ML19210A608

Responds to NRC 771019 Request for Summary of Stress Analyses Performed for TMI-1 Decay Heat Pump Shafts Re 80 Gallon Per Minute Operation.Requests Submittal by 771102 Re Operation at 135,550 & 3000 Gallons Per Minute

ML19210A608
Person / Time
Site:	Three Mile Island
Issue date:	10/27/1977
From:	Herbein J METROPOLITAN EDISON CO.
To:	Reid R Office of Nuclear Reactor Regulation
References
GQL-1477, NUDOCS 7910300644
Download: ML19210A608 (7)

Text

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October 27, 1977

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4 y M.M.g , ?"r Directer of Nuclear Reector Regulation "1%I .%,rg i N *s t Attn: R. W. Reid, Chief operating Reactors 3 ranch :To. L f U. S. Nuclear Regulatory Commission C.%',g/ itV Washington, D.C.

20555.-Eear Sir: Three Mile Island Nuclear Station, Unit 1 (TMI-1)

Docket No. 50-289 Operatiae License :To. DFR-50 Decay Heat Pump Shafts During a telephone conversation on October 19, 1977, be.tveen the :TEC (Mr. Zwet::ig) and Met-Ed ("r. Stevens), the NRC asked Met-Ed to provide a summary of the stress analyses perfomed for the "MI-1 Decav Heat Pump Shafts. The NRC requested that the information applicable to 80 gpm operation be provided by October 26. 1977, and that information pertinent to operation at 135, 550, and 3,000 gom be provided by November 2, 1977.

The information given belov is primarily directed at answering the NRC questions regarding 80 gpm operation. However, information for the full f1cv range is also discussed to scme extent to show trends for loed and stress changes as a function of flow rate. Further information regarding operation at 135, 550, and 3,000 g n vill be provided in the ITovember 2,1977 submittal.

1.Steadv-State and Alternatine Loads for Bendine and Torsion a.By dine Loads (1) Steadv-State At a flow of 80 gpm, Worthington calculated a maximum steady-ste.te radial load of 260 pounds acting at the centerline of the impeller (which corresponds roughly to the end of the shaft). Loads for higher f1cvs vill be less than this value. Worthington indicated that these load calculations are based en codels developed by t'aem using test data for their pumos.

1492 154 779040044

..R. W. Reid, Chief October 27, 1977 GQL 1h77..(2) Alternating Worthington has indicated that there is no reliable methcd for accurately calculating actual alternating radial loads at lov flow rates such as 80 and 135 gpm.

Hevever, an upner beund was established by MPR by calculating the radial load required to bend the shaft enough for the i=peller and wear rings to contact.

This maximum load was calculated to be 670 pounds acting at the centeriine of the impeller.

b.Torsion (1)Ste ady-S tate The maxi =um steady-state torsional load occurs when the pump is drawing the maxi =um horsepcver,'..c., at the maximum pump flev rate being considered (3,000 gyn). The horsepower at 3,000 gpm is 3h0 HP; this corresponds to a torque of 11,900 lb-in.

At a flow rate of 80 gpm, the horsepcVer is approxtnately 150 HP, which co esponds to a torque of 5,250 lb-in.

(2)Alternating With typical pumps, alternating torques are norually about 1% of steady-state torques . This vould give very low torsional stresses, so an upper bound estimate of alternating torque was calculated by MPR by assuming that the full pump discharge head periodically acts across one impeller vane. This resulted in an upper bound alternating torque of 11,300 lb-in. , i.e. , essentially 100% of full power steady-state torque and is much higher than expected to realistically occur.

s It is assumed that these upper bound loads could occur at either 80 or 135 gpm. but that they should not occur at flows over 20% of the rated f1cv, i . e . , ov er c00 gpm.

2.Stress Concentration Facters a.Bending Per Page 108 of " Stress Concentration Design Factors", by R. E. Peterson, John Wiley and Sens, Inc. ,1953, the stress concentration factor for this type of keyvay in bending is 1.79 b.Torsion Per Figure 100 in " Stress Concentration Design Factors", the theoretical stress concentration factor in torsion for a sharp-edged keyway ccult. be over 5 Thererore, a fatigue strength reduction factor of 5 0 was uled, in accordance with ASME Section III Design Rules, which state that no fatigue strength reduction f actor greater than 5.0 need be used.

14r942 155 R. W. Reid, Chief October 27, 1977 GQL 1k77 3.Calculated Stresses for 80 zem Operation Using the above estimated loads and stress concentratien/ strength reduction factors , the following steady-state and alternating stresses are calculated to be acting at the keyway during operation at 80 gpm:

.The steady-state bending stress due to a steady 260 pound losd is a.1,270 psi.

b.The alternating bending stress due to a fluctuating 670 pound lead is 3,260 usi.

c.The steady-state torsional stresa due to a 5,250 lb-in. steady torque is 9,270 psi. 1/

d.The alternating torsional stress due to a torque varying between 0 and 11,300 of 9,980 psi). l'p 0 to 19,960 psi Ci.e. , an alternating torsional stress lb-in. varies frc h.Cemnarison of Alternatine Stresses with Faticue Endurance Limit The alternating bending and torsional stresses can be cc=bined, and result in an effective alternating stress intensity of 21,660 psi. 2/ Per the ARMCO catalog, the endurance limit for 17 h PH in the H1150 heat treat condition is 90,000 psi for both 107 and 100 cycles.Since the endurance limit does not change when 8 going from 107 to 10 cycles, it is reasonable to conclude that a true end limit has been reached, and that 90,000 psi also applies to cycles over 10grance cycles.Folleving standard ASME Section III practices, the allevable alternating stress range for this high cycle range would be a factor of 2 less than the best-fit tested value, i.e., vould be 90,000/2 = h5,000 psi.

Since the upper limit combined alternating stress of 21,660 psi is well belev this value, no fatigue damage is calculated to occur.

1/ " Formula for Stress and Strain", hth Edition, McGraw Hill, Inc. , Copy Right 1965, Page 19h, Chapter 9, Table 9, Raymond J. Reark.

2_/ " Formulas for Stress and Strain", hth Edition, McGraw Hill, Inc., Ray =ond J. Beark, Copy Right 196;, Ne 95, Table 2, Case No. 3 for axial stress combined with sheer.

1492 156 R. W. Reid, Chief h-October 27, 1977 G1L lh77 5.Vibration Analvses

'a'.Bending , In the pump design reviev =eeting of Septe=ber 20, 1977, Worthingtcr stated that the calculated critical speed for lateral vibrations is about 5,300 rp=.

This speed is =uch higher than the nc=inal rotating speed of 1,800 rp=, and no problems are expected due to letaral vibrations.

This was confirmed by the 1cv vibraticnal levels detected during recent pump tests, as previously reported to the NRC.

b.Tcrsion In the design reviev =eeting of Septe=ber 20, 1977, Worthington indicated that their experience has been that motor driven pu=ps which do not incluie a gear reducer are not subject to torsional vibration prob 1'=s.

Further, Worthington indicated that there are no realistic mechanisms for imposing significant torsional loads. Accordingly, torsional critical speed analyses have not been perfer=ed to date. Further, it is noted that the vibration measurements of the pump indicate that no significant vibrations are occuring. However, to provide further assurance that torsional critical speeds are not a problem, an analysis vill be initiated and is expected to be cc=pleted by Nove=ber 30, 1977 c.Allevance for Resenance Effects As discussed above, no specific allevance has been made for resonance effects. If the torsional critical speed is calculated to be close to sc=e driving frequency, then possible resonance effects will be evaluated.

In addition to the request for the su==ary of stress analysis provided seove five (5) questions were askad by the NRC during the telephone conversation of October 19, 1977 and are answered below:

1.Hor many spare shafts are en hand?

Three (3), two (2) with dccu=entation, one (1) without.

2.Do spare shafts have keyvays of the proper radii?

Yes, all three (3) shafts have 0.06 inch radii in the keyway.

3 (a) Would installation of spare shafts require shutdown?

Under present Technical Specification, yes because of the estimated time for shaft replacement.(b) What is the time required to replace a shaft and restore the pump to its proper condition?

Our best estimate is 70 hours8.101852e-4 days <br />0.0194 hours <br />1.157407e-4 weeks <br />2.6635e-5 months <br /> for one pump providing no problems are encountered.

1492\57 R. W. Reid, Chief October 27, 1977 GQL lh77 (c) Are we lookine at the scst expeditious ways to make replacement?

For example, pumo replacement?

Yes, several ways to replace the pump shafts have been considered including pu=p replacement. The answer to 3c. and b.

above are a result of considering these options.

h.Our Cetober 15, 1977 letter rafers to replication and hardness testing.

Is this for the installed or spare shafts?

It is for the installed shafts.

In our Istter of October 5,1977, G;L 1368, Met-Ed indicated its intention to investigate the feasibility of metallegraphic replication and hardnesc testing to verify the heat treated condition of the installed TMI Decay Heat Pump Shafts. We have scheduled, concurrent with the vibration and UT testing during the first week of November, the folleving tests to verify to the extent possible, the heat treated condition of the installed shafts.

a.A BHN Hardness Test vill be performed en bcth shafts at the pump coupling end.

b.A Resistivity Measurement Test vill also be conducted on both shafts with a "TEV0" tester.

c.A Magnetic Permeability - Eddy Current Tests vill be performed.

The results of the above three tests should give us a good indication of the heat treat condition of the existing shafts.

5 What is Met-Ed's specific proposal for shaft replacement, crifice removal, and " piggy back" operation?

Based on the results of the three tests described in k. above, Met-Ed vill evaluate both long term operation with the present shafts and the advisability of scheduling shaft replacement during the 1978 refueling. Since receval of the recirculation line orifices can be acec=clished without requiring unit shutacvn it has been scheduled during November as a routine maintenance item.

No change to the long term recirculation c7eling (piggy back) mode of operation is deemed necessary or planned as a result of this current analysis of the decay heat pump shaft adequacy.

777, Met-Ed cemmitted to As a result of a meeting with ti.s NRC en Septembe7

'1 acec=plish vibrational testing and ultrasonic inspect 1cn of the installed decay heat pump shafts on a monthly basis. This ce= nit =ent was documented in GQL 123h dated Septenber 8,1977. Subsequent testing and inspecticn results have all met

• the acceptance criteria agreed to by Met-Ed and the NRC.

Based on the excellent test and inspection results , information resulting frc= the evaluation of pump design and decay heat system design which is documented in GQL 1368 dated octeLsr 5,1977, and su==arized above, and discussions with the NRC by telephone on Octeuer 19, 1977 Met-Ed requests that the ec==itment be =cdified frc=

a monthly inspection and test as follevs:

\f\ .R.W. Reid, Chief- 6-October 27, 1977 GQL IL77.a.Conduct quarterly vibration testing and ultrascnic inspection with an -~ allevance of + 25% of the inspection interval. The inspection interval allevance has been established consistent with ?MI-l Technical Specification Section h and current regulatory practices. b.Conduct vibration testing and ultrasonic inspection when the decay heat pumps are shutdevn after accu =ulating in excess of one bour run time in any mode or ccmbination of modes. This inspection vill commence within five working days of pump shutdev1. This inspection frequency is to ec==ence with the first test being conducted during the first week in November. All paramete s of the vibration test and ultrasonic inspection except frequency of the inspec;ien are to remain *ke same as those specified in GQL 123h, dated September 8,1977. Sincerely,.l >p',s'j'[ / J. G. Herbein Vice president JGH:WEP:tas \k-}}