MANE-6960: Friction, Wear, and Lubrication of Materials
Wear of a Gas
Turbine Friction
Damper
Research Project
Bill Beckman
12/9/2013
Introduction
Many engineering applications either rely on or have a by-product of vibrational
damping due to contact between two or more surfaces. Frictional damping is an
effective way to reduce the relative motion between components by dissipating
vibrational kinetic energy as heat. The focus of this paper is on dry friction dampers to
combat blade vibration in turbomachinery, specifically in the turbine section of axial flow
gas turbine applications.
During engine operation, turbine components are sensitive to rotational speeds,
component thermal environment, and gaspath aerodynamic loading. A combination of
static stresses along with dynamic stresses due to high cycle fatigue loading is a
primary concern for turbine components. Excitation can lead to rapid crack growth in
highly stressed locations and ultimately failure of the part. Due to constraints on design
space and upstream structural geometry, airfoil components often must operate close to
natural frequency excitation. The use of friction dampers can mitigate excitation
amplitudes significantly if component mode crossings cannot be designed out of the
engine operating range.
The specific configuration investigated in this paper is known as a cottage roof
damper. This design has advantages over dampers applied to airfoil sections, as a
separate cast part is inserted underneath the blade hub during assembly. This allows
for ease of worn part replacement, can be used with solid airfoils as opposed to hollow
applications, and minimizes system imbalance and loads due to lower rotational
distance from the engine centerline. While effective component damping can be
attained, contact loads due to rotational speeds and heat generation eventually results
in wear of one or both components. Due to high costs of turbine airfoil parts, the
preferred wearing surface is typically that of the damper itself. Improving wear
properties of the damper must take into consideration possible impacts to the blade
wear rate itself. Merely increasing hardness of the damper runs the risk of wearing the
blade part.
Theory and Methodology
Figure 1 is a schematic view of a typical cottage roof under-platform friction
damper and turbine blade. The damper is assembled underneath the airfoil hub, and
between adjacent blade members.
Figure 1: Schematic View of Friction Damper and Turbine Blade
The damper is free to move within the pocket; however during operation
centrifugal loads radially seat the damper against adjacent blade platforms. The relative
motion between damper and blades results in two directions of loads to be concerned
with: loads normal to the contact surface, and tangential sliding loads. Stiffness, friction,
stick vs slip conditions, and damper rolling are all inputs that must be taken into
consideration for damper effectiveness and resulting harmonic response. Because of
the complex nature of the system, optimizing a damper is typically done for one critical
mode or resonance. Energy dissipation is often graphically represented and studied by
means of hysteresis loops.
Figure 2: Hysteresis Loop of a Dry Friction Damper
Taken from www3.imperial.ac.uk/medynamics/joints/jointsproject
Figure 2 shows one cycle of loading and unloading for a dry friction damper. The
steep vertical slopes are understood to represent pre-sliding motion of the elastic
deformation at the area of contact. Stiffer components result in steeper curves of elastic
non-sliding loading. The system quickly transitions into a micro-slip condition and finally
to a macro slip where most of the relative motion between the components is seen.
Rapid progression between stick, microslip, and macroslip regimes along with varying
system operational boundary conditions makes frequency response and wear difficult to
predict.
This oscillating macroslip travel results in minor fretting and adhesive wear
development on the interfacing surface of the damper. On the microscopic level, high
asperity contact loads and pressures can plastically deform and ultimately oxidize or
dislodge particles away from the rubbing surfaces. Since movement in this system is
confined in a tight space, oxidation and material transfer can aid in wear resistance due
to local hardening. High temperature of the turbine environment as well as frictional heat
generation causes softening of the damper material which then transfers some particles
to the counter surface of the blade. Continuous rubbing between the fresh damper
surface and the transferred layer on the blade can cause a microcomposite hard layer
on both the blade and damper. As long as loads remain relatively low, which is a
reasonable assumption as the mass of the damper is small, this hardened
microcomposite layer can increase damper life.
The interfacing alloy materials are similar in make-up and standard surface
finishes have roughness on the order of 1 to 4 micrometer, it’s safe to assume the wear
regime is primarily one of an adhesive sliding wear. To focus purely on wear rates, the
macroslip tangential motion of the damper is investigated further.
Results and Discussion
A reasonable life expectation for a commercial application is on the order of
20000 hours of engine operation before inspection and overhaul. Considering this,
Table 1 shows the assumed boundary conditions used for the analysis.
ω
v
ρ
m
r
α
rotational speed [rpm]
initial damper volume [cm^3]
damper density [g/cm^3]
damper mass [g]
radius of rotation [m]
damper to counter surface contact angle [°]
10000
0.78
8.35 to 8.85
6.54 to 6.94
0.25
30
Table 1: Boundary Conditions
Damper density, and therefore mass, have been applied for varying materials. In
order to increase hardness, density typically increases and as a result weight of the
component increases. In industry numerous high temperature super alloys have been
developed to accommodate gas turbine requirements of different temperature, wear
resistance, corrosion resistance, and ductility. The total radial load of the damper due
to centripetal forces and the resulting normal force between the contacting surfaces
were found to be:
𝐹𝑟 = 𝑚 ∗ 𝑟 ∗ 𝜔2 ,
range between 1794 N and 1902 N
𝐹𝑛 = 𝐹𝑟 ∗ cos(𝛼),
range between 1554 N and 1647 N
The blade vibration was assumed to be 1000 Hz and the relative tangential
movement of the contacting surfaces was set to 25 micrometers per cycle for a pure
macroslip sliding condition. For the purpose of this paper, relative movement
oscillations of the damper and counter surface were assumed to follow the dynamic
cyclic loading of the blade itself. The resulting movement over the lifetime of the
wearing component was determined to be 1800 meters of sliding distance, L. These
assumptions may be considered conservative as the entire duration of engine operation
wouldn’t remain at a constant speed of 10000 rpms or dwell at a components natural
frequency causing continuous damper sliding and rapid wear. Designers target
excitation modes to cross engine operating ranges only during transient speeds as it’s
not ideal to dwell at critical component frequencies. These critical speeds would exist
somewhere between steady state ground idle, and cruise or takeoff speeds. For
adhesive wear conditions the volumetric wear rate is defined by Archard’s law:
Vw = K ∗
Fn ∗ L
H
𝐾=
𝐴𝑤 𝑛
=
𝐴𝑟 𝑁
The typical range of wear coefficient K is between 10-8 for lubricated inconsistent
materials to 10-3 for dry sliding common materials. These values and the equation
above support the idea that the actual wearing area, Aw, is a function of asperity contact
and fracture. Aw is also considerably smaller than the real contact area, Ar.
Considering the material combination and surface finish of the damper wearing surface
and the counter surface of the blade platform, a reasonable assumption for the wear
coefficient is taken to be 10-4. This value reflects the system of cast surface finishes,
closely related metal alloy materials, and non-lubricated dry sliding contact.
Table 2 shows the range of hardness values used for high wear resistant Nickel
and Cobalt-Chromium alloys. Taking into consideration a high temperature
environment (scaled to roughly 600°C), six types of bulk material were investigated
along with a varying hardness by means of a carburized material. Surface hardness
modification by means of gas carburization can be seen in Figure 3 as it depreciates
from 515 BHN to 165 BHN per worn depth of the component. With the wear
coefficient, normal force, hardness, and length of sliding distance determined, Archard’s
law was then used to determine volumetric wear rate of each alloy. A discrete numerical
approach was analyzed per sliding distance. Damper mass, normal load, density, and
hardness were varied to better understand volumetric wear rates of this system.
BHN
[kg/mm^2]
165
235
305
Ni alloy
375
445
515
Co-Cr alloy
Table 2: Hardness Values
BHN Hardness [kg/mm^2]
Carburization Hardness
600
500
400
300
200
100
0
0
0.0005
0.001
0.0015
0.002
0.0025
0.003
Material depth, [m]
Figure 3: Carburization hardness variation per depth
0.0035
Figure 4: Volumetric Wear Rate per Cycle
As expected the softer Nickel alloy material at 165 BHN can be seen in Figure 4
to have the highest volumetric wear rate per cycle. However the 165 BHN Nickel alloy
also has the largest variation in wear rate over the life of the part. As the hardness of
the material increases from 165 BHN to 515 BHN, the initial wear rate and variation at
end of life progressively decreases. For the materials of constant hardness, density and
change of volume of the damper are the primary variables impacting wear rates due to
decreasing normal loads, Fn. The carburized material shows an interesting upward
trend plot. Due to the bulk of the material having a lighter density, the wear rate up to
total sliding distance of about 240 meters is less than that of the 515 BHN Cobalt
Chromium Alloy. This equates to about 2667 hours of operation. The wear rate of the
carburized material increases in a nonlinear fashion, but never reaches the highest
wear rate of the bulk material.
165 BHN
19.8%
235 BHN
14.5%
305 BHN
11.5%
375 BHN
9.6%
445 BHN
8.2%
Table 3: Damper Volume Loss at End of Life
515 BHN
7.2%
Carburized
9.8%
Figure 5: Damper Volume
Table 3 and Figure 5 highlight the change of the damper volume over the life of
operation. Mass of the damper is essential for sufficient damping of the turbine blade.
As the mass of the damper changes sufficient damping of high stress locations may be
at risk of increasing frequency amplitudes. For constant hardness alloys the volume
reduction isn’t quite linear. This indicates there isn’t a rapid deterioration in material if
the contact area remains constant as material is worn away. Due to relatively low loads
and local contained movement, oxidation and transfer of melted material at the interface
will likely benefit wear rates and preserve damper volume beyond what’s shown in
Figure 5. Due to the large loss in volume of the 165 BHN and 235 BHN Nickel alloys,
these materials are not preferred for this application. Up to about 500 meters of total
sliding distance, the carburized material retains its volume similarly to that of the 515
BHN alloy. This is equivalent to about 5556 hours of engine running.
Comsol Multiphysics was used to understand the stresses seen by the bulk
material of the carburized damper under initial loading and end of life loading conditions.
A plastic model was run as peak von Mises stresses far exceeded yield strength of the
alloy. Figure 6 shows the highest peak stress to be 2.7 GPa after the first tangential
motion of 25 micrometers. Over time the damper material is worn away and the normal
loading is reduced at the end of life of the part. Figure 7 highlights damper stress to be
1.5 GPa for this condition. This is profoundly reduced from initial loading, but plastic
deformation remains due to the high rotational speed and centrifugal loading of the
sytem.
Figure 6: Damper Stress under First Cyclic Loading
Figure 7: Damper Stress at End of Life Loading
In conclusion, the evidence shown supports the use of a gas turbine damper
comprised of a high temperature Cobalt-Chromium alloy material. This is primarily due
to improved wear resistance and high hardness values at conditions up to 600°C.
Archard’s law for wear further shows that gas carburization can improve damper
component life while keeping weight to a minimum by retaining lighter bulk material
properties.
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