ML20090K382

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Forwards Tdi Diesel Generator Rept on Evaluation of Emergency Diesel Generator Crankshafts at Shoreham
ML20090K382
Person / Time
Site: Shoreham File:Long Island Lighting Company icon.png
Issue date: 04/24/1984
From: Britt L
LONG ISLAND LIGHTING CO., TDI (TRANSAMERICA DELAVAL, INC.) OWNERS GROUP
To: Bordenick B, Repka D
NRC OFFICE OF THE EXECUTIVE LEGAL DIRECTOR (OELD)
References
NUDOCS 8405240189
Download: ML20090K382 (1)
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  • P.O. BOX 628 = W ADING RIVE R, NEW YORK 11792

%w wa. - n m wr v mt_ TE L. (516) 929 8300 s' yy April 24, 1984 Via Federal Express Bernard M. Bordenick, Esq. David A. Repka, Esq. U.S. Nuclear Regulatory Commission Washington, D.C. 20555 Gentlemen: In accordance with our committments, copies of the following report were issued during the week of April 16, 1984. Distribution is indicated on the attached service list. TDI Diesel Generator Report on the Evaluation of Emergency Diesel Generator Crankshafts at Shoreham Very truly yours, C L.F. Britt Supervisor of Regulatory Services Nuclear Operations Support Department LFB/it cc: NOSD SR2 01 8405240189 840424 PDR ADOCK 05000322 S PDR FC 8682.2 L /

LONG ISLAND LIGHTING COMPANY / FAC'O JSearmcaaw SHOREHAM NUCLEAR POWER STATION P.O. DOX 610. NORTH COUNTRY RO AD e WADING RIVER. N.Y.11792 m Direct Dial Number April 20, 1984 TDI-17 Mr. H. R. Denton, Director Office of Nuclear Reactor Regulation U. S. Nuclear Regulatory Commission Washington, D. C. 20555

Dear Mr. Denton:

Please find enclosed fifteen (15) copies of the TDI Diesel Generator Report on the Evaluation of Emergency Diesel Generator Crankshafts at Shoreham. An additional five (5) copies of the report are being forwarded to Battelle Pacific Northwest Laboratories, per NRC request, under cover of this letter. This report has been prepared and is being submitted on behalf of the TDI Diesel Generator Owners' Group. Although the subject of this report is the crankshafts at both the ,Shoreham and Grand Gulf stations, at this time only the Shoreham crankshafts are addressed. The report will be supplemented at a later time to include Grand Gulf. Should you have any questions, please do not hesitate to contact us. Very truly yours, QR W. J. Museler Technical Program Director TDI Diesel Generator Owners' Group JTM/vf enclosure y N cc: Owners' Group Executives Owners' Group Alternates C. Berlinger R. Caruso (15 enc 1) Battelle Northwest (5 encl) l l bec: M. H. Milligan T. Ellis (- aba@T;:rMr!asic)) A. Ear 1ey S. Parkhurst (2 encl) Owners' Group Site Reps. f C. K. Seaman M. Lowrey J. T. Murphy C. Woodard (NRC Site Rep) C. Mathews J. C. Kammeyer i C. Wells C. Ray l J. M. Hart l R. L. Gill

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Pa. O a. to r at es cw.,a w L +,4 's e., w. ,t i f n 'c4.'. FaAA P4-3-16 PA0 7396/PPJ 0331nt EVALUATION OF EMERGENCY DIESEL GENERATOR CRANKSHAFTS AT SHOREHAM AND GRAND GULF NUCLEAR POWER STATIONS The report is final pending confirmatory reviews required by FaAA's OA operating procedures. l O l Prepared by Failure Analysis Associates Palo Alto, California Prepared for 'r[1 TDI Diesel Generator Owners Groups 1 \\ o,i April 19, 1984 p OV s a.c a.tt . tos v.ta. m. ....+ i . :se '. . i.

i l STATEMENT OF APPLICABILITY This report addresses tne structural integrity of the crankshaf ts in I Transamerica Delaval Inc. DSR-48 engines at the Shoreham Nuclear Power Station l and DSRV-16-4 engines at the Grand Gulf Nuclear Power Station. In view of possible differences in generators, flywheels, and engine operating l conditions, the results do not necessarily apply to other engines of the same model. These plant-specific dif ferences, where they exist, will be evaluated in sw* rate reports. l O l O -iA-

EXECUTIVE SUIMARY v DSR-48 13-INCH BY 12-INCH CRANKSHAFTS AT SHOREHAM IRICLEAR POWER STATION The structural integrity of the replacement 13-inch by 12-inch diameter crankshafts installed in the emergency diesel generators at the Shoreham Nuclear Power Station has been extensively evaluated by testing and analy-sis. Conventional analytical techniques typically utilized by the diesel engine industry show that 13-inch by 12-inch crankshafts comply with DEMA requirements. Angular displacements of the free end of the crankshaf t, stress ranges in the most highly stressed crank pin fillets, and the range of output torque at the flywheel were measured at and above full-rated load. The tor-siograph neasurements of twist showed that the crankshafts meet the DEMA requirements. In addition, the strain gage measurements of maximum bending and torsional stress and calculations of maximum stress by a modal superposi-tion analysis showed that the crankshafts have a factor of safety in fatigue of 1.48 without taking into account any benefit of shot peening the crank pin ( fillets. The factor of safety was determined from the measured endurance i limit of the original 13-inch by 11-inch crankshafts that cracked in fatigue. The measured shaft response was in close agreement with that pre-dicted by the modal superposition analysis. J The replacement crankshafts are suitable for unlimited operation in the [ emergency diesel generators at SNPS. l 1 l, i l Y

PA0 7396 s_-) Task No. 03310A TABLE OF CONTENTS Page STATEMENT OF APPLICARILITY............................................... i EXECUTIVE

SUMMARY

ii PART A

1.0 INTRODUCTION

TO REVIEW 0F DSR-48 13-INCH BY 12-INCH CRANKSHAFT..... 1-1 Section 1 References............................................... 1-2 2.0 COMPLIANCE OF CRANKSHAFT WITH DIESEL ENGINE MANUFACTURERS ASSOCIATION RECOMMENDATIONS........................................ 2-1 2.1 Review of TDI Torsional Critical Speed Analysis............... 2-1 2.1.1 Natu ral F requenci es.................................... 2-2 2.1.2 Nomi n a l St re s s e s....................................... 2-2 2.2 Review of Stone & Webster Engineering Corporation T o r s i o g r a p h Te s t.............................................. 2-3 ('_s) 2.2.1 Na t u ral F requen ci e s.................................... 2 4 s/ 2.2.2 N om i n a l St re s s e s....................................... 2 4 i m l 2.3 Nominal Stresses for Underspeed and Overspeed Conditions...... 2-5 l Sect i on 2 Re f e re n ce s............................................... 2-6 3.0 F AT I GUE AN AL YS IS OF CR ANK SHAFT..................................... 3-1 3.1 Cra nk sha f t Dynami c Torsi onal Analysi s......................... 3-1 3.1.1 To r s i o n a l M o d el........................................ 3 - 1 l 3.1.2 H a rmo n i c Loa d i n g....................................... 3-2 l~ 3.1.3 Comparison of Calculated Response With Test Data....... 3-3 l 3.2 C ra n k s h a f t St re s s Ana ly s i s.................................... 3 4 3.2.1 F i ni te El eme nt Mode 1................................... 3 4 3.2.2 Stres ses Due to. Torsional Loadi n g...................... 3-6 3.2.3 St resses Due to Gas Pres sure Loadi ng................... 3-7 3.2.4 Compa ri son of Stresses with Test Data................,. 3-7 3.3 Crank sha ft Fati gue Fail u re Ma rgi n............................. 3-8 3.3.1 Stresses in Replacement Crankshaf ts.................... 3-8 3.3.2 Endurance Limit for Fail ed Crankshaf t................... 30 3.3.3 Endurance Limit for Replacement Crankshaf ts............ 3-10 3.3.4 Factor of Sa fety Against Fatigue Fail ure............... 3-11 l l Sect i on 3 Ref e re n ce s............................................... 3-12 l (_-) -iC-i l l l t

g) (G TABLE OF CONTENTS CONTINUED page 4.0 D I SCU S S I ON AND CON CL U S I ON S......................................... 4-1 Section 4 References............................................... 4-2 PART A APPENDIX.......................................................... A-I PART B

5.0 INTRODUCTION

TO REVIEW 0F DSRV-16 4 13-INCH BY 13-INCH CRANKSHAFT... 5.1 I n d u s t ry E x p e i e n c e...............,............................. Section 5 References................................................ 6.0 COMPLIANCE OF CRANKSHAFT WITH DIESEL ENGINE MANUFACTURERS ASSOCIATI09 RECOMMENDATIONS......................................... 6.1 Revi ew of TDI Torsional Criti cal Speed Analysi s................ -s [\\- / i 6.1.1 Natural Frequencies..................................... 6.1.2 Nomi n al St re s s e s........................................ 6.2 Review of Stone & Webster Engineering Corporation Torsiograph Test............................................... 6.2.1 Nat u ra l F re q ue n c i e s..................................... 6.2.2 N om i n a l S t r e s s e s........................................ 6.3 Nominal Stresses for Underspeed and Overspeed Conditions....... Section 6 References................................................ 7.0 CRANKSHAFT DYNAMIC TORSIONAL ANALYSIS............................... 7.1 To r s i o n a l Mo d el................................................ l 7.2 Harmonic Loading............................................... 7.3 Compa ri son of Cal cul ated Response With Test Data............... I Section 7 References................................................ 8.0 DISCUSSION AND CONCLUSIONS.......................................... OV -1D-

O PART A: REVIEW OF DSR-48 13-INCH BY 12-INCH CRANKSHAFT i O O

(]

1.0 INTRODUCTION

TO REVIEW W DSR-48 13-INCH BY 12-INCH CRANKSHAFT \\J As a result of fatigue dama5e in the crankshafts of three emergency diesel generator sets at Shoreham Nuclear Power Station, replacement crank-shafts of current design have been installed. The principal difference is an increase in crankpin diameter from 11 inches to 12 inches. This report pre-sents Failure Analysis Associates' findings on the adequacy of the replacement crankshafts in the emergency diesel engines at Shoreham Nuclear Power Station. A detailed investigation of the orginal crankshaf t, which attributed failure to high cycle torsional fatigue resulting from inadequate design, was previously conducted by Failure Analysis Associates (FaAA) [1-1]. An analysis of the replacement crankshafts, conducted prior to dynamic testing, was also performed by Fa AA [1-2]. The installation of the replacement crankshaft was required to meet the recommendations of the Diesel Engine Manufacturers Association (DEMA). In Section 2.0, the torsional calculations of Transamerica Delaval Inc. (TDI) [1-3] and the torsiograph test results of Stone & Webster Engineering Corpora-(,/ tion (SWEC) [1-4] are reviewed for compliance with the DEMA stress allowables. In Section 3.0, a detailed analysis of the factor of safety against i fatigue failure is performed. A torsional dynamic analysis is used to compute I nominal torsional stresses at each crank throw. A three-dimensional finite element analysis of a quarter section of a crank throw is then performed to obtain the local stresses in the crankpin fillet. The computed stresses are compared to dynamic strain gage measurements to verify the models. In turn, the models are used to verify that strain gages have been placed in locations of maximum stress. Finally, the measured stresses are used to compute a fac-tor of safety against fatigue failure for the replacement crankshafts. This is accomplished by comparing the measured stresses with the endurance limit for the replacement crankshafts, n 1-1

A Section 1 References 1-1 " Emergency Diesel Generator Crankshaft Failure Investigation, Shorehan Nuclear Power Station," Failure Analysis Associates Report No. Fa AA 83-10-2.1, October 31, 1983. 1-2 " Analysis of the Replacement Crankshafts for Emergency Diesel Generators, Shoreham Nuclear Power Station," Failure Analysis Associates Report No. FaAA-83-10-2.2, October 31, 1982. 1-3 Yang, Roland, " Proposed Torsional and Lateral Critical Speed Analysis: Engine Numbers 74010/12 Delaval-Enterprise Engine Model DSR 48 3500 XW/4889 BHP at 450 RPM". Transamerica Delaval Inc., Engine and Compres-sor Division, Oakland, California, August 22, 1983. 1-4 Bercel, E., and Hall, J.R., " Field Test of Emergency Diesel Generator 103," Stone 8 kebster Engineering Corporation, April 1984. O 1 O O 1-2

l /] 2.0 COMPLIANCE OF CRANKSHAFT WITH DIESEL ENGINE MMtVFACTURERS ASSOCIATION V RECOMENDATIONS l The specifications for the replacement crankshafts required that the rec-l ommendations of the Diesel Engine Manufacturers Association, DEMA [2-1], be followed. These recommendations state: In the case of constant speed units, such as gen-erator sets, the objective is to insure that no ham-ful torsional vibratory stresses occur within five percent above and below rated speed. For crankshaft <-, connecting shafts, flange or coupling components, etc., made of conventional materials, torsional vibratory conditions shall gener-l ally be considered safe when they induce a superim-l posed stress of less than 5000 psi, created by a single order of vibration, or a superimposed stress of less than 7000 psi, created by the sumation of the major orders of vibration which might come into phase periodically, l i In August, 1983, Transamerica Delaval Inc. (TDI) performed a torsional q critical speed analysis of the replacement crankshafts [2-2]. References to (V TDI analysis in the body of this report all reference this effort. In Section 2.1, this analysis will be reviewed for compliance with the above allowable stresses. The inappropriate T values employed in the original analysis of n the 13 inch by 11-inch crankshaft were replaced with the correct values for this analysis. In January, 1984, Stone & Webster Engineering Corporation, SWEC, conducted a torsiograph test on a replacement crankshaft at Shorehan Nuclear Power Station [2-3]. In Section 2.2, the test results will be com-l. pared with the above allowable stresses. 2.1 Review of TDI Torsional Critical Speed Analysis I Diesel generator torques due to dynamic response are usually calculated in two steps. First, the torsional mode shapes and natural frequencies of vibration are calculated. Second, the dynamic forced vibration response due to gas pressure and reciprocating inertia loading is calculated. TDI calcu-lated the response at 100% of rated level of 3500 kW. Ov 2-1

O 2.1.1 Natural Frequencies The first step in a torsional critical speed analysis is to determine the natural frequencies of the crankshaft. The engine speed at which a given or-l der resonates may then be calculated. The diesel generator is modeled as a system of lumped mass moments of inertia interconnected by torsional springs, as shown in Figure 2-1. The inertia and stiffness values are shown in Table 2.1. j It has long been standard practice in the diesel engine industry to solve this eigenvalue problem by the Holzer method [2-4]. This method has been used for at least 40 years [2-5], and thus is well established. TDI used the Holzer method to calculate the system's first three natural frequencies, which are shown in Table 2.2. The first natural frequency was found to be 38.7 Hz, which produces fourth order resonance at 581 rpm. 2.1.2 Nominal Stresses p) The second step in a torsional critical speed analysis is to determine 5v the dynamic torsional response of the crankshaft due to gas pressure and re-ciprocating inertia loading. The first order is a harmonic which repeats once per revolution of the crankshaft. For a four-stroke engine, harmonics of or-der 0. 5, 1.0, 1.5, 2.0, 2.5... exist. TDI performs this calculation for each order of vibration up to 12 separately. For each order, the applied torque at a cylinder due to gas pressure and reciprocating inertia is calculated. The values of this torque for each order are usually normalized by dividing by the piston area and throw radius. The normalized value for the nth order is referred to as T. The values of T for signifcant orders used by TDI are n n shown in Table 2.3. These values may be compared to those recommended by Lloyd's Register of Shipping, LRS, [2-6]. It is found that TDl's values are higher than LRS's values for low orders and lower for high orders.

However, the largest single order was measured to be within 5t, of those computed using TDI's values of T.

The response is then calculated by one procedure if the n harmonic is at resonance and - by another if the hartrionic is away from reso-nance. I i v 2-2 i

A (3 At resonance, the torsional vibration amplitudes would increase indefi-nitely in the absence of damping. The solution is obtained by balancing the energy input with the energy loss due to damping. TDI used an empirical form of hysteresis damping due to friction. The purpose of this calculation is to ensure that the diesel generator could be brought up to operating speed with-out undergoing excessive stresses as critical speeds are passtd. Observations have shown that excessive vibration during startup does not occur [2-3]. Since the engine runs at 450 rpm and the fourth order critical speed is SRn rpm, the calculated response at resonance will not be further considered. Away from resonance, the torsional vibrations reach a steady-state level even without the aid of any damping. The magnitude of this response for each structural mode and loading order is calculated as the product of a dynamic amplification factor and an equivalent static equilibrium amplitude, The equivalent static equilibrium amplitude is computed using a modal load and modal stiffness [2-7] for the nth order harmonic and given mode shape. The nominal shear stress, r, in the 12-inch pin of Crankpin No. 8 for each order is then calculated from the dynamic torque, T, using i = Tr/J, where r is the V) pin radius and J is the polar moment of inertia. l l TDI calculated the response for the first three modes and plotted the results for only the fi rst mode since higher modes produce much smaller ( stresses. The nominal shear stresses for the significant orders are shown in l Table 2.4 It is seen that the largest single order stress of 2980 psi at rated load and speed for the fourth order is well below the 5nn0 psi DEMA al-lowable. l TDI does not calculate the associated phase angle with the response of l each order, so that it is not possible to calculate the combined response. The measured combined response will be compared with the allowable in the next I section. l l 2.2 Review 'of Stone & Webster Engineering Corporation Torsiograph Test i Torsiograph tests are commonly used to confinn torsicnal vibrational cal-culations. The test is usually performed in two stages. The first stage is ,O V 2-3 s ~

I performed without load at variable speed and is used to determine the location V of critical speeds. Critical speeds may also be determined while operating at a fixed speed and observing the frequency content of the response. The second stage is performed at rated speed of 450 rpm with variable load, and is used to confirm the forced vibration calculations. 2.2.1 Natural Frequencies The frequency content of the torsional vibration signal at 450 rpm showed a resonance at 38.6 Hz. This value is in excellent agreement with T01's com-puted value of 38.7 Hz. 2.2.2 Nnminal Stresses The torsiograph provides the angular displacement response of the free end of the crankshaft. This displacement may be decomposed into components corresponding to each order. The peak-to-peak response may also be obtained. The nominal shear stress, r, in Crankpin No. 8 may be established from [a the amplitude of free-end vibration by assuming the shaft is vibroting in the first mode. The nominal she?r stress is then found to be 4562 psi per degree of free-end vibration from the TDI analysis [2-2]. SWEC tabulated the single order and peak-to-peak response for both 3500 kW (100% of rated load) and for 3800 kW (109% of rated load). These values have been factored to obtain nominal shear stresses and are shown in Table 2.5. The results at 100% load show that the largest single order has a stress of 3108 psi which is well below the OEMA allowable of 5000 psi. The total stress of 6626 psi is also shown to be below the DEMA allowable of 70nn psi. At 3800 kW the stresses of 3242 psi for a single order and 6875 psi for combined response are also lower than 5000 psi and 7000 psi respectively. At 3900 kW the corresponding stresses are 3287 psi and 6958 psi by linear extra-polation. However, the 3900 kW level is a two hour overload rating at which the engine is not required to operate continuously. A 2-4

f~~b (j The measured response at 3500 kW is in agreement with that calculated by TDI and shown in Table 2.4. The measured values are somewhat higher than the calculated values. 2.3 Nominal Stresses for Underspeed and Overspeed Conditions Strict interpretation of DEMA regulations requires the consideration of torsional stresses at conditions other than operating speed. During normal standby diesel generator testing at SNPS the units are synchronized to the Long !sland power distribution grid to simulate full load and two-hour over-load conditions. At this time the frequency characteristics of the grid assure that the speed and associated frequency vary less than 1 Hz or 1.6 percent speed. During other testing and potentially during a LOOP /LOCA event, the unit speed is controlled by the Woodward Governor. Testing at the SNPS site, during which step changes in load were produced by starting or stopping various pumps, revealed the largest variations in speed to be -3% to +25 t,h i / associated with increasing or decreasing load respectively. These step a changes were the sa.me order of magnitude of those calculated to occur during a LOOP /LOCA. The time lag associated with the unit's ability to return to 450 rpm was likewise found to be less than 3 seconds. Since speed variations associated with load step changes cannot be produced at full or two-hour overload power conditions due to grid connection, the modal superposition method was used to calculate the effects. The free-end vibration amplitude was first calculated by the model superposition method, and then the nominal torsional stress was calculated to be 9562 psi per degree of free-end rotation [2-2]. The T values used in the model super-n position analysis were assumed to be equal to those obtained at 3500 kW and 450 rpm. The maximum nominal torsional stresses at 428 rpm (957, rated speed) and also at 473 rpm (105% rated speed) have been calculated to equal the DEMA limit of 7000 psi within 3%, which reflects the uncertainty in the T values n at these speeds. Thus, within the accuracy of the analysis, compliance with DEMA is obtained. Furthermore, the very small potential time during which such conditions could actually occur with the demonstrated performance of the V governor precludes fatigue damage. 2-5

Section 2 References 2-1 Standard Practices for Low and Medium Speed Stationary Diesel and Gas Engines. Diesel Engine Manufacturers Association, 6th ed.,1972. 2-2 Yang, Roland, " Proposed Torsional and Lateral Critical Speed Analysis: Engine Numbers 74010/12 Delaval-Enterprise Engine Model DSR-48 3500 KW/4889 BHP at 450 RPM." Transamerica Delaval Inc., Engine and Compres-sor Division. Oakland, California, August 22, 1983. 2-3 Bercel, E., and Hall, J.R., " Field Test of Emergency Diesel Generator 103," Stone & Webster Engineering Corporation, April 1984. 2-4 Thomson, William T., Theory of Vibration with Applications. Second edi-tion, Prentice-Hall,1981. 2-5 Hartog Den, Mechanical Vibrations. Third edition, McGraw-Hill,1947. 2-6 Lloyd's Register of Shipping, Guidance Notes on Torsional Vibration Char-acteristics of Main and Auxillary 011 Engines. 27 Crai g, Roy R., Jr., Structural Dynamics: An Introduction to Computer Methods. Wiley, 1981. !O l v 2-6 ~

'I TABLE 2.1 $TIFFNESS AND INERTI AS FOR TDI HOLZER ANALYSIS Inertia Inertia Stiffness Location (ib. ft. sec2) (ft.Ib./ rad) i Front Gear 6.8 58.1 x 106 Cylinder No. 1 49.2 84.7 x 106 Cylinder No. 2 47.9 84.7 x 106 Cylinder No. 3 47.9 84.7 x 106 Cylinder No. 4 47.9 g4,7, Ins Cylinder No. 5 47.9 84.7 x 106 Cylinder No. 6 47.9 84.7 x 106 Cylinder No. 7 47.9

84. 7 x 106 Cylinder No. 8 50.1 76.9 x 106 Flywheel 1100.1 276.8 x 106 Generator 2650.4 i

f l l I G 2-7 ,_-.n---4 nn--,,,--,._..._,.._.v..,,, n ,,n,,.mn,,.,,...,,_,,.,,y,.,_-,,_,,,,,,_--,n.,,,,w._,,,.m,n,,- ,eg. ,,,.--we__.--

TABLE 2.2 TORSIONAL NATURAL FRE00ENCIES FROM TDI ANALYSIS Natural O Frequency (Hz) i 38.7 2 92*9 3 116.7 + lO l l 2-R 1

i TABLE 2.3 TORSIONAL LOADINGS FOR TDI ANALYSIS Order Torsianal Loading, Tn (psi) l l 1.5 129.5 I l 2.5 71.7 3.5 42.8 4.0 27.7 4.5 23.8 5.5 12.8 e t e i 2-9 3 n ,.,.. _... _.. - - _.. _. -.... _ _ _ _ _ _. _ _. _. - ~. - _ - _... _ _ _ _ _ _. _. _ _ _.... _ -,., - -... _.... _.... - _.. -.

O TABLE 2.4 SINGLE-ORDER NOMINAL SHEAR STRESSES FROM TDI ANALYSIS i Amplitude of Order flominal Shear Stress (psi) ( 1.5 1606 i l 2.5 1064 i l 3.5 452 l l 4.0 2980 4.5 565 S.5 10R0 DEMA Allowable l for Single Order 5000 lO l s [ i l l i 1 6 O 2-10 vr--,---, .,,n,---------- ._n,-,,,,,n-,,-,,-_ . - - - - - - - - - ~ ~ - - - - - -

TABLE 2.5 NOMINAL SHEAR STRESSES CALCULATED FROM SWEC TORSIGGRAPH TEST Order Amplitude of free-end Amplitude of Nominal rotation (degrees) Shear Stress (psi)* At 3500 kW At 3800 kW At 3500 kW At 3800 kW 1.5 0.171 0.187 1635 1788 2.5 0.130 0.140 1243 1339 3.5 0.058 0.061 555 584 4.0 0.325 0.339 3108 3242 4.5 0.064 0.067 612 643 5.5 0.127

0. 136 1214 1300 DEMA Allowable fer a Single Order 5000 5000 O

16 peak to peak 0.693 0.719 6626 6875 DEMA Allowable 1/2 peak to peak 7000 7000 Amplitude of nominal shear stress is calculated to be 9562 psi per degree of free-end rotational amplitude. O G/

l i 1 i I f I EE Front gear I / Crank no.1 // -4 2 M Crank no. 2 2 a. a S 6' M Crank no. 3 = 3 i - m j ,i y $,[3' Crank no. 4 3 eu o s F M Crank no. 5 t M S M Crank no. 6 ) 5 4 l g-Crank no. 7 5 a j 3 Crank no. O 1 t ( l Flywheel l m l I I. U i Generator e

3.0 FATIGUE ANALYSIS OF CRANKSHAFT v In Section 2.0 it was found that the replacement crankshaf ts satisfy the DEMA nominal stress recommendations for both 3500 kW and 3900 kW. The stresses for a single order were considerably below the 5000 psi that is recommended as an allowable. However, the stresses for combined orders were quite close to the 7000 psi that is recommended as an allowable. While the DEMA limits are believed to contain an intrinsic (though unspecified) safety margin, a fatigue analysis of the crankshaft was undertaken to determine the true margin. First, a dynamic torsional analysis of the crankshaft is performed to determine the true range of torque at each crank throw. This model is co :- pared with SWEC test data for the amplitudes raf free end vibration, measured with the torsiograph, and for range of torque near the flywheel, measured with strain gages on the shaft. Second, a finite element model of a one quarter crank throw is used to p compute the local stresses in the fillet region. Torsional and gas pressure b loading cases are considered. The results of this analysis are compared with strain gage test results, which were measured in the fillets of Crank Throw Nos. 5 and 7. Third, the fatigue endurance limit is established for the replacement crankshaft by first obtaining the endurance limit for the failed crankshaf ts, and then assessing the dif ferences between the failed and replacement crank-l shafts. The endurance limit is compared with values provided in the litera-l

ture, j

Finally, a factor of safety against fatigue failure is computed. 3.1 Crankshaft Dynamic Torsional Analysis 3.1.1 Torsional Model FaAA developed a dynamic torsional model of the crankshaf t to overcome limitations in TDI's conventional forced vibration calculations. For 3-1

instance, the TDI method does not compute the phase relationship between the various orders or modes, so it is not possible to compute the true summa-tion. The actual maximum stress is a direct result of this sununation. 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 ground to represent the effect of the grid on dynamic response during synchronous operation. This spring constant was found to be 1.409 x 106 ft.-lb./ radian based on generator specifications. This constant is set close to zero to represent SNPS emergency 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). The 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 l l repeat themselves every two revolutions of the crankshaft. 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, reciprocating inertia, and f rictional 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 [ robe through the air start valve. A top dead v 3-2

At, J center, TDC, mark for Cylinder No. 7 was simultaneously recorded by a probe on I the flywheel. The pressure data at 100% 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 mechani-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. 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 \\ / U approximately 820 lbs. This mass causes reciprocating inertia torque on the crankshaft. The effect of this torque was combined with the gas pressure l 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 l dividing by the piston area and throw radius. The resulting normalized l torqucs for the most significant orders are shown in Table 3.2. I 3.1.3 Comparison of Calculated Response With Test Data I 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]. O

V 3-3

() The calculated amplitude of free-end displacement is compared to the SWEC test measurements in Table 3.3. It is seen that the agreement is close for all significant orders. The vector sununation listed represents half the maximum peak-to-peak displacement range. The model also calculates the range of torque at each crank throw, which is shown along with the corresponding nominal shear stress (t = Tr/J) in Table 3.4 The computed torque range near the flywheel was found to be 312 ft-kips compared with the measured value of 357 ft-kips [3-2]. The strain gages were placed close to the flywheel hub, and thus were expected to give higher values. The apparent stress concentration factor is 1.14 The com-puted torque as a function of crank angle for each crank throw is shown in Figure 3-2. The computed and measured torques at the flywheel are shown as a function of crank angle in Figure 3-3. 3.2 Crankshaft Stress Analysis 3.2.1 Finite Element Model i The nominal crankshaft stress values calculated from the dynamic model are considerably less than the actual maximum stresses in the crankshaft. Those nominal values would prevail if the crankshaft were a long circular cylinder. Stresses in the real crankshaft are greatly influenced by its complex geometry and by stress concentrations, especially at the fillet radii between the main journal and web and the crankpin and web. In this section, l maximum stresses and their location are determined with particular attention i to the crankpin fillet. The multi-throw crankshaft under investigation consists of a series of crankpins and main journals interconnected through webs. Typical structural dimensions of one throw are shown in Figure 3-4. The main journal is 13.0 inches in diameter; the crankpin is 12.0 inches in diameter with a web thickness of 4.5 inches. Fillet details are also shown in Figure 3 4 The following material properties, corresponding to the A!SI 104? crankshaf t steel, were used in the analyses: 34 .. _. _ _ _ _. ~ _ -., _ _

Young's Modulus: E = 30.0 x 106 psi Poisson's Ratio: v = 0.3 A crankshaft throw is subjected to loads of two basic types: (1) torque transmitted through the throw, which is influenced by the output power level and by the torsional vibration response of the crankshaf t and (2) connecting rod forces applied to the crankpin and reacted at bearing supports. Linear elastic analyses were performed using the computer program MARC, K.1-1 Version, from MARC Analysis Research Corporation. Generation of the geometric input data and the post-processing graphics was performed using PATRAN-G and PATMAR developed by PDA Engineering. One throw of the crankshaft was analyzed by applying a static unit twist on the main journal. Since all throws are geometrically identical, a single model with appropriate loads and boundary conditions can be used to represent approximately any throw. Three existing planes of local symmetry were employed in the analysis to keep the O finite element model to a feasible size without compromising the accuracy of the results. These planes of symmetry are shown schematically in Figure 3-5 along with the coordinate system. The first plan of symetry is the vertical plane passing through mid sections of the crankpin, the web, and the main journal. The second and third planes of symetry are orthogonal to the first at the mid distance between two adjacent webs in the crankpin and the mair journal, respectively. Thus, only the portion of the crankpin, the web, and the main journal contained within these planes of symmetry was modeled. This model uses eight node, three dimensional, isoparametric brick elements with linear interpolation, capable of modeling an arbitrarily dis-torted cube. Each node in an element has three translational degrees of freedom. Because the state of stress in the vicinity of the fillet is of greatest interest and stress gradients are highest there, a finer mesh was used in this region. Figure 3-6 shows the three-dimensional model, with node and element numbers omitted for clarity, along with the coordinate system adopted in the model. O As adjustment to account for mesh refinement was obtained by comparing b 3-5

finite element stresses for a step shaf t with data reported from Peterson [3-3]. It was found that a factor of 1.08 needs to be applied to the finite element stresses. 3.2.2 Stresses Due to Torsional Loading There is no set of boundary conditions that can be applied to the model that will represent exactly the physical crankshaft under torsional loading. Therefore, two separate sets of boundary conditions (Table 3.5) were ana-lyzed. Boundary conditions for Case 1 represent antisymmetric behavior of the main journal to torsional loading in the axial (x) direction and those for Case 2, symmetric behavior. For both cases, transmitted torque was simulated by applying a unit rotation about the axis of the main journal in the third plane of symmetry of Figure 3-5. Figure 3-7 illustrates the relative crank throw orientations best approximated by each of the two boundary condition cases. Those crankpin fillets adjacent to a throw on the same side of the main journal and in the C same plane are best represented by Case 1 boundary conditions. Those adjacent to a throw on the opposite side of the main journal and in the same plane are more closely approximated by the Case 2 boundary conditions. Stresses in fillets not represented by either of these situations (i.e., adjacent to a throw not in the same plane) will fall between the two cases considered. Stresses obtained from applying the unit torsional rotation were scaled to represent maximum positive and negative torques of 251,600 and -144,600 ft. lbs. at Cylinder No. 5. These stresses were then scaled by a factor of 1.08 to account for the slight finice element underprediction of the fillet stresses, due to the size of the elements used. From the eight element integration points, stresses were extrapolated to the surface. For Case 1, Figure 3-8 shows the circumferential variation, and Figure 3-9 shows the axial variation of maximum principal stress for both the peak-positive and peak-negative torque conditions. Figures 3-10 and 3-11 show similar variation for Case 2. 3-6

(D V All stress values that have been presented are for the positive z side of the crankshaft, as viewed in Figure 3-4 In the crankpin fillet, this has also been designated as the O' to 180* portion. 3.2.3 Stresses Due to Gas Pressure Loading Near TDC the pressure in a cylinder causes a high vertical load on the crankpin. This load n.ay be calculated from the pressure loading and recipro-cating inertial loading. The pressure load is calculated from the area of the piston and peak pressure of 1680 psi. The reciprocating inertial load is obtained from the 820 lbs. of reciprocating weight and peak acceleration of 74.1 g. The reciprocating inertial load subtracts from the pressure load at TDC. The pressure loading was applied to the model as a distributed load on the topmost three lines of noded points on the crankpin. Two types of boun-dary conditions were applied. In both cases synsnetry planes 1 and 2 (see Figure 3-5) were modeled by symmetric boundary conditions. In the first case the third plane was modeled as a fixed support, and in the second case it was modeled as a pinned support. The actual moment in the main journal is greater than zero (pinned support) but less than the fixed-end moment (fixed support ). The moment in the main journal may be estimated by treating the crankshaf t as a continuous beam with simple supports at the main bearing loca-l tions. From this analysis it was determined that the moment in the main jour-l nal was 0.63 times the fixed-end moment. Since stresses for the fixed-end case were very small, the stresses due to the vertical loading were calculated as 0.37 times the stresses for the simply supported case. The maximum stress occurs in the 180' location and was found to be 15.5 ksi. The distribution of stress around the crankpin is shown in Figure 3.12. 3.2.4 Comparison of Stresses with Test Data The SWEC test [3-2] recorded data from strain gages in the fillets of Crank Throw Nos. 5 and 7. These gages were placed in the locations where p stresses are a maximum due to torsional loading. The measured stresses are compared with those calculated by the finite element model in Tables 3.6 and 3-7

O 3.7. Good agreement is found between the test data and computed results. The maximum principal stress range of 44.9 ksi was measured in Crank Throw No. 5 and it is bounded by the two finite element results of Case 1 and Case 2. At Crank Throw No. 7, the measured stress is slightly higher than the computed stress from Cases 1 and 2. The finite element stresses for vertical loading are in agreement with stresses measured in TDI's static test [3 4] on an inline 6-cylinder 13-inch by 11-inch crankshaft. TDI determined the maximum stress due to vertical loading, after factoring for the difference in crankpin area, to be 16.3 ksi at the 180* location. At this location, the torsional stresses are less than half of their maximum values. Also, at the location of maximum torsional stresses, the vertical bending stresses are measured to be 7.8 ksi and com-puted to be 8.1 ksi. At the No. 5 location, the transmitted torque is quite low during firing (see Figure 3-2), and thus, the highest stresses are not affected by vertical bending stresses. 3.3 Crankshaft Fatigue Failure Margin The factor of safety against fatigue failure in the replacement (12-inch crankpins) crankshafts is calculated in this section. The stress levels in the replacement crankshafts are computed from strain gage test data. The endurance limit is first established for the failed crankshaf ts (11-inch crankpins) from strain gage test data. This endurance limit is then scaled to account for the higher ultimate tensile strength of the replacement crankshaft. The effect of shot peening the replacement crankshaf ts provides an additional margin against fatigue failure. 3.3.1 Stresses in Replacement Crankshafts The replacement crankshaft was instrumented with strain gages in the fillet locations of Crankpin Nos. 5 and 7 and tested under oparational condi. tions at 3500 kW (100% rated load) and 450 rpm (100% rated speed). TN high-est stresses were measured in Crankpin No. 5. A dynamic modal of the crank-shaft confirms that this pin undergoes the greatest range of torque. Three-dimensional finite element models of a quarter crank throw show that the OU 3-8

( strain gage rosette was placed in the location of highest stress, both within the fillet and around the crankpin. The following strains were measured at 3500 kW: Strain Gage Maximum Winimum 5-1 (Compression) -195pc 288uc 5-2 (Bending) 695pc -410uc 5-3 (Tension) 737pc -610u c To account for the simultaneous ef fects of shear and bending, the stress state is represented by equivalent stresses using Sine's method [3-5]. For a biaxial stress state, the equivalent alternating stress, Sga' and equivalent mean stress S,, are given by: q S

  • (S qa ai y + Sd)l/

S S ~ and Sqm

  • Smi m2

+S o where S and S are the alternating camponents of principal stress, and S ai a2 q and S are the mean components of principal stress. From the test m2 report [3-2], the equivalent alternating stress, Sqa, and equivalent mean stress, Sqm, on Crankpin No. 5 were calculated to be: Sqa = 24.6 ksi S = 4.8 ksi qn Equivalent stresses, S and S are those alternating and mean uniaxial qa qm, stresses that con be expected to give the same life as the given multiaxial stresses. 3.3.2 Endurance Limit for Original 13-inch by 11-Inch Crankshaf t The original 13-inch by Li-inch crankshaf t was instrumented with strain gages in the fillet location of Crankpin No. 5. This fillet had previously a experienced a fatigue crack during performance testing. Af ter the test, the ) 3-9 1

r (v) three-dimensional finite element models of a quarter section of a crank throw showed that the strain gage location was placed close to the location of maximum stress. The measured stress range is used to establish the endurance limit in this analysis as a conservative assumption, although the actual maximum stress range is revealed by the finite element model to be about 15% higher at a nearby location. From the test report [3-6], the following strains were measured at 3500 kW: Strain Gage Maximum Minimum 5-1 (Tension) 1118uo -707uc 5-2 (Bending) 773pr -459uc 5-3 (Compression) -389p c 266,. c The equivalent alternating stress, S and equivalent mean stress, qa, Sqm, were calculated to be: Sqa = 33.7 ksi Q Sqm = 10.9 ksi From the test logs, it was determined that the shaft had experienced 273 nours at equal to or greater than 100% load, or about 4x 106 cycles. By using Miner's rule and typical slopes of S-N curves, it was determined that the endurance limit for this mean stress was 32.4 ksi. The ultimate tensile strength for these crankshaf ts averaged 96 ksi. A line representing this endurance limit is shown on the Goodman diagram [3-7] in Figure 3-13. This line is bounded by two lines showing the endurance limit for full scale crankshaf ts based on other test data [3-R]. 3.3.3 Endurance Limit for Replacement Crankshaf ts The repl a r.ement crankshafts have a minimum tested ultimate tensile strength of 103 ksi. The endorance limit scales linearly with ultimate ter.- sile strr!ngth. On this basis, the endurance limit for the replacement cran.- shaf ts is shown in Figure 3-13. OV 3-10

The fillet regions of the replacement crankshafts have been shot peened. The effect of shot peening may produce widely differing increases in fatigue endurance limit; however, a conservative range of values of this in-crease is from 5% to 20% [3-9]. The endurance limit for the replacement crankshafts, assuming a 20% increase from shot peening, is shown in Figure 3-13. 3.3.4 Factor of Safety Against Fatigue Failurt The factor of safety against fatigue failure of the replacement crank. shafts is 1.48 when the effect of shot peening is not considered. The effect of shot peening is to increase the apparent endurance limit. At 3800 kW, the strain gage test data [3-2] on the replacement crank-shaft shows that the stress level is 4% greater than it is at 3500 kW. At 3900 kW it would be about 5% greater than it is at 35M kW. Thus, there is an adequate safety margin against fatigue failure at the specified diesel genera-tor set two-hour-per-24-hour period rating of 3900 kW. O o LJ 3-11

O Section 3 References 'w/ 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 3-3 Peterson, R.E., Stress Concentration Factor. Wiley & Sons, New York, 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 1

3-6 Bercel, E., and Hall, J.R., "Fiele Test of Emergency Diesel Generator 101," Stone & Webster Engineering Corporation, October 1983. 3-7 Collins, J.A., Failure of Materials in Mechanical Design. Wiley,1981. 3-8 Nishihara, M., and Fukui, Y., " Fatigue Properties of Full Scale Forged and Cast Steel Crankshafts," Transactions of the Institute of Marine Engineering. Series B on Component Design for Highly Pressure-Charged Diesel Engines, Londor., January 1976. 3-9 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, Arnerican Society for Metals, April 1980. l l-t ( 3-12

TABLE 3.1 NATURAL FREQUENCIES FOR DSR-48 13-INCH BY 12-INCH CRANKSHAFT Mode Matural Frequency (Hz) 1 2.93* 2 38.73 3 92.94 4 116.67 5 184.33

  • For SNPS emergency bus operation the natural frequency of the first mode is zero (i.e.,

rigid body mode), and the natural frequencies of the higher modes are not significantly altered. O TABLE 3.2 TORSIONAL LOADING FOR Fa AA ANALYSIS Order Torsional Loading, Tn (P5;} 1.5 112.0 2.5 77.0 3.5 48.0 4.0 33.0 4.5 26.2 5.5 15.5 3-13

TABLE 3.3 FREE-END VIBRATION AT 100% LOAD FOR DSR-48 13-INCH BY 12-INCH CRANKSHAFT Amplitude of vibration (degrees) r FaAA Analysis SWEC Test [3-2] 0.5 0.065 0.056 1.0 0.001 0.005 1.5 0.177 0.171 2.0 0.000 0.001 2.5 0.142 0.130 3.0 0.001 0.001 3.5 0.061 0.058 4.0 0.340 0.325 4.5 0.069 0.064 5.0 0.031 0.034 5.5 0.122 0.127 6.0 0.014 0.008 6.5 0.014 0.016 7.0 0.002 0.002 7.5 0.001 8.0 0.015 Vector Summation 0.662 0.693 0 3-14

s TABLE 3.4 ~- TOROUE RANGE AT 100% LOAD FOR DSR-48 13-INCH BY 12-INCH CRANKSHAFT Amplitude of Torque Range Nominal Shear Location (ft. Ibs.) Stress (psi) 4th Order Total 4th Order Total Between Cylinder No. I and 36.6 x 103 167.1 x 103 648 2955 Cylinder No. 2 Between Cylinder No. 2 and 69.0 x 103 184.5 x 103 1220 3263 Cylinder No. 3 Between Cylinder No. 3 and 100.0 x 103 271.1 x 103 1768 4794 Cylinder No. 4 t' ~)

,/

Between Cylinder No. 4 and 129.0 x 103 309.8 x 103 2282 5479 Cylinder No. 5 Between Cylinder No. 5 and 155.6 x 103 396.2 x 103 2752 7006 Cylinder No. 6 Between Cylinder No. 6 and 178.8 x 103 327.3 x 103 3162 5788 Cylinder No. 7 Between Cylinder No. 7 and 198.6 x 103 329.7 x 103 3512 5830 Cylinder No. 8 Between Cylinder No. 8 and 214.2 x 103 311.8 x 103* 3792 5514 Flywheel SWEC test [3-2] computed the torque range to be 357.1 x 103 ft. lb. This indi-cates a stress concentration factor of 1.145 due to the proximity of the gage to the flywheel hub. v 3-15 0

\\ TABLE 3.5 DISPLACEMENT BOUNDARY CONDITIONS FOR TORSIONAL LOADING (REFER TO FIGURE 3-5) Case 1 flodal Degrees of Freedom X Y Z 1 Fixed Fixed Free 2 Free Fixed Fixed 3 Free Prescribed

  • Prescribed
  • O Case 2 Ilodal Degrees of Freedom p)

X Y Z 1 Fixed Fixed Free 2 Free Fixed Fixed 3 Fixed Prescribed

  • Prescribed *
  • Prescribed displacements were used on synrnetry plane 3 to simulate torsional load on the main journal.

O 3-16 _ ~ - - ...._. - - _ ~.

O O O Table 3.6 COMPARISON BETWEEN FINITE ELEMENT MODEL TORSIONAL LOADING REstiLTS AND TEST RESilLTS FOR PIN NilNBER 5 Peak Positive Torquel Peak negative Torque 2 Range of Range of Principal Stresses (ksi) Principal Stresses (ksi) Principal Equivalent Stress Stress 't "2

  • 1
  • 2 (ksi)

(ksi) Finite Element 20.7 -18.0 10.4 -11.9 32.6 52.9 i Case 1 l Finite Element 29.2 1.2 -0.7 16.8 46.0 45.1 Case 2 Strain Gage [4] 26.2 -2.9 4.9 -18.7 44.9 49.3 i Peak positive torque = 251.6 = 103 ft.-lb. 2 Peak negative torque = -144.6 = 103 ft. lb.

O O O Table 3.7 COMPARISON BETWEEN FINITE ELEMENT MODEL TORSIONAL LOADING RESULTS AND TEST REStiLTS FOR PIN NtfMBER 7 Peak Positive Torquel Peak IIegative Torque 2 Range of Range of i Principal Stresses (ksi) Principal Stresses (ksi) Principal Equivalent Stress Stress 'I

  • 2 "I

2 (ksi) (ksi) I i Finite Element 18.7 -16.3 7.3 -8.3 27.6 43.9 Case 1 l ) Finite Element 26.5 1.1 -0.5 11.8 38.3 37.5 i j Case 2 i Strain Gage [4] 23.4 -8.9 2.8 -14.1 37.5 44.5 i l I 1 Peak positive torque = 251.6 x 101 ft.-lb. 2 peak negative torque -144.6 e 103 ft.-lb. i 1 l l

i 1 l l i e PRE 550;E (c5ic, t ..u. I r i sx < .eu,l I .i' n e P le

  • 84 +

./ ,a n. \\ f / N. s. e: nz: ia: m C8 's-A N.~. E ' D r

c. E E S.

Figure 3-1. Measured pressure versus crank angle at 130'. load. l FeAA-84-3-16

l I [ V V O( I

  • t4El'GIM = 1IO O MAtlMUM = 173 0 2'

Ml41'9Pt = %F % 3' MI 4 IMIPt

  • F. I 9

'o "o 2 . 3 4 R n g. e a i i Yz Yg i to-1 1 De

== l o .a a g g r: r s v a. v. p. ,6. i,. n. z. i,. i i,. c su 1,. Cmeet AgCLE 10fCRFFSt renom A4Ctf fDf04Ff59 [ f a) Cylinder 1 to 2. b) Cylinder 2 to 3. l 1 g M4EIMUM = 207 0 MaulMU I = IPO O 3' Ml'elMUN = IOR 0 . R< MiselMUM = -808 0 -o O 1 7 2 X I t x ?< I G G I 1 i i Y Ya i C.a' L.a j w. w. o ,T' .? l .i -1 ) ** b g.s e 6.

  • '.,a
v. 3

. ~.. - 8. A. .',. n 1 i

i..

A. c~ . s. 3,. r aww agrer enf r.f t e. rgggw Avre f #DF, gf p e.e e c) Cylinder 3 to 4. d) Cylinder 4 to 5. t l Figure 3-2. Dynamic model torsional response at 100% load for Shoreham 13-inch by I?-inch crankshaft.

. _ _ _ _. -.. ~.. _. _ _.. O O O est a, .,u,,i s m, .. a.

    • t s *
  • wig..m

..s. i,,, 7 j [ e v., 1 ~* \\ m m,. n, -} + t-i U,n ..,, in .a. > .r a. . ~,,, ..a i..,, ..a,,. . s. i ., s. r,.,, swetr orret e s,.. = eg ava avri e perwree. t a j e). Cylinder 5 to 6. f) Cylinder 6 to 7. 4 t 1 o i 4 te s i am.,. gis a

  • ,4 t t,*t

/.sa a l t< .m a m i v i w *, mivis

.i i

1 n n 't l l 8 I { t l 4 a. f

  • =

Y., .u-u 1 = I \\ 1

u...

c s } ) l I i l Y ~ 7 -. - /: I. s....... l t g) r.ylinder 1 to 8. h) Cylinder 8 to flywheel. I ] l inure 3-? (<ontinued). l 4 i i w 3 -m -.-.m-

d b on y 357 f t-k ip s* 5 w 3 O E N y O y 1 I I I I l i 1 0.00 180.00 360.00 540.00 720.00 CR ANK ANGLE (degrees) a) Measured. i O 2 i s i = 311 f t-kips- ~ O Et O t i i i i i i i 1 0.00 180.00 360.00 540.00 720.00 CRANK ANGLE (degrees) b) Calculated. Figure 3-3. Comparison of measured and calculated torque near the flywheel.

  • See test for explanation of difference.

FsAA-84-3-16

= __ O ( t oos 0 l ? e a bK site a N .3/4 m

1. tars Detail A t,e,,

~ a> 1/16 -.- - fe j[ L: ce e e h. O e g o

e. e e o 2

e m4 4 e l l l i i t.s a - s ,r / e en - I \\ N, / I s g i -{ --f __M_ ) ,N / t l / f_ 'y l l i I l l I _ _ ___,I ' ~ ~~Y' ~~ ~ ~ i l Detail A i 8 \\ \\ l \\ f s \\ l ~# '" e sn I '._ i e l l .....a to S O figure 3-4. Typical structural dimensions of the crankshaft. FsAA-84-3-16

i O O .l 1 4 j First plane og symmetry + Web u l -Second plane / of symmetry i ,1 Main journal N Is. j N s N ) 1 f s w- 'N $ of c.renkshaf t , ', ', ',-k ( .Q k crankpin 27 " ;,.', <', i il s 1 \\ "' f " Crankpin 3 SE ,,s s : / r- ?. ,' /' i Third plane of symmetry ' / s \\ X i ) t' N i +Z i i .M i I 1 = 1 i u, \\ 1 I Iiqure 1 's. Typ u. a l ' ' f I"" ',, it;e < rankshaf t. showinq e nordinate axes and plane. or synetry. l l

-l i. O x n'% ,D~ 1 t x4 ? x N Q ', * ' / ys N' i-

  • /

v',\\ . w>A / 1 J \\ \\, hl_ ' ' l rrg. c !O .\\ .' \\ V.'. v' 's / / i t AN N i } i l Figure 3-6. Three-dimensional view of tne finite element model. 4 O FsAA-84-3-16 l L ~-

L Section,' F modeled 1 3 i (V i I i t i i I 1 q t d I (a)l Boundary Condition, Case 1 l I I I i 8 i I i l r l { d i _ Section _ ~ ~ ~ ~ [ modeled ~, (b) Boundary Condition, Case 2 Generator Go eene. End Enc Cyhnder 6 1 I I I i ,2r ,1 j, i i '2 i L I l_ I Cylinder 5 Cylinder 4 (1 denotes fillet location best approaimated b> Case 1 2 denotes fillet location best approximated t'> Case :) (c) Actual Figure 3-7. Schematic of relative crank throw orientations ft" (a) F.E. Model, Ease 1; (b) F.E. Model, Case 2; and (c) tN actual crankshaft. FaAA-84-3-16 L

50 i i i i i i i 3 %/ i l Web ' O.14 -*- 40 - 1. 0 I I I i a 30 F g e* Crank pin ~ l y f Fillet maximum W 20 - l E ~ E l" g 20 L_______ w O g Midsurf ace maximum en 0 wg Midsurf ace minimum C --.- -* ~ =.

e. - = =-

m -10r-Fillet minimum -20 0 20 40 60 80 100 120 140 160 180 200 220 ANGULAR POSITION (degrees) Figure 3-8. Circumferential variation of maximum principai 'cr torsion Case 1 boundary conditions. O F e A A -s e 3 16

I d 50 i i i i i 4 4 I Web l0.14+ l 40 - l i 1.0 i l I t [ 30 - o -1 Crank pin 8 i M Mw f 20 -- ~ l I f G E f Maximum stress 7 [ 10 - O s w c:2 I m 0 w 5 i g" l b \\ Minimum stress -10t- -20 O.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 20 SURFACE DIST ANCE (in ) I rigare 3-9. Asial variation of meitman principal stress for it" tic' Case 1 boundary conditions. O F e A A 8 4 3 16 .--.,__.,_m,_____..m..-_.y__._..--pw._,..y _.- _.yge,_,-m.,,,,-_,,.7--__,._%,-,y..-r_---._ --w-,.- ..--=,,v.,

30 i Peak positive fFillet torque conditions 20 O M 10 Mw E E d Crank pin midpointg a 0 l l l l ^ \\l l l l O W N-10 6 E D" Peak negative torque conditions Fille t ~20 O 20 40 40 80 100 120 140 160 180 ANGUL AR POSITION AROUND CRANK PIN (degrees) (0 is top of pin at TDC) I Figure 3 10. Circumferential variation of manins-principa' stress for torsion Cast 2 boundary conditions. O FaAA 84 3 1(

l 3 3 4 3 4 I I 1 i i ) Web 0.14 + I 40 k l - 1. 0 g I j 30 - T g ) Crank pin w mmf 20 " m .J{ Maximum stress W [ 10 w \\ m O W w tw Minimum stress -10 -20 O.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 2.2 SURFACE DISTANCE (in.) i l Figure 3 11. Axial variation of maximum principal stress for torsica Case 2 boundary conditions. O FeAA 84 3-16 _ _, _ _ _ -. - -. _ - _ _ _.. _.. _ _..... - _. _ _ _ _ _ _ _ _ _ _ _ _ _ _.. _. ~..

0 \\ J I i. I 20 I i i j r i i J l 1 4 I ~ 7, Fillet 5 10 ""*I*"*T i m us i E I 9-e Midsurf a ce j maximum .4.#,*, g" Fillet minimum ~~ i '~ j E i l g, y,,_ __ s. c.....------------ r ------ / \\Midsurf a c e E minimum l 4 i 1, l i -10 I I I I I O 45 90 135 180 = ANGULAR POSITION (degrees) 1 D. e ? Fiquro 1-17. ('ir< umferent ial variat ion of mavimum print ipal stress for qas pre',sure loading. t u e I I l

O O O 50 47.0 i 3 --20% increase in endurance limit due to shot peening 5 40 Factor of safety is 1.75 (s's- - --Endurance limit for repfscement crankshaft 39.2 2 t " 36.5 g w.thout shot peening w E Factor of safety is 1.48 9 s N ' 'Q \\ Stress endurance limit from test i n f ailed crankshaf t e 30 's \\ 's z 5 'ss\\ i s\\ z 1 g Stress s j from test onN\\ un j replacement N g, crankshaft 's s 20 ) ' s h' s I i Z t s data for other full scale %s l W Range of fatigue test j l j 'sg s i s crankshaf ts with UTS of Ws 10 's 1 = \\ l I UTS for failed crankshaf t - 8 I O I O 20 40 80 80 100 ' s 120 I EOUIVAl ENT MEAN STRESS (ksi) N UTS for repiscement i I cransk sha f t u \\ l I s qvar e-1-11 Gneniman diaqram f or-re plat e nent t ran6 <.ha f f

  • i I

I l

OG 4.0 DISCUSSION Als CONCLUSIONS DSR 48 engines with 13. inch by 12. inch crankshaf ts are in diesel gener. ator set service at seven other locations as shown in Table 4.1. This data 4 i shows that there has been extended service (long enough to produce more than 107 stress cycles) on several engines with RO% to 94% load, and limited ser. vice at 100% load. The fillet regions of Crankpin Nos. 5 through 8 of the Shorehan re. placement crankshaf ts were eddy current tested after 102 to 114 hours of operation at 100% or greater load (see Table 4.2). No relevant indications were found. Thus, there are no cracks in the high stresses fillet regions. The drawing of the replacement crankshaft has been certified by the American Bureau of Shipping for compliance with their rules [41] for sizing of the pins, journals, and webs. ] The following conclusions are made: l. The design calculations on the 13. inch by 12. inch crankshafts performed by TDI are appropriate and show that the crankshaf t stresses are below DEMA recommendations for a single order. Combined stress is not calculated by this method, but may be determined by torsiograph testing. 2. The SWEC torsiograph test results show that the 13. inch by 12 inch crankshaft stresses are below the DEMA recommended levels for both single order and combined orders for both 3500 kW (100% rated load) and 3P00 kW. A linear extrapolation to 3900 kW also shows compliance. 3. The factor of safety against fatigue failure was found to be 1.49 if the effect of shot peening the fillet regions is ignored and is even greater if the shot peening of the Shoreham crankshaf ts is j considered. 4. The replacement crankshaf ts are suitable for unlimited operation in the emergency diesel generators at SNPS. O I 41

Section 4 References 4-1 American Bureau of Shipping, Rules for Building and Classing Steel Vessels. New York, 1984 O O 4-2

b TABLE 4.1 l AVAILABLE LOGGED HollRS OF OPERATION OF DSR-48, RATEn 3500 kW 9 450 RPM i Kilowatt Total Average Serial Ratin9 9 Hours Date Load other Loads and l Number Location 450 rpm Logged Logged Reported Hours Reported l >3500 kW for 114 hrs, i 74010 SNPS 3500 36R 4-1-84 74011 430 4-1-84 >35n0 kW for 116 hrs. I 74012 345 4-1-84 >3500 kW for 11n hrs. 75005 K00$HENG, 3600 246 3-15-84 Mostly 75006 TAlWAN 221 3-15-84 100% l 75007 368 3-15-84 l 75008 299 3-15-84 76010

OHUBA, 3500 19800 3-17-84 76011 SAV0!

23300 3-17-84 76012 ARABIA 23800 3-17 84 76013 19700 3-17-84 76014 23500 3-17-84 3000/3200 kW for 9000 hrs. 76026

ONElZA, 3515 16204 3-17-84 76027 SAUDI 12428 3-17 84 76028 ARAB A 14978 3-17-84 78029 U.of 3500 0180 3-15-84 1100 kW 78030 TEXAS 5385 3-01-R4 1100 kW 78044 WADI 3515 10882 3-17-84 2200/3000 kW 78045
DAWASIR, 10832 3-17-84 2200/3000 kW 78046 S. ARABIA 11212 3-17-84 2200/3000 kW 3300 kW for 6200 hrs.

79002

RAFHA, 3515 12667 3-16-84 3200 kW for 825n hrs.

79003 SAUDI 11655 3-16 84 32n0 kW for 560n hrs. 79004 ARABIA 13186 3-16-84 A0001

RABIGH, 3515 10196 3-16.R4 2700 kW 80002 SAU01 10245 3-16-84 2800 kW 80003 ARABIA 11602 3-16-84 2800 kW O

4-3 l

I 1 TABLE 4.2 } N00R5 0F OPERAtl0N OF 5HOREHAM REPLACEMENT CRANK 5 HAFT 5 AT TIME OF E00Y CURRENT TESTING j NotIRS OF OPERAT10N l AT LOADS a 100". R ATED DIESEL GENERATOR AT ALL LOADS inAn y 101 36 8 114 ] 102 319 lop l i 103 34 5 110 l j 44

i r i l l f I l t I l t t l APPENDIX I i COMPONENT DESIGN REVIEW -l t l l t h l e l i i I ( l l b + I i i ,__n_.,.-

t/ nn.n3 31oA COMPONENT DES 18N REVltW CRAW 5 HAFT Classtf1catton A PART No. 03 310A Completion 3/5/R4 PRIMRV FUNCT10N: The crankshaft converts reciprocating motion, component inertial forces and gas pressure piston forces to rotary motion and torque at the output flange. FUNCTIONAL ATTRISUTis: 1. Structural stiffness of the crankshaf t must be suf ficient to maintain acceptable states of stress in the crank pin web and main journal areas and to maintain system natural frequencies which are suf fletently removed from engine operating speeds. The crankshaft design should also be sufficient to withstand normal main bearing misalignments inherent in service. 2. The journal area of the main and connecting rod (crank pin) hearing must be suf ficiently large for proper bearing oil film pressure but the journal length must be suff tetently short to prevent end wear of the bearing sleeves. (O,/ 3. The material of the crankshaf t and the surface finish should be sufficient to resist fatigue crack initiation, j SPECIFIED STANDAR0$t 1. IEEE i 2. ASTM 3. DEMA EVALUAfl0N: 1. Review TDI calculations and tests 2. Conduct engine test of 13 m 12 shaf t 3. Conduct modal superposition and Holzer torsional analyses of: MNS ((R 48) $NPS a. RV16) b. C. Midland (RV-12) d. $an Onofre (RV 20) 4 Conduct finite element analysis of R.4R 12.tnch cranbrin fillett i O 5. Compare measured and calculated stresses R.4n 13 a 12 shAf t b i

l [ l l 6. Compare measured and calculated output torque and free end torstograph l l traces for R 48. l 7. Compare stress levels with endurance limit for R.48 84 Compare nominal stresses of R.48 and RV.16 with those recommended by [ other standards, b. Compare nominal stresses of RV.12 and RV 20 with those recommended by various organizations. l 9. Complete final report on SNPS and GGNS crankshaf t integrity. 10. Complete final report on Midland RV.12 and San Onof re,1V.20 i REVIEW TDI ANALYSTS: l 1. Emperimental stress analysis (static) of D5R.46 crankshaf t t i 2. Torstograph tests 3. Holzer Table calculations l I INFORM 4Tl0N REQUIRED. 1. TOI drawings for 0$R.48 and RV engines O 2. Test reports for 05R.48 and RV engines i l 3. Original Holzer calculations and revisions for R.48 and RV.16. RV 12 and RV.20 engines 4a. Emperimental pressure vs. time curve for R.48 and RV.16 engines, b. Esperimental pressure vs. time curve for RV.12 and RV.20 engines. l l l l I l t l i l /N - - -}}

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