1Effect of Shaft Microcavity Patterns for Flow and Friction Control on Radial Lip Seal Performance...

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Effect of Shaft Microcavity Patterns for Flow andFrict ion Cont rol on Radial Lip Seal PerformanceAFeasibilit y StudyKat her ine H. Warren

a& Lyndon Scot t St ephens

a

aBearings and Seals Laborat ory, Universit y of Kent ucky, Lexington, Kent ucky, 40506, USA

Available online: 30 Oct 2009

To cite this art icle: Kat her ine H. Warren & Lyndon Scott St ephens (2009): Eff ect of Shaft Microcavi ty Patt erns for Flow andFrict ion Contr ol on Radial Lip Seal Perfor manceA Feasibil i t y St udy, Trib ology Transact ions, 52:6, 731-743

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Tribology Transactions, 52: 731-743, 2009

Copyright C Society of Tribologists and Lubrication Engineers

ISSN: 1040-2004 print / 1547-397X online

DOI: 10.1080/10402000903097361

Effect of Shaft Microcavity Patterns for Flow and Friction

Control on Radial Lip Seal PerformanceA FeasibilityStudy

KATHERINE H. WARREN and LYNDON SCOTT STEPHENS

Bearings and Seals Laboratory

University of Kentucky

Lexington, Kentucky 40506, USA

It has been shown that deterministic microfeatures on the

shaft of a radial lip seal impact seal behavior. This work seeks

to determine whether it is feasible to control lubricant pump-

ing direction and enhance pump rate with microcavities. The

effect of nickel film triangular cavity orientation on seal perfor-

mance, in particular the flow direction, the pumping rate, and

the friction torque, is investigated experimentally. Cavity shape,

area fraction, and depth are held constant while cavity orienta-

tion is varied. The oil drop test results are compared to those

for conventional seals; i.e., plain stainless steel shafts and shafts

with an electroplated nickel surface but no micro-cavities. It

was found that shafts with surface texture designs can control

the pumping direction and increase the sealing capability via

enhanced pump rates by up to eight times that of stainless steel

shafts. Preferential orientations pumped oil toward the widerend, or base, of the triangular cavities while patterns in neutral,

or nonpreferential, orientations were found to reverse pump.

The presence of microcavities reduced the friction torque by

as much as 51% when pumping and in all cases reduced the

operating temperatures. In some cases, the microcavities also

reduced the friction torque 813% when the seal was operating

in a starved condition.

KEY WORDS

Lip Seals; Hydrodynamic Lubrication; Surface Modification;

Deterministic Microfeatures; Reverse Pumping; Experimental

Results

INTRODUCTION

Radial lip seals are one of the most widely used type of dy-

namic seal and serve one of two purposesseal a fluid in or seal

Manuscript received August 21, 2008

Manuscript accepted May 1, 2009

Review led by Alan Lebeck

contaminants out. The basic components of a radial lip seal are

shown in Fig. 1 and include the elastomer (rubber lip), metal

casing, rotating or reciprocating shaft, garter spring, and sealed

fluid. The shaft outer diameter is slightly larger than the inner

diameter of the elastomer, which introduces an interference fit

between the two. This interference, together with the radial force

due to the garter spring and the beam effect of the elastomer,

maintains the fit of the elastomer to the shaft in light of variations

such as misalignment and changes in the system over time such as

wear.

Once in operation, the motion of the shaft begins to wear

the elastomer at its apex and a region known as the sealing zone

develops at the elastomer/shaft interface as shown in Fig. 2. The

profile of the elastomer in this region is shown flat in the sketch

of Fig. 2 for generality, but deformations will occur depending on

the specifics of a given seal. The formation of small irregularitieson the elastomer in this region (see inset (a) of Fig. 2) as it

wears has been credited for the ability of a radial lip seal to seal.

The irregularities, or microasperities, generate a hydrodynamic

pressure distribution that lifts the elastomer away from the shaft

such that a lubricating film develops and the asymmetric shear

deformation of the asperities in the circumferential direction

creates a small microscopic pump rate that redirects leakage

back into the sealed cavity resulting in the sealing phenomenon

known as reverse pumping (Hirano and Ishiwata (1); Jagger (2);

Kammuller (3); Kawahara and Hirabayashi (4); Muller (5)).

Without proper asperity formation, leakage in the direction

opposite of the desired seal may occur. Paige and Stephens (6)

characterized the microasperities on the elastomer of testedlip seals, noting their size, shape, and distribution. The tests

conducted in the course of that work were based on recommen-

dations of Horve (7), including the experimental design, the shaft

surface roughness, and the seal alignment parameters. Other

studies showed that the shaft surfaces of conventional lip seals

tend to an equilibrium roughness after completing a run-in time

and reaching a steady-state condition (Gawlinski (8)).

The use of deterministic microasperities on one surface of two

parallel plates was shown to increase the film thickness between

the plates and improve lubrication when one was in motion

relative to the other (Anno, et al. (9)). Researchers began to

731


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732 K. H. WARREN AND L. S. STEPHENS

Fig. 1Cross section of a typical lip seal; Chicago Rawhide 2000 (Paige

(16)).

consider the use of deterministic microasperities on lip seal shaft

surfaces (see inset (b) of Fig. 2) and how they might impact seal

performance in conjunction with, or instead of, nondeterminis-

tic asperities that develop on the elastomer. Otto investigated the

use of triangular asperities on radial lip seal shafts with the in-

tention of enhancing lip seal interface hydrodynamics, observing

that if the flow within the sealing zone could be controlled, then

seal leakage, wear, and temperature could be controlled (Otto

(10)). He found that appropriate asperity dimensions and pat-

terns enabled this control and showed improvement of both seal

efficiency andseal life. Otto was restricted to certainasperity sizes

due to the manufacturing capabilities available at the time and he

did not investigate the effects of cavities on the shaft surface.

A detailed theoretical analysis is beyond the scope of this

article, but the experimental work completed is built upon the

theoretical work of previous researchers. Much of the analysis

of Salant (11) and Salant and Flaherty (12) is considered in un-

derstanding what is occurring physically in the current testing.This includes the deformation of the elastomer lip due to the

pressure distribution at the interface as well as the possibility of

a starved operation due to an ingested meniscus. This ingested

meniscus condition may occur if a surface has a significantly high

reverse pumping rate such that little or no lubricant remains

in the interface. Siripuram and Stephens (13) conducted a nu-

merical study on the effects of deterministic asperity and cav-

ity shape, size, orientation, and distribution on the lubrication

performance of thrust surfaces and determined the sensitivity

of the friction coefficient and leakage rate to those parameters.

Hadinata and Stephens (14) presented a study with similar deter-

ministic asperity considerations as Siripuram and Stephens but

for radial lip seal application and developed an elastohydrody-

namic model for the seal with deterministic features on the shaft

surface. Impellizzeri (15) also used numerical models to inves-

tigate the effect of deterministic microfeatures on lip seal shafts,

including both asperities and cavities. The use of mass-conserving

Jakobsson-Floberg-Olsson (JFO) boundary conditions in these

works enabled the evaluation of the effect of the features not only

on the leakage rate of the system but also on the relationship be-

tween feature orientation and leakage. The work of Impellizzeri

(15) included a case study on equilateral triangular features that

predicts that triangular asperities will pump the lubricant toward

the apex of the triangle while the cavities will pump toward the

triangle base.

The current work presents experimental results for a nickel

film of oriented triangular cavities on the shaft surface of a ra-

dial lip seal. Previously tested radial lip seal patterns did not

Fig. 2Sealing zone.


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Effect of Shaft Microcavity Patterns 733

Fig. 3Sample manufacturing process. Adapted from Venkatesan (18).

include cavities, only triangular asperities, and did not consider

all the orientations tested here. Smaller dimensions are possi-

ble in this work as well due to a UV photolithography processdeveloped at the University of Kentucky Bearings and Seals

Laboratory. The feasibility of controlling the lubricant pumping

direction and enhancing the pump rate with triangular microcav-

ities is considered and observations on the effect of these cavi-

ties on seal performance are made. The results are compared to

those of conventional radial lip seals (bare stainless steel shafts)

and electroplated nickel shafts that have no deterministiccavities.

The performance parameters of interest include the pumping di-

rection of the cavities as well as the pumping rate, the friction

torque, and the temperature.

SAMPLE MANUFACTURE

Fabrication of the shaft surfaces to be tested is not the focus

of the current work, but a brief overview of the manufacturing

process is provided here for completeness. The ultraviolet (UV)

photolithography and nickel deposition process of Kortikar (17)

for flat surfaces was modified for application to cylindrical sur-

faces and is shown in Fig. 3. Stainless steel rings machined to

specifications recommended by Horve (7) serve as the shaft sub-

strate. The process starts with the preparation of this substrate

including anodic dissolution and cleaning, followed with the ap-

plication of photoresist (SU-8 10), pattern masking, and UV ex-

posure. After a post-bake process is complete, the photoresist is

developed. With the photoresist in place, nickel is electroplatedonto the steel substrate through a series of emersions including

C12, woods strike, and nickel sulfamate. Finally, the photoresist

is removed, leaving the shaft coated with nickel microfeatures as

shown in Fig. 4. The stainless steel and solid nickel surfaces tested

are shown in the figure as well as the orientations of the triangu-

lar shaft features. Those triangles oriented in the same direction

as shaft rotation are labeled cavities leading and defined to be

at a 0 rotation. Triangular cavities with apexes rotated +/90

are referenced as cavities to air side and cavities to oil side,

respectively. Those surfaces with triangular cavity apexes rotated

180 such that they point in the direction opposite of the shaft

rotation are labeled cavities lagging.

The selection of the equilateral triangle as the microfeature

to evaluate in this work is based upon the previous numerical

studies mentioned above (Siripuram and Stephens (13); Hadi-nata and Stephens (14); Impellizzeri (15)). As shown in these

works, preferential pumping requires an asymmetric geometry

and that geometric orientation has a significant influence on the

pumping rate. The manufactured triangular microcavities evalu-

ated in the present study are designed to be 5 m deep with a

base of 107 m and a height of 78 m. The center-to-center spac-

ing for the cavities-to-air and cavities-to-oil patterns is 150 m

in the circumferential direction and 114 m in the axial direc-

tion with the apexes staggered. For the cavities leading and the

cavities lagging patterns, the center-to-center spacing is 150 m

axially and 114 m circumferentially with apexes staggered. The

microcavity dimensions, depth, and spacing result from the evalu-

ation of previous work as well as consideration of manufacturing

capabilities.

EXPERIMENTAL SETUP

The experimental setup was developed by Paige (6) and con-

sists of a tribometer with lip seal assembly adaptors as shown in

Fig. 5. A ring with the desired test surface is installed on the ro-

tating adapter and the elastomer seal is installed at the top of the

stationary oil bath. The elastomer seals were acquired in groups

from the same production batches in an effort to reduce seal-to-

seal variability. All seals are made of nitrile rubber and have the

same geometric dimensions and tolerancing. The seal and shaftare aligned such that the seal tilt and the total eccentricity meet

the desired specifications (Horve (7)). As installed, the air side

or barrel angle of the seal is measured at approximately 20 and

the oil side angle is measured at approximately 59. These an-

gles are in the typical installed air side and oil side ranges of 20

35 and 4070, respectively, and satisfy the requirement that the

oil side angle be larger than the air side angle to achieve reverse

pumping (Horve (7)). The bath is filled with SAE grade 20W-

50 oil and the shaft is pressed into the elastomer. A graduated

cylinder connected to the oil bath by tubing is placed above the

elastomer/shaft interface, creating a pressure of approximately

5.38 kPa (0.78 psi). The tests are conducted at a rotational speed


5/14


Fig. 4Patterns tested and resulting shaft wear tracks.

of 750 rpm. These and other test details are summarized in

Table 1.

A data acquisition system is used to record the parameter

measurements of interest including speed, friction torque, and

temperature all with respect to time. A load cell is used to mea-

sure the friction torque in the system. The temperatures are mea-

sured using three type T thermocouples placed 120 apart on theseal housing, one in the graduated cylinder of oil, one in the am-

bient oil supply, and one on a surface away from the test for a

baseline room temperature measurement. The viscosity of the

oil changes with temperature and is therefore calculated from

the temperature measurements recorded during testing. This cal-

culation is achieved using the following viscosity equation from

Booser (19) where is the kinematic viscosity in centistokes and

Tis the temperature in degrees Kelvin.

loglog(v+ 0.7) = 8.7257 3.3565 log T [1]

The viscosity curve generated for the 20W-50 oil used in these

tests is shown in Fig.6. Therange of temperatures for steady-state

operation at 750 rpm is indicated in the figure as well. Over the

Fig. 5Test setup (Paige (16)).


6/14


TABLE 1TEST PARAMETERS

Seal material Nitrile rubber

Shaft material Stainless steel

Nickel electroplating

Shaft O.D. 139.7 mm (5.5 inches)

Shaft surface 0.250.50 m Ra

Lubricant SAE 20W-50Seal tilt (0.05) 0.026-0.047

Total eccentricity ( 0.0100 in) 0.0406-0.2413 mm

(0.0016-0.0095 in)

Seal pressure 5.38 kPa (0.78 psi)

Sump level Full

Test speed 750 rpm

Seal run time (@ 750 rpm) 660 h

Radial load 64.6-147.8 N/m (5.9-13.5 oz/in)

Viscosity (@ 40C) 170 mm2/s (cSt)

temperature variation seen in these tests, the viscosity changes

no more than approximately 20 mm

2

/s. However, the measuredtemperature values used in this calculation are those of the seal

housing close to, but not at, the elastomer/shaft interface. The

temperature and therefore the viscosity changes at the interface

are potentially higher as oil moves through and/or is pumped out

of this region.

Two types of tests are conductedone for seals with patterns

expected to reverse pump, i.e., seal, and one for seals with pat-

terns expected to forward pump. Forward pumping is defined as

an enhanced leakage that pumps the lubricant out of the oil bath

and through the interface to the environment side of the seal and

can be used to exclude contaminants. If the seal reverse pumps,

then a series of oil drop tests are performed as shown in Fig. 7.

Five of the six surface types tested in this study are evaluated us-ing oil drop tests where known amounts of room-temperature oil

(250, 500, 1000, 1500 L) are alternately injected on top of the

elastomer/shaft interface using a digital pipette and the time re-

quired for the seal to pump this oil through the interface into the

oil bath is recorded. The result is a reverse pumping rate. A spec-

ified amount of time is allowed to pass between the recovery of

the system from one oil drop and the injection of the next. If a

seal forward pumps, then the room-temperature oil is added to

the graduated cylinder connected to the oil bath and the time it

takes for that oil level to drop a specified amount is recorded, re-

sulting in a forward pumping rate. This test is conducted for the

only forward-pumping shaft in this study, triangular cavities ori-ented to the oil side of the seal. The amount of time required to

complete any set of tests varied from seal to seal depending on

the time required for the seal to come to steady-state tempera-

ture before the graduated oil injections began and the individ-

ual pumping rates of each seal. In this set of 18 shafts, 6 h was

the minimum amount of time required to complete the testing at

750 rpm and 60 h was the longest required time. The pumping

rates resulting from these tests are one parameter used to com-

pare shaft performance and, in the case of reverse pumping, seal-

ing capability.

RESULTS

Shaft Condition

Triangular cavities in four orientations were analyzed. The

test results for bare stainless steel and electroplated nickel shafts

with no deterministic cavities were also included in the analysis

for baseline comparison. Figure 4 shows a representative sample

of each orientation considered as previously discussedthe cav-

ity apex toward the environment or the air side of the seal, the

cavity apex toward the oil side of the seal, the cavity apex leading

the direction of rotation, and the cavity apex lagging the direc-

tion of rotation. Also seen in the figure is the final wear track

developed on each shaft during testing. For the cavity patterns,

the dark background outside of the wear track is the electro-

plated nickel with the lighter stainless steel substrate showing atthe bottom of each triangular cavity. The nickel in the wear track

region has been worn and polished by the elastomer. The final

wear track width after the completion of all testing was measured

and ranged from an average of 674 217 m on the cavities ori-

ented to oil shafts to over 1000 m on the electroplated nickel (no

Fig. 6Viscosity curve for 20W-50 oil indicating temperature range of operation.


7/14


Fig. 7Oil drop test; figure concept from Muller and Nru (20).

cavities) and cavities oriented to air shafts (1091 233 m and

1024 31 m, respectively). The neutral cavity patterns of tri-

angles leading and lagging had similar final wear track width av-

erages of 829 158 m and 811 124 m, respectively. The

shafts with the widest average final wear track width, electro-

plated nickel (no cavities) and cavities oriented to air; also had

the highest average starved friction torque. These results are not

surprising since operation in a starved or partially starved condi-

tion can result in contact between the lip and shaft and therefore

lead to greater wear.

An optical profilometer was used to analyze both the tested

and untested regions of the shaft surfaces as shown in Fig. 8. The

region outside of the wear track is the unworn or untested sur-

face, and the region within the wear track is the worn or tested

surface. The average surface roughness, Ra, of each of these ar-

eas was measured. For shafts with cavity patterns, the average

Ra unworn = 2.425 m with a standard deviation of 0.23 m and

Ra worn = 2.375 m with a standard deviation of 0.29 m. The

approximate 5 m depth of the cavities is reflected in these sur-

face roughness measurements and some surface wear is evident.

Flow

As expected, the baseline stainless steel shafts reversepumped. The shafts with electroplated nickel but no determinis-

tic cavities were also found to reverse pump. The cavity patterns

in preferential pumping orientations on the shaft surface

i.e., +/90controlled the direction of the lubricant flow by

pumping oil toward their base. Therefore, the cavities oriented

toward the air side of the seal reverse pumped, and cavities

oriented toward the oil side forward pumped. Figure 9 shows a

sketch of what is occurring at the interface for a reverse-pumping

pattern during an oil drop test injection. The asymmetric contact

pressure distribution of the elastomer is shown as well as the

circumferential lubricant flow due to shaft rotation (Qin) and

the flow due to reverse pumping (Qout). It has been theorized

by Hadinata and Stephens (14) and Impellizzeri (15) that the

Fig. 8Surface roughness.


8/14


Fig. 9Reverse pumping representation.

physical phenomenon responsible for the directional flow isthe asymmetric pressure distribution under the lip due to the

asymmetric geometry of the triangular microfeatures. These

previous works considered only a solution of the Reynolds

equation and the pressure-driven flow and did not consider the

possibility of deformation and/or bulging of the elastomer into

the cavities nor the possible presence of either rubber or nickel

debris as the seal wears. Otto (10), who investigated triangular

asperities, saw evidence in his testing that the elastomer did

deform around microfeatures and between circumferential rows

such that contact was made with the features.

The cavities in neutral orientations at 0 and 180 reverse

pumped in both orientations. The direction of the pumping for

these neutral orientation patterns is attributed to the nondeter-

ministic asperities on the elastomer and its asymmetric contact

pressure distribution. The corresponding pumping rates are

shown in Fig. 10 where reverse pumping is represented by a neg-

ative sign. These results are averages over multiple tests on threeshafts of each pattern and therefore error bars indicating the

standard deviation are also included. The standard deviation for

many of the results, including the friction torque and tempera-

ture values presented subsequently, is large. Though every effort

was made to replicate test specimens and conditions, each elas-

tomer/ring combination is unique and as such the unique wear

process for each combined with the inherent variability of the

elastomers results in the variations seen in testing. The standard

deviation of the pumping rate for the forward-pumping pattern,

triangular cavities oriented toward the oil side of the seal, was

notably less than that of the other deterministic patterns as seen

in Fig. 10. As the only pattern to forward pump, this pattern is the

only one that had a continuously lubricated interface. The other

patterns experienced periods of starvation due to reverse pump-

ing when little to no oil was present in the interface, most likely

resulting in additional wear and therefore greater variability.

-600

-500

-400

-300

-200

-100

0

100

200

PumpingRate(m

icroL/min)

Cav Oil

SS Ni Cav Air Cav Lead Cav Lag

Fig. 10Pumping rate.


9/14


TABLE 2PERFORMANCE COMPARISON TO BASELINE STAINLESS STEEL CASE

% Lubricated Friction Torque % Starved Friction Torque Factor of Pumping Rate Increase

Reduction (over Stainless Steel) Reduction (over Stainless Steel) (over Stainless Steel)

Nickel 18.59 5.33 (torque increase) 2.06

Triangular cavities to air 51.17 2.01 (torque increase) 8.13

Triangular cavities to oil 3.66 N/A 1.18

Triangular cavities leading 26.72 8.15 7.92

Triangular cavities lagging 19.45 13.79 8.37

Based on average values from multiple tests on 18 shafts.

The pumping rates for the four shaft surfaces with deter-

ministic cavities are higher than the rate for the conventional

stainless steel shaft and the rates of the three patterns that reverse

pumped (cavities to air, cavities leading, and cavities lagging)

are also higher than that of the nickel shaft. The deterministic

patterns increase reverse pumping by as much as eight times

that of stainless steel as summarized in Table 2. An increased

reverse pumping rate is considered in this context to indicate a

greater sealing capability of the seal. The stainless steel pump

rate is due only to elastomer asperities, which provide minimal

hydrodynamic lift and result in the lowest pumping rate of the

tested patterns. As predicted by Hadinata and Stephens (14), the

cavity patterns, when pumping, provide greater hydrodynamic

lift than the bare stainless steel, resulting in an increase in film

thickness, which generates a greater flow rate per the relationship

Qz = h3

12

P

zdx [2]

where Qz is the flow rate, h3 is the film thickness, is the dynamic

viscosity, Pis the pressure, andx and z are the circumferential and

axial directions, respectively. An increase in film thickness will

also reduce the shear stress on, and therefore deformation of, as-

perities found on the elastomer and for a conventional seal where

the elastomer is running on an unmodified stainless steel shaft

may result in a lower pumping rate. Here, however, the increase

in reverse pumping due to deterministic features dominates and

results in a reverse-pumping rate increase. The forward-pumping

pattern, though still exhibiting a greater pump rate than stainless

steel, does not show as significant an increase in the pump rate

as the other cavity patterns since the cavities in this case must

overcome the action of the reverse-pumping elastomer asperities

and the related asymmetric contact pressure distribution of the

elastomer.

Friction Torque

Figure 11 shows the first 5 h of a typical friction torque re-

sponse for each orientation compared with that of the baseline

stainless steel case. Each response includes an initial peak in

friction torque due to the start of system rotation. The reverse-

pumping patterns also exhibit a second friction torque peak that

occurs when a small reservoir of oil that remains on top of the

interface after installation pumps through the interface into the

sump and the seal first operates in a starved condition. The

friction torque response of each shaft type initially decreases

steeply, but this decrease becomes more gradual over time. Af-

ter the system reaches an equilibrium temperature, the gradual

response is attributed to the relaxing of the interference fit be-

tween the elastomer and shaft as the elastomer wears. Also seen

in the response of the reverse-pumping patterns are periodic

drops in friction torque. These drops correspond to the oil in-

jections carried out in the course of the previously described oil

drop tests used to measure the pumping rates.

A representative set of these oil drop tests for a baseline

stainless steel shaft and a shaft with cavities oriented to air is

shown in Fig. 12. Lower amplitude noise is seen when reverse

pumping occurs during these tests and also in the response of

the forward-pumping seals (as shown in the cavities-to-oil plot of

Fig. 11). The additional noise seen when not pumpingi.e.,

operating in a starved conditionis indicative of stick-slip fric-

tion occurring in the absence of full hydrodynamic lubrication.

An overshoot in the friction torque response can be seen at

the end of the oil drop tests when the seal initially returns to

operation with a starved interface and is likely due to the inertia

of the shaft assembly mass. This figure also shows the significant

reduction in friction torque for a reverse-pumping pattern when

pumping as compared to that of the conventional stainless steel

shaft surface. The width of each friction torque drop reflects the

amount of oil injected on the air side of the interface. For larger

injections of oil, the seal requires more time to pump the fluid

through the interface, resulting in the wider periods of reduced

torque. As seen in the figure and previously discussed in the

Experimental Setup section, four different volumes of oil were

injected throughout the course of testing. The enhanced pump

rate of the textured shaft is also apparent in Fig. 12.

The data of this work suggest that all of the reverse-pumping

seals tested, including those with unmodified shafts, may be op-

erating with ingested menisci when in a starved condition be-

tween the oil drop test injections. This is not surprising for seals

with modified shafts due to their significant pump rate increases.

Though conventional radial lip seals (no shaft modifications) gen-erally operate with a full lubricant film (Jagger (2)), the position

of the meniscus during operation may vary and mixed lubrica-

tion conditions have been reported in the literature (Jagger (21);

Salant (22); Horve (7)).

Figure 13 shows the average friction torque measured for

each type of shaft tested. For patterns that exhibit reverse pump-

ing, both the friction torque when operating with a lubricated

interface as well as that when operating in a starved condition are

shown. These lubricated and starved values are averages of the

friction torque while pumping when the oil is injected at the inter-

face during an oil drop test and while operating starved between

oil injections, respectively. The values included in averaging


10/14


0 50 100 150 200 250 3000

0.5

1

1.5

2

2.5

3

3.5

4

4.5

Time (minutes)

FrictionTorq

ue(N-m)

Cav Air

SS

0 50 100 150 200 250 3000

0.5

1

1.5

2

2.5

3

3.5

4

4.5

Time (minutes)

Ni

SS

0 50 100 150 200 250 3000

0.5

1

1.5

2

2.5

3

3.5

4

4.5

Time (minutes)

Stainless Steel and Cavities to Oil

Cav Oil

SS

0 50 100 150 200 250 3000

0.5

1

1.5

2

2.5

3

3.5

4

4.5

Time (minutes)

FrictionTorque(N-m)

Stainless Steel and C avities Leading

Cav Lead

SS

0 50 100 150 200 250 3000

0.5

1

1.5

2

2.5

3

3.5

4

4.5Stainless Steel and Cavities Lagging

Cav Lag

SS

Stainless Steel and Nickel (No Cavities) Stainless Steel and Cavities to Air

Time (Minutes) Time (Minutes)

Time (Minutes) Time (Minutes)

Time (Minutes)

Stainless Steel and Cavities to Oil Stainless Steel and Cavities Leading

Stainless Steel and Cavities Lagging

FrictionTorque(N-m)

FrictionTor

ue(N-m

FrictionTorque

(N-m)

FrictionTorue

(N-m

FrictionTorque(N

-m)

Cav Air

SS

Ni

SS

Cav Oil

SS

Cav Lead

SS

Cav Lag

SS

Fig. 11Friction torque response samples.

are taken from the data acquired once the system has reached

equilibrium; i.e., when a steady-state operating temperature is

achieved. The friction torque reduction results summarized in

Table 2 are based on these data. As shown in the figure, stainless

steel friction torque values exhibit minimal change when the lu-

bricant is introduced to the interface. The polished stainless steel

shaft surface has very few asperities, leaving only those devel-

oped on the elastomer to contribute to the pumping action when

the interface is lubricated and therefore causing little change in

the hydrodynamics of the seal. The nickel shafts without deter-

ministic cavities exhibit a slightly larger starved friction torque

value and a lower lubricated friction torque value for a torque

reduction when pumping larger than that seen in the case of plain

stainless steel (see Table 2). The presence of electroplated nickel

introduces additional nondeterministic microasperities not

present on the stainless steel shaft, resulting in better hydro-

dynamic effects; i.e., larger film thickness and better lubricant

pumping such that this greater difference between the lubricated

and starved friction torque is seen. The absence of deterministic

features, however, means that the direction and the rate of

flow are dependent upon the contact pressure distribution and

elastomer asperities.

The shafts with a triangular cavity pattern oriented to the at-

mosphere side of the seal result in a friction torque value very


11/14


2100 2120 2140 2160 2180 2200 2220 22400

0.5

1

1.5

2

2.5

3

235 240 245 250 255 260 265 270 2750

0.5

1

1.5

2

2.5

3

Lower amplitude noise wheninterface is lubricated.

Friction torque overshoot

Stainless Steel Cavities to Air

Time (minutes) Time (minutes)

FrictionTorque(N-

m)

FrictionTorque(N-

m)

Fig. 12Friction torque response during oil drop tests.

similar to that of the stainless steel and nickel shafts when oper-

ating with a starved interface. However, once the oil is introduced

to the interface, the cavities-to-air pattern shows the greatest

torque reduction of all the tested shafts, with values significantly

lower than those of either nickel or stainless steel. This reduction

demonstrates the superior hydrodynamics of the cavities over the

nondeterministic asperities of the stainless steel and nickel shafts

as well as the result of directing the enhanced flow.

Shafts with triangular cavities oriented to the oil side of the

seal forward pump, resulting in a continuously lubricated inter-

face such that operation in a starved condition does not occur.

The lubricated friction torque of these shafts is lower than that of

the stainless steel shafts but is the highest of all shafts with elec-

troplated nickel on the shaft surface. As discussed previously, this

pattern has a low pump rate, which is the result of a smaller film

thickness and therefore may operate without the benefit of full

hydrodynamic lift such that the friction torque is not drastically

reduced. The higher friction torque values may also be due to

these seals maintaining a higher radial lip force throughout test-

ing since the interference of the elastomer will not be reduced

to the same extent as the patterns subject to greater wear due to

operation in a starved condition.

The nonpreferential patterns, cavities leading and lagging,

have friction torque values when reverse pumping that are lower

Fig. 13Friction torque measurements.


12/14


SS

Ni

Cav Air

Cav Oil

Cav Lead

Cav Lag

70

75

80

85

90

95

Temperature(degC

)

Fig. 14Seal temperature during testing.

than that of stainless steel and nickel but not reduced as signif-

icantly as that of the shafts with cavities oriented to air. These

nonpreferential patterns do, however, demonstrate the lowest

friction torque values of the tested shafts when operating in a

starved condition. The lack of lubricant guidance by these neu-

tralpatterns may result in better lubricant retention in thecavities

even though high pump rates are seen due to the hydrodynamic

lift the cavities provide. As summarized in Table 2, the neutral

cavity patterns not only increase sealing capability by approxi-

mately eight times that of a bare stainless steel shaft, they also

reduce friction torque by 813% in the starved condition, which

will occur during normal operation when sealing. These results,

which increased sealing capability and reduced operational fric-

tion torque, are desirable in lip seal performance.

Temperature

Figure 14 shows the temperature of each seal during testing.

The reduction in temperature for patterns with deterministic cav-

ities as compared to the baseline stainless steel case ranged from

3 to approximately 9C. The life and performance of a nitrile rub-

ber seal is sensitive to temperature, with a 14C decrease having

the potential to double the life of a seal (Horve (7)). The reduc-

tions seen in these experiments may not be significant enough to

double seal life but could contribute to some life extension. Thelowest temperatures are seen for the cases of cavities leading and

lagging, which are the patterns that also had the lowest starved

operating torque and two of the highest pumping rates.

DISCUSSION

Conventional radial lip seal elastomers were used as manufac-

tured in the experiments of this work. These seal designs are not

optimized for the enhanced pump rates and lubricant flow con-

trol exhibited by the microcavities on the modified shaft surfaces.

Varying features of the seal such as the radial load and elastomer

contact angles would enable the seal to be designed to fully ben-

efit from the presence of microcavities on the shaft. It is probable

that seal performance parameters such as friction torque would

be found sensitive to these types of seal modifications.

The results indicate that seals with higher wear, signified by

larger final wear track widths, have lower friction torque values

when the sealing zone is lubricated. There is a twofold explana-

tion for this result. First, as the elastomer wears, the interference

fit between the seal and the shaft is reduced, resulting in a lower

load carried by the film and a corresponding larger film thickness

and lower shear stress on the film. Second, those seals with larger

wear track widths due to higher wear will have more rows of cavi-

ties engaged to carry the radial load. This again results in a larger

film thickness as the load per cavity is reduced. The seals with

larger final wear track widths also result in higher friction torque

when the sealing zone is starved. One possible explanation for

this refers again to the wider track engaging more cavities, but in

a starved condition this results in more contact in the interface

and an increase in friction torque.

The images in Fig. 4 and the plots of Fig. 11 and Fig. 12 are

representative of the samples tested in the course of this work.

They are examples that reflect the average trends as previously

discussed. When individual locations or tests are considered, a

great deal of variation is seen. Since an integral part of lip seal

operation is the wearing of the elastomer, this is not a surpris-

ing observation. The images of Fig. 4 show the final wear track atone circumferential location on one ring of each surface tested.

Taking a closer look at the individual variations reveals interest-

ing behavior in the seals that does not always correspond with

the performance results as might be expected. For the case of

triangular cavities oriented to the air side of the seal, all three

rings tested had similar final wear track widths with an average

width of 1024 m and a standard deviation of only 31 m. This

might imply that the seals wore similarly and would therefore

perform similarly. The lubricated friction torque of each cavities-

to-air seal supports this inference with very similar measured

values. One of these seals, however, had a significantly higher

reverse-pumping rate approximately twice that of the other two


13/14


cavities-to-air seals, with the highest reverse-pumping rate of all

seals tested. This resulted in a higher starved friction torque

and higher operating temperature for this seal. In contrast, the

cavities-to-oil seals had an average final wear track width of

674 m and a standard deviation of over 200 m. This large de-

viation was due to one cavity-to-oil seal having a final wear track

width significantly larger than the other two cavities-to-oil seals.This seal, however, performed similarly to one of the seals with

a narrower track while the other narrow tracked seal exhibited a

greatly reduced pumping rate and lower friction torque and op-

erating temperature. As previously stated, the neutrally oriented

cavity patterns of leading and lagging triangles had similar aver-

age wear track widths of 829 and 811 m, respectively, with stan-

dard deviations of 158 and 124 m. Individually, however, one

seal of each set of three resulted in a shaft wear scar noticeably

wider or narrower than the other two as reflected by the large

standard deviations of the measurements. This variation in final

wear track width seemed to correspond with a higher and lower

pumping rate, respectively, but had no appreciable impact on the

operating temperature of the seal and resulted in no obvious cor-

relation with variations in friction torque.

Much like the variations seen in final wear track width when

individual seals are considered, variations in seal response are

seen when looking at individual sets of test data as shown in

Fig. 11 and Fig. 12. As the wear track images of Fig. 4 were

single pictures representing sets of surfaces, the data plotted

in these figures is a representative set selected from multiple

available sets for each of the eighteen rings tested. Certain

trends are consistent from ring to ring such as the previously

discussed initial torque decrease and the additional noise seen

in the friction torque response when the seal is operating in a

starved condition. From test to test, however, variations are seen

in the details of the response of each seal as the lubricant moves

through the interface. While reverse pumping during an oil drop

test, the friction torque response of some seals exhibits a positive

slope or increase, some display a slight decrease, and others

have a level friction torque response with a slight convexity.

These variations are dependent on how each shaft surface and

microcavities (if applicable) work with the developed elastomeric

asperities to move the oil in, out, and through the interface. As

the lubricant moves through the interface, the friction torque will

decrease and as a result the operating temperature of the seal

will also decrease. As the temperature decreases, the viscosity

of the oil will increase and tend to increase the friction torque.

This continuous interplay between the friction torque, the tem-perature, and the viscosity is dynamic and at any given instant in

time, in any particular interface location, the magnitude of each

increase and decrease will vary with those specific conditions.

These local variations combine with other influences to result

in the response of the system as measured. One such additional

influence impacting the performance of each ring/test is seal

hydrodynamics. Over the course of testing, the lip seals will

operate in different lubrication regimes. The polished regions

of nickel in the wear track scars indicate that there are periods

during seal operation when the elastomer is in contact with the

shaft surface. This would be expected during startup when the

seal may be operating in the boundary lubrication regime as well

as during operation in a starved condition when the seal would

be expected to be operating in a mixed lubrication regime. It is

assumed that the seals will operate in full hydrodynamic lubri-

cation when pumping, but this may not always be true. When a

seal is operating with a forward-pumping pattern, the interface is

always lubricated since the seal is continuously pumping oil from

the sump to the air side of the seal. However, polished bandsof nickel can be seen in the wear track of the forward-pumping

shaft surfaces. This indicates that there may be periods of mixed

lubrication even when the interface is apparently fully lubricated.

Despite these variations, some of which are significant, the

overall operation of each seal is controlled by the cavity pattern

on the shaft. This demonstrates the ability of surface textures to

dominate both the effects of the naturally forming elastomer as-

perities and the inherent uncertainty of radial lip seal systems.

These results confirm the feasibility of using deterministic micro-

cavities to control and enhance radial lip seal performance.

CONCLUSIONS AND FUTURE WORK

This work demonstrates the feasibility of triangular microcav-

ities on a radial lip seal shaft surface to impact various seal per-

formance parameters including lubricant flow direction, pumping

rate, and friction torque. The following conclusions can be drawn.

1. The shaft surface texture can be designed to dominate the

elastomer pumping of a lip seal and control the direction of the

seal leakage. In particular, triangular microcavities oriented

in preferential patterns pump the lubricant toward the cavity

base; i.e., a triangular pattern oriented with apexes toward the

air side of a seal will reverse pump and those oriented to the

oil side will forward pump.

2. The experimental results of the previous conclusion werepredicted by Impellizzeri (15) using numerical models. This

verifies the utility of those models for surface design and

performance trend prediction.

3. The lubricant pumping rate and therefore the sealing capabil-

ity of the seal are enhanced by the presence of microcavities.

The rates are increased over that of the baseline stainless

steel case as well as that of the plain electroplated nickel case

when reverse pumping.

4. Microcavities reduce the friction torque of the system when

pumping lubricant and in some cases when operating in a

starved condition.

5. A reduction in temperature is seen when operating with

microcavities on the shaft.6. Trends between shaft wear and friction torque were seen

for seals operating with shaft microcavities. Higher starved

friction torque and lower lubricated friction torque is seen

with increased wear (larger final wear track widths).

A demonstration of the control of lubrication parameters via

the use of deterministic microcavities on the shaft surface of a

radial lip seal gives rise to further study considerations such as life

tests, seal design, and different applications. With the lower fric-

tion torque values and reduced operating temperatures seen with

some of the patterns tested, longer seal life would be expected.

However, long-term tests up to and possibly exceeding 1000 h


14/14


on seals operating with microcavities are necessary to determine

the overall impact of the features on seal life. Any benefit gained

from the use of microcavities would have to be weighed against

the detriment of a shorter seal life if the assumed life extension

does not occur. This work suggests that current lip seal design

may not be optimized for the enhanced pump rates and lubricant

flow control exhibited by the microcavities. Therefore, a redesignof the basic lip seal may be needed to fully benefit from the

inclusion of microfeatures on the shaft surface. Design consider-

ations could include, but are not limited to, changes in the garter

spring force and/or the interference fit as well as adjustments to

the barrel and oil side angles of the elastomer. An evaluation of

specific needs and operating conditions in various applications

is needed since it could present opportunities for utilizing seals

with microfeatures on the shaft surface given current lip seal

designs. Installation of a seal that leaks, or forward pumps, could

be advantageous in applications where contaminant exclusion

is of great importance, such as when protecting grease-filled

bearings. The extremely high sealing capability of some of the

cavity patterns may be beneficial when sealing high-pressure

systems. Lastly, an exploration of the use of microcavities in

other suitable rotary or reciprocating seals beyond that of the

elastomeric radial lip seal should be considered.

ACKNOWLEDGEMENTS

The authors wish to thank the Army Research Office (grant

number DAAD 19-02-1-0198), the Timken Company, the U.S.

Department of Education, and Valvoline for their support of and

contributions to the project that made this work possible.

REFERENCES

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Shaft Seals, Proc. Conf. on Lubrication and Wear, Inst. of Mech. Eng., pp

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1Effect of Shaft Microcavity Patterns for Flow and Friction Control on Radial Lip Seal Performance...

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