Failure Analysis Case Studies II Episode 8 ppt

233 8.0 - 7.2- 6.4 - 5.6 - 4.8 - 4.0 - 3.2 - 2.4 - 1.6- 0.8 - 0- - Axial coil j i 1 Transverse coil jl 3, 'd :. hi A I I I I I I I I I1 .i . _ ___._ : * _ 8 2 .a B 2.0 - 1.8 - 1.6 - 1.4 - 1.2 - 1.0 - . . . . . (a) 1/2 inch flange size Axial coil Transverse coil (b) 1 inch flange size Fig. 10. Frequency response of node 33 showing peak displacement for the transverse and axial coil con- figurations. 234 Table 2. Natural frequencies for 1 in. flange connections, with axial coil, showing the effects of reducing tube height and flange size; (A) infield configuration, (B) modified configuration 1 11.5 2 22.5 3 35.9 4 45.5 5 111.6 6 147.8 7 299.3 8 344.1 9 353.1 10 356.3 16.5 53.4 100.0 132.3 144.8 246.2 300.2 354.0 356.3 428.4 REFERENCES 1. AI-Asmi, K., Seibi, A,, Samanta, B. and Siddiqui, R., Investigation into the Failure of Pressure Transmitter Impulse Line. 2. Callistcr, Jr, W. D., Material Science and Engineering An Introduction, 3rd ed. John Wiley & Sons, Inc., 1994. 3. Hertzberg, R. W., Deformation and Fracture Mechanics of Engineering Materials, 2nd ed. John Wiley & Sons, Inc., 1983. 4. Shigley, J. S. and Mischke, C. R., Mechanical Engineering Des@, 5th ed. McGraw-Hill, 1989. Sahma Booster Station, Internal Report ME/SQU/July, 1995. Failure Analysis Case Studies II D.R.H. Jones (Editor) 0 2001 Elsevier Science Ltd. All rights reserved 235 MALFUNCTIONS OF A STEAM TURBINE MECHANICAL CONTROL SYSTEM J. H. BULLOCH* and A. G. CALLAGY Power Generation, ESB., Lower Fitzwilliam Street, Dublin 2, Ireland (Received 13 October 1997) Abstract This paper is aimed at elucidating the cause of a series of malfunctions involving the bending or breaking of main steam turbine throttle valve spindles which occurred at service times ranging from hundreds to several thousand hours in a number of 270 MW steam raising units. It was clearly established, by two distinct approaches (one engineering, one micromechanistic) that the stresses which produced these malfunctions were bending in nature and were the result of out-of-alignment deflections. In the case of the bent spindles the stresses were very high and approached flow strength levels of around 8000 MPa while the broken spindles were the results of fatigue initiation and subsequent growth from a thread root (stress concentration) location on the spindle. Using relevant fatigue crack propagation data for the valve spindle material at 300°C it was demonstrated that fatigue failures occurred at spindle deflections of between 0.9 and 1.6 mm. Finally, it was demonstrated that the fatigue breakage problem could be significantly reduced, especially at the lower end of the valve spindle deflection range, by a combination of re-profiling the thread root and shot peening. 0 1998 Published by Elsevier Science Ltd. All rights reserved. Keywords: Fatigue, fatigue crack growth, fatigue markings, plastic deformation, power-plant failures 1. INTRODUCTION During the past decade or so certain ESB stations that operate 270 MW units have encountered operational problems involving main steam turbine throttle valve spindles. Basically the problems were identified as: (i) valve sticking as a direct result of spindle undergoing permanent plastic deformation or bending; (ii) spindle fracturing during operation near the top of the valve spindle in the threaded section. The present study considered four separate spindle failures which occurred over a three year period and involved: (i) a detailed failure analysis or micromechanic assessment; and (ii) a basic engineering stress analysis approach. or 2 MECHANICAL CONSIDERATIONS A detailed view uf the valve spindle arrangement is shown in Fig. 1 and in this instance the valve was in the closed position. All failures occurred at position A at the root of a spindle thread while the spindle bending problem was observed at position B. At these positions the working temperature was assessed at about 300°C. Also, the forces acting on the valve spindle are illustrated in Fig. 1 between piston C and the actuating lever where the vertical force is the result of steam pressure and the angular force was caused by the angle of the actuating or lifting lever. The resultant force is a sideways bending force on the valve spindle and it was calculated that the relationship between the bending stress uB, and value of spindle deflection 1, at position A could be expressed as follows: ~g = 342(%), (1) a Author to whom correspondence should be addressed. Reprinted from Engineering Failure Analysis 5 (3), 235-240 (1998) 236 11 \M SPRING ff"i HOUSING CLEARANCE n 3 0.25MM ,- I ll&i POSITIONB B VALVE SPINDLE I STEAM ' TO TURBINE Fig. 1. Valve spindle system. where uB was in MPa and 1 was in mm. At the valve spindle bending location, Le. position B, the value of uB was 1.37 times greater because it was 118 mm further away from the applied force. In the stainless steel valve spindle, bending occurred when the material flow stress was attained, Le. half the sum of the yield and tensile stress, viz., 830 MPa, and hence the maximum stress at position A (see Fig. I) was around 600 MPa or a spindle deflection value of 1.75 mm. As this coincided with a machined IS0 thread, a local stress concentration factor (SCF) of around 4 was prevalent. At position B an area of wear, around 0.5 mm deep (see Fig. 1) was observed on one side of the valve spindle surface indicating the presence of a significant bending action or stress. This bending stress could have been the result of: (i) clearance between the piston and cylinder wall, position C (see Fig. 1) or (ii) misalignment of the value actuating housing. However, the facts that (i) clearance related bending stresses are negligible at unit loads approaching 200 MW and (ii) failure occurred at high loads, suggested that the failures observed at position A were the result of spindle deflections or bending stresses caused by misalignment or possible thermal distortion in the valve spindle system. 3. FAILURE CONSIDERATIONS From a detailed examination of the broken spindle fracture surfaces it was assessed that failure was the result of a ductile fatigue crack extension process which was initiated at the thread root 23 7 -05 I Fatigue Crack Fig. 2. Fatigue initiation site at thread root location rhread Root location Fig. 3. Coarse fatigue striations. Striation spacing around 15 pm location (see Fig. 2) and propagated by ductile striated crack growth (see Fig. 3). In three of the four failures surface crack initiation was helped by the presence of near surface titanium based non- metallic inclusions. The coarse nature of the striation spacing (see Fig. 3) suggested a high stress, low cycle fatigue situation. Indeed the short service lives of the four broken spindles, viz., 33, 136, 167 and 3 16 days added credence to this suggestion. 4. FATIGUE CRACK GROWTH ASSESSMENT In this section relevant real material fatigue crack growth data at 300°C was used in an effort to assess the valve spindle deflection values required to cause the real “in field” short service time failures observed over the past few years. Amzallag and Maillard [l] have reported fatigue crack growth results at 300°C for a similar martensitic stainless type bolting steel and the upper bound data was described as: da dn (2) - = 1.05 x 10-6(AK)2, 23 8 where daldn = fatigue crack growth rate in mm per cycle and AK = stress intensity range = 1.12Ao ,/Z in MPa Jm. In an attempt to determine the number of fatigue cycles based on the unit power load a typical load profile is shown in Fig. 4. From this figure it can be seen that over a one day period there was one large unit power change and about four transient power changes which were around half this value. Using a range of spindle deflection values (bending stresses), AKand thus the fatigue crack growth characteristics, over a range of defect sizes, could be assessed. Also, taking the spindle failure criteria as bein that fatigue crack length where K approaches the materials fracture toughness of - 250 MPa &, the number of fatigue cycles or time to spindle failure was obtained. The relationship between valve spindle deflection and time to failure is illustrated in Fig. 5. From Fig. 5 it can be seen that the four fatigue failures occurred over a valve spindle deflection range of 0.9 to 1.6 mm and resided somewhat below the maximum spindle deflection possible at this location of 1.75 mm. It is clear that certain facts have emerged which strongly indicate that spindle breakages occurred due to a high stress, low cycle, ductile, fatigue process, viz.: (i) high bending stress or deflections are needed to bend spindles; (ii) significant surface indentations of about 0.5 mm were observed; UNIT LOAD (MW 150 I I 1 8 16 24 TIME (HOURS) Fig. 4. Typical load profile. VALVE SPINDLE DEFLECTION (MM) 1.5 1.0 FAILURE NO/ 2 FAILURE NO/ 4 200 400 TIME TO FAIL-AYS) Fig. 5. Time lo fatigue failure versus spindle deflection. 239 VALVE SPINDLE DEFLECTION (MM) UNIT POWER 0 Fig. 6. Schematic of spindle deflection versus unit power and calculated cycles to fatigue failure. (iii) the failures occurred over very short and unacceptable service durations; and large spindle deflections occurred at high unit loads as failures occurred in this regime and could only have come from out of alignment of the actuating mechanism. As a result Fig. 6 is semi-schematic in nature and illustrates the spindle bending and spindle fatigue breakage regions together with the notional increase in valve spindle deflection with unit power and the number of fatigue cycles required to cause spindle failure. It is clear that reducing the spindle deflection increased the working life of the spindle; indeed at deflections approaching 0.6 mm it was predicted that the spindle life was 21,200 cycles or almost 12 years. 5. DISCUSSION AND RECOMMENDATIONS It has been reasonably demonstrated that the series of spindle failures were the result of a high stress, ductile fatigue process that was caused by significant out-of-alignment of the actuating mechanism during service. During an outage, an exercise was conducted to determine the “cold” amount of misalignment of an actuating mechanism where a fatigue failure occurred. The measured amount was 0.38 mm which was more than 40% of the lower end valve spindle deflection level of 0.9 mm required to cause fatigue failure. As such, it is not difficult to envisage that thermal distortions during the hot “on-load” excursion could easily account for spindle deflections attaining the critical range necessary for fatigue failure. Such a recurring failure in a critical plant component needed to be urgently addressed in an effort to obviate or at least mitigate the problem. In the present instance it was suggested that the high stress situation at the spindle threaded location be reduced by two actions; firstly, changing the thread profile to a rounded thread (e.g. NF000-032 type thread) which had an associated stress concentration factor which was about 25% lower than the present square IS0 thread and secondly by introducing significant compressive stresses (which need to be overcome before fatigue can occur) by shot peening. The influence of these actions on the valve spindle deflection-time to fatigue failure relationship is illustrated in Fig. 7. From Fig. 7 it can be seen that both shot peening and re-profiling the thread significantly increased the valve spindle life at spindle deflection levels of around 1 mm. Indeed at 0.9 mm spindle deflection the spindle service life was increased from less than one year to around 8 years. The service temperature in the region of valve spindle failure was estimated to be around 300°C and it is known that thermal relaxation of the compressive stresses can occur at high temperature. However, recently Gauchet et al. [2] have reported encouraging results where significant 240 8 DEFLECTION I I. 2000 4000 TIME TO FAILURE(DAYS) Fig. 7. Time to fatigue failure versus spindle deflection. compressive stresses (by shot peening) have remained in HP heater water chambers after 5 years at 270°C. Finally, it has been shown that thread re-profiling and shot peening can mitigate fatigue failure in valve spindles. Also regular checks of actuating system out-of-alignment should be carried out together with good insulation at key locations which should help minimise thermal distortions. REFERENCES 1. Amzallag, C. and Maillard, J. L., 9th Inter. Conf. Structural Mechanics in Reactor Technology, Smirt, Lausanne, Switzerland, Aug. 1987, Vol. F, pp. 173-180. 2. Gauchet, J. P., Reversat, C., Leguernic, Y., Lebrun, J. L., Castax, L. and Barrallier, L. K., EDF Feedback on French Feedwater Plants Repaired by Shot Peening, EPRI Inter. Conf. Welding & Repair Technology for Fossil Power Plants, March 1994, Williamsburg, Virginia, U.S.A. Failure Analysis Case Studies /I D.R.H. Jones (Editor) 0 200 1 Elsevier Science Ltd. All rights reserved 24 1 FATJGUE FAILURE OF HOLD-DOWN BOLTS FOR A HYDRAULIC CYLINDER GLAND C. TAO,* N. XI, H. YAN and Y. ZHANG AVlC Failure Analysis Center, PO Box 81-4, Beijing, 100095, P.R. China (Receiced 22 September 1997) Abstract-A hydraulic-cylinder gland system used in aircrah failed by leaking because the hold-down bolts broke in the course of a trial run. The metallographic examination of the fracture surface and the stress calculations for the bolts are described in this article. The investigation showed that the failure was caused by fatigue and the reason for failure was considered in relation to the processing, surface condition and assembly of the bolts. Measures to increase the fatigue strength of the bolts are proposed. CJ 1998 Elsevier Science Ltd. All rights reserved. 1. INTRODUCTION A hydraulic-cylinder gland system used in aircraft failed by leaking in the course of a trial run. The gland was fixed with one hold-down and four hold-down bolts. Three of the four bolts broke in service. The bolts were manufactured by turning and threading from 17-4PH steel. The nominal composition of 17-4PH is OCr-17NiltCu-4Nb and typical mechanical properties are yield strength = 1200 MPa, tensile strength ab = 1300 MPa after solution heat treatment at 1040"C, then water quenching and tempering for 4 h at 495°C. This paper describes an analysis of the nature and the causes of fracture as well as preventive measures for avoiding fatigue failure of the hold-down bolts. 2. METALLOGRAPHIC EXAMINATION A schematic drawing of the hold-down and bolts is shown in Fig. 1. The positions of the hold- down bolts are indicated by 1#, 2#, 3# and 4#. Each bolt head was cross drilled with two assembly holes at right angles to one another. The 3# bolt broke away in the middle of the threaded portion. The 1# and 2# bolts broke away in the head between the assembly holes and the shoulder transition radius. General views of the fracture surfaces taken in the scanning electron microscope are shown in Fig. 2. The fracture surface (Fig. 2(a)) of the 3# bolt was characteristic of a typical fatigue fracture, i.e. there was a crack initiation zone, a fatigue crack propagation zone and a final ductile fracture zone. The fatigue crack initiated at one position in the thread root at a machining mark. The crack propagated towards the far edge of the thread. The origin mne was rough (Fig. 3) and had many radial lines. The propagation zone was smooth and there were distinct fatigue striations (Fig. 4). In comparison with the fatigue surface, the final ductile zone was smaller and was around 20% of the total cross sectional area. According to the above-mentioned features, the fracture surface is characteristic of fatigue. The final ductile fracture zone was typically dimpled. No material defects were found in the fatigue origin zone. The fracture surfaces of the 1# and 2# bolts initiated at the edges of the assembly holes, as shown in Fig. 2(b) and Fig. 2(c). A schematic of the fatigue fracture sites is shown in Fig. 5. The cracks obviously propagated towards the root of the bolt until the remaining cross section became unable to support the load and failed by fast fracture. In comparison with the macro-fracture surface of 3# *Author to whom correspondence should be addressed. Reprinted from Engineering Failure Analysis 5 (3), 24 1-246 (I 998) 242 Fig. 1. Schematic of hold-down and bolts (numbered 14). Fig. 2. Fracture surfaces of bolts. (a) l#; (b) 2#; (c) 3#. Fracture origins are marked with an arrow. bolt, the fracture surfaces of bolts 1# and 2# were rather rough with propagation radial lines, but no fatigue beach marks. However, the photographs of the fracture surfaces taken in the SEM at high magnification showed that there were fatigue striations (Fig. 6). Mixed zones between the dimples and the fatigue striations existed in the fracture surfaces of bolts 1# and 2#. [...]... cycles to failure for each weight stack setting Fully reversed equivalent stress (MPa) Cycles to failure 23 47 70 93 1 I6 140 163 186 210 233 256 279 303 326 349 373 396 419 442 Weight stack setting (Ib) 1,337,234,019,763 6,465,350,374 285 , 787 ,954 31,259,1 18 5,617,1 38 I , 38 1,747 422,137 151,134 6 1,077 27,1 58 13,047 6 68 1 3609 204 1 1200 73 1 4 58 295 195 ~ 15 30 45 60 75 90 105 120 135 150 165 180 195... McGraw-Hill, 1 986 4 Ryder, G H., Strength of Materials, Macmillan Press Limited, 1969 Failure Analysis Case Studies II D.R.H Jones (Editor) 0 2001 Elsevier Science Ltd All rights reserved 255 Fatigue failure analysis of a leg press exercise machine P.J Vernon, T.J Mackin" Unicersity of' Illinois at Urbana-Champaign, Department of Mechanical and Industrial Engineering, 1206 West Green SI., Urbana, IL 6 180 1,... 146-7 [8] Peterson RE, Sines G editors Metal fatigue New York: McGraw-Hill Book Company, Inc., 1959 pp 296-7 [l] [2] [3] [4] Failure Analysis Case Studies II D.R.H Jones (Editor) 0 2001 Elsevier Science Ltd A11 rights reserved 267 Failure analysis of rubber fuel pipes in aero-engines Guoru Fu Beijing Aeronuuticul Technoiogy Research Center, Box 9203-16, Beging 100076, China Received 19 May 19 98; accepted... 90 105 120 135 150 165 180 195 210 225 Weight Stack Setting (lb.) Fig 9 Machine usage over a three week period 240 255 270 285 263 Table 1 Bend stress incurred in the adjustment pin at each weight stack setting Weight stack Ob) 15 30 45 60 75 90 105 120 135 150 I65 180 195 210 225 240 255 270 285 Fatigue bend stress (MW 19 38 58 17 96 115 135 154 173 192 212 23 1 250 269 289 3 08 327 346 366 2.6 Bending... form (4) Consider the use of forging for producing the basic bolt shape Failure Analysis Case Studies I1 D.R.H Jones (Editor) 0 200 1 Elsevier Science Ltd AI1 rights reserved 247 ANALYSIS OF A VEHICLE WHEEL SHAFT FAILURE J VOGWELL Department of Mechanical Engineering, University of Bath, Bath BA2 7AY, U.K (Received 24 April 19 98) Abstract-This paper describes an investigationwhich was carried out on... sensitivity for steels, Fig 8 [ ] In this case the notch root radius was 0 .8 mm (& in.) and the approximate ultimate tensile strength of the steel was 324 MPa as calculated from equation (1) Using these values and Fig 8 produces a notch sensitivity value of y = 0. 58 Using equation (3) and the value for q, the fatigue stress concentration factor (K,) was calculated to equal 1. 38 The fatigue stress concentration... cycles to failure at each stress level Using equation (8) and the values from Table 2 the cumulative damage incurred by one three week loading block was 0.057 Therefore, a total of 17.7 loading blocks would be required to cause failure of the adjustment pin This translates into approximately 53 weeks of typical use before failure According to the purchase records of the machine the actual failure occurred... law (5) to determine the number of cycles to failure at each weight level Using the following constants for AIS1 1005 steel [5]: A E 87 8 MPa, b r -0.13 and equation (5), the number of cycles to failure were calculated at each weight level Table 2 summarizes the equivalent fully reversed stresses using the Goodman equation and the corresponding cycles to failure according to Basquin’s law for each weight... 19 98 Abstract This paper describes a series of examinations and the fracture analysis of burst rubber fuel hosepipes on aero-engines The bursting of the hosepipes was caused by fatigue failure and major reasons are improper design of hosepipe structure, low fatigue bearing capacity and severe deformation in mounting 8 1999 Published by Elsevier Science Ltd All rights reserved K e y w o r k Aircraft failures;... Unicersity of' Illinois at Urbana-Champaign, Department of Mechanical and Industrial Engineering, 1206 West Green SI., Urbana, IL 6 180 1, U.S.A Received 21 August 19 98; accepted 9 September 19 98 ~~ Abstract The following paper is an engineering failure analysis of an adjustment pin used in a leg press exercise machine The pin is used to allow adjustment of the machine for people of different heights It was . 1 983 . 4. Shigley, J. S. and Mischke, C. R., Mechanical Engineering Des@, 5th ed. McGraw-Hill, 1 989 . Sahma Booster Station, Internal Report ME/SQU/July, 1995. Failure Analysis Case Studies. basic bolt shape. Failure Analysis Case Studies I1 D.R.H. Jones (Editor) 0 200 1 Elsevier Science Ltd. AI1 rights reserved 247 ANALYSIS OF A VEHICLE WHEEL SHAFT FAILURE J. VOGWELL. Williamsburg, Virginia, U.S.A. Failure Analysis Case Studies /I D.R.H. Jones (Editor) 0 200 1 Elsevier Science Ltd. All rights reserved 24 1 FATJGUE FAILURE OF HOLD-DOWN BOLTS FOR