ML19270B896
| ML19270B896 | |
| Person / Time | |
|---|---|
| Site: | Perry, Catawba, Grand Gulf, Comanche Peak, Midland, 05000000, Shoreham |
| Issue date: | 03/23/1984 |
| From: | Eisenhut D Office of Nuclear Reactor Regulation |
| To: | Gilinsky, Palladino, Roberts NRC COMMISSION (OCM) |
| Shared Package | |
| ML19270B897 | List: |
| References | |
| TASK-AS, TASK-BN84-063, TASK-BN84-63 BN-84-063, BN-84-63, NUDOCS 8403140249 | |
| Download: ML19270B896 (5) | |
Text
6 dca) 4/es March 23, 1984 Docket Nos.:
50-322 and 50-329/330 MEf10RANDUM FOR: Chairman Palladino Commissioner Gilinsky Commissioner Roberts Commissioner Asselstire Commissioner Bernthal FROM:
Darrell G. Eisenhut, Director Division of Licensing
SUBJECT:
BOAPD NOTIFICATION 84-063 REPORTS SUBMITTED OF THE TRANSAMERICA DELAVAL, INC. (TDI) OWNERS GPOUP In accordance with procedures for Board Notifications, the following information is being provided directly to the Commission. The appropriate Boards and Parties are being informed by copy of this memorandum. This information is relevant to all facilities that have diesel generators manufactured by TDI including Midland and Shoreham which are currently before the Commission.
On March 12, 1984, the TDI Owners Group submitted for staff review the Design Review Report of Connecting Rod Bearing Shells for TDI Enterprise Engines.
This report is included as Enclosure 1.
On March 13, 1984, the TDI Owners Group submitted the TDI Diesel Generator Rocker Arm Capscrew Stress Analysis Report. This report is included as.
The staff is reviewing both reports as part of its overall assessment of TDI diesel generators.
Oricinn rima by Darrell G. Eisenhut Darrell G. Eisenhut, Director Division of Licensing
Enclosures:
As stated cc w/ enclosures:
See next page TDI:Pp T I:
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The Commissioners cc w/ enclosures:
SECY (2)
OPE EDO OGC Parties to the Proceeding ASLB Shoreham (Brenner,Ferguson, Morris)
Catawba (Kelly, Foster,Purdom)
Perry (Bloch, Bright,Kline)
Comanche Peak (Block, Jordan,McCollom)
Midland (Bechhoefer,Cowan, Harbour)
ASLAB Shoreham (Rosenthal,Edles,Wilber)
Catawba (Rosenthal, Moore,Wilber)
Perry (Kohl, Buck,Edles)
Midland (Kohl, Buck, Moore)
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DISTRIBUTION LIST FOR BOARD NOTIFICATION Catawba Units 1&2, Docket Nos. 50-413/414 William C. Potter Jr.,Esq.
Comanche Peak Units 1&2, Docket Nos. 50-445/446 Dr. Paul W. Purdom Midland Units 1&2, Docket Nos. 50-329/330 Mr. Paul Rcu Perry Units 1&2, Docket Nos. 50-440/441 Harold F. Reis, Esq.
Shoreham Unit 1, Docket No. 50-322 W. Taylor Reveley III, Esq.
Nicholas S. Reynolds, Esq.
Dr. Peter F. Riehm Atomic Safety and Licensing Dr. Richard F. Foster Mr. Jesse L. Riley Board Panel Leca Friedman, Esq.
Ken Robinson, Esq.
Atomic Safety and Licensing Eleanor L. Frucci, Esq.
Alan S. Rosenthal, Esq.
Appeal Panel Steve J. Gadler, P.E.
Cherif Sedky, Esq.
Docketing and Service Section Mr. R. J. Gary Ralph Shapiro, Esq.
Document Management Branch Stewart M. Glass, Esq.
Mr. Frederick J. Shan MHB Technical Associates Mr. Marc W. Goldsmith Jay Silberg, Esq.
Palmetto Alliance Robert Guild, Esq.
Ms. Mary Sinclair Dr. Jerry Harbour Mr. Lanny Alan Sinkin Martin Bradley Ashare. Esq.
Mr. Bruce L. Harshe Ms. Barbara Stamiris Edward M. Barrett, Esq.
Samuel A. Haubold, Esq.
Howard A. Wilber, Esq.
Charles Bcchhoefer, Esq.
Mr. Wayne Hearn Mr. Donald R. Willard Ms. Lynne Barnabei Ms. Susan Hiatt Mr. Frederick C. Williams Howard L. Blau, Esq.
Renea Hicks, Esq.
Richard P. Wilson, Esq.
Peter B. Bloch, Esq.
Dr. W. Reed Johnscn Ms. Nora Bredes Dr. Walter H. Jordan Lawrence Brenner, Esq.
Mr. James R. Kates ACRS Members Mr. Glenn 0. Bright Mr. Frank J. Kelley Dr. Rober C. Axtmann Herbert H. Brown, Esq.
James L. Kelley, Esq.
Mr. Myer Bender James E. Brunner, Esq.
Dr. Jerry R. Kline Dr. Max W. Carbon Mr. Ronald C. Callen Christine N. Kohl, Esq.
Mr. Jesse C. Ebersole Jota G. Cardinal, Esq.
Stephen B. Latham, Esq.
Mr. Harold Etherington Gerald Charnoff, Esq.
James A. Laurenson, Esq.
Dr. William Kerr Myron M. Cherry, p.c.
Dr. J. Venn Leeds, Jr.
Dr. Harold W. Lewis John Clewett, Esq.
Mr. Howard A. Levin Dr. J. Carson Mark Hon. Peter Cohalan Stephen H. Lewis, Esq.
Mr. William M. Mathis Mr. John T. Collins Terry J. Lodge, Esq.
Dr. Dade W. Moeller Barton Z. Cowan, Esq.
Ms. Karen E. Long Dr. David Okrent Dr. Frederick P. Cowan Dr. Emmeth A. Luebke Dr. Milton S. Plesset T. J. Creswell, Esq.
Mr. Wendell H. Marshall Mr. Jeremiah J. Ray Gerald C. Crotty, Esq.
Mr. Brian McCaffrey Dr. Paul C. Shewmon Mr. James E. Cummins Dr. Kenneth A. McCollom Dr. Chester P. Siess James B. Dougherty, Esq.
J. Michael McGarry III, Esq. Mr. David A. Ward Mr. Jay Dunkleberger Janine Migdeni Esq.
Mr. Anthony F. Earley, Jr.
Mr. Marshall E. Miller Gary J. Edles, Eso.
Michael I. Miller,lEsq.
Mrs. Juanita Ellis Dr. Peter A. Morris Peter S. Everett, Esq.
Mr. Chris Nolin Donald T. Ezzone, Esq.
Fabian G. Palomino, Esq.
Jonathan D. Feinberg. Esq.
Spence Perry, Esq.
Dr. George A. Ferguson William L. Porter, Esq.
k LIST OF ADDRESSES RECEIVING MATERIAL ON THE FOLLOWING DOCKETS CATAWBA Mr. H. B. Tucker, Vice President North Ccrolina MPA-1 Mr. F. J. Twogood Mr. J. C. Plunkett, Jr.
Mr. Pierce H. Skinner North Carolina Electric Membership Corp.
Saluda River Electric Cooperative, Inc.
Mr. Peter K. VanDoorn Mr. James P. O'Reilly Spence Perry, Escuire Mark S. Calvert, Esq.
COMANCHE PEAK Robert A. Wooldridge, Esq.
Mr. Honer C. Schmidt Mr. H. R. Rock Mr. A. T. Parker MIDLAND Mr. J. W. Cook Stewart H. Freeman Ms. Julie Morrison Mr. R. B. Borsum Mr. Don van Farrowe Resident Inspector's Office Mr. Paul A. Perry Mr. Walt Apley Mr. James G. Keppler Mr. Ren Callen Dr. Steven J. Poulos Billie Pirner Garde P. C. Huang Mr. L. J. Auge Mr. Neil Gehring Mr. I. Charak Clyde Herrick Mr. Patrick Bassett
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2 PERRY Mr. Murray R. Edelmen Donald H. Hauser, Esc.
Resident Inspector Mr. Jenes G. Keppler SHOREHAM Mr. M. S. Pollock Resident Inspector /Shnrehan NPS Energy Research Group, Inc.
Mr. Janes Rivello Ezra I. Bialik Dr. H. Stanley Livingstone
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ENCLOSURE 1 Failure twestERe4 AmeD ast7ALLuscrAi CONSJLiANT$
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- mic ALTO Catrosue.a tax:3.sts.ewswoo TiLta 7042i6 F a AA-M- 3 PA0 7396/LAS-M&T-3A DESIGN REVIEW OF CONNECTING ROD BEARING SHELLS FOR TRAMSAMERICA DELAVAL ENTERPRISE ENGINES THIS REPORT IS FIKAL PENDING CONFIRMATORY REVIEWS REQUIRED BY FaAA's QLIALITY ASSURANCE PROCEDURES Preparad by Failure Analysis Associates Palo Alto, California Prepared for TDI Diesel Generator Owners Group March 12, 1984
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1 PA0 7396 Task No. 03340B TABLE OF CONTENTS Page STATEMENT OF APPLICABILITY...............................................
ii E X E C U T I V E SU MM AR Y........................................................ i i i 1.0 INTR 000CTION......................................................
1-1 Sect i on 1 Re f e ren:e s..............................................
1-3 2.0 EXAMINATION OF BEARING SHELLS.....................................
2-1 2.1 Nondest ru ct i v e Exami nat i on..................................
2-2 2.2 Dest ru cti ve Exami nati on.....................................
2-5 2.3 El e ct ron Mi c r o s c o py.........................................
2-5 2.4 Chemi cal C ompos i ti on........................................
2-6 2.5 Tensile Properties..........................................
2-6 Se cti on 2 R e f e re n c e s..............................................
2-8 3.0 J OU R N AL ORB I T AN AL Y S I S............................................
3-1 3.1 DSR 48, 11-inch Crankpin....................................
3-1 3.2 D SR 48. 12-i n c h Cra n k p i n....................................
3-2 3.3 U SRV-16 4, 13-i n ch C ra n k pi n................................. 3-2 3.4 Inte rpretati on of Journal Orbi t Analysi s....................
3-3 Se c t i on 3 Re f e renc e s.............................................
3-4 4.0 F I N ITE ELEMENT STRESS ANALYSI S....................................
4-1 4.1 OSR 48, 11-inch Crankpin....................................
4-1 4,,2 OSR-48, 12-i n ch Cra n k pi n....................................
4-4 4.3 DSRV-16-4, 13-i n ch Cr a nkpi n.................................
4-5 Se ct i on 4 Re f e re n c e s..............................................
4-6 5.0 FATIGUE ANALYSIS..................................................
5-1 5.1 Fati gue Li fe Cal cul ati on....................................
5-1 5.2 Thres hol d St ress Int ensity Range............................ 5-3
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5.3 Acceptance Criteria For Connecting Rod Bearings.............
5-4 Se ct i on 5 R e f e re n c e s..............................................
5-5 6.0 DI SCUSSI ON AND CONCLU SIONS........................................
6-1 APPENDIX: Ta s k De s cri p ti on.............................................. A-1 i
i
.i STATEMENT OF APPLICABILITY The analysis and conclusions in this design review apply to connecting rod bearing shells used in the Transamerica Delaval, Inc. standby diesel a
generators at Long Island Lighting Company's Shoreham Nuclear Power Station and at Mississippi Power & Light's Grand Gulf Nuclear Station.
Other Owners Group members may, at a future date, request rev1ew of other individual diesel generator units.
At that time this analysis may expand, and be revised to the extent necessary to be applicable to those units.
O r
.h EXECUTIVE
SUMMARY
CONNECTING R0D BEARING SHELL REPORT This report addresses the design and expected life of the Transamerica Delaval, Inc. (TDI) cast aluminum connecting rod bearing shells used in DSR 48 and *)SRV-series standby diesel generators.
The re>crt was prepared for the TDI Diesel Generator Owr.ars Group as one of a series of reports on generic components of tnose diesel engines ;n nuclear instGlations., the generically ta*med Phase I components.
The design review shows that the new 12-inch diameter connecting rod bearing shells recently installed in the DSR-48 diesel generators, and the
- 3-inch diamet'fr bearing snells used in the DSRV-series diesel generators, are predicted to have a fatigue life of 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> at full load, whi ch is greater than ten times that required during the 40-year service life of a nuclear power station.
On this basis it is concluded that the connecting rod bearing shells will function reliably in nuclear standby applications.
Four of the forty-eight 11-inch diameter connecting rod bearing shells originally installed in the Shoreham Nuclear Pcuer Station (SNPS) DSR-48 diesel engines had developed cracks after about 600 to 800 hours0.00926 days <br />0.222 hours <br />0.00132 weeks <br />3.044e-4 months <br /> of operation, including about 250 hours0.00289 days <br />0.0694 hours <br />4.133598e-4 weeks <br />9.5125e-5 months <br /> of fuli-load operation.
The cracks w.re found in the course of inspections following post-crankshaft-failure disassembly.
Although these cracks did not affect engine operation, all the SNpS DSR 48 diesel engines were modified with new crankshafts that used larger, 12-inch diameter, connecting rod bearugs.
Laboratory investigation of the cracked 11-inch bearings showed them to be of the proper chemistry and ultimate tensile strengt%
Metallurgical and analytical evaluation of the bearing shells irdicated three causes contributing to the observed cracking:
(1) the geometry of the connecting rods and the bearing shells which resulted in unsupported areas at the ends of the bearing shells, (2) high peak oil film pressure, and (3) edge loading of the bearings which resulted in concentration of the ope. rating loads on the.
unsupported bearing ends.
In addition-scaraing electron microscooy of the fracture surface of one of the cracked bearings originally installed in the
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3, Shoreham DSR 48 engines revealed voids abeut 0.025 inch in diameter that appeared to be the initiation sites for the fatigue crack. Such voids are not uncommon for cluminum castings of the type used in connecting rod bearing shells.
The replacement 12-inch bearing shells and connecting rods installed with the new crankshafts in the SNPS standby diesel generators are of a different design than the original components.
The design modi fication addresses each of the conditions identified as contributing causes of the cracking.
First, the geometey of the connecting rods and the bearing shells
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was changed to provide complete support to the bearing shell ends.
- Second, the increase in diameter to 12 inches reduced the peak oil film pressure.
Third, although the edge loading was not affected, the two previous changes reduce the stress caused by the edge loading to an allowable level for the bearing shell material.
The effect of this improvement was quantified by finite element stress analysis of the 11 and 12-inch bearing shells, using the results of journal orbit analysis to determine the detailed loading of the bearing shells.
The maximum tensile stress found in the 11-inch bearing was reduced by about 50% in the 12-inch configuration.
Two analyses were performed to determine the effect of the stress reduction on the fatigue resistance of the new 12-inch bearing shells.
A stress vs. number of cycles equation predicted that, based on the observed life of the 11-inch diameter bearing shells, the 12-and 13-inch shell fatigue life should be approximately 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> at full load, which is over ten times the usage expected over the 40-year service life of the nuclear standby diesel generators. An alternative analysis demonstrated that the decrease in the stress range is sufficient to prevent fatigue cracks, which indicates an infinite fatigue life for the bearing shells.
Based on fracture mechanics analysis, an acceptance criterion for discontinuities in the aluminum was established.
Voids up to 0.050 inch in diameter will not compromise the fatigue performance of the 12-inch and 13-inch connecting rod bearing shells in DSR a8 or DSRV-16-4 standby diesel generators. In addition to the standard; manufacturer's recommended periodic bearing shell inspections
[1],
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radiographic NDE of the bearing shells, on a sampling plan, will be recommended to assure compliance with the acceptance criterion.
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N, St r u rf References 1.
Transamerica Delaval Instruction Manual, Moce! DSR 48 Diesel Engine, Tecial Nos. 74010 - F64, 740'.i - 2605, 74012 - 2606, Transamerica Delaval Inc.. Engine and Compressor Division.
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1.0 INTRODUCTION
Transamerica Delaval Inc.
(TDI) Enterprise diesel engincs for are equipped with journal connecting rod bearings which are hydrodynamically lubricated.
The bearing surf ace itself is installed in the connecting rod in the form of shells made from cast aluminum-6% tin, Alcoa [1-1] alloy B850 with an inner surface layer of electroplated lead-base babbitt. F'gure I shows the schematic representation of a connecting rod bearing half shell and indicates the nomenclature used to describe its features.
The engines originally supplied to SNPS had 11-inch diameter bearings, but were subsequently modified with new crankshafts which had 12-inch bearings.
DSRV-16-4 diesel generators at Grand Gulf use 13-inch diameter bearings.
The prima ry function of these connecting rod bearing chells is to provide a siiding surface between the connecting rod and the crankpin of the crankshaft through the formation of a hydrodynamic oil film.
The connecting rod bearing shells transmit the cylinder firing pressure to the crankshaft through this oil film, converting the energy of the combustion process into torque to turn the electrical generator.
In order to perform this function,
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the bearing shells must have sufficient fatigue life ' and wear resistance to tolerate normal operating conditions.
This is requisite for service in emergency standby powar systems.
Obviously a larger connecting rod bearing has more area over which to distribute cylinder firing load, which results in a lower bearing pressure.
The bearing materials must exhibit low friction to tolerate possible momentary rubbing contact with the crankshaft during starting of the engine, before the sliding oil film is fully developed.
The bearing naterial should be resistant to possible corrosion caused by the chemical composition of the lubricating oil.
The running surface of the bearing shell should be constructed of a material which is tolerant to the presence of foreign particles, thereby minimizing journal wear.
The bearing design must be such that during operation key parameters -- including oil supply pressure, peak oil film pressure, minimum oil film thickness, and oil film temperature rise
-- are within acceptable limits for specified diesel engine application and for required life.
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f-While disassembling the three TDI diesel engines at the Shoreham Nuclear Power Station (SNPS) to replace the original crankshafts with those of modified design using 12-inch bea rings, four out of the 48 connecting rod 11-inch bearing shells were observed to be damaged.
All of the damaged connectir.g rod 11-inch bearing shells were in the upper position of the pair of connecting rod bearing shells in each connecting rod.
The upper position is more highly loaded, being subject to combustion forces.
The most severe damage was suf fered by the No. 5 upper bearing (associated with the No. 5 cylinder) of diesel generator set DG103.
This bearing is shown in Figure 2.
This half shell was fractured into two separate pieces near one end of the bearing.
Three other upper 11-inch bearing shells contained cracks through their thicknesses.
These cracks were in the same relative location as the fracture surface on the broken bearing but had not intercepted the bearing end to cause complete fracture.
The cracked bearing shells were in the No. 3 and No. A positions in DG103, and in the No. 4 position in DG102.
As pa r' the design review process for the connecting rod bearing shells, an investigation was undertaken to determine the cause of the cracking of these four upper connecting rod 11-inch bearing shells.
This cracking had occurred after approximately 600 to 800 hours0.00926 days <br />0.222 hours <br />0.00132 weeks <br />3.044e-4 months <br /> of operation, including about 250 hours0.00289 days <br />0.0694 hours <br />4.133598e-4 weeks <br />9.5125e-5 months <br /> of engine operation at or above 100". load.
In diesel engines of this size, normal connecting rod bearing life would be expected to exceed 20,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />. [1-2]
As a result of the replacement of the crankshafts, connecting rods, and associated connecting rod 11-inch bearing shells in the SNpS TDI DSR-48 diesels, changes occurred in the geometry and the size of the connecting rods and connecting rod bearing shells. The effects of these changes are to extend the expected life of the bearings, as is shown in this design review.
The approach taken is to characterize the bearing material in terms of its mechanical properties and to use the techniques of journal orbit analysis, finite element stress analysis, and fracture mechanics to determine the performance of this material as applied to connecting rod bearing shells in TDI diesel engines.
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Aluminum Company of Ame ri ca, Alcoa Aluminum Design Data, Pittsburgh, Pennsylvania, 1977.
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R.
Ewing (Manager of Engineering, Heavy Bearings, Imperial Clevite Inc., Engine Parts Division), private communication with L. A. Swanger (FaAA), November 2, 1983.
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Connecting rod bearing shell design and nomenclature (schematic).
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Broken connecting rod bearing, DG103, No. 5.
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2.0 EXAMINATION OF BEARING SHELLS Visual examination of connecting rod bearing shells from the original 11-inch diameter configuration of the SNpS DSR-48 TDI diesel engines showed that, except for the four baaring shells containing fatigue cracks, the remaining bearings were generally in serviceable condition.
The contact patterns in the babbitt overlay revealed significant edge loading of some of the bearing shells.
Contact patterns on the back of the bearing shells revealed that the ends were not supported by the bores of the connecting rods.
Replacement SNPS 12-inch bearing shells installed with the 12-inch crankpin crankshaft, showed similar edge loading in the babbitt contact patterns.
In addition, the No. 2 upper connecting rod hea ring from DG102 showed a baboitt removal pattern which was found to be due to reduced adhesion of the babbitt to the aluminum substrate.
Analysis, presented in later sections, demonstrates neither condition is expected to adversely impact the expected life of the bearing.
Visual inspection of 13-inch connecting rod bearing shells from the Grand Gulf Nuclear Station DSRV-16-4 engines showed some edge leading effects on the bearings and some areas of overlay cavitation, neither of which is a significant factor in the predicted bearing life.
Scanning electron microscopy of the fracture surface of c e of the cracked SNPS 11-inch connecting rod bearing shells showed that the fracture probably originated at surface pores approximately 0.020 inch to 0.030 inch in diameter.
An acceptance criteria is presented in Section 5.0 to detect unacceptably large voids in any new bearing shells.
The tensile properties of specimens taken from the cracked SNPS 11-inch bearings shells showed that the ultimate tensile strength of the material met current TDI [2-1] specifications.
However, since only subsized tensile specimens could be obtained from the actual bearing shells, it is difficult to determine whether or not ductility specifications were met with the material from the cracked bearings.
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2.1 Nondestructive Examination CanecM ng rod bea ring shells from three di f ferent sources were examined visually. The first group of shells were the original connecting rod bearing shells from the SNPS DSR 48 diesels with 11-inch trankpin crankshafts.
The second group were bearing shells from SNPS DSR 48 diesels with the replacement 12-inch crankpin crankshafts. The last source of bearing shells was Grand Gulf Nuclear Station operated by Mississiopi Power & Light; four pairs of connecting rod bearing shells from the DSRV-16 4 engine with a 13-inch crankpin crankshaft were examined.
The original 11-inch SNPS bearings, with the exception of the cracked shells, appured to be in serviceable condition, showing the expected polishing of the babbitt overlay.
The polishing had occurred in the most highly loaded areas of the bearing.
The amount of scoring of the bearing surface resulting from circulating solid particles in the lubricant was minimal, indicating that the engines were internally clean.
There was no evidence of any chemical attack of the babbitt overlay, indicating that the lubricating oil had remained nonacidic and was uncontaminated by acidic combustion products or by coolant leaks into the oil system.
The majority of these bearings showed a polishing pattern in the babbitt that was wider at the ends of the bearings, covering almost 90* of arc, than it was in the middle where it covered less than 45' of arc.
Also the intensity of the wear was higher at the Edges of the bearing than in the center cf the bearing. This pattern is indicative of edge loading which results when the journal axis is not perfectly parallel with the bearing surface.
This causes the journal to approach the bearing more closely at the bearing ends, increasing the proportion of the firing pressure carried on the bearing ends. This asymmetry is considered and assumed in the life prediction of the new 12-inch bearings, and acceptable bearing life is found even in its presence.
Visual examination of the backs of the original 11-inch SNPS bearings showed that the ends of the bearing were not supported by the bores of the -
connecting rods. This was a consequence of the large 1/4-inch chamfers on the 2-2
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connecting rod bores. Figure 3a shows a cross-sectional representation of the contact between the connecting rod and the connecting rod bearing and indicates the unsupported ends of the bearings with the original 11-inch cr:1kpin crankshaft.
This large chamfe has been reduced to 1/16 inch in the new 12-inch connectir.g rod 'esign.
The three cracked 11-inch bearings which had not completely fractured had cracks approximately four inches in length near one end of the be a ri ng.
A crack was apparent on both the I.D. and the 0.D. of the bearing shell; these two indications appeared to coincide and thus to represent one through-crack in the bearing shell.
Dye penetrant testing of these bearing shells indicated that these visual features were cracks.
Radiography of one of the 11-inch bearing shells containing this indication, the No. 4 upper connecting rod bearing shell from DG102, also indicated the presence of a discontinuity or crack in the bearing material.
The second category of bearings to be visually inspected consisted of the replacement connecting rod bearing shells which were installed at SNPS with the new 12-inch diameter crankpin crankshafts.
After 100 hours0.00116 days <br />0.0278 hours <br />1.653439e-4 weeks <br />3.805e-5 months <br /> of testing at full load, DG102 was partially disassembled for inspection.
At that time several of the connecting rod 12-inch bearing shells were removed for visual inspection, dye penetrant inspection, and measurement of wall thickness.
The contact patterns on the I.n.
of the bearing were evaluated to determine load distribution across the length of the bearing. A number of the bearing shells showed clear indications of edge loading in the polishing pattern on the babbitt, but not to a degree that would indicate impaired bearing performance during the life of the diesel generator uisit.
In addition, the No. 2 upper connecting rod bearing from OG102 showed a pattern of babbitt removal at one end.
Examination of this region by optical microscopy at 40X magnification showed that the babbitt removal was occurring e
in very localizd regions and that the babbitt which remained on the surface,
between the localized regions of removal had no sign of babbitt fatigue or 2-3 1
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cracking.
In addition, at the bottom of the pits left by babbitt remova*., the machining marks on the I.D.
of the aluminum bearing material were clearly visible.
On the remainder of the babbitt surface of this bearing shell there is a pattern c,f very small blisters in the babbitt.
The mode of removal was celamination in areas of weak adhesion of the babbitt overlay to the underlying aluminum substrate.
This condition is not significant as far as the performance of the connecting rod bearings in the engine is concerqed, but is primarily a cosmetic surface condition.
Over the normal wear life of the connecting rod bearings, the babbitt layer, which is at most 0.002-inch thick, will be worn completely away in the highly loaded regions.
Also, up to 0.003 inch of the underlying aluminum material can be worn away, for a total reduction in thickness of 0.005 inch [2-2].
The contact pattern on the back of this bearing shows that, with the change to the new connecting rods with a small 1/16-inch chamfer (see Figure 3b), full support of the bearing back has been achieved.
The third category of bearings to be visually inspected was from the Grand Gulf Nuclear Station DSkV-16-4 diesel engines.
These bearings are reported by Grand Gulf to have experienced about 1200 hours0.0139 days <br />0.333 hours <br />0.00198 weeks <br />4.566e-4 months <br /> of total engine operating time, of which approximately 315 hours0.00365 days <br />0.0875 hours <br />5.208333e-4 weeks <br />1.198575e-4 months <br /> was at or above 100". load.
The bearings in the upper position showed normal babbitt contact patterns in the most highly loaded regions.
In some cases there was evidence of edge loading of the
Dearings but,
in those examples examined, even less than for the 11-inch or 12-inch SNps bearings.
The bearings from Grand Gulf did show light to moderate scoring of the overlay.
In addition, there were a few isolated areas of overlay removal. These areas were, however, not in the most highly loaded region of the bearing and probably represent areas of cavitation.
This apparent cavitation is confined to the babbitt overlay and shows nn evidence of progressing into the underlying aluminum substrate.
It had no effect on the function of the bearings.
The contact pattern on the back of the DSRV-16-4 connecting rod Dearing shells from Grand Gulf Nuclear Station shows that the connecting rod bore is,
providing essentially full support of the bearing shell 0.0..
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2.2 Destructive Examination Destructive examination was confined to the original 11-inch diameter bearing shells from SNPS. Two of the bearing shells which had cracked but not separated were destructively analyzed to expose the fracture surfaces for detailed examination.
At the FaAA Palo Alto laboratory, the No. 4 upper connecting rod bearing shell from DG103 (containing a crack approximately four inches long) was subiected to destructive examination.
Two axial-radial cuts through the fracture surface were made from the end of the bearing containing the crack.
This freed the major portion of the fracture surface for seqaration and examination.
The fracture extended from the I.D of the bearing completely through to the 0.0..
The No. 3 upper connecting rod bearing shell from DG103 was initially examined by TOI in Oakland.
Suf ficient force was applied to the cracked bearing to complete fracture to the bearing edge, freeing the fracture surface for examination.
Again the crack was shown to be a through-crack from the
- I.D. of the bearing shell to the 0.0..
In addition, the shape of the crack fronts at both ends of the crack showed that the I.D. edge of the crack was leading the 0.0. edge of the crack, suggesting that the crack had initiated at the 1.0. of the bearing shell.
2.3 Electron Microscopy A portion of the fracture surface which was removed from the No. 4 upper connecting rod 11-inch be--ing from DG103 was examined by scanning electron microscopy. This examinatior revealed significant near-surface pores which are the probable initiation sites for the cracking.
These pores are approximately 0.020 inch to 0.030 inch in diameter.
Examples of these pores on the fracture surface are shown in Figure 4 The brittle character of the R850 aluminum alloy prevents it from yielding very much information about the nature of the cracking process.,
However, the orientation of the crack relative to the pores that were 2-5 4
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discovered is consistent with those pores being the initiation sites for the fatigue crack.
The internal surface of the pores, being rounded and showing signs of dendritic structures, indicates that their most likely origin is from shrinkage associated with solidification of the castings from whi ch the bearing shells are made.
This shrinkage may also be assisted by dissolution of gases, such as hydrogen, from the liquid aluminum during solidification.
As such, the pores would be a normal effect of the manufacturing process by which the castings were made.
2.4 Chemical Composition To confirm that the 11-inch connecting rod bearings are made from the specified material, Alcoa B850, a sample of aluminum material from the No. 4 upper connectica end bearing from DG103 was submitted to Metallurgical Testing Corporr;ien fe emical analysis.
The results of the chemical analysis, as well as the specificetion for alloy BS50 [2-3], are given in Table 1.
The results indicate that the specified alloy was used in the manufacture of these bearings.
2.3 Tensile Properties Mechanical properties samples were cut from the end of the subject 11-inch bearings containing the cracks, between each parting line and the fracture surface.
The specimens were 1/4-inch gage di amete r, 1-inch gage length per ASTM B-557-81 [2 4], the largest that could be obtained from the finished bearing, and they were oriented parallel to the axis of the bearing perpendicular to the plane of the fracture.
Ten specimens were prepared and tested according to ASTM Standards.
Eight of the specimens were from the No. 4 upper connecting rod besring shell of DG103 and two of the mechanical properties test specimens were from the
~
No. 3 upper connecting rod bearing shell of DG103.
The results are listed in Table 2.
Ultimate tensile strength ranged from 23.7 ksi to 28.1 ksi with elongations ranging from 0.40% to 0.88%.
Only one of the ten test specimens met the apparent original design requirement [2-1] for tensile strength and -
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none met the elongation requirement.
When compared with TDI's current specification requirements [2-1], all ten samples met the tensile strength criterion, but again, none met the elongation requirement.
the largest that could be taken from finished The w v
bearings, but werc.:1e-nal f the size of samples that would be taken from unfinished castings fo. quality assurance. ASTM Standard B-557-81 states that elongation values obtained from smaller specimens may be less than those obtained from larger specimens.
The ultimate tensile strength results indicate that the bulk cast aluminum tearing material is suitable for its intended application.
The re;,rted ductility values are not significant, since they were measured on sub-size specimert.
If full-size specimens could tave been used, it is expected that the ductility would have been satisfactory.
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Section 2 References 2-1.
C.
Matthews and G.
King (Transamerica Delaval Inc.,
Engine and Compressor Division), private communication with L. A. Swanger (FaAA),
October 4, 1983.
2-2.
TRANSAMERICA DELAVAL INSTRUCTION MANUAL, Model DSR-48 Diesel Engine, Serial Nos.
74010 - 2504, 74011 - 2605, 74012 - 2606, Transamerica Delaval Inc., Engine and Compressor Division.
2-3.
Aluminum Company of America, Alcoa Aluminum nesign Data, Pittsburgh, Pennsylvania, 1977.
24 ASTM Standard B-557, " Tension Testing Wrought and Cast Aluminum and Magnesium Alloy Products," ASTM,1981.
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TABLE 1 CHEMISTRY OF DG103, No. 4 UPPER CONNECTING ROD BEARING Results B850 Nominal Composition of Analyses
( ".)
(%)
A1 90.0 balance Sn 6.0 5.26 Cu 1.0 1.86 Ni 2.0 1.38 Mg 1.0
.77 Fe
.36 Si
.25
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Ti
.12 O
e e
S O
2-9
=,=,=7+..
- - -~g.----
TABLE 2 TENSION TEST RESULTS FOR DG103 CONNECTING R00 BEARING SHELLS Test No.
Position U.T.S.
Elongation (Kst)
(percent) 1 No. 3 Upper 25.7 0.80 2
No. 3 Upper 23.7 0.40 3
No. 4 Upper 25.2 0.70 4
No. 4 Upper 25.7 0.76 5
No. 4 Upper 26.5 0.76 6
No. 4 Upper 26.1 0.56 7
No. 4 Upper 26.7 0.72 8
No. 4 Upper 26.9 0.54 9
No. 4 Upper 28.1 0.88 10 No. 4 Upper 26.1 0.68 Specification (1976) [2-1]
27.0 2.00 Specification (1983) [2-1]
23.0 2.00 Note: Results are from 1/4-inch diameter test specimens.
Specifications are for 1/2-inch diameter test specimens.
The smaller test specimens result in lower elongation results, but the tensile strength results are unaffected by this difference in size.
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Connecting rod and bearing configuration original 11-inch journals
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Scanning electron microscope fractographs of DG103, No. 4.
Voids are approximately 0.020-inch to 0.030-inch.
FaAA-84-3-1 2-12
3.0 JOURNAL ORBIT ANALYSIS The pressure distribution in tne hydrodynamic oil film acting on the connecting rod bearing shells was determined by journal orbit analysis [3-1, 3-2].
In their original configuration, the DSR-48 engines at SNps developed a peak oil film pressure of 29,700 psi.
Replacement of the original 11-inch diameter journal crankshaft, connecting rods, and connecting rod bearing shells with the 12-inch model resulted in a reduction of the peak oil film pressure to 26,800 psi.
For the DSRV-16-4 engine, the peak oil film pressure in the connecting rod bearings is 25,800 psi.
~
An industry guideline [3-3] for the bearing material used by TDI in its diesel engines recommends a maximum oil film pressure of 26,000 psi.
These industry guidelines are not absolute maximum allowable values.
Some engine manufacturers successfully operate engine sleeve bearings above indust ry
, guidelines in specific applications.
In addition, the finite element method and fracture mechanics analysis of the connecting rod bearings, performed by FaAA, is a more detailed analysis than is performed by engine builders and bearing suppliers in ti.2,;ourse of normal applications engineering.
This detailed analysis provides the basis for the calculated bearing fatigue life.
3.1 DSR-48, 11-inch Crankpin Imperial Clevite Inc.,
a major independent manufacturer of sleeve bearings, was engaged to compute the loading on the connecting rod bearings.
Journal orbit analysis [3-4] was used to determine the characteristics of the hydrodynamic oil film formed between the bearing and the crankpin.
Examples of those characteristics that were determined are the thickness and the pressure profile in the oil film throughout the 720* of the engine operating cycle.
Journal orbit analysis is a technique for solving the differential equations describing the formation of an oil film, taking into account the dynamic character of the loading of a bearing in a reciprocating internal combustion engine.
Data supplied to Imperial Clevite for the computation included relevant engine and bearing dimensions and design features obtained 3-1 4
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..--....-..w..
from the direct measurement of some of the engine components.
Data also included operating parameters, such as engine speed; power output; lube oil temperature, pressure, and viscosity; peak cylinder pressure; and mechanical efficiency.
The reciprocating and rotating masses required for the analysis were obtained by direct weighing of components from the SNPS engine, including determination of the center of gravity of the connecting rod so that the appropriate allocation of weight could be made to the rotating and reciprocating components.
Input data used in the analysis is shown in
~
Table 3.
A summary of the results of the journal orbit analysis for the SNPS engines is also shown in Table 3.
For the original engine configuration with the 11-inch crankpins, the maximum oil film pressure acting on the bearing was found to be 29,700 pounds per square inch [3-4].
3.2 OSR-48,12-inch Crankpin Journal orbit analysis was also performed for the current configuration of the SNPS diesels with 12-inch diameter crankpins.
The only change necessary in the input data was the change in shaft diameter.
This has the effect of distributing the applied load over a greater bearing area and reducing the pressure in the hydrodynamic oil film, and hence reducing the stress acting on the connecting rod bearing shell.
The exact reduction in pressure was determined by rerunning the journal orbit analysis for the 12-inch diameter crankpin.
The result shown in Table 3 is that the maximum
^
oil film pressure has been reduced to 26,800 psi [3 4].
3.3 DSRV-16-4, 13-inch Crankpin Using Imperial Clevite's proprietary code, journal orbit analysis was performed by FaAA on the connecting rod bearing shells for the DSRV-16-4 engine as represented by those installed at the Grand Gulf Nuclear Station of Mississippi power & Light.
Input data for this journal orbit analysis is -
given in Table A.
Also shown in Table 4 is the result of the journal orbit 3-2 4
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analysis. The peak oil film pressure in the OSRV-16-4 is 25,800 psi.
3.4 Interpretation of Journal Orbit Analysis The peak oil film pressure in the OSRV engine is lower than that in the DSR engine primarily because the greater cass of the connecting rod assembly and the two pistons adds an inertial component to the load which opposes the firing pressure of one of the cylinders and reduces the peak in the oil film pressure.
The firing order in the DSRV-se.-ies engines is such that only one piston on each crankpin fires during one revolution of the crankshaft.
This prevents firing pressure from two cylinders firing only 45' apart from being added together.
Thus, the bearing is loaded by the firing of only one cylinder at a time.
The larger journal diameter, 13 inches for the DSRV versus 12 inches for the currently configured DSR, also contributes to a reduction in the peak oil film pressure.
The net result of the differences in inertial forces and in crankpin diameter accounts for the difference in the peak oil film pressures computed for the inline and the DSRV engines.
Since both the OSR-4P and DSRV-16 4 peak oil film pressures were quite close to the industry guideline of 26,000 psi for the peak oil film pressure, both engines were analyzed in more detail by the finite element and fracture mechanics techniques described in the following sections.
3-3 i
---a
Section 3 References 3-1.
Ross, J. M. and R.R. Slaymaker, " Journal Center Orbits in Piston Engine Beari ngs,"
SAE Paper
- 690114, Society of Automotive Engineers, Warrendale. Pennsylvania,1969.
3-2.
Hollander, M.
and K.
A.
Bryda, " Interpretation of Engine Bearing Performance by Journal Orbit Analysis," SAE Paper 830062, Society of Automotive Engineers, Warrendale, Pennsylvania,1983.
3-3.
W.
A.
Yahraus (Mana,ar of Product Analysis, Imperial Clevite Inc.,
Engine Parts Division), private communication with L.
A.
Swange-(FaAA), October 4, 1983..
34 Journal orbit analyses of TDI Enterprise R 48 Diesel Engine performed by Imperial Clevite Inc. for FaAA, October 6,1983.
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TABLE 3 JOURNAL ORBIT ANALYSIS OF CONNECTING R00 BEARINGS TDI DSR-48 Diesel Engines Shoreham Nuclear Power Station Input Data - TDI Enterprise Diesel Engine Brake Horsepower 4,880 b.h.p.
Cylinders 8
Bore 17.000 inches Stroke 21.000 inches Compression Ratio 11.57:1 Connecting Rod C/L-C/L length 46.125 inches Reciprocating Weight 799.4 pounds Rotating Weight 432.3 pounds Shaft Diameter 11.000 inches /12.000 inches Radial Clearance
.0045 inch Effective Length 3.1885 (x2) inches Grooving 360' 011 Pressure 55 psig Oil Temperature 165'F Peak Pressure 1,680 psig Mechanical Efficiency 885 Operating Cycle 4-stroke Output inch Journal Maximum 011 Film Pressure 29,745 psi At Bearing Angle 2 degrees Output inch Journal Maximum 011 Film Pressure 26,780 psi At Bearing Angle 2 degrees Imperial Clevite Inc. Recommended Bearing Loads, Solid (Wrought) Al-6% Sn [3-3]
Maximum 011 Film Pressure, Stationary Diesel Engines, Intermittant Service 26,000 psi
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TABLE 4 JOURNAL ORBIT ANALYSIS OF CONNECTING ROD BEARINGS TDI DSRV-16-4 Diesel Engines Grand Gulf Nuclear Power Station Input Data - TDI Enterprise Diesel Engine Brake Horsepower 9,750 b.h.p.
Cylinders 16 Bore 17.000 inches Stroke 21.000 inches Compression Ratio 12.0:1 Connection Rod C/L-C/L Length 46.125 inches Reciprocating Weight 800 pounds / cylinder Shaft Diameter 13.000 inches Radial Clearance
.0070 inch Effective Length 3.0325 (x2) inches Grooving 360' Oil Pressure 52.5 psig Oil Temperature 158'F Peak Pressure 1, A50 psig Mechanical Efficiency 88%
Operating Cycle 4-stroke Output inch Journal Maximum 011 Film Pressure 25,815 psi At ' Bearing Angle 5 degrees Imperial Clevite Inc. Reconmended Bearing Loads, Solid (Wrought) Al-6% Sn [3-3]
Maximum 011 Film Pressure, Stationary Diesel Engines, Intermittant Service 26,000 psi 4
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4.0 FINITE ELEMENT STRESS ANALYSIS Finite element stress analysis was used to compare the magnitude, orientation, and location of the largest tensile stress in the connecting rod bearing shells for the three bearing geometries analyzed: DSR-48 with 11-inch trankpins, DSR-48 with 12-inch crankpins, and DSRV-16-4 with 13-inch crankpins. The purpose of this calculation was to obtain, the ratio of maximum tensile stress for use in a subsequent fracture mechanics analysis.
In an earlier report, FaAA-83-12-9 [4-1], a simplied model for computing this stress ratio was used.
For the present design review, a more sophisticated analytical model was used, which also demonstrated that the stress ratio had been accurately calculated.
The maximum tensile stress in the DSR 48 11-inch crankpin case was located at approximately the site of crack initiation, oriented perpendicular
~
to the crack surf ace, and with a magnitude of 17,700 psi. The DSR 48 with 12-inch crankpins had the maximum tensile stress in the same location, and in the same direction, but with a magnitude of only 8,800 psi.
The DSRV-16-4 connecting rod bearings are subject to an even lower stress: 8,200 psi.
Although the absolute values of stress found in this report differ somewhat from those presented in FaAA-83-12-9, and are more accurate, the reduction in stress is predicted to be by the same factor, and the prediction of the amount of increased bearing life is therefore identical.The ratio of tensile stress in the replacement 12-inch bearing shell (with improved support from the connecting rod) to the tensile stress in the original 11-inch bearing shell (with unsupported ends) was 0.497, compared with the ratio 0.495 computed with the simplified model.
4.1 DSR-48, 11-inch Crankpin Finite element stress analysis of the connecting rod bearing shell was performed using the ANSYS code.
In order to properly accommodate the input conditions, the geometry, and the expected stress distribution pattern, a three-dimensional finite element model was required.
The three-dimensional analysis was simplified by taking advantage of the essentially radial symmetry -
of this configuration.
The connecting rod bearing shell was modelled as an 4-1 s
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axisymmetric ring of aluminum with elastic properties taken from the Alcoa design data manual.
The connecting rod was not modelled completely, but was represented by a ring of steel outside of the connecting rod bearing shell, sufficiently thick such that the boundary conditions on the exterior surface of this steel ring had a negligible effect on the behaviour of the aluminum connecting rod bearing shell contained within it.
A mathematical boundary, concentric with the geometric center of the bearing shell, was constructed within the forged steel connecting rod, and the material outside of the mathematical boundary was neglected, since it would have a negligible effect on the bearing.
The steel ring, being over three times the thickness of the aluminum bearing shell, and being made from steel with three times the elastic modulus of the bearing shell, was significantly stiffer than the aluminum itself such that the deformation of the aluminum shell was insensitive to the behaviour of the material outside of the mathematical boundary.
This assumption simplified the geometry of the problem considerably and allowed a more efficient means of solution without impacting the required accuracy of the stress analysis.
The interference fit between the connecting rod bearing shell and the steel connecting rod was modelled by imposing an external radial displacement on the mathematical boundary such that the circumferential stress in the aluminum connecting bearing shell was apJroximately at the yield stress of the aluminum material, about 20,000 pounds per square inch [4-2].
It is known from experience that, when assembled into connecting rods, the connecting rod bea ring shells do reach the yield stress because they are permanently reduced in diameter.
The outer boundary of the steel ring was then fixed at the computed displaced condition, as one of the boundary conditions operating on the system.
The inner boundary condition was expressed in terms of the oil film pressure acting on the I.n.
of the bearing.
This oil film pressure varied both axially and circumferentially within the bearing.
The theoretical axial distribution of pressure, as determined by journal orbit analysis, would have been a parabolic distribution with the maximum pressure at the center of each of the bearing lands on either side of the oil groove. However, visual examination of connecting rod bearings which had run both in DSR 48 and DSRV-16-4 engines showed that the babbitt contact 4-2
patterns, and hence the load on the bearings, was concentrated at the ends of the bearings.
Allowance was made for this by skewing the pressure distribution toward the enos of the bearings.
In order to match observations made on the bearings, a skewed load was deieloped by numerical integration which placed 80% of the applied load on the outer 30% of the bearing length.
This approach approxirrates the worst condition of edge loading seen on any of the bearings inspected.
Many of the bearings did not have this degree of edge loading and would not develop the peak stresses that were found by the
'inite element method stress analysis.
~
The oil film pressure profile also varies circumferentially with the nximum pressure being near the axis of the connecting rod at what is referred to as a bearing angle of 0',
which is in the direction of the piston from the center of rotation at the large end the connecting rod.
A model was developed in which the actual circumferential variation of pressure was expressed as a
Fourier series of sinu;oidal pressure distributions around the circumference of the bea rings.
Each term in the Fourier series was solved separately by the ANSYS routine, and then, by the principle of superposition applicable to problems in linear elasticity, the summation of these solutions was obtained.
This summation represents the solution to the problem with the original circumferential variation which was represented by the Fourier series.
Figure 5 shows the mesh that was used for finite element analysis of the connecting rod bearings used in the original SNPS engines.
Figure 6 shows the most important result of the finite element analysi s, an isostress plot of the axial stresses acting on the bearing shell.
The highest axial stress,17,700 psi, occurs at the 1.3.
at a node approximately 0.88 inches from the end of the bearing.
This tt the mxin.um tensile stress in the connecting rod bearing shell.
It occurs closE to the origin of the fatigue cracks which were found in several of these bearings and is oriented perpendicular to the fracture surface.
The importance of the stress analysis of the original connecting rod bearings used with the 11-inch shaft, and with the unsupported bearing ends.
from the original connecting rods,. is in its use for computing a ratio of t
4-3 i
- m. _ _ m m _
tensile stresses with the currently installed connecting rod bearings used with the 12-inch diameter shaft and with r.ew connecting rods which completely support the ends of the bearings.
The finite element stress analysis of the current configuration is reported in the next section.
4.2 DSR-48, 12-inch Crankpin Analysis similar to that described for the 11-inch diameter connecting rod bearing shells was performed for the connecting rod bearing shells installed with the 12-inch crankpin.
The model was similar, but was modified in three respects.
The model for the connecting rod reflects the smaller chamfer, which allows the connecting rod bore to completely support the bearing shell.
This is shown in Figure 7, which depicts the finite element mesh for the 12-inch model, By comparison with Figure 5, the change in chamfer on the connecting rod can be seen.
The model I.D. was increased from 11 inches to 12 inches. Finally the boundary conditions were modified. First the external displacement was recalculated for this model to prestress the bearing shell to 20,000 ps1 in the circumferential direction.
Also the boundary conditions on the I.D.
of the bearing shell were those from the journal orbit analysis of the new configuration.
The significant results of this stress analysis are shown in Figure 8 which displays an isostress map of the axial stresses in the connecting rod bearing shell in the current SNPS configuration.
The maximum tensile stress still occurs at the same node on the I.D. of the bearing, but the magnitude of the maximum tensile stress normal to the potential fracture surface has been decreased to R,800 psi, or to a value 0.497 of the maximum stress in the original 11-inch connecting rod bearing shells.
The maximum tensile stress in the connecting rod bearing shells currently installed in the TDI diesel engines at Shoreham is only one-half of that which was acting on the original bearings that cracked after about 250 hours0.00289 days <br />0.0694 hours <br />4.133598e-4 weeks <br />9.5125e-5 months <br /> of full load operation.
The reduction in this tensile stress results from a reduction in the calculated peak oil film pressure, which is a direct consequence of a larger journal diameter, and also it is due to the ~
elimination of the unsupported bearing ends via the reduction of the chamfer 4-4 i
)
in the bore of the iw connecting rods.
The complete st;pport of the bearing backt elim' nates the cantilevered loading which led to the high stresses in the original bearingj.
The r4tio of tensile stress in the current 12-inch
'mring shell, relativ't to the tensi'.e stress in the original 11-inch bearing shell, 0.497 is used in Sectior, 5,0 to predict the improveme it in fatigue life for thre current 12-inch bearing shells.
4.3 C.V-Ift-4,13-inch Crankpin The stress distribution in the connecting Pod cearing shells for the
~
DSRV-16-4 engine was determined by finite element analysis corresponding to that used for the DSR-48 engine.
Similar models and Issumptions were used as discussed in the previcus sections.
In Figure 9 thr.' mesh use<J for tne 13-inch c.rinkpin mocel is shown, and in Figure 10 the isoster:ss plots for the axial strtss in the connecting rod bearing shell are shown. The numerical result is
' that the maximum tensile stress in the bearing shell occurs et the 1.n. and h35 o a value of 8,200 psi.
Since this maximum tensile stiass for the DSRV engine is slightly lower than the maximun tensi~t e stress in the current
- onfiguration of the DSR 48 engine, the resistance of the bearings to fatigue cracking will be similar in both cases.
4-5 7_
Section 4 References 4-1.
Failure Analysis Associates, " Analysis of the Replacement Connecting Rod Bearings for Emergency Diesel Generators, f atigue Life Prediction, Shoreham Nuclear Power Station," FaAA-83-12-9, Palo Alto, California, December 15, 1983.
4-2.
Aluminum Company of America, Alcoa A':uminum Design Data, Pittsburgh, Pennsylvania, 1977.
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i
- i. 7.1:
4
84/5/5 17.650 E!
&s 76 67 ss 43 40 31 22 16 7
Y L2 Zvst
- 5 75 66 57 as 59 50 ti il 6
ANGLs90 ENUW=1 84 74 65 56 47 58 2s 20 10 2
85 75 64 55 46 57 ts is s
3 l 27 l la s
3 81 ft 65 54 45 56 8
/
7 62 55 44 55 26 17 it 7s 70 61 52 at 52 25 is 4
[as ss 7s 6s 60 58 41 55 24 14 3
EPLT ANSYS 1
at 77 ss
!s 50 45 34 25 15 V-16 CQN ROD SEARING ANA;., CRANK = 422., SKEWED LOA 05 Figure 9.
13-inch trankpin, DSRV-16-4 FEM Model FaAA-84 1 4-11
84/5/3 ANGLE =
0 17.756 El 4000 d'
. L.-
3 t.
zv.i ANGL=90 As-16000 9 -12000 C=-8000 D=-4000 Eso F=4000 J-
}
E
- L r
_ r_.
L-y y
-I r
s r
,N
.u_
V/
7 T
L.
.L 3
L SY ANSYS 5
r DA Y v ~e
- A 1
V-16 CON ROD EEARING ANAL. CKANK s 422. SKEVED LOADS Figure 10.
c,13-inch Crankpin, DSRV-16-4 y
Longitudinal Tensile Stress FaAA-84 1 4-12
5.0 FATIGUE ANALYSIS Using an equation presented in Fuchs and Stephens, Metal Fatigue in Engineering [5-1], the ratio of the maximum tensile stre.sses computed by the finite element analysis was used to compute a conservative estimate of the ratio of the fatigue lives for two stress states.
By this calculation, the DSR-48 diesel engine with completely supported 12-inch I.D. bearings was shown to resist fatigue cracking for at least 152 times the fatigue life of the original 11-inch SNPS configuration, as was also calculated in FaAA-83-12-9.
The DSRV-16-4 diesel should also have a similarly long fatigue life, of 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> at full load.
Fracture mechanics calculations for the original and current configura-tions of the SNPS DSR-48 diesel engines confirmed the improved fatigue resistance.
The stress intensity factor range, AK, for the original 11-inch DSR-48 bearing shells exceeded the approximate threshold value for fatigue crack growth, while AK for the current 12-inch DSR-48 connecting rod bearing shells is significantly below the threshold value for the same casting quality (void size).
The stress intensity factor range was used to compute an acceptance criterion for bearing shell castings in the current DSR 48 and DSRV-16-4 engines, a maximum discontinuity size of 0.050 inch.
The bearing shells installed at Shoreham have been nondestructively inspected by radiography to assure that only bearing shells satisfying the acceptance criterion are installed in the upper connecting rod bearing position.
5.1 Fatigue Life Calculation In the elastic strain, high-cycle-f ati gue region (number of cycles 6
greater than 10 ), the behavior of aluminum can be described by the equation
[5-1]:
af (2N)D c
=
a stress amplitude where c
=
a of fatigue strength coefficient
=
5-1
=. _,_y:
number of cycles N
=
fatigue strength exponent b
=
The stress amplitude for the connecting rod bearings is one-half of the maximum tensile stress computed by the finite element stress analysis described in Section 4.0.
In using the ratio of the stress amplitudes to compute the ratio of the number of cycles to failure, the coefficient af drops out of the expression.
The value for b is in the range of -0.06 to -0.14
-0.14, which yields the The most conservative computation is to use b
=
smallest change in N for a given change in c
- a f (2N ) 0.14 1
al 0.497
=
=
f (2N ) gg4
,2 a
o thus, y
[2 152
=
New bearing (12-inch) where subscript 1
=
Old bearing (11-inch) and subscript 2
=
This calculation predicts that the new bearings should not fail by fatigue until they have experienced 152 times the number of cycles that failed the original bearings.
The connecting rod bearings are subjected to one stress cycle 'a every two rotations of the crankshaft, or 225 cycles per minute. Four original 11-inch bearing shells were found to be cracked after approximately 250 hours0.00289 days <br />0.0694 hours <br />4.133598e-4 weeks <br />9.5125e-5 months <br />, or 3.4 x 106 cycles at full load.
The new bearing shells would not be expected 8
to exhibit cracking until after 5 X 10 cycles, or 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> of full-load operation have occurred.
The 38,000 hour0 days <br />0 hours <br />0 weeks <br />0 months <br /> life at full load also is greater than ten times the number of hours that the SNPS diesel generators will accunulate over a 40-year service life.
5-2
4 5.2 Threshold Stress Intensity Range An alternative analysis for determining whether or n( \\ fatigue failure would occur in the replacement connecting rod bearings was performed.
This analysis involved the computation of the stress intensity factor range AK based on the stresses applied to the bearing and on the assumed void sizes present.
The stress intensity factor range was computed by using the BIGIF fracture mechanics code [5-3].
Computing as on the basis of the results of the finite element stress analysis for the 11-inch bearing shells originally installed in the SNPS DSR-48 diesel engines, AK is approximately 2.9 ksi vincnes.
For the new bearing shells currently installed in those engines, the reduction in the stress field causes the stress intensity factor range to be reduced to 1.44 ksi vintnes.
The threshold value of AK for growth of a pre-existing void in fatigue, obtained by comparison to other aluminum alloys, is estimated to be approximately 2.0 ksi vincnes
[5-13 Therefore, since the AK value for the new bearing shells is below the threshold value for growth of pre-existing voids, the presence of 0.020-inch to 0.030-inch voids should not cause the initiation of fatigue cracking.
The results of the finite element stress analyses of the connecting rod bearing shells in the DSRV-16-4 engines reported in Section 4.3 showed that the maximum tensile stress in those bearing shells is slightly less than the maximum tensile stress in the currently configured DSR 48 engine connecting rod bearing shells.
Therefore, since the BIGIF fracture mechanics fatigue analysis is based on linear elasticity theory, the predicted fatigue life in the bearings in the DSRV engines would also be 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br />.
Also, the stress intensity factor range in the DSRV engines will be less than the stress intensity factor range for the current DSR-48 engine, or less than 1.44 ksi vincnes.
For both engines, the stress intensity factory range is below the threshold value of approximately 2.0 ksivintnes.
5-3 1
..-.-meme-,
-..m
t 5.3 Acceptance Criteria for Connecting Rod Bearings Compared to the original configuraticn installed in the DSR-48 engines, there is a factor of 152 increase in the expected fatigue life of the bearing shells, and a 50". decrease in the stress intensity factor range. Because of the significant reduction in the tensile stress in the connecting rod bearing shells of the current configuration, the size of voids which can be tolerated is larger than would have been acceptable in the original bearings. Using the BIGIF fracture mechanics code to calculate the stress intensity factor range that would remain below the approximate threshold value of IX,
2.0 ksi (inenes, yields an acceptable void size of 0.050 inch in the highly stressed areas of the upper connecting rod bearings.
The BIGIF analysis used to generate this acceptance criterion is conservative in that the voids actually seen in the bearing material are essentially spherical with rounded interior surfaces.
The BIGIF analysis assumes that the voids are sharp-edged and behave like sharp cracks from the onset of fatigue.
This is conservative in that no credit is taken for the increased stress required to initiate a sharp fatigue crack from a typical casting void.
The critical zone of the connect!ing rod bearings, to which the 0.050-inch maximum discontinuity acceptance criterion applies, was determined by the region of the connecting rod bearing shell in which the tensile stress exceeds one-hal f of the maximum tensile stress in the bea ri ng.
By examining the outputs from the ANSYS finite element models of the DSR-48 and DSRV-16-4 connecting rod bearing shells, it was determined that this critical zone encompasses a band on each end of the bearing beginning 0.4 inch from the bearing end, extending inward toward the oil groove to a point 1.4 inches from the bearing end. This critical zone is also centered circumferentially on the bearing shell, extending circumferentially on either side of the center 2.5 inches.
Outside of this critical zone, and in the lower connecting rod bearing shells which are much less highly loaded than the upper bearing shells, the acceptable voio size is a calculated 0.250 inch.
5-4
.~._:-_._.,-.=n.:=
=
Section 5 References 5-1.
H.O. Fuchs and R.I. Stephens, Metal Fatigue in Engineering, John Wiley
& Sons, New York, 1980.
5-2.
Failure Analysis Associates, " Analysis of the Replacement Connnecting Rod Bearings for Emergency Diesel Generators, Fatigue Life Prediction, Shoreham Nuclear Power Stati on," Fa AA-83-12-9, Palo Alto, California, December 15, 1983.
5-3.
Besuner, P.M., S. A. Rau, C.S. Davis, G.W. Rogers, J.L. Grover, and D.C.
Peters, "BIGIF:
Fracture Mechanics Code for Structures," (Manuals 1, 2, and 3) EPRI Technical Report NP-1830, Failure Analysis Associates Report, April 1981.
O e
O 5-5 1
6
= --
4-
6.0 DISCUSSION AND CONCLUSIONS The adequacy of the 12-inch and 13-inch connecting rod bearing shells in TOI DSR and DSRV-series diesel engines respectively has been evaluated by analysis of the loading conditions, the state of stress computed by finite element analysis, and the fatigue life computed by fracture mechanics analysis of the bearing shells.
With the configuration change associated with the replacement of the crankshafts, connecting rods, and bearings in the SNPS DSR-48 diesel engines, the babbitt-overlayed cast aluminum connecting rod bearing shells supplied by TDI are suitable for their intended purpose.
Minor surface imperfections in the babbitt overlay will not degrade the suitability of the bearing shells.
The analysis was calibrated against experience with the original 11-inch diameter bearing shells installed in the SNPS DSR-48 diesel engines.
This experience, in which four upper connecting rod bearing shells cracked after approximately 600 to 800 hours0.00926 days <br />0.222 hours <br />0.00132 weeks <br />3.044e-4 months <br /> of operation including about 250 hours0.00289 days <br />0.0694 hours <br />4.133598e-4 weeks <br />9.5125e-5 months <br /> at full load, has established the fatigue life of the 12-inch bearings currently installed in the SNPS OSR 48 diesel engines, and in the 13-inch bearings in Grand Gulf Nuclear Station DSRV-16-4 engines, to be 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> of engine operation at full load, provided the bearing shells meet the acceptance criterion of Section 5.3.
By going to the larger bearings and decreasing the connecting rod chamfer size, the stress range which led to fatigue cracking of the original connecting rod bearings has been reduced by a factor of two, thereby raising the predicted fatigue life of the bearings to 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> at full load. The connecting rod bearings in the DSRV-series engines, as exemplified by the DSRV-16 4 diesels at the Grand Gulf Nuclear Station, have a stress distribution which very nearly matches the stress distribution in the Shoreham DSR a8 bearing shells. Thus, the same conclusion applies to the DSRV engines:
that is, that the expected fatigue life of the bearings should be on the order of 38,000 hours0 days <br />0 hours <br />0 weeks <br />0 months <br /> of full load operation.
9
]
6-1
.,.7.;==m..,-..--n=.
. = - - -
3:., --.
.,_~ - _.: ;
In the case of Shoreham, the testing projections developed by the architect / engineer and based on NRC requirements show that tne engines will be run at full load approximately 100 hours0.00116 days <br />0.0278 hours <br />1.653439e-4 weeks <br />3.805e-5 months <br /> in every two-year cycle.
Over the expected 40-year life of the power station, or 20 two-yea r cycl es,
approximately 2,000 total hours of full load running should be accumulated.
The ratio of the stress level in the connecting rod bearings currently installed at Shoreham and at Grand Gulf, relative to the stress level in the original 11-inch bearings at SNPS, provides a criterion for the acceptance of bearing shells.
Calculation of the acceptable size of a normal casting void
~
which would remain below the computed threshold stress intensity factor range shows that discontinuities up to 0.050 inch may be allowed in the critical zones of the upper connecting rod bearings.
Inspection to assure compliance with the acceptance criterion can be accomplished by radiography.
Shoreham Nuclear Power Station has developed a procedure that detects discontinuities that could result in rejection of a bearing shell.
A sampling procedure will be recommended to th TDI Owners Group for radiography of bearing shells for purposes of quality revalidation.
t 6-2 t
.--~_..--_-g.-
OR-03-340B-1 Appendix COMPONENT DESIGN REVIEW TASK DESCRIPTION CONNECTING R00 BEARING SHELLS Classification A PART NO 03-3408 Completion 3/1/84 PRIMARY FUNCTION:
The connecting rod bearing shells provide the oscillating sliding surface between the connecting rod and the trank pin through the formation of a hydrodynamic oil film.
They transmit the cylinder firing pressure to the crankshaft through the oil film, converting the force into torque.
FUNCTIONAL ATTRIBtJTES:
1.
The bearing shells must have sufficient fatigue life and wear resistance to tolerate normal operating conditions for the intended service.
2.
The bearing material must be of low friction to tolerate possib'e momentary contact with the crankshaft during starting of tt ?
engine, and the surface of the bearing shell should be constructe.t of a material which is tolerant to the presence of foreign particles minimizing journal wear.
3.
The dimensions must be manufactured with sufficient accuracy to obtain the proper interference fit in the connecting rod, and to establish the specified clearance between the bearing shell and the crankshaft.
4 The bearing must be designed so that during operation key parameters including oil supply pressure, peak oil film pressure, minimum oil film thickness, and oil film temperature rise are within acceptable limits for the specified diesel engine application and required life.
5.
The bearing material should be resistant to possible corrosion due '
to chemical composition of lube oil.
A-1
-NM
- 4
y-h M '
N1We M. t\\
UN+5 TC'D*""
e
+
SPECIFIED STANDARDS: None EVALUATION:
1.
Dbtain cylinder pressure vs. crank angle data from DSR-48 test and compare to assumptions for previous bearing shell design review.
2.
Review cylinder pressure vs. crank angle for DSRV-16 4 design.
3.
Perform journal orbit analysis of DSR 48 design.
4 Perform finite element analysis of DSR-48 design.
5.
a.
Make fracture mechanics life estimate of DSR-48 design.
b.
Determine maximum void size in bearing castings for radiograph inspection acceptance criteria.
6.
Perform journal orbit analysis of DSRV-16-4 design.
7.
Perform finite element analysis of DSRV-16 4 (if required by Item 6).
8.
a.
Make fracture rechanics life estimate of DSRV-16-4 design (if required by Item 6).
b.
Determine maximum void size in bearing castings for radiogrcph inspection acceptance criteria (if required by Item 6).
9.
Make physical examination of used DSRV-16 4 bearing shells frcm GGNS to determine elastic deflection patterns.
10.
Evaluate effects of babbit adhesion and thickness variations.
- 11. Complete report on DSR t8 and DSRV-16 4 bearing shells in SNPS and GGNS engines.
12.
a.
Determine differences, if any, between DSRV-16 4 and DSRV-12-4, DSRV-2D 4 Conduct necessary design review steps, issue final report covering all engines.
b.
Evaluate possible preventive maintenance and monitoring procedures (i.e., oil sample particulate / chemical analysis, periodic visual inspection).
A-2
-,-w,p
_~ RE
,MM -'
goo
REVIEW TDI ANALYSES:
1.
Obtain any available journal orbit analyses.
2.
Review any bearing f ailure analyses.
INFORMATION REQUIRED:
1.
Manuf acturer's drawings of bea ri ngs, connecting rods, crankpin journals.
2.
Cylinder firing pressure versus time for DSRV-16-4.
3.
Lubrication oil specifications.
4.
Connecting rod rotating and reciprocating weights.
e 4
0 A-3
.-,~~.-n=.,.
.c
_.. = _ _.
March 1984 1
EMERGENCY DIESEL GENERATOR ROCKER ARM CAPSCREW STRESS ANALYSIS
/
Prepared for:
TDI OWNERS GROUP Prepared by:
Stone & Webster Engineering Corporation
< ~,
v
,tfQOkO i fpil 9Yff
t EMERGENCY DIESEL GENERATOR ROCKER AR?i CAPSCRD7 STRESS ANALYSIS Prepared for TDI EMERGENCY DIESEL GENERATOR OWNERS GROUP MARCH, 1984 Prepared By STONE & WEBSTER ENGINEERING CORPORATION O
_a TABLE OF CONTENTS Section Title Page 1
APPLICABILITY l
2 EXECUTIVE
SUMMARY
2 3
OBJECTIVES 3
4
SUMMARY
OF SERVICE CONDITIONS 4
5 METHODS OF ANALYSIS 5
5.1 Determination of Applied Stresses 5.2 Determination of Endurance Limits 5.3 Thread Distortion Analysis 5.4 Thermal Stress Evaluation 6
DISCUSSION OF " 2TLTS 9
7 CONCLUSIONS 11 APPENDIX A STRESS
SUMMARY
APPENDIX B FATIGUE / LOADING DIAGRAMS APPENDIX C COMPONENT DRAWING APPENDIX D TASK DESCRIPTION APPENDIX E REFERENCES
/
_Y a
...~.j,..
., ~
SECTION 1 APPLICABILITY This report is applicable to the TDI Nuclear Stand-by Service Diesel Generators utilized r.t the Shoreham Nuclear Power Station.
Other TDI Nuclear Stand-by Service Diesel Engine Rocker Arm Capscrews, as part of the TDI Owners Group Design Review / Quality Revalidation effort, will be evaluated seperately.
Reviews will include identification of the specifie Japscrew design installed and a ecmparison to the Shoreham's analysis as to differences in materials and loading conditions, if any.
?
j i
~
SECTION 2 EXECUTIVE
SUMMARY
An analysis was conducted for two rocker arm capscrew designs utilized in the TDI Nuclear Stand-by Service Diesel I
Generators.
This analysis served to evaluate the various rocker arm capscrew product improvements incorporated by the manufacturer.
The designs considered were the original " straight section" type and the modified " necked down" design as detailed in Reference 1.
The purpose of this analysis was to evaluate these capscrew designs based on the criteria referenced in the Component Design Review Task Description, (see Appendix D).
This analysis concluded that both original and modified rocker arm capscrews are adequately designed for the given service conditions.
G mi
... ~ _... _
...._s.-
- _ _ _ _ ~ _.... - _.
SECTION 3 OBJECTIVES The objectives of this analysis are to evaluate the functional attributes of the TDI rocker arm capscrews as detailed in the Component Design Review Task Description (see Appendix D).
The capscrews under consideration include the original and modified TDI designs as detailed in Reference 1.
Task Description details include:
1.
Evaluate the stresses at the minimum cross-sectional area (A(min)) resulting from capscrew preloads, cam follower acceleration loads, valve spring loads and residual cylinder pressure forces.
2.
Compare the total resultant stress to yield and endurance limits for the capscrew.
3.
Evaluate the thread specification for resistance to distortion and creep.
4.
Compare the material utilized in the rocker arm capscrew to ASTM A-193.
The details of this analysis are provided in SWEC calculation number 11600.60-245.1-M1 (Reference 2).
Ql
' t
~,
v.
. - - - - -. -.. ~..
r...%
....p.
op._
p
,5
_-=
SECTION 4
SUMMARY
CF SERVICE CONDITIONS The rocker arm capscrews serve to transmit camshaft follower acceleration loads, valve spring loads and residual cylinder
~
pressure forces from the rocker arm shaft to the cylinder heads.
To achieve these design objectives, the rocker arm capscrews are preloaded with a specified torque.
This produces a tensile load p) on the capscrew which is and subsequent preload stress (S y
converted to a clamping force on the rocker arm shaft.
I-addition to the preload stress, a cyclic fatigue stress (S ) is induced on the capscrews due to the above referenced b
loads.
This cyclic stress alternates about the total applied mean stress (S ) as shown in Appendix B, Figure 1.
t 6
_4
~ _ _. _. _
=
SECTION 5 METHODS OF ANAYLSIS 5.1 Determination of Applied Stresses In determining the stresses applied to the orignal and modified rocker arm capscrews, the tensile preload resulting from the applied torque is first calculated by the relation py= T/.2d Eq. 1 (Reference 3)
F where T= applied torque, LB -in f
d= nominal stud diameter, in due to torque T F y= stud preload, LBf p
The resulting tensile load is taken to act uniformly over the minimum cross sectional area (A(min)) of the capscrew so that the resulting preload stress is maximized.
For the original design, A(min) is located at the capscrew thread root whereas for the modified design A(min) is at a smooth, polished " necked down" section between the threads and the capscrew head (see Appendix C Figure 1).
The tensile loads produced in the capscrews are converted to a clamping force on the rocker arm shaft which serves to resist the residual exhaust pressure, valve spring and cam-follower acceleration loads.
The residual exhaust pressure loads are calculated by multiplying the cylinder pressure at the exhaust valve opening point (130*ATDC) by the valve face area.
The valve spring forces are calculated by multiplying the full spring deflection by the applicable spring constant.
These forces are transmitted to the intermediate rocker arm via the intermediate pushrod.
In determining the camshaft follower acceleratica
loads, the follower velocity and acceleration are calculated by taking first and second derivatives of the follower displacement as a function of can rotation.
To obtain the maximum load condition, the mass of followers is multiplied by the maximum value of the acceleration.
The cam follower loads along with the pressure and spring loads are resolved into a single vector acting on the intermediate rocker arm as detailed in Appendix B Figure 4, As detailed in Reference 2, the total preload on the rocker arm shaft exceeds the resultant forces by a factor of nine (9).
For components proloaded in this manner, the cyclic stress due to an alternating load is given by:
/2 Eq. 2 (FefereMe h Sb=+Frb
-A (min)
's Where sb = alternating stress Frb * '?
E
= alternating load r
b E *E b m and K3 = spring rate for the stud K,= spring rate for the section joined by the capscrews F = mar.inum resultant force l,
t
~
and K are As per reference 3, the spring rates Kb determined by:
2 Kb = frd E and K
= 2rrd E 4L L
where d = nominal capscrew dameter, in E = modulus of elasticity, LB /in f
L = length of joined sections, in p) and de alternadng stress Once the preload stress (S y
(+S i are determined, the total applied mean stress (S ) is b
t defined as:
St" pl* b i
o o
s er#.
1 v._.=
w gg%%e,....y..
g.e...g_.=
~
5.2 Determination of Endurance Limits The endurance limits (S ) f r the original and modified e
designs are determined by:
S
= K,K.
K K b
q.
( e erence 3) e d
e f
e S
the endurance limit of a rotating beam specimen K,'== component surface factor where g
K3 = component si.te factor K
= reliability factor e
temperature factor Kd = modifying factor for stress concentration K
=
e Kf = miscellaneous effects factor The endurance limit of the two designs is conpared,as well as the total applied mean stress (St) and the yield strength (S )
of the material in order to determine the suitability of the component design for the given service.
s e
O
., =..
s.3 Thread Distortion Analysis To determine if thread distortion occurs during preload application, the maximum no2 mal stress (Sn) at the capscrew minimum area is calculated by:
I 2+S 2 Reference 3 S
- S pl *
[ pl s
n g(
2 2
where S
= torsional stress due to tightening s
16T Per Reference 3, S
=
s 3
frdr where T = 75% of the applied torque, in-LB f dr = root or minimum diameter Based on the maximum normal stress theory, thread distortion and subsequent failure during preload application will not occur providing Sn4Sy.
5.4 Thermal Stress Evaluntion Thermal stresses in the rocker arm capscrews are qualatatively evaluated by considering the physical constraints of the joined components and the coefficients ci' thermal expansion for the materials utilized.
As detailed in Reference 2, the thermal stress (Stl) in the capscrew is defined as:
Stl =
(
ra -
cs) Ecs where X = coefficient of thermal expansion 'F-Tra =
chr.nge in temperature of the rocker arm
- assembly,
'F
(,.)
Tes =
change in temperature cf the
- capscrew,
'F 2
Ecs =
modulcus of elasticity, LB /in f.
SECTION 6 DISCUSSION OF RESULTS er Reference 5, fatigue failure in bolts and studs is Aded if the total applied mean stress (At) is below the yield point for the component material and the endurance strength exceeds the cyclic stress by an acceptable margin.
As per Appendix A Tables 1 and 2, the total applied mean stress (St' taken at the minimum cross sectional areas A(min) is 41.0 ksi and 50.8 ksi for the original and modified stud designs respectively.
The higher value of S f r the modified design is primarily due t
to the " necked down" nature of the minimum area.
Per Reference 4, the yield strength for th2 stud material is approximately 105 ksi.
As a result, the criteria for a fatigue resistant design, namely S b
and S 6
are satisfied for both t
y b
e designs.
This is graphically depicted by Appendix B, Figures 2 and 3.
As shown in Appendix A, Table 1, the modified design utilizes an endurance strees of 37.6 ksi with a cyclic stress of
+0.6 ksi (see Appendix B, Figure 3) and therefort, satisfies the second criteria for a fatigue resistant design.
This increase in endurance strength is primarily attributed to the location of the minimum area.
As detailed in Reference 2, A(min) is located in a polished section of the capscrew, free of any significant stress concentration factors.
This allows for K and K to be set equal a
e
)
,#**4
/
- 1
to 1 in Eq. 3, thereby increasing the endurance limit (Ref. 2,3).
Note K, and K are less than 1 for the original design since the e
minimum area is located at an unpolished, high stress concentration section (thread root) of the capscrew.
During engine stand-by and operation, minimal thermal and creep effect.s are present in the rocker arm capscrews.
Due to the relatively uniform temperature distribution within the rocker arm assemblies, (i.e.
Tra = T:s) effects due to thermal stresses are negligible.
Additionally, since the capscrew is confined to the rocker arm assembly and experiences relatively low temperatures, effects due to creap and stress relaxation are also negligible.
Since the yield strength of the material is not exceeded, thread distortion will not occur for the given service conditions and installation procedures.
Additionally, bending stresses on the capscrews are negligible due to the relatively low cyclic forces acting on the intermediate rocker arm (see Appendix B, Figure 4).
The material specifiec for the modified rocker arm capscrew is AISI 4140 hardened to 25-30 (Rockwell).
This material meets or exceeds the requirerents of ASTM A193 Grade B7.
.. ~..
^
SECTION 7 CONCLUSIONS Reference 2 provides an analysis of the total applied mean stress (S ) and the endurance limit for two TDI rocker arm t
capscrew designs.
Based on this analysis, both original and
~
modified designs (Reference 1) are adequate for the intended service.
However, the modified design is more fatigue resistant due to the relatively high component endurance limit.
The threads utilized for both original and modified rocker arm capscrew designs adequately resist distortion.
Additionally, the material utilized for the modified rocker arm capscrews meets or exceeds the requirements of ASTM-A193.
t
.i
, a i
APPENDIX A - STRESS SUPJmRY TABLE 1: STRESS SUPJRRY - ORIGINAL DESIGN TABLE 2: STRESS
SUMMARY
- MODITIED DESIGN
\\
,4
p l
TABLE N 1
i STRESS
SUMMARY
i i
ORIGINAL DESIGN l
l S PL SB SE ST KSI KSI KSI KSI 40.5 0.5 8.7 41.0 h
S PL-PRELOAD STRESS l
S e - ALTERNATING STRESS S E - ENDURANCE LIMIT S T - TOTAL APPLIED MEAN STRESS I
d t
e i
l (L 1
s,
TABLE N2 2 STRESS
SUMMARY
~
MODIFIED DESIGN SPL
i 50.2 t0.6 37.6 50.8 t
I i
f SPL-PRELOAD STRESS Ss - ALTERNATING STRESS SE - ENDURANCE LIMIT ST - TOTAL APPLIED MEAN STRESS 9
6 0
6 I
4 tw.
I
~
~
.=
APPENDIX B - FATIGUE / LOADING DIAGRAMS FIGURE 1:
CYCLIC FATIGUE LOAD DIAGRAM FIGURE 2:
FATIGUE DIAGRAM - ORIGINAL DESIGN FIGURE 3:
FATIGUE DIAGRAM - MODIFIED DESIGN FIGURE 4:
INTERMEDIATE ROCKER AR!!-FORCE DIAGRAM 4
a t_. #.
FIGURE N21 CYCUC FATIGUE STRESS DIAGRAM i
n t
i A
Onn f 'V V
V ()
s 72 ST N~
N-1 ST-TOTAL APPLIED MEAN STRESS
$B ALTERNATING STRESS U
t
+4
-,e.. ym..
.e
- w+ e w
.,---*.+--.-w-w.
- -w, e---
a
~
~ ~
FIGURE N22 FATlGUEDIAG GOODMAN LINE ORIGINAL DESIGN 10-8--
Ui 6-x IW
(.h 4-POTENTIAL FAILURE ZONE 2-
_ _.. _ _ _ _ _.o 2O 4O 6O 80 lbO 1bO I
UTS Ksi Ss- 0.5 Ksl
.o*
- O ST=41 Oksi 4
..-n
.,~..-. ~. ---. - - -.
a FIGURE N23 FATIGUE DIAG GOODMAN LINE MODIFIED DESIGN 36-30--
i 24-Di 18-x tW 12I-POTENTIAL FAILURE ZONE 6"
. -._ _ _ _ _._ - 9 I
IO 40 60 BO 10 0 lbO UTS-Ksl Se-* O.6 Ksi
,;b.
ST 50.8Ksi
s.
= _ _.
FIGURE N24 INTERMEDIATE ROCKER ARM FORCE DI AGRAM 6*y FvS+Fve FR 21 o "O"
> ~ ~,
(FCF 27 i
e I
I SCALE: 1"=1000 LB F
- FORCE VECTOR
: FORCE LINE EXTENSION "O'
- FORCE INTERSECTION POINT t
,t wm.,
APPENDIX C - COMPONEN" DRAWING FIGURE 1:
MODIFIED ROCKER ARM CAPSCREW
~
a*98W hem
i I'
t FIGURE NR1 MODIFIED ROCKER ARM CAPSCREW i
l r
I,
(
F
(
3 s
NOTES
- 1. NO SCALE
- 2. REF TDI DWG O2-390-04-OJ l
-=a.,
--i---.+,,--,--m,.ww.,-.%-,
, ese APPENDIX D - TASK DESCRIPTION COMPONENT DESIGN REVIEW: DR-03-390G-1
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M'
,g
~
.m n.
,...s
-.w - - -_ --- -
- a...x a m m s..~. s o s. * *. ~.
DR-03-390G-1 COMPONENT DEEIGN REVIEW T A 5 r'. DEECRIPTION ROCKER ARM CAPSCREWE Classification B pART NO. 02-290G Coracl et i on 3/1/BA PRIMARY FUNCTION:
The rocker-arre
,t
';.c r e ws t ra nstal t resultant l ead s f rota the valve sprir.gs, valve opening p-are cushr ods, and rocker a ria assemblies to the sub cover and cy l i nde-r herds.
~
FUNCTIONAL ATTRIBUTES:
1.
The rocker arra capscrews raust have sufficient strength to withstand the necessary preload and oscillation lorde without fat 2gue cracking, unacceptacle preload relar.ation or t i. r e., d distortion.
SPECIFIED STANDARDS:
None E:t iraat ed EVALUATION:
Ggypletion 1.
Deserraine the stud d i rae ns 2 or.s f rora existing des s pr:
2/17/84 drawings and evaluate the stress at the rai ni ta ura cross-sect 2onal area result 2ng t rora the eppl2ed preloads ass urai ng uni f or ra t h 'eed lub-ication and load distribution.
St res: concent rat ion f a::t ors f o t-the thread root area will ce a nc 1 '.tc ed in the analysis for the previous TDT de w s p n.
2.
Det erraine t he stresses exper2enced at t h e :.i n 2, "--
O/20/84 area resulting frora push rod rest lon, velve spring deflection and valve openang pressure.
3.
Det erraine t he total resultant bolt stceso end 2/20/84 compare to yield and endure,nce 12mits.
4 Compe.re capscrew design and raat e r a a l specif2catlon to ASTM A-193.
5.
Evaluate the thread specificat son for resi st a nc e to distortion and creep.
6.
Perform sirailar analysis on the previous uni f ce ra cross sect 2cn capscrew des 2gn.
REVIEW TDI ANALYSES:
1.
Review any TDI st ress analyses associated with accign/
material changes.
INFORMAT10N REQUIRED:
Est isaat ed C20 Die.ti2D 1.
Caoserew preload (hold-down force)
- / 9 / 84 2.
Capscrew lubrication 2/9/84 3.
Capscrew design drawings and material specific at tons 2/9/84 4.
Rocker arin geometry and drawings, 2/9/84 f
5.
Valve spring constants, free len;th, cornpres sed O/?.'84
~
'y, '
length 6.
Oper at ing loads on the raoscrews 7.
Valve pop-open pressure in cylinder 2/25/84 e
~
- e p= rr
APPENDIX E - REFERENCES Reference 1:
TDI Drawing 02-390-04-0G - Original Design TDI Drawing 02-390-04-0J - Modified Design Reference 2:
SWEC Calculation 11600.60-245.1-M2 Reference 3:
Mechanical Engineering Design, J.E. Shigley, 3rd edition
\\
Reference 4:
Engineering Properties of Steel, ASM 1982 Reference 5:
Simple Diagrams Aid in Analyzing Forces in Bolted Joints, Assembly Engineering, G. Meyer 1972
. = -.,
m
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