ML20100P247
| ML20100P247 | |
| Person / Time | |
|---|---|
| Site: | Shoreham File:Long Island Lighting Company icon.png |
| Issue date: | 09/17/1984 |
| From: | FAILURE ANALYSIS ASSOCIATES, INC., LONG ISLAND LIGHTING CO. |
| To: | |
| References | |
| OL-I-017, OL-I-17, NUDOCS 8412140083 | |
| Download: ML20100P247 (64) | |
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984 UNITED STATES OF AMERICA D
NUCLEAR REGULATORY COMMISSION Before the Atomic Safety and Licensing Board 3
In the Matter of
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LONG ISLAND LIGHTING COMPANY
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Docket No. 50-322(OL) i
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(Shoreham Nuclear Pown.r
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Station, Unit 1)
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J CRANKSHAFT EXHIBITS TESTIMONY OF ROGERT L. McCARTHY, PAUL R. JOHNSTON, EUGENE MONTGOMERY AND SIMON K. CHEN D
AND TESTIMONY OF EDWARD YOUNGLING AND FRANZ PISCHINCER 3
AND TESTIMONY OF CLIFFORD WELLS, DUANE JOHNSON, HARRY WACHOB, CRAIG SEAMAN, DOMINIC CIMINO AND N. K. BURRELL D
l 0412140083 840917 FDR ADOCK 05000322 0
PDR VOLUME III Exhibits 17 - 26
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NUCLEAR REGULATORY COMMIS$10N I(
- b~Ab Docket No KO' A-O_Le_ Official Exh. Ms-la the mtttar of _bt1 Iand $b]
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IDENTIF1 D RECElVE3 Applicant Intervenot _/
REJECTED
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itegorter D
LILCO, August 14, 1984 g
UNITED STATES OF AMERICA NUCLEAR REGULATORY CCMMISSION Before the Atomic Safety and Licensing Board g
In the Matter of
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LONG ISLAND LIGHTING COMPANY
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Docket No. 50-322(OL)
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(Shoreham Nuclear Power
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Station, Unit 1)
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CRANKSHAFT EXHIBITS D
C-1 Evaluation of E=ergency Diesel Generator Crankshafts at Shoreham and Grand Gulf Nuclear Power Stations prepared for TDI Diesel Generator Owners Group dated May 22, 1984 (hereinafter " Owners Group Crankshaft Report"), Figure 3-4 D
C-2 Specification for Diesel Generator Sets, Shoreham Nuclear Power Station - Unit 1, Spec. No. SH1-89, Revision 2, January 26, 1983, page 1-20.
C-3 U.S. Nuclear Regulatory Commission Regulatory Guide 1.9, D
Revision 2 December 1979.
C-4 IEEE Standard Criteria for Diesel-Generator Units Applied as Standby Power Supplies for Nuclear Power Generating Stations, Std 387-1977.
C-5 Transcript of July ll, 1984 meeting of the TDI Diesel D
Generator Owners Group, pages 124-25.
C-6 Available Logged Hours of Gperation of DSR-48, Rated 3500 KW
@ 450 RPM.
C-7 TDI Diesel Generator Run Historv - Shoreham Nuclear Power D
Sation - Unit 1 - August 6, 1984 C-8 Results of non-destructive examinations of replace =ent crankshafts at Shoreham after 100 hours0.00116 days <br />0.0278 hours <br />1.653439e-4 weeks <br />3.805e-5 months <br /> of operation at full load or greater.
O C-9 American Bureau of Shipping, Rules for Building and Classing Steel Vessels (1983), 5 37.17.1.
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C-10 American Bureau of Shipping, Rules for Building and Classing Steel Vessels (1983), Table 34.3.
C-ll TDI Crankshaft Drawing Number 03-310-05-AC.
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C-12 American Bureau of Shipping Reports on Castings or Forgings of Replacement Crankshafts.
C-13 American Bureau of Shipping letter to TDI dated May 3, 1984.
C-14 Diesel Engine Manufacturers Association Standard Practices for Low and Medium Speed Stationary Diesel and Gas Engines (1972 ed.), pages 53-56.
C-15 TDI Proposed Torsional and Lateral Critical Speed Analysis, August 22, 1983.
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C-16 Field Test of Emergency Diesel Generator 103 with 13 x Crankshaft, April, 1984.
C-17 Owners Group Crankshaft Report.
C-18 Crankshaft Torsional Stress Calculations for 8L 17 x 21 Engine-Generator Set, July 19, 1984.
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C-19 Table 2.2 from Owners Group Crankshaft Report showing natural frequencies from TDI analyeis.
C-20 Table 2.4 from Owners Group Crankshaft Report showing single order nominal stresses from TDI analysis.
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C-21 Table 2.5 from Owners Group Crankshaft Report showing nominal stresses calculated from torsiograph.
C-22 Crankshaft Torsional Stress Calculations for 8L 17 x 21 Engine-Generator Set, July 19, 1984, page 11.
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C-23 Figure 3-3 from Owners Group Report showing comparison of measured and calculated torque.
C-24 Tables 3.6 and 3.7 from Owners Group Crankshaft Report showing comparison between analytical and test results.
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C-25 Figure 3-13 from Owners Group Crankshaft Report showing fatigue endurance limit of replacement crankshafts on Goodman diagram.
C-26 Oberg and Jones, Machinery's Handbook (18th Ed.) pages 352-53: Shigley, Mechanical Engineering Design
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(McGraw-Hill) pages 212-13; Rothbart (editor), Mechanical J
Design and Systems Handbook (McGraw-Hill) page 18-4 2
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C-27 Engineering and Design Coordination Report No. F-46109G.
C-28 Military Specification No. 13165B, Amendment 2, June 25, 1979.
C-29 LILCO Operational Quality Assurance Reports (EDG 102 and
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103 Crankshafts).
C-30 Metal Improvement Company Certificate of Shot Peening (EDG 102 and 103 Crankshafts).
C-31 Certificate of Non-Destructive Testing Issued by Krupp Stahl AG (EDG 102 and 103 Crankshafts).
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C-32 LILCO Magnetic Par-icle Testing and Liquid Penetrant Testing Records (EDG 102 and 103 Crankshafts).
C-33 LILCO Ultra Sonic Testing Rec 1rds (EDG 102 and 103 Crankshafts).
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C-34 H. Fuchs and R. Stevens, Metal Fatigue in Engineering (1980) at pages 226-227; H. Uhlig, corrosion and Corrosion Control at pages 132-133.
1 C-35 Metal Lnprovement Company Certificate of Shot Peening (EDG 101 Crankshaft).
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C-36 LIICO Operational Quality ' Assurance Reports (EDG 101 Crankshaft).
i C-37 Certificates of Non-Destructive Testing Issued by Krupp Stahl AG (EDG 101 Crankshaft).
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LILCO Magnetic Particle Testing,(Liquid Penetrant Testing C-38 and Ultra Sonic Testing Records EDG 101 Crankshaft).
C-39 Kirk, Behavior of Peen-Formed Steel Strip on Isochronal Annealing, Proceedings of the Second International Conference on Shot Peening at page 231, (May, 1984).
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Fa AA.84 3 16 k ' Mj b..-.--PAO 7396/PRJ-03310 M
O EVALUATION OF EMERGEEY DIESEL ENERATOR CRANKSHAFT 5 AT 510RENAM AMD GRAls ER.F IWCLEAR POWER STATIONS O
O The report is final, pending confirmatory reviews required by FaAA's QA operating procedures.
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Prepared by Failure Analysis Associates 3
Palo Alto, California 4
3 Prepared for TDI Diesel Generator Owners Groups
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May 22,1984 l
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STATDENT OF APPLICA81LITY This report addresses the structural integrity of the crankshaf ts in Transamerica Delaval Inc. 05R-48 engines at the Shoreham % clear Power Station p
and 05RV-16-4 engines at the Grand Gulf nuclear Power Station.
In view of possible differences in generators, flywheels, and engine operating conditions, the results may not necessarily apply to other engines of the same model.
These plant-specific differences, where they exist, will be evaluate::
in separate reports.
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O PRJ:R03310A-ig/geg 05/14/84 DRAFT C
EnECUTIVE 335WtY DSR-4813-IEN ST 12-INCN OtANKSHAFTS AT 5HORENAM lEl CLEAR POWER STATION O
The structural integrity of the replacement 13-inch by 12-inch diameter crankshafts installed in the emergency diesel generators at the Shore 5a-Nuclear Power Station has been extensively evaluated by testing and analy-sis.
Conventional analytical techniques typically utilized by the diesel O
engine industry show that 13-inch by 12-inch crankshafts comply with DEM*
j requirements. Angular displacements of the free end of the crankshaft, stress ranges in the most highly stressed crank pin fillets, and the range of output torque at the flywheel were measured at and above full-rated load.
The tor-4
.O slograph measurements of twist showed that the crankshafts meet the DEM*
i requirements.
In addition, the strain gage measurements of maxi...um bending and torsional stress and calculations of maximum stress by a modal superposi-tion analysis showed that the crankshafts have a factor of safety in fatigue O
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of 148 "ithout taking into account any benefit of shot peening the crant pin fillets.
The factor of safety was determined from the measured +ndurance limit of the original 13-inch by 11-inch crankshafts that cracked in
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fatigue.
The measured shaf t response was in close agreement with. that p re -
dicted by the modal superposition analysis.
The replacement crankshaf ts are suitable for unlimited operation at the rated load and speed in the emergency diesel generators at SNPS.
3 DSRV-16-413-IKH BY 13-INCH CRANKSHAFTS AT stAls GULF filCLEAR STATION The structural integrity of the 13-inch by 13-inch diameter crankshaf ts installed in the emergency diesel generators at the Grand Gulf Nuclear Statior j
has been evaluated by testing and analysis.
Conventional analytical tec m -
ques typically utilized by the diesel engine industry show that 13-inch by 13-inch diameter crankshafts comply with DEMA requirements. Angular displace-ments of the free end of the crankshaf t were measured at and above full-rated load at TDI. The torstograph measurements of twist taken during f actory tests v
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PRJ:R03310A-ig/ges 05/14/84 ORAFT
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showed that the crankshafts meet the DEMA requirements.
The measured shaf t response was in close agreement with that predicted by the modal superposition analysis.
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The DSRV-16 4 crankshaft is sensitive to operating speed and tne balance of cylinder firing.
Torsiograph tests of several engines should be conducted to determine the range of crankshaft response per.7itted by T3; specified balance limits and the governor characteristics.
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The oil holes in the main journals numbers 4,
6, and 8 are more critical in torsion than are the crankpin fillets and should be inspected.
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j PRJ:R03310A-ig/geg 05/14/84 DRAFT 7
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PA0 7396 Task No. 03310:
1 TABLE OF CONTENTS l
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STATEMENT OF APPLICABILITY..............................................
1 11 E X E C U T I V E SUMMA R Y........................................................
PART A
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1.0 INTRODUCTION
TO REVIEW OF DSR-4813-INCH BY 12-INCH CRANKSHAFT.....
1-1 Section 1 References...............................................
1-?
2.0 COMPLIANCE OF CRANKSHAFT WITH DIESEL ENGINE MANUFACTURERS
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ASSOCIATION RECOMMENDATIONS........................................
2-1 2.1 Review of TDI Torsional Critical Speed Analysi s...............
2-1 2.1.1 6 a t u ra l Frequen c i e s....................................
2-2 2.1.2 kom i n a l St r e s s e s.......................................
2-2 2.2 Review of Stone & Webster Engineering Corporation To r s i o g r a p h Te s t..............................................
2-3
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2.2.1 ha t u ra l Frequ e n c i e s.................................... 2
- 2.2.2 Nom i n a l St re s s e s.......................................
2 4 2.3 Nominal Stresses for Underspeed and Overspeed Conditions..,...
2-5 Se c t i on 2 Re f e re n c e s...............................................
2-f
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3.0 FAT !GUE AN AL YSI S OF CR ANKSH AFT.....................................
3-3.1 Cranksha f t Dynamic Torsional Analysis.........................
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3.1.1 T o r s i o na l Mo de 1........................................
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3.1.2 Harmonic loading.......................................
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3.1.3 Comparison of Calculated Response With Test Data.......
3-3 3.2 Cra nk s ha f t St res s Analy si s....................................
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3.2.1 F i n i t e El eme nt Mo d e 1...................................
3 3.2.2 Stres ses Due to Torsional Loading......................
3-6 3.2.3 Stresses Due to Gas Pressure Loa ding...................
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3.2.4 Compa ri son of St resses wi th Test Cata..................
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3.3 Crank sha f t Fati gue Failu re Ma rgin.............................
3-8 3.3.1 Stresses in Replacement Crankshafts....................
3-5
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3.3.2 Endurance Limit f or Failed Crankshaf t................... 30 3.3.3 Endurance Limit for Replacement Crank sha f ts............
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3.3.4 Factor of Sa f ety Against Fatigue Failure...............
3-11 Se c t i on 3 Re f e re n c e s...............................................
3 - 12
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PRJ:R03310A-ig/ges 05/14/84 otAFT TABLE OF CONTENTS CDNTINUED Page 4.0 D I SC U S S I ON AND C ON C L U S I ONS......................................... 4 - 1 h
Se c t i on 4 Re f e re n c e s...............................................
4-2 FART B
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5.0 INTRODUCTION
TO REvlEW 0F DSRV-16 413-INCH BY 13-INCH CRANKSHAFT..
5-1 5.1 I n d u s t ry E x p e t e n c e............................................ 5 - 1 Se c t i o n 5 Re f e r e n c e s...............................................
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6.0 COMPLIANCE.0F CRANKSHAFT WITH DIESEL ENGINE MANUFACTURERS ASSOCIATION RECOMMENDATIONS........................................
6-1 6.1 Revf ~ af TDI Torsi onal Cri ti ct.1 Speed Analysi s............... 6-1 6.;,1 Na t u ra l Frequen ci e s....................................
6-2 6.1.2 Nomi n a l St re s s e s....................................... 6 - 2 6.2 Review of TDI Torsiograph Test................................
6-3
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6.2.1 Na t u ra l Frequ e n ci e s.................................... 6 4 6.2.2 Nomi n a l St r e s s e s....................................... 6 4 6.3 Nominal Stresses for Underspeed and Overspeed Conditions......
6-5 Se c t i o n 6 Re f e r e n c e s............................................... 6 - 6
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7.0 CR ANKSH AFT DYNAM IC TOR SI ON AL AN AL YSI S..............................
7-1 7.1 To r s i on a l Mod e 1............................................... 7 - 1 7.2 Harmonic Loading..............................................
7-2 7.3 Comparison of Calculated Response With Test Data..............
7-3 Section 7 References...............................................
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8.0 DI SCUS S I ON AND CONC L U S I ON S.........................................
8-1 Appendi x A - Component Ta s k De s c ri pt i on.................................. f - !
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PART A:
REVIEW OF DSR-48 13-IIICH BT 12-IEH CIUutKSHATT
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1.0 INTRODUCTION
10 EVIEW (F B5R-4813-!KH Bf 12-IRCH OtANKSHAFT As a result of fatigue damage in the crankshafts of three emergency diesel generator sets at Shoreham helear Power Station, replacement crank-shafts of current design have been installed. The principal difference is aa
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increase in crankpin diameter from 11 inches to 12 inches.
This report pre-sents Failure Analysis Associates' findings on the adequacy of the replacement crankshafts in the emergency diesel engines at Shoreham % clear Power Station.
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A detailed investigation of the orginal crankshaft, which attributed failure to high cycle torsional fatigue resulting from inadequate design, was previously conducted by Failure Analysis Associates (FaAA) [1-1]. An analysis of the replacement crankshafts, conducted prior to dynamic testing, was also performed by FaAA [1-2].
l The installation of the replacement crankshaft was required to meet the l
reconenendations of the Diesel Engine Manufacturers Association (DEMA).
In Section 2.0, the torsional calculations of Transamerica Delaval Inc. (TOI)
[1-3] and the torsiograph test results of Stone & Webster Engineering Corpora-tion (SWEC) [14] are reviewed for compliance with the DEMA stress allowables.
i In Section 3.0, a detailed analysis of the factor of safety against I
fatigue failure is performed. A torsional dynamic analysis is used to compute i
nominal torsional stresses at each crank throw.
A three-dimensional finite r
j element analysis of a quarter section of a crank throw is then performed to i
obtain the local stresses in the crankpin fillet.
The computed stresses are compared to dynamic strain gage measurements to verify the models.
In turn, the models are used to verify that strain gages have been placed in locattoas of maximum stress.
Finally, the seasured stresses are used to compute a f ac-tor of safety against fatigue failure for the replacement crankshafts.
This is accomplished by comparing the measured stresses with the endurance lin:
for the replacement crankshaf ts.
1-1
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Section 1 References 1-1 " Emergency Diesel Generator Crankshaft Failure Investigation, Shorehe-Aclear Power Station," Failure Analysis Associates Report No. Fa AA-83 10-2.1, October 31, 1983.
1-2 " Analysis of the Replacement Crankshafts for Emergency Diesel Generators, Shoreham Actear Power Station," Failure Analysis Associates Report to.
Fa AA-83-10-2. 2, October 31, 1983.
1-3 Yang, Roland, " Proposed Torsional and Lateral Critical Speed Analysis:
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Engine h abers 74010/12 Delaval-Enterprise Engine Model 05R 43 350';
Transamerica Delaval Inc., Engine and Coms es-sor Division, Oakland, California, August 22, 1983.
1-4 Bercel, E., and Hall, J.R.,
" Field Test of Emergency Diesel Gene ato-103," Stone & Webster Engineering Corporation, Maren 1984
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1-2
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2.0 CIMPLIAKE W CRAfESMAFT WITH DIESEL ENGIME imRIFACTURERS ASSOCIATION REC (NOEleAT10R$
The purchase specifications for the diesel generator sets required that the recommendations of the Diesel Engine Manufacturers Association, DEC 1
[2-13, be followed. These reconenendations state:
In the case of constant speed units, such as generator sets, the objective is to insure that no harmful torsional vibratory stresses occur within five percent above and below rated speed.
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For crankshafts, connecting shafts, flange or coupling components, etc.,
made of conventional materials, torsional vibratory conditions shall genea-ally be considered safe when they induce a superim-posed stress of less ?han 5000 psi, created by a
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single order of vibration, or a superimposed stress of less than 7000 psi, created by the sussnation of the major orders of vibration which might come into phase periodically.
In August, 1983, Transamerica Delaval Inc. (TDI) performed a torsional
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critical speed analysis of the replacement crankshafts [2-2).
References to TDI analysis in the body of this report all reference this effort. In Se:: tor 2.1, this analysis will be reviewed for compliance with the above allowaMe stresses.
The inappropriate T values employed in the original analysis o' n
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the 13-inch by ll-inch crankshaf t were replaced with the correct values f o-this analysis, in January, 1984 Stone & Webster Engineering Corporation, SWEC, conducted a torsiograph test on a replacement crankshaft at Shorena-Nuclear Power Station [2-3].
In Section 2.2, the test results will be co -
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pared with the above allowable stresses.
2.1 Review of TDI Torsional Critical Speed Analysis Diesel generator torques due to dynamic response are usually calculate:
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in two steps.
First, the torsional mode shapes and natural frequencies o' vibration are calculated.
Second, the dynamic forced vibration response doe to gas pressure and reciprocating inertia loading is calculated.
TDI calcu-lated the response at 100% of rated level of 3500 kW.
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2-1
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2.1.1 Natural Frequencies The first step in a torsional critical speed analysis is to determine the natural frequencies of the crankshaf t.
The engine speed at which a given order resonates may then be calculated. The diesel generator is modeled as a
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system of Iveped mass moments of inertia interconnected by torsional spaings, as shown in Figure 2-1.
The inertia and stiffness values are s h ow*.
In Table 2.1.
It has long been standard practice in the diesel engine indust y to
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solve this eigenvalue problem by the Holzer method [2 4].
inis metho: has been used for at least 40 years [2-5], and thus is well established.
TDI used the Holzer method to calculate the system's first three nats.
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ral frequencies, which are shown in Table 2.2.
The first natural frequen:y was found to be 38.7 Hz, which produces 4th order resonance at 581 rpm.
2.1.2 Nominal Stresses
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The second step in a torsional critical speed analysis is to determne the dynamic torsional response of the crankshaft due to gas pressare anc re.
ciprocating inertia loading.
The 1st order is a harmonic which repeats on:e per revolution of the '.. snkshaf t.
For a four-stroke engine, harmonics o' o*.
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de r 0. 5, 1. 0, 1. 5, 2. 0, 2. 5... ex i s t. TDI performs this calculation for ea:"
order of vibration up to 12.0 separately.
For each order, the applied torde at a cylinder due to gas pressure and reciprocating inertia is calculate:.
The values of this torque for each order are usually normalized by dividing by the piston area and throw radius.
The normalized value for the nth order is
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referred to as i.
The values of T for signifcant orders used by TDI are n
n shown in Table 2.3.
These values may be compared to those recommended by Lloyd's Register of Shipping, LRS [2-6].
It is found that TDl's values a e higher than LRS's values for low orders and lower for high orders.
Howeve*,
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the largest single order was measured to be within 51 of those computed as'a; TDi's values of T.
The response is then calculated by one procedure if tne n
harmonic is at resonance and by another if the harmonic is away from reso.
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nance.
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2-2
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At resonance, the torsional vibration amplitudes would increase indefi-nitely in the absence of damping.
The solution is o6tained by balancing the
' energy input with the energy loss due to damping. TDI used an empirical forr
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of hysteresis damping due to friction. The purpose of this calculation is tc j
ensure that the diesel generator could be brought e to operating speed witn-out undergoing excessive stresses as critical speeds are passed. Observations have shown that excessive vibration during startup does not occur [2 3).
Since the engine runs at 450 rpm and the 4th order critical speed is 580 rpe,
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the calculated response at resonance will not be further considered.
l Away from resonance, the torsional vibrations reach a steady-state level even without the aid of any damping. The magnitude of this response fo-
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each structural mode and loading order is calculated as the product o# a dynamic amplification factor and an equivalent static equilibrium amplitude.
The equivalent static equilibrium amplitude is computed using a modal load and modal stiffness [2-7] for the nth order harmonic and given mode shape.
The nominal shear stress, t,
in the 12-inch pin of Crankpin No. 8 for each order
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is then calculated from the dynamic torque, T, using i = Tr/J, where r is the pin radius and J is the polar moment of inertia.
TDI calculated the response for the first three modes and plottee the
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results for only the first mode since higher podes produce much smaller stresses. The nominal shear stresses for the significant orders are sno n in Table 2.4 It is seen that the largest single order stress of 2980 pst at rated load and speed for the 4th order is well below the 5000 psi D E ". -
allowable.
TDI does not calculate the associated phase angle with the response of each order, so that it is not possible to calculate the combined response.
The measured combined response will be c6mpared with the allowable in the nee
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section.
2.2 Review of Stone & liebster Engineering Corporation Torstograph Test Torstograph tests are comenonly used to confirm torsional vibrationai
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calculations.
The test is usually performed in two stages.
The first stage 2-3
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b is performed without load at variable speed and is used to determine the location of critical speeds.
Critical speeds may also be determined while operating at a fixed speed and observing the frequency content of tne response. The second stage is performed at rated speed of 450 rpm with v a -i -
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able load, and is used to confirm the forced vibratim calculations.
2.2.1 Natural Frequencies The frequency content of the torsional vibration signal at 450 rp-
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l showed a resonance at 38.6 Hz.
This value is in excellent agreement witr TDI's computed value of 38.7 Hz.
2.2.2 Nominal Stresses
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The torstograpg provides the angular displacement response of the free end of the crankshaf t.
This displacement may be decomposed into components corresponding to each order. The' peak-to-peak response may also be obtained.
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The nominal shear stress, t, in Crankpin No. 8 may be established f ro-the amplitude of free-end vibration by assuming the shaf t is vibrating in the first mode. The nominal shear stress is then found to be 9562 psi per de; ee of free-end vibration from the TDI analysis [2-2].
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SWEC tabulated the single order and pea k -to-pea k response for beta 3500 kW (100% of rated load) and for 3800 kW (109t of rated load).
Tnese values have been factored to obtain nominal shear stresses and are sho - ir.
Table 2.5.
The results at 1001 load show that the largest single order has a
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stress of 3108 psi which is well below the DEMA allowable of 5000 psi.
The total stress of 6626 psi is also shown to be below the DEMA allowable of 70 %
psi.
At 3600 kW the stresses of 3242 psi for a single order and 6875 psi fo-combined response are also lower than 5000 psi and 7000 psi respectively.
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3900 kW the corresponding stresses are 3287 psi and 6958 psi by linear extea-polation.
However, the 3900 kW 1evel is a two-hour overload rating at au c-the engine is not required to operate continuously.
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The measured response at 3500 kW is in agreement with that calculated by TDI and shown in Table 2.4.
The measured values are somewhat higher than the calculated values.
2.3 Hominal Stresses for thderspeed and Overspeed Conditions g
Strict interpretation of DEMA regulations requires the consideration of torsional stresses at conditions other than operating speed.
During norma' standby diesel generator testing at $NPS the units are synchronized to the O
Long Island power distribution grid to simulate full load and two hour ove -
load conditions.
At this time the frequency characteristics of the grid assure that the speed and associated frequency vary less than 2 1 Hz or 1.6%
speed.
O During other testing and potentially during a LOOP /LOCA event, the unit speed is controlled by the Woodward Governor.
Testing at the SNP5 site, during which step changes in load were produced by starting or stopping various pumps, revealed the largest variations in speed to be -31 to +2 O
associated with increasing or decreasing load respecti vely.
These step changes were the same order of magnitude of those calculated to occur da-in; a LOOP /LOCA.
The time lag associated with the unit's ability to return to 450 rpm was likewise found to be less than 3 seconds.
Since speed variations associated with load step changes canno* be produced at full or two-hour overload power conditions due to grid conne: tion, the modal superposition taethod was used to calculate the effects.
The f ee-end vibration amplitude was first calculated by the modal superpositio-3 method, and then the nominal torsional stress was calculated to be 955: psi per degree of free-end rotation [2-2]. The T values used in the modal super-n position analysis were assumed to be equal to those obtained at 3500 W ano 450 rpm. The maximum nominal torsional stresses at 428 rpm (95: rated speeo) and also at 473 rpm (105% rated speed) have been calculated to equal the DEC g
limit of 7000 psi within 231, which reflects the uncertainty in the Tn '8bd'5 at these speeds.
Thus, within the accuracy of the analysis, compliance wit's DEMA is obtained.
Furthermore, the very small potential time caring =nich such conditions could actually occur with the demonstrated performance of the
)
governor precludes fatigue damage.
2-5 3
1
)
Section 2 References 2-1 Standard Practices for Low and Medium Speed Stationary Diesel and Gas Engines. Diesel Engine Manufacturers Association, 6th ed.,1972.
22 Yang, Roland, " Proposed Torsional and Lateral Critical Speed Analysis:
Engine Numbers 74010/12 Delaval-Enterprise Engine Model DSR 48 35"4 KW/4889 BHP at 450 RPM."
Transamerica Delaval Inc., Engine and Compres-sor Division, Oakland, California, August 22, 1983.
2-3 Bercel.
E.,
and Hall, J.R.,
" Field Test of Emergency Diesel Generato-
)
103 " Stone & Webster Engir.eering Corporation April 1984 2-4 Thomson, William T., Theory of Vibration with Applications.
Second ett-tion, Prentice-Hall,1981, 2-5 Hartog, Den, Mechanical Vibrations. Third edition, McGraw-Hill,19c'.
2-6 Lloyd's Register of Shipping, Guidance Notes on Torsional Vibration Cha -
acteristics of Main and Auxillary Oil Engines.
2-7 Craig, Roy R., Jr., Structural Dynamics:
An Introduction to Comp;ter Metnods. Wiley, 1981.
I a
3 1
i b
2-6 l
~~
-, ~
O
)
)
TABLE 2.1 STIFFNESS Am INERTIAS FOR TDI HOLZER ANALfSIS
)
Inertia Inertia 5tiffuss Location (Ib. ft, sec2)
(ft. Ib./ rad)
Front Gear 6.8 58.1 = 106
)
Cylinder No. 1 49.2 84.7 = 106 Cylinder No. 2 47.9 84.7 = 106 Cylinder No. 3 47.9 84.7 = 106
)
Cylinder No. 4 47.9 84.7 = 106 Cylinder No. S 47.9 84.7 = 106 Cylinder No. 6 47.9 84.7 = 106 Cylinder No. 7 47.9 84.7 = 106 Cylinder No. 8 50.1 76.9 = 106 Flywheel 1100.1 276.8 = 10E Generator 2650.4 s
i 2-7 i
)-
)'
TABLE 2.2 TOR $10NAL MTURAL FREQUENCIES FROM TDI ANALY$l$
)
Natural Frequency (Hz)
)
1 38.7 2
92.9 3
116.7 4
i I
r l
l
'l
)
2-8 i
3 TABLE 2.3 TOR $10NAL LOADINGS FOR TDI ANALYSIS Q
Order Torsional Landing T, (psi) 1.5 129.5
)
2.5 71.7 3.5 42.8 4.0 27.7 4.5 23.8 4
i 5.5 12.8
~
i
?
l
- 3 1
1 3
2-9 l'
l
)-
)
TABLE 2.4 SINGLE.0RDER NOMINAL $ HEAR STRESSES FROM TDI ANALYSIS
)
Amplitude of Nominal Shear Stress (psi) 1.5 1606
)
2.5 1064
- 3. 5 452 4.0 2980
)
4.5 565 5.5 1080 DEMA Allowable for Single Order 5000
)
)
)
)
2-10
)
O O
TABLE 2.5 l
NOMINAL SHEAR STRESSES CALCULATED FROM SWEC TOR 5!0 GRAPH TEST g
Order Amplitude of free-end
%11tude of nominal rotation (degrees)
Shear Stress (psi)*
At 500 kW At 500 kW At 500 kW At 200 kW O
1.5 0.171 0.187 1635 1785 2.5 0.130 0.140 1243 1339 1
- 3. 5 0.058 0.061 555 58a 4.0 0.325 0.339 3108 3242 4.5 0.064 0.067 612 643 l
5.5 0.127 0.136 1214 1300 O
DEMA Allowable for a Single Order 5000 5000 I/ peak to peak 0.693 0.719 6626 6875 2
O DEMA Allowable 7000 7000 1/ peak to peak 2
o' Amplitude of nominal shear stress is calculated to be 9562 psi per degree free-end rotational amplitude.
i i
l l
i b
2-11
-)
d h
i i
a 1
e
~.
e n
e e
4:
?
t t
8.
8 I
I I
I s
f f
i f
j i
j j
i 11 3
5 5
i a
i a
a i
1 l
l R ot ational inertia Torsional spring Figure 2-1.
TDI dynamic model.
)
b 1
4 F S A A - 8 4 18'
)
3.0 FATleE 414 LYSIS OF CRAIESMAFT In Section 2.0 it was feend that the replacement crankshaf ts satisfy the DEMA nominal stress recommendations for both 3500 kW and 3900 kW.
The stresses for a single order were considerably below the 5000 psi tha-is j
)
recossmended as an allowable.
However, the stresses for combined orders were quite close to the 7000 psi tnat is recommended as an allowable.
While tne DEMA limits are believed to contain an intrinsic (though unspecified) safety margin, a fatigue analysis of the crankshaft was undertaken to determine the
)
true margin.
First, a dynamic torsional analysis of the crankshaft is performed to determine the true range of torque at each crank throw.
This model is co -
i pared with $WEC test data for the amplitudes of free end vibration, measare:
)
with the torsiograph, and for range of torque near the flywheel, measure with strain gages on the shaf t.
Second, a finite element model of a one quarter crank throw is used to
)
compute the local stresses in the fillet region.
Torsional and gas pressare loading cases are considered. The results of this analysis are compared with strain gage test results, which were measured ir. the fillets of Crant Th-o-Nos. 5 and 7.
)
Third, the fatigue endurance limit is established for the replacemen:
crankshaft by first obtaining the endurance limit for the failed crankshaf ts, and then assessing the differences between the failed and replacement crans-shafts.
The endurance limit is compared with values provided in the litera-
)
ture.
Finally, a factor of safety against fatigue failure is computed.
3.1 Crankshaft Dynamic Torsional Analysis
)
3.1.1 Torsional Model FaAA developed a dynamic torsional model of the crankshaf t to oveacome limitations in TDl's conventional forced vibration calculations.
For
)
3-1
)
3 f
O instance, the TDI method does not compute the phase relationship between the various orders or modes, so it is not possible to compute the true summa-j tion.
The actual maximum stress is a direct result of this sunenation.
Fur-thermore, the TDI method always predicts maximum stress in Crankpin No. 8,
.g which is pnerally true for a single order in the first mode but not true for the combined response of all orders and modes.
The dynamic model developed used the same idealized lumped inertia and torsional spring model as the TDI analysis (Figure 2-1 and Table 2.1) with one O
additional spring placed between the generator and ground to represent tne effect of the grid on dynamic response during synchronous operation.
Tnts spring constant was found to be 1.409 x 106 ft.-lb./ radian based on generato-D' specifications.
This constant is set close to zero to represent 5 'i,,
3 emergency bus operation.
The first five torsional natural frequencies for the replacement cran -
shaft are shown in Table 3.1.
The first natural frequency was found to be
- 2. 93 Hz due to the connection to the grid.
For operation on the SNPS 3
emergency bus the first natural frequency is 0 Hz (rigid body mode).
Tne other natural frequencies are in agreement with those computed by TM and measured by SWEC.
When the diesel generator is running at a given speed and powe-le.e',
g the forced vibration problem is steady-state where both load and response repeat themselves every two revolutions of the crankshaf t.
To mo:et tne dynamic response, a model superposition analysis [3-1] was used with ha moatt load input.
The calculation of the harmonic loads will be discussed in tne O
neat section.
3.1.2 Harmonic Loading To calculate the harmonic loading on a crankshaf t it is necessary to 3
consider gas pressure, reciprocating inertia, and frictional loads.
The gas pressure loading may be obtained from pressure versus crank angle data. This pressure was measured in the SWEC test [3-2].
The pressure was measured in Cylinder No. 7 by inserting a probe through the air start valve.
A top dea:
3-2 3
D 3
center TDC, mark for Cylinder No. 7 was simultaneously recorded by a probe on the flywheel.
The pressure data at 1005 load was reduced by FaAA to obtain the pressure curve shown in Figure 3-1.
The torque produced by this pressure may then be calculated as a func-tion of crank angle.
The mean value of this torque should be the tora;e required to produce 3500 kW divided by the mechanical efficiency.
A mechan 1-cal ef ficiency of 1.0 was obtained, rather than the expected 0.88.
The dif-ference is probably explained by either the pressure measurements betn; too 3
low or by the TDC being shif ted.
Peak pressures were measured in all tne cylinders to ensure that all cylinders were balanced.
Normally, the excess torque above that required to run the engine a*.
3500 kW is dnsipated by friction.
In this case, because the pressure curve produced the correct power without friction, friction was not applied.
Tne effects of pressure being too low and not applying friction are expected to largely cancel each other.
The reciprocating mass of the connecting rod and piston was found to be approximately 820 lbs.
This mass causes reciprocating inertia torque on the i
crankshaft.
The ef fect of this torque was combined with the gas pressJ"?
I torque.
l)
The total torque was then decomposed into its sine and cosine harmoci:s I
corresponding to each order. These torque harmonics were used in the steady-state analysis.
The magnitude of the torque harmonics are normalized by dividing by the piston area and throw radius.
The resulting normalize:
torques for the most significant orders are shown in Table 3.2.
3.1.3 Comparison of Calculated Response With Test Data i
The response due to the first 24 o-ders and all 11 modes is calcalatec using modal superposition with 2.57, of critical damping for each moce.
Tne actual value of damping used has little effect on the response since the i
orders are not at resonance at 450 rpm.
The SWEC test report stated that tne measured damping in the system was 2.67, [3-2].
b 3-3
(
)
D 3
The calculated amplitude of free-end displacement is compared to the
$WEC test measurements in Table 3.3.
It is seen that the agreement is close for all significant orders.
The vector sunnation listed represents half the maximum peak-to-peak displacement range.
The model also* calculates the range of torque at each crank thro,
which is shown along with the corresponding nominal shear stress (t Tr/J) in
=
Table 3.4.
The computed torque range near the flywheel was found to be 3
312 f t-kips compared with the measured value of 357 f t-kips [3-2]. The straie.
gages were placed close to the flywheel hub, and thus were expected to give higher values.
The apparent stress concentration factor is 1.14.
The co.-
puted torque as a function of crank angle for each crank throw is show-in Figure 3-2.
The computed and measured torques at the flywheel are show as a a
J function of crank angle in Figure 3-3.
3.2 Crankshaf t Stress Analysis
)
3.2.1 Finite Element Model The nominal crankshaf t stress values calculated from the dynamic mo:e' are considerably less than the actual maximum stresses in the cranssnaft.
Tnose nominal values would prevail if the crankshaf t were a long cir:J a-cylinder.
Stresses in the real crankshaft are greatly influenced by its complex geometry and by stress concentrations, especially at the fillet racii between the main journal and web and the crankpin and web.
In this se
- io".
maximum stresses and their location are determined with particular attentio-
)
to the crankpin fillet.
The m;1tt-throw crankshaf t under investigation consists of a se ies o' j
crankpins and main journals interconnected through webs.
Typical structural dimensions of one throw are shown in Figure 3-4.
The main jouraa' is
)
13.0 inches in diameter; the crankpin is 12.0 inches in diameter with a wed thickness of 4.5 inches. Fillet details are also shown in Figure 3-4 The following material properties, corresponding to the A!S: 10C
)
crankshaf t steel, were used in the analyses:
3-4
)
O-O Young's Modulus: E = 30.0 x 106 psi Poisson's Ratio: w = 0.3 A crankshaft throw is subjected to loads of two basic types:
(1) torque transmitted through the throw, which is influenced by the oatpsi power ~ 1evel and by the torsional vibration response of the crankshaf t an:
(2) connecting rod forces applied to the crankpin and reacted at be a
- i a. ;
supports.
O Linear elastic analyses were performed using the computer program M*t~,
K.1-1 Version, from MAP.C Analysis Research Corporation.
Generation of the geometric input data and the post-processing graphics was performed usH; PATRAN-G and PATMAR developed by PDA Engineering. One throw of the crannsna't was analyzed by applying a static unit tilst on the main journal.
Sin:e a'l throws are geometrically identical, a single model with appropriate loads an:
boundary conditions can be used to represent approximately any throw.
Three existing planes of local symmetry were employed in the analysis to keep tne O
finite element model to a feasible si e without compryntsing the accura:j of the results.
These planes of symmetry are shown scheratically in Figs e 3 5 r
along with the coordinate system. The first plan of synenetry is the ve-tical plane passing through mid sections of the crankpin, the web, and the ma'n g
journal. The second and third planes of synenetry are orthogona' to tne first at the mid distance between two adjacent webs in the crankpiG and the mai-journal, respectively.
Thus, only the portion of the crankpin, the wet, aa:
the main journal contained within these planes of symmetry was modeled.
O This model uses eight node, three-dimensi onal, isoparametric bri:.
elements with linear interpolation, capable of modeling an arbitrarily dis.
torted cube.
Each node in an element has three translational deg*ees c'
freedom.
Because the state of stress in the vicinity of the fillet is o' O
greatest interest and stress gradients are highest there, a finer mes*
as used in this region.
Figure 3 6 shows the three-dimensional model, witn nos and element numbers omitted for clarity, along with the coordinate syste-adopted in the model.
O As adjustment to account for mesh refinement was obtained b, compa ing 3-5 O
)
)
finite element stresses for a step shaf t with data reported from Peterson (3-3].
It was found that a factor of 1.08 needs to be applied to the finite element stresses.
3.2.2 Stresses Due to Torsional Loading There is no set of boundary conditions that can be applied to the mode' that will represent exactly the physical crankshaft under torsional loading.
l
)
Therefore, two separate sets of boundary conditions (Table 3.5) were ana-lyzed. Soundary conditions for Case 1 represent antisynenetric behavior of the main journal to torsional loading in the amial (a) direction and those for Case 2, synenetric behavior. For both cases, transmitted torque was simulate:
by applying a unit rotation about the axis of the main journal in the tnirc
)
plane of syawnetry of Figure 3-5.
Figure 3-7 illustrates the relative crank throw orientations best approximated by each of the two boundary condition cases.
Those crankpin
)
fillets adjacent to a throw on the same side of the main journal and in the same plane are best represented by Case 1 boundary conditions. Those adjacent to a throw on the opposite side of the main journal and in the same plane a e more closely approximated by the Case 2 boundary conditions.
Stresses 'i-fillets not represented by either of these situations (i.e., adjacent to a
)
throw not in the same plane) will fall between the two cases considered.
Stresses obtained from applying the unit torsional rotation were scale:
to represent maximum positive and negative torques of 251,600 and -144,60:
)
f t.-lbs. at Cylinder No. 5.
These stresses were then scaled by a facto of 1.08 to account for the slight finite element underprediction of the fillet stresses, due to the size of the elements used.
From the eight element integration points, stresses were entrapolate:
)
to the surf ace.
For Case 1 Figure 3 8 shows the circumferential variatioc, and Figure 3-9 shows the axial variation of maximum principal stress for botn the peak-positive and peak-negative torque conditions.
Figures 3-10 and 3-11 show similar variation for Case 2.
)
3-6
)
q 4
O All stress values that have been presented are for the positive z side of the crankshaft, as viewed in Figure 3 4.
In the crankpin fillet, this has also been designated as the 0* to 180* portion.
3.2.3 Stresses Due to Gas Pressure Loading Near TDC the pressure in a cylinder causes a high vertical load o* tne crankpin.
This load may be calculated from the pressure loading and re:1 pro.
3 cating inertial loading. The pressure load is calculated from the area of the piston and peak pressure of 1680 psi.
The reciprocating inertial loa: is obtained from the 820 lbs. of reciprocating weight and peak acceleration o' 74.1 3 The reciprocating inertial load subtracts from the pressure load at TDC.
The pressure loading was applied to the model as a distributed load o-the topmost three lines of noded points on the crankpin.
Two types of boan-da ry conditions were applied.
In both cases synenetry planes 1 and 2 (see O
Figure 3-5) were modeled by symmetric boundary conditions.
In the first case the third plane was modeled as a fixed support, and in the second case it was modeled as a pinned support..The actual moment in the main journal is greate*
than zero (pinned support) but less than the fixed-end moment (fixe support ).
The moment in the main journal may be estimated by treatin; t%
3 crankshaf t as a continuous beam with simple supports at the main bearing lo:a-tions. From this analysis it was determined that the moment in the main joga-nal was 0.63 times the fixed-end soment.
Since stresses for the fixed-en:
case were very small, the stresses due to the vertical loading were cal:ulate:
O as 0.37 times the stresses for the simply supported case. The maxim e stress occurs in the 180' location and was found to be 15.5 ksi.
The distribation o' stress around the crankpin is shown in Figure 3.12, 3
3.2.4 Comparison of Stresses with Test Data The SWEC test [3 2] recorded data from strain gages in the fillets o' Crant Throw hos. 5 and 7.
These gages were placed in the locations =nere stresses are a maximum due to torsional loading.
The measured stresses are
)
compared with those calculated by the finite element model in Tables 3.6 and 37 9
)
i
)
3.7 Good agreement is found between the test data and computed results. The maximum principal stress range of 44.9 ksi was measured in Crank Throw No. 5 and it is bounded by the two finite element results of Case 1 and Case 2.
At Crank Throw No. 7. the measured stress is slightly higher than the comp;te:
)
stress from Cases 1 and 2.
'The finite element stresses for vertical loadtng are in agreement wite stresses measured in T01's static test [3-4] on an inline 6 cylinder 13-ince by 11. inch crankshaft.
TDI determined the maximum stress due to vertical loading, after factoring for the difference in crankpin area, to be 16.3 asi at the 180* location. At this location, the torsional stresses are less tna-half of their maximum values.
Also, at the location of maximum torsiona' stresses, the vertical bending stresses are seasured to be 7.8 ksi and co -
)
puted to be 8.1 ksi.
At the No. 5 location, the transmitted torque is quite low during firing (see Figure 3 2), and thus, the highest stresses are not af fected by vertical bending stresses.
)
3.3 Crankshaf t Fatigue Failure phrgin The factor of safety against fatigue failure in the repla:ement (12. inch crankpins) crankshaf ts is calculated in this section.
Tne steess levels in the replacement crankshaf ts are computed from strain gage tes:
)
data.
The endurance limit is first established for the failed crannsna'is (11. inch crankpins) from strain gage test data. This endurance limit is tne-scaled to account for the higher ultimate tensile strength of the repla:e,e-:
crankshaft.
The effect of shot peening the replacement tranksha'ts provides
)
an additional margin against fatigue failure.
3.3.1 Stresses in Replacement Crankshaf ts The replacement crankshaf t was instrumented with strain gages in tne
)
fillet locations of Crankpin Nos. 5 and 7 and tested under operational con: -
tions at 3500 kW (100% rated load) and 450 rpm (2001 rated speed). The hign-est stresses were measured in Crankpin No. 5.
A dynamic model of the cran -
shaf t confirms that this pin undergoes the greatest range of torque.
Tnree-
)
dimensional finite element models of a quarter crank throw show that the 3-8
)
)
)
strain page rosette was placed in the location of highest stress, both witnin the fillet and around the crankpin. The following strains were measured at 3500 kW:
)
Strain Eage Maximan Miniane 5-1 (Compression)
-195pc 288uc 5-2 (Bending) 695uc
-410uc 5-3 (Tension) 737ue
-610we
)
To account for the simultaneous effects of shear and bending, the stress state is represented by equivalent stresses using Sine's metho:
[3-5).
For a biaxial stress state, the equivalent alternating stress, 5;,,
)
and equivalent mean stress, S,, are given by:
q
+Sy)A/2 S,=(5
-S 5
q and S,= 5,,
+ S,,
)
q where 5,, and 5,2 are the alternating components of principal stress, an: Sq are the mean components of principal stress.
From the test and 5,,
report [3-23, the equivalent alternating stress, S,,
and equivalent maa-q
)
stress, S,, on Crankpin No. 5 were calculated to be:
q S, = 24.6 ksi q
S, = 4.8 ksi q
)
Equivalent stresses, S, and S,, are those alternating and meam uniaca' q
q stresses that can be expected to give the same life as the given multiania!
stresses.
)
3.3.2 Endurance Limit for Original 13-Inch by 11-inch Crankshaft The original 13-inch by 11-inch crank shaf t was instrumented witn st air gages in the fillet location of Crankpin Nr.
5.
This fillet had previoasty
)
esperienced a fatigue Crack during performance testing.
After the test, the 3-9
)
D
(
)
three-dimensional finite element models of a quarter section of a crank throw showed that the strain gage location was placed close to the location of maximum stress. The measured stress range is used to establish the endurance limit in this analysis as a conservative assumption, although the actua' maximum stress range is revealed by the finite element model to be aboat 151 g
higher at a nearby location.
From the test report [3-6), the follo** m; strains were measured at 3500 kW:
Strain Gage Maximum Minimum 5-1 (Tension) 1118vc
-707uc 5-2(Bending) 773pc
-459uc 5-3 (Compression)
-389uc 266sc 3
The equivalent alternating stress, S,, and equivalent mean stress, q
Sqn, were calculated to be:
S, = 33.7 ksi q
J S
10.9 ksi qm From the test logs, it was determined that the shaft had experienced 273 no -s at equal to or greater than 100: Ioad, or about 4-
106 cycles.
By us r; Miner's rule and typical slopes of 5-N curves, it was determined that t't e 3
endurance limit for this mean stress was 32.4 ksi.
The ultimate teasite strength for these crankshaf ts averaged 96 ksi.
A line representing tes endurance limit is shown on the Goodman diagram [3-7] in Figure 3-13.
D This line is bounded by two lines showing the endurance Itmit for f.'!
scale crankshafts based on other test data [3-8].
3.3.3 Endurance Limit for Replacement Crankshaf ts D
The replacement crankshafts have a minimum tested ultimate teas le strength of 103 ksi.
Th= endurance limit scales linearly with ultimate te -
sile strength.
On this basis, the endurance limit for the replacement craa. -
Shaf ts is shown in Figure 3-13.
D 3-10 D
O O
The ft11et regions of the replacement crankshafts have been shot peened.
The effect of shot peening will produce increases in fatigue endurance limit (3.g].
3.3.4 Factor of Safety Against Fatigue Failure The factor of safety against fatigue failure of the replacement cran. -
j shaf ts is 1.48 when the effect of shot peening is not considered.
At 3800 kW, the strain gage test data [3-2] on the replacement crana -
shaft shows that the stress level is 41 greater than it is at 3500 kh.
At 3900 kW it would be about 55 greater than it is at 3500 kW.
Thus, there is a*
adequate safety margin against fatigue failure at the specified diesel geneaa-l tor set two-hour-per-24-hour period rating of 3900 kW.
l O
i O
O O
'O O
3-11 O
O P
Section 3 References 3-1 Timoshenko, S.,
D.H.
Young, and W. Weaver, Jr., Vibration Problems in Engineering. Fourth edition. Wiley, 1974 O
3-2 Bercel.
E., and Hall, J.R., " Field Test of Emergency Diesel Generator 103," Stone & Webster Engineering Corporation, April 1984 3-3 Peterson, R.E.,
Stress Concentration Fa ctor.
Wiley 8 Sons, Ne-Yor..
1974 3-4 "R-48 Crank Crankshaf t Stress Analysis," Transamerica Delaval Inc. Repo-t O
ho. CR-01-1983.
3-5 Fuchs, H.O., and Stephens, R.I.,
Metal Fatigue in Engineering.
- Wilej, 1980.
3-6 Bercel.
E.,
and Hall, J.R.,
" Field Test of Emergency Diesel Generator O
101," Stone 8 Webster Engineering Corporation, October 1983.
3-7 Collins, J. A., Failure of Materials in Mechanical Design. Wiley, 1981.
3-8 Nishihara, M., and Fukui, Y., " Fatigue Properties of Full Scale Forge and Cast Steel Crankshaf ts," Transactions of the Institute of Marine g
Engineering.
Series 8 on Component Design for Highly Pressure-charge:
Diesel Engines London, January 1976.
3-9 Burrell, N.K., " Controlled Shot Peening to Produce Residual Compressive Stress and Improved Fatigue Life," Proceedings of a Conference on Rest::-
ual Stress fo-Designers and Meta 11urgists, American Society for Metals.
April 1980.
D D
e D
3-12 D
1
)
)
TASLE 3.1 IIATURAL FREQUENCIES FOR D5R 48 131NCH BY 12-INCH CRANKSHAFT
)
Instural Frequency (Hz) 1 2.93*
)
2 38.73 3
92.94 4
116.67
)
5 184.33
- For $NPS emergency bus operation the natural frequency of the first mode is zero (i.e.,
rigid body mode). and the natura!
frequencies of the higher modes are not significantly altered.
)
TABLE 3.2 TORSIONAL LOADING FOR FaAA ANALYSIS
)
Order Torsional Landing. Tn (PSII
)
1.5 112.0 2.5 77.0
- 3. 5 48.0 4.0 33.0
)
4.5 26.2 5.5 15.5
)
3-13
)
J 3
taste 3.3 FREE-END VIBRATION AT 1001 LOAD FOR 05R-48 13-INCH 8Y 12-INCH CRANKSMAFT
<3 Amplitude of Vibration (degrees)
Order FaAA Analysis SWEC Test [3-2]
- D 0.5 0.065 0.056 1.0 0.001 0.005
- 1. 5 0.177 0.171 2.0 0.000 0.001
- 2. 5 0.142 0.130
)
3.0 0.001 0.001 3.5 0.061 0.058 4.0 0.340 0.325 4.5
- 0. 06 9 0.064 I) 5.0 0.031 0.034
- 5. 5 0.122 0.127 6.0 0.014 0.03E 6.5 0.014 O.016
[)
7.0 0.002 0.032
- 7. 5 0.001 8.0 0.015
. vector Summation 0.662 0.693
- )
J J
3-14 4
)
)
TABLE 3.4 TORQUE RANGE AT 1001 LOAD FOR j
p$R.4813. INCH BY 12. INCH CRANKSHAFT
)
Amplitude of Torque Range IIO"I"8I 3h
Location (f t. Ibs.)
Stress (psi) 4th Order
- Total, 4th Order Total
)
between Cylinder No. 1 36.6 = 103 167.1 = 103 648 2955 Cylinder No. 2 Between Cylinder No. 2 69.0 = 103 184.5 = 103 1223 3263 Cylinder No. 3 Between Cylinder No. 3 100.0 = 103 271.1 = 103 1763 479 Cylinder No. 4 Between Cylinder No. 4 29.0 = 101 309.8 = 103 2282 5:75 Cylinder No. 5 Between Cylinder No. 5 155.6 = 103 396.2 = 103 2752
'Oh
)
Cylinder No. 6 Between Cylinder No. 6 178.8 = 103 327.3 = 103 3162 575!
Cylinder No. 7
)
Between Cylinder No. 7 198.6 = 101 329.7 = 103 3512 5 9."
Cylinder No. 8 Between Cylinder No. 8 214.2 = 103 311.8 103' 3792 551' Flywheel
- 5WEC test [3 2] computed the torque ran9e to be 357.1 = 103 ft..Ib.
inis in:t-Cates a stress concentration factor of 1.145 due to the proximity of the gage to the flywheel hub.
)
3 15
)
)
h TABLE 3.5 DISPLACEMENT BOUNDARY CONDITIONS FOR TORSIONAL LOADl%
(REFER TO FIGURE 3-5)
Case 1 S
l Nodal Degrees of Freedom Synnetry P1ane D
I 1
Fixed Fixed Free 2
Free Fixed Fimed l
g 3
Free Prescribed
- Prescribe
- D Case 2 D
Nodal Degrees of Freedom Syssetry Plane X
Y Z
D 1
Fixed Fixed Free 2
Free Fixed Fi xe:
3 Fixed Prescribed
- Prescribed
- y
- Prescribed displacements were used on synenetry plane 3 to simulate torsio*d' load on the main journal, D
3-16 D
Table 3.6 COMPARISON BETWEEN FINITF ELEMENT MODEL TORSIONAL LOADING REStJLTS AND TEST RF5tfL T5 FOR PIN MIMRER 5
(
r Range W Range of I
Peak Itegative Torque Peek Positive Torque Principal Stresses (kst)
Principal Stresses (kst)
Principal Equivalent Stress Stress
- 1
'2
- I
'2 (kst)
(kst)
Finite Element 20.7
-18.0 10.4
-11.9 32.6 52.9 Case 1 Finite Element 29.2 1.2
-0.7
-16.8 46.0 45.1 Case 2 Strain Gage [4]
26.2
-2.9 4.9
-18.7 44.9 49.3 i Peak positive torque = 251.6 = 103 ft. lb.
3 ft.-lh.
2 Peak negative torque = -144.6 = 10
v v
v-v v
v-m Table 3.7 COMPARISON BETWFEN FINITE FLEMENT MODEL TOR $IONAL LOADING RESULTS AND i[$T RIStiliS FOR PIM MIMRER 7 z
Range of Range of i
Peak Itegetive Torgue Peak Posittwe Torque Princtpal Stresses (kst)
Principal Stresses (kst)
Priactpal Equivalent Stress Stress
- 2
- 2 (kst)
(kst)
Finite Element 18.7
-16.3 7.3
-8.3 27.6 43.9 Case 1 Finite Element 26.5 1.1
-0.5
-11.8 38.3 37.5 Case 2 Strain Gage [4]
23.4
-8.9 2.8
-14.1 37.5 44.5 i Peak positive torque - 251.6 = 103 ft.-Ib.
3 ft.-lb.
2 Peak negative torque
-144.6 = 10 j
i
,0 l
O O
PRES 5URE (PSIG; I
eese< t s
I e tte a
g g
i ii i
ll' i
i
'me-l
/g 0
t I
1 I
r <
lJ.lj'\\
O I
ss
.A
- 0; if0:
18 O
CR A% - AN.E (DEGREES-l E
Figure 3-1.
Measured pressure versus crank angle at 100t load.
O J
3 1
Fa A A ad 16 i
G
9 9
9 1
i ate.llsam. 580, 0 na.sl.eem. s.p) O.
n
,i in.
s s.
n
-o t
.s I
=
G.
=--?
Tt
?~
-i.t T*
o 7.
T.
2
.s i..
- a..
.w s.
(
7,.
7,.
(444g agrtt eDir.tf $3 re..g gartp rDirertgo l
b) Cylinder 2 to 3.
a) Cylinder I to 2.
"'483'em 707 0 I
sp0 0 t'
MIgleam. ice 0 see:I*a.m.
aleta m-see O
- n 3<
-a
-*2 R
n
~
3 l
.G '
~
l T1 l
ft L~
q~
i.t o'
it l
n' e
.t
.w to l
g
.ti, w e i..
+
_r 1
.,........n..
.ie d) Cylinder 4 to 5
}
e
() Cylinder 1 to 4 l
model torsional response at 100% load for Shoreham 13 inch by 17 inch crankshaft.
l Fiqure 3-?.
Dynaml(
I L_._.__
l i
'N l
-)
l I
3 8
'?
,e 3,.)
e E
k k
7 N
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Q
(
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e.
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8 8
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v
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(
es cc
't.
..8.y so.
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's w s,.
s Z;
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e.,
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c Kr 3C Et i1 ry v
Et EP Is m
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er p m
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g
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.c r." '#
4 5 e t s..'43~,.
- sete, a
- e:
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.C.
C:.
end CC
)
v
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3 6
6 3
'2 5
3 8
m C
i 8
f, f 1
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i
a It, 5r II EI 37 PF
.s. !
FF 37 C
3 m
1 E
es sin er er se i.
er e er e.
en er.*
or es er er er e
,e ee
.Sta l$dle ass M sC4 J avan
.O l e t $ e l h. 4 s
- 3ece:,
.s==.
Fe A A-84 t o O
T f
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b
)
SS T ft-k e p s*
m E
i 5o" W
l
)
I i
i 1
1 i
i i
S.00 180.00 980.00 640.00 F20.00 CRANK ANGLE (degrees)
)
a) Nessured.
)
2 4
2
?
l b
311 f1-k i p s -
E O
i t
)
i e
a i
1 1
1 n
9.00 180.00 940.00 640.00 720.00 CRANK ANGLE (deyees)
)
b) Calculated.
)
Figure 3-3.
Casparison of measured and calculated torque near the flywheel.
- 5ee test for explanation of difference.
)
r.aa.ma.$.,a
N
.see oE a,d
)
" '8 Detail A J
t 4
3
,,,e 1
~
iE e:
l
)
8
- I
- 3 at a-e l
I I,
s
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s o. - /
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e 1
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e I
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1 Detail A l
t l
\\
s f
)
e s. / s._
l
,j ' - ' ' "
\\
l l
'88S"
~4s_
=
to s
)
Figure 3 4.
Typical structural dimensions of the crankshaf t.
Fe A A.ed.3. t e
)
/
A 9
L C
S.=3 C
-Se I
Y 3
DCE e
a w
e>
m a
c U*
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y G
G 8
5 Mo W
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,'!*"% 7, -. 'u, y
/ ' "27,{,
. '" r.
e s
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d N
5
. t..,; ?l.:
- p..
n
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)
>=
li; f
L
/
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/
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j
/
/
J
- sf. 'Q's%s e
A. W e
s
/ '\\$'+
g g
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E
(_ \\ A gs
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s
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y
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o
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g
==
e-2 e
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a C
sr.'
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e, U
S
- g W
.=.
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w
>=
FeAA*S4-3-18
]
D e
I!-
s g'.1 sii 1 $
/'
t new
\\'\\\\
h rm\\'.'N s
s
,y
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i
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D Figure 3-6.
Three-disensional view of the finite element model.
D F s A A
- S d
- 3 16 D
O
.O l_ Section _l r
een e -
l t
I, I
l I
I f
I l,
I
'O l
i i
8
.I l
i l
I I
a f
a d
I L
I L
,Q
, Soundary Cbndition, Case 1 (a)
)
3 I
F 7
O F
i l
_I j
i e
I 1
I l
l l
I O
i y
i b Section _
~
l j'modesee i 1
(b) Soundary Condition. Ca.e 2 O
I Ge. erne
- Generator Enc End Cyl6nder 6 i
3
,1 j i, I
I 6
',i O
l l-L I
~
criineer a criander 4 (1 don'etes fillet location best e.,rosameted by case 1
..onet e.... e,.......n to.t e.,r....a t.d,, C.. e,,
(c) Actves O
Figure 3-7.
Schematic of relative crank throw orientations for (a) F.E. flodel, Case 1; (b) F.E. Indel Case 2; and (c) the actual crankshaft.
F a A A
- S4 10 0
)
E i
i i
i e
i i
e i
i we.
. 0.14-*
40
)
' 1.0 b
~
~
j
)
Crank pin e6 g
Fillet maximum 20 -
l N
e S
20 L.
[
I N
6 Midsurf ace msnimum l
g I
so 0
W 1
Midsurf ace minimum d
i
)
C I,7_______--_______,
1
-10 N
i Fillet minimum
-20 O
20 40 40 80 100 120 140 160 180 200 220 i
l ANGUL AR POSITION (degrees)
Figure 3-8.
Circumferential variation of manimum principal fc-torsion Case 1 boundary conditions.
Fa A A-84 3 1s
)
I j
2 i
i i
ie i
i i
Web 10.14 -*
40 i
l t
' 1.0 I
f 0
)
}
D'"
.e W
~
)
20 -
l m
t i
eo 7 aximum stress M
E
[
10 -
)
mu Ee3 to O
w
)
N I
w Minimum stress
-10
)
-20 O.2 0.4 0.6 0.8 t.0 1.2 1.4 1.6 18 20 22
)
SURFACE DISTANCE (in )
Figure 3-9.
Axial variation of maximum principal stress for to-sie-Case 1 boundary conditions.
FeAA-84-3-10 m
)
s0 litet Peak positive tefeue senditions
)
30
=
w 10
)
aw g
5 m
N Crank pin asidpoint E
O a
N Z
1 E
e.
W N-10
)
c Peak negative Fillet torque conditions
-20 O
20 40 00 SO 100 120 140 160 180 ANGUL AR POSITION AROUND CR ANK PIN (degrees)
(0 la top of pin at TDC)
Figure 3-10.
Circwnferential variation of maatmc principa stress for torsion Case 2 boundary conditions.
)
)
F aA A - 8 4 16
)
O f
e s0 Web 0.14
- 1.6 g
40 t
l 1.0 l
3 m-f ~
-1 Crank pin i
e m
5
- r
~
3 f
o f
Maximum stress N
{
10 3
I
.u 6a Dm O
l)
W j
6 l
Minimum stress i
-10 b
1 l
l
-20 O.2 04 0.6 0.8 1.0 1.2 1.4 1.6 18 20 22 SURFACE D!$TANCE On) g Figure 3-11.
Amial variation of maximum principal stress for torsion Case 2 boundary conditions.
Fe A A-se 16
llll1lll.
o u
g m
n i
u d
m a
e o
in e
l u
i m
em 0
e feu I
0 r
wm t
1 u
e-oi s
l dn s
.l ii e
i r
.,F-Mm p
\\
L r
s sp-a g
e u
c r
am' m p" o
u f
f re s
ei 6
s n
d Mmg,-
3 e
io r
m 1
t u
)
s m
s C.
t l
e e
i a
a e
p l
r l a i
g iFm c
e n
d i
(
rp N
O mu I
m 0T l
9I u
S a
O m
P f
=
o R
A n
o L
i U
t G
a i
N r
A a
v 5
l 4
a i
tn er
,/
e f
m,******* s m
so
=
r i
C 7
O 1
1 oru q
i F
0 0
0 2
1 1
g s* eawag a.2g3a.
a
=I> b. ? 3 llll1 l
l
)
I.
j I
.o S
.a =
{
t
)
.'a I
F!
y n
i I
I I
Du 8
U a
C C
t
=
v g
s
)
s.
=
=
/
I,
- ~
E
~
W E
~
=
/ /
h
.n W
L
,/
=
)
23, 31 fg'
-5.
o
=
g
- =,
l m_
j a315
//,'
I
~
l u
=..
e,
- =.
[<'<.:.
=
a
..-e-
)
12* 3:
==
35 j
3m*
//
_i,25 g
S
,m
/ja,t
.=.
A
..=
. =.
, /,,..n.
a
)
/ ',,
i
~
- ~21 i
=.
i
- .a
.. 9 O*..
/
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EDUS
- 1
,u n
)
I.
e g wf 9
M (iem as3 mis Om:1vMu311v AN37vAiriO3
)
FaAA-44-3-18
)
)
)
4.0 DISCUSS 10H AIS GNELUSIONS DSR-48 engines with 13-inch by 12-inch crankshafts are in diesel gener-ator set service at seven other locations as shown in Table 4.1.
This date
)
shows that there has been extended service (long enough to produce more thaa 10' stress cycles) on several engines with 80% to 941 load, and limited ser-l vice at 100% load.
)
The fillet regions of Crankpin Nos. 5 through 8 of the Shoreha-re-placement crankshaf ts were eddy current tested after 102 to 114 hoses o'
operation at 100% or greater load (see Table 4.2).
to relevant indicatices were found. Thus, there are no cracks in the high stresses fillet regions.
)
The drawing of the replacement crankshaft has been certified by t't e i
American Bureau of Shipping for compliance with their rules [4-1] for sizing of the pins, jcurnals, and webs.
The following conclusions are made:
)
1.
The design calculations on the 13-inch by 12 inch tranksha'ts performed by TDI are appropriate and shp that the cra9.she't stresses are below DEMA reconenendations for a single orce.
Combined stress is not calculated by this method, bat ma.,
be
)
determined by torsiograph testing.
2.
The SWEC torsiograph test results show that the 13-inch by 12-in:n crankshaf t stresses are below the DEMA receirnended levels for beim single order and combined orders for both 3500 kW (100*. rate:
Ioad) and 3800 kW.
A linear extrapolation to 3900 kW also sho s compliance.
)
3.
The factor of safety against fatigue failure was found to be 1.49 if the effect of shot peening the fillet regions is ignored and is even greater if the shot peening of the $noreham cranksha'is is considered.
)
4.
The replacement crankshafts are suitable for unlimited ope atio-in the emergency diesel generators at SNPS.
)
4-1
)
)
)
Section 4 References 41 American Bureau of Shipping, Rules for Building and Classing Steel Vessels, New York, 1984
)
)
)
)
)
)
)
4-2 1
)
)
TABLE 4.1 AVAILABLE LOGGED HOURS OF OPERATION OF OSR 48, RATED 3500 KW 9 450 RPM Kilowatt Total Avera9e Serial Rating 9 Hours Date Load Other Loads and
)
Number Location 450 rpm Logged Logged Reported Hours Reported 74010
$NPS 3500 368 3-21-84
>3500 kW for 114 hrs.
74011 430 2-13-84
>3500 kW for 116 nes.
74012 345 3-14-84
>3500 kW for 110 nes.
)
75005 K00SHE'4G,
3600 246 3-15-84 Mostly 75006 TAlWAN 221 3-15-84 1001 75007 368 3-15-84 7500S 299 3-15-84 76010
- DHU8A, 3500 19800 3-17-84 76011 SAUDI 23300 3-17-84
).
1 76012 ARABIA 23800 3-17-84 l
76013 19700 3-17-84 l
76014 23500 3-17-84 3000/3200 kW for 9000 nas.
76026
- ONE12A, 3515 16204 3-17-64 76027 SAUDI 12428 3-17-84
)
76028 ARA 81A 14978 3-17-84 78029 U.of 3500 8180 3-15-84 1100 kW 78030 TEXAS 5385 3-01-84 1100 kW 78044 WADI 3515 10882 3-17-84 2200/3000 kW 78045 OAWAS!R, 10832 3-17-84 2200/3000 kW
)
78046 S. ARA 8!A 11212 3.17-84 2200/3000 kW 79002
- RAFHA, 3515 12667 3-16-84 3300 kW fo 6200 n s.
3200 kW f o 8250 nrs.
79003 SAUDI 11655 3-16-84 3200 kW for 550^ n-s.
79004 ARA 8IA 13186 3-16-84
)
80001
- RA81GH, 3515 10196 3-16-84 2700 kW 80002 SAUDI 10245 3-16-84 2800 kW 80003 ARA 8IA 11602 3-16-84 2800 kW
)
)
4-3
)
)
b l
i J
)
TABLE 4.2 HOURS OF OPERATION OF $NORENAM REPLACEMENT CRANKSHAFTS AT TIME OF ED0Y CURRENT TESTING I
)
HOURS OF OPERATION AT LOADS > 100: RATE; DIESEL GENERATOR AT ALL LOADS LOAD 101 368 114
)
102 281 102 103 345 110
)
)
)
)
)
)
44
)
lD 3
DR-03-3101 C94PONENT DESIGN EVIEW CRANKSHAFT Classificatioa A PART NO. 03-310A Completion 3/5/B:
O PRilt4AY FINICTION:
The crankshaft converts reciprocating motion, component inertial forces an:
gas pressure piston forces to rotary motion and torque at the output flange.
FUNCTIONAL ATTRIBUTES:
p 1.
Structural stif fness of the crankshaf t must be suf ficient to mainta'-
acceptable states of stress in the crank pin web rad main journal areas and to maintain system natural frequencies which are sufficiea.tly removed from engine operating speeds.
The crankshaft design sno.1 :
also be sufficient to withstand normal main bearing misalignments g
inherent in service.
2.
The journal area of the main and connecting rod (crank pin) bearin; must be suf ficiently large for pr'oper bearir.g oil film pressure bat tne journal length must be suf ficiently short to prevent end wear of tne bearing sleeves.
S 3.
The material of the crankshaft and the surface finish should be suf ficient to resist f atigue crack initiation.
SPECIFIED STAlWARDS:
1.
IEEE D
2.
ASTM 3.
DEMA EVALUATION:
E 1.
Review TDI calculations and tests 2.
Conduct engine test of 13 x }2 shaft 3.
Conduct modal superposition and Ho12er torsional analyses of :
a.
$NPS (R 48) p b.
GGNS (RV-16) c.
Midland (RV-12) d.
San Onof re (RV-20) 4 Conduct finite element analysis of R 4812-inch crankpin fillets 5.
Compare measured and calculated stresses R-4813 m 12 shaft D
O
/
O
'f 6.
Compare seasured and calculated output torque and free end torstograph traces for R-48.
s 7.
Compare stress Tevels with enourance limit for B-48 Sa.
Compare nominal stresses of R-48 and RV-16 with those reconenended by O
other standards.
b.
Compare nominal stresses of RY-12 and RV-20 with those reconrnended by various organizations.
9.
Complete final report on SNPS and GGNS crankshaft integrity.
O 10.
Complete final report on Midland RV-12 and San Onofre RV-20 REVIEW TDI ANALYSES:
1.
Experimental stress analysis (static) of DSR-46 crankshaf t 2.
Torsiograph tests 3
3.
Holzer Table calculations IWORMATION REQUIRED:
1.
TDI drawings for 05R-48 and RV engines 2.
Test reports for DSR-48 and RV engines
\\
3.
Original Holzer calculations 2nd revisions for R-48 and RV-16. RV-12 and Rv-20 engines 44.
Experimental pressure vs. time curve for R 48 and RV-16 engines.
3 b.
Emperimental pressure vs. time curve for Rv-12 and Rv-20 engines.
D D
D
.