 represent two examples
of turbine generator machine trains.
In Figure 3, the thrust bearing is located in the machine’s front standard or pedestal. The rotor assembly is
restricted from movement at this point. In this example all rotor expansion is toward the generator. Only
one differential expansion measurement is required due to unidirectional growth of the entire rotor
assembly.
In Figure 4, the thrust bearing is located at pedestal number two, between low pressure and high pressure
casings. The rotor assembly is restricted from movement at this point. The high pressure rotor grows or
expands towards the governor end of the machine, while the low pressure rotor expands or grows toward
the generator end of the machine from the thrust assembly.
Due to the bipolar growth of the rotor assembly, two differential expansion measurements are required.
One measuring high pressure rotor differential expansion and one measuring low pressure differential
expansion.