Wear of Elastomer Seals
Analysis of Seal Profile
Devin Murray
December 9, 2013
MANE 6960
Friction, Wear and Lubrication of Materials
Professor Ernesto Gutierrez-Miravete
Table of Contents
Abstract ......................................................................................................................................................... 2
Background ................................................................................................................................................... 2
Problem Summary ........................................................................................................................................ 2
Seal Profiles .................................................................................................................................................. 3
Theory ........................................................................................................................................................... 4
Methodology ................................................................................................................................................. 5
Material Properties ........................................................................................................................................ 7
Analysis Results ............................................................................................................................................ 7
Conclusion .................................................................................................................................................... 9
Direct References ........................................................................................................................................ 10
Research References ................................................................................................................................... 10
List of Figures
Figure 1: Standard O-Ring Seal Profile ........................................................................................................ 3
Figure 2: Quad Ring Seal Profile .................................................................................................................. 3
Figure 3: ‘Quad-O-Dyn’ Seal Profile............................................................................................................ 4
Figure 4: Standard O-Ring Seal Profile FE Model ....................................................................................... 5
Figure 5: Quad Ring Seal Profile FE Model ................................................................................................. 6
Figure 6: Quad-O-Dyn Seal Profile FE Model ............................................................................................. 6
Figure 7: Standard O-Ring Seal Profile FE Model Results .......................................................................... 7
Figure 8: Quad Ring Seal Profile FE Model Results .................................................................................... 8
Figure 9: Quad-O-Dyn Seal Profile FE Model Result .................................................................................. 8
List of Tables
Table 1: FE Model Material Properties......................................................................................................... 7
Table 2: FE Model Compression Results ..................................................................................................... 9
Table 3: Seal Profile Wear Resistance ........................................................................................................ 10
1
Abstract
This paper investigates elastomeric dynamic seals in fluid power systems in an effort to further
understanding the effect of seal profile on performance in high cycle applications. Standard torus shaped
o-rings are commonly used as dynamic seals in industry on reciprocating pistons inside cylinders due to
their standard and robust nature. However, seal manufacturers are constantly developing and testing new
seal profile geometries to improve seal life by reducing wear that occurs during cycling. Test data and
manufacturer recommendations support elastomer seals with engineered profiles having superior
performance when compared to standard o-rings in high cycle endurance tests. This paper evaluates the
standard o-ring seal profile against two non-standard engineered seal profiles to discover the attributes
which reduce abrasive wear during cycling. A recommendation for further study of engineered seals is
also made in the conclusion of the paper.
Background
Seals in fluid power systems are broken into two distinct categories; dynamic and static. Static seals are
non-moving seals typically used to prevent leakage at system boundaries. Dynamic seals are used in
moving components, such as cylinder pistons, motors shafts, pumps shafts and valve slides and can move
either in linear or rotational directions. Dynamic seals can be mounted in two orientations, the first where
the seal moves with part of the body (as in piston seals moving in a cylinder) or secondly where the seals
remains stationary as the body moves or rotates past it (motor shaft seals).
Typically, static seals are simple o-rings of various materials designed to fit in their gland with a high seal
compression to provide a zero leakage boundary. When choosing static seals, engineers are generally not
concerned with wear characteristics of the seal due to lack of motion and will make design choices based
on material compatibilities with system fluid, system temperature, system pressure and parts
commonality. Consequently, when making design choices for dynamic seals, an engineer must consider
all of the same attributes that govern selection of static seals as well as taking into consideration the type
of dynamic motion, differential pressure across the seal, number of expected cycles, seal compression and
acceptable cross-seal leakage rates. Success or failure of dynamic seals hinges on their ability to provide a
fluid pressure boundary while allowing dynamic motion of component parts. Poorly selected seals or
poorly designed hardware can quickly lead to severe wear in dynamic seals which results in unacceptable
leakage across the seal negatively affecting component performance.
Problem Summary
For fluid power systems in military, industrial, and commercial applications a major driver of both cost
and system performance is corrective maintenance. For dynamic seals in fluid power system, seal wear
and failure is a main contributor to required component maintenance. Therefore, it is extremely beneficial
in all dynamic sealing applications to attempt to minimize seal wear in an attempt to maximize the
operational time of components and minimize maintenance labor and material costs.
Dynamic seal wear is considered an abrasive wear process due to the fact seals are generally made of
elastomer material which is much softer material than the metallic surfaces they travel across. Factors
contributing to wear include surface finish of the metallic cylinder and piston gland, lubrication, speed of
travel, compression, seal material, and seal profile.
The project will isolate and address seal profile from the other variables contributing to dynamic seal
wear. Different seal profiles will be studied in a theatrical piston and cylinder sealing application in an
effort to further understanding of why some engineered seal profiles exhibit superior performance in high
cycle dynamic sealing applications while others fail at much higher rates.
2
Seal Profiles
Three seal profiles were selected for analysis and constitute different selections from the Reference (a)
Minnesota Rubber Company product line. These seals were selected due to each being available in a
homogeneous NBR material compositions and the fact that all seals are designed to fit in standard
dynamic seal glands therefore having identical seal compression.
Figure 1: Standard O-Ring Seal Profile
The Figure 1 o-ring seal profile is easily the most universal and widest used fluid power system seal for
both dynamic and static applications. Advantages of using a standard o-ring seal in any given application
are the wide array of available material choices, industry standard gland and seal sizes, low cost and
general availability in industry. Depending on seal compression, and the distance, type and speed of
travel, standard o-rings are known to perform satisfactorily in dynamic sealing applications. A common
failure mode of standard o-rings in a dynamic application is extrusion of the seal into the gap between the
piston and cylinder which can quickly lead to substantial seal damage and failure. As much of the
potential for seal extrusion can be prevented by proper seal gland design and the use of back-up rings, this
mode of failure is not considered in the evaluation performed by this paper as it is not pertinent to long
term seal wear.
Figure 2: Quad Ring Seal Profile
3
The Quad Ring seal design in Figure 2 is proprietary to the Minnesota Rubber Corporation and is
recommended as an alternative to standard o-rings in difficult high pressure and high cycle dynamic
sealing applications. The four ‘lobes’ of the seal allow it to be used in a wider temperature and geometric
tolerance range than a standard o-ring and provide two points of contact for sealing.
Figure 3: ‘Quad-O-Dyn’ Seal Profile
The Quad-O-Dyn seal design in Figure 3 is also a proprietary product of the Minnesota Rubber
Corporation and is recommended for the most difficult dynamic sealing applications. This seal has six
‘lobes’, two in the center which provide the primary sealing surface and four secondary lobes which
provide back-up sealing surfaces depending on direction of differential pressure.
Theory
Of the three major type of wear; adhesive, abrasive and 3rd body, only abrasive and 3rd body wear
generally pertain to elastomeric seals in a real world environment. For the purposes of this evaluation, 3 rd
body wear will be disregarded. This is because 3rd body wear introduces outside particles to the seal and
cylinder system, which can cause not only wear, but significant gouging and tearing of seals which can
result in failure. As 3rd body wear would only confound differences between sealing profiles, this
mechanism of seal wear is disregarded.
This leaves abrasive wear as the only mechanism by which the different seal profiles will be evaluated
against one another. As a given elastomeric seal travels over a metal cylinder surface, abrasive wear
occurs through cutting, ploughing or fragmentation of the softer seal material. Abrasive wear can be
measured by the amount of mass removed from the wear body, which in seal applications results in
decreased compression. Seal compression itself comes from the original thickness of the unreformed seal
being squeezed between two metallic surfaces, creating a decreased seal thickness and a fluid pressure
boundary due to the seal material attempting to expand filling flow potential flow paths. Therefore, as
more wear occurs in a seal, less seal compression exists due to removal of material and the greater the
chance for leakage across a seal. The amount of material removed through abrasive wear depends upon
the contact area of the compressed seal against the cylinder as well as the overall contact force and
friction. As all seals evaluated in this paper are assumed to be in the same fluid system, constructed of the
same material and assumed to be moving at identical speed on identical cylinder surfaces, friction can be
taken out of the evaluation of the seal profiles. We are left with contact area and contact force (which can
be evaluated as contact stress) as mechanisms to evaluate the different seal profiles.
4
It should be noted that the surface finish of the cylinder coupled with the system fluid have large impacts
on abrasive wear due to their influence on seal friction. For example, a given dynamic seal in a hydraulic
fluid system is expected to have a much longer seal life than the same seal in an air system (unlubricated)
due to the inherent lubrication of the seal by hydraulic fluid trapped between surface asperities in the
cylinder. Consequently, surface finish of cylinders can be both too rough and too smooth, with both
having detrimental effects on seal life. When a cylinder is too rough, extreme abrasive wear mechanisms
take over and can quickly ‘chew up’ a dynamic seal causing failure. Conversely, if a cylinder surface
finish is too smooth, not enough system fluid (or applied lubricant) is trapped in the valleys of the
cylinder surface profile which leads to high friction and subsequent heat generation and seal wear.
Reference (b) recommends 10-20 micro-inch surface finish for dynamic o-ring applications.
Methodology
To form a more complete understanding of the effect of seal profile in seal wear, representative finite
element (FE) models of the three seal profiles were constructed in COMSOL Multiphysics Ver. 4.2a. Half
of each seal profile was modeled and dimensions were constrained to represent a seal that is 0.5 in thick
and 0.5 high in all models. As exact dimensions for the Quad Ring and Quad-O-Dyn seal profiles are
unknown, the FE models are considered representative models used only for proof of concept.
O-Ring Seal
Cylinder
Figure 4: Standard O-Ring Seal Profile FE Model
5
Quad Ring
Seal
Cylinder
Figure 5: Quad Ring Seal Profile FE Model
Quad-O-Dyn
Seal
Cylinder
Figure 6: Quad-O-Dyn Seal Profile FE Model
6
Material Properties
Seal material was assumed to be nitrile butadiene rubber (NBR) and material properties were derived
from Reference (c). Cylinder material was assumed to be steel and properties were taken from previous
FE models developed for MANE: 6960 Friction, Wear, and Lubrication.
Table 1: FE Model Material Properties
Property
Seal (NBR) Cylinder (Steel)
Young’s Modulus (E)
5x106 Pa
2x1011 Pa
Poisson Ratio (ν)
0.4
0.3
Density (ρ)
1000 kg/m3
7850 kg/m3
Analysis Results
Seal compression was applied to each FE seal profile model to gauge contact area and contact force.
Reference (b) recommends 10 to 15% seal compression for dynamic sealing applications in fluid power
systems. Considering each seal is modeled based on a 0.5 inch thickness seal, recommended compression
is:
0.50 𝑖𝑛𝑐ℎ𝑒𝑠 × 10% = 0.050 𝑖𝑛𝑐ℎ𝑒𝑠
-to0.50 𝑖𝑛𝑐ℎ𝑒𝑠 × 15% = 0.075 𝑖𝑛𝑐ℎ𝑒𝑠
The 10% compression level was initially selected for analysis; however significant model error occurred
when compression was applied due to the large deflections in the NBR seal material. As a consequence,
seal compression was reduced to 1% (0.005 inches). As results are interpreted comparatively between seal
profiles, this reduction in seal compression has no impact other than to reduce the total magnitude of
compression stress and was found to be an acceptable measure to allow analysis to continue. Model
results are shown below:
Figure 7: Standard O-Ring Seal Profile FE Model Results
7
Figure 8: Quad Ring Seal Profile FE Model Results
Figure 9: Quad-O-Dyn Seal Profile FE Model Result
8
Table 2: FE Model Compression Results
Seal Profile
Peak Contact Stress
Standard O-Ring
1.86 x 105 Pa
Quad Ring
3.78 x105 Pa
Quad-O-Dyn
2.95 x 105 Pa
As can be seen in Table 2, the different seal profiles result in substantially different contact stresses when
each seal is compressed an equal amount (0.005 inches). Contact stresses are intuitively determined by
the force required to compress the seal (equating to amount of seal material deformed) and the surface
area over which the compressing force is applied (equating to the deformed seal area in contact with the
cylinder).
Conclusion
The results of the seal profile analysis provide crucial insight into how each seal is affected by wear
during dynamic travel. Intuitive observation applied to analysis results and the geometry of the seal
profile will be used to qualitatively rank the seals from best to worst in their resistance to wear.
Standard O-Ring
The standard o-ring profile has the lowest overall peak contact stress in large part due to the
contact area of the deformed seal and the cylinder being the greatest of the three seals evaluated.
This results in a seal which is the weakest initially in overall sealing strength (i.e. the standard oring is the least capable of the three seals in withstanding differential pressure without leakage).
The large contact area of the seal with the cylinder also presents the worst prospect with regard to
abrasive wear during dynamic travel. As the standard o-ring wears and material is removed, a
significant decrease in seal compression can be expected which ultimately leads to failure of the
o-ring to contain differential pressure and leakage rates will reach unacceptable levels.
Quad Ring Seal
The Quad Ring seal has the highest overall peak contact stress which is due to this seal profile
having the smallest contact area with the cylinder of all profiles evaluated. The high contact
stress results in a very strong seal strength; however this will also result in increased wear of the
sealing ‘lobes’ when compared with the Quad-O-Dyn seal. With all facts considered, the Quad
Ring should perform better than the standard o-ring due to the small contact area of the seal
exposed to abrasive wear and the geometry of the seal being such that wear at the two ‘lobes’
should not greatly decrease the seal compression. This is because the majority of the compressed
seal material resides in the ‘main body’ of the seal and is unaffected by wear. Conversely, the
Quad Ring is likely to fail sooner due to wear than the Quad-O-Dyn seal due to the comparatively
higher contact stress.
Quad-O-Dyn Seal
The peak contact stress of the Quad-O-Dyn seal profile falls roughly in the middle of the standard
o-ring and Quad Ring stresses. As the general seal geometry of the Quad-O-Dyn seal profile is
very similar to the Quad Ring, it is apparent that the same positive aspects of the engineered
profile are maintained (strong seal strength, low contact area exposed to abrasive wear, and
maintaining seal compression despite wear at seal ‘lobes’). Therefore, due to the decreased
contact stress when compared to the Quad Ring, the Quad-O-Dyn seal is expected to wear at a
decreased rate resulting in a greater seal life prior to failure.
9
Table 3: Seal Profile Wear Resistance
Seal Profile
Resistance to Wear
Standard O-Ring
Worst
Quad Ring
Moderate
Quad-O-Dyn
Best
The interpretation of the analysis results support manufacturer claims as well as independent test data
conducted by industry users of the seals. Quad-O-Dyn seals have been found to exhibit very long
operational lives in comparison to other seal profile designs in very challenging dynamic applications.
As can be seen in the comparison of the three seals evaluated in this paper, seals which have several
smaller ‘lobes’ as sealing surfaces have both increased sealing strength and expected seal life. Further
study should be undertaken to determine if adding additional seal lobes (on the order of four or five
‘lobes’) further increases seal life at an acceptable depreciation of sealing strength. A point of diminishing
returns will undoubtedly be reached in terms of the number of ‘lobes’ that can be added before the seal
performs no better or potentially worse than the Quad-O-Dyn seal. Additionally, manufacturing practices
must be evaluated to determine if current molding technology will support further refinement of the
engineered seal profile.
Direct References
(a) Minnesota Rubber and Plastics © 2013; www.mnrubber.com; date visited December 8, 2013
(b) Parker O-Ring Handbook, ORD 5700, Copyright © 2007 Parker Hannafin Corporation,
Cleveland Ohio
(c) Designer Data, Nitrile Butadiene Rubber © 2013; www.designerdata.nl; date visited December 1,
2013
Research References
(a) Friction and wear of elastomer seals; M. Gawlinski, Wroclaw University of Technology;
Archives of Civil and Mechanical Engineering Vol. VII, No. 4, 2007
(b) Friction and wear of elastomers and sliding seals, Booklet of theses; Nandor Bekesi, Budapest
University of Technology and Economics, 2011
(c) Tribology of Sealing elastomers; Braham Prakash, Division of Machine Elements Lulea
University of Technology, 2012
(d) Wear of Polymers – Special Issue Paper 401; BJ Briscoe - Department of Chemical Engineering
and Chemical Technology, Imperial College of Science, Technology and Medicine, London, UK;
and SK Sinha - Department Mechanical Engineering, National University of Singapore,
Singapore, 2002
10