The American Society of Mechanical Engineers

W A RREN —M O DERN LARGE STEAM T U R B IN ES FO R G EN ERA TO R D RIVE 69
Tests indicate th at for a given bucket wheel the amplitude of
resonant vibration generally increases in proportion to the load
being carried. Statistical experience accumulated by an analysis
of results obtained on about 10,000 wheel-years of operation
indicates quite definitely th at the liability of reaching vibration
stresses in the buckets on such wheels which will produce failure
is reduced to the vanishing point when the buckets are made
strong enough in relation to the loads carried. I t is essential, of
course, that the other details of design and construction be uniformly
and properly carried out.
R o t o r s
The rotor, aside from the buckets, might be considered as the
aggregate of the shaft and the wheels which carry the buckets,
together with the rotating elements of the packings.
Rotors for 1800-rpm turbines are generally made of a shaft
with wheels separately shrunk on and packing rings. The
wheels in the low-temperature section are keyed to the shaft.
In the higher-temperature sections it has been found necessary
to be sure th at the wheels are able to leave the shaft under the
combined influence of centrifugal force and a sudden increase in
temperature, and at the same time stay central and drive the
shaft. This is accomplished by a so-called “pin bushing.”
The bushing is keyed to the shaft, and the wheel hub is attached
to the bushing itself by a number of radial pins.
Two refinements which have resulted in great improvement
in the operation of this type of turbine are: (1) Making the
keys discontinuous, th at is, a separate key for each wheel, which
does not extend under the packing ring between adjacent wheels;
and (2) the undercutting of the packing sleeves to reduce the
heat passed into the shaft due to accidental rubbing of the packings.
3600-Rpm T u r b i n e s
Rotors for 3600-rpm turbines are generally machined from a
solid forging. This obviates the difficulties associated with
separate wheels and packing sleeves, and secures a short rigid
rotor which permits operation of such high-speed turbines below
the calculated critical speed. These rotors have shown great
stability under changing load and tem perature conditions.
This rotor construction also possesses an im portant inherent
advantage in the ease with which it is heated due to its extended
surface. This gives it an autom atic “end loosening” effect in
that during starting the axial clearanes are opened up without
change of the thrust-bearing position. The close proximity of
the diaphragms prevents the rotor’s cooling faster than the shell
on shutting down and so prevents a tightening of the clearances.
A refinement in manufacture which has materially contributed
to smoother operation of turbines under service conditions
F i g . 2 5 T y p ic a l H o r iz o n t a l J o in t F l a n g e a n d B o l t in g
F i g . 2 6
M a g n i f i e d P h o t o g r a p h o f T u r n e d a n d M i l l e d T h r e a d s
o n S t u d
is the so-called “heat stabilizing” of the shafts. Every shaft or
rotor on the turbines under consideration receives this test and
treatm ent which is as follows:
After a shaft or rotor is rough-machined it is put in a special
lathe in an electrically heated furnace and slowly heated while
being revolved, and it is carefully indicated for “runout” during
the heating. M ost normal shafts will increase their runout
5 to 15 mils in this process until a tem perature of 700 F to 900 F
is reached, a t which some kind of a readjustm ent takes place,
probably on the machined surfaces, which permits the shaft to
come back to a “hot runout” of not more than 1 to 2 mils more
than its “cold runout,” and on being cooled its runout does not
change greatly. On putting such a shaft through a second heating
its runout will not go through this cycle, but will go to the
hot-runout condition directly. Some shafts will have a large
hot runout which cannot be corrected this way. Such a shaft or
rotor m ust be reheat-treated, and if this does not correct the
condition, it m ust be rejected, otherwise the balance of the tu r­
bine in w'hich it is used would be seriously different in the hot
and cold condition and would change w ith load.
All wheel and solid rotor forgings m ust be metallurgically
sound, and are checked following final machining by the magnaflux
method.
B o l t i n g F l a n g e s
The horizontal flanges by which the upper and lower halves
of the shells are joined are perhaps one of the most im portant
parts of the turbine design. This joint m ust remain tight despite
tem perature and pressure conditions and changes. I t must
permit of being dismantled, and reassembled, and still remain
tight; and this w ithout the aid of gaskets.
Fig. 25 shows a typical shell flange. The bolts are put in as
close to the inner diam eter as possible, and still leave sufficient
metal between the inner flange wall and the bolt hole, generally
around 1 in. as a minimum. This requires a flange of great depth.
The flange is given a sizeable “toe” or portion beyond the bolt.
Studies carried out several years ago on rubber models indicated
definitely th at this configuration resulted in relatively low bolt
stresses while a t the same time it kept the joint tight a t the
cylinder bore.
The nuts are made cylindrical, w ith the “hex” on top, which
permits a smaller hex and hence smaller wrenches. This in
turn permits a closer spacing of the bolts than is possible with
nuts of the conventional type.
Joints of this type when properly “scraped to fit” appear to
hold steam tight when the stress in the bolts is about IV 2 to 2
times th a t required to overcome the bursting pressure on the