Fracture Failure Analysis of AISI 304L Stainless Steel Shaft

Engineering Failure Analysis 36 (2014) 155–165

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Engineerin Engineering g Failure Failure Analysis Analysis j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m / l o c at at e / e n g f a i l a n a l

Fracture failure analysis of AISI 304L stainless steel shaft ⇑

Sh. Zangeneh , M. Ketabchi, A. Kalaki Mining and Metallurgical Metallurgical Engineering Engineering Department, Department, Amirkabir University of Technology, Technology, Tehran, Iran

a r t i c l e

i n f o

Article history: Received Received 8 June 2013 Received Received in revised form 13 September September 2013 Accepted 20 September 2013 Available online 2 October 2013 Keywords: Failure analysis Agitator shaft Finite element analysis AISI 304L stainless steel

a b s t r a c t

Fracture failure analysis of an agitator shaft in a large vessel is investigated in the present work. This analysis methodology focused on fracture surface examination and finite element method (FEM) simulation using Abaqus software for stress analysis. The results show that the steel shaft failed due to inadequate fillet radius size and more importantly marking defects originated during machining on the shaft. In addition, after visual investigation of the fracture surface, it is concluded that fracture occurred due to torsional–bending fatigue during operation.  2013 Elsevier Ltd. All rights reserved.

1. Introduction

In order to mixing of fluids, mechanically stirred vessels are widely used for variety of purposes such as homogenizing single or multiple phases in terms of concentration of components, physical properties, and temperature. Processing during mechanical mixing occurs under either laminar or turbulent flow conditions, depending on the impeller Reynolds number, defined as R e = qND2 /l. For Reynolds numbers below than about R e 6 10, the process is laminar which is called as creeping flow. Fully Fully turbulent turbulent conditio conditions ns are achieved achieved at Reynolds Reynolds numbers numbers higher higher than about about Re P 104, and the flow which has has a Reynolds number between these two regimes would be considered as transitional flow [1] [1].. Typically, a large vessel consists of three main parts: agitator shaft with impeller, top structure with motor and gearbox and fixed vessel to foundation including anchor bolting. The agitator shaft consists of two parts, namely, upper shaft and lower shaft. These two parts are tightly connected connected by means of a rigid coupling to construct the main shaft [2] [2].. General view of the investigated shaft is illustrated in Fig. 1. 1. The agitator shaft was mainly made of an austenitic stainless stainless steel to resist the corrosive media in the vessel. Since shafts are subjected to fluctuating loading of combined bending and torsion with various degrees of stress concentration, the main problem would fundamentally be fatigue loading [3] [3].. In general, shafts are an important component used for power transmission in machinery and mechanical equipment. Failures of such components and structures have engaged scientists and engineers extensively in an attempt to find their main causes and thereby offer methods for their possible prevention [4–7] [4–7].. In this study, failure investigation of an agitator shaft of a typical large vessel was considered. considered. The aim of this study is to understand the main main reason of fracture and avoid the loss of product and time due to shaft failure and happening of similar cases.

2. Experimental procedure

The failed shaft was inspected macroscopically and microscopically by means of optical and scanning electron microscopy (SEM) while a great great care was taken taken to avoid avoid damage damage of fracture fracture surfaces. surfaces. Atomic absorptio absorption n spectrom spectrometry etry was ⇑

Corresponding author. Tel.: +98 918 385 3445. E-mail address: [email protected] [email protected] (Sh. (Sh. Zangeneh).

1350-6307/ 1350-6307/$ $ - see front matter  2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.engfailanal.2013.09.013

156

Sh. Zangeneh et al./ Engineering Failure Analysis 36 (2014) 155–165

Electric motor

Top structure

Upper shaft Flange connection

Lower shaft

Baffles (4 each at 90˚)

Impeller (3 blades)

Fig. 1. General arrangement of the vessel with agitator.

employed to determine chemical composition of the alloy. Room-temperature tensile and impact (Charpy) tests were performed in conformity with the ASTM E-8 and ASTM E-23 standards requirements, respectively. Test-samples for impact (Charpy) toughness are made in compliance with the ASTM-E23 of the size 10  10  55 mm, with the U notch of 2 mm depth and with the 1 mm radius on top of the notch. The tests were made on instrumented Charpy pendulum and the obtained results showed that in all the tested specimens, major fracture energy (85%) was used for crack initiation while its minor portion (15%) was spent on the crack propagation. To determine the hardness of the alloy, Brinell hardness (HRB) measurements were carried out on several points from the outer surface to the central zone of the polished surface of the shaft.

3. Results and discussion 3.1. Chemical composition

According to manufacturer’s documents, the shaft was made from AISI 304L stainless steel. Table 1 shows the atomic absorption spectroscopy analysis which is compared with the related standard [8]. As it can be seen the chemical composition of the failed shaft meets the specified requirement.

3.2. Mechanical properties

Table 2 summarizes the measured results of uniaxial tension tests. Compared to the qualified values, the shaft had a good quality in ductility, the average yield stress 274 ± 15 MPa and ultimate stress 595 ± 10 MPa. The all values of impact toughness were above the minimal required values of 116 J. It has to be mentioned that scattered mechanical data is mainly due to inhomogeneity of the cast material. In addition, the hardness of various sites of the failed shaft was determined and listed in Table 2. Vividly, it can be seen that the measured hardness met the standard ones [9]. Table 1

Chemical composition of failed shaft and AISI 304L stainless steel.

Composition

Carbon

Manganese

Phosphorus

Sulfur

Silicon

Chromium

Nickel

Nitrogen

Iron

Failed shaft 304L stainless steel

0.029 0.03 max.

1.05 2.00 max.

0.035 0.045 max.

0.011 0.030 max.

0.44 0.75 max.

18.1 18.0–20.0

9 8.0–12.0

– 0.10 max.

Balance Balance

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Sh. Zangeneh et al. / Engineering Failure Analysis 36 (2014) 155–165 Table 2

Mechanical properties of failed shaft.

Mechanical properties

0.2% Yield strength (MPa)

Ultimate tensile strength (MPa)

Elongation (%)

Hardness (HRB)

Charpy test (J)

Failed shaft

274 ± 15

595 ± 10

95 ± 3

110 ± 8

116 ± 3

Fig. 2. Fracture surface of the failed shaft at flange connection.

3.3. Fracture surface interpretations

As shown in Fig. 2, the fracture surface has the typical characteristics of a fatigue failure. After crack initiation, propagation of the crack happened over about 3/4 of the shaft cross section which was caused by rotational-bending load. It can be deduced that fatigue is a high-cycle, low-stress type and that propagation occurred under low nominal stresses covering of more than 3/4 of the shaft cross section by fatigue crack [10]. In general, three different locations of fracture were evaluated; surfaces of the shaft, radial zone and final fracture zone. High number of machining grooves which were formed during manufacturing process were revealed by careful macroscopic examination of the exterior surface of the agitator shaft ( Fig. 3a). More surface roughness leads to less fatigue strength due to the fact that valleys of rough surfaces act as stress concentration sites. Under higher magnification, some micro cracks due to improper machining process originated on the surface of agitator shaft ( Fig. 3b). Since a very large fraction of fatigue life is spent in the initiation of crack in high cycle fatigue, fatigue life decrease severely when such micro cracks exist on a shaft [11]. Radial zone contains both many radial coarse lines directed to center and concentric circles about center. In early phases of crack initiation, each crack propagates at different planes ( Fig. 4a) and eventually some steps occur between neighbor planes named ratchet marks. Ratchet marks formation is one of the important factors showed low stress with high stress concentration as illustrated in Fig. 4b. Fracture surface of radial zone with typical fatigue striation is shown in Fig. 5a. Some sections flattened by the rubbing of crack surfaces during the compressive component of the stress cycle are shown in Fig. 5a. Based on Bates–Clark equation (Eq. (1)) [12], stress intensity factor (SIF) during cyclic crack growth can be easily calculated by knowing elastic modulus and striation spacing, which for the shaft case, the values of E and striation spacing (according to p Fig. 5b) were 193 GPa and approximately 1.7 lm, respectively. Thus stress intensity would be calculated 102 MPa min in the failed shaft.

Striation Spacing

2

¼ 6ðDK =E Þ

ð1Þ

To calculate Dr, a simple geometry such as the one shown in Fig. 6a was chosen i.e. a cylindrical shaft with a surface crack with elliptical shape, in a plane perpendicular to the shaft axis (direction of loading). The crack shape is defined by means of the lengths a and b representing the semi-axes of the ellipse ( Fig. 6b). The stress intensity factor for the geometry and mode of loading is:

where

Dr is

  ðp ffiffi ffi ffiÞ

ast

¼ Y

DK

a Dr D

pa

the axial stress range, a is the crack length and Y 

ð2Þ a  D

is a dimensionless function given by following equation:

158


Fig. 3. (a) Macroscopic examination of the exterior surface of the shaft revealed a number of machining grooves and (b) microscopic examination shows micro cracks on the surface of failed shaft.

 a Y

 ¼ D

0:473



a 3:286 D

 þ

a 14:797 D

2

12

a D

a D

2

14

        

:

ð3Þ

The asterisk is used to indicate only one geometric parameter (the crack depth a ) is required, i.e., it is a simplified approach in which the aspect ratio a/b and the curvilinear coordinate s are not considered. This was taken, since a representative geometry (the cylinder with a straight-fronted edge crack) is employed, and the global character of the fracture criterion is achieved, respectively [13]. Therefore, striation fatigue based on SEM observation shown in Fig. 5b was taken nearly 2 mm of shaft surface.   So, a value is nearly 2 mm, and also the diameter of the failed shaft under investigation was ast a 100 mm. As a result, Y D and axial stress range Dr would be 4.14 and 311 MPa, respectively.Final fracture zone has a diameter of about 25 mm. The size of this zone provides information on how big the load. That is, the smaller the size is, the lower the load applied. As it can be clearly seen in Fig. 2, when it comes to the comparison, the size of final fracture zone is smaller than the shaft diameter. Thus, the nominal applied loads are fairly low and the shaft material is highly ductile and tough. Note that because of a distance between the shaft center and zone center, if the shaft is subjected to bending accompanied by torsion, the final fracture zone relocates from center to edge of the shaft. Under higher magnification, one can see the typical dimple features ( Fig. 7a). Carbide/matrix interface which is shown in ( Fig. 7b), was very weak contributed to interfacial separation. In this case, carbides were responsible for the void nucleation. 3.4. Calculation of torque acting on the shaft

Torque calculation on the shaft during process determined through delivered power to a given fluid at a constant rotational speed [2]. The power consumed by a mixer can be obtained by the following equation: P

¼



N p qn3 D g c

5

!

ð4Þ

Sh. Zangeneh et al. / Engineering Failure Analysis 36 (2014) 155–165

159

a

Cracks initiation zones

b

Ratchet Marks

Fig. 4. (a) Cracks initiation zones on flange connection and (b) ratchet marks.

Fig. 5. (a) Striation fatigues at fracture surface, some flattened surface indicated by red arrow and (b) striation spacing. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)

160


a D

b a 2b

Fig. 6. Geometry of analysis: (a) 3D view of the shaft and (b) 2D sketch of the cracked section.

where q is density of fluid (lb/ft 3), D a is impeller diameter (ft), n is the speed of impeller (rps), g is Newton’s law conversion factor and N p is power number which depends on Reynolds number ( N Re), Froude number ( N Fr ) and impeller type based on following equation:

N p

¼



2

nDa q n2 Da ; ; S 1 ; . . . ; S 9 / l g

!

ð5Þ

As given in Eq. (5) the power number, N p, also is a function of ratio of tank to impeller diameter ( T /D), height of impeller above vessel floor to impeller diameter ( Z /D), length of impeller blade to impeller diameter ( L/D), impeller diameter to width of blade ( D/W ), tank diameter to baffle width ( T /B), liquid depth in vessel to impeller diameter ( H /D), number of impeller blades (n), pitch/angle (Degree) and number of baffles ( n), all of which included to the above equation as shape factors S 1 ,. . .,S 9 respectively. In general, if Reynolds number goes over 10 4, based on available information [2] N p will be constant and independent of liquid viscosity. Furthermore, changes in N Fr have not any significant effect on N p. Therefore, Eqs. (4)(5) can be rewritten as independent of N Re and N Fr .

N p

¼ n Pg ¼ /ðS ; . . . ; S Þ D q 3

c 5 a

1

9

ð6Þ

Regarding the available information [1], power number for three flat blades is 2.58. For pitched blade turbines, changing the blade angle h would change the power number by N p  (sinh)2 [2]. As a result, when blades make a 45  angle with shaft, N p is 0.4 times the value when blades are parallel with a shaft, which means the power number will be changed to 1.032 based on S 8 shape factor. Some necessary data taken from the specification sheet of vessel are given in Table 3. The torque can be calculated 372 NM in accordance to P = xT = 2 pnT .

161


Fig. 7. (a) Final fracture zone of the failed shaft shows the typical dimple features and (b) interfacial separation of carbide/matrix interface.

Table 3

Necessary data extracted from the specification sheet of vessel.

Definition

Impeller diameter

Viscosity

Density

Acceleration of gravity

Speed of impeller

Values

3.44 ft

1.34  103 lb/ft s

69 lb/ft3

32.17 ft/s2

1.17 rps

3.5. Stress analysis by Abaqus program

In this study, the commercial Abaqus software was employed to analyze stress propagation at start-up regime of the shaft during mixing operation. The 3D geometry model of the agitator shaft is shown in Fig. 8a. As shown, all parts such as shafts, flange connection and bolts (Fig. 8b) used to fasten flange were modeled and assembled to this analysis. Mechanical behavior of shaft material was imported into the software by calibration tool in order to use exact data for the simulation. Bolt loads defined by concentrate force in their length to fasten flanges. In order to define the contacts among bodies and position them relative to each other in only one global coordinate system in which shafts and bolts would be identical with practical places in the process, contact surfaces and contact pairs are also defined. Before performing the simulation, several assumptions were made: (1) shaft weight was not to be taken into consideration, since during mixing operation, the axial (upward) forces created by impeller due to blade angle of 45  are opposite to the shaft weight, (2) drum pressure acts as a hydrostatic pressure equally over the shaft and has no effect on the general yielding, thus, it helps to increase the yield of the shaft material. (3) operation temperature (216  C) is a moderate temperature for the shaft material and at this temperature the properties of the material do not change considerably but ductility and fracture toughness increase moderately at high temperatures up to around 350  C and (4) the material is assumed to be isotropic and homogeneous. In order to study numerical simulations of the agitator shaft, hexahedral brick elements C3D20R which are the quadratic reduced-integration elements were adopted. These elements are not susceptible to shear locking, even when subjected to complicated states of stress. Therefore, these elements are also generally the best choice for the most general stress/displace-

162


Fig. 8. (a) 3D geometry model of the agitator shaft and (b) intake flange connection.

Fig. 9. Stress contours at different stages of the analysis during the start-up of the shaft calculated according to the Von Mises’ criterion in the incremental times of (a) 10 4, (b) 3  10 4, (c) 7  10 4 and (d) 103 s, respectively.

ment simulations. The mesh is graded in a way such that there is a higher mesh density at shaft to flange connection. This improves the accuracy of the solution around the shaft without tremendously increasing the computational time. Mesh in the all parts generated by sweep method. Abaqus/CAE creates swept meshes by internally generating the mesh on an edge or


163

Fig. 10. Stress contours at different stages of the analysis during the start-up of the shaft with a single machining groove with a depth of 0.2 mm calculated according to the Von Mises’ criterion in the incremental times of (a) 10  4, (b) 3  10 4, (c) 7  10 4 and (d) 10 3 s, respectively.

face and then sweeping that mesh along a sweep path. Normally, a higher mesh density provides for higher accuracy but also increases the computational time, therefore, a trade-off between time and accuracy becomes crucial. In this case, three different mesh densities were investigated. The shaft were initially meshed with 7855 elements with a higher mesh density closer to the shaft to flange and then mesh density changed from 10,165 to 11,705 in order to reach successful convergent. After obtaining the torque (372 NM), it was applied to a reference point 2 (Rp-2) which was coupled to lower position of the shaft (impeller location). Also, constant rotating speed of 70 rpm defined to the reference point 1 (Rp-1) in the upper position of the model (electric motor location). In addition, some constraints were applied for conformity to the real shaft operation. Dynamic analysis in Abaqus/Standard uses implicit time integration method to calculate the transient dynamic response of a system. These time integration operators are implicit, which means that the operator matrix must be inverted and a set of simultaneous nonlinear dynamic equilibrium equations must be solved at each time increment. This solution is done iteratively using Newton’s method [14]. Fig. 9a–d shows stress contours at different stages of the analysis during the start-up of the shaft calculated according to the Von Mises’ criterion in the incremental times of 10 4, 3  104, 7  104 and 103 s, respectively. As can be seen in Fig. 9a–d, stress waves propagated into the upper shaft in a very short period of time, however due to abrupt change in the cross section of shaft to flange, stress raised at shaft/flange connection until the flow stress reached to the bolts and then the stress relaxed by flowing through the bolts to the lower shaft caused the shaft to rotate. Although stress highly increased up to 101 MPa at the location of crack initiation at shaft to flange connection in a short period of time (7  104 s), this can be neglected by considering mechanical properties of the shaft. In general, fracture of the shaft originates at points of stress concentration either inherent in design or introduced during fabrication or operation. In this case, the design features that cause concentrated stress is inadequate fillet radius size. Furthermore, stress concentration produced during fabrication as a result of machining grooves. Depending on the variables of machining technique and the quality of the cutting tool, machining can lead to fairly sharp grooves. To understand the effect of machining grooves on stress concentration, a single machining mark with a depth of 0.2 mm was considered based on roughness measurements. Surface roughness used here was based on the difference between the highest peak and the lowest one. Fig. 10a–d shows stress contours at different stages of the analysis along with a single machining mark with a depth of 0.2 mm at start-up of the shaft calculated according to the Von Mises’ criterion in the incremental times of 10 4, 3  104, 7  104 and 103 s, respectively. Fig. 11a–b shows the effect of 0.2 mm machining groove on stress concentration during the start-up of the shaft at the time of 7  104 s at higher magnification. Also, it can

164


Fig. 11. Effect of a single machining groove with a depth of 0.2 mm in a start-up of the shaft at the time of 7  10 4 s (a) side view and (b) root of machining mark.

Fig. 12. Stress contours at steady state operation of the shaft calculated according to the Von Mises’ criterion.


165

be seen that such a circumferential groove increased stress concentration to as much as 50% the smooth one at start-up time. The number of such machining marks with the size of more than 0.2 mm originated during fabrication process ( Fig. 3a), so each one play a crucial role in decreasing fatigue life and eventually contributed to failure of the shaft. In general, the major defects that originate during machining are machining grooves, grinding burns and cracks. Machining grooves are extremely dangerous in critical rotating parts, with life limited by fatigue. Because of their stress-raising effect, machining defects are the preferred sites for fatigue crack initiation. A large number of service failures occur due to improper machining. Static analysis in Abaqus/Standard was carried out to understand stress distribution during steady state operation of the shaft. Fig. 12 depicts stress contours in the assembled agitator shaft. As it is clear due to inadequate fillet radius size, stress raised at shaft/flange connection, but the obtained maximum stress, i.e. 6 MPa, is not comparable with the one occur during startup. Therefore, an increase in stress at shaft/flange connection might occur mostly in the service during startup or other transient conditions. 4. Conclusion

Experimental and numerical simulation of the agitator AISI 304L stainless steel shaft was studied. Results showed that mechanical properties (uniaxial tension, impact and hardness tests) and chemical composition of the failed shaft were in acceptable range. Calculations showed that torque applied to the shaft during mixing operation at constant speed of 70 rpm was approximately 372 NM. Stress analysis based on the boundaries conditions mentioned earlier showed that in the smooth shaft at 7  104 s of starting time, stress raised up to 101 MPa within shaft to flange connection. This value in the shaft with a single machining groove with depth of 0.2 mm increases as much as 50%. In addition, Static analysis showed that stress rose at flange/shaft connection up to 6 MPa in steady state operation. On the basis of all the above stated, it can be concluded that inadequate fillet radius size and more importantly machining grooves on the shaft surface was the main cause of the shaft failure. References [1] Perry RH. Chemical engineering handbook. 7th ed. New York: McGraw-Hill; 1997. [2] Paul EL, Atiemo-Obeng VA, KrestaSM. Handbook industrial mixing science and practice. Hoboken (New Jersey): John Wiley & Sons; 2004. [3] Vincent L, Roux JL, Taheri S. On the high cycle fatigue behavior of a type 304L stainless steel at room temperature. Int J Fatigue 2010;38:84–91. [4] Lancha AM, Lapena J, Serrano M, Gorrochategui I. Metallurgical failure analysis of a BWR recirculation pump shaft. Eng Fail Anal 2000;7:333–46. [5] Bhaumik SK, Rangaraju R, Parameswara MA, Venkataswamy MA, Bhaskaran TA, Krishnan RV. Fatigue failure of a hollowpower transmission shaft. Eng Fail Anal 2002;9:457–67. [6] Sattari-Far I, Moalemi M. Failure of stainless digester shafts in a paper production plant. Eng Fail Anal 2003;10:675–82. [7] Fahir Arisoy C, Basman Gokhan, KelamiSesen M. Failure of a 17-4 PH stainless steel sailboat propeller shaft. Eng Fail Anal 2003;10:711–7. [8] ASTM A 276–92, Specification for stainless and heat-resisting steel and shapes. [9] Courtney TH. Mechanical behavior of materials. 2nd ed. NewYork: McGraw-Hill; 2000. [10] Jian Ping J, Guang M. Investigation on the failure of the gear shaft connected to extruder. Eng Fail Anal 2008;15:420–9. [11] Dieter GE. Mechanical metallurgy. 3nd ed. New York: McGraw-Hill; 2001. [12] Bates RC, Clark WG. Fractography and fracture mechanics. Trans ASM 1969;62:380–9. [13] Athanassiadis A, Boissenot JM, Brevet P, Francois D, Raharinaivo A. Linear elastic fracture mechanics computations of cracked cylindrical tensioned bodies. Int J Fract 1981;17:553–66. [14] Harewood FJ, McHugh PE. Comparison of the implicit and explicitnite element methods using crystal plasticity. Comput Mater Sci 2007:481–94.

Fracture Failure Analysis of AISI 304L Stainless Steel Shaft

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