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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