DESIGN AND OPERATING
CHARACTERISTICS OF JOURNAL
BEARINGS

Practically all major bearings used in high speed
equipment of the turbine type are sleeve type bearings operating with small clearances to the rotating
shaft. They require a continuous supply of clean oil
at normal pressure ranging from 10 to 25 psi.

The primary motion is that of one surface sliding
over another. There are two basic types of lubrication for this bearing.

Boundary lubrication applies when the lubricant
acts primarily to fill surface discontinuities or depressions on the rotating shaft. The supporting
bearing acts basically as an agent for making the
surfaces slippery to reduce fiction. Intermittent
metal-to-metal contact occurs and the load and
shaft speeds are limited to prevent seizure and failure of the bearing by scoring.

Fluid film lubrication applies when the lubricant is
utilized as a high pressure medium to actually lift
the shaft and form a layer or film of oil of a definite
thickness, keeping the rotating shaft surface and the
bearing surface separate at all times.

This can occur when two sliding surfaces form a
converging wedge and are supplied with a continuous supply of oil. Note carefully that cylindrical
sleeve bearings inherently provide a converging
area as oil is drawn under the shaft due to the difference in diameters of the shaft and the bearing and to
the fact that the shaft or journal tends to rest in the
bottom of the bearing. The viscous drag of the rotating shaft at high speed tends to draw more lubricant
into the decreasing space at the bottom of the journal than can be passed under the journal causing an
increasingly higher pressure to be developed sufficient to lift and sustain the shaft on the oil film at all
times in normal operation.

This is sometimes referred to as the pumping action
of the shaft. Resulting oil films may vary from
0.001 to about 0.006 inch at their point of minimum
thickness. Oil films for most turbine applications
probably are 0.001 to 0.002 inch in minimum thickness. The oil film thickness will generally depend
upon the load, viscosity of the oil, shaft speed, and
to some extent, bearing clearance. Thus, scratches, bumps, foreign particles, etc., of
greater height or diameter than the minimum oil
film thickness will cause a breakdown in the oil
film and resultant scratching or scoring of the shaft.
If sufficient heat is generated a bearing failure will
result. Note, however, that a shaft at rest squeezes the oil
from under it and metal to metal contact exists, or at
best boundary type lubrication exists. Therefore, in the first quarter of a turn, the shaft
tends to "climb" the bearing in the direction against
rotation due to the higher frictional forces in the absence of the oil film at the bottom of the shaft and
some wear results. This first quarter of a turn represents about the only possibility of causing wear in a
normal bearing with a clean oil supply.

Note from Figure 1 that at normal speed and load, 
the combined effect of the journals' weight and hydraulic pressure of the oil film is to move the journals slightly off the center of the bearing in the
direction of rotation. The length of the arrows
shown represents the magnitude of the oil film
pressure and will be seen to be a maximum slightly
to the bottom side of the point of minimum film
thickness.

Note also that oil film pressure covers about one-half of the shaft's circumference, being greatest off
the bottom center in the direction of rotation, and
that a slightly negative pressure actually exists just
past the minimum oil film point compared to the oil
supply pressure. Note that little or no pressure is developed in the upper half of the bearing except at the downcoming
side of the journal. As a general rule of thumb, the maximum film pressure will be approximately three times the unit
loading pressure in the bearing. This means that for a bearing loading of 200 psi, the maximum oil film pressure might be 600 psi for a
bearing with good alignment.

For misaligned bearings, the maximum oil film
pressure may increase considerable with attendant
higher oil film temperatures. The oil flow to this type of bearing is usually limited by an orifice of a size to result in an overall temperature rise of about 30°F in the oil flowing
through the bearing at its rated load.

The maximum design limit for the maximum oil
film temperatures may be as high as 240°F. A bearing will probably operate satisfactorily with a maximum oil film temperature as high as 300°F. 350°F
is certain to cause failure. Present designs result in
most bearings operating with maximum oil film
temperatures of 160 to 200°F.

   Fig. 1 Action of Cylindrical Sleeve Bearings

Of the total oil flowing through the bearing a relatively small amount is used to maintain the oil film
and the rest simply tends to wash some of the heat
from the upper half of the bearing journal. If no orifice were used in the oil supply line, the bearing
would, itself, limit the flow of oil; however, the increased flow of oil would not affect oil film temperature in any significant amount.

Without the bearing orifice, the flow would increase more than desirable if bearing clearances become excessive and in the event of failure of the
bearing, would increase enough to possibly result
in oil starvation at other bearings. The flow of oil through a bearing of the type under
discussion will increase about as the cube of the
bearing clearance; that is, doubling the bearing
clearance would cause about 8 times the flow. Increased bearing clearance would have some effect in reducing bearing temperatures but would
have little effect on the power losses absorbed in the
bearing. Bearing clearances of specified amount will result
in best operation of the bearing and have the least
adverse effect on the overall oil system.

The proper bearing clearance will depend upon the
surface speed of the journal primarily; i.e., the journal speed and diameter. A journal surface speed of
over 6000 feet per minute is considered a high
speed journal. For a 3-inch shaft this would be
approximately 7600 rpm. In a 6-inch shaft this
would be approximately 3800 rpm; and in a 10-inch
shaft, approximately 2200 rpm.

Optimum bearing clearances are as follows:

0.001 inch per inch of shaft diameter up to
2000 rpm

0.0013 inch per inch of shaft diameter up to
4500 rpm

0.002 inch per inch of shaft diameter over
4500 rpm

Clearances may increase to about twice their minimum recommended clearances without seriously harming bearing performance; however, it is not
wise to allow bearings to wear to this degree because of the effect upon the oil system, alignment to
connected equipment, and close radial clearances
throughout the turbine. Axial grooves are sometimes used in bearings serving to spread the oil uniformly the length of the
bearing but should not be in the loaded area. Circumferential grooves are used mainly near the
ends of the bearings to serve as collection points for
drain oils and routing their flow or for sight flows.
These grooves control drain oil, making it easier to
prevent oil leakage from the bearing. A large circumferential groove in the center of the
upper half is used on bearings known as "overshot"
bearings merely serving to pass increased oil flow
for cooling the journal. This is not a common bearing. Circumferential grooves are seldom used in the
lower half and when used are usually meant to decrease the bearing area deliberately for the purpose
of increasing bearing loading.

The tendency for modern turbine design has been to
shorten bearing lengths and raise bearing loadings
slightly to increase the margin against instability at
higher speeds. Design bearing loadings vary from a
minimum of 100 psi to a maximum of about 250
psi. Bearings are seldom built with an effective length
greater than the shaft diameter, and it is usually less. A bearing loading is generally defined as being the
weight of the journal divided by the projected area
of the bearing. The projected area is the bearing diameter multiplied by the babbitted length (less
width of any grooving). Bearing stability has become a factor of concern on
higher speed units and many bearings have been
developed to increase stability.

For some conditions of load, oil viscosity, and
speed a bearing may become unstable and the journal become affected by what is known as an oil
whip condition.

Oil whipping is characterized by an additional and
undesirable cycle or half-circular motion of the
shaft usually at about one-half running speed frequency or, if the first critical speed is anywhere
near one-half speed, at the first critical speed. Large
vibrations result, particularly at critical speed frequency.

Oil whipping is due to hydrodynamic forces within
the bearing and is not fully understood. It is, however, associated with one or a combination of lightly loaded bearing, high speeds, and oil viscosity.
The hydrodynamic frictional forces on a lightly
loaded shaft seem to cause the shaft to set up an independent motion connected with climbing part
way up the bearing and falling back at frequencies
generally connected with the shaft critical speed.

When oil whip starts, it is thought to begin at frequencies about equal to twice the value of the first
critical speed. As whipping develops it excites the
natural vibration frequency of the shaft and
changes to or near the first critical speed frequency.
Sometimes both frequencies persist. For this reason, the rotor critical speed is designed
to be appreciably above one-half speed thus discouraging an oil whipping condition.

Bearings such as elliptical bearings, pressure pad; 
axial groove, 3-lobe, and tilting pad have been developed for their stable characteristics and will be
described in their basic details.

Tin-based babbitt metal is the prevailing bearing
bore material for turbine bearings because of its superior anti-seizing properties and affinity for absorbing oil into its surface. A good grade of babbitt
for turbine bearing applications contains about 88
percent tin, 7 1/2 percent antimony, 4 percent copper and the very small balance of lead. Tin based
babbitt will melt at about 460°F. It will begin to
soften above 240°F.

Some application of aluminum alloys, particularly
with tin, have been made successfully although avoided for turbine application due to difficulties in
pouring sound bearings, inability to operate satisfactorily in the presence of even minute particles of
dirt or other contaminants and their characteristic
of causing extensive journal damage in case of failure. The major advantage of this material for bearing application is its high fatigue strength at high
temperatures.

TYPES OF JOURNAL BEARING
CONSTRUCTION

Bearings may be generally classified as thick liner
or thin liner bearings.

The thick liner bearing is normally made as a part of
a larger metal housing with a babbitted bore as the
bearing surface. This type bearing has the major advantage of being suitable for complete repair after
long wear or damage in most cases. It can be re-poured and restored to its original bore and configuration or altered to different bore dimensions to
suit altered journal sizes.

The thin liner bearing is now in wide use consisting
of two parts where the thick liner bearing uses only
one. The bearing is made in a thin shell liner carrying the babbitted bore which inserts into the liner
housing. The thin liner is then the only part requiring replacement and is less expensive for purchasing and storing replacements.

The thin liner is subject to variation in bore shape
during assembly. It may, for example, have a tendency to pinch inward at the joints. In order to
counteract this tendency, thin shell liners are purposely sprung outward slightly by about 0.002 to
0.007 inch to insure that this does not occur and that
the liner conforms to the bore of the housing in
which it is installed. The thin liner cannot be babbitted satisfactorily due to distortion and is regarded as an expendable item.

Thin liner designs perform equally as well as the
thick liner bearings and have been applied for shaft
sizes up to 9 inches in diameter.  A bearing may fit into its casing with a cylindrical
tongue and groove fit thus having its position and
alignment determined by the bore of the casing.

Many bearings are manufactured with a spherical
shaped surface on their outer diameter intimately
matching a similar shape of the casing bore in
which it is installed. This allows a certain amount of
adjustment in the alignment of the bore of the bearing when it is installed to suit the shaft axis.

Many larger bearings incorporate a double spherical fit of this type obtained by an additional spherically constructed ring fitting between the bearing
and the bearing casing, sometimes called a ball liner. This, of course, allows more flexibility of bearing alignment when the spherical surfaces are kept
in good condition and the bearing is properly
installed.

Bearing caps which bolt over bearing housings are
manufactured so that the bore exactly conforms in
shape to the outer diameter surface of the bearing
housings; however, the horizontal joint surface of
the upper half is purposely machined back to insure
that it "pinches" or firmly presses the bearing housing which should not be free to move after assembly. This "pinch" fit will be between 0.001 and
0.005 inch depending upon the size of the bearing.

A QUESTION FREQUENTLY ASKED

Why should any great care be taken in bolting
spherical or ball seated bearings since they are self-aligning?

The spherical or ball seat bearing allows the bearing to align to the true axis of the shaft within reasonable variations when assembling the bearing if
done properly. Once the bearing cap is bolted, the
bearing is "pinched" or held firmly in a fixed position relative to the line of the shaft. In fact, this bearing can be unintentionally misaligned in the
process of assembly by installing bearing cap bolts
in the improper sequence and manner.
 

JOURNAL BEARING DESIGNS AND
THEIR CHARACTERISTICS

Cylindrical Bearings

A plain cylindrical bearing is a universally used
type of sleeve bearing, usually without any special
grooving except at the joint to introduce lubricant
into the bearing. This bearing has a relatively high load capacity;
however, it is the least stable bearing of the major
types when operated at high speeds and relatively
light loading. Cylindrical bearings operate with fairly high oil film temperatures at high surface speeds, considerably more than elliptical bearings for example. Larger high speed shafts generally will not incorporate cylindrical bearing design for this reason.

An overshot cylindrical bearing design, previously
explained, is sometimes used at higher speed applications providing considerably more oil flow for
cooling the shaft and resulting in some reduction of
power losses in the bearing. Wide application of cylindrical bearings is made in
equipment which is subject to changes in the direction of journal loading for different load conditions
such as reduction or speed increasing gears. All the foregoing material regarding bearing characteristics, design, film pressures, etc. is generally
applicable to cylindrical bearings.

Modern turbines tend to use elliptical bearings for
larger shaft sizes and axial groove or pressure type
bearings for smaller sizes of shafts and light load
operating conditions, but cylindrical bearings are
applied very frequently when operating conditions
are suitable.

Bearings invariably incorporate oil seals or bearing
deflectors to prevent oil leakage along the shaft and
down the faces of bearing housings. These may consist simply of a series of close fitting rings or
teeth installed around the shaft at each end of the
bearing housing usually of bronze or aluminum or
other material with preferred rubbing characteristics (less heat and rubs out easily). A leaded bronze
material is frequently used. In any case, the bearing
deflectors will present a high resistance path to oil
leakage and incorporate drain grooves in the bottom half to drain oil collected in the grooves between teeth back into the bearing housing. These
grooves, of course, should not be allowed to become plugged with foreign material such as joint
compounds. Deflector clearances to the shaft will
run 0.008 to 0.010 inch on the radius.

In this connection, the turbine shaft often incorporates a stepped shoulder or a thick vane type projection slightly larger than the journal which acts as a
"slinger" to throw oil from the shaft into the bearing
casing to reduce oil leakage along the shaft.

Oil deflectors, being of different material from the
steel bearing housings, sometimes begin allowing
oil leakage at their horizontal joint or around their
fit into the bearing housings after long service due
to the greater expansion characteristics of the deflector material causing it to be plastically crushed
in a small degree by the stronger and smaller expanding bearing housing eventually resulting in a
distorted and poorly fitting deflector. It is usually
best to renew the deflector or field modifications
can be made to compensate for this condition if the
problem is recognized.

Elliptical Bearings (See Figure 2)

The bore of the elliptical bearing, although not a
true elliptical shape, has this general characteristic
since its horizontal diameter is appreciably greater
than its vertical diameter.

This is best visualized by examining the method
used in manufacturing the bearing. Shims are
placed at each joint between the upper and lower
halves and the bearing is then bored with a true cylindrical shape. Following the boring operation, the
shims are removed; resulting in a bearing bore with a horizontal clearance almost twice the vertical
clearance when installed.

Oil flow through the elliptical bearing is rather substantial, resulting in a relatively cooler running
bearing. It is not sensitive to changes in bearing
clearance and has high load capacity. Oil film temperatures are generally less than other commonly
used designs.

The elliptical bearing is usually applied where a
combination of large shaft sizes, high speed, and substantial load is found. The elliptical bearing may also be used with an
"overshot" design as explained earlier for cylindrical bearings. The effect of the elliptical shape is to increase the
angle of convergence (wedge characteristic) between the bearing and shaft surface as will be seen
by reference to Figure 3, causing a stabilizing pressure wave in opposite quadrants making this bearing more stable than the cylindrical bearing.

Note that the journal will locate off the center of the
bearing as determined by the combined effects of
the lower oil film and the pressure waves at opposite quadrants. Note that the pressure characteristic in the lower quadrant of

       Fig. 2 Elliptical Bearing

the bearing would be expected to be more pronounced than in the upper
quadrant due to the higher convergence of oil flowing into the bottom of the bearing. This will cause
the journal to ride slightly higher in the bearing.

Pressure Type Bearings (See Figure 4)

The pressure bearing is a modified form of the cylindrical bearing design originally designed and applied to provide an additional stabilizing pressure
in the upper half, tending to hold the shaft at its normal position in the lower half in a more stable manner.

The pressure bearing is usually associated with
lightly loaded shafts or shafts more apt to be subject
to light loads due to changing alignment. It is used
mainly on the smaller shaft sizes.

A rather wide groove is provided in the upper half
bearing which, in contrast to usual grooving, is terminated in a sharp shoulder or dam about 45 degrees beyond the top vertical centerline in the
direction of rotation.

Note that oil is fed at both joints of the bearing
which may be thought of in a general manner as
separate oil supplies for the lower half oil film and
the upper half pressure pad or groove.

 Fig. 3 Operating Characteristics of the
      Elliptical Bearing

The pressure pad dam should not be beveled or
have its square edge broken or rounded with bearing scrapers or equivalent which is quite frequently
done.

The viscous drag of the shaft causes oil to flow
along the groove until it is abruptly stopped by the
dam resulting in a relatively high static pressure being developed to resist undesired vertical instability of the shaft.

The resulting upper half oil pressure is greatest over
the dam gradually reducing to smaller values away
from this point. Maximum pressure may be 100 psi
or more. The average pressure in the upper half is
more apt to be 20 to 50 psi.

The width and depth of the pressure groove may
vary depending upon the application, the width
usually being about one-half the bearing length,
and the depth 0.020 to 0.030 inch.

The edge of the pressure dam, being quite sharp, may tend to be damaged if the oil supply contains
foreign particles such as dirt.

Fig. 4 Pressure-Type Bearing

Note specifically that the pressure bearing is effective only for a given direction of rotation. Sometimes replacement bearings of wrong rotation are
installed through error.

Axial Groove Bearings (See Figure 5)

The axial groove bearing bore contains grooves
running parallel to the shaft.

The design may use either a cylindrical or elliptical
bore. The number of grooves may vary and generally are unequally spaced depending upon the application.

The effect of the axial grooves, as might be expected, is to produce an unequal pattern of oil film
pressures around the circumference of the shaft and
it will be noted from Figure 5 that oil is fed to the
bearing at each axial groove through holes serving
as oil supply orifices.

The effect of this design is to produce a bearing of
relatively high stability.

Fig. 5 Axial Groove Bearing

Depth of the grooves has little effect upon the bearing performance; however, the width of the
grooves is important since the groove width has a
direct effect on the load capacity of the bearing.
Consequently, the radius around the edges of the
grooves will similarly affect bearing performance
and should not be chamfered or rounded with bearing scrapers, etc.

Each groove utilizes a small triangular chamfer at
each end which controls the total flow of oil to the
bearing. Modification to these chamfers in any
form will, of course, have an immediate effect upon
oil flow and bearing drain temperatures.

Chamfer size variation will have very little effect
upon the oil film temperature however, as axial
grooved bearings operate with higher oil film temperatures and lower minimum film thickness in the
loaded areas inherently due to the discontinuity in
film pressures caused by the axial grooves.

The load capacity, although lower than many other
bearing designs, is nevertheless much higher than
the maximum allowable design loadings used in
bearing applications.

Note in Figure 6 that the effect of the axial grooves
is to break up the pattern of film pressure with two
resulting stabilizing pressures at each side of the
lower half bearing. By visualizing the effect of the
change in pressure films in the event the journal
tended to become unstable and ride up the side of
the bearing, it can be seen that the high film pressure area above each lobe would also be higher due
to the increase in the oil film convergence at the
higher point, yet the effect of the groove would be
to decrease the film pressure under the shaft, thus
producing a net downward stabilizing force.

As would be expected, the grooves in the lower half
provide relatively small percentages of the bearing
flow due to the effects of the film pressure in this
area.

Maximum film temperatures may run between
160° F and 240 F dependent upon the speed and load, but the higher value would not be considered
good design practice.

Oil drain temperature is not a good indication of the
bearing film temperature although a change in oil
temperatures at the same load conditions would, of
course, indicate a change in the bearing condition.

Elliptical bore axial groove bearings are applied, 
usually incorporating 6 grooves and a single circumferential groove at the center of the bearing acting as an oil supply in each axial groove. The flow
will be larger and drain temperatures will be reduced.

Tilting Pad Bearings

The tilted pad bearing is a relatively late design
bearing principally applied where shaft stability is
a problem. Its application has been largely confined
to larger multi-span units where the problems of
changing alignment across relatively short spans
result in a bearing location unusually susceptible to
variable and light loads.

Since this problem generally does not appear in industrial applications and due to the relatively high
expense of this bearing, it is seldom encountered in
this application.

Fig. 6 Operating Characteristics of the
        Axial Groove Bearing

The bearing consists of a series of pads or segments
around the journal which are babbitted on their inner surface and which are free to tilt or pivot to adjust themselves to the shaft in such a way that a
positive film pressure pattern builds up at each pad.

Tilting pad bearings have high load capacity and
run with an oil film temperature generally higher
than in conventional bearings. Power losses and oil
drain temperature are characteristically higher than
in conventional bearings.

The tilting pad bearing will result in a bearing stability at bearing loads of 20 psi or considerably less.
Other types of bearings require loadings of 75 psi or
more to perform satisfactorily.

An interesting characteristic of tilting pad bearings
is that in some cases the hydrodynamic effects may
cause the journal center to be located off the vertical
centerline of the bearing in a direction counter to
rotation, considering the bottom half bearing, a
condition which normally will not occur in conventional bearings.

A Summary of Bearing Characteristics

In each heading bearings are listed in order of superiority.

INSPECTION AND MAINTENANCE
OF BEARINGS AND JOURNALS

Bearing journals should be round, of uniform diameter from end-to-end, and of good surface finish.

Roundness can easily be determined with micrometers although more careful checking is necessary
than is usually done. Egg-shaped journals will produce a twice-per-revolution vibration which can
cause high vibration and be of essentially destructive nature to bearings. Journals have been found
with other configurations such as a 3-lobe shape
producing a 3-times per revolution vibration. A
run-out check carefully taken on a bearing journal
while rotating in its own bearings will establish its
roundness.

A bearing journal taper must be found with micrometers.

Journal surface condition should, of course, be kept
at the best possible finish. Light circumferential
scratches do no great harm to bearing operation
when less than 0.010-inch wide, providing the
edges are thoroughly smoothed. Axial scratches are
very detrimental.

Hand stoning of journals should be confined to a
circumferential direction of stoning and should be
limited to just enough to remove projecting nicks, raised edges of scratches, etc.  Only oil stones of
fine finish and in good condition should be used.
Excessive stoning of bearing journals does no good
and may seriously affect vibration characteristics
of the unit. Emery cloth and files should not be
used. Hand dressing of journals should be confined
to thorough checks for and removal of local high
spots or upsets above the journal surface since low
spots can obviously not be removed.

Damaged journals can sometimes be resurfaced using a high carbon spray and subsequently grinding
to the desired diameter and finish. However, journals operating at surface speeds above 6000 ft/min.
should not be spray metalized generally speaking.
These journals are usually turned undersized and
used with bearings of suitable undersized bore.

The condition of bearing journals quickly gives a
good clue to cleanliness of oil systems, and perhaps
to the nature of past maintenance practices. Bearing
journals should be carefully protected at all times
during maintenance activities against dirt, mechanical damage, and rusting.

Loaded areas of journal bearings should be examined judging by contact indications visible in the
bearing. Visual checks for cracked babbitt or loosening of babbitt in the shell, small imbedded metal
particles, and wiping should be made.

Bearing surfaces can be carefully smoothed using a
bearing scraper in good condition after removing
imbedded foreign material. It is common to scrape
high spots indicated by local contact markings.

Never use emery cloth or similar material to
smooth babbitted surfaces since abrasive particles
should be kept away from bearings.

Condition of bearing joints is of primary importance to proper operation and clearance. Never use
files to smooth bearing joints and emery cloth or
equivalent is undesirable. Oil stones in good condition should be used for this purpose, although the
outer edges of bearing and bearing casing joints
may be carefully smoothed or broken with a fine
file to remove upset metal due to bumps. Joint compounds should not be used on bearing joints although light application of slow drying compound
is used on bearing cap joints. Heavy applications
commonly are very undesirable since the compound tends to prevent the proper pinching action
of the bearing cap on the bearing, and squeezes out
into bearing passages, orifices, drains, etc.

The outer diameter of bearings, whether cylindrical
or spherical, should be carefully checked and all
nicks, high spots, or galling areas removed to allow
the proper fit and adjustment to bearing casings in
the case of spherical or ball seat bearings. Corresponding fits in bearing casings should be similarly
checked.

Bearing clearances are commonly determined by
measuring the bearing bore with bearing halves assembled after carefully checking the bearing joints.
Vertical and horizontal readings are necessary and
quadrant readings desirable, both front and back, to
evaluate bearing clearance and distortion. Note Figure 7.

In place bearing clearance determined by lead or
fuse wire is usually satisfactory on smaller bearings. The fuse wire may be placed axially along the
top of the journal, or if side or quadrant clearance is
desired, placed circumferentially over the journal
at each end. The upper half bearing is then carefully
bolted in its assembled condition and removed after
which the thickness of the compressed wire is measured. Care should be taken that fuse wire is not
over 0.010 inch or so greater in diameter than the expected bearing clearance as excessive thickness of the fuse wire will introduce excessive forces
when the bearing is bolted, causing incorrect readings and possibly crushing of babbitt or bearing liner distortion in the case of thin liners. Plastigage
strips are very convenient and satisfactory for
smaller bearings.

It is sometimes helpful to measure bearing shell
thickness at each end as a clue to bearing wear or
alignment, particularly on gear bearings.

Note from Figure 7 that a bearing pinch check represents a method of determining the amount of
tightness or looseness existing in the fit between the
bearing housing and its upper casing. The greatest
of care is required as it is necessary to measure
within 1/2 mil or less between these irregularly
shaped parts. Shims are placed between the bearing
cap joints in such a manner that no distortion is
introduced by tightening joint bolts. Fuse wire, or
perhaps plastigage, is placed axially across the top
of the upper half bearing housing and on the bearing cap joint immediately next to the bearing housing at each side. The bearing cap is then bolted in
place and removed after which the compressed
thickness of the fuse wire across the top to the bearing housing is compared to the shim thickness used
and the strips placed at the horizontal joint.

    Fig. 7 Bearing Diameter and Pitch

THRUST BEARINGS AND THEIR
OPERATING CHARACTERISTICS
 (See Figures 8-11)

The turbine thrust runner is usually shrunk and
keyed to the turbine shaft and further secured by
locknuts or other devices. It may be machined integral with the shaft.

In any case the thrust runner is always positioned
precisely in axial location since it is the reference
point for practically every axial dimension and
clearance associated with the rotor assembly.

The axis of the thrust runner must be truly perpendicular to the shaft axis with a maximum axial total
indicator run-out of 0.0005 inch or less. The thrust
runner surfaces must be flat and of polished finish
comparable to shaft journal finish. All the precautions for protecting and maintaining good surface
condition and finish previously discussed for shaft
journals apply for the faces of the thrust runner or
runners.

Since most turbines will normally exhibit a positive
thrust force in a downstream direction, the thrust
runner taking this load is called the active or loaded
thrust runner.

Many turbines will exhibit a reverse thrust load for
some conditions of operation, usually associated
with light load operation, which will result in the
rotor operating against the inactive thrust face.

The total available clearance for the rotor to move
axially between thrust faces with the thrust bearing
installed is necessarily small because of the very
small axial clearances between rotating and stationary parts throughout the turbine steam path.
This clearance is limited to only that sufficient to allow formation of thrust bearing oil films with possible very slight increase for cooling oil flow.
Thrust bearing clearances will normally fall between 0.005 and 0.015 inches, depending upon the
application. Thrust runners must be of sufficient thickness and
strength to prevent deflection under load. Similarly, the thrust housing into which the thrust plates fit
must be free from deflection under load and machined such that thrust plates do not deflect.

Note in Figure 8 that both types of thrust bearing
construction shown involve fits or bolted connections to the shaft journal bearing assembly. Consequently, any looseness or abnormal axial
movement of the journal bearing housing will allow undesired movements of the thrust bearing assembly and the turbine rotor. Thus, the journal
bearing "pinch" action is an unusually important
consideration on journal bearings associated with
the thrust assembly.

Thrust runners of some types may be subject to
loosening in operation, particularly if not properly installed, and the thrust runner should be inspected
with this in mind.

                
           Fig. 8 Thrust Bearing Construction

Thrust bearings operate with hydrodynamic oil
films separating the babbitted thrust plate and the
steel thrust runner. Note carefully that the relationship between thrust bearing and runner supplies no
inherent converging characteristic to oil flowing
between the moving and stationary surfaces as is
the case in shaft journal bearings. To effect a strong
or "stiff' oil film the thrust plate surfaces must be
suitably designed to cause the converging oil characteristic which is necessary.

Babbitted thrust plates are subject to the same type
of inspection applied to journal bearings. Contact
pattern, imbedded foreign material; wiped, cracked
or loosened babbitt are all items of concern. Contrary to journal bearing practice, thrust bearings
cannot be hand scraped or finished any appreciable
amount without damaging the bearing since the
babbitted surfaces many times are deliberately tapered in two directions to facilitate formation of the
oil film.

The presence of nicks, burrs, scratches, etc., on the
steel backing of thrust plates or their fit surfaces in
the thrust housings, cannot be tolerated without seriously affecting thrust bearing alignment, oil film
temperatures, and load capacity.

Most thrust bearings are designed for a given direction of rotation and cannot be interchanged between active and inactive ends.

It can be seen that the alignment of the thrust bearing housing is a very important consideration and, 
since it is usually intimately associated with an adjacent journal bearing, the proper installation and
method of bolting the journal bearing should be
realized.

Thrust forces accumulating on turbine rotors are
the result of steam forces on turbine buckets or
wheels and of net forces of the steam acting at shaft
shoulders. Abnormal conditions, such as blocking
of the steam path due to bucket damage or deposits
must also be considered when designing thrust bearings. In applications involving a flexible coupling to the driven equipment, a coupling thrust
may be experienced due to friction in the coupling
teeth.

The thrust capacity will be determined by the net
area in the thrust bearing face which, in turn, will
require various amounts of physical space available to build the necessary size of parts. The inner
diameter of the thrust bearing must, of course, be
larger than the shaft diameter so that the outer diameter must be increased to allow greater capacity.
This, in turn, will generally cause greater concern
for deflections which might be experienced under
load. The net thrust area is calculated as being the
babbitted area between the inside and outside diameters less the area used for grooving or between
segments or pads which might be incorporated in
the design.

The number of pads or segments used in a thrust
bearing will be selected to limit the lengths of oil
film paths and thereby limit the maximum temperature of the oil film. Thus, a thrust bearing is actually
designed as a series of smaller thrust lands or pads around the shaft, each designed within limitations
of oil film temperature and thickness. The circumferential length of a thrust bearing land is never
made greater than twice the radial length. This results in thrust bearing designs incorporating anywhere from 6 to 12 lands or pads.

            Fig. 9 Double-Runner Thrust Bearing

Tapered Land Thrust Bearings

Tapered land thrust bearings are similar in appearance to flat land thrust plates since their surface is
divided into a number of pads or segments by radial
oil grooves.

Each land is tapered in a circumferential direction
so that the rotation of the thrust runner will wipe oil
into the contracting wedge shape to build up load
carrying oil films.

The total amount of taper, as well as the taper angle, 
is usually made larger at the inside diameter due to
differences in surface speeds of the thrust runner to
result in equalizing oil flows and temperatures
across the land.

The tapered section of each land usually extends
over 80 to 90 percent of the length of the land, the remainder being flat. The flat section is to provide a
load carrying surface at very low speeds before
strong oil films can be developed.

   Fig. 10 Single Runner Thrust Bearing

Tapered land bearings are made with an even number of lands to facilitate machining of plates. The
horizontal split can then be made through the oil
grooves.

Radial oil grooves are dammed at the outer circumference of the pads to control the amount of oil
flowing through the bearing and its temperature to
some extent. Oil is introduced at the inner diameter.

In general, smaller tapers result in higher load carrying capacity and greater oil film thickness. However, increasing the taper slightly will result in
more oil being pumped through the bearing with
some reduction in temperature.

The overall taper in tapered land thrust bearings
will usually fall from 0.0025 to 0.007 inch, depending upon the application. Maintenance personnel
sometimes will readjust shims behind tapered land
trust plates to reduce clearances in the thrust bearing due to excessive clearances. If wear on the
thrust bearing has reduced the land taper appreciably, the bearing should be renewed as it can be
seen that it will begin to approach the far smaller
load characteristics of the flat land bearing.

Due to the compound taper, it is almost impossible
to hand scrape the bearing satisfactorily or re-pour
and refinish the bearing without special fixtures.

Tapered land thrust bearings are normally applied
for expected loadings of 200 to 500 psi. They can be
applied for 450 to 700 psi with some margin in the
smaller sizes where deflections under high loads
are of less concern. Tapered land thrust bearings in
good alignment will withstand loads of 1000 psi or
greater without failure.

The main disadvantage of the tapered land thrust
bearing is its relative sensitiveness to misalignment
which will seriously reduce its capability.

As a matter of interest, it has recently been found
that copper backed thrust plates (in place of steel) result in very substantial increased capacities in the
bearing. This is due to the higher heat conductivity
characteristics of the copper allowing the film temperatures to run cooler and more uniformly. If local
hotspots exist in the bearing surface, it can be seen
that temperature distortion would cause the bearing
capacity to be seriously impaired.

Flat Land Thrust Bearings

This bearing, as the name implies, incorporates a
flat face facilitating its manufacture and reducing
its cost. Radial grooves are provided in the bearing
face forming pads or segments and supplying oil
for lubrication and cooling.

The flat land thrust has relatively low load capacity
and is applied where thrust loads are small and consistent with the bearing capability. It, therefore, 
functions primarily as a locating device with limited load capacity.

The bearing surface does not result in formation of
a definite converging characteristic to the flow of
oil necessary to form a stiff oil film. An oil film of
sorts is established by radii or contours at the edges
of oil grooving primarily.

Flat land thrust plates are normally supplied for expected loadings of 75 to 100 psi although some
margin remains at 150 psi loading.

Tilting Pad Thrust Bearings

Tilting pad thrust bearings, of which the Kingsbury
Bearing seems to be the most widely applied, differ
from the flat or tapered land bearing due to the fact
that each pad is an independent segment free to tilt
about a pivot. The pivot point is usually a hardened
spherical surface behind each pad allowing tilting
in both radial and circumferential directions or a
combination of the two.

The spherical pivot of the pads bear against a series
of leveling plates which act to distribute the thrust
load uniformly around the bearing casing. Point’s of
contact are hardened in the supporting plate arrangement.  A misaligned will cause the tilting of the leveling plates in the high load area to force the
remaining plates to move in the direction of the
thrust collar on the low load area, thus, distributing
the load.

The greatest advantage of tilting pad thrust bearing
is this ability to adjust to a misaligned condition.

Oil flow is introduced near the shaft and flows radially outward. Due to large spacing between the
pads these bearings require larger oil flows than the
flat or tapered land type.

Due to the relatively numerous parts required and
small contact areas in the leveling rings and their
supporting base ring, the bearing is subject to compressive yielding or deflection at the load contact
points under high loads, accounting for the apparent increase in thrust clearance frequently found.
Under high loads, the pads are subject to deflection
or "crowning" with increases in film temperature
resulting in reduced capacity.

Tilting pad bearings are normally designed to expected loads of 200 to 400 psi. The ultimate capacity is 600 to 900 psi.

       Fig. 11 Tilting Pad Thrust Bearing (Partial View)

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