ML20100R078

Intervenor Exhibit I-46,consisting of Discussing & Forwarding Info & Repts Re Tdi Diesel Generator Crankshafts

ML20100R078
Person / Time
Site:	Shoreham File:Long Island Lighting Company icon.png
Issue date:	10/01/1984
From:	Dynner A KIRKPATRICK & LOCKHART
To:	Blanding H AMERICAN BUREAU OF SHIPPING
References
OL-I-046, OL-I-46, NUDOCS 8412170303
Download: ML20100R078 (100)

Text

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KraxPArnIcx, Locun.urr, Hru., CHRISTOPHER & PHru.IPS cm...,m co.,-.n-1900 M SrREET, N. W.

WASHINGTON, D. C. 20006 s4as suscaza.r. AvENUs Tur.arnows:(sos) 4ss-rooo isoo or.rvan acII. naso MIAMI, F14RIDA 33138 FITT5BcRGE, PENNSY1.VANIA !s222 mm moe mira er (sos) or4-eus g,,),,,,,,,,

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Mr. Howard C. Blanding cyh 8

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'iechnical Division, Machinery Department 65 Broadway New York, New York 10006

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Dear Mr. Blanding:

During the July 16 depositions of Robert Woytowich, Robert Giuffra and yourself by Suffolk County, ABS personnel 2

testified that the' ABS, in issuing its May 3, 1984 letter to Transamerica Delaval Inc. ("TDI") concerning the Shoreham crankshafts, relied on strain gage test measurerrents, service experience and the shotpeening of the crankshafts.

You also i

stated that if any of this material information supplied to the ABS by TDI were incomplete, incorrect or misleading, the ABS would have to reconsider the conclusions stated in the May 3 letter.

In this connection we are notifying you'of the following information:

1.

Page 24 of TDI's submittal to ABS stated that a-conservative minimal value of the increase in the fatigue j

endurance limit of the replacement crankshafts from shotpeening is 20 percent.

The ABS used this 20 percent increased fatigue limit value in reaching the conclusions stated in the May 3 letter.

TDI's Manager of Engineering testified in his deposition, however, that TDI had recommended against shotpeening the replacement crankshafts on the DSR-48 engines at Shoreham based upon experience and on the opinion of TDI.'s metallurgical consultant that shotpeening would not provide more than a 5 percent improvement in the fatigue strength of the crankshaft (see document 5 below at pages 45-48).

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' EruxPATmICX, EWA*T, Hrzz, CumssroFumm & Pnzz.r.rrs Howard C. Blanding July 25, 1984 Page Two Two of the replacement crankshafts were determined to have been. inadequately shotpeened in the lower third of the fillet (see documents 1 and 2 below).

The surface texture of the shotpeening was reported to appear more like grit blasting than shotpeening, that is, as if it had been " gauged (sic] by sharp particles instead of dimpled by round, smooth particles" (see~ document 3 below at pages 63-70).

These.two crankshafts were re-shotpeened (see document 3 below at 64-65).

2.

ABS used 1700 psi in its calculations for the maximum firing pressure.

TDI Factory Test Logs reported firing pressures for the DSR-48 engines at Shoreham in excess of 1700 psi, and as high as 1800 psi (see document 6 below).

TDI's Manager of Engineering testified in his deposition that the maximum firing pressure in the cylinder during operation of.the DSR-48 engine was.approximately 1800 psi at operation at overload of approxi-i mately 3900.kW (see document 5 below at 128-29).

The Engine Cylinder Pressure Log for a DSR-48 engine at Shoreham shows firing pressure in excess of 1700 psi (see document 7 below).

Mr. Museler, then the Technical Manager of the TDI Owners Group program and Director, Office of Nuclear, of Leag Island Lighting Company, stated at a March 22, 1984 meeting between the staff of the Nuclear Regitlatory Commission and-the TDI Owners Group that the " normal ' firing pressure on the DSR-48 engines at Shoreham is 1670 psi and 1750 psi at overload -(see document 9 below).

TDI'.s Instruction Manual states that che. firing pressures a.le considered normal if within plus or minus 75 psi of the average for all cylinders and that the firing pressure limits for sustained operation is 200 pai between any two cylinders (see document 10 below).

3.

Mr. Giuffra testified in'his deposition to the effect that prohlane wi*b = Mfa=al argine weuld aurface within 100 hours0.00116 days <br />0.0278 hours <br />1.653439e-4 weeks <br />3.805e-5 months <br /> of operation.

One of the original 11 x 13 inch crankshafts on the TDI DSR-48 engines at Shoreham fractured after 718 hours0.00831 days <br />0.199 hours <br />0.00119 weeks <br />2.73199e-4 months <br />, 22 of which were at the 2-hour overload rating (h3850 kW).

The l

other two 11 x 13 inch crankshafts in the DSR-48 engines at Shoreham were found to have cracks after 646 and 818 hours0.00947 days <br />0.227 hours <br />0.00135 weeks <br />3.11249e-4 months <br /> of operation, 19 and 23 of which were at the 2-hour overload rating (see document 3 below at page 15).

4.

Page 28 of TDI's submittal to the ABS showed that, as of April 1, 1984, EDGs 101, 102 and 103 at Shoreham had 114, 116 and 110 hours0.00127 days <br />0.0306 hours <br />1.818783e-4 weeks <br />4.1855e-5 months <br /> of operation at 3500 kW and above.

One of the ABS deponents testified that he didn't know how many of those 1,

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Krunnrancx, Locrwimr, En.r., Cuarsromrum & Parz.z.rrs j

Howard C. Blanding July 25, 1984 Page Three hours were at-loads actually above 3500 kW.

Failure Analysis

-Associates reported that, as of April 30, 1984, Shoreham EDGs 101, 102 and 103 (with the replacement crankshafts) had 91.5, 70.5 and 47.5 hours5.787037e-5 days <br />0.00139 hours <br />8.267196e-6 weeks <br />1.9025e-6 months <br /> of operation at loads between 100 and 110 percent load, and 6.5, O and 7 hours8.101852e-5 days <br />0.00194 hours <br />1.157407e-5 weeks <br />2.6635e-6 months <br /> of operation at 110 percent load or greater, respectively (see document 4 below).

Page.28 of TDI's submittal to ABS referred to operating hours of DSR-48 engines at various installations.

This informa-tion does not show whether any of the crankshafts had been inspected with dye penetrant, eddy current or other means of nondestructive examination for indications, or what the results were if such inspections had been made.

4 5.

The ABS relied on the strain gage measurements submitted by TDI.

During the test on the replacement crankshaft, the firing pressure was measured by inserting a piezoelectric pressure transducer (AVL 5007) in the air start valve of cylinder number 7.

FaAA calculated the torque produced by this pressure and obtained a mechanical efficiency of 1.0 (see document 8 below).

Firing pressures also were measured in the e

compression test cocks of cylinder numbers 5 and 7 with two piezoelectric pressure transducers (PCB Model ll2A).

SWEC reported that the strain measurements were accurate to within 15 percent and that the output torque measurements had a probable error of 5 to 8 percent (see document 11 below at 7-3, 6-2, 6-3).

In the strain gage test on the failed crankshaft, different model transducers (PCB Model lllA) were installed in the compression test cocks of cylinder numbers 5 and 7.

SWEC concluded from the recorded pressure measurements that the dynamic measurements were affected by the gaseous column and flow path geometry associated with the pressure test cock.

These cylinder pressure measurements were unacceptably low (see document 3 below at 52).

6.

FaAA performed a mode superposition analysis of the dynamic response of the replacement crankshafts at.3500 kW and obtained average torsional stress values of 3300 psi due to the 4th order and 5640 psi due to the summation of the orders (see document 3 below at 60, 62).

5640 psi exceeds the allowable stress levels under the 1984 ABS rules (see sheet 6 of 6 to ABS calculations).

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r KraxPATmICX, ENwwA*T, Hsz.r., CantsToenza & Parz.z.rps Howard C. Blanding July 25, 1984

-Page Four 7., The Shoreham engines are expected to run at the loads listed in document 12 below.

Enclosed are the following documents or information referred to above which were supplied to'Suffolk County by either the Nuclear Regulatory Commission, TDI or Long Island Lighting Company:

1.

Stone & Webster Engineering Corporation's ("S&W")

Engineering & Design Coordination Report ("E&DCR") No. F-46109G, dated 9/16/83; 2.

Interoffice memorandum dated September 20, 1983; 3.

Pages 15-16 and 47-70.of Franklin Research Center's Technical Evaluation Report, Evaluation of Diesel Generator Failure at Shoreham Unit 1, Final Report, Failure Cause Evalua-tion, dated April 6, 1984; 4.

Tables of loads on EDGs 101, 102 and 103;.

5.

Deposition testimony of G. E. Trussell, TDI's Manager of Engineering, May 7, 1984, pages 1, 45-48, and 128-129; 6.

Test bed data for the TDI DSR-48 engines; 7.

Appendix F to Shoreham Preoperational Test Results Review and Approval, Engine Cylinder Pressure Log 'for EDG 101; 8.

Evaluation of Emergency Diesel Generator Crankshafts at Shoreham and Grand Gulf Nuclear Power ~ Stations, Failure Analysis Associates, May 22, 1984, pages 3-2, 323,'3-12 and figure 3-1.

9.

Statements of William J. Museler at the March 22, 1984 meeting between the Nuclear Regulatory Commission staff and the TDI Owners Group, pages 70-72.

10.

TDI's Instruction Manual, Appendix II, Operating Pressures and Temperatures at 8-3.

11.

Bercel, E. and Hall, J.R.

" Field Test of Emergency Diesel Generator 103," Stone & Webster Engineering Corporation, April 1984, pages 6-2, 6-3, 7-3.

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KraxrArarcx, LocxHAmT, Erz.T., Cumrsrorumm & Parz.r.res Howard C. Blanding July 25, 1984 Page Five l

12. - Shoreham Final Safety Analysis Report, Table 8.3.1-1.

Very truly yours, f,

Alan Roy Dyr er ARD/ss Enclosures 1

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C,,;'8En t$ne fillet areas of the new diesel crankshafts, purchased in escordance with E&DCR T-46109C.

These holidays have been d:.spositioned p

Qais functionally acceptable by CI, however, recent analysis performed t

by Pa11ure Analysis Associates indicate that 1002 peening coverage is beneficial.

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Metal Improvements Co.,

a local firm with extens peening of crankshafts, will perform the rework.ive exper:ence d. shot The fillet areas shall be repeened in accordance with the requirements of MIL-s-131653 to assure 100t coverage of the fillet areas.

Peening shall be performed by Metal Improvements Co. on site and the crankshaft inspected by OQA for 1001 peening at the fillet areas.

Refer to attached procedure.

CI QC inspection of journal bearing masking is required pr: or to com:.encing shot peening.

CI approval for shot peening procedure has been obtained TELECC" DATE!

AFfrNS WOpv Usero spretricancy gx t ge{ written approval to be filed wit.

(AESPONCWG PART Q65 VERFIED SY R/ R at C.l.o s e o u ::.

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September 12, 1983 To:

John Karreneyer From:

Ken Kropf

Subject:

Diesel Generators TDI S/N 74010-12 Holidays in Shot-Peening on Crankshaft 1693, PCf 8152, HTl 821487

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There are two areas on top of f 1 Pin directly adjacent to the crank pin and at the outer edge of the crank radius that have Holidays in the Shot-Peening.

These holidays are in a relatively low stressed area of the crankshaft.

I have looked at these areas and disposition them functionally i

acceptable.

The TDI procedure for shot peening the LILCO crankshaft is also

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l.'600.17 aomo SUBJECT 3:ISEL CRA:."ASF_ATT TILLITS

• CATE' SEPTIM3ER 12. 1983 EINING FROM GARY V.

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preliminary analysis performed by FaAA -(as w...

fate) finds-tha: the 1 ewer 1/3 ef the reentry fillet at the crankshaft pin jun= tion is the mes; critical area with respeen to crankshaft failure:

The reascns given are:

(1) High stress at loading (the crankshaf: =ay be near a harmonic at its leaded running.

speed.

(2) A residual stress (deter =ar.ed by x ray diffraction) a: the fillets caused apparently by machining.

(3) Surface finish at the fillets.

Although the finish does =ct appear to be rough to the naked eye (or by feal). I have seen scanning electron macresc=pe phetegraphs of cracks in initial stages of propaga:icn.

The cracks appear ::

be initiating at ene of the radially machined

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!= order :: reduce the residual tensile stresses at the fillet (and also to reduce the degree of valley alignrent with the pr =ary tensile stress planes), we have specified peening to be perfer:ed (at TO:).

However, review of the as-received cranks fcund tha:

they were inadequately peened at loca : ens of interest (tc FaAA; e.g.

the lower 1/3.cf the fillet.

FaAA recc== ended that the peening be redene by a " Metal : provement Co.'

, specialists in centrolled peening.

MIC already performs peening to crankshaft fillets for other diesel ma-cturers.

M: 's equipment was transportec to the site and the peening was performed in'a building erected to house the crankshaft.

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-I-MIC reported that the fatique life of a prcperly s ened crankshaf t can be increased by 100% ty proper p2ening.

CC Checks - Pee.i-c

.100% peening is defined as areas where the surface is totally di= pled.

They have invented a "peenscan meshed" to verify 100% peened surfaces.

The process utilizes an ultraviolet solu:1cn (the peccess is semewhat similar := a UV =agnaflux check) which the UV sensi :ve solution is-applied to the area Oc be peened.

1004 peening is defined by a :::al remeval ef the ultraviole sensitive solution.

An ultraviolet light is used := verify its remeval.

MIC-GL - Cerversatiens

.I questioned the MIC engineer en repeening previously peoned surfaces and his reply was that peening typically teccmes de:: mental to fatique life only if the peening abrades a *.rtt cal section.

He also stated tha: their equipment is designed for accessability and con:ref and that their Operaters are experienced enough so that abrasion of peened sections is not a problem.

Attached is a cample of MIC's shet: you will note its uniformly round gecmetry which'is much "better" than ec=mercial cleaning shot.

The tape which the shot-is attached to is similar to duct tape and is used for masking areas specified as et to be peened.

z Overall. I was impressed with the cc=pany's centrol and expertise.

The ecmpany essentially wrote MIL-5-13155 (which we inveked en the crankshaf t).

MIL-S-13165 is much nere definitive overall in the peening process than any SAE specifica: en I have seen.

I requested that they send a pr duct brochure to Boston describing their peccesses and any data they they have concerning the beneficial effects for other applications.

Other applications they cited T.ay be reduction of residual stress en turbine shafts, and reduction of residual tensile stresses where stress ccrresi n crackgpq is a concern.

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.aM N G.V.

Luther P.S.

The new crankshafts reentry fillet geometry is -uch better than the old one.

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TECHNICAL EVALUATION REPORT EVALUATION OF DIESEL GENERATOR FAILURE AT SHOREHAM UNIT 1 FINAL REPORT, FAILURE CAUSE EVALUATION I NRCCONTRACTNO.N NRC DOCKET NO. 50-322 FRC PROJECT C5506 NRC TAC NO. -

FRC ASSIGNMENT 20 FRC TASK 426 Prepared by Franklin Research Center Author: R. C. Herrick 20th and Race Streets Philadelphia, PA 19103 FRC Group Leader:

S. Ah:ned Prepared for Nuclear Regulatory Commission Lead NRC Engineer: R. J. Giardina Washington, D.C. 20555 April 6. 1984 This report was prepared as an account of work sponsored by an agency of the United States Government. Neitner the United States Government nor any agency thereof, or any of their employees, makes any warranty, expressed or implied, or assumes any legal liability or responsabahty for any third party's use, or the results of such use, of any information, appa-ratus, product or process disclosed in this report, or represents that its use by such third party would not infringe pnvately owned nghts.

Prepared by:

Reviewed by:

Approved by:

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Principal Author Project Manager Department Director (Acting) fM4 Date:

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test program, including all mechanical, electrical and qualification tests, was completed for all three diesel generators."

In August 1983, all three diesels underwent a cylinder head stud replacement program, and one diesel generator completed the high load retest (15].

LILCO's summary continues, stating that "one remaining demonstration of diesel capability was scheduled prior to ftsel load; the integrated emergency core cooling system and emergency diesel generator operational demonstration."

LILCO re. ported (15] that as of the August 12, 1983 crankshaft fracture, the diesel generators had accumulated 2182 hours0.0253 days <br />0.606 hours <br />0.00361 weeks <br />8.30251e-4 months <br /> of operation as follows:

DG 101 - 646 hours0.00748 days <br />0.179 hours <br />0.00107 weeks <br />2.45803e-4 months <br /> DG 102 - 718 hours0.00831 days <br />0.199 hours <br />0.00119 weeks <br />2.73199e-4 months <br /> DG 103 -- 818 hours0.00947 days <br />0.227 hours <br />0.00135 weeks <br />3.11249e-4 months <br />.

In response to a request for information by the NRC regarding the total number of operating hours on each diesel generator and the total number of hours at 3900 kW or greater, LILCO responded [16] as follows:

Total Operating Hours for Each DG Unit at 2-hour Overload Rating

(> 3850 kW)

Total Operating Hours (These hours included in total on Each DG Unit operating hours)

DG At At At At Unit TDI Shoreham Total TDI Shoreham Total 101 128 518 646 3

16 19 102 30 688 718 3

19 22 103 40 778 818 3

20 23 3.2.5.2 Review of Conditions at the Time of Crankshaft Failure In response to a request for information by the NRC about the test procedures in use at the time of the crankshaft failure, LILCO responded with the following description of the test [17) :

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' Cylinder heads on DG 102 were replaced under R/RR R43-1001 with new design stress relieved heads. With all eight cylinders equipped with the new heads, the 102 DG was cun for 12 hours1.388889e-4 days <br />0.00333 hours <br />1.984127e-5 weeks <br />4.566e-6 months <br /> to allow hot tocquing of the exhaust header bolts and air start valve nuts. Following this run, a retest of the engine was begun under-8.7-R43-042.'

The specific scope of the retest under this 8.7 Form was to:

1.

Verify proper diesel generator start to synchronous speed and cated voltage in less than 10 seconds.

2.

Verify prop -c DG operation for four hours at the continous load rating.

3.

Verify proper DG operation for 2 hours2.314815e-5 days <br />5.555556e-4 hours <br />3.306878e-6 weeks <br />7.61e-7 months <br /> at the two hour overload rating.

Refee to the response to NRC Request for Information II.2, pages 10.5 through 10.17, for a copy of the retest procedure 8.7-R43-042, as completed up until the time of the failure of DG 102.*

On page 10.1 of Reference 18, LILCO provided the following detailed j

descetption of the events just prior to the failure:

"The dtesel generator prior to the performance of 8.7-R43/42 was in its

{

normal standby condition. An interim operating instruction was performed l

to ensure peopec breakee positions, proper valve lineup and coecect initial conditions. The diesel engine was started from its remote location, the main control room. Proper starting, acceleration to synchronous speed and eated voltage within 10 seconds was verified by the test eng sneer and the OQA inspector. Plant Operator syncheonized the 1

diesel generator to BUS 102 by closing ACB 102-8 and then proceeded to increase the diesel generator load to 3500KW in less than 60 seconds.

Once at the 3500KW300KVar load the operator was instructed to maintain ents lud foe four hours. He was instructed that any deviations, caused ny the MLCO ge td, away from 3500KW/300Kvacs should ce coerected.

Anocner plant operator was stationed in tne engine :com with verbal communt:attena estaolished between operators via headsets. Du ing tne course of the four hour full load cun, a LILCO technician was also stationed in the diesel engine room with the task of recording all pectinent test infacmation every 30 minutes. No aonormal esadings were ooserved by eitner Operator not was the data written down by the ecnnt:ian found to be out of its normal operating :ange as specified by tne engine manuf acturee for this size load.

Since this test was handleo similar to a Station Surveillance vrocedura no special test equipment was utilized for data recording. All data written down was taken of f of normal plant gauges either in the main conteol room or in the diesel engine room. The two exceptions were tne genecatoes hearing temperature and tha generator stator temperature, both j Frankhn Research Center m,

t TER C5506-426 material in the above calculation yields a minimum allowable crankpin diameter of d = 10.84 in.

This is less than the 11.00-in actual crankpin diameter. In summery, if full credit were taken for the actual crankshaf t material properties, the ll.00-in crankpin diameter of the diesel generators just meets the minimus ABS crankpin requirements as shown by calculations performel as a part of this review.

ABS Paragraph 34.17.4, Solid Crankshaft Web Dimensions In order to provide adequate bending stiffness in the web, ABS requires that the web dimensions satisfy the following inequality:

wt 2; 0.35 d where w = 21.0 inches, width of web t = 4.5 inches, web thickness d = 11.00 inches, crankpin diameter.

Thus, (21) ( 4. 5) 2;0.35(11.0) 425 3;42.4 and the inequality appears to be satisfied.

(The values for w and t were acquired informally by a telephone conversation with Dr. Wells, FaAA, on February 9,1984, and are assumed to be sufficiently accurate for this calculation.)

e Summary of ABS Rules Application Comparison of the TDI 13 x 11 crankshaf t design with the ABS rules indicated that the crankshaf t geometrical proportions were within the ABS rules, but the dynamic stresses in the crankpin were not.

Thus, the ABS rules are significantly more conservative with respect to harmonically induced stress than the DEMA rules, or about half the DEMA recommended limits. Again, this reflects the conservative design believed to be required for safety at sea, and probably is derived from the culmination of long-term experience in that industry.

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l TER-C5506-426 4.2.4 Summary of Crankshaft Desion Review The review of this section on crankshaft methodology and design may be summarized by the following statements:

o The DEMA rules (2] are not design specifications and Standards.

Supplementary specifications and standards are required.

It is advisable to employ the more comprehensive direct or modified o

direct solution of the mathematical model equations for torsional dynamics. With the present development of computer methods and accessibility of computer systems, the direct solution methods are not more labor intensive than the present computerized tabular methods and do provide more comprehensive design assistance.

o TDI used Tn values for torsional excitation that are very low compared to values recognized in the industry since at least-1942 (36].

o The TDI crankshaft (11 x 13) does not meet the DDtA or ABS rules for dynamic stress when the revised TDI values of Tn are employed.

4.3 REVIDi OF CRANKSHArf DYNAMIC TESTING Dynamic testing of the crankshaft is regarded in this review as the essential element of the failure investigation because it is only through carefully conducted measurements that the actual engine dynamics and local component stresses are confirmed. Accordingly, great attention was paid to each aspect of the test program.

Dynamic testing of DG 101 using an instrumented crankshaft was performed on September 20 and 28, 1983 at the Shoreham Nuclear Power Station. Reviews of preparations and procedures and an account of test observations were reported previously (ll.

j Instrumentation for the measurement and recording of vital dynamic data included that are shown in Section 3.2.4.1 Since the completion of testing, the recorded data were reduced and i

l reported (37] by Stone and Webster, and the implications for the crankshaft 4,.,

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TER-C5506-426 failure investigation were reviewed and reported (29] by FaAA. This section is primarily a review and evaluation of the reported test data [37] and the failure investigation conclusions [29] that were reached.

4.3.1 Instrumentation, Signal Conditioning, and Data Recording Reference 37 provides a description and statement of applicability of transducers employed, including those for strain, torque, torsional shaf t displacement, cylinder pressure, generator voltage and current, linear vibration, and the combination of crankshaf t position and rotational speed. A table listing their pertinent characteristics and applicable ranges is also shown. The instrumentation was evaluated and its installation observed by the reviewer at the time of dynamic testing.

For the most part, data output from the transducers was good. Earlier problems of strain gages and data transmitters on the rotating crankshaf t were largely corrected before completion of testing on September 28, 1983, although the reported data [37] do include noisy, but apparently functioning, Serain gage signals, e.g., on the No. 7 crankpin fillet. Also, the transducers for cylinder pressure seemed to function satisfactorily but appeared to provide pressure data lower in value than the actual pressure. The application of instrumentation in these environments is difficult and the experienced experimental test engineer anticipates certain aberrations in these data channels.

Indeed, the essence of the test engineer's work is to plan and conduct the test to maximize the good data extracted. Data from the strain gages on the crankshaft were telemetered to nonrotating receivers and were conditioned and recorded along with the other data oi. a 14-channel, FM mode tape recorder. With proper planning of signal channels prior to a test run, this af forded an opportunity to record simultaneous events on parallel channels. The signal conditioning and recording equipment are described in Reference 37.

The application of transducers, signal conditioning, and data recorders was reviewed and found to be satisfactory.

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'4 3 2 Calibration Procedures Measured values are not necessarily more accurate than analytical estimates; experimental measurements are only as accurate as the accuracy of their calibration, and then only if the proper instrument was chosen for the task.

4.3.2.1 Strain Gage and Torque Bridge Calibration Fillet strain gages and the torque bridge (employing strain gage) were L

calibrated by the shunt resistance method, wherein a precision resistor of known value is shunted in succession across the available arms of the bridge circuit.

Shunt resistance of the strain gages provides calibration not only of the strain gages, but also of the conditioning circuitry and recording equipment.

However, it calibrates the gage only for measurement of surf ace strain in ths metal on which the gage is located. This is sufficient calibration for the crankpin fillet gages which were for the measurement of surf ace strain.

. Calibration of the torque bridge, which used :..

in gages, required additional pre.cedures because the measured quantity was that of shaf t torque and not strain at a point. Consequently, the test engineers employed static torque tests and test operation of the engine at zero electrical output to confirm the calibration of the torque bridge.

The static torque test yielded measured torque plotted against applied mechanical torque as shown in Figures A-10 and A-ll of Reference 37.

Considerable hysteresis is noted in these figures due to L.he friction in the engine and possibly due, in part, to strain gages that are not fully exercised following their installation.

Industry experience has shown that the relationship would be much raore linear in actual operation, where the bearing surf aces would be operating on developed oil films to greatly reduce hysteresis due to friction in the engine, and the strain gages would become

• exercised" for greater linearity.

1 The zero-output tests of the instrumented engine are discussed in Section i

A.2.2 of Reference 37, which includes a table of values measured at four

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l TER-C5536-426 electrical loads. The normalized values of "kW/1000 lb-f t" showed a spread of l

+4% and -44 about an arithmetic mean value. Using linear regression, the mean ratio of the measured values of "kW/1000 lb-f t" was calculated by Stone and Webster to be 1.21.

Although Reference 37 explains this to be the stress concentration in the shaf t on which the strain gages are mounted, evaluation during this review indicated that the actual stress concentration is on the order of 1.16 and that the balance of the factor is due to the experimental measurement spread of the "kW/1000 lb-f t" values previously discussed.

Shifts in zero reference of the data recordings were investigated as a part of the data analysis as discussed in Section A.3 of Reference 37.

The overall error due to static strain ranged from 1.0 to 4.21.

Thus, the static offset does affect the calculation of principal stresses by a small percentage because these are based upon both the static and instantaneous cyclic stress.

It should be noted, however, that the stress range of the cyclic stresses is not affected by this offset.

4.3.2.2 Calibration of the Torsional Vibration Displacement Transducer The torsional vibration transducer is the unit attached to the gear case end of the crankshaf t for the direct measurement of vibrational amplitude.

Sections 3.2.2, 4.3, 6.3, and A.4 of Reference 37 describe th2 application and calibration of this unit, wherein calibration is performed easily by means of fixed limits on displacement built into the unit.

A problem that arose with the use of the transducer was corrected during data reduction. As described in Reference 37, an internal L..cer selection switch mained set to a 10-Hz cutoff frequency. This attenuated all signal ~

components above 10 Hz.

Data reduction procedures were developed to amplify the attenuated signal components in an effort to correct the error. The procedure was reviewed, and the results of the error-correcting efforts shown in Reference 37 were evaluated and found to be satisfactory.

l 4.3.2.3 Calibeation of Accelecometers Sections 3.2.5 and 4.5 of Reference 37 cover the application and calibra-tion of accelecometers for linear vibration measurement. The accelerometers

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were calibrated with the use of a BsK Model 4291 calibrator, which could serve as a transfer standard from the National Bureau of Standards.

This revie*> showed that any use of this transfer standard capability was not stated in Reference 37 and that the calibration source was not known.

Although these data were not necessary in forming a conclusion regarding the i

cause of failure, calibration of the accelerometers and other instrumentation should nevertheless be traceable to the National Bureau of Standards.

4.3.2.4 Calibration of Cylinder Pressure Instrumentation Sections 3.1.4, 3.2.3, 4.4, and 7.3.5 of Reference 37 describe the measurement of cylinder pressure and its calibration. Time history pressure measurement was attempted by means of precalibrated piezoelectric transducers installed in the compression test cocks of engine cylinders 5 and 7.

Calibra-tion of the data signal circuitry between the transducer and the tape recorder was performed using the B&K Model 4291 calibrator mentioned previously.

The cylinder pressure measurements were unacceptably low. Efforts by Stone and Webster and FaAA following these tests c9ncluded that the gas flow path geometry (see Figure A.4, Reference 37) was responsible. Accurate cylinder pressure measurement was not necessary in this test for conclusions regarding the cause of failure.

4.3.3 Review of the Experimental Data Dynamic tests of engine operation were run at zero-output load (variable speed tests) and at loads of 100 kW, 1695 kW, 1706 kW, 1750 kW, -2250 kW, 2550 kW, and 3500 4W, with constant speed (450 rpm) operation. Data foe these tests were reduced by Stone and Webster and are presented as charts in References 29 and 37.

The test data as presented (24, 37) are dominated by presentations of torque and crankpin fillet strain. Torque, as presented in Figure 4-21 of Reference 29, is characterized by a 30-H2 oscillation of varying amplitude superimposed upon a steady-state value. Torque oscillatory amplitudes for

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TER-C5506-426 3500-kW operation reach a value of 1175,000 ft-lb (350,000 ft-lb, peak-to-peak torque range) superimposed upon a steady torque of 57,000 ft-lb.

Note that this cyclic torque is a little over 3 times (6 times for peak-to-peak range) the steady torque required to produce an electrical output of 3500 kW from the generator. This single amplitude ratio of 3 stands in contrast to the ABS rules [3] where the single amplitude dynamic component is expected to bw on the order of the value of the steady-state component (power transmitted).

This is explained as follows. Refer to Section 4.2.3.4 of this report and note that the allowable crankpin single-order torsional stress, using the example of ABS Grade 4 steel, is 12679 pai. For the 100% load rating of the diesel generators (3500-kW output), the tagine torque at the flywheel shaf t (torque bridge location) s 57,040 f t-lb, which yields a crankpin torsional

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shear stress of 2619 psi. This is very close to the limiting torsional stress a

level allowed by the ABS rules. This example was calculated for the 1004 load rating of 3500 kW, the maximum load in the torsional dynamic tests performed.

For the intermittent 3900-kW diesel generator loads projected for actual service, the steady state, and cyclic stresses would be proportionally higher.

I The engine firing rate is 30 Hz.

This engine firing rate is sufficiently close to the first mode torsional natural frequency of 35.5 Hz to produce the large dynamic response in the absence of significant damping.

Measured fillet strains on Crank No. 5 varied to a maximum peak-to-peak

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range of 1800 microstrain (1800 x 10 inch /inen) as reported for stain gage 5-1 in rigure 4-21 of Reference 29.

Table 6-2 of Reference 29 eports the major principal stress component of the measured strains to be 57,300 psi at Crank No. 5, corresponding to a measured total peak positive torque of 230,000 ft-lb (cyclic and steady-state) and negative torque of -153,000 ft-lb.

4 In the abaence of direct access to the data and data reduction instrumen-tation, observation of the diesel generator tests plus analytic investigations of the data reported in References 29 and 37 performed during this review provide basic arrangement with the range and characteristics of torque and crankpin fillet stress reported by References 29 and 37.

Note that these are measured values subjected to the measurement errors discussed previously.

However, it does appear that these values are accurate to within 110%.

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TER-C5506-426 In addition to indicating high cyclic torques and stresses in the crankshatt, the test program yielded the following observations, with which this review concurs:

o The rotor-stator electrical coupling within the generator which acted to couple the electrical load inertia to the engine dynamic model produced varying generator output current at 3.75 Hz when connected to the electric power grid, but did not contribute to the failure of the crankshafts.

o Operation at 0.8, 0.9, and 1.0 power factors at the 2500-kW load range indicated that operation in this range of power factor did not contribute to the crankshaf t failure or dynamics of the system.

o The 30-Hs major dynamic response of the engine is not compounded by any oc c M L-*aeaction with the electric loads, electric power grid, oc plant loads.

o The sudden initiation of plant loads was observed to cause a smooth-orderly response of the engine and generator and was not seen to cause cyclic fluctuations.

o connection of the diesel generator to the electric power grid was observed to be smooth and without significant. transients, although it is realized that considerable care was taken at the time to make a proper connection. Connection of gee.; tors to the electric power grid without adequate synchronizatior..:an be damaging.

4.4 REVIEW OF FaAA DYNAMIC MODEL AND CRANKSHAFT STRESS ANALYSIS 4.4.1 Dynamic Response Model In the course of the failure investigation, FaAA prepared and used a digi *al computer dyna.nic response model. From a discussion,* it was learned that the model is generally of the raode-superposition type discussed in Section 4.2.2.1 of this report Reference 29 indicates that the model used ene same basic lumped-parameter (inertias and spring constants) model as formulated by TDI (Appendix A) with the addition of the rotor-stator equivalent spring constant and the electrical load inertia (see Figure 2 also).

i

• Discussion with Dr.

P. Johnston, FaAA, during test of DG 101 on January 7, 1984, at the Shoreham Nuclear Power Station.

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l TER-C5506-426 FaAA's computer model output, as indicated by Figures 5-3 through 5-6 of Reference 29, has a remarkable similarity in character and amplitude to the values measured by the engine test. FaAA did not initially provide, in its report [29), the list of Tn values employed in its mathematical model. When it was suggested that the Tn values would be valuable for comparison to TDI's design values and to those from other published sources, FaAA made them available.* Table 8 includes FaAA's values with accepted values from Lloyds Register and Ker Wilson which were included here from Table 5 to facilitate comparison. Comparison with values employed by TDI was made using the TDI values of Table 6.

In these comparisons, it was observed that the FaAA values compared favorably with those of Lloyds Register and Ker Wilson. The FaAA values were more chan twice TDI's design values (TDI 1974-1975 list in Table 6) in the critical range of orders 4.0, 4.5, and 5.0, and even greater for other orders. Thus, the Tn values for FaAA's mathematical model for which FaAA reported [29) excellent agreement of computed dynamic response with that experimentally measured further confirms the validity of published Tn values over that employed by TDI for design.

Even if FaAA's excitation had been prepared only to achieve the same dynamic responde amplitudes as measured in the engine tests, the model would have provided a highly useful interpolation function in portraying the dynamic action at points not available for measurement.

As discussed in Section 4.2.2.1, computer models following f rom the direct solution of the dynamic equations are very powerful in describing the full dynamics and interactions of a system. FaAA's computse model confirms this. The first task for the model was the prediction of the available cyclic life of DG 101 remaining throughout the course of diesel qenerator testing on i

September 20 and 28, 1983. Here, initial dynamic response data measured at the beginning of each test session were introduced to the computer model for comparison and prediction of the available life cycles remaining.

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• Telephone call from Dr. P. Jonnston, FaAA, March 9, 1984.

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TER-C5506-426 Table 8.

Comparison of FaAA's Tn Values with Those of Lloyds and Ker Wilson Vibration Lloyds Order FaAA*

Register **

Ker Wilson ***

0.5 74.5 80.0 77.0 1.0 86.0 88.0 79.0 1.5 75.1 83.0 75.0 2.0 75.6 69.0 66.0 2.5 54.0 57.5 55.0 3.0 12.3 47.5 43.0 3.5 37.7 38.5 32.0 4.0 28.7 30.5 25.0 4.5 24.7 23.6 19.0 5.0 20.7 18.0 15.0 5.5 16.9 13.8 11.0 6.0 13.8 10.5 8.9 6.5 11.2 8.5 7.3 7.0 9.4 6.8 6.0

• Calculated independently by FaAA.

Includes effects of reciprocating masses.

• • From Table 5.

Not known what ef fects, such a's reciprocating masses, are included.

• • • From Table 5.

Values for cylinder pressure only.

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TER-C5506-426 4.4.2 Crankshaft Stress Analysis 4.4.2.1 Finite-Element Model FaAA formulated a finite-element structural mathematical model using three-dimensional, eight-node, isoperametric elements to represent one throw of the crankshaft. With application of torques from the dynamic response analysis, the model had the capability to indicate the highly stressed points in the complex crankshaf t geometry. Unless extremely fine element grids are employed, finite-element models generally underestimate the stress concentra-tion at local regions. Accordingly, FaAA used the same element distributions in an axisymmetric model of the same diameter and fillet radius so that the lack of stress concentration definition could be assessed by comparison to well-established values (381 The ratio of the established value and the finite-element stress concentration factor was used as a multiplier for the final stresses predicted in the fillet region by the crankshaf t throw finite-i element model. This was reviewed and found acceptable. The alternative method of using many more elements in the fillet would have been much more costly in both modeling and computer run time.

FaAA did not include a description of its method of torsional load application in its report (29). However, when it was shown that the method of i

torsional load application employed by FaAA in the finite-element model was needed to complete the review of FaAA's crankshaf t analysis, Dr. Wells (FaAA) provided a verbal description of the torsional loading method during a document review at the Shoreham Nuclear Power Station on March 8, 1984. The loading method was said to consist of a unit angular displacement applied to the journal end of, the crank-throw finite-element t odel, plus a lateral displacement constraint applied to the side of the journal to represent the lateral constraint provided by the journal bearing. The axial location of the lateral constraint representing the journal-bearing reaction was said to have been varied to study its ef fect upon the computed stresses in the crankpin i

tillet. This ef fect was said to be relatively small. During the review of the crank-throw finite-element analysis and method of loading, it was noted that the unit angular displacement method of torsional load application along with the lateral displacement constraint to induce the journal-bearing

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4.4.2.2

, Bending Stresses-Or. pages 6-8 and 6-11 of Reference 29, FaAA discussed an investigation of bending stresses in the finite-element model due to an effective pisto i load at top-dead-center. When the associated bending stresses were indicated to be on the order of 4500 psi, as compared to approximately 40,000 psi for the torque load, the contribution of the connecting rod load in consideration of the fillet stresses was considered to be negligible, especially when the maximum fillet stresses occured when the crank was 130 degrees or so after top-dead-center.

j Bending stresses, however, did appear to play a part in the stressing of the fillet as indicated by Figure B-100 in Reference 29.

This bending action i

h9 wever, appeared to be local bending in the web and crankpin as part of the gross torsional loading. Consequently, it became a part of the stress concen-trating mechanism that caused the highly stressed region to develop at an angle of approximately 130 degrees from the 12 o'cicck position on the crankpin.

4.4.2.3 Crankshaft Stress Analysis Summary The usefulness of a comprehensive stress model is readily apparent. The stresses predicted by the finite-element model appear to be in good agreement with experimentally measured values, even acknowledging the fact that the experimentally measured values contain an inherent erroe band of up to about

+104.

Although the use of finite-element models for theoretical analysis, as weil as for extending experimental investigations to regions not measuraole, is to be strongly encouraged, the validity of the f ailure investigation was considered during this review to be most celevant in the experimental measurement of crankshaf t fillet stresses in actual engine operation.

Analytic techniques, such as the dynamic medel and the finite-element crank-throw model, while quite powerful, were looked upon in this review as supplemental and confirming investigations.

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i TER-C53r)6-476 4.5 AEVIEW OF REPLACEMENT CRANKSHAFT DESIGN Following failure of the crankshaft of the Shoreham diesel generators, the engine ma. ufacturer, TDI recommended the use of an impeoved crankshaf t design, designated the 13 x 12 crankshaf t.

Whereas both the failed crannshaft (13 x 11) and the recommended replacements had 13-inch main journal diameters, 3

the replacement crankshaf t featured an increase in the crankpin diameter from 11 to 12 inches, as well as an increase in the crankpin fillet radius from I

one-half to three-quarters of an inch. Analyses of the replacement crankshaf t by FaAA [39) and TDI [40] are reviewed in this section of the report.

4.5.1 Review of Analysis by Transamerica Delaval, Inc (TDI)

TDI used the same method of analysis as shown in Appendix A for the analysis of the original 13 x 11 crankshaf ts, with the exception that they substituted the Tn values shcwn in Group 4 of Table 6 of this report. Here the Tn value for the 4th order is 27.62 as, compared to the previous value of 13.30.

4 In summary, although the critical 4th order Tn excitation value was doubled, the following considerations produced a reduction in the calculated stress for comparison to the DENA-recommended valuest o The larger crankpin permitted a 22% reduction in crankpin nominal torsional stress.

o The increased natural frequency from 35.5 to 38.7 Hz reduced the dynamic magnitier foe a 30-Ha excitation from 3.51 to 2.51.

This yielded a 4th order stress of 2990 psi as calculated by TDI for cornparison to the DEMA recommendation of <5000 psi.

l 4.5.2 Review of FaAA Dynamic Resoonse Analysis and Crankshaft Stress Analysis 4.5.2.1 Response Analysis FaAA employed its computer dynamic model using mode superposition to analyze the dynamic response with all significant modes considered.

Inertia and spring constant elements for the model are shown in Table 3-1 of Reference 39, and the resulting natural frequencies are shown in Table 3-4 of that sa:ne da 59-

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. - - _ _. l TER-C5506-425 reference. The dynamic response was computed by FaAA for " full load".

The use of the term " full load" does not carry full definition since the design rating of the diesel generator is 3500 kW, but it is expected to operate at 3900 kW for short periods. For this review, 3500-kW generator output is inferred to be " full load".

Comparison of TDI and FaAA dynamic stress values to the DEMA recommenda-tions follows:

Average Torsional Average Torsional Stress (psi) Due Stress (psi) Due to Method of Analysis to 4th Order Summation of Orders TDI Analysis 2990 FaAA Modal Superposition 3300 5640 DEMA Reconumendation

<5000

<7000 comparison of these stresses to those updated stresses for the 13 x 11 crankshaft, as shown in Table 7 of this report, indicates reductions in stress by a factor of 1.79.

Comparison of these stresses to the ABS rules, similar to that shown in Section 4.2.3.4 of this report, indicates that the ABS rules may or may not be satisfied depending upon the interpretation that would be approved by the ABS following its review.

Assuming that an ABS Grade 4 steel *was used for the crankshaft, the ABS allowable stress for a single harmonic is 2680 psi (see Section 4.2.3.4),

whereas the calculated stress (TDI) is 2990 psi. Thus, TDI's stress of 2990 psi and FaAA's stress of 3300 psi were both in excess of the ABS allowable stress for a single harmonic using a nominal ABS Grade 4 material.

The actual mechanical properties of the replacement crankshaf t material, however, were shown by the quality control documents at the Shoreham plant to be those provided in Table 9.

Whereas Appendix B shows an ABS Grade 4 material to have an ultimate tensile of 83,000 psi, the minimum ultimate tensile i

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TER-C5506-426 Table 9.

Properties of Replacement 13 x 12 Crankshafts Mechanical Prooerties Yield Ultimate Production Crankshaft Point Stress F.longation Area Brinnel Sample Number M

(pei)

(%)

(%)

Hardness Location 693 (DG 103) 58,310 100,360 25.0 54.1 205 59,470 106,460 24.0 58.9 212 694 (DC 102) 57,290 101,820 25.0 50.9 210 58,310 106,460 25.0 48.7 215 695 (OG 101) 52,650 100,800 24.0 50.9 205 Top 48,590 100,800 23.0 49.8 210 Bottom chemical Analysis Crankshaft C

Si Mn P

S Cr Al Number Heat

(%)

(%)

(%)

(1)

(%)

(%)

(%)

693 (DG 103) 821-487 0.50 0.05 0.70 0.006 0.010 0.63 0.003 694 (DG 102) 821-487 0.50 0.05 0.70 0.006 0.010 0.63 0.003 695 (DG-101) 811-167 0.46 0.12 0.65 0.010 0.008 0.69

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TER-C5506-425 l

L strength of the replacement crankshaf t materials as shown in Table 9 is 100,360 psi. To take full advantage of this material, an allowable value of 3090 psi for a single harmonic could be presented to Aas for approval in accordance with Note 4 of Ass Table 34.3 in Appendix 3.

If full advantage of the material is to be taken, then it is also appropriate to use the full calculated dynamic response due to a single harmonic exciting factor. TDI's stress of 2990 psi was calculated using only the first mode response. Although TDI's analysis does show a small response for the second and third modes of torsional vibration, the second and third modes are seen to add very little to the first mode stress of 2990 psi. Thus, should the interpretation of the ABS rules discussed above be accepted by ABS, TDI's single harmonic stress would be within the ABS limits. However, FaAA's calculated stres of 3300 psi for a single harmonic excitation, based upon a somewhat higher value of Tn and upon greater modal participation, would not.

A8S also requires that the total vibratory stress from all harmonic excitation not exceed 150% of the allowable stec:: fer a single harnenic exciting factor. For a nominal ABS Grade 4 material, this allowable stress is 4020 psi.

For the interpretation of the ABS cules to use the full properties, the allowabie stress is 4640 psi. TDI's total stresses cannot be compared to these ABS allowables because their analysis methods do not facilitate such summation of stresses. FaAA's calculated torsional stress for the summation of excitation orders is 5640 psi, which is well beyond even the interpreted ABS allowabie stresses.

4.5.2.2 Crankshaft Stress Analysis FaAA used the finite-element method of analysis reviewed in Section 4.4.2 of this report to compute the stress magnitude and distribution for the replacement crankshaft.

Stresses were reported to be reduced from the previous cyclic principal stress range of 60,000 psi to a range of 37,000 psi. This constitutes a reduction by a factor of 1.78 to a cyclic range that is only 56% of the former cyclic range. The reduction was due to the larger crankpin and increased

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TER-C5506-426 stif fness with resulting increased natural frequencies as previously discussed, and was supplemented by the increase in the crankpin fillet radius from one-half to three-quarters of an inch. The analysis was considered to be acceptable.

4.5.3 Crankshaft Shotpeening FaAA reported [39] that shotpeening was introduced to the crankshaft processing in an effort to assure a " consistent, high level of compressive residual stress in the surf ace and to eliminate machining marks." The report continued by' stating that the fillets "will be inspected by a high-resolution, eddy-current method after the break-in run."

Shotpeening has a long history of use in closing microscopic surface cracks and establishing a surface layer of the material in compressive stress.

Although the basic idea is good, it was noted during the review that while various levels of'shotpeening are available, no description of the process was provided.

]

Accordingly, the NRC arranged for a document review at the Shoreham Nuclear Power Station on March 8, 1984, during which quality control documents pertaining to crankshaf t shotpeening were reviewed, and an informal discussion was held with Dr. Wells of FaAA.

It was learned from Dr. Wells that two of the three replacement crankshafts, Nos. 693 (DG 103) and 694 (DG 102), arrived f rom TDI with the crankpin fillets already shotpeened.

The crankshafts were inspected and the results of the inspection are described by Stone and Webster Engineering Corporation's Coordination Report No. F-46109-G (41] as follows:

" Problem

Description:

Delaval has identified ' holidays',or lack of peen coverage in the fillet areas of new diesel crankshaf ts purchased in accordance with E&DCR F-46109-C.

These ' holidays' have been dispositioned as functionally acceptable by TDI, however, recent analysis performed by Failure Analysis Associates indicates that 100% peening coverage i s beneficial. "

In conjunction with the review of documents on March 8,1984, photographs of the original shotpeening supplied by TDI were reviewed. Although the

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TER-C5506-425 photographs did not provide the desired detail, the photographs gave an impression of surface texture more like grit blasting than shotpoening, i.e.,

the surface appeared to have been gauged by sharp particles instead of dimpled by round, smooth particles. Although the photographs provided only a limited view of the fillet surfaces, this evaluation of the initial shotpoening concurs with the results of the inspection (41] by Stone and Webster Engineering Corporation.

Stone and Webster's Coordination Report No. F-44109-G [41] provided a recommended solution as follows:

" Problem Solution: Since the crankshaf ts are delivered to the site, Metal Improvement Company, a local firm with espertise in shotpoening will perform the rework. The fillet areas shall be repeened in accordance with the requirements of MIL-S-13165B to assure 100% coverage of the fillet areas. Peening shall be performed by Metal Improvement Company on site and the crankshaf ts inspected by CQA for 100% peening at the fillet areas."

Accordingly, LILCO Repair / Rework Request R/RR R43-1632 specified shotpeening to include the following parametees:

o Shot sizes MI-550 o Intensity, 0.006-0.010, Almen "C" test strips o MIL-S-131658, Amendment 2.

Quality control documents were reviewed and indicated that the Almen test strips for the repeening provided readings within the specified intensity of 0.008 to 0.010 inch (are height) with the exception of one test strip which was measured at 0.011 inch.

Photographs of the,repeened surface were reviewed and show an improvement in surf ace texture, indicating an improvement in the quality of shotpeening of the crankpin fillets.

Crankshaf t No. 695 for DG 101 was received at the Shoreham plant direct f rom the supplier, Krupp-Stahl in Germany, without shotpeening. Crankshaft No. 695 was shotpeened at the Shoreham plant to the same specifications as those described for crankshaft No. 693 and No. 694 above. Records reviewed at the Shoreham plant showed that the Almen test strips for crankshaf t No. 695 i

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i-TER-C5506-426 nhotpeening indicated that the intensities remained within the specified range of 0.008 to 0.010 inch are height.

Shotpeening of'this intensity is shown by Figure 4 to induce a compressive stress to a depth of from 0.027 to 0.034 inch, with the induced stress distributed as shown in Figure 5.

Figures 4 and 5 are taken from Reference 42.

The purpose of shotpoening is to induce a compressive stress in the saterial at the surface of the crankpin fillets. Since the smooth surface is being disturbed by the particle impacts, it is necessary, once shotpoening is begun, to assure that the shotpoening coverage is uniform and of an intensity, with the right size of smooth shot, to achieve a suitable depth of material in compressive stress. Otherwise, improper shotpeening could serve as a source of added stress concentrations.to make the crankshaft more susceptible to fatigue.

The actual peened surface were not available for inspection in the course of this reviews therefore, this evaluation was made using the specified parameters, recorded Almen test strip measurements, and photographs of the peened surface. The shotpoening performed at the Shoreham plant is acceptable for the new cranashaft (No. 695) not subjected to shotpoening in advance and will serve to increase the fatigue life of the crankpin fillets.

Inspection of crankshaf t Nos. 693 and 694 revealed inadegaate initial shotpeenings for these crankshafts, the rework shotpeening discussed above would be sufficient to counter the undesirable effects of the previous shotpeening, provided that the shotpeened surfaces that were photographed and made available for this review were representative of all crankpin fillet shotpoening. With this provision, the rework shotpoening is acceptable.

As an alternative to shotpoening, a surface layer under compressive stress can be induced into crankpin fillets by colling techniques. This is accomplished by pressing a rolling element against the fillet surface with suf ficient force to produce stresses in the fillet surface material that are just.beyond the yield point. With the proper design of rolling element, the distribution of induced compressive stresses can be controlled to an ideal

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Distribution of Stress in Shotpeened Beam with No External read (from Reference 42]

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TER-C5506-426 profile of magnitude and depth in addition to providing a smooth fillet

. surface for optimum fatigue resistance. Fillet rolling provides many advantages; however, there are fillet geometeries for which it is difficult to design a coller, e.g.,

recessed fillets similar to those of the TDI crankshafts.

In addition, the technique requires the proper machinery to 1

hold, load, and rotate the crankshaft and roller. Where this technique is possible, benefits follow. Lacking the means, shotpoening is recommended.

4.5.4 Summarv of Reolacement Crankshaft Desion The increase in crankpin diameter from 11.00 to 12.00 inches provided a significant crankpin stress reduction by reducing the direct torsional stress in the crankpin due to larger cross section and by stiffening the shaft to produce a higher natural frequency and thus reduce the dynamic multiplication factor.

Stresses calculated by 7DI and FaAA were within the DEMA [2] recommenda-tions for a single harmonic excitation. FaAA's summation of stresses for all excitation orders was also within DEMA's recommended values. TDI's analysis did not permit comparison of total stresses with those recommended by DEMA.

TDI's crankpin stress for single harmonic excitation does not satisfy the ABS limiting values [3] for ABS Grade 4 steel, except through an interpreta-tion of the rules in which full advantage of the crankshaft material properties is taken. Such interpretation would require study and approval by ABS.

TDI's analysis did not permit the comparison of total stress due to summation of orders with the ABS allowable values. Crankpin torsional stresses calculated by FaAA for both single harmonic excitation and summation of orders were in excess of ABS allowable values, including the higher allowable values' determined by an laterpretation of the ABS rules that used the full material properties of the crankshaf t material.

Crankpin fillet shotpeening of the replacement crankshafts was evaluated through the review of documentation and photographs at the Shoreham plant.

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TER-C5506-426 Crankshafts No. 693 (DG 103) and No. 694 (DG 102) were found to have been previously shotpeoned by TDI. When inspection at the Shoreham plant indicated that the initial shotpeening was unsatisfactory, the crankpin fillets were repeened at the Shoreham plant. Crankshaf t No. 695 (DG 1011, received direct

]

from the supplier in Germany, was not initially shotpoened by TDI and was shotpeened only at the Shoreham plant. The crankshafts could not be inspected directly, and the shotpoening was evaluated only through the review of documentation and inspection of photographs of local regions. The shotpeening and rework shotpoening perfor: sed at the Shorehas plant were found to be acceptable insofar as the photographs inspected are representative of all shotpcened surfaces.

It must be noted that all of the TDI and FaAA stresses reviewed herein pertain to the 3500-kW electrical output loading (100% design load) and not to the short-tecs 3900-kW load required by the Shoreham plant.

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TER-C5506-426 S.

CONCLUSIONS Based on the findings of the failure investigation reviewed herein, it is concluded that:

o The crankshaf t of diesel generator 102 failed in high cycle fatigue.

o Sufficient cause for the high cycle fatigue failure was crankshaft design based upon exceptionally low values of cyclic torque excitation (Tn) coupled with a natural frequency fairly close to the dominant excitation frequency.

o The specified design standards were not definitive and contributed to the failure by not providing design review material by which the design would have been evaluated and found to be in question price to the diesel generator's application as safe shutdown equipment.

With respect to the replacement crankshaf t design, it is concluded that:

o The combined static and dynamic effects of a 1.00-inch increase in crankpin diameter from 11.00 to 12.00 inches serve to reduce the crankshaf t stresses calculated by TDI and FaAA to within the DEMA recommended values for single order excitation and for summation of order excitation.

o Although stresses from TDI's analysis for the replacement crankshaft do not satisfy the ABS rules for a single harmonic using a nominal Grade 4 material, they would just meet an interpretation of the ABS rules for a single harmonic wherein the actual properties of the crankshaft material are used. Bowever, such interpretation of the ABS rules is subject to review and approval by ABS.

{

TDI did not present an analysis by which their summation of stresses o

from all orders can be compared to the ABS limiting value for that j

condition.

o FaAA's crankshaf t analysis predicts higher dynamic stresses due to (1) the use of slightly larger amplitudes of excitation (Tn values) than those used by TDI and (2) the superposition of modes resulting from the direct solution of the equations of crankshaf t dynamics.

Vioratory stresses computed by FaAA do not satisfy the ABS requirements for a single vibratory order or for the summation of orders, even considering an interpretation of the ABS rules to fully use the mechanical properties of the crankshaf t steel.

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All analysis of stresses performed by TDI and FaAA pertained to the o

3500-kW full load condition and not to the 3900-kW short-term overload required by the Shoreham plant.

Crankpin shotpoening of one crankshaft and rework shotpoening of two o

crankshafts performed at the Shoreham plant were found to be acceptable only insofsr as the evaluation from documents and photographs of localised shotpoened areas is representative of all crankpin fillet areas.

From the broad evaluations performed in the course of this review, it is summarily concluded that a set of standards more definitive than DEMA's

" Standard Practices for Low and Medium Speed Statiocacy Diesel and Gas Engines" is required for diesel engines essential for safe shutdown of the Shoreham plant; that "Itules for Building and Classing Steel Vessels" by the American Bureau of Shipping is representatire of definit.ive standards for safety at sea; and that, with the possible exception of TDI's stress for a single harmonic, the stresses evaluated in this review do not meet the requirements of the ABS standard.

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DESIGN REYlEW OF TDI R-4 AND RY-4 SERIES EMERGENCY DIESEL GENERATOR CYLINDER BLOCKS AND LINERS This report is final, pending confirmatory reviews required by FaAA's QA operating procedures.

Prepared by Failure Analysis Associates Palo Alto, California Prepared for TDI Diesel Generator Owners Group June 1984

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TABLE 1-1 ENGINE 101 LOAD HISTORY SHOREHAM NUCLEAR POWER STATION Hours at Load, L (1)

Total Event

Hours, and Date L(75 75cL<100 L=100 100<L(110 L>110 All Loads j

Original Crankshaf t 19.0 634 Hours 164.0 262.5 188.5 Crankshaft replaced Restart 12/29/83 Testing Hours 78.0 179.0 20.0 91.0 4.5 372.5 Outace 3/18/84 51ocx Inspection 3/20/84 Qual. Tes-: g Hours 43.0 10.0 29.5 5

2.0

?!

4/10/84 Total 285.0 451.5 238.0 91.5 25.5 10);.5 1

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TABLE 1-2 ENGINE 102 LOAD HISTORY SHOREHAM NUCLEAR POWER STATION Hours at Load, L (1)

Event Total and Hours.

Date L(75 75cL(100 L=100 100(L(110 L>110 All Loads Original Crankshaft Hours 83.0 325.0 259.0 22.0 699 Crankshaft Reclaced Restar: 12-22-83 324 Testing Hours 34.5 183.0 36.5 70.0 Outage 2/09/84 Bloca inspection 2/10/84 110 Qual. Testing Hours 90.0

3. 5 16.0 0.5 81c:t Inspection. 3/08/84

'1*I Total Hours 207.5~

511.5 311.5 92.5 i

1-9 l

1

TABLE 1-3 ENGINE 103 LOAD HISTORY SHOREHAM NUCLEAR POWER STATION Hours at Load, L (5)

Total Event Hou rs,

and Date L(75 7kL(100 L=100 100<L(110 L>110 All Loads Original Crankshaft 23.0 815 Hours 103.0 432.0 257.0 Crankshaft Replaced Restart 12/17/53 Testing Hours 67.0 170.5 69.0 34.5 6.0 347 Outage 3/11/94 Blocx Inspection 3/11/84 Qual. Testing Hours 64.5 5.5 24.5 13.0 1.0 108.5 31ock Failure 4/14/84 3!o:n Inspection 4/15/84 Total Hours 234.5 608.0 350.5 47.5 30.0 1270.5 1-10 i

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BEFORE THE ATOMIC SAFET'l AND LICENSING BOARD 2

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In the matter of

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LONG ISLAND LIGHTING COMPANY

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(Shoreham Nuclear Power Station, )

DOCKET NO. 5 0-3 2 2-OL Unit 1)

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3 10 DEPOSITION OF GERALD EDGAR TRUSSELL g

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,Ql May 7, 1984 12 3 '

VOLUME I - Morning Session g

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14 15 16

2 17 18 REPORTED BY:

l MARION G. KOLB, CSR NO. 4381

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i 24 2 5 TOOKER f ANTZ

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26 CERTIFIED SHORTHAND REPORTERS 9r 681 MARKET STREET, SUITE 925 27 SAN FRANCISCO, CALIFORNIA 94105 A

415/392-065J

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FREDERIC R. TOOKER KEMBLE ANTZ

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l 45 1

A.

Yes.

And what is that name?

2 Q.

i 3

A.

Shot peening.

Did Delaval recommend that the replacement 4

Q.

[

crank shaft be shot peened?

5 6

A.

No.

Did Delaval recommend that the replacement 7

Q.

crank shaft not be shot peened?

8 9

A.

I don't recall.

Who was responsible in your organization for 10 Q.

supplying the replacement crank shaf t to LILCo?

11 Supply -- can you give me the question one more 12 A.

13 time?

Who was responsible in your organization for 14 Q.

the' supplying of the replacement crank shaft for LILC07 15 I am asking who.

And it may be more than one person.

16 As to the supplies, the parts manager.

17 A.

I Who was responsible for giving the 18 Q.

Yes.

recommendation as to whether or not the replacement crank 19 shaft should be shot peened?

20 21 A.

I was.

And do you now recall what your recommendation 22 Q.

,as in that regard?

23 w

recommended against I

My recollection is that I

24 A.

i 1 25 shot peening.

Why did you recommend against shot peening?

t j 26 Q.

The detailed drawing for that part did not call A.

, 27 29 for snot peening.

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Man Francisco 94105 415/392-0650

3;l I.

1 Q.

Who prepared the detailed drawing for that part?

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2 A.

I don't know.

3 Q.

Was it Delaval who supplied the detailed l'

4 drawing for that part?

l, 5

A.

Yes.

6 Q.

Was there any discussion within the Delaval 7

organization concerning whether or not the detailed 8

drawing should or should not provide for shot peening of 9

the replacement crank shaft?

10 A.

Yes.

11 Q.

What was the basis for the conclusion that it 12 should not contain the requirements for shot peening?

13 A.

The basis for that conclusion lay in an opinion 14 that mechanical improvement by shot peening did not 15 substantially improve the f atigue strength of the crank i

16

.thaft.

17 Q.

Did it improve the strength of the crank shaf t 18 at all?

l 19 A.

Yes.

20 Q.

Are there disadvantages to shot peening the al crank shaft?

22 MR. SMITH:

You are talking about the specific i

23 shaft in question here, I assume?

24 MR. DYNNER:

Yes, right.

25 THE WITNESS:

No.

26 MR. DYNNER:

Q.

So as I understand your 17 testimony -- please correct me if I'm wrong -- there are ti no disadvantages to the shot poening in this crank shcft, 700KER & ANTZ 681 Market Street San Francisco 94105 415/392-0650

- ~ ' ' - -

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47 t

1 there was an advantage in that it somewhat increased the 2

strength of the crank shaft, and yet ye : recommended j

3 against shot peening; is that correct?

4 A.

That's correct.

I S

Q.

On what was that recommendation based?

i 3-6 MS. TARLETZ:

Asked and answered.

7 MR. SMITH:

I will join in that objection.

ij 8

MR. DYNNER:

Q.

Aside from the fact that the e i 9

detailed drawings did not call for the shot peening.

'j 10 MR. SMITH:

The question has been asked and 11 answered.

12 MR. DYNNER:

I don't think so.

13 THE WITNESS:

What is the question?

14

" MR. DYNNER:

Q.

The question is On what was 15 your recommendation against shot peening based aside from 16 your prior testimony that -- when I asked the question 17 previously -- that it was based upon the f act that the 18 design drawings did not call for shot peening?

19 MR. SMITH:

Well, note my objection to the form 20.

because I don't think that was -- I think the record will 21 show that that was not the only basis against the 22 recommendation that the witness has already testified to.

23 THE WITNESS:

The recommendation against shot 24 peening was based in part on, A, the experience that shot in the 23 peening did not provide a substantial improvement 26 fatigue strength of the shaft, and in part on a 27 discussion with, I believe it was, P rofessor Wallace.

23 Q.

Well, what did P rof essor Wallaca have to say w.

=

TOOKER & ANTI 681 Market Street San Francisco 94105 415/392-0650

eu i

i about the shot peening?

1 A.

I'm going to have to paraphrase the thing, but 1

I believe Jack indicated to us that the shot peening j

technique is section sensitive and since we were involved i

here with a heavy section, the improvement would not be i

substantial.

F Q.

What does "section sen,sitive" mean?

j i

3 A.

I would like to give an example that would I

)

provid'e a comparison.

)

Shot peening a thin piece of steel,of the same specifications of the crank shaft would substantially 1

improve its fatigue strength while applying the same i

surf ace improvement technique to a thick section, like a i

i crank shaft, would not provide a substantial improvement 3

in the fatigue strength of the piece.

i 5

MS. TARLETZ:

Could I have that answer read T

back, please.

3 (Question and answer read.)

1 MS. TARLETZ:

Thank you.

3 MR. DYNNER:

Q.

Mr. Trussell, what do you mean L

by a substantial improvement?

1 A.

Something more than five percent.

3 Q.

Did anyone disagree with your recommendation 4

against shot peening the replacement crank shaft?

5 A.

Are you asking for a specific name?

5 Q.

Anyone.

7 A.

Someone did.

3 Q.

Who?

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TOOKER & ANTZ 681 Market S treet San Francisco 94105 415/392-0650

128 i

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Unat load level was postulated in connection 2

with the stress analysis of the AE piston performed by f

1 3

Mr. Scshouri?

4 MR. SMITH:

What do you mean " postulated"?

3 MR. DYNNER:

Q.

Assumed, taken into 6

censideration, utili=ed.

7 THE WITNESS:

Yes, I believe it was as high as 1

1 9

3 2,000 PSI combustion :oned pressure.

I'

}

1R. DYNNN'R:

Q.

Is that the pressure inside l

l 0

the cylinder head, this cylinder you are talking acout?

l t

1 A.

Yes.

l 2

Q.

And does the pressure in the cylinder of the

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DSR 4 8 engino ever reach 2,000 PSI?

j i

4 A.

I think you ha ve to -- I don 't know which 1

5 engino you are talking about.

l 6

Q.

DSR 4 8 engine.

7 A.

• Not to my knowledge.

l 6

Q.

Well, acw high does the pressure in the 9

cylinder reach in the DSR 48 enginc?

O MS. TARLET:s Under what application?

1 MR. DYNNER:

Q.

Can you answer that question?

2 MR. SMITH:

I join in that oojection.

'!ha t is maximum pressure in 3

MR. DYNNER:

Q.

i i

l 4

the cylinder during operation of the DSR 48 engine of the 5

same type as at Shoreham?

6 MR. S:!ITH:

At any timu?

7

!R. DYNNER

taximum pressure in the cylinder.

,o THE WITNESSt I think ap,aroximately 1900 PSI.

.I A

Toc.; E.1 & /.37:

661 :tarhut Street San Francisco 94105 415/392-0650

-,m 129 1

1R. DYNNER:

Q.

And is that full load?

l,_ 5 r e E hwuWM_-

2 A.

I believe that is overload.

r-.

3 Q.

All right.

What is the overload condition at 4

wnien that pressure is reached in the engines at Shoreham, 5

the kw rating, if you know that?

6 A.

I believe it's approximately 3900 kw.

i

)

7 MR. SMITH:

It is now 5 :31 according to my 8

watch.

In any event, if you have one or two more 9

questions on this line, go ahead, but I am going to call g

.0 an end to the day.

l

.1 MR. DYN!!ER:

That covers the questions along

.2 chis particular line.

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~ R i t **lf'ENTROLi.c0 " ~ ' " ' ' I ~ X.HOREHAM i RYJISk N CLEAR POWER STATIOt4 RTUP FORM 8.3 4 Precoerational Test Results Review and Acoroval 1 R43A System No. 1, s7 2.- Preoperational Test No. A107_00 d e-.nn -su m.,,_ 3;

3. -System EDG 10l j ud ificatinn c

m 4. Test Engineer William J. (*nnie /; Lead Startup Engineer M w u.14 wj 5. Attached for your review are: 6. Preoperational Test Results and Analysis E_ / ) Prepared (py/P rfc med Gy M /F j db*l& k W ' 9 /h f l$ 4 16wed ny' ,/ / ib h 4 t[ML N Star t".anag et piovt Prcoperational Test Approval / Release For Performance t t i (S tartup 8 8.1) System Checkout & Initial Operations Tests a Test Change !!ctice(s) (Startup 4 8. 2) A Tosg Results attached are Approved by the /~'TG 7. pe ion h Y t LJ JTG Chat man / [ Op h s fianag' r / As '/b ff 13lP4 Y blu ~ I Date A visory Opqtati s gineer A-70Rr 4 L Start anage g A09106 ~

f PT.307.004A-2 (~*

.g cnTaca.E0 j

p; laiormatica OnN APPENDIX F ENGINE CYLINDER PRESSURE LOG Step No. 8.8.2 //!8f Date 88N Time M @I ENGINE / CYLINDER PRESSURE (PSIG) t I nSo 2 lLYo 3 J L40 / 4 /44 c j 5 /6@ 6 fyoo 7 /dTo 8 /900 Gen. Lead (KW) 37* Var Loading (KVAR) A COU vt2A1 Data taken by: v l 1 i l ~ mm NOV 81983 37 A09J42 L . ~.

7.. L r, i O O O C U = l e-CD \\ l P. l m O w.

2 . je i e i I u. Fa AA-84-3-16 g PA0 7396/PRJ-03310A i EVALUATION OF DEERGE! ICY DIESEL GENERATOR CRANKSHAFTS AT SHOREHAM Ale GRAle EJLF IWCLEAR P(hiER STATIONS The report is final, pending confirmatory reviews required by FaAA's QA operating procedures. 1 Prepared by Failure Analysis Associates Palo Alto, California Prepared for TDI Diesel Generator Owners Groups i i May 22,1984 I l

getance, the TDI method does not compute the phase relationship between the ,arious orders or modes, so it is not possible to compute the true sumna-tion. The actual maximum stress is a direct result of this sungnation. Fur-thermore, the TDI method always predicts maximum stress in Crankpin No. 8, which is generally 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 additional spring placed between the generator and tround to represent the effect of the grid on dynamic response during synchronous operation. This ' spring constant was found to be 1.409 x 10' ft.-lb./ra:tian based on generator specifications. This constant is set close to zaro to represent SNPS en.ergency bus operation. The first five torsional natural frequencies for the replacement crank-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 emergency bus the first natural frequency is 0 Hz (rigid body mode). Tne other natural frequencies are in agreement with those computed by TDI and measured by SWEC. When the diesel generator is running at a given speed and power level, the forced vibration problem is steady-state where both load and response repeat themselves every two revolutions of the crankshaf t. To model the dynamic response, a model superposition analysis [3-1] was used with harmonic load input. The calculation of the harmonic loads will' be discussed in the next section. 3.1.2 Harmonic Loading To calculate the harmonic loading on a crankshaf t it is necessary to consider gas pressure, reciprocatirg 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 dead 3-2 ,.,,,w

J 1 f er. TDC, mark for Cylinder No. 7 was simultaneously recorded by a probe on t the flywheel. The pressure data at 1001 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 torque required to produce 3500 kW divided by the mechanical efficiency. A mechant-cal efficiency of 1.0 was obtained, rather than the expected 0.88. The dif-ference is' probably explained by either the pressure measurements being too low or by the TDC being shifted. Peak pressures were measured in all the , cylinders to ensure that all cylinders were balanced. I Normally, the excess torque above that required to run the engine at 3500 kW is dissipated by friction. In this case, because the pressure curve produced the correct power without friction, friction was not applied. The 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 crankshaft. The effect of this torque was combined with the gas pressure torque. The total torque was then decomposed into its sine and cosine harmonics corresponding to each order. These torque harmonics were used in the steady-state analysis. The magnitude of the torque harmonics are normalized by divioing by the piston area and thro-radius. The resulting normalized torques for the most significant orders are shown in Table 3.2. I 3.1.3 Comparison of Calentated Response With Test Data The response due to the first 24 orders and all 11 modes is calculated using modal superposition with 2.5% of critical damping for each mode. The actual value of damping used has little effect on the response since the orders are not at resonance at 450 rpm. The SWEC test report stated that the measured damping in the system was 2.6% [3-2]. 3-3 l l

Section 3 References 3-1 Timoshenko, S., D.H. Young, and W. Weaver, Jr., Vibration Problems in Engineering. Fourth edition. Wiley, 1974 3-2 Bercel, E., and Hall, J.R., " Field Test of Emergency Diesel Generator 103." Stone & Webster Engineering Corporation. April 1984 f 3-3 Peterson, R.E., Stress Concentration Factor. Wiley & Sons, New Yorx, 1974 3-4 "R-48 Crank Crankshaft Stress Analysis," Transamerica Delaval Inc. Report No. CR-01-1983. 3-5 Fuchs, H.O., and Stephens, R.I., Metal Fatigue in Engineering.

Wiley, 1980.

3-6 Bercel, E., and Hall, J.R., " Field Test of Emergency Diesel Generator 101," Stone & Webster Engineering Corporation. October 1983. 37 Collins, J.A., Failure of Materials in Mechanical Design. Wiley,1981. 3-8 Nishihara, M., and Fukui, Y., " Fatigue Properties of Full Scale Forged and Cast Steel Crankshaf ts," Transactions of the Institute of Ma rine Engineering. Series 8 on Component pesign for mgnty Pressure-Lnarged Diesel Engines, London, January 1976. 39 Burrell, N.K., " Controlled Shot Peening to Produce Residual Compressive Stress and Improved Fatigue Life," Proceedings of a Conference on Resid-ual Stress for Designers and Metallurgists American Society for Metals, April 1980. / 3-12

3 PRE 550RE (PSIG) it .am. \\ ? I 1 .me. I t l asi l i l u f i, l i = 4 l CRAN < ANS E (DEGREES I load. Measured pressure v'ersus crank angle at 100 Figure 3-1. FsAA-84-3-14 e 6 .-_ -. - - - - - - - _ ~... -,,

p,. g,ma 8 5, i b-s 0 l 1 6 O l O O I 5 C E 6 O m. 1 l' i 3 ( tt l c ; e L, I I t t I i l 4 r 1 1 1 k - - - ~ -, - .wy,r,___,~_-,,,,-c-.--_ m.-,. y,_,,,w,my,- ,,v, ,y-

ORIGINAL UNITED STATIS OF AMERICA NecLEAR RIGULATORY COMMISSION in the matter at TRANSAMERICA DELAVAL INCORPORATED O'*" N#- CWNERS GROUP MEETING r p.. Locanon: Wading River, N. Y. p,,. 1 - 197 ] Date Thursday, March 22, 1984 1 TAYLoE ASSOCIAT13 c wi soonas

• M iSuess. 1 w Sansseus
• eemam O C. m e 33 2BbMSB

I 70 1 that, I hope, but I think Paul has some input that would 2 suggest that perhaps the operating conditions are not 3 similar. And the last thing you mentioned, inspection 1 4 and test, to me, that's going to be a very Laportant 1 8 aspect of all this, and I don't see inspection and test e addressed in the piston skirt report. In other words, ? what's going to be done subsequently, to verify that j l these are indeed behaving satisf actorily on the basis S 8 of inspection and test. 10 ;' That's a key element of the ownership program plan i a 11 I but it's not a matter that's identified, or addressed J f u specifically for these piston skirts at this time. I D MR. MUSELER: Well, what we certainly--with respect i 14 to the inspections, I think we mentioned that seven 13 of the ten pistons have already been inspected after is a hundred hours at full pcwer, and roughly three hundred 17 total hours, which is mote than, much more than the is cycle they're going to see, at least in the first eighteen 1

• 18 months.

i O The intention there is to use the--and those inspection s 21 are, you know, are a matter of record. So our intentions, l 23 if those inspections started to show us that we were O experiencing problems, might get back to this final 24 loop that has to be closed when early inspections are D 'done, in terms of whether it tells you something that i j

r 71 1 contradicts your original premise. With respect to the 8 Kodiak, Alaska engine, in point of fact, that engine, 8 we think typica?.ly runs at about to percent full load, l 4 and with respect to -- and so it doesn't run at a hundred I which is what we've beeen running shorehas

percent, 1

8 for at least a hundred hours. But relative to the at service these engines actually have to see, it's more I 8 I severe service than these engines actually see in nuclear service, at least certainly for the first cycle, and to ' probab1y for several cycles. II And the reason I say that is, that a, is that i 12 one of these engines, or all of these engines, if it 18 ever had to perform its intended function, would see 14 a high load for a very short period of time, and then I la would drop well below the level that the Rodiak engine la runs at continuously. So while there's same--you know-- 17 there's certainly a difference between going the high is I load for a short period of time. 19 If you look at the histograms of that engine versus a what these engines actually have to do, we think Kodiak 21 is significant from that standpoint, even though it's a not run at a hundred percent. m The R5 engine, on the other hand, ran, I guess, l about six hundrad hours at horsepower per cylinder, l instead of six hundred, it runs up around nine hundred,

7 m 72 1 or something. And with respect to peak pressures--because %T 8

t guess they're not always analogous--with respect to T

3 peak pressures, though, it does run at a significantly y T 4 higher peak pressure. Our normal pressure is 1670, and 8 at the overload rating, is 1750 in the Shoreham engines. 8 That engine, in order to run at that high horsepower, 7 ' runs typically around 2000 psi. So while there aren't a a lot of pistens that have experienced that, there are at least the two that we've examined, that have gone i f through six hundred hours at 2000 psi, and I don't know 10 f the percentages, but it's probably in the neighborho'od II of thirty percent higher peak pressure, peak firing 13 i pressures than our engines, and that data does give 14 l us a considerable amcunt of confidence. I'd rather have a hundred pistons that have seen a lot of service, but 18 based on what we've seen so far, I think we have a high 17 l l degree of confidence. Is l And we're not basing it--let me say again, we're is not basing--and the report says that, I believe, cartainly O in the front--that we're basing the conclusion--our 21 conclusion is that those pistons are unlikely to crack, O and that further, you know, Dr. Wells points out in D the report, that any cracks would not propagate. But 24 we don't want them to crack, and we think that, not O based on the analysis alone, but based on the information

I 6 g 4, i e e em a C I 3 i ft O t s b i

I e L Transamenca Delaval 1253 ! f;llE5 I o L1 I instruction Manual Model DSR-48 Diesel Engine Serial Nos. 74010-2604 74011-2605 74012-2606 l LONG ISLAND LIGHTING COMPANY Shoreham Nuclear Power Station I Unit No.1 i TransameriCa 39 aVaL 'nC. Engine anc Compressor Jivision

Instruction Manual a-a APPENDIX 11 OPERATING PRESSURES AND TEMPERATURES 1 PRESSURES The following pressures should be present for starting: Starting Air Supply 250 psi ...................................... 17.6 kg/sq cm Starting Air Header 250 psi ...................................... 17.6 kg/sq cm While running at rated speed. the operating pressures should be as follows: psi in-hg kg/sq cm Lucricating Oil' 50 -55 101.8 - 112.0 3.52 - 3.87 Lubricating Oil at Turbocharger inlet 20 - 25 40.7 - 50.9 1.41 - 1.76 Jacket Water 10 - 30 20.4 - 61.1 0.70 - 2.11 Fuel Oil 20 - 30 40.7 - 61.1 1.41 - 2.11 TEMPERATURES While running under rated load. the outlet temperatures should be as follows: Lubricating Oil out of Engine

• 1708 F - 1808 F (76.6' C - 82.2' C)

Jacket Water out of Engine 1708 F - 1808 F (76.6' C - 82.2' C) EXHAUST TEMPERATURES. The exhaust temperatures shown on the " Factory Test Results" page are the average for att cylinders during factory test under local ambient conditions. Temperatures in the field, therefore, may exceed this average temperature. Exhaust temperatures may be considered normalif within plus or minus 50' F of the average taken for all cylinders. Temperatures, hign or low, exceeding this range should be investigated (see Section 7). The exhaust temperature limits for sustained operation is 150' F between any two cylinders and 1100$ F maximum. FIRING PRESSURES. Firing pressures may be considered normalif within plus or minus 75 psi of the average for allcylinders. High or low pressures exceeding this range should be investigated '(see Section 7). The finng pressure limits for sustained operation is 200 psi between any two cyhnders. NOTES. Operating pressures and temperatures listed are estactished as a guide to proper operation. Except as noted for exhaust temperatures and finng pressures, they should be held to within plus or minus 10 percent. Sudden Changes in readings require immediate investigation and correction. When making adjustments as a result of a high or low cylinder' exhaust temperature, or finng pressure. both temperatuIe and pressure readings must be taken into account when determining tne proper corrective action. l 'When using SAE 40 fuencating oilin engine. e...,,m...., o9 l

n, o FIELD TEST OF EMERGENCY DIESEL GEhTRATOR 103 WITH 13 x 12 CRANKSHAFT Prepared for SHOREHAM NUCLEAR PokTR STATION LONG ISIAND LIGHTING COMPANY by E. BERCEL J. R. HALL l' APRIL 1984 Approved by: ~ unmible E=gineer w itesponsible Engineer E. Bercel P.. Hall STONE & WEBSTER ENGINEERING CORPORATION B1-1160037-1

r i I' I 6.2 PRINCIPAI, STRESS CAI.CUI.ATIONS The synchronously averaged strain signals described in Section 6.1 were digitized inte 1024 data points in the Nicolet 660 B. The digitized strain data from the three elements of each strain sage rosette were read by an HP 9826 computer. The applicable static strain component computed as ~ discussed in 6.1 was added to each data point and the time history of the principal stresses, and the maximum shear stress, as well as the bending and torsion related stresses, were computed for the 0.4-second (1.5 strain i cycle) time window of the strain samples. l 6.3 OUTPUT TORQUE The torque measurements were made by measuring the strain in the output i shaft as described in Section 3.1.2. Of the two torque bridges installed, one was subject to some reduction in signal quality, but the other one (TQ9) provided excellent dynamic signals throughout the test. Using the j technique described in 6.1 the torque-power relationship was determined in terms of micro-strain /kW. The torque vs output power curves are presented in Tigure A-27 and A-28 in Appendix A. The obtained value was very close to the one calculated from the known material properties and geometry of the output shaft at the section where the strain gage bridges were located. The measured value for the torque power relationship was 63.2 micro-strain per l 1000 kW generated power (four-arm bridge, four times the actual strain in 1 the shaft). Using the vendor specified figure of 96 percent for generator efficiency, the calculated torque-power relationship is 63.2 micro-strain i 31-1160037-1 6-2

F per 1000 kW of generated power. Since the measurement involved very low level signals the probable error is estimated at 5 to 8 percent. The torque measurements at various selected load levels were analyzed in both the time domain and the frequency domain. Synchronous averaging was us ed "in the time domain analysis. In the frequency domain, synchronous averaging was employed in some of the analysis to obtain phase measurements relative to the firing top-dead-center of cylinder No. 7. These measure-seats were obtained in the 500 Hz range which provided a 0.8 second sample. Ensemble averaging was used for general spectrum analysis to a'ecurately display a region of resonance in the spectrum. Synchronous averaging tends to attenuate the data near a point of resonance in the spectrum as a result of the random nature of the spectrum there. The frequency range used was 50 Hz. 6.4 TORSIONAI. VIBRATION The torsional vibration data were analyzed in the time and frequency domains using the same techniques as described in 6.3. Synchronous time averaging, of the torque and torsional signals together, was also performed to display the two phenomena in relation to each other. 6.5 CYI.INDER PRESSURES Synchronous time averaging of all three cylinder pressures was performed. Frequency domain analysis was also done to measure the amplitude and phase of the 7.5. Hz and 15.0 Hz components of the cylinder pressure pulse, wcich B1-1160037-1 6-3

1 sechanical variables and output power and the results of the variable speed test are illustrated in Figures 3-1 through 3-11. Figures 3-12 through B-78 contain the time domain data of the mechanical variables, including the calculated principal stresses and transient phenomena. The time-domain records of the electrical variables are in the third group in Tigures B-79 througfa B-86. Finally, the frequency domain plots are presented in the fourth group in Figures B-37 through B-96 for both mechanical and electrical variables. In each group the figure numbers are arranged in ascending order with generated power to assist the' reader. 7.2.1 Strain Measurements i In comparison to the dynamic strain, the static component of the measured strain was small. Since the dynamic range of the instrumentation had to 1 accommodate the total strain, the static strain components were in the botton 5 percent of the total measurement range. Nevertheless, the proce-dure described in Section 6.1 enabled the measurement of those components to satisfactory accuracy (about +/- 5 percent). The values determined for a the various load levels are given in Table B-1. f Measurements 5-3 and 7-1 represent the tensile components while 5-1 and 7-3 l are the compressive strain components. The dynamic strain records are presented in the time domain only (Tigures B-12 through B-18, B-24 through i B-30, and B-37 through B-43). Each of those records represents 48 strain cycles averaged synchronously over a period of 12.8 seconds. To facilitate analysis, all records have been plotted to the same scale with a rero average and have been triggered at the same point in time. The time of the B1-1160037-2 7-3

l SNPS-1 FSAR TASLE 8.3.1-t EIGERGEseCV DIESEL GEDGERATOR SYSTEst REQUIRED t GADS A86 80AM101UII C0ltsCIDENT DEteAte Dhamber RecsJtred Nameplate Total Design Sasis Loss Loss of Maxtaum Colncident Rating Plant of Coolant Accident Offatto Power Demand (K1lowatt) " * ' " Functton (9tpl M*r 0-10 Iten 10 atin on (Hot S tanew) 0G-809 DG-102 D4-903 Core Spray Pump 9250 2 1* I 998 998 Resteaal temat Removat Pump 1250 4 2* 1 2 999 999 1998

• g Servtce Water Pump 450 4

2 2 3 "* 358 358 796 s RSSVS and CRAC Water Chiller 292 4 2 2 2 235 235 470

• l RSSVS and CRAC Water Cheller Lube Oil Pump

.25 4 2 2 2 0.2 0.2 0.4

• RSSVS Cheller Circ.

Water Pump 75 4 2 2 2 60 60 820

• R85VS Chtller Cond. Water 20 4

2 2 2 16 16 32

• RBSVS Unit Cooter 30 8

4 4 4 96 96 2 2 42.5 82.5 82.5 RSSVS Eutuust Fan 100 3 2 Reactor ButIdeng Exhaust Booster Fan 7.5 2 8 8 9 6 ,6 R85VS Filter Reheat Colt 6.6 kW 2 1 1 1 6.6 6.6 RBCLCW Carc. Pump 100 3 2 2 2 SO SO SO I Diesel Generator Air 12 82 12 Compressor to 6 I Olesel Generator Fuel I Ost Transfer Pump .5 6 2 2 2 0.4 0.4 0.4 Diesel Generator Jacbet 72 "" 72 "" 72 " " Water temater 36 kW G Diesel Generator Jacket 2. 5 " 2.5"**

2. 5 " " l water steep Warm Pump 2.5 kW 3

0tese1 Generator Lsee 20 " *

• 20 " "

20 " " OeI teeater 20 hw 3 OteseI Generator Before 4 o s. 4.. s. 4.. n & After tube OII Pump 5 3 Diesel Generator tteater 4.2 kW 3 4.2 " " 4.2**" 4.2"" Battery Charger (825 V) 60 kva "" 3 2 2 2 20 25 87 120 V ac Instrument Power 100 kva DG 108 3 2 2 2 SO SO 40 100 kva DG 102 50 kva DG 103 320 V Nonemergency Feeds 65 kVa X"* 52

• OIeseI Generator Room Vont Supply Fan 20 3

2 2 2 16 16 16 i Battery Room Vent Supply Fan 2 3 2 2 2 1.6 1.6 1.6 l Control Room Air Condttlon-Ing Ungt 40 2 9 9 1 33.9 33.9 l Control Room Vent Sooster Fan 7.5 2 1 1 1 6.0 6.0 l t of 4 Revision 31 - August 1983

SNPS-9 FSAR TASLE S.3.1-1 (CON 7*0) thoutper Secmatred stamepla t e Total Desegn Saola css Loss of stau taue Coincident RatSng Paant or Coolant Acetdont offa to Power Demanse futlowett) * "

• Functton (Hp) u*e O-SO men to men on Otot stenew)

M-to t M -102 M-103 Eeergency suetchgear. ReIay & evter aooms air Corus4-ttoning untt 40 2 9 1 1 33.9 33.9 ISC A9r Corusa t eoneng 40 h W t 1 1 40 ". * *

• Unit 15C A8r Cooeed 30 hW t

t 9 30 ". * *

• Condenser Eeergency 5=ttceigaar ReIay &

RS5v5 ChatIer Soom Embaust 1 1 8.O 8.O g Computer Soomms Emhaust Fan SO 2 1 Fan 3 2 1 1 0 2.4 2.4 Screenwell Enhaust Fan 80 2 1 1 1 S.O S.0 Screenwell Interpossng a kva t t t 9 O.8 Relay Panel teCC Room ventilatton .75 2 1 1 0.5 0.5 l LPCI M-C Set soon vente-lation 3 4 2 2 2 2.4 2.4 4.8 Untt Cooler teCC OSS Room t.5 9 1 1 1 9.2 1 Spent Fuel Pool Coollng Water Pump 30 2 1 1 24

• 24 "*

8 L000 Levei Pump (C5. aHe, sePCI. RCIC) 7.5 4 4 2 4 12.0 12.0 I Atmospher8c Cont. - thKs. Reconhecer SOS kW 2 4 108 "* tOS

• sesiv-LCS temators G.E kw 4

26. 4 ***

M5iv-LCS Slowers 4.4 3 7* 3.5 "* Radiatto.,teonttortng 1 to 4.8 3.2 Lighteng (Egatwalent kW) 407.2 kw X* 190 *** 227.2

• Fence Securtty Light 9ng 60 kW I*

34 2g Beactor Protectton System N-C 5e t *** 25 2 2 20 " ** 20 " ** l Reactor Protect ton System Sackup Transformer 25 kva 3 1 20 * *

• Sattery Charger 224 W 3 kva 4

2. 4 "a 2,4 ***

Uninterrietthle Power (Wital Sus) ** 37.5 kva t t 9 t 30 Unenterruptthle Power (Security & r - et-ca t t ons ) *** 20 kva 1 1 8 1 16 Sattery Charger (Secursty a,us Co==unscatlon) 20 kva ** *- 9 4 l Uninterruptible Power (Computer Bus) *- 20 kva 9 9 9 9 16 1 206.1

• 206.t * -

Control Rod Dr Iwe Pump *** 250 2 Drywell Coot eng System f an *** 25 8 4 80

• 90
• 2 of 4 Sev t s t on 3 8 - Augp.as t 1983 L

w' - J

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• 2 1

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I l SasPS-t fSaa TASLE 8.3.1-4 (CONT *0) Total kW (Pelor to SO altmates) 3400.3 3364.6 3000.7 0 -9300.3 0 solemas sente 4 Loads O O + 70 0 Plus Isole 94 Loads 340s.3 3364.6 3640.5 Total kW (Atter to etnutes) NOTES: eartrag the O-90 etraste period af ter a deslyt laests loss of coelent accleont thea teum coincident demand setourt rws l (LOCA) from manufacturer's data for the C5. nee, service water pumps, motor-generator sets, nesv5 l =- sci touat e toeds alwen are Chllier earnt ts, and all motors greater than SCO @, On loss of ofistto power. et is necessary to go to a cold setutasaura consett ton if 98-903 does rent start. slace the three regatred service neater pumps esllt not he awellatale. Dente that only teso service aseter pumpe are respaires$ for a deelgra O taasts LOCA condit ters. (Only one pump la connected automatically to DG-SO3. the other may tse connected manually only. ) Teso uralts are started ces OG-lO3. One unit is shut doesn istiert it is determined ishlcet section of the system will tse used. Paese nonclass IE ewts are not rosystred for a safe shutdoesn. Leading indicatast for iserious modes of operatican le destraele, although not essential. All reselntrag components are Class IE. Deinlous safe samatdoesn respatrements f or a suctlert line breek. Actual pump respairements depenst ort lareek location (see Section G.3.3). E IndtCates load re@etred. l These loeds are tripped Intenttonelly (automatically) on a LOCA. These loads are riot storeally operating enes recalve no automatic start signal af ter a LOCA. l

• *
• These nonsafety related loads have seal-in type control circuits that drap east art a loss of of f et te posser prior to connecting to the diesel generators.

These SIDW*s are connected to their respecttwo diesel Inuses knut dshernet operate upon a LOCA.

• • " the load to be carried by the 30-G 5ets consist of certain motor-operated valves. On Unit 903, one set operates at full 3 load and one set operates unloaded these loads are automatically trlpped ashort diesel generator starts.

These loads are prevented from starting until 10 strustes af ter a LOCA signal. Loads leposed Dy t,attery chargers are based on the oc loading of the bettery chargers. 4 of 4 Sewtston 39 - August 1983 __}}