Handbook of Lubrication Episode 2 Part 3 ppt

Dual basket filter — Woven gauze or perforated metal element with changeover valve;
elements easily removed for hand cleaning when system in operation. Magnets can be
incorporated. Normally 50 µm and above.
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The table shows that only large particles will settle within practical time limits. Water settling
is hastened by raising the bulk temperature to 70°C after which it can be drained from the
bottom of the reservoir.
Filtration
This is the most universal method and many filter materials and designs are available.
Filters should be selected so that under clean conditions and maximum working viscosity
the pressure loss does not exceed 0.3 bar (5 psi). Cleaning is normally recommended when
pressure loss increases by 1 bar. Filters usually fall within the following types.
Line strainer — Woven gauze or perforated metal element; easily removed for hand
cleaning when system stopped. Normally 150 µm and above.
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Mechanically cleaning filter — Interleaved radial plates plough the dirt from the gaps
between metal discs when the filter pack is rotated, which can be while the system is
operating. May be motorized. Periodically drain contaminant from sump when system is
stopped. Normally 150 µm and above.
Various other designs of mechanically cleaning filter, such as wire wound and back
flushing, are available.
402 CRC Handbook of Lubrication
Disposable element filter — Element of materials such as treated paper, felt, and nylon
easily replaced when system stopped. Dual versions with changeover valve permit element
replacement when system operating. Normally below 50 µm.
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Centrifuging
Installed in a by-pass circuit. Removal of sediment and water can be carried out while

the oil system is in operation. For maximum efficiency, the oil should be centrifuged at
70°C with provision of an inline oil heater.
COOLERS
Constant oil temperature desirably enables constant flow and pressure control as these are
both affected by changes in viscosity. Most machine designers recommend that the working
temperature of the oil be 40°C. High-ambient temperatures, heat generation from bearings
and gears, and machine and oil pump inlet power all transfer heat into the oil. The amount
of heat is normally specified by the machine designer or based on experience with similar
units. This commonly amounts to an oil temperature increase through the machine in the
region of 10°C which needs to be removed by a cooler. Prolonged working temperatures
above 60°C shorten oil life.
Water and air are the common cooling mediums. When considering water, its cleanliness,
corrosion characteristics, hardness, and pressure will affect selection of cooler materials.
Temperature, quality, and quantity of cooling medium available are important in obtaining
the most efficient cooler. Pressure losses for oil or water through the cooler should not
exceed 0.7 bar.
Cooler types frequently used are as follows:
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Shell and tube — Oil through shell, water through tubes. Requires space for tube removal.
Plate — Oil and water between alternate plates. Compact and readily separated for
cleaning.
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Radiator— Oil through tubes, air over tubes motivated by fan.
TEMPERATURE CONTROL
Along with heaters and coolers, controls are required to cope with inevitable temperature
fluctuations.
ReservoirHeater
The control thermostat or thermostatic valve probe, Figure 3, should be positioned near
the pump suction and at about middepth of the oil in the reservoir to sense the average

temperature. Ensure that the thermostat is never exposed or placed near any localized hot
spot such as a heater.
Oil Cooler
One control method is to regulate the water/air flow, the other to regulate the flow of oil
to be cooled. Water flow can be governed by a hand control valve as its temperature usually
only fluctuates on a seasonal basis. Automatic control of air is essential as its temperature
fluctuates daily.
Automatic control utilizes a direct-acting modulating valve in the cooling water supply
line which is controlled by a sensing element in the cooling oil outlet. The effect on oil
temperature is not instantaneous but is generally acceptable for industrial systems. Cooling
water pressure should be reasonably stable, otherwise a pressure regulating valve will be
required.
Alternatively, where instantaneous response to control is vital and/or where air is the
cooling medium, the cooler should be provided with a bypass line and a control valve to
divert flow into the bypass. The valve is of a three-way type with overlapping ports, Figure
4. Mixing of the oil streams within the valve produces an average temperature and a
thermostatic element detects any deviation in temperature and corrects the valve position.
PRESSURE CONTROL
A pressure control valve is necessary to spill-off surplus oil from the pump, to regulate
any flow variation due to temperature fluctuations, and to accommodate any changes in
demand from the machine being lubricated. The following are typical methods of control.
Spring-loaded relief valve — Provides coarse control and is sensitive to viscosity changes.
Generally used on smaller, simple systems.
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Direct-operated diaphragm valve — The diaphragm chamber is connected so that system
pressure is transmitted to the diaphragm. The spring counter-balancing the diaphragm load
is adjustable and determines the system pressure. Any change in demand will tend to vary
the system pressure and the diaphragm, sensing this, will reposition the valve to adjust the

spill-off rate. This valve will maintain the pressure within acceptable limits provided the
viscosity remains reasonably stable.
Pneumatically controlled diaphragm valve — The diaphragm is air actuated via a control
instrument. This valve is normally selected when very accurate control is necessary or if
the system operating pressure is too great for the direct-acting valve diaphragm.
Header tank — While space requirements often preclude this simple form of control,
the procedure is to place a tank at the required height. Filled directly with a line teed off
the main pump supply, the tank is fitted with an overflow connection. This method ensures
the continual change of tank contents to maintain the oil at system temperature. As an added
advantage, if the system pumps fail the tank will discharge its contents via the fill connection
to the equipment being lubricated. Pump check valves will prevent oil returning directly to
the main reservoir.
Providing different pressures within the same system involves the use of pressure-reducing
valves. These normally comprise a restricting orifice, or valve opening, which is controlled
by imposing the outlet pressure on the valve control diaphragm.
EMERGENCY EQUIPMENT AND SYSTEM CONTROL
If a machine must continue to run after a failure in the lubrication system main pump, it
is imperative to arrange fully automatic starting of a second, and maybe a third pump. This
can be done by use of flow or pressure-operated switches, or both. Control panel lights
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dished end is included in the tank sizing calculations. As some air will be absorbed in the
oil, it will be necessary from time to time to add air. Acheck valve should be fitted in the
air supply line to prevent oil from entering the air main and an air regulator installed to
avoid accidental overpressurizing.
SYSTEM PIPING
Sizes of interconnecting pipes should be considered in relation to oil viscosity, velocity,
and resultant friction losses. Pipes should be large enough to prevent cavitation in pump
suction lines, to avoid undue pressure drop in pump supply lines (minimizing pump drive

power), and to avoid backup in drain lines.
Suction
Pipe runs should be short. Right angle bends and tee pieces should be kept to a minimum.
Nominal bore of pipe to be one size larger than supply.
Supply
Friction loss due to viscosity frequently outweighs velocity considerations, particularly
with heavier oils. Figure 6 is based on a friction loss of 0.1-m head per m of pipe and
restricted to velocities under an acceptable 2 m/ sec. The viscosity at specified operating
temperature should be used to determine pipe nominal bore.
To determine the friction loss in pipework, multiply the length by 0.1 m head. Friction
loss caused by fittings and valves can be determined by converting them to equivalent pipe
lengths in Table 2. These, together with pressure losses through the filter and cooler, will
determine the system losses.
Drain
Drain pipes should be sized to run not more than half full so as to encourage escape of
entrained air, provide space for any foam and give a margin of safety. Drain pipes should
be vented. Flow rate, slope, and viscosity govern their size. Aminimum slope of 1 in 40
is essential but use should be made of all available drop.
The pipe nominal bore may be determined from Figure 7. At startup, pipework and any
408CRC Handbook of Lubrication
FIGURE 6. Supply line sizing.
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Design Principles
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Copyright © 1983 CRC Press LLC
JOURNALAND THRUSTBEARINGS
A.A. Raimondi and A.Z. Szeri
INTRODUCTION
This chapter applies hydrodynamic lubrication theory to the analysis and design of self-

acting fluid film journal and thrust bearings, in contrast to earlier chapters which emphasize
lubrication theory and solution techniques. Most of the material has been summarized in
design charts and tables for estimating performance of a variety of applications.
It is not within the scope of this chapter to recommend bearing proportions, allowable
temperature rise, etc. These are left to the designer to decide on the basis of experience and
test. The charts provided here will serve for performance calculations on many representative
bearings; similar information is available in the literature for a variety of designs. Computer
programs are also available for studying design parameters for specific applications.
LUBRICANTPROPERTIES IN BEARING DESIGN
The lubricant property of greatest concern in fluid film bearings is the absolute viscosity,
or just viscosity, μ. Its SI unit is Pa · sec (Pascal second), and in English units it is usually
expressed in lb
F
· sec/in.
2
(reyn). The ratio of absolute viscosity to density (ρ) is termed
the kinematic viscosity,
␯
=μ/ρ.It is measured in m
2
/sec in SI units and commonly in
in.
2
/sec in English units. Table 1 contains conversion factors for commonly used viscosity
units.
Increasing temperature lowers the viscosity of lubricating oils as shown in Figure 1 for
typical industrial petroleum lubricants in the various ISO viscosity grades. The viscosity of
a number of other fluids is given in Figure 2.
Average Viscosity
In numerous applications, the temperature rise in the bearing film remains relatively small.

However, in estimating bearing performance on the basis of classical (isothermal) theory,
the calculations should employ an effective viscosity compatible with the mean bearing
temperature rise.
1.19
This calculation might be based on the assumptions that:
1. All heat, H, generated in the film by viscous action is carried out by the lubricant.
2. The lubricant which leaves the bearing by its sides has a uniform average temperature
T
s
= (T
i
+ ΔT/2), where ΔT = T
o
– T
i
is the mean temperature rise across the
bearing.
This mean temperature rise, ΔT, can be calculated from a simple energy balance which
gives:
ΔΤ ϭ H/[
ρ
c(Q Ϫ Q
s
/2)] (1)
For typical petroleum oils, ρc = 112 lb
F
/in.
2
F (139 N/cm
2

C) and for water ρc = 327
lb
F
/in.
2
F (406 N/cm
2
C).
Rather than assuming a uniform effective viscosity at a mean bearing temperature, using
the actual variation in viscosity resulting from temperature changes as the lubricant flows
through the bearing will result in considerably improved accuracy in calculating performance.
However, this increases the complexity of the solution, usually requires special computer
Volume II 413
Copyright © 1983 CRC Press LLC
Flow Transition
Two basic modes of flow occur in nature: laminar and turbulent. Flow transition from
laminar to turbulent in bearings is preceded by flow instability in one of two basic forms:
(1) centifugal instability in flows with curved streamlines, or (2) parallel flow instability
characterized by propagating waves in the boundary layer.
Instability between concentric cylinders was studied by Taylor.
3
He found that when the
Taylor number (Ta = Re
2
C/R = Rω
2
C
3
/
␯

) reaches its critical value of 1707.8, laminar
flow becomes unstable. The equivalent critical reduced Reynolds number is √
—
C/R Re =
41.3. The instability manifests itself in cellular, toroidal vortices that are equally spaced
along the axis.
As the Taylor number is increased above its critical value, the axisymmetric Taylor vortices
become unstable to produce nonaxisymmetric disturbances, and turbulence eventually makes
its appearance.
4
If the Reynolds number reaches 2000 before the Taylor number achieves
its critical value, turbulence is introduced rapidly
5
without appearance of a secondary laminar
flow.
Eccentricity plays a role in defining critical conditions as covered in the chapter on
Hydrodynamic Lubrication (Volume II). A positive radial temperature gradient in the clear-
ance space, such as found in journal bearings at the position of minimum film thickness, is
also destabilizing,
6
as is heat generation by viscous dissipation.
7
These statements draw
support from experimental journal bearing data.
8
The local critical Reynolds number R
h
=
Rωh/
␯

seems to be in the 400 to 900 range.
2
Accepting R
h
= 900 for onset of turbulence,
the critical value of global Reynolds number is approximately Re = 900 ␯
-
/(1 − ⑀
). Here,
␯
-
is the ratio of the lowest value of the kinematic viscosity in the film to its value at the
leading edge. Thus, for a fourfold decrease in viscosity and an 0.8 eccentricity ratio, the
critical global Reynolds number is Re = 1125. A global value of 1000 has been used in
later examples as a criterion for onset of turbulence.
In thrust bearings, it was found
9
turbulent transition takes place within the range 580 <
Re < 800, where Re = U
a
h
a
/
␯
is calculated on average conditions. Reference 10 reports
agreement, but after replacing the average film thickness with the minimum film thickness
in Re.
Turbulence
Turbulence is an irregular fluid motion in which properties such as velocity and pressure
show random variation with time and with position. Once a relationship is established between

the mean flow and Reynolds stresses, averaged equations of motion and continuity can again
be combined to yield an equation in the (stochastic) average pressure p
-
:
(3)
Calculations in later sections of this chapter make use of a linearized theory
11
for turbulence
functions k
x
and k
z
. The main contribution of isothermal turbulence to bearing performance
is a significant increase in both load-carrying capacity and power loss.
Thermal Effects
If the bearing is large or if loading conditions are severe, pointwise variation of viscosity
in the lubricant film is significant. Assuming negligible temperature variation in the axial
direction, thermohydrodynamic (THD) journal bearing lubrication is represented by the
following equations of pressure and temperature:
12
(4)
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Copyright © 1983 CRC Press LLC
(5)
Here, ␾
−
the mean dissipation and the turbulent functions k
x
, k
2

, and F, as well as the velocity
temperature correlation V’t’
—
,depend on the turbulence model used. Aminimum list of
nondimensional parameters that characterize bearing performance according to THD theory
must include
(6)
Consideration here is limited to varying S, Re, and Pe while keeping all other parameters
constant at the values given in Table 2 in an example which covers transition from laminar
to turbulent operation.
12,13
Asignificant result of THD theory is the strong effect of turbulence on bearing temperature
illustrated in Figure 3. Transition to turbulence (occurring at D ~ 25 cm) is beneficial for
limiting bearing temperatures, especially at low loads. Coefficient of friction is strongly
dependent on bearing specific load in the laminar regime (Figure 4), but this dependence
lessens as diameter is increased.
Inertia Effects
Lubricant inertia can have a significant effect on bearing performance if Re (C/R) у1.
2
While the equations of motion are nonlinear when convective inertia is retained, the problem
becomes tractable as pointwise lubricant inertia is replaced by its average value, obtained
via integration across the film.
14,15
The averaged equations of motion and continuity combine
in a single equation in lubricant pressure. For journal bearings:
(7)
418 CRC Handbook of Lubrication
Table 2
PARAMETERS FOR THD
SAMPLE SOLUTION

Copyright © 1983 CRC Press LLC
Equation 7 was made nondimensional through:
The entries of Table 3 were calculated from Equation 7 for a 160° partial arc journal bearing
(L/D = 1.0).
Dynamic Properties of Lubricant Films
Figure 5 represents an idealized configuration where rotor weight (2W) is supported on
two bearings. Under steady load W(Figure 5b), the journal center O
js
is displaced from the
bearing center to the steady operating position shown.
Rotor response to a small excitation, say imbalance, assuming the bearings to be rigid
supports, will be as shown in Figure 6. (The same curve applies with roiling contact bearings.)
Such rotors cannot be operated at the critical speed and can become “hung” on the critical
when attempting to drive through. When hydrodynamic bearings arc used, the lubricant film
adds another spring (in addition to the shaft spring in bending) and, importantly, considerable
damping. Two effects can be noticed in the rotor response curve: (1) critical speed is lowered
below that calculated for rigid supports, and (2) vibration amplitude is reduced.
In this example, the excitation is imbalance and occurs at running speed. In practice,
exciting frequencies can be different from the shaft speed: magnetic pulls, gear impacts,
out-of-round shaft, steam, or aerodynamic forces
16
on turbine or compressor blades, etc.
The latter has been known to cause large self-excited vibrations. In addition, lubricant films
themselves can originate destructive self-excited vibrations. Aclassical case is oil whip at
slightly less than one-half running speed.
17
Another oil film phenomenon is a self-excited
vibration at exactly one-half (or other exact submultiple) of the running speed known as
subharmonic resonance.
18

The rotor-shaft configuration of Figure 5 is reduced to a simple dynamical system of
springs and dashpots in Figure 7. Amass W/g (one half the rotor weight) can be imagined
to be concentrated at O
js
, the steady running position of the journal. If some excitation F
420CRC Handbook of Lubrication
FIGURE 5.Dynamical elements of rotor-shaft configuration.
Copyright © 1983 CRC Press LLC
422 CRC Handbook of Lubrication
Table 3 (continued)
JOURNAL BEARING PERFORMANCE INCLUDING INERTIA AND TURBULENCE (L/D = 1.0,
ββ
= 160°)
Copyright © 1983 CRC Press LLC
424CRC Handbook of Lubrication
(9)
Steady state response (x,y, phase angle) to excitation F can then be obtained by solving
the above linear equations. Equally important, existence of any self-excited vibrations can
be investigated by performing a conventional stability analysis
18
on the homogeneous equa-
tions obtained by setting the right hand side of Equation 9 equal to zero. When the oil film
is only one element in a more complicated linear dynamical chain, it can be incorporated
in the basic equations of motion for the vibrating system, as shown. All eight oil film
coefficients are required in order to make accurate dynamical analyses of rotor-shaft
configurations.
In some instances, lubricant inertia should also be accounted for in linear representation
of the oil film.
15
This effect becomes important when Re (C/R) у1, and oil film force

Equation 8 should be amended as follows to include acceleration (inertia) coefficients:
(10)
Table 3 gives approximate D inertia coefficients for a 160° partial-arc bearing.
THRUSTBEARINGS
Fixed-Type Thrust Bearings
The fixed-pad slider bearing in Figure 8 is the most basic configuration. In practice, this
type is commonly constructed as shown in Figure 9a with a flat area following a machined
taper to support the sliding surface when it is at rest. Basic design charts for these config-
urations can be found in Reference 19.
If the pads are arranged in an annular configuration with radial oil distribution grooves,
a complete thrust bearing (Figure 10) is achieved. Approximate performance calculations
of this bearing can be made by relating the rectangular slider bearing (width B, length L)
FIGURE 8.Fixed-pad slider bearing.
Copyright © 1983 CRC Press LLC
Pivoted-Pad Thrust Bearings
This type (Figure 13) employs a supporting pivot at the center of film pressure ( in
Figure 8) developed on the surface of a slider bearing. While performance is theoretically
identical to a fixed-pad bearing designed with the same slope, the pivoted type has the
advantages of (a) self-aligning capability, (b) automatically adjusting pad inclination to
optimally match the needs of varying speed and load, and (c) operation in either direction
of rotation. Theoretically, the pivoted pad can be optimized for all speeds and loads by
judicious pivot positioning, whereas the fixed-pad bearing can be designed for optimum
performance only for one operating condition. Although pivoted pad bearings involve some-
what greater complexity, standard designs are readily available for medium to large size
machines.
When the thrust bearing is divided into a set of pie-shaped segments, circumferential
length of each sector at its mean radium (R
mean
) is commonly set approximately equal to the
426CRC Handbook of Lubrication

FIGURE 12.Spiral groove bearing.
FIGURE 13.Pivoted-pad thrust bearing.
Copyright © 1983 CRC Press LLC