Liquid lubrication uses a fluid, normally an oil, to reduce friction and wear between moving surfaces. The lubricant can also remove heat, protect surfaces against corrosion and contamination, and help carry or distribute loads within a bearing.
Liquid-film lubrication can operate under hydrodynamic or hydrostatic conditions. In a hydrodynamic bearing, relative movement and bearing geometry generate pressure within the lubricant film. In a hydrostatic bearing, pressurised lubricant is supplied from an external source.
Under full-film hydrodynamic lubrication, the moving surfaces are separated by a continuous lubricant film. During starting, stopping, low-speed operation or overloading, sufficient film thickness may not be maintained and the bearing can operate in mixed or boundary lubrication, where some surface contact may occur.
Viscosity is one of the most important lubricant properties because it influences film formation, frictional losses and bearing performance. See the Viscosity page for further engineering data.
For students and engineers, the key principle is that a successful fluid-film bearing attempts to keep the opposing solid surfaces separated by lubricant. Bearing performance therefore depends on establishing and maintaining sufficient film thickness under the expected speed, load, temperature and lubricant viscosity.
For students and engineers, the key principle is that a successful fluid-film bearing attempts to keep the opposing solid surfaces separated by lubricant. Bearing performance therefore depends on establishing and maintaining sufficient film thickness under the expected speed, load, temperature and lubricant viscosity.
The design and operation of a liquid lubrication system should consider:
It is not possible to address these factors on this page.. Please refer to the links provided below
Petroff's equation provides an friction value for an unloaded journal bearing i.e with the shaft concentric with the journal. This assumes no end load and does not allow for end leakage..
f = 2 . p 2 . (m . n / p ) . r / c
Torque to rotate shaft .. T (Nm)
= W. f .r
= W.[ 2 . p 2 . (m . n / p ) . r / c ] . r
= W.[ 2 . p 2 . (m . n . L. 2. r / W ) . r / c ] .r =
4.p 2.r3.m . n .L / c
Power to rotate shaft... = T.ω = T.2.p .n
> P (Watts) =
8 . p 3 .n 2 . r 3 . L . m / c
= p 3 .n 2 . d 3 . L . m / c
The equivalent power for a rotating thrust bearing ...
P (Watts) = 2 . p 3 .n 2 . m . ( r2 4 — r1 4 ) / t
Journal and thrust and linear bearings can be hydrostatic in that the load carrying capacity results
from externally generated lubricant pressure. These type of bearing do not depend on
the relative motion of the bearing surfaces for lubrication and so they are effective at zero and
very low velocities. This type of lubrication does not involve the necessity of metal
contact during start-up and shut down.
Hydrostatic lubrication systems are generally expensive to engineer and are liable to problems
with controlling the lubrication supply
Note; Hydrodynamic Lubrication is extremely complex.
These notes provide only outline explanations of the principles involved...
In hydrodynamic lubrication the fluid is assumed not to slip at the interface with the
bearing surfaces i.e. the fluid in contact with the bearing surfaces moves at the same velocity as the surface.
Over the thickness of the fluid there is a velocity gradient depending on the relative
movement of the bearing surfaces. If the bearing surfaces are parallel (or concentric ) the action
motion of the lubricant will not result in a pressure which could support any bearing load. However
if the surfaces are at a slight angle the resulting lubrication fluid velocity gradients will be such that
a pressure results from the wedging action of the bearing surfaces... Hydrodynamic lubrication depends
upon this effect... Note; This principle is similar to the lift in water skiing / aqua planing ..
The operation of hydrodynamic lubrication in journal bearings is illustrated below. Before the rotation commences
the shaft rests on the bearing surface. When the rotation commences the shaft moves up the bore until an equilibrium condition
is reached when the shaft is supported on a wedge of lubricant. The moving surfaces are then held apart by the pressure generated
within the fluid film. Journal bearings are designed such that at normal operating conditions the continuously generated fluid pressure supports
the load with no contact between the bearing surfaces. This operating condition is known as thick film lubrication and results in a very
low operating friction and extremely low bearing load
Boundary lubricating conditions occur when the lubricant film is insufficient to prevent surface contact. This occurs at rotation start-up, a slow speed
operation or if the load is too heavy. This regime results in bearing wear and a relatively high friction value. If a bearing is operated under
boundary lubricating conditions special lubricants are needed.
The operation of hydrodynamic lubrication for thrust bearings is enabled by various design options including tilting
pads, taper lands and step bearings. The tilting pads provide the most ideal Hydrodynamic lubrication conditions
as shown on the figure below..
It is generally desirable to achieve hydrodynamic lubrication in bearings or the following reasons;
Hydrodynamic lubrication depends on at least three of dimensionless numbers ...
( m .n / p ) , ( D / h ) , and (L/D)
The relationship between the bearing friction coefficient and the bearing modulus
is shown in the figure below
| Equipment | Bearing | Max Pressure | m | m n /p |
| MPa | Pa. s | |||
| Automobile /Aircraft Engines | Main | 5 - 12 | 0,007 | 3,67 x 10-8 |
| Crankpin | 10-23 | 0,008 | 2,50 x 10-8 | |
| Wrist pin | 14 -35 | 0,008 | ||
| Gas and Oil Engines | Main | 3.5 - 8 | 0,02 | 5,00 x 10-8 |
| Crankpin | 7 -12 | 0,04 | 2,50 x 10-8 | |
| Wrist pin | 8-14 | 0,065 | ||
| Marine Engines | Main | 3.5 | 0,03 | 5,00 x 10-8 |
| Crankpin | 4 | 0,04 | 3,67 x 10-8 | |
| Wristpin | 10 | 0,05 | ||
| Stationary steam engines | Main | 1,5 - 3 | 0,015 -0,06 | 5,00 x 10-8 |
| Crankpin | 4 -10 | 0,03 -0,08 | 1,50 x 10-8 | |
| Wristpin | 12 | 0,025 -0,06 | ||
| Reciprocating pumps and compressors | Main | 2 | 0,03 | 7,33 x 10-8 |
| Crankpin | 4 | 0,05 | 3,67 x 10-8 | |
| Wristpin | 7 | 0,08 | 3,67 x 10-8 | |
| Steam Turbines | Main | 0,5 - 2 | 0,002-0,016 | 25,0 x 10-8 |
| Rotary Pumps and Motors | Shaft | 0,5 -1.5 | 0,025 | 50,0 x 10-8 |
A design constraint to keep thick film (full hydrodynamic)
is to ensure the bearing modulus (m n /p ) >= 1.09 x 10-9
McKee established the following relationship using small bearings . This does not allow for end leakage
f = 19,56 . ( m .n / p ). ( D /C) + k
k is obtained from the diagram below but can be approximated as 0.002 over an L/D ratio 0 0.75 - 2.6
The Sommerfeld Number is a dimensionless parameter used in lubrication analysis.
S = ( m .n / p ). ( D /C) 2
This parameter has been used as the abscissa for a number of design curves. The ordinate can
be selected to allow the friction value, film thickness, oil leakage, temperature rise etc
to be determined. Design curves have been produced of various variables against the Sommerfeld using computer
techniques by A.A Raimondi and J.Boyd of Westinghouse Research Labs(ASLE Transactions Vol 1 No 1 April 1958).
These graphs include compensation for end leakage and eccentricity.
The illustrative design curve has been included below. Detailed
journal and thrust bearing designs should be completed using the relevant specialist
sources of information or software..
In comparing the value of f(D/C) resulting from the Petroff, McKee and the Raimondi-Boyd data the
values are in close agreement for the higher values of the Sommerfeld number (above 0.5) i.e the
lightly loaded bearings..
| Speed Range | Shaft Diameter | Diametrical Clearance |
| mm | mm | |
| Below 600 RPM | 25 | 0.025 - 0.05 |
| Above 600 RPM | 25 | 0.03 - 0.10 |
| Below 600 RPM | 40 | 0.03 - 0.08 |
| Above 600 RPM | 40 | 0.05 - 0.12 |
| Below 600 RPM | 50 | 0.04 - 0.09 |
| Above 600 RPM | 50 | 0.06 - 0.14 |
| Below 600 RPM | 80 | 0.05 - 0.11 |
| Above 600 RPM | 80 | 0.08 - 0.17 |
| Below 600 RPM | 100 | 0.06 - 0.13 |
| Above 600 RPM | 100 | 0..09 - 0.20 |
| Below 600 RPM | 125 | 0.07 - 0.14 |
| Above 600 RPM | 125 | 0.10 - 0.22 |
| Below 600 RPM | 150 | 0.08 - 0.15 |
| Above 600 RPM | 150 | 0.12 - 0.24 |
| Below 600 RPM | 200 | 0.09 - 0.17 |
| Above 600 RPM | 200 | 0.14 - 0.27 |
Liquid lubrication uses a fluid, normally oil, to reduce friction and wear between moving surfaces. The lubricant may also remove heat, protect surfaces and support loads within fluid-film bearings.
Hydrodynamic lubrication occurs when relative motion and bearing geometry generate pressure within a lubricant film sufficient to separate the moving surfaces and support the applied load.
Hydrostatic lubrication uses lubricant supplied under external pressure to separate and support the bearing surfaces. It can therefore provide a supporting fluid film even when there is little or no relative motion.
Viscosity affects the ability of the lubricant to form and maintain a load-carrying film. It also influences frictional losses, heat generation and the operating characteristics of the bearing.
The Sommerfeld number is a dimensionless parameter used in hydrodynamic journal bearing analysis. It combines operating and geometric parameters and is commonly used with bearing design curves.
Boundary lubrication occurs when the lubricant film is insufficient to completely separate the moving surfaces. Some surface contact can then occur, increasing friction and wear compared with full-film lubrication.