NLGII
Deve~opment of a Bench Test MethodoDogy for Performance Eva~uat~on
of Energy Effic~ent Qndustr~a~ Gear OHs
Ajay Kumar Harinarain, Samuel Pappy, lnder Singh, A K Jaiswal, F Sayanna, B Basu, R K Malhotra. Indian Oil
Corporation Ltd, R&D Centre, Faridabad, Haryana, India.
Presented at NLGI’s 80th Annual Meeting, June, 2013, Tucson, Arizona, USA
2.0 Introduction
3.0 Basis of Test Methodology
There is an enormous amount of interest in saving
energy in the current age looking to the depletion of
fossil fuels and the time to develop potential alternative
energy sources. The maxim behind this is “Energy Saved
is Energy Generated”. This is true for almost all energy
intensive equipments (Automotive as well as indus
trial), and efforts are on to redesign the lubricant using
viscosity modification as well as an additive approach.
Generally for industrial gear applications, use of gear
oils formulated with friction modifiers has been generally
the main approach followed by most of the lubricant
majors. Although a number of laboratory bench scale
test methods are available for evaluating the various
load carrying parameters of lubricants, reliable methods
for predicting the energy efficiency characteristics are
rather limited with respect to the laboratory and field
data correlation. Attempts have been made by many
researchers to establish test methods for the evaluation
of energy efficiency characteristics of gear oils [1 -3].
Among these, ASTM axle efficiency test [2] for energy
efficient automotive gear oils and CEC ECOTRONS test
method [4] are in the process of adoption for the evalu
ation of EE characteristics. The ASTM test however is
very cumbersome and expensive and not many labora
tories can acquire it. Besides, it is not possible to get a
complete assessment of the performance using a single
test method, so a combination of test methods in form
of a test methodology is being used for the development
of an energy efficient industrial and automotive gear oil.
The basic requirements of the methodology are to use
reliable, simple and accurate laboratory methods which
could assess the various aspects of the energy efficiency
characteristics quickly. The present work describes the
various evaluations carried out in the development of this
laboratory test methodology.
In general, gear contacts work under the boundary
and EHD lubrication regimes. It is therefore important
to assess the gear lubricants to predict the overall
effect under these regimes on the gear efficiency. So it
was decided to follow a step by step approach to first
assess the lubricants under different regimes of lubrica
tion, using different gear oils formulated and with known
performance. Various laboratory bench tests were used.
The SRV friction and wear test [6] to assess the fric
tional characteristics, EHL Ultra Thin Film measurement
system for the mapping of the oil film thickness over a
wide speed range, from 20 mm/sec to 4 metres/sec, by
optical interferometry [7] and the traction performance
under simulated slide roll ratio as experienced in a gear
contact, and the Block on Ring test [8] machine for
measurement of boundary friction and the temperature
rise in a simulated gear contact were used for assess
ing the factors responsible for the Energ~y efficiency.
Finally a gear box properly instrumented to precisely
control the test parameters using the popular FZG gear
test machine was selected to validate the results since
it was best suited for this purpose.
40 Description of Test Equipments
1. EHL Ultra Thin Film Measurement System is
a computer-controlled instrument for measuring
the film thickness and traction coefficient (friction
coefficient) of lubricants in the elastohydrodynamic
(EHL) lubricating regime. It is as shown in Figure 1
(General schematic), Figure 2 (Film thickness) and
Figure 3 (Traction). In the Film thickness mode, the
instrument can measure lubricant film thickness
down to 1 nm (1 millionth of a millimeter) with a
precision of +/- 1 nm. In the traction mode, the
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NLGI SPOKESMAN, MAY/JUNE 2014
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Traction coefficient can be measured at any slide!
roll ratio from pure rolling up to 100% sliding. The
instrument measures these lubricant properties in
the contact formed between a steel ball and a rotat
ing glass or steel disk. The contact pressures and
shear rates in this contact can be maintained close
to those experienced by the oil in gear boxes. The
EHL system measures film thickness and traction
coefficient in the EHL contact formed between a
3!4 inch steel ball and a rotating 100 mm diameter
disc. For film thickness measurements, the disc is
made from glass and has a chromium and silica
coating on the working face. For traction measure
ments, the disc is made from polished bearing steel.
The ball is mechanically loaded against the under
side of the disc and can be allowed to rotate freely
or can be driven to induce sliding between the ball
and disc. The load is controlled automatically and is
variable between 0 and 30 N. This gives maximum
contact pressures between the ball and disc of up to
approximately 1 .1 GPa with a steel disc and 0.7 GPa
with a glass disc.
The ball and disc are independently driven by DC
servomotors. With standard gearing, the maximum
rolling speed is 4 m/s and the minimum rolling speed
is 25 mm!s. The ball can be allowed to idle freely
in nominal pure rolling, or a drilled ball and drive
shaft can be fitted. The ball can then be driven at
any desired slide!roll ratio, the required ball and
disc speeds being determined automatically by the
control software. The traction force is measured by
a high sensitivity torque transducer between the ball
motor and the ball. The oil sample to be tested is
contained in a reservoir constructed from a single
stainless steel block. Heaters are fitted to allow
measurements at temperatures from ambient up to
150°C; however, our tests were carried out at test
temperatures of 60°C representative of the tempera
tures experienced in an industrial gear box.
The lubricant film thickness is measured by optical
interferometry. The contact is illuminated by a white
light source directed down a microscope through
the glass disc onto the contact. Part of the light
is reflected from the chrome layer on the disc and
part travels through the silica layer and any fluid
film and is reflected back from the steel ball. The
recombining light paths form an interference image
which is passed into a spectrometer and then into a
high resolution CCD camera. The camera image is
captured by a video frame grabber and analyzed by
the control software to determine film thickness. The
thin film software takes data from the spectrometer,
determines the wavelength of maximum constructive
interference and hence the lubricant film thickness.
The traction software allows the user to select rolling
speeds and slide! roll ratios and logs data from the
ball torque and load transducers and calculates the
Figure 2
—
Optical Film Thickness Apparatus
©PCS Instruments 1999
Sectional view through pot showing disc and ball drive assemblies
Figure 1
—
Film Thickness and_Traction Rig Assembly
Figure 3 —Traction Test Apparatus (Steel on Steel)
,
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VOLUME 78, NUMBER 2
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resulting traction coefficient. These values are mea
sured and stored in the PC using a data acquisition
system.
Load
Friction
~
—*—
Block
—~—
—
Ring
Steel or Ceramic
–
Steel
FALEX 1 Test Rig
Figure 4 — Falex Block on Ring test rig
Figure 5
—
2. Falex Block on Ring test rig is used to simulate
the gear contact (line) under boundary regimes. It is
mainly used for the friction reduction capabilities of
the gear oils chosen for the study. The machine con
sists of a variable speed motor which drives a shaff
on which a ring is rotated, against which a stationary
rectangular test block is loaded. The schematic is
shown in Figure 4. The Ring and block are made up
of Standard Hardened Gear Material. The test block,
which is held stationary against the revolving ring,
is restrained from horizontal movement, The load is
accurately maintained throughout the test, using a
loading lever on which a load cell is mounted to mea
sure the resulting friction between the block and ring.
The oil is kept in a standard enclosed test chamber
and is carried by the rotating ring to the zone of con
tact. As a result of the frictional heating the tempera
ture rises, and is measured during the test. Higher is
the temperature, higher is the friction coefficient, and
lower is the expected energy efficiency.
3. SRV friction and wear test is used to measure
the boundary frictional characteristics of the oils
in a point contact configuration as shown in
Figures 5 & 6. A steel ball of 10 mm diameter made
of bearing steel (1 OOMnCr5) is oscillated against a
flat steel disc under 200 N load, 1 mm amplitude of
oscillation at 50 Hz frequency, and the temperature
of the oil is kept at 50°C. The friction is being mea
sured over a test duration of 1 hour, and recorded
on the computer.
SRV Test machine (Ball on Flat Disc)
I
Figure 6 — SRV Test configuration
Figure 7
—
Modified FZG Test schematic
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NLGI SPOKESMAN, MAY/JUNE
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4. FZG Test Rig: The principle of power circulation of
the FZG test rig is favourable for the measurement
of friction losses, since by construction, the electric
motor overcomes the losses. The power consumed
by the motor is the sum of: (a) The internal losses of
the motor and the friction losses in the coupling etc.
(which can be reasonably taken as constant) (b) The
frictional losses between gear teeth and the bearings
of the machine (which is affected by oil formulation).
Thus the effect of the oil on friction losses can be
evaluated by the measurement of electric power [4].
It avoids the more delicate and cumbersome task
of measuring torques of rotating shafts. This can
precisely distinguish the narrow differences in energy
consumed with two oils.
The principle of the test is to measure the electric
power consumption using a precise microproces
sor controlled energy meter at three standard FZG
test load stages (4, 6 and 8) at speeds of 1500 rpm
maintaining the oil at a constant oil temperature of
80°C (using heating coils and cooling water) for a
running period of 1 hour at each load stage. The test
condition enabled us to assess the effect of both
friction modifier (boundary) and viscosity
(hydrodynamic) effect for the automotive gear oils.
A schematic of the test setup is given at Figure 7.
Studies were conducted on five different oils (Table 1)
taking a VG320 Mineral oil (Oil A) as the baseline
and comparing two fully synthetic VG 220 oils, Oil B
and Oil C, as well as two VG320 semi synthetic
and Friction modified oils (Oils D, E,) with it. The film
thickness mapping at 20N load (0.48 GPa contact
pressure), 60°C, and speed ranging from 20 mm!
sec to 4 metres per sec was done for assessment
of film thickness as shown in Figure 8. The SRV fric
tion studies in an oscillating sliding contact at 200N
load, 50 deg C, 50 Hz, 1 mm for 1 hour are shown
Table 1
i~’r~
1
2
3
4
5
A
B
C
D
E
~
Mineral
Synthetic
Synthetic
Part Synthetic
Part Synthetic
~r
‘
~‘~~Iie~tri
VG 320
VG 220
VG 220
VG 320
VG 320
~
Film thickness @ 20N, 60 Deg C, 0-4mlsec Speed
1100
1000
900
800
~ 700
600
.M
500
~5 400
U-
300
200
100
0
O.OOE+OO
5.QOE-O1
1.OOE+OO
1.50E+OO
2.OOE+OO
2.50E+OO
Disc Speed, rn/sec
Figure 8
—
Film thickness data by Optical Interterometry
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VOLUME 78, NUMBER 2
3.OOE+OO
3.50E+OO
4.OOE+OO
4.50E+OO
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at Figure 9. The Traction data at 30N load, speeds
ranging from 20 mm/sec to 4 metres per second
at 30% Slide Roll ratio corresponding to a spur
gear meshing is shown in Figure 12. The Block on
Ring Frictional performance in terms of oil tempera
ture rise and the frictional coefficient is shown in
figures 10 & 11, and the reduction in Energy con
sumed in FZG gear test (in %) shown at figure 13.
5~O Test Observations & Discussions
From the bench tests as shown in the results and
figures, we have the following observations that can
be made regarding the oil performance.
(1) The Film thickness and Traction coefficient trends
match with each other, this is due to lower pressure
viscosity coefficient for the synthetic oils, and
SRv @ 200N,
50 Deg
c and
1 Hour
0.
0. 12~
C
Co
0)
0.11
E
0.10
0
C
0
0.09
.;//~
.
.
. .
~.
/~
C)
U-
~.
0.08
0.07
00:20:00
00:10:00
OllA
Figure 9
—
00:30:00
00:40:00
00:50:00
ollc
Oil D
Oil E
OlIB
01:00:00
SRV Frictional Studies
fl/~r —~,&~.rfl
– .—
-~w~W ~
-~ .~
..
r,.,~
~
Block on Ring @ 200 Ibs, 1125 rpm and 1 Hr
140
130
120
110
0
Cl)
100
Cl)
Ctl))
90
80
70
liE 60
50
_______
—Co Initial Temp
_
-~
Co Final Temp
40
—
30
20
10
0
—______
Oil A
Oil B
Oil
c
Oil D
Different Oils
Figure 10
—
Block on Ring Test Gear Simulation
I~W~
—29 NLGI SPOKESMAN, MAY/JUNE 2014
z
Oil E
Temp Rise
V
NLGfl
intermediate values for the semi-synthetic oils as
compared to the mineral oils. This is indicative that
the Synthetic oils will experience lower viscous drag
and be energy efficient as compared to the other
oils in the EHD regime prevalent in the POD Zone.
will determine the end performance of the oil in a
gear contact.
(4) Oils D & F being both semi-synthetic and friction
modified when subjected to the FZG gear efficiency
test, where the combined effect of Viscous friction and
the Boundary friction is taken into account, show the
lowest energy consumption as compared to VG 320
oils. In case of the gear contacts, which works under
Boundary, mixed and FHD regimes of lubrication
(where the films are thin and nearing the surface
roughness of the contact surfaces), the friction modi
fication thus determines the overall efficiency of the
equipment. Due to this the Oils D & F, despite being
inferior to Synthetic oils B & C in terms of viscous
friction, exhibit higher energy efficiency, due to the
predominant boundary effect in the gear contacts.
(2) The Boundary Friction studies using the SRV and the
simulated gear contact using the block on ring test
machine following the same trends. Here the two fric
tion modified oils D & F have the lowest temperature
rise as well as the friction coefficient. The synthetic
oils B and C do not have a friction modification so the
friction coefficient with these oils are the highest. The
friction coefficient for the Mineral Oil VG 320 based
Oil A is intermediate to the Oils D & F and Oils B & C.
(3) From 1 & 2 above, it is expected that a combined
effect of the viscous friction and boundary friction
Block on Ring @ 200 Ibs, 1125 rpm and 1 Hr
0.09
0.08
0.07
~ 0.06
0
C)
~ 0.05
0.04
Figure 11
—
,
Different Oils
Block on Ring Friction coefficient comparison
~
~
~ ~/
~ ~
~
~V ~
~
~
Traction Coefficient @ 30N, 100 Deg, 30% SSR and 1 Hr Duration
8.OOE-02
Oil D
Oil E
7.OOE-02
~
~ 6.OOE-02
—~
OilA
Oil B
—
~
ollc
~; ::::::: ~~__
~
2.OOE-02
0
0.5
1
1.5
2
2.5
3
3.5
4
Disc Speed, rn/sec
Figure 12
~
—
Traction coefficient studies
~*flfl
cfl
a~3!a~.
—~ —~.
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VOLUME 78, NUMBER 2
fl fl~
4.5
a,,
6.0 Conclusions
NLGI
FZG % Reduction in energy consumed vis a vis
(1) Gear Lubricants work under
different regimes of lubrica
tion, and using a Tribotesting
methodology for the study
of these can enable us
to understand and work
towards energy efficiency
using lubricants.
Reduction in energy consumed vis a vis Oil A
(2) Simulated testing using the
-10
-9
-8
-7
-6
-5
-4
-3
-2
-1
0
SRV Test rig and the Block
Figure 13 Reduction in Energy Consumed in FZG Test vis a vis Oil A
on Ring test rig for the
Boundary regime efficiency
and the EHD film thickness and Traction studies for
5. Mechanical testing in the FZG gear test rig accord
the contribution of the viscous friction towards the
ing to DIN ISO 14635 (DIN 51354, method A /8,3 /
fluid film energy efficiency can offer good insight and
90 and A/i 6, 6 / 90)
enable development of lubricants.
6. http://www.optimol-instruments.de/content/
—
(3) The above combination of the Tribotesting tech
niques can enable development of gear lubricants
with high energy efficiency without the need for
expensive field trials.
produkte/srv%C2%AEtechnologieplattform/index.
php?lang=en SRV Test system.
.
7. http://www. pcs-instruments.com/ehl/ehl.shtml, EH L
Ultra Thin film measurement system.
(4) Maximization of energy efficiency is possible by
using a combined approach of viscosity optimiza
tion (to reduce the losses due to viscous drag) and
Friction modifiers (boundary friction effects).
8. http ://www,falex. com/pdf/FalexBlockRing. pdf Falex
Block on Ring test system
ABOUT THE AUTHOR
7.0 References
1. Greene, A.B. Risdon T.J., “The effect of molybdenum
containing oil soluble friction modifiers on engine
fuel economy and gear oil efficiency” SAE Paper No.
811187.
2. Douglass C. Porrett, Stephen D. Miles, Edward F.
Werderits and Donald L. Powell Development of
a laboratory axle efficiency test” SAE Paper No.
800804.
3. Facchiano, D L, Johnson R.L., “An examination of
synthetic and mineral based gear lubricants and
their effects on energy efficiency.” NLGI Spokesman,
1985, Vol. 48,11, pp. 399-403.
4. “The ECOTRONS test method for the assessment
of the ability of lubricants to reduce Friction losses in
Transmissions, Ill CEO Symposium 1989.
Ajay Kumar Harinarain Indian Oil
Corporation Ltd. Harinarain earned
his B.S. in Mechanical Engineering from
the M S University, Baroda India (1988),
Masters in Technology (Industrial Tribol
ogy) from the Indian Institute of Technol
ogy, Madras (Chennai) India (1990), and
Executive Masters in Business Adminis
tration (EMBA) from S P Jam Institute for
Management & Research (2008). He is
currently Senior Research Manager (Grease & Tribology) at Indian
Oil Corporation Ltd, R&D Centre, Faridabad India. He has over 23
years of experience in tribotesting of lubricants and fuels, and the
development of test methodologies for performance evaluation of
lubricants and greases. He has experience in field trials in steel!
cement/power plants etc. Ajay is ranked amongst the Deans Top
Five (2nd rank) during Executive Masters in Business Administration
(EMBA) from S P Jam Institute for Management & Research. He is
a Life Member of the Tribology Society of India, also Joint Secretary
of the Tribology Society of India, and is a member of STLE.
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