Pressure vessel lug design

7/27/2019 Pressure vessel lug design

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Journal ofPressure Vessel

Technology Design Innovations

Proposed Design Criterion for Vessel

Lifting Lugs in Lieu of ASME B30.20

Dennis K. WilliamsSharoden Engineering Consultants, P.A.

P.O. Box 1336,

1153 Willow Oaks Trail,

Matthews, NC 28106-1336

e-mail: [email protected]

This paper describes a method for evaluating the structural ad-equacy of various lifting lugs utilized in the erection and up right-ing of large pressure vessels. In addition, the analysis techniquesare described in detail and design guidelines for vessel lifting aretendered. The statutory and provincial regulations in both theUnited States and the province of Alberta, Canada are also re-viewed and discussed with respect to the too often utilized phrasefactor of safety (FOS). The implied implications derived fromthe chosen FOS are also outlined. A discussion is presented as tothe applicability of the ASME safety standard B30.20 entitled,Below the Hook Lifting Devices (1999, ASME, New York) andas to the severe shortcomings of the safety standard in its attemptto delve into the design of lifting devices, especially when appliedto lifting lugs on large and heavy-weight pressure vessels. Exem-

plar lugs on vessels are defined and the finite element analysesand closed form Hertzian contact problem solutions are presentedand interpreted in accordance with the proposed design criteria.These results are compared against the very limited design infor-mation contained within ASME B30.20. Suggestions for the revi-sion and applicability of the safety standard are presented anddiscussed in light of the examples and technical justification pre-sented in the following paragraphs. In addition, the silence of thissafety standard on the very large contact stresses that are wellknown to exist between a lifting pin and clevis type geometry isalso discussed. Because of the limited number of repetitive load-ing cycles that vessel lifting lugs actually experience during theservice life of a vessel, a recommendation is made to either clearlyexclude vessel lifting lugs from the scope of ASME B30.20 or tospecifically include a separate design and analysis section withinthis standard to properly address the mechanical and structuraldesign issues applicable to pressure vessel lifting lugs.

DOI: 10.1115/1.2716439

Introduction

The basic approach of the current study is to first define themode of failure against which any design criteria and/or standard

must try to protect during the design phase of a vessel lifting lug.

Second, the approach selects a failure criteria from those com-

monly discussed in the literature, such as maximum shear stress

theory, maximum octahedral shear stress theory, and maximum

principal stress theory, that most closely matches the mode of

failure defined in the first step of the basic approach. Third, the

method adopts an achievable factor of safety based on the cho-

sen failure theory from step two of the approach and applies the

FOS against the respective failure stress. Finally, a well-defined

design criterion for vessel lifting lugs is outlined based on the

basic approach presented herein and applied to the statutory and

provincial regulations contained within 29CFR1926 OSHA regu-

lations 1 and the Occupational Safety and Health Act of theProvince of Alberta, Canada 2 .The purpose of this paper is to provide technical insight into the

applied mechanics evaluation of an exemplar pressure vessel lift-

ing lug and proposed design criteria in lieu of the limited design

requirements contained within ASME B30.20 3 . The currentwork is restricted to the evaluation of the lug in the vicinity of thelifting pin. The subject lifting lug of this paper is one whose

design load capacity is 700 metric tons 1,544,000 lbf . Thisload capacity is not uncommon in the petroleum refinery industryfor a number of specialty types of ASME B&PV Code, Section

VIII 4 reactor vessels. On many project designs, there are atleast two types of imposed design bases for the lifting and han-dling equipment. The first of these design bases is an internallygenerated or self-imposed design basis. The second of these de-

sign bases is one that may be classified as externally generateddesign basis. For purposes of this discussion, internal and ex-ternal refer to an organization within the design engineering or-ganization hence internal or to an outside authority having juris-diction hence external .

The internal design basis for the lifting lug can further be de-fined by either internally generated design and analysis criteria or

by externally generated codes and safety standards. The internallygenerated criteria most often attempt to define an allowable setof component stresses that restrict the computed bending, bearing,and shear stresses within the lifting lug critical sections as deter-

mined by both experience and empirical data. Although there areno uniform set of criteria among the numerous engineering designprofessionals throughout the U.S. and Canada, it is this authorsexperience that one guideline, which is often quoted, is the limi-

tation of the bending stress to one third of the yield strength of thelug material assuming a one-piece forged design . The additionalcomponent stresses and the associated allowable stressess varywidely across-the-board, depending on the particular design engi-

neering group and their given experience. The externally gener-ated or imposed design standard on lifting lugs often falls on thelimited criteria contained within the ASME safety standard

B30.20 3 entitled, Below the Hook Lifting Devices. This pa-per addresses the fallback position employed by many engineer-ing organizations in attempting to utilize the design criteria con-tained therein and the inherent pitfalls of such a practice,

particularly when applied to large lifting lugs for pressure vessels.The second major group of design basis criteria originates from

Contributed by the Pressure Vessel and Piping Division of ASME for publication

in the JOURNAL OF PRESSURE VESSEL TECHNOLOGY. Manuscript received January 23,

2006; final manuscript received September 7, 2006. Review conducted by David Raj.Paper presented at the 2002 ASME Pressure Vessels and Piping Conference

PVP2002 , Vancouver, British Columbia, Canada, August 59, 2002.

326 / Vol. 129, MAY 2007 Copyright 2007 by ASME Transactions of the ASME

wnloaded From: http://pressurevesseltech.asmedigitalcollection.asme.org/ on 10/05/2013 Terms of Use: http://asme.org/terms


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an outside source and most often takes the form of regulatoryrequirements defined by federal, state, and/or provincial authori-ties having jurisdiction. For the purposes of this paper, only theregulations imposed in the U.S. and the Province of Alberta,Canada are discussed, although the subject discussion is one thathas universal implications. This becomes even more apparent asfabricators across the globe are providing ASME Code 4 pres-

sure vessels intended for installation in both the U.S. and Canada.The specific design lifting requirements imposed by the U.S. De-partment of Labor, 29CFR1926 OSHA regulations 1 are out-lined and the implementation of a 5 to 1 factor of safety FOSis also explored. Finally, the Provincial Regulations containedwithin the Occupational Safety and Health Act of the Province ofAlberta, Canada 2 are also reviewed and discussed with furtherinterpretation of a very specific five-to-one factor of safety on theultimate tensile strength of the chosen lug material.

The analysis of the 700 metric ton lifting lug is conducted intwo separate manners. The first is the analysis and prediction ofthe contact stresses, the general stress field within the critical sec-tions of the lug, and the associated contact area between a closetolerance fitting lifting bolt/pin utilizing the techniques presentedby Timoshenko and Goodier in the Theory of Elasticity 5 andsummarized for direct application by Young

6

. The lug geometryitself is a simple clevis type design of uniform thickness as shownin Fig. 1. The second analysis utilizes the finite element method toanalyze the effects of the contact stresses within the lug and alsoutilizes the calculated results to serve as a basis for the proposedlug design criteria. The proposed lug design criteria must achievetwo goals. The first is to ensure a safe and practical design. Thesecond is to provide a design that is in full compliance with thedefined regulatory requirements defined herein. The technicalbases for achieving this two-fold objective are outlined in theparagraphs that follow.

ASME B30.20: Below the Hook Lifting Devices

ASME B30.20 had its beginning in December 1916, when an

eight-page Code of Safety Standards for Cranes was presented tothe annual meeting of the ASME. Because of changes in design,advancement in techniques, and the general interest of labor andindustry in safety, an American National Standards Committeewas formed under the joint sponsorship of ASME and Naval Fa-cilities Engineering Command ASME 3 . According to theforeword provided within ASME B30.20 3 , the Standard pre-sents a coordinated set of rules that may serve as a guide to gov-ernment and other regulatory bodies responsible for the guardingand inspection of the equipment falling within its scope. Further-more, the foreword to the referenced ASME Standard 3 urgesadministrative and regulatory agencies to consult the B30 Com-mittee prior to rendering decisions on disputed points. The fore-

word concludes with the protective remark that, revisions to theStandard do not imply that previous editions were inadequate.

Within the Introduction to the ASME Standard 3 , the Stan-dards Committee states that they fully realize the importance ofproper design factors. In addition, the Standards Committee statesthat, they will be glad to receive criticisms of this Standardsrequirements and suggestions for its improvement, especiallythose based on actual experience in application of the rules 3 .The purpose of this paper is to explore the design requirementscontained in the Standard 3 and to provide an exemplar of theapplication of the suggested rules and some well defined and

quantitative criticisms of the Standard.Within the scope of ASME B30.20 3 is the construction oflifting devices. Some engineers interpret the 700 metric tons lift-ing lug under consideration to fall within the scope of the subjectStandard 3 . Although Chapter 20-1 entitled Structural and Me-chanical Lifting Devices 3 does not specifically categorize thelifting lug under consideration, many design organizations citethis chapter for use as a design guideline or requirement. In par-ticular, paragraph 20-1.1.1 entitled General Construction, out-lines the design requirements for a lifter as follows:

The load bearing structural components of a lifter shall bedesigned to withstand stresses imposed by its rated load plus theweight of the lifter, with a minimum design factor of three, basedupon yield strength of the material, and with stress ranges that donot exceed the values given in ANSI/AWS D14.1 [7] for the appli-cable conditions [3].

Before proceeding, there are several observations that must behighlighted concerning the preceding design requirement. First,the placement of a design requirement under the heading ofConstruction is not consistent with the organization of manyother ASME Codes and Standards 4 in that design criteria areclearly labeled as such and are also segregated from the construc-tion i.e., fabrication requirements. Second, the design require-ment contained within the referenced Standard 3 only gives theengineer a very vague idea at best as to which particular com-puted stresses must be compared against essentially the materialsyield stress divided by three. Third, the choice of an allowablestress of sorts by the Standard 3 that is based on yield impliessome sort of failure criteria that would preclude the initiation ofyielding within the lug material when subjected to its maximumrated load. Fourth, the selection of a yield based failure criteria

would most likely also imply that the anticipated failure modewould be one of a ductile nature of course assuming the selectionof a linear, elastic, homogeneous, and isotropic lug material .Fifth, the utilization of a large capacity vessel lifting lug willtypically only be subjected to a very limited number of liftcycles, as once the vessel is moved from the fabricators shopand uprighted in the field, it generally is there for the remainder ofits intended design life. Finally, the subject design problem clearlyinvolves contact stresses between a shackle pin and a clevis hole,which obviously implies very large contact stresses that are highlylocalized, which are not addressed by this or any other standardknown to this author. It is these implications in combination withother regulatory requirements that are explored in some detail inthe paragraphs that follow.

OSHA Rigging Equipment Regulations in the United

States

The U.S. Department of Labor, through the regulations speci-fied by its Occupational Health and Safety Administration

OSHA , specifies the general safety requirements for riggingequipment for material handling. Throughout the U.S. Code ofFederal Regulations CFR , numerous references are made toslings, wire rope, hooks, shackles, and other forms of materialhandling and other forms of lifting equipment 1 . In particular,the regulations that are of most importance during the uprightingand lifting of a pressure vessel are found in Title 29, Part 1926,

Fig. 1 Lifting-lug geometry

Journal of Pressure Vessel Technology MAY 2007, Vol. 129 / 327



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Subpart 251 entitled Rigging Equipment for Material Handling 1 . This and other OSHA Regulations may now be easily foundon the Internet at URL http://www.osha.gov.

Although there are numerous similarities between ASMEB30.20 3 and 29CFR1926.251 1 regarding the inspection ofthe lifting equipment including, but not limited to lifting lugsprior to use, there exists a striking difference in the design cri-teria for the handling devices. The OSHA requirements specifi-cally define that the safe working loads of shackles and specificidentifiable products be designed with a safety factor of not lessthan 5, although the safety factor is not explicitly defined therein

1 . The FOS of 5 is repeatedly echoed throughout this and otherOSHA Regulations regarding handling equipment.As in the proceeding section, there are several observations that

must be discussed concerning the preceding design requirement.First, the OSHA design requirement 1 gives the engineer nosuggestion or proposal as to which particular computed stressesmust be compared against essentially the materials ultimate ten-sile stress divided by five. This assumes that the failure load,however, is proportional to the ultimate tensile failure stress orvice versa in the application of the FOS as recommended byBoresi et al. 8 . Second, the choice of an allowable stress ofsorts by the Regulation 1 , which is based on what is assumed tobe a FOS of 5 applied to the ultimate tensile strength, impliessome sort of failure criteria that would be based on a predomi-nant component stress or computed maximum principal stresswhen subjected to its maximum rated load. Third, the selection ofan ultimate tensile strength based failure criteria would mostlikely also imply that the anticipated failure mode would be one ofa brittle nature even though the lug material is most assuredlyclassified as a ductile material under ordinary loading conditions .Fourth, similar to the parley on ASME B30.20, the utilization of alarge-capacity vessel lifting lug will typically only be subjected toa very limited number of lift cycles, for the reasons previouslycited. Finally, the subject design problem clearly involves contactstresses between a shackle pin and a clevis hole, which obviouslyimplies very large contact stresses that are highly localized, whichare not addressed by this or any other regulation known to thisauthor.

Alberta Safety and Health Act Requirements

The Province of Alberta, Canada, through the regulations speci-fied by its Occupational Health and Safety Act, Regulation 448/83

2 , specifies the general safety requirements for rigging equip-ment for material handling. Similar to the OSHA Regulation 1previously discussed, within the Alberta Regulation 2 , numerousreferences are made to slings, wire rope, hooks, shackles, andother forms of material handling and other forms of lifting equip-ment. In particular, the regulations that are of most importanceduring the up-righting and lifting of a pressure vessel are found inPart 8, entitled, Rigging, paragraph 139 2 . This and other Al-berta Regulations may now be easily found on the Internet at URLhttp://www.gov.ab.ca

Once again, there are numerous similarities between ASMEB30.20 3 , 29CFR1926.251 1 , and Alberta Regulation 448/83

2 AR 448/83 regarding the inspection of the lifting equipment

including, but not limited to lifting lugs prior to use. The AlbertaRegulation, however, begins to define the design criteria of thehandling devices more specifically than the OSHA regulations andtakes on a different basis for the FOS than even ASME B30.20

3 . AR 448/83 requirements specifically state that the maximumsafe working load of rigging or rigging equipment must not ex-ceed20% of the ultimate breaking strength of the weakest com-ponent of the rigging. As with the OSHA regulations, a FOS of 5appears and is repeatedly echoed throughout this Regulation 2regarding handling equipment.

The pertinent observations that must be discussed concerningthe preceding design requirement are as follows. The design re-quirement 2 gives the engineer the latitude to formulate a failure

criteria based on the ultimate breaking strength, which suggestsat least an ultimate tensile strength divided by five allowablestress. Again however, this assumes that the failure load is propor-tional to the ultimate tensile failure stress in the application of thefive-to-one FOS. Second, the choice of an allowable stress ofsorts by the Regulation 1 that is based on what is assumed to bea FOS of 5 applied to the ultimate tensile strength implies somesort of failure criteria that would be based on a predominantcomponent stress or computed maximum principal stress whensubjected to its maximum rated load. Third, the selection of anultimate tensile strength based failure criteria would most likely

also imply that the anticipated failure mode would be one of abrittle nature. Fourth, similar to the discussion on ASME B30.20,the utilization of a large capacity vessel lifting lug will typicallybe subjected to only a very limited number of lift cycles, for thereasons previously cited. Finally, as with all the other regulationsand standards previously cited, the very large contact stresses,which are highly localized, are not addressed by this regulation aswell.

Classical Contact Stress Evaluations

The problem presented by the design of a vessel lifting lug isfirst and foremost one created by the Hertzian contact of a liftingpin supported by a shackle in this case with that of the liftinghole in a clevis-type lifting lug. The resulting stresses of the great-est magnitude are indeed those created by the pressure exerted by

the pin on the clevis over a limited area of contact. The Hertziancontact stress problem is treated and discussed in detail in Boresi

8 Chap. 14 , Timoshenko 5 and Young 6 . It is these contactstresses that are not, however, addressed in any of the Standardsor Regulations previously discussed in the preceding paragraphsof this paper 13 . The only plausible justification for the silenceby these Standards and Regulations on the obvious existence ofthe contact stresses may be as explained in the following byBoresi 8 :

Most load resisting members are designed on the basis ofstress in the main body of the member, that is, in portions of thebody not affected by the localized stresses at or near a surface ofcontact between bodies. In other words, most failures (excessiveelastic deflection, yielding, and fracture) of members are associ-ated with stresses and strains in portions of the body far removed

from the points of application of the loads.Appendix A contains a sample calculation for the

700 metric ton lifting lug contact stress evaluation. The geometricand material properties are defined within the calculation proper.The chosen material for the lifting lug is ASME 508 Grade 3

Class 2 with an ultimate tensile strength of 90,000 psi and yield

strength of 65,000 psi. The lifting lughole diameter is 8.543 in.

within which a pin diameter of 8.460 in. must be inserted. The

thickness of the lifting lug is 11.50 in.The relatively simple evaluation of the contact stresses con-

tained in Appendix A is not applicable close to the edges wherethe contact boundary begins. This sharp interface area can pro-

duce highly localized stresses in excess of 100,000 psi 8 for the700 metric tons loading under consideration. The depths at whichthe effects of the contact stresses are significant range from thesurface of contact

for the maximum principal stress

to an ap-

proximate depth of 1.6 in. for the maximum shear stress and themaximum octahedral shearing stress . Nevertheless, the results ofthe contact stress evaluation contained in Appendix A reveal atleast two important pieces of information. The first is that theaverage contact stress for the loading, geometry, and materials of

construction defined herein are 40,000 psi. Keeping in mind thatthis is a compressive stress and that when compared to a yield

strength of at least 65,000 psi, this is far below those levels ofbearing and/or compressive stress allowed within other ASMECodes 4 . The second important piece of information is the cal-culated width of the assumed rectangular contact area. For the

subject design, this width is calculated to be 4.35 in., which

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equates to a half angle from top dead center of 30 deg for thecontact area. This value is and can be utilized in a more detailedfinite element analysis of the lifting lug and represents the angulardimension over which the lifting load may be applied.

Classical Failure Theories and the Mystery of the Uli-

mate Factor of Safety

Clearly in the design and analysis of a lifting lug for a vessel

that weighs 700 metric tons, the engineer must have an excep-tional understanding of the possible ways by which the lug mayfail to perform its function. In determining the possible modes of

failure, the engineer must also establish the failure criteria bywhich the design will be judged. In the present study, the modesof failure for the lug must include not only the common staticcauses, such as bending, shear, and bearing, but also the effects ofa dynamic or shock loading due to the lift itself. It is this consid-eration of a potential dynamic load generated as a result of thesudden loss of tension in a cable or sling that drives the need forsome form of a FOS to be employed to the applied static load orresulting computed stress. Furthermore, the selection of a failurecriteria for the lifting lug must be predicated based on this dy-namic load consideration and not the initiation of yield per se, asthe lug will only be utilized a very few number of times in itsdesign life i.e., certainly no more than ten times . As will beshown, it is not only the mere specification of a single FOS butalso the failure criterion that determines the ultimate factor of

safety of a given lifting lug design.Boresi 8 states, There is considerable but not necessarilyconclusive evidencethat when a member fails by general yield-ing at ordinary temperatures, the significant quantity associatedwith the failure is shearing stress. Two of the most widely uti-lized failure criteria that address general yielding include themaximum shearing stress Trescas criterion and the maximumoctahedral stress von Mises criterion theories. A third criterion tobe considered is the maximum principal stress theory of failure

Rankines criterion . Before proceeding, however, the engineermust remember that in a uniaxial state of stress, the critical fail-ure values for each of the defined theories are achieved simulta-neously in a simple tensile test.

Trescas criterion states that inelastic action in any point in apart initiates when the computed maximum shearing stressreaches one-half of the materials yield strength. Although themaximum shear stress failure theory is best suited for ductile ma-terial behavior in which relatively large shearing stresses are de-veloped 8 , for the problem at hand, the maximum shearing stresswould be limited to no more than 32,500 psi prior to the applica-tion of any FOS.

The maximum octahedral shear stress failure theory states thatinelastic action in any point in a part initiates when the computedoctahedral shearing stress reaches 0.471 times the materials yieldstrength. In many ductile materials utilized within the pressure

vessel industry, the octahedral shearing stress criterion predicts theinitiation of yield better and with less conservatism than does theTresca criterion. For the problem at hand, the maximum octahe-

dral shearing stress would be limited to no more than 30,641 psiprior to the application of any FOS.

The maximum principal failure theory is one that may be easilyemployed to establish either the initiation of yield for a brittlematerial or one that may be employed to establish a guard againstfracture through the use of the ultimate tensile strength UTS andan imposed FOS. In fact, even though the chosen material for avessel lifting lug should be one of high ductility and possess goodimpact properties, the failure mechanism due to a sudden accel-eration i.e., a dynamic load in most cases will be one of a brittlenature. Therefore, Rankines criterion, when modified to utilizethe UTS, provides both an easy to use criterion and one that isconsistent with the expected mode of failure. For the problem at

hand, the maximum principal stress would be limited to 90,000 psidivided by the chosen FOS. This also implies that the minimumprincipal stress must be addressed as a separate matter. This isbecause the failure mode that must be guarded against is one, firstand foremost, that would be tensile in nature and tend to open anypreexisting cracks in the lifting lug material.

Criteria for Consideration in the Design and Analysis of

Pressure Vessel Lifting Lugs

In an effort to more fully understand the stress field within the

700 metric tons lifting lug, a finite element analysis was per-

Fig. 2 Lifting-lug finite element mesh




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formed. The lug geometry was as shown in Fig. 1, and the refinedelement mesh was as shown in Fig. 2. The base of the lug wasfixed against translation in all three coordinate directions, as theelements chosen to model the geometry were a three-dimensionalsolid with three-degrees of freedom at each node i.e., translationsin the x, y, and z directions . The 700 metric tons load was evenly

distributed within a 30 deg half angle from top dead center oneach side of the symmetry plane and continued through the thick-ness of the lifting lug. The chosen contact area was confirmedindependently by the calculations presented in Appendix A of this

paper, which was previously discussed. Several mesh densitieswere employed until the stress results converged to within 3% ofthe more course mesh density.

The results of the finite element analysis were decomposed intoall of the constituent component stresses. In addition, the resultswere also reviewed in light of the three failure criteria previouslydefined and outlined above. The stress contours showing the cal-culated maximum stress intensities i.e., twice the value of themaximum shear stresses , the von Mises stresses and the maxi-mum principal stresses are included in Figs. 35, respectively. Allof the contours reflect a highly concentrated respective combinedstress at the geometric discontinuity between the lug hole bore andorthogonal outside surface. This is both attributed to the contactload discontinuity and the reality of the geometric/load combina-tion. This area is very small and is not anticipated to reflect theoverall load carrying capacity of the lug, regardless of the failurecriterion employed.

The calculated maximum shear stress as shown in Fig. 3 is40,000 psi i.e., one-half of the calculated stress intensity .When compared to the Tresca criterion allowable stress of one-

half of yield i.e., 32,500 psi , this represents an overage of25%without employing any factor of safety. As the distance from theapplied contact load increases, the stress contour reveals that anoverall maximum shear stress throughout the body of the lifting

lug quickly decreases to a value of 13,40017,800 psi. In orderto fully evaluate the FOS for this criterion, these values must becompared to one-half of the yield, which results in a FOS of aslow as 1.82 on the initiation of yield a short distance away from

the proximity of the contact load application. It is recognized that

the subject lifting lug design would have a much higher load

carrying capacity than that predicted by this criterion due to the

strain hardening capacity of the chosen forging material.

The calculated equivalent stress i.e., von Mises as shown inFig. 4 is 69,194 psi. When compared to the equivalent allow-able stress of 1.0 times yield i.e., 65,000 psi , this represents anoverage of 6.5% without employing any factor of safety. As thedistance from the applied contact load increases, the stress contour

reveals that an overall maximum equivalent stress throughout the

body of the lifting lug quickly decreases to a value of23,10030,800 psi. In order to fully evaluate the FOS for thiscriterion, these values must be compared to the one times yield,

which results in a FOS of as low as 2.11 on the initiation of yield

a short distance away from the proximity of the contact load ap-

plication. As before, it is recognized that the subject lifting lug

design would have a much higher load carrying capacity than that

predicted by this criterion due to the strain hardening capacity of

the chosen forging material.

The calculated maximum principal stress as shown in Fig. 5 is

53,276 psi. When compared to the equivalent allowable stressof 1.0 times yield i.e., 65,000 psi , this represents a margin of18% without employing any factor of safety. As the distancefrom the applied contact load increases, the stress contour reveals

that an overall maximum principal stress throughout the body of

the lifting lug quickly decreases to a value of 16,400 psi. Inorder to fully evaluate the FOS for this criterion, these values may

be compared to the one-time yield, which results in a FOS of

3.96 on the initiation of yield. As stated previously however,this is not a good predictor of yield and is most well suited for use

with the ultimate tensile strength in this case. Proceeding on this

basis, these values must be compared to the UTS, which results in

a FOS of 5.49 on the ultimate failure load assuming totallyelastic response, which is a conservative predictor . In contrast,

the calculated minimum principal stress, as shown in Fig. 6, is

found to be approximately 61,292 psi, which is reflective of the

compressive contact stress between the lifting pin and lug hole.

Fig. 3 Stress-intensity contour




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Again it is recognized that the subject lifting lug design wouldhave a much higher load carrying capacity than that predicted bythis criterion due to the reasons previously stated.

Based on the results of the detailed finite element analysis of

the 700 metric ton lifting lug, the following design criterion istendered for consideration. First, in an effort to align the chosenfailure criterion with the most significant mode of failure i.e.,from a dynamic load , the maximum principal stress failure crite-

ria should be utilized. Although attempts by the ASME B30.20Safety Standard highlights the use of a one-third yield criterion,there is no single published failure criterion known to this authorthat can address all of the aspects of the lug design with thismethod. In this authors opinion, the simple limitation of a singletype of component stress for example, a bending stress to someallowable value hardly defines a failure criterion for use in thecurrent or future design of lifting lugs. Furthermore, applying this

Fig. 4 Von Mises stress Contour

Fig. 5 Max principal Stress Contour




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limitation i.e., the one-third yield criterion on the compressivestresses present in the contact area, simply does not work nor willit work in a triaxial state of stress. Second, in conjunction with thechosen criterion, the maximum principal stresses should be calcu-lated for the given working load and lug geometry. The loadshould be distributed over a contact area as determined by eitheran acceptable contact stress technique such as that presented inAppendix A or by a nonlinear finite element contact stress analysisprocedure. The resulting maximum principal stresses in the bodyof the lug that are located slightly beyond the area of load appli-cation should then be compared to an allowable stress value equal

to 20% of the UTS of the lug material. This will be consistentwith both domestic and Canadian regulations regarding the imple-mentation of a well defined FOS of 5 on fracture and ultimatebreaking strength. Finally, the minimum principal stresses shouldbe computed and compared to simply one times the specifiedminimum yield strength of the lug material. This allowable isconsistent with those specified in other ASME Codes 4 and thecompressive stresses do not pose the same crack opening hazardfound in purely tensile stresses. Utilizing this method of analysisnot only achieves a clear utilization of a highly recognized FOS,but also addresses the issue of contact stresses about which theRegulations and Standards identified herein have remained for-ever silent.

Summary and Conclusions

A method for evaluating the structural adequacy of various lift-ing lugs utilized in the erection and uprighting of large pressurevessels was presented. The analysis techniques were described indetail and design guidelines for vessel lifting lugs were tendered.The statutory and provincial regulations in both the United Statesand the province of Alberta, Canada, were also reviewed and dis-cussed with respect to the too often utilized phrase factor ofsafety FOS . Hopefully, the introduction of a clearly definedFOS of 5, when utilized with the maximum principal stress failurecriterion, will serve as a constructive criticism to the very limiteddesign criterion given in the current ASME safety standardB30.20 3 entitled, Below the Hook Lifting Devices and may

be applied to lifting lugs for large and heavyweight pressure ves-sels in future design standards. Because of the limited number ofrepetitive loading cycles that vessel lifting lugs actually experi-ence during the service life of a vessel, a recommendation is madeto either clearly exclude vessel lifting lugs from the scope ofASME B30.20 3 or to address the design aspects in a separatestandard to be developed at a later date. Based on the resultspresented herein, it is hoped that a more realistic assessment of thefailure modes and the proper selection of a failure criterion will befurther studied and revised as necessary by the ASME B30.20Committee, thereby leaving less to chance for the less experienced

design engineer of rigging and materials handling equipment.

NomenclatureCE material constant for contacting bodies

D1 diameter of lifting lug hole

D2 diameter of lifting shackle pinE1 modulus of elasticity for lifting lug

E2 modulus of elasticity for lifting shackle pinKD geometric constant for contacting bodies

L length of contact; thickness of lifting lug

P load to be lifted; maximum safe working load

bb width of rectangular contact area

p load per unit length of contact

1 Poissons ratio for lifting lug

2

Poissons ratio for lifting shackle pincmax maximum calculated contact stress

Appendix A: A Proposed Design Criterion for Vessel

Lifting Lugs in Lieu of ASME B30.20

First, we will evaluate the contact stresses in the lifting lugbased upon the one-fifth of the ultimate tensile strength designcriteria contained within the Alberta OS&H Act.

Lifting Cover Contact Stresses, SA-508 Gr 3 Cl 2.

Cylinder in a Cylindrical SocketTable 33, Case 2c, Formulasfor Stress & Strain [6].

Fig. 6 Min. Principal Stress Contour




8/8

Provides an explanation of Table 33 and the notation used.

Diameter of top cylinder: D28.46in

Poissons ratio forthe top cylinder:

20.3

Modulus of elasticity

for the top cylinder: E2

29.7106

lbf

in2

Diameter of bottom socket: D18.543in

Poissons ratio forthe bottom socket:

10.3

Modulus of elasticityfor the bottom socket: E127.810

6 lbf

in2

Loading: P1.54338106 lbf

Length of cylinder: L11.50in

KDD2 D1

D1 D2KD =870.768 in

CE1

1

2

E1+

1 2

2

E2CE =6.337108

in2

lbfLoad per unit length:

pP

Lp=1.342105

lbf

in

Width of rectangular contact area:

bb1.60p KD CE bb=4.354 in

Stress:

cmax0.798 pKD CE

cmax=39354lbf

in2

With the UTS of SA-508 Gr 3 Cl 2 of 90000 psi, this is anacceptable material for the lifting lug.

Note: See Timoshenko 5 and Sague 9 for technique andformulas utilized above. ASME Paper Lifting Lug.mcd

References 1 U. S. Department of Labor, Occupational Safety & Health Administration,

2002, Code of Federal Regulations, Title 29, Part 1926, Safety & Health

Regulations for Construction, Subpart 251, U. S. Government Printing Office,

Washington, DC. 2 Province of Alberta, Canada, 2000, Alberta Regulation 448/83, Occupational

Health & Safety Act, Queens Printer for Alberta, Edmonton, Alberta, Canada. 3 ASME, 1999, Below-the-Hook Lifting Devices, ASME, New York, ASME

B30.201999. 4 ASME, 2001, ASME Boiler & Pressure Vessel Code, Rules for Construction of

Pressure Vessels, Division 2-Alternate Rules, ASME, New York. 5 Timoshenko, S. P. and Goodier, J. N., 1970, Theory of Elasticity, 3rd ed.,

McGraw-Hill, New York. 6 Young, W. C., 1989, Roarks Formulas for Stress & Strain, 6th ed., McGraw-

Hill, New York.

7 AWS, 1997, Specification for Welding Industrial and Mill Cranes & OtherMaterial Handling Equipment, American Welding Society, Miami ANSI/AWS

D14.11997. 8 Boresi, A. P., Sidebottom, O. M., Seely, F. B., and Smith, J. O., 1978, Ad-

vanced Mechanics of Materials, 3rd ed., Wiley, New York. 9 Sague, J. E., 1978, The Special Way Big Bearings Can Fail, Mach. Des.

50 21 , 113117.


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Pressure vessel lug design