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ZPlus, LLC

Burlington, NC 27215
ZPlus Tech Brief #11
Internal Combustion Engine Lubrication
Lubrication in an Ideally Designed Engine
It is a matter of fact that using PCEO (Passenger Car Engine Oil) to lubricate all of the different parts of an engine
is a necessary compromise. In an ideally designed engine, each part requiring lubrication would be given the exact
lubricant that best suited the nature of the friction being minimized. The basic principle of lubrication is to keep two
bearing surfaces that move relative to each other separated, to minimize wear. Ideally, this would be accomplished in
all IC (Internal Combustion) engine bearings with an oil flm thick enough to eliminate any contact, but this is not always
possible. In an actual engine, the lubricating oil flm thickness will range from a continuous flm of oil in Hydrodynamic
Lubrication, (as in crankshaft bearings), to a trapped and thinner flm of oil in EHD (Elasto-HydroDynamic) Lubrication
(found in wrist pins and some high-performance engine crankpin bearings), with actual rubbing contact through the oil flm
found in Boundary Lubrication which happens in some engines with fat lifters. Although in reality these three lubrication
modes are part of a continuous range of oil flm thickness, they each place different demands on the oil, so we will
examine them independently.
System Load in psi Mode of Lubrication
Crankshaft Bearing 1700 Hydrodynamic
Crankpin Bearing 3300 Hydrodynamic
Wrist Pin 5000 Sliding Hydrodynamic
Cam/Lifter Contact >150,000 EHD/Boundary
The complex relationships between these factors are fgure 1 - IC Engine Bearing Systems and Loads
illustrated in a function called the Stribeck Curve, shown
in purple on the graph in fgure 2. Named after German scientist Richard Stribeck (1861-1950), this curve plots the
relationship between the three major factors which determine oil flm thickness in bearing systems. Overlaid on this curve
is the relative Coeffcient of Friction curve in blue and relative Wear in green.
The horizontal x-axis in fgure 2 is a value which increases with rotational speed or oil viscosity, and decreases with an
increase in load. As the rotational speed of the shaft increases or the viscosity of the oil is increased, the resulting oil
flm thickness increases along the purple curve to a higher value. If the load increases, the flm thickness decreases
along the purple curve to a lower value. It is important to note the system oil pressure is not included in this chart, as
the system oil pressure only needs to be adequate enough to replace oil that leaks out of bearing clearances. Pressure
above this minimum value does not affect the oil flm thickness in a hydrodynamic bearing, as we explain in the
hydrodynamic lubrication section of this paper. This leakage is affected by oil viscosity, bearing clearance and pressure.
In fgure 2, the oil flm thickness curve x-axis is separated into three regions: Hydrodynamic, Mixed, and Boundary. In
hydrodynamic lubrication, the two bearing surfaces are completely separated by an oil flm. In the mixed region, the oil
flm thickness drops into the range of the roughness of the bearing surfaces. The peaks of the surface which give it its
roughness are called asperities. As you move towards the left in this mixed region, the asperities begin to contact. The
left hand region is boundary lubrication, in which the interaction of the asperities between the two bearing surfaces is
increasingly the cause of friction. You will notice in fgure 2 there are dashed lines superimposed on the left side of
Beardmore, Roy, http://www.roymech.co.uk/Useful_Tables/Tribology/Liquid_Lubrication.htm, 2004
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January 13, 2009
Figure 1 gives average loading values for different engine
bearing systems and the mode of lubrication they employ.
To better understand the different modes of lubrication,
we will examine the relationship between flm thickness,
oil viscosity, load, and speed of movement of the frictional
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January 13, 2009
the graph. They represent the decrease in both coeffcient of friction and wear in the presence of the proper anti-wear
additive such as ZDDP. Without the anti-wear additive, a decrease in either rotational velocity or viscosity, or an increase
in the load ultimately results in destruction of the journal and bearing.
Hydrodynamic Lubrication
Defnition: A system of lubrication in which the shape and relati ve motion of the sliding sur faces causes the
formation of a fuid flm having suffcient pressure to separate the surfaces.
The crankshaft, crankpin, camshaft bearings, and in some engines rocker shaft bearings, utilize a plain bearing insert
and journal which are kept separated by an oil flm operating in a hydrodynamic lubrication mode. Contrary to the popular
misconception, the journals in these bearings are not kept from rubbing on the bearing inserts by the engine oil system
pressure. Using the crankshaft bearings as an example, the oil is introduced into the bearing journal clearance only at
a hole in the top bearing insert, although in many engines there is also a groove to distribute the oil half-way around the
journal. In this bearing, the oil system pressure is actually trying to push the crankshaft away from the oil hole toward
the lower bearing half. If it were not for the presence of hydrodynamic lubrication, the crankshaft would be pushed down
into contact with the lower bearing insert. Since crankshaft bearings are not sealed, oil leaks out the side clearance of
the bearings, and the engine oiling system merely keeps an adequate supply of oil in the bearing. The hydrodynamic
lubrication does the actual work of keeping the bearing and journal separated.
In order to illustrate the principle of hydrodynamic lubrication, consider the bearing and journal system of an IC engine
crankshaft journal and bearing, illustrated in fgures 3, 4, and 5. The size differential between the bearing shell and
journal has been exaggerated to make the principle clear. For the purposes of the example, we are loading the
crankshaft vertically opposed to the oil hole. Our initial conditions are that of an engine which has been recently built,
but has no oil between the journal and bearing surface.
(Our example starts with dry bearings in order to make the point more clearly but in practice, all correctly assembled
engines have the bearings pre-lubricated at assembly, and an engine which has been run will have a thin flm of oil in the
bearings. In both of these conditions, the bearings will operate in a mixed hydrodynamic/boundary lubrication mode until
the oil pressure rises and re-establishes a continuous supply of oil, enabling full hydrodynamic lubrication.)
With the crankshaft stationary, as illustrated in fgure 3, the frst application of pressure on a cylinder power stroke forces
the journal downwards into contact with the bearing. As the crankshaft attempts to spin in the absence of oil, the journal
attempts to climb the left side of the bearing due to the mutual friction between the two, as shown in fgure 4. Of course
Noria Corporation, Dictionary of Tribological Terms, 1328 E. 43rd Court, Tulsa, OK 74105
fgure 2 - The Stribeck Curve (Plus Coeffcient of Friction and Wear Curves)
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January 13, 2009
without oil, the journal and bearing will be in rubbing contact,
quickly causing damage to both.
To the system in fgure 4, we then introduce a supply of oil.
The oil characteristics in play in hydrodynamic lubrication are
cohesion (how well the molecules adhere to each other),
adhesion (how well molecules stick to other molecules), and
viscosity (an oils resistance to fow).
If an oil has poor cohesion, it would have poor flm strength
and the flm could be easily separated. If the oil had
absolutely no adhesion to the crankshaft journal, then an oil
flm would not form on the crankshaft journal. The journal
would then be forced under load to the bottom bearing
insert and attempt to climb it just as if there was no oil. This
pinpoints oil adhesion as a fundamental part of hydrodynamic
oil lubrication. Viscosity is the dominant factor in an otherwise
suitable oil which determines the resulting flm thickness.
As shown in fgure 5, the adhesion of the oil to the spinning
journal will cause the oil to be continually dragged into the
point of contact between the journal and the bearing insert.
The oil is entering the contact point from the left in our
example, so the oil flm thickness will be greater on the left
side of the contact than on the right side. This gives the oil
flm a shape called a hydrodynamic wedge. With the thickest
layer on the left, the pressure in the wedge pushes the journal
up and to the right of the load center line.
As the hydrodynamic wedge pushes the journal away from the
bearing insert, the pressure in the wedge will drop due to the
increased clearance between the journal and bearing insert.
At the point where the upward force of the wedge equals the
load, the hydrodynamic lubrication is in equilibrium. In this
equilibrium, the oil pressure in the hydrodynamic wedge part
of the oil flm can be many times the engine oiling system
Since the journal in this system is the pumping component,
the pressure in the hydrodynamic wedge is proportional to the
speed of rotation. This means the slower a journal rotates,
the lower the hydrodynamic pressure becomes, resulting in a
thinner oil flm, as shown in the Stribeck Curve of fgure 2.
This is one reason why lugging an engine at very low rpms
and high load can cause bearing damage. In a lugging
condition, the oil flm being generated by the hydrodynamic
action can be too thin to sustain the load, and may allow
journal-to-bearing contact.
As is the case with all gas or liquid compression systems,
energy is used compressing the oil in the wedge, and the oil is heated in the process. This energy is supplied by the
engine, and the heat generated in the process adds to the heat in the oil from oil shear. Both must be dissipated in
the sump.
The correct viscosity oil is a critical factor to maintaining adequate hydrodynamic lubrication. The factors determining
the choice of the correct oil are the maximum load, bearing width, and rpm range in which the bearing will operate. High
viscosity oil will allow for high loading at a lower rpm range. Low viscosity oil will allow for higher rpm operation with low
shear heating and less power loss.
fgure 3 - Crankshaft Stationary
fgure 4 - Crankshaft Spinning Without Oil
fgure 5 - Crankshaft Spinning with Oil
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January 13, 2009
Factors Which Determine Hydrodynamic Oil Film Thickness
Specifed Oil Viscosity
Hydrodynamic oil flm thickness is directly proportional to the viscosity of the oil, which in turn is related to the cohesion
of oil molecules to each other in combination with the average molecule size. The easiest way to gain load safety margin
in the form of additional oil flm thickness, is by increasing the viscosity of the oil. The correct choice of oil viscosity is
determined by the design load, bearing clearance, and rpm range of the engine.
In general, the higher the load, the higher the appropriate viscosity. The larger the bearing clearance, the higher the
appropriate viscosity. The higher the average rpm range, the lower the appropriate viscosity. Low viscosity oil with low
shearing losses will increase energy effciency and prolong oil life. More viscous oil will allow for heavier loading or lower
rpm use, but suffer excess oil shearing under high rpm duty.
Engine manufacturers have suggested viscosities which are designed to take these factors into account, as well as
factoring in seasonal variations in operating temperatures. Racing oils may seem to violate these constraints if you judge
them solely by their viscosity ratings, which can be as high as SAE 70 weight. Racing engines cannot escape from the
laws of physics, and these highly viscous oils will suffer from extreme shear, heating, and degradation when the engine is
operated at high rpms. This consequence is unimportant in this application, because the oil is often discarded after mere
hours of use. Street vehicles operate with very different oil life requirements.
Temperature of the Oil
Oil temperature directly affects the viscosity of the oil, as discussed in ZPlus Tech Brief #10. This lowering of viscosity
with temperature increase translates directly into a reduction of flm thickness. This characteristic is fortunate for a
commuter on a cold morning, considering how little oil pressure is available in a cold engine upon start-up. When cold,
residual oil in the bearing will have a much greater viscosity than it will at the engines full operating temperature, which
lowers the chance of bearing damage. A hot re-start of an engine poses more of a risk of bearing damage, due to the
fact the hot oil flm will be thinner than a cold oil flm, and may not adequately separate the journal and bearing until proper
oil system pressure is re-established.
Rotational Speed of the Journal
As the adhesion of the oil to the spinning journal drags the oil molecules into the hydrodynamic wedge, the cohesion of
the molecules will cause a buildup of pressure in the wedge. This increased pressure will directly translate into increased
flm thickness.
Peak Pressure Due to Loading of the Bearing (Not Oil System Pressure)
A doubling of load will result in a halving of flm thickness, if you do not factor in the or pressure dependent viscosity of
the oil. Practically speaking, the viscosity of most PCEO will double from zero pressure to 10,000 psi.
Peak pressures in
the crank bearings of a normal IC motor are less than 2000 psi, so the of an average oil will cause a useful but minimal
10% increase in flm thickness, relative to that of its uncompressed viscosity.
Polarit y of the oil
The polarity of the oil directly affects its adhesion to the
parts it is supposed to lubricate. Polar substances can
be mutually attractive, which enhances the adhesion of
the two.
As shown in fgure 6, when a drop of water is placed in
contact with surfaces of different polarities, the resulting fgure 6 - Polarity Affecting Adhesion
interaction of the two determines the degree of wetting.
Totten, G.R., Fuels and Lubricants Handbook: Technology, Properties, Performance and Testing, ASTM Manual Series, ASTM International, West
Conshocken, PA, 2003, pg. 189
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The greater the adhesion, the more the liquid is attracted to the surface, and the more it will spread out on the surface.
The amount of adhesion can be determined by the resulting angle at the junction of the drop as it wets the surface. This
wetting can be expressed as a contact angle , with the greatest adhesion giving the lowest angle. This contact angle is
also determined by the surface fnish, which is discussed in the next section.
Highly polar oil displays greater adhesion to both bearing and journal, resulting in greater boundary layer thickness and
higher ultimate flm rupture strength, all other characteristics being equal. The cam, lifter, crankshaft, and bearing insert
are made from highly polar metals. PCEO polarity is dependent on the polarity of both the base stock and additive
package. Mineral-based Group I, II and III base stocks are moderately polar, with polarity decreasing in general as
additional refning is added. There are some base stocks, such as Group IV PAO, which are very non-polar. Some
Group V oil stocks, like PE (polyol ester) and PAG (polyalkaline glycol), are highly polar (see ZPlus Tech Brief #10 for
more oil information).
Without polar additives, Group IV PAO base stock will exhibit a lower degree of wetting of engine parts compared to
other base oils. This is one of the reasons that most current true synthetic Group IV based PCEO is formulated with the
addition of another more polar base stock, such as a Group V polyol ester. Of course, the addition of a polar additive
such as a detergent or an anti wear additive such as ZDDP will also increase the overall polarity of the oil. Without the
addition of these polar substances, a PAO oil flm can be more easily displaced due to the lower adhesion and reduced
hydrodynamic wedge development. This can result in metal-to-metal contact between the bearing journal and insert.
A highly polar liquid will display more thorough wetting of the bearing interface (high adhesion) and displace less under
load, other factors being equal.

Surface Finish
A perfectly smooth bearing surface, although impossible to attain, would have no reserve oil capacity, as well as minimal
mechanical oil adhesion. The microscopic surface imperfections of asperities and small pits greatly increase the surface
area of the journal, and allow for greater oil flm adhesion. The limiting factor on the height of these surface imperfections
is the thickness of the oil flm itself. If the asperities are taller (above the average surface) than the oil flm is thick, then
contact will be made between the rotating and stationary asperities, causing wear.
Surface Finish (RMS) Smooth (<10nm) Low (20nm) High (50nm)
Contact Angle 33 28 22
Recovered Mass (mg) 5 8 10
The DCA used in this University of California fgure 7 - Surface Finish vs. Oil Adhesion
Santa Barbara study measures the adhesion
between oil and metal strips of varying roughness, by measuring the contact angle or amount of wetting which occurs
between the two. The recovered mass in the chart is the amount of oil which adhered to the test metal strip due to
the combination of polarity and the surface roughness. Since the roughness was the parameter that was varied in the
test, the difference between the results is due solely to the surface fnish. The rougher the surface fnish, the more oil
was retained. Although increasing roughness will increase oil adhesion, there is a practical limit. At some amount of
roughness, the height of the asperities will begin to approach the thickness of the oil flm, reducing the effective oil flm
thickness. For each bearing system, there is an optimum surface fnish which will have maximum oil adhesion without
causing contact between the two bearing surface asperities under load.
Engine bearing manufacturers have long known there is an optimum surface fnish for the bearing journals which
encourages the formation and adhesion of an oil flm. The importance of the surface fnish is demonstrated by Clevite
regarding the directionality of the fnish left by cylindrical grinding and polishing. The process of
grinding or polishing at the microscopic level is one of imbedding a hard chunk of abrasive into the relatively soft metal,
and then moving it parallel to the surface to remove or push over the metal. Some metal will be removed, and some fows
plastically, leaving a directional scaled appearing surface aligned in the direction of the fnishing rotation. This surface
feature can help retain oil by mechanically increasing the adhesion of the oil flm to the bearing journal and providing
microscopic oil reservoirs. These scales also can appear as microscopic ramps, perhaps helping hydrodynamically
pump oil outward on the journal.

Victoria Broje and Arturo A. Keller, Advanced Oil Spill Recovery in Marine Environments, Bren School of Environmental Science and Management,
University of California, Santa Barbara, 2005
Figure 7 details the adhesion of oil to steel of
increasing roughness as determined by oil
adhesion experiments using a DCA (Dynamic
Contact Angle) analyzer.
For these reasons, the direction of grinding and polishing relative to the direction of bearing rotation during operation is
critical to the performance of the bearing/oil system.
Width of the Bearing
In a plain bearing of infnite width, there could be no
side-leakage of the high pressure in the hydrodynamic
wedge. Since all IC engines have bearings of relatively small
width compared to the
shaft diameter (typ.<0.5 dia.), and fairly large clearance
between the journal and bearing insert relative to the size of
the oil molecules, there is considerable side-leakage. This
leakage allows the high-pressure oil in the hydrodynamic
wedge to leak out along the length of the bearing.
As a result of the side leakage, there is a bell-shaped
pressure gradient of this oil pressure wedge as viewed
perpendicularly to the shaft, shown on fgure 8. In this
diagram, the red zone of highest pressure is in the center
of the bearing, causing a downward pressure gradient as
you move to either edge of the bearing. The net result is
a reduction in the peak pressure of the bearing surface
covered by the wedge. The load-bearing capability of a
hydrodynamic wedge is proportional to both bearing width
and journal rpm, so a wider bearing will support adequate
hydrodynamic lubrication at a lower speed. The correct width
for a bearing is one which allows for a hydrodynamic oil flm thick enough to support the load with an adequate safety
margin, without being so wide as to incur excessive oil shear power loss.
What is the Actual Minimum Film Thickness Needed to Keep the Bearing and Journal Apart?
Figure 9 is a graph plotting measured crankshaft bearing load and minimum oil flm thickness as a function of crank

The data in the chart shows graphically how the crankshaft journal does not stay centered in its bearing when loaded.
Paranjpe, Rohit S., Tseregounis, Spyros I. and Viola, Michael B. (2000), Comparison between Theoretical Calculations and Oil Film Thickness
Measurements Using the Total Capacitance Method for Crankshaft Bearings in a Firing Engine, Tribology Transactions, 43:3, pp. 345-356
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fgure 8 - Hydrodynamic Pressure Profle with Side Leakage
fgure 9 - Crankshaft Load and Oil Film Minimum Thickness vs. Crankshaft Angle
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Note the chart does not specify at what location that minimum occurs, just the angle of the crankshaft in its rotation
relative to TDC (Top Dead Center) when the minimum oil flm thickness occurs.
By this we mean the minimum flm thickness will, of course, be located near the middle of the bearing cap during the
power stroke, and at the top near the oil hole during the intake stroke. But both events cause a relative minima on the
graph. The data in the chart was obtained by testing with 5W-20 oil, and results show an average minimum flm thickness
of 0.000264. The test engine was a GM 3800 V6, with radial main bearing clearance of 0.0012. This means in the initial
test, due to radial forces in the crankshaft, the actual journal spends most of its time rotating off-center in the bearing
housing an average of 0.000936, which is most of the 0.0012 clearance! The critical piece of information in this chart
is the absolute minimum oil flm of 0.000047, which occurs in this particular bearing position at around 300. Retesting
using 10W-30 oil gave a minimum of 0.000054, and 20W-50 oil resulted in a minimum flm thickness of 0.000070. This
means in this particular engine using 5W-30 oil, if the asperities on either the crankshaft or bearing insert were higher
than 0.000047 total, there could be contact between them at this time. Once the asperities of the journal contact those
on the bearing, some will break off and be carried in the oil as debris which may cause additional damage. This is one
of the reasons the current high-performance engine design emphasizes the roundness, surface fnish, and accuracy of
machining of journals and bearings which are supported by hydrodynamic lubrication.
Since the sum of the machining, bearing manufacturing tolerances, and installation tolerances far exceed this minimum oil
flm thickness, why does a crankshaft bearing not fail immediately upon startup? The answer is the bearing insert surface
material is relatively soft and compliant, and when an engine is frst fred up, the crankshaft DOES make contact with
enough force to literally pound the high spots and misalignments into a close ft with the journal. Once they have been
reformed to a height less than the oil flm thickness, the crankshaft is fully supported by the hydrodynamic flm.
It is obvious that any and all precision employed during assembly benefts bearing life.
Oil Shear
Figure 10 shows the relative velocity at different
layers in the oil flm. Referring to the illustration,
the oil flm directly in contact with the bearing
insert is adhered to the metal of the bearing and
does not move. In a liquid, molecules can be
moved relative to each other, so the layer of oil
molecules adhering to the stationary layer does
have some freedom of movement. The layer of
oil at the crankshaft journal is adhered to the
journal, and will move at the full rotational speed
of the shaft. The intermediate layers of oil,
starting with the stationary bearing insert layer,
have gradually increasing velocity crossing the
thickness of the flm towards the crankshaft.
These oil molecules, in the act of moving past
each other, will adhere to each other then break
free. This action of pulling the oil molecules
apart is called shearing, and requires energy
proportional to the viscosity of the oil. The
constant shearing of the oil around a spinning
crankshaft absorbs energy and is called viscous
drag. The energy of the shearing (or the shear
energy) is dissipated as heat in the oil.
The heat caused by shear is a primary factor in
the choice of oil viscosity. If too high viscosity oil
is chosen for a specifc bearing clearance and
shaft speed, the oil will be heated by the shearing
to a temperature which may cause
decomposition of either the base oil or an
additive. The longest, highest weight molecules
fgure 10 - Oil Shear Between Moving Journal and Bearing
in the oil are subject to the most shear, and it is often the VI (Viscosity Improver) additive molecules which suffer the
greatest degradation as a result of shearing. If the VI molecules degrade, the oil loses the effect of the VI additive and will
drop in viscosity as well as show a greater change in viscosity with temperature change.
In order to gain an appreciation of how important shear is to effcient engine design, consider that in one paper
it was
calculated that a standard European 2 liter 4 cylinder engine running 7500 rpm using 10W-30 oil is stated to waste
more than 4000 watts (5.4 horsepower) shearing the oil in the bearings alone. The same engine in the same operating
conditions using 0W-20 oil would waste 2700 watts (3.9 horsepower).
Summarizing Factors Affecting Hydrodynamic Oil Film Thickness
A hydrodynamic oil flm will support a journal until a critical design point of: increasing load, too low rotational speed, or
insuffcient oil viscosity. At that point, the oil flm will be thinner than the asperities are tall. Once the asperities of the
journal contact those on the bearing, there is the potential for damage to both surfaces.
Sliding Hydrodynamic Lubrication
The pushrod ends, rocker ball, rocker arm ends and oil pump gears operate in a sliding, relatively lightly-loaded
hydrodynamic/hydrostatic lubrication regime. In a manner similar to rotating hydrodynamic lubrication, the flm of oil
between the parts is pressurized by the actual contact pressure. In this case, the pressurized oil moves from side to
side or back and forth in an arc, creating a hydrodynamic wedge of oil which keeps the surfaces separate. As the oil flm
attempts to squeeze out ahead of the applied pressure, the movement reverses which reverses the direction of the oil
wedge, keeping the oil (except for side leakage) in the bearing gap. In these systems, there is an ample supply of oil to
replenish the flm on the edge of the moving parts. Due to the fact there is relatively little shear in the small movements,
these systems would beneft from an oil with high viscosity. In normal gear operation, the lubrication is mixed boundary/
elasto-hydrodynamic. In a gear oil pump, one gear is driven and the other is actually driven by the oil flm itself and is not
coupled to a load other than the compressed oil, resulting in relatively low loading.
The lifter bore, piston rings, skirt and valve stem also operate in a lightly-loaded sliding contact hydrodynamic regime.
Unlike the pushrod ends and rocker balls, the movement of these parts relative to the stationary engine block is large. To
keep power losses low, these parts would beneft from a low-to-moderate viscosity oil chosen for its temperature stability
and low shear loss. Due to the relatively large amount of contamination to which these parts are subject, the oil must
have a good amount of chemical stability as well.
EHD (Elasto-HydroDynamic) Lubrication
Defnition: The opposing sur faces are separated but there occurs some interaction bet ween the raised solid
features called asperities, and there is an elastic deformation on the contacting sur face enlarging the load
bearing area whereby the viscous resistance of the lubricant becomes capable of supporting the load.
For roller and ball bearings: In rolling element bearings, the elastic deformation of the bearing (fattening) as
it rolls, under load, in the bearing race. This momentary fattening improves the hydrodynamic lubrication
EHD lubrication occurs when a volume of oil in a flm is trapped by surface characteristics or the velocity of the bearing
surfaces. Since it cannot escape, the load pressure of the bearing is transmitted through the trapped flm of oil, and
can momentarily become great enough to exceed the strength of the base metal, causing deformation of the metal. In
other EHD lubrication scenarios, under suffcient pressure, oil with a large factor will have a viscosity increase and in
doing so transfer the load between bearing members without metal-to-metal contact. An important aspect of EHD, as
the defnition points out, is the pressure which can be transferred through the trapped or thickened lubricant flm can be
great enough to deform the two metal bearing surfaces. In doing this, a small contact patch will broaden into a larger
one, spreading the pressure over a larger area. This lowers the effective pressure per unit area, which can allow the flm
to carry the load without rupture. This characteristic is valuable in rolling contact bearings, such as ball or roller bearings,
where the point or line contact can cause extreme localized pressures and temperatures at the contact point.
Taylor , R.I., Lubrication, Tribology & Motorsport, Shell Global Solutions (UK), Cheshire Innovation Park, PO Box 1, Chester, CH1 3SH, UK, 2002-01-
3355, pg. 10
Noria Corporation, Dictionary of Tribological Terms 1328 E. 43rd Court Tulsa, OK 74105
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Crank and crankpin bearings usually operate in hydrodynamic lubrication regime where the peak pressures are lower, and
cause little viscosity increase.
In some high-performance engines, especially at high rpms, the crankpin bearings can effectively employ EHD
lubrication, since the connecting rod is more fexible than the engine casting in the main bearing area. This allows the
crankpin to effectively distort the rod bearing material, thereby enlarging the contact area.
The cam/lifter interface, and in certain diesel engines, heavily-loaded wrist pins are the only engine systems where
the oil is subjected to suffcient load for the pressure/viscosity factor to be a major factor in flm thickness. Some
loads typical of those found in an IC engine are shown in fgure 1. Even 5000 psi at a heavily-loaded wrist pin bearing
surface is insuffcient to radically change the oil viscosity. The cam/lifter interface, which is a sliding point or line contact
interface, operates in a mixed elasto-hydrodynamic/boundary lubrication regime. This system would operate best with a
very high viscosity oil with a high , or grease, either with the appropriate anti-wear agent. Many engines with low spring
pressures or mild cam profles have low enough lifter foot pressures to employ EHD lubrication alone at all cam angles
(refer to fgure 11). Modern engines with roller lifters employ EHD lubrication at the rolling point of contact between the
lifter and camshaft.
Boundary Lubrication
Defnition: Form of lubrication between two rubbing surfaces without development of a full-fuid lubricating flm.
Boundar y lubrication can be made more ef fecti ve by including additi ves in the lubricating oil that provide
a stronger oil flm, thus preventing excessive friction and possible scoring.
In order to better understand the lubrication modes employed in the cam lobe/lifter contact, examine fgure 11, which
plots the valve lift, velocity, acceleration and oil flm thickness of the cam lobe/lifter interface. The oil flm thickness is
consistent in the base circle area of rotation, as the lifter slides on a wedge of oil preceding the contact point. This is a
fully-fooded hydrodynamic lubrication mode.
As the lifter begins to climb the entrance ramp of the lobe, the wedge shaped gap formed by the spherical lifter foot and
cam lobe trap oil. As the follower climbs the lobe, and the spring compresses, the lifter foot pressure increases, and
at some point puts the trapped oil flm into an EHD lubrication mode. As the lifter foot continues to climb the cam lobe
entrance ramp, the increasing spring pressure reduces the oil flm thickness, reaching minima at 145 and 215. The
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January 13, 2009
fgure 11 - Cam and Lifter Characteristics and Timing
Loughborough University study data concluded the oil flm thickness at these
minima was less than 2x10
meters (0.2 microns) or about 0.000008, which
is 8 micro inches.
The study concluded of the cam lobe/lifter contact: The peak transient contact
pressures remain in the gigapascal range.
One gigapascal is equal to
145,037.738 pounds per square inch. This contact pressure was calculated for
what we would consider to be typical of the valve train in a low-performance
engine, with only 135 pounds of valve open lifter foot pressure.
A common classic and high-performance engine found in the U.S. is the Chevrolet
350 engine. We obtained a fat-lifter camshaft with lifters in order to get an idea
of the lifter foot pressure which could be found in one of these engines. Using an
optical comparator, we examined the foot of the GM lifter and cam lobe. Figure
12 shows a simplifed view of the relationship between the two, with both the
sphericity of the lifter foot and the taper of the cam lobe exaggerated for clarity.
The lobe taper and spherical lifter foot profle, together with the lifter offset from
the cam lobe center line, induce a spinning/skidding motion to the lifter. This
motion reduces wear by several methods. First, it spreads the contact point over a
large circular area on the lifter foot instead of concentrating it on a single patch. It
also helps to bring lifter metal with a fresh oil coating into the contact point. This
ensures that an effective EHD flm is maintained in lightly-loaded conditions, and in
heavily-loaded conditions, that fresh anti-wear agent is brought to the contact point.
The geometry of the lifter/cam lobe contact causes a tight mating of the two, which
encourages the contact patch to be as broad as possible, as well as encouraging
the formation of an EHD oil wedge. The GM lifter foot was ground with a 36 radius
spherical profle, and the cam lobe had a face taper of 0.003 across the 0.450 cam lobe width. This means the angle of
the cams lobe surface is 0.387 relative to the axis of the cam, and the spherical foot of the lifter will touch the lobe at a
point with a tangent of 0.371. With little or no pressure, the two meet at a point. But, under the high loading of the valve
spring, the resulting plastic deformation of the cams surface and lifter foot cause the point contact to lengthen into a line
or rectangular patch contact. The cam and lifter we obtained are often used with springs which can have as much as 335
pounds of pressure at full lift. With 1.5:1 ratio rocker arms, that results in 500 pounds of lifter foot pressure, or 3.7 times
that used in the Loughborough study. We measured a static contact patch of 0.3 mm by 1.3 mm under the 500 pound
load. This would indicate the metal of the cam lobe and lifter were each being deformed by several microns, and the
resulting equilibrium pressure was in the 200,000-400,000 psi range. The actual pressure will be limited by the amount of
deformation of the cam lobe and lifter. This is because in response to pressure the contact patch gets larger, spreading
the pressure over a larger area, which in turn lowers the peak pressure per unit area.
Due to the diffculty of actually measuring the area of the contact point as it moves dynamically, as well as the complex
deformation of the cam and lifter metal in response to pressure this great, we cannot know the exact pressure at the
contact point using our GM parts. We can state with confdence that a 500 pound lifter foot pressure will cause a contact
pressure not less than that stated in the Loughborough study using a 135 pound lifter foot load, and possibly up to 3.7
times as much. This would mean the contact pressure is in the range of 145,000-536,000 psi. At these pressures, the
deformation of the metal will limit the actual pressure per unit area to a value less than the yield limit of the metals.
We have seen a number of failed lifters and cams, both in OEM engines and in engines employing aftermarket fat cam/
lifter systems which have aggressive ramps and high spring pressures. This indicates to us, unlike the low-performance
engine used as an example in the Loughborough study, effective EHD flms are not being sustained in many high-
performance engine applications. In these cases of cam lobe destruction, it is clear that regular API SM oils are not able
to provide EHD flm lubrication at the points in the cam rotation where the oil flm is minimum thickness. If there is no
EHD oil flm, the only mode of lubrication which can inhibit wear under these conditions is boundary. In boundary
lubrication mode, the protection for the two bearing surfaces will be provided by the anti-wear agent in the oil, which has
historically been ZDDP. Current API SM oils do indeed still have ZDDP, but the evidence says there is not a suffcient
amount to protect high-performance engines from wear. The characteristics of ZDDP cause it to decompose into a
glassy high-strength flm in the presence of the high pressure and temperatures at the cam lobe and lifter contact point.
This flm, while not as low-friction a lubricant as an oil flm, is the last line of defense against wear, and effectively forms a
sacrifcial layer to protect the base metal of the cam lobe and lifter foot. Since it is sacrifcial, the rate of flm build will be
M Kushwaha and H Rahnejat, Transient elastohydrodynamic lubrication of fnite line conjunction of cam to follower concentrated contact, Wolfson
School of Mechanical and Manufacturing Engineering, Loughborough University, Loughborough, UK, pg. 2882
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ZPlus Tech Brief #11
Copyright 2008 ZPlus, LLC
January 13, 2009
fgure 12 - Ideal Cam / Lifter Contact
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ZPlus Tech Brief #11
Copyright 2008 ZPlus, LLC
January 13, 2009
partially dependent on the amount of ZDDP available in the oil. High spring pressures or rpms will wear away the flm at a
greater than normal rate, increasing the required amount of ZDDP needed to ensure low wear on the cam lobes and
lifter foot.
Lubricating an Engine in Reality
Back here in reality, all of our engines are designed with a single lubrication system which is required to handle ALL of the
lubrication chores in all engine systems. In addition, the oil is asked to move many kilowatts of heat from critical engine
parts to the sump, where it is expected to transfer the heat to the sump walls and return cooled for more abuse. As if this
wasnt enough of a job, the oil is also supposed to capture all blow-by and atmospheric moisture condensates, absorb,
neutralize and store them until the next oil change, and protect all internal components from corrosion. The oil is asked
to perform these chores for 3000 to 7500 miles or more, depending on how often you empty the sump. With the most
recent innovations in precision engine design and fuel injection, there is a large push in the industry, especially in Europe,
for sealed-for-life oil charges in the sump. We can expect PCEO to evolve with engine design towards this goal.
The quality and performance of todays base oil stocks far exceeds anything available in the past. Unfortunately, the
change in engine design and the resulting change in oil composition means the newest oils do not necessarily contain the
best additive package for older engines. Specifcally in the area of anti-wear agents for the boundary lubrication regime,
the newest oils exhibit less anti-wear capabilities than do the heaviest duty oils containing high amounts of ZDDP.
Obviously, a single oil cannot optimally lubricate all engine systems, while at the same time performing all of the other
jobs it is asked to do. This means that like all other engineered substances, it is the product of compromises in design.
In many cases, the very best oil to use is the one that provides optimum protection for the weakest part of your specifc
engine. In the case of fat-tappet engines, especially those with high valve-spring pressures, you must be careful to
ensure an adequate level of the anti-wear agent ZDDP.