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Failure analysis of a gas turbine compressor
Engineering Failure Analysis 18 (2011) 474–484
Contents lists available at ScienceDirect
Engineering Failure Analysis journal homepage: www.elsevier.com/locate/engfailanal
Failure analysis of a gas turbine compressor G.H. Farrahi a, M. Tirehdast b, E. Masoumi Khalil Abad c,⇑, S. Parsa d, M. Motakefpoor e a
School of Mechanical Engineering, Sharif University of Technology, Tehran, Iran Materials Science and Engineering Department, Sharif University of Technology, Tehran, Iran Mechanical Engineering Department, McGill University, Montreal, Canada d Rei Power Station, Tehran, Iran e Golpaygan College of Engineering, Golpaygan, Iran b c
a r t i c l e
i n f o
Article history: Received 8 September 2010 Received in revised form 25 September 2010 Accepted 30 September 2010 Available online 28 October 2010 Keywords: Gas turbine Compressor disk Failure Fatigue Finite element method
a b s t r a c t During the shut down period, a 32 MW gas turbine experienced a severe failure accompanied by a loud noise near its second natural frequency at 4200 rpm. After opening the turbine casing, it was revealed that the disks of stages 16 and 17 of the compressor had been fractured and all of the stationary and rotary blades of stages 14–18 of the compressor had been detached from the dovetail region of the disks. The degree of damage was such that repairing the compressor was not economical, and thus, the compressor was no longer able to be used. Diagnostic work was carried out using different finite element models and fractography analysis. Analysis showed that multiple cracks had been initiated in the interface of the disks and shaft by the fretting fatigue mechanism and had been propagated by fatigue mechanism. Finally, unexpected and/or excessive forces or impact loads had led to the final brittle fracture of the disks. Some recommendations are proposed for preventing similar failures in the future. Ó 2010 Elsevier Ltd. All rights reserved.
1. Introduction A gas turbine, used for generating electric power, is a complex system with numerous rotary and stationary parts working under precisely controlled working conditions. In a gas turbine unit, the inlet ambient air is compressed by passing through several stages of stationary and rotary blades and can then be used both in the combustion chamber and for cooling purposes. The compressed air that enters the combustion chamber is mixed with fuel and is ignited to provide a high pressure, high velocity, and high temperature gas flow that is able to drive the turbine shaft at high rotary speeds. However, due to the precise design conditions of gas turbine units and the high rotary speeds at which they operate, the malfunction of one component can lead to severe damage to the entire unit. In between, the rotary and stationary parts of the turbine section, such as blades and disks, are more prone to failure because they work in a corrosive environment under a high temperature gas flow with a high pressure gradient. Several failures in the turbine section were reported and studied by different authors, with the aim of preventing future failures by improving the mechanical design, designing new materials, or proposing guidelines for better maintenance and utilization of gas turbine units. The failure mechanism of the gas turbine due to damage in turbine disks or blades is studied in [1–7] by using visual inspection, macro and micro fractography, and numerical mechanical analyses. In these studies, the fatigue fracture, existence of region with high stress levels, fretting, foreign object damage, and material degradation due to surface erosion and oxidation were identified as the main failure mechanisms.
⇑ Corresponding author. E-mail address: [email protected] (E. Masoumi Khalil Abad). 1350-6307/$ - see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.engfailanal.2010.09.042
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However, gas turbine damage due to compressor failure has also been reported and investigated in few studies. The failure of the compressor disk of stage 17 of a J79 engine was investigated by using metallurgical tests in [8] and the stall was determined as the main failure mechanism; in this case, when the compressor was operating at rotational speeds well below the nominal design speed, the air density in the last few stages was very low, and the blades stalled, which led to aerodynamically induced vibrations resulting in fatigue failures and total engine destruction. The failure of a Ti6Al4V compressor blades because of tight contact between the blade root and the disk in dovetail region and low wear resistance of the blade was studied in [9] by using experimental tests and numerical simulation. Two gas turbine failures due to compressor blades breakage at different compressor stages were reported and studied in [10]. One of the latest damages to a gas turbine due to failure of its compressor disk was reported in a 32 MW gas turbine in Rei electrical power plant, Iran, which led to severe damage to the turbine shaft and stator casing [11]. This paper explores the failure mechanism of this damage through visual inspections, SEM fractography and numerical modeling in order to propose some recommendations for the utilization and maintenance of similar gas turbine units to avoid future incidents.
1.1. Background After 2 days of operation at nominal speed of 5000 rpm, a 32 MW gas turbine was shut down. Meanwhile, near the second natural frequency of the turbine rotor, around 4200 rpm, the turbine vibration increased suddenly and the turbine experienced a severe failure accompanied by a loud noise. This turbine was manufactured by ACEC Company and was in service for 67,481 h before this failure [11]. After disassembling the turbine casing, it was revealed that the disks of stages 16 and 17 of the turbine’s compressor were fractured, and all of the compressor’s stationary and moving blades in stages 14–18 were deformed or broken from the dovetail region, as shown in Figs. 1 and 2. A relatively thick layer of dust had been deposited on the surface of the compressor blades of the initial stages, and several compressor blades in stages 11–14 were twisted. The degree of damage was such that repairing the compressor was not economical, and thus, the compressor was no longer able to be used. Six months after this failure, a similar failure occurred in another gas turbine unit and raised alarm about probable failures in other units. Therefore to prevent further failures, a failure root cause analysis was carried out using numerical modeling and fractography analysis.
Fig. 1. (a) The turbine compressor after failure; (b) and (c) zoomed views from the front and back of the stages 15–18; (d) the fractured part of the stage 17.
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Fig. 2. Turbine stator after failure.
1.2. Compressor disk fracture Fig. 1c shows the compressor disks of stages 16 and 17, which looks like a horseshoe as a result of the failure. As can be seen in Fig. 3a and b, a radial crack, which was initiated from the interface of the disk at stage 16 and the shaft, was opened due to shearing of the disk on the stator during failure. Significant plastic deformation and local buckling could be observed
Fig. 3. (a) The fracture surface of the stage 16; (b) zoomed view of the fatigue crack propagation zone (zone A); (c) and (d) zoomed view of disk web with chevron lines pattern (zone B).
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Fig. 4. The cracks initiated form the disk dovetail region.
in the disk, which formed as the disk was opened during the failure (Fig. 1b). In contrast, as shown in Fig. 1c and d, the disk of stage 17 was broken into two parts due to the opening of the radial cracks. No significant plastic deformation was observed in this disk. As shown in Fig. 4, several macro-cracks were nucleated from the dovetail region of the disks of stages 16 and 17. These cracks had different lengths, varying from 1 cm to 30 cm, and were initially propagated in the radial direction and then in the circumferential direction. 2. Fractography The fracture surfaces of the two disks are quite similar and can be divided into two zones (Fig. 3). As shown in Fig. 3, the darker and smaller zone, represented by A, contained fatigue beach marks and radial lines originating from different points at the interface of the disk and shaft, while the fatigue striations are clear in the SEM image of zone A (Fig. 5). These observations demonstrate the dominance of the fatigue fracture mechanism in formation of this zone [2]. Furthermore, the different origins of radial lines suggests that the fatigue cracks were initiated as a result of the fretting fatigue mechanism and propagated under the alternative loading condition and fatigue mechanism. In contrast, as can be seen in Fig. 3, zone
Fig. 5. Zoomed views of zone A, which was created by the fatigue fracture mechanism.
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Fig. 6. Zoomed views of zone B, which was created by the brittle fracture mechanism.
B occupied nearly 96 percent of the entire fracture surface and had a bright and facet surface; the web region of the disk was covered with the chevron line patterns (Fig. 3c and d). The SEM images of this zone, shown in Fig. 6, resemble typical cleavage fracture features. Based on these fractography observations and the sudden failure of the gas turbine unit, the cleavage fracture was the dominant mechanism in the final fracture of the disk of stages 16 and 17. On the other hand, because of the high ductility of the disk material and relatively high temperature working conditions of the disk of stages 16 and 17, the sudden failure could have occurred as a result of a sudden overload and/or impact loading conditions [12]. 3. FEM modeling A gas turbine is a complex system with numerous rotary and stationary parts, which work precisely under controlled conditions. Comprehensive theoretical and experimental knowledge about various fields of mechanical engineering, such as structural mechanics, dynamics and vibration, aerodynamics, heat transfer, and combustion is required for the efficient designing, manufacturing, and utilization, of a gas turbine unit, and is also necessary for proper repair and maintenance of these units. Also, due to this broad range of parameters, performing a proper and comprehensive failure analysis depends on careful selection of the mechanical and metallurgical parameters. Failure reports, visual inspections, and metallurgical investigation, plus knowledge and experience of the failure with the behaviour and mechanism of gas turbines are essential sources for writing different failure scenarios and choosing proper mechanical factors for mechanical and mathematical analysis. In this study, the following facts were extracted based on the failure report, visual observations, and fractography analysis, which were discussed in the previous sections: 1. 2. 3. 4. 5. 6. 7.
Failure occurred near the second natural frequency of the rotor. Severe vibration was observed before the failure. All of the blades of the stages 14–18 were fractured or broken. The fatigue cracks were initiated from the interface of the disk and the shaft. There are cracks that were initiated from the dovetail region of the disk. Some of the blades in stages close to the compressor’s bleed valves were twisted. The disks were assembled on the shaft using shrink-fit method.
Facts 1, 2, and 3 suggest that modal analysis should be performed to study the role of vibration in the failure. While there is the possibility of the turbine failure as a result of high stress or contact pressure levels in the compressor disks, facts 3, 4, 5,
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and 7, the mechanical analysis should be performed to obtain the stress distribution in the compressor disk under normal working conditions. Because of complexity of the proposed analysis, the authors could not find an analytical method to analyze and solve the mechanical models. In such situations, numerical methods, such as finite element method (FEM), can be implemented as a robust tool for solving the structural problems. In this method, a structure is divided into smaller parts called elements, which are connected to their neighbouring elements through a limited number of points, namely nodes. Solving a system of equations, based on the analysis type, the unknown degrees of freedom of each node will be obtained. For static structural analysis, the problem of obtaining the displacement field is solved by obtaining the solution of the following system of equations:
½KfXg ¼ fFg
ð1Þ
where [K] is the stiffness matrix of the structure and {X} and {F} are the nodal displacement and nodal force vectors, respectively. Having the displacement field inside the structure, the stress and strain distribution in the structure can be calculated easily. In the present study, ANSYS [13] commercial finite element software was implemented to create and to solve proposed models for the turbine failure. 3.1. Stress distribution in compressor disk Under normal working conditions, the compressor disks are subjected to rotor torque, aerodynamic forces, and centrifugal forces resulting from the rotation of the disk itself and the attached rotary blades. In present work, the disk rotation was modeled by applying a global angular velocity on the disk nodes, and the effect of blade rotation was modeled by an equivalent ring attached to the disk rim that applies the same centrifugal force as the rotary blades. The outer diameter of the ring obtained as:
r2o;r ¼ r2o;d þ
F c;b
ð2Þ
ptx2
where r 2o;r and r 2o;d are, respectively, the outer diameter of the equivalent ring and the compressor disk, Fc,b is the centrifugal force of the blades under normal working conditions, t is the disk thickness in the dovetail region and x is the angular velocity of the disk. By implementing fundamentals of turbo-machinery theory, the aerodynamic forces on the disk were estimated under incompressible, one dimensional, isentropic fluid flow with constant pressure amplification ratio (q) at each compressor stage. Using the equations of the balance of the angular momentum for a control volume around the compressor disk, shown in Fig. 7, we have [14]:
s ¼ m_ rV a ðtan a3 tan a2 Þ
ð3Þ
Fig. 7. The flow velocity triangles.
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_ is the mass flow, Va is the average axial velocity, a2 and a3, are the where s is the applied torque from fluid flow on the disk, m angles of entering and exiting flow respect to the control volume, respectively, and r is distance between the middle of the blade to rotation axis. Substituting the estimated values for these variables, based on the turbine features in Table 1, the applied torque on the disk was calculated to be about 8.1 kN m. Two 2D finite element models, one for investigating the stress distribution inside the compressor disk under normal working conditions, and the other for obtaining the contact pressure distribution in the interface of the disks and shaft, were designed and solved. Except for the small region corresponding to the dovetail region, the disk geometry can be obtained by revolving its cross section around the rotor axis, as shown in Fig. 8; hence, in the present study, the disk was modeled by using a two dimensional FE model. Plane 83, a 2D, eight nodes element with the capability of modeling axi-symmetric structures with non-axi-symmetric loads was selected to mesh the disk cross section [3]. However, Plane 82 element type was used to obtain the contact pressure distribution as Plane 83 is not applicable in contact problems. Thus, in the latter model, the aerodynamic force was neglected as it is negligible compared to the centrifugal force. The material properties of the turbine disk are given in Table 2. To model the effect of the shrink-fit process for assembling the disks on the shaft, a temperature load of 150 °C was applied on the disk nodes [1]. 3.2. Modal analysis The probable contribution of the resonant at natural frequency of the rotor was investigated through performing a 3D modal analysis of the rotor. Because of the geometrical complexity of the rotor, modal analysis of the entire rotor was a time consuming and computationally expensive effort; therefore, the model complication was reduced by using a simplified 2D model of the rotor, which was meshed by 2D Plane 82 element type. The 3D model of the rotor was simply created Table 1 The failed compressor characteristics. Pressure ratio
Angular velocity (rpm)
Input temperature (°C)
Output temperature (°C)
Input pressure (kPa)
Number of blades of stages 17 and 16
10
5000
20
300
100
81
Fig. 8. 2D meshed model of the compressor disk.
Table 2 The mechanical properties of compressor disk. Young modulus (GPa)
Heat expansion coefficient (1/°C)
Density
240
14 106
7860
kg m3
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Fig. 9. 3D meshed model of the rotor for modal analysis.
by sweeping the 2D elements around the rotor axis, as shown in Fig. 9. In the present study, the rotor bearings were modeled using the spring and damper elements located perpendicular to the rotor axis. Based on the characteristic curve for four groove bearing type [15], the spring-damper element constants, representing spring and damper values, were obtained as, K = 1.3 1010 N/m and C = 3 108 N s/m. The blades were modeled using an equivalent ring, as described in previous section. After imposing the boundary conditions on the spring-damper elements, the modal analysis of the rotor was performed based on the eigenvalue method.
Fig. 10. (a) Von-Mises stress distribution, (b) contact pressure distribution, (c) contact type in compressor disk under working conditions.
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3.3. Blade and disk contact analysis The contact pressure distribution at the interface of the disks and the blades under normal working conditions was investigated by performing a 2D finite element. Because of the cyclic symmetry of the disk only a sector of the disk that contains one blade was selected for modeling, and the nodes on the two edges of the disk were coupled instead. Eight node Plane 82 element type, with plane stress element formulation, was implemented to mesh the model. In order to model the contact between the blade and disk, contact elements were placed at the interface of the two components. The aerodynamic and centrifugal forces on the disk were modeled, respectively, by applying a line pressure on the blade edge and a constant angular velocity on all the model nodes. At the end, a non-linear static analysis was performed to solve the model.
4. Results Distribution of Von-Mises stress and contact pressure in the compressor disk as well as the contact type in the interface of disk and shaft are shown in Fig. 10. It can be seen that the maximum Von-Mises stress occurs in the interface of the disk and the shaft near the blade edges, while the maximum contact pressure as well as the transition point between sticking and slipping contact types occur somewhere inside the disk. It is worthwhile to notice that the transition point between sticking and sliding friction types happens near the fretting fatigue crack nucleation sites in the failed compressor disk. The result of the modal analysis shows that the rotor passes two natural frequencies at 40 Hz and 71 Hz before working in nominal working frequency, at 85 Hz, which agrees with the observed natural frequencies in practice. The rotor modal shape at the second natural frequency is shown in Fig. 11. It can be seen that the compressor disks of the stages 15–18 are experience the maximum deflection near the second natural frequency. Fig. 12 shows the distribution of Von-Mises stress and contact pressure in the dovetail region of the disks. High contact pressure as well as a high gradient of contact pressure, point A, can be observed at a point near the origin of peripheral cracks in the failed disk (Fig. 4).
Fig. 11. The rotor mode shape at second natural frequency.
Fig. 12. Contact pressure distribution in dovetail region of the compressor disk.
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5. Discussion During turbine operation, the position of the transition point, which is located near the point with maximum contact pressure in Fig. 10b, was changing as a result of the rotor vibration; therefore, the points close to the transition point were subjected to both slipping/sticking contact types with small alternative movement, respect to the shaft, under high contact pressure. This situation led to nucleation of the fretting fatigue cracks in the interface of the disks and shaft. These fretting fatigue cracks were propagated in a small region under alternative stress field by fatigue mechanism before final brittle fracture, which can be seen in Fig. 3. However, by observing a ductile disk working under low nominal stress levels, which is in contrast with the expected final ductile fracture mechanism in small fraction of the entire fracture surface, it was concluded that an unusually high amplitude load and/or impact load was the main cause of failure. However, due to severe failure of the compressor, the exact sources of this unusual loading condition could not be determined, but some of probable sources will be mentioned later in this section. High contact pressure with high gradient in the compressor disk, seen at point A of Fig. 12, provided a suitable situation for nucleation of fretting fatigue cracks. The growth and opening of these cracks led to a reduction of applied pressure by the disks on the blades in the dovetail region that could lead to blade fracture and/or escaping of the blades and the subsequent highly unbalanced force on the compressor, especially near the natural frequency of the rotor. Existence of unbalanced aerodynamic forces that are produced by the flow disturbances, such as surge and stall, around the compressor disk during the start up and shut down periods increase the stress level in the compressor component. Such unbalanced forces can be controlled by different techniques such as controlling the fluid flow in compressor by opening specific valves, namely bleed valves, at a certain angular velocities. The malfunction of bleed valves, which was reported in the failure report of the failed gas turbine [11], led to surge in the compressor. Some of the compressor rotary blades near the bleed valves were plastically twisted, while some blades were detached from the disk and were found in the turbine exhaust, as a result of the surge, which subsequently applied a high unbalanced force on the compressor rotor. The result of the modal analysis shows that the rotor passes its second natural frequency at 71 Hz, near the turbine speed at failure. From the rotor mode shape, Fig. 11, it can be observed that the compressor disks of the stages 15–18 experience the maximum deflection near the second natural frequency. Hence, by increasing the vibration amplitude near the rotor second natural frequency, the fracture of some blades due to surge in the compressor or escaped blades from the cracked dovetails led to unusual impact loading on the compressor that was followed by sudden brittle failure of the disk of stage 17. Finally, by shearing of the compressor rotary components on the stator casing, which produced a loud noise, all of the blades of the stages 14–18 were damaged and broken. 6. Conclusion In this study, comprehensive failure analysis was carried out in order to investigate the failure mechanisms that led to the failure of a 32 MW gas turbine. The turbine failed due to fracture of its compressor disks at stages 16 and 17. It was shown that the fatigue cracks were initiated at the interface of the shaft and the disks by fretting fatigue mechanism and propagated under the fatigue mechanism. Unexpected high amplitude and/or impact forces, applied to the compressor rotary components, led to the sudden brittle fracture of the compressor disks of stages 16 and 17. Unbalanced forces due to fractured or escaped blades, and/or fluid flow disorders, such as surge, plus the increasing of the vibration amplitude of the rotor near the second natural frequency of the rotor were the probable sources of the applied unusual loading condition on the compressor rotary components. Finally, by shearing of the compressor rotary components on the stator casing, which was accompanied by a loud noise, all of the blades of the stages 14–18 were damaged and broken and the turbine failed. Using the findings of this study, the following recommendations were proposed to avoid similar failures in future: 1. 2. 3. 4.
Regular inspections of the rotor to detect the cracks in the disks. Implementing new instrumentation for detecting and controlling the compressor surge. Monitoring the vibrating behaviour of the compressor by using diagnostic systems. Using proper air filters and regular cleaning of the compressor blades.
Acknowledgement For this work we are grateful for the support of Rei Power Station. References [1] Carter TJ. Common failures in gas turbine blades. Eng Fail Anal 2005;12:237–47. [2] Hou J, Wicks BJ, Antoniou RA. An investigation of fatigue failures of turbine blades in a gas turbine engine by mechanical analysis. Eng Fail Anal 2002;9:201–11. [3] Khajavi MR, Shariat MH. Failure of first stage gas turbine blades. Eng Fail Anal 2004;11:589–97. [4] Kubiak J, Urquiza G, Rodriguez JA, González G, Rosales I, Castillo G, et al. Failure analysis of the 150 MW gas turbine blades. Eng Fail Anal 2009;16:1794–804.
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[5] Shlyannikov VN, Iltchenko BV, Stepanov NV. Fracture analysis of turbine disks, computational–experimental background of the operational decisions. Eng Fail Anal 2001;8:461–75. [6] Chang J-C, Yun Y-H, Choi C, Kim J-C. Failure analysis of gas turbine buckets. Eng Fail Anal 2003;10:559–67. [7] Poursaeidi E, Aieneravaie M, Mohammadi MR. Failure analysis of a second stage blade in a gas turbine engine. Eng Fail Anal 2008;15:1111–29. [8] Song K-S, Kim S-G, Hwang Y-H. Failure of the J79 engine compressor blade due to stall. J Fail Anal Prevent 2007;7:212–7. [9] Kermanpur A, Sepehri Amin H, Ziaei-Rad S, Nourbakhshnia N, Mosaddeghfar M. Failure analysis of Ti6Al4V gas turbine compressor blades. Eng Fail Anal 2008;15:1052–64. [10] Infante V, Silva JM, de Freitas M, Reis L. Failures analysis of compressor blades of aeroengines due to service. Eng Fail Anal 2009;16:1118–25. [11] Operational and technical documents of the ACEC gas turbine units available in Rei Power plant; 2008. [12] Fail Anal Prevent. 1998 ed.; 1998. [13] ANSYS Inc. user manual, version 12. [14] Dixon SL. Fluid mechanics and thermodynamics of turbomachinery. 4th ed. Elsevier; 1998. [15] Ehrich F. Handbook of rotordynamics. New York: McGraw-Hill; 1992.
Contents lists available at ScienceDirect
Engineering Failure Analysis journal homepage: www.elsevier.com/locate/engfailanal
Failure analysis of a gas turbine compressor G.H. Farrahi a, M. Tirehdast b, E. Masoumi Khalil Abad c,⇑, S. Parsa d, M. Motakefpoor e a
School of Mechanical Engineering, Sharif University of Technology, Tehran, Iran Materials Science and Engineering Department, Sharif University of Technology, Tehran, Iran Mechanical Engineering Department, McGill University, Montreal, Canada d Rei Power Station, Tehran, Iran e Golpaygan College of Engineering, Golpaygan, Iran b c
a r t i c l e
i n f o
Article history: Received 8 September 2010 Received in revised form 25 September 2010 Accepted 30 September 2010 Available online 28 October 2010 Keywords: Gas turbine Compressor disk Failure Fatigue Finite element method
a b s t r a c t During the shut down period, a 32 MW gas turbine experienced a severe failure accompanied by a loud noise near its second natural frequency at 4200 rpm. After opening the turbine casing, it was revealed that the disks of stages 16 and 17 of the compressor had been fractured and all of the stationary and rotary blades of stages 14–18 of the compressor had been detached from the dovetail region of the disks. The degree of damage was such that repairing the compressor was not economical, and thus, the compressor was no longer able to be used. Diagnostic work was carried out using different finite element models and fractography analysis. Analysis showed that multiple cracks had been initiated in the interface of the disks and shaft by the fretting fatigue mechanism and had been propagated by fatigue mechanism. Finally, unexpected and/or excessive forces or impact loads had led to the final brittle fracture of the disks. Some recommendations are proposed for preventing similar failures in the future. Ó 2010 Elsevier Ltd. All rights reserved.
1. Introduction A gas turbine, used for generating electric power, is a complex system with numerous rotary and stationary parts working under precisely controlled working conditions. In a gas turbine unit, the inlet ambient air is compressed by passing through several stages of stationary and rotary blades and can then be used both in the combustion chamber and for cooling purposes. The compressed air that enters the combustion chamber is mixed with fuel and is ignited to provide a high pressure, high velocity, and high temperature gas flow that is able to drive the turbine shaft at high rotary speeds. However, due to the precise design conditions of gas turbine units and the high rotary speeds at which they operate, the malfunction of one component can lead to severe damage to the entire unit. In between, the rotary and stationary parts of the turbine section, such as blades and disks, are more prone to failure because they work in a corrosive environment under a high temperature gas flow with a high pressure gradient. Several failures in the turbine section were reported and studied by different authors, with the aim of preventing future failures by improving the mechanical design, designing new materials, or proposing guidelines for better maintenance and utilization of gas turbine units. The failure mechanism of the gas turbine due to damage in turbine disks or blades is studied in [1–7] by using visual inspection, macro and micro fractography, and numerical mechanical analyses. In these studies, the fatigue fracture, existence of region with high stress levels, fretting, foreign object damage, and material degradation due to surface erosion and oxidation were identified as the main failure mechanisms.
⇑ Corresponding author. E-mail address: [email protected] (E. Masoumi Khalil Abad). 1350-6307/$ - see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.engfailanal.2010.09.042
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475
However, gas turbine damage due to compressor failure has also been reported and investigated in few studies. The failure of the compressor disk of stage 17 of a J79 engine was investigated by using metallurgical tests in [8] and the stall was determined as the main failure mechanism; in this case, when the compressor was operating at rotational speeds well below the nominal design speed, the air density in the last few stages was very low, and the blades stalled, which led to aerodynamically induced vibrations resulting in fatigue failures and total engine destruction. The failure of a Ti6Al4V compressor blades because of tight contact between the blade root and the disk in dovetail region and low wear resistance of the blade was studied in [9] by using experimental tests and numerical simulation. Two gas turbine failures due to compressor blades breakage at different compressor stages were reported and studied in [10]. One of the latest damages to a gas turbine due to failure of its compressor disk was reported in a 32 MW gas turbine in Rei electrical power plant, Iran, which led to severe damage to the turbine shaft and stator casing [11]. This paper explores the failure mechanism of this damage through visual inspections, SEM fractography and numerical modeling in order to propose some recommendations for the utilization and maintenance of similar gas turbine units to avoid future incidents.
1.1. Background After 2 days of operation at nominal speed of 5000 rpm, a 32 MW gas turbine was shut down. Meanwhile, near the second natural frequency of the turbine rotor, around 4200 rpm, the turbine vibration increased suddenly and the turbine experienced a severe failure accompanied by a loud noise. This turbine was manufactured by ACEC Company and was in service for 67,481 h before this failure [11]. After disassembling the turbine casing, it was revealed that the disks of stages 16 and 17 of the turbine’s compressor were fractured, and all of the compressor’s stationary and moving blades in stages 14–18 were deformed or broken from the dovetail region, as shown in Figs. 1 and 2. A relatively thick layer of dust had been deposited on the surface of the compressor blades of the initial stages, and several compressor blades in stages 11–14 were twisted. The degree of damage was such that repairing the compressor was not economical, and thus, the compressor was no longer able to be used. Six months after this failure, a similar failure occurred in another gas turbine unit and raised alarm about probable failures in other units. Therefore to prevent further failures, a failure root cause analysis was carried out using numerical modeling and fractography analysis.
Fig. 1. (a) The turbine compressor after failure; (b) and (c) zoomed views from the front and back of the stages 15–18; (d) the fractured part of the stage 17.
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Fig. 2. Turbine stator after failure.
1.2. Compressor disk fracture Fig. 1c shows the compressor disks of stages 16 and 17, which looks like a horseshoe as a result of the failure. As can be seen in Fig. 3a and b, a radial crack, which was initiated from the interface of the disk at stage 16 and the shaft, was opened due to shearing of the disk on the stator during failure. Significant plastic deformation and local buckling could be observed
Fig. 3. (a) The fracture surface of the stage 16; (b) zoomed view of the fatigue crack propagation zone (zone A); (c) and (d) zoomed view of disk web with chevron lines pattern (zone B).
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Fig. 4. The cracks initiated form the disk dovetail region.
in the disk, which formed as the disk was opened during the failure (Fig. 1b). In contrast, as shown in Fig. 1c and d, the disk of stage 17 was broken into two parts due to the opening of the radial cracks. No significant plastic deformation was observed in this disk. As shown in Fig. 4, several macro-cracks were nucleated from the dovetail region of the disks of stages 16 and 17. These cracks had different lengths, varying from 1 cm to 30 cm, and were initially propagated in the radial direction and then in the circumferential direction. 2. Fractography The fracture surfaces of the two disks are quite similar and can be divided into two zones (Fig. 3). As shown in Fig. 3, the darker and smaller zone, represented by A, contained fatigue beach marks and radial lines originating from different points at the interface of the disk and shaft, while the fatigue striations are clear in the SEM image of zone A (Fig. 5). These observations demonstrate the dominance of the fatigue fracture mechanism in formation of this zone [2]. Furthermore, the different origins of radial lines suggests that the fatigue cracks were initiated as a result of the fretting fatigue mechanism and propagated under the alternative loading condition and fatigue mechanism. In contrast, as can be seen in Fig. 3, zone
Fig. 5. Zoomed views of zone A, which was created by the fatigue fracture mechanism.
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Fig. 6. Zoomed views of zone B, which was created by the brittle fracture mechanism.
B occupied nearly 96 percent of the entire fracture surface and had a bright and facet surface; the web region of the disk was covered with the chevron line patterns (Fig. 3c and d). The SEM images of this zone, shown in Fig. 6, resemble typical cleavage fracture features. Based on these fractography observations and the sudden failure of the gas turbine unit, the cleavage fracture was the dominant mechanism in the final fracture of the disk of stages 16 and 17. On the other hand, because of the high ductility of the disk material and relatively high temperature working conditions of the disk of stages 16 and 17, the sudden failure could have occurred as a result of a sudden overload and/or impact loading conditions [12]. 3. FEM modeling A gas turbine is a complex system with numerous rotary and stationary parts, which work precisely under controlled conditions. Comprehensive theoretical and experimental knowledge about various fields of mechanical engineering, such as structural mechanics, dynamics and vibration, aerodynamics, heat transfer, and combustion is required for the efficient designing, manufacturing, and utilization, of a gas turbine unit, and is also necessary for proper repair and maintenance of these units. Also, due to this broad range of parameters, performing a proper and comprehensive failure analysis depends on careful selection of the mechanical and metallurgical parameters. Failure reports, visual inspections, and metallurgical investigation, plus knowledge and experience of the failure with the behaviour and mechanism of gas turbines are essential sources for writing different failure scenarios and choosing proper mechanical factors for mechanical and mathematical analysis. In this study, the following facts were extracted based on the failure report, visual observations, and fractography analysis, which were discussed in the previous sections: 1. 2. 3. 4. 5. 6. 7.
Failure occurred near the second natural frequency of the rotor. Severe vibration was observed before the failure. All of the blades of the stages 14–18 were fractured or broken. The fatigue cracks were initiated from the interface of the disk and the shaft. There are cracks that were initiated from the dovetail region of the disk. Some of the blades in stages close to the compressor’s bleed valves were twisted. The disks were assembled on the shaft using shrink-fit method.
Facts 1, 2, and 3 suggest that modal analysis should be performed to study the role of vibration in the failure. While there is the possibility of the turbine failure as a result of high stress or contact pressure levels in the compressor disks, facts 3, 4, 5,
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and 7, the mechanical analysis should be performed to obtain the stress distribution in the compressor disk under normal working conditions. Because of complexity of the proposed analysis, the authors could not find an analytical method to analyze and solve the mechanical models. In such situations, numerical methods, such as finite element method (FEM), can be implemented as a robust tool for solving the structural problems. In this method, a structure is divided into smaller parts called elements, which are connected to their neighbouring elements through a limited number of points, namely nodes. Solving a system of equations, based on the analysis type, the unknown degrees of freedom of each node will be obtained. For static structural analysis, the problem of obtaining the displacement field is solved by obtaining the solution of the following system of equations:
½KfXg ¼ fFg
ð1Þ
where [K] is the stiffness matrix of the structure and {X} and {F} are the nodal displacement and nodal force vectors, respectively. Having the displacement field inside the structure, the stress and strain distribution in the structure can be calculated easily. In the present study, ANSYS [13] commercial finite element software was implemented to create and to solve proposed models for the turbine failure. 3.1. Stress distribution in compressor disk Under normal working conditions, the compressor disks are subjected to rotor torque, aerodynamic forces, and centrifugal forces resulting from the rotation of the disk itself and the attached rotary blades. In present work, the disk rotation was modeled by applying a global angular velocity on the disk nodes, and the effect of blade rotation was modeled by an equivalent ring attached to the disk rim that applies the same centrifugal force as the rotary blades. The outer diameter of the ring obtained as:
r2o;r ¼ r2o;d þ
F c;b
ð2Þ
ptx2
where r 2o;r and r 2o;d are, respectively, the outer diameter of the equivalent ring and the compressor disk, Fc,b is the centrifugal force of the blades under normal working conditions, t is the disk thickness in the dovetail region and x is the angular velocity of the disk. By implementing fundamentals of turbo-machinery theory, the aerodynamic forces on the disk were estimated under incompressible, one dimensional, isentropic fluid flow with constant pressure amplification ratio (q) at each compressor stage. Using the equations of the balance of the angular momentum for a control volume around the compressor disk, shown in Fig. 7, we have [14]:
s ¼ m_ rV a ðtan a3 tan a2 Þ
ð3Þ
Fig. 7. The flow velocity triangles.
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_ is the mass flow, Va is the average axial velocity, a2 and a3, are the where s is the applied torque from fluid flow on the disk, m angles of entering and exiting flow respect to the control volume, respectively, and r is distance between the middle of the blade to rotation axis. Substituting the estimated values for these variables, based on the turbine features in Table 1, the applied torque on the disk was calculated to be about 8.1 kN m. Two 2D finite element models, one for investigating the stress distribution inside the compressor disk under normal working conditions, and the other for obtaining the contact pressure distribution in the interface of the disks and shaft, were designed and solved. Except for the small region corresponding to the dovetail region, the disk geometry can be obtained by revolving its cross section around the rotor axis, as shown in Fig. 8; hence, in the present study, the disk was modeled by using a two dimensional FE model. Plane 83, a 2D, eight nodes element with the capability of modeling axi-symmetric structures with non-axi-symmetric loads was selected to mesh the disk cross section [3]. However, Plane 82 element type was used to obtain the contact pressure distribution as Plane 83 is not applicable in contact problems. Thus, in the latter model, the aerodynamic force was neglected as it is negligible compared to the centrifugal force. The material properties of the turbine disk are given in Table 2. To model the effect of the shrink-fit process for assembling the disks on the shaft, a temperature load of 150 °C was applied on the disk nodes [1]. 3.2. Modal analysis The probable contribution of the resonant at natural frequency of the rotor was investigated through performing a 3D modal analysis of the rotor. Because of the geometrical complexity of the rotor, modal analysis of the entire rotor was a time consuming and computationally expensive effort; therefore, the model complication was reduced by using a simplified 2D model of the rotor, which was meshed by 2D Plane 82 element type. The 3D model of the rotor was simply created Table 1 The failed compressor characteristics. Pressure ratio
Angular velocity (rpm)
Input temperature (°C)
Output temperature (°C)
Input pressure (kPa)
Number of blades of stages 17 and 16
10
5000
20
300
100
81
Fig. 8. 2D meshed model of the compressor disk.
Table 2 The mechanical properties of compressor disk. Young modulus (GPa)
Heat expansion coefficient (1/°C)
Density
240
14 106
7860
kg m3
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Fig. 9. 3D meshed model of the rotor for modal analysis.
by sweeping the 2D elements around the rotor axis, as shown in Fig. 9. In the present study, the rotor bearings were modeled using the spring and damper elements located perpendicular to the rotor axis. Based on the characteristic curve for four groove bearing type [15], the spring-damper element constants, representing spring and damper values, were obtained as, K = 1.3 1010 N/m and C = 3 108 N s/m. The blades were modeled using an equivalent ring, as described in previous section. After imposing the boundary conditions on the spring-damper elements, the modal analysis of the rotor was performed based on the eigenvalue method.
Fig. 10. (a) Von-Mises stress distribution, (b) contact pressure distribution, (c) contact type in compressor disk under working conditions.
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3.3. Blade and disk contact analysis The contact pressure distribution at the interface of the disks and the blades under normal working conditions was investigated by performing a 2D finite element. Because of the cyclic symmetry of the disk only a sector of the disk that contains one blade was selected for modeling, and the nodes on the two edges of the disk were coupled instead. Eight node Plane 82 element type, with plane stress element formulation, was implemented to mesh the model. In order to model the contact between the blade and disk, contact elements were placed at the interface of the two components. The aerodynamic and centrifugal forces on the disk were modeled, respectively, by applying a line pressure on the blade edge and a constant angular velocity on all the model nodes. At the end, a non-linear static analysis was performed to solve the model.
4. Results Distribution of Von-Mises stress and contact pressure in the compressor disk as well as the contact type in the interface of disk and shaft are shown in Fig. 10. It can be seen that the maximum Von-Mises stress occurs in the interface of the disk and the shaft near the blade edges, while the maximum contact pressure as well as the transition point between sticking and slipping contact types occur somewhere inside the disk. It is worthwhile to notice that the transition point between sticking and sliding friction types happens near the fretting fatigue crack nucleation sites in the failed compressor disk. The result of the modal analysis shows that the rotor passes two natural frequencies at 40 Hz and 71 Hz before working in nominal working frequency, at 85 Hz, which agrees with the observed natural frequencies in practice. The rotor modal shape at the second natural frequency is shown in Fig. 11. It can be seen that the compressor disks of the stages 15–18 are experience the maximum deflection near the second natural frequency. Fig. 12 shows the distribution of Von-Mises stress and contact pressure in the dovetail region of the disks. High contact pressure as well as a high gradient of contact pressure, point A, can be observed at a point near the origin of peripheral cracks in the failed disk (Fig. 4).
Fig. 11. The rotor mode shape at second natural frequency.
Fig. 12. Contact pressure distribution in dovetail region of the compressor disk.
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5. Discussion During turbine operation, the position of the transition point, which is located near the point with maximum contact pressure in Fig. 10b, was changing as a result of the rotor vibration; therefore, the points close to the transition point were subjected to both slipping/sticking contact types with small alternative movement, respect to the shaft, under high contact pressure. This situation led to nucleation of the fretting fatigue cracks in the interface of the disks and shaft. These fretting fatigue cracks were propagated in a small region under alternative stress field by fatigue mechanism before final brittle fracture, which can be seen in Fig. 3. However, by observing a ductile disk working under low nominal stress levels, which is in contrast with the expected final ductile fracture mechanism in small fraction of the entire fracture surface, it was concluded that an unusually high amplitude load and/or impact load was the main cause of failure. However, due to severe failure of the compressor, the exact sources of this unusual loading condition could not be determined, but some of probable sources will be mentioned later in this section. High contact pressure with high gradient in the compressor disk, seen at point A of Fig. 12, provided a suitable situation for nucleation of fretting fatigue cracks. The growth and opening of these cracks led to a reduction of applied pressure by the disks on the blades in the dovetail region that could lead to blade fracture and/or escaping of the blades and the subsequent highly unbalanced force on the compressor, especially near the natural frequency of the rotor. Existence of unbalanced aerodynamic forces that are produced by the flow disturbances, such as surge and stall, around the compressor disk during the start up and shut down periods increase the stress level in the compressor component. Such unbalanced forces can be controlled by different techniques such as controlling the fluid flow in compressor by opening specific valves, namely bleed valves, at a certain angular velocities. The malfunction of bleed valves, which was reported in the failure report of the failed gas turbine [11], led to surge in the compressor. Some of the compressor rotary blades near the bleed valves were plastically twisted, while some blades were detached from the disk and were found in the turbine exhaust, as a result of the surge, which subsequently applied a high unbalanced force on the compressor rotor. The result of the modal analysis shows that the rotor passes its second natural frequency at 71 Hz, near the turbine speed at failure. From the rotor mode shape, Fig. 11, it can be observed that the compressor disks of the stages 15–18 experience the maximum deflection near the second natural frequency. Hence, by increasing the vibration amplitude near the rotor second natural frequency, the fracture of some blades due to surge in the compressor or escaped blades from the cracked dovetails led to unusual impact loading on the compressor that was followed by sudden brittle failure of the disk of stage 17. Finally, by shearing of the compressor rotary components on the stator casing, which produced a loud noise, all of the blades of the stages 14–18 were damaged and broken. 6. Conclusion In this study, comprehensive failure analysis was carried out in order to investigate the failure mechanisms that led to the failure of a 32 MW gas turbine. The turbine failed due to fracture of its compressor disks at stages 16 and 17. It was shown that the fatigue cracks were initiated at the interface of the shaft and the disks by fretting fatigue mechanism and propagated under the fatigue mechanism. Unexpected high amplitude and/or impact forces, applied to the compressor rotary components, led to the sudden brittle fracture of the compressor disks of stages 16 and 17. Unbalanced forces due to fractured or escaped blades, and/or fluid flow disorders, such as surge, plus the increasing of the vibration amplitude of the rotor near the second natural frequency of the rotor were the probable sources of the applied unusual loading condition on the compressor rotary components. Finally, by shearing of the compressor rotary components on the stator casing, which was accompanied by a loud noise, all of the blades of the stages 14–18 were damaged and broken and the turbine failed. Using the findings of this study, the following recommendations were proposed to avoid similar failures in future: 1. 2. 3. 4.
Regular inspections of the rotor to detect the cracks in the disks. Implementing new instrumentation for detecting and controlling the compressor surge. Monitoring the vibrating behaviour of the compressor by using diagnostic systems. Using proper air filters and regular cleaning of the compressor blades.
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