ML20045H756
| ML20045H756 | |
| Person / Time | |
|---|---|
| Site: | 05200003 |
| Issue date: | 07/07/1993 |
| From: | WESTINGHOUSE ELECTRIC COMPANY, DIV OF CBS CORP. |
| To: | |
| Shared Package | |
| ML19303F729 | List: |
| References | |
| WCAP-13759, NUDOCS 9307210231 | |
| Download: ML20045H756 (56) | |
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WESTINGHOUSE CLASS 3 UCAP-13759 WESTINGHOUSE PROPRIETARY CLASS 2 VERSION EXISTS AS WCAP-13758 HIGH INERTIA ROTOR TEST PHASE 3 REPORT E (C) WESTINGHOUSE ELECTRIC CORPORATION 19.9_3 A heense is reserved to the U.S. Govemment under mntract DEACO3-90SFto495-O WESTINGHOUSE PROPRIETARY CLASS 2 This document contains informaton propnetary to Westinghouse Electne Corporabon; it is submitted in contdence and is to be used eclely for the purpose for which it is lumished a -d retumed upon request. This document and such informabon is not to be reproduced, transmitted, disclosed or used otherwise in whole or in part without authorizabon of Wesbnghouse Eiectne Corporabon, Energy Systems Business Unit, subject to the legends contained hereof. GOVERNMENT LIMITED RIGHTS: (A) These data are subtrutted with hmited nghts under Govemment Contract No. DE-ACO3-90SF18495. These data may be reprodaced and used by the Govemment with the express krnitabon that they will not, without wntten permisoson of the Contractor, be used for purposes of - manufacturer nor dsclosed outade the Govemment; except that the Govemment may declose these data outade the Govemment for the followmg purposes, if any, provided that the Govemment makes such declosure subject to prohibition egenst further use and disdosure: (1) This 'propnetary data
- may be disdosed for evaluation purposee under the restncbons above.
(11) - The ' proprietary data' may be dsclosed to the Electric Power Reseant institute (EPRI), electric utility representatrves and their drect consultants, excludng droct cornmercial competitors, and the DOE Nabonal Laboratones under the prohibitons and restriebons above. (B) This notice shall be marked on any reproduchon of these data, in whole or in part. 00 WESTINGHOUSE CLASS 3 (NON PROPRIETARY) EPRI CONFIDENTIAUOBLIGATION NOTICES: NOTICE: 1E 20 3 O4 O5 O CATEGORY: AEB DC ODDeOF O O DOE CONTRACT DELIVERABLES (DELIVERED DATA) Subject to specified excephons, disclosure of this data is restricted unbl September 30,1995 or Design Certification under DOE contract DE-ACO3 90SF18495, whichever is later. Westinghouse Electric Corporation Energy Systems Business Unit Nuclear And Advanced Technology Division P.O. Box 355 Pittsburgh, Pennsylvania 15230 I i l
@ 1993 Westinghouse Electric Corporation
' All Rights Reserved ,
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WESTINGHOUSE CLASS 3 Contents Paoe Abstract................................................................................................ iv Lis t o f Tabl e s . .. . . . . . . . . . . . ..... . . . .. . . . . . . .. . . . . . . . . ... . . . . .. . .. . . .. . . . . . ... . . .. . . . . . . .. . .. . . .. . . v Lis t o f Fi g u re s . .. . .... .. . . .. . . . .. . .. .. .. . . .. . . .. . . .. .. .. . . .. .. .. . ... ... .. .. . ... . ... . .. . . . . . . . .. . .. . . . vi
- 1. I n tro d u cti o n . . . . . . .. . . . .. . .... . .. . .. . . .. . . .. . . .. . . .. . . .. .. .. . .. . . .. . . .. . . .. . . .. . . .. .. .. . . .. . .. . . . . . . .. . . 1-1
- 2. Summary.............................................................................................. 2-1 2.1 C o n cl u sio n s. .. . . .. . . ...... .. .. . .... . . .. . . . . . . .. . . .. .. .. . . .. . .. . . ... . ... . .. . . .. . . .. . . .. . . .. .. 2-1 2.2 R e co mme ndations ............................... ................. .. ... ..... ..... .. ..... . 2-2
- 3. Testing With Half Inch Radial Gap ...................... ................................. 3-1 3.1 Mod i ficatio ns ................................. .................. ......................... . . . 3-1 3.2 B e aring Test Re sults ............. ........................... ..................... ....... 3-3 3.3 Dis cu s s io n . . . . .. ...... . .... .. . .. . . ... . ... ... . . .. . . .. . ... .... . .... .. .. .. ... . .. .. .. .. . . . . . . . . . 3-7
- 4. Ackn owledg e me nts ........... ................ . ..................... ......................... 4-1
- 5. R e fe re n ce s . ... ....... . . .... ..... . ... .. .... .. .. . ....... . . .... ....... . . .. .. .. ..... . .... .. . .. . ... . . . . . 5-1 Tables............................................................................................ 6-1 Figures................................................................................... 7-1 Appendix - Bearing Test Data from All the Tests in Chro nolog ical Ord er............................................ ....................... ........ 81 i
t til
Abstract Phase 3 of the high inertia rotor project for the Advanced Passive 600-MW nuclear power plant has been completed. Loss measurements were made over a full range of speeds on the friction dynamometer and bearlhg test facility. These tests were successful in demonstrating a significant reduction in the losses due to replacement of the radial bearing pads by a half-inch radial clearance. Other testing showed that rotation direction and thrust load have little effect on the losses. Starting coefficient of friction and lubricating water film development speed were measured. Several recommendations have been made for further reduction of the losses by optimization of the design. t t iv
List of Tables Table 1 Thermocouples Connected to the Data Logger Table 2 Results of Loss Measurements for Forward Direction Tests with Half-Inch Radial Gap Table 3 Results of Loss Measurements for Reverse Direction Tests with Half-Inch Radial Gap
- Table 4 Results of Loss Measurements for Reverse Direction Tests at Different Thrust Loads with Half-Inch Radial Gap Table 5 Results of Loss Measurements for Forward Direction Tests at Different Thrust Loads with Half-inch Radial Gap Table 6 Summary of Power Losses at 1750 rpm with Half-inch Radial Gap Table 7 Effect of Replacement of Radial Beating Pads by Half-Inch Radial Gap on Averaged Power Losses at 1750 rpm Table 8 Summary of Power Losses at Different Speeds and Thrust Loads Table 9 increase _in Power Loss due to increasing Thrust Loa (
a.c 2 Table 10 Summary of Starting Torques and Speeds Table 11 Summary of Starts and Operating Time Table 12 Average Temperatures at Maximum Speed l l l l l l V 1
List of Figures Figure la Cylindrical shroud for producing a half-inch radial gap around the high inertia test rotor Figure ib Details of the cylindrical shroud Figure 2a AP600 high-inertia rotor and test housing with half-inch gap Figure 2b AP600 high-inertia journal and bearing assembly Figure 3 Cylindrical shroud for half-inch radial gap Figure 4 Installing the cylindrical shroud in the test housing Figure 5 Test rotor with Micarta spacer for transporting the rotor / housing erssembly to the test facility Figure 6 Schematic of thermocouple locations on the eleven thrust shoes Figure 7 Non-contacting displacement transducers mounted between the coupling and the thrust-loading cylinder Figure 8 Non-contacting displacement transducers mounted near the large support bearing housing Figure 9 Test housing with support framework holding the non-contactinc displacement transducers Figure 10 Test housing with instrumentation Figure 11 End view of test housing Figure 12 Optical tachometer Figure 13 Test data at 1754 rpm with half-inch radial gap Figure 14 Support bearing temperatures and displacement gauge readings for the test shown in Figure 13 Figure 15 Schematic plot of temperatures in the thrust bearing at 1754 rpm with half-inch radial gap vi
Figure 16 Variation of power losses from torque and speed and from flow and temperature rise with rotation speed and direction for tests with half inch radial gap Figure 17 Influence of thrust load on power loss for tests with half-inch radial gap - Figure 18 Examples of torque and speed oscilloscope traces for testing with 10000 lb thrust load. (Upper) Start No. 4. (Lower) Slow running after start No. 4.
- Figure 19 Influence of replacement of radial bearing pads by a half-inch radial gap [
on power loss Figure 20 Influence of rotation direction on the pressure distribution with radial position for tests with half-inch radial gap,1752 rpm forward and 1744 rpm reverse Figure 21 Influence of removal of radial bearing pads on the pressure distribution with radial position,1724 rpm with radial pads installed (Phase 2) and 1748 rpm with radial pads replaced by a half-inch radial gap (Phase 3) Figure 22 Variation of presure at two radial positions with speed showing influence of replacement of radial bearing pads by a half-inch radial gap t 1 ill vil
- 1. Introduction This project is part of an extensive program, which has the objective of developing the design for a greatly simplified pressurized water reactor (PWR) plant for power generation. Major improvements are anticipated in safety, licensing certainty, life-cycle cost, and construction schedule, compared to older designs. The overall program is intended to complete the conceptual development of an advanced passive 600-MWe PWR plant (referred to as the AP600). An important part of the program is the development of a high inertia reactor coolant pump for the AP600 plant.
The most effective way of providing flow during coastdown of a pump during a loss of power transient was to add rotational inertia to the shaft inside the pressure 2 boundary of the pump. The required inertia was found to be 5000 lb-ft to provide the needed coastdown time. To achieve this inertia with minimum drag loss a high inertia rotor at the impeller end incorporates a depleted uranium insert and also functions as the runner for the thrust bearing. Because of the size, unique construction, and high operating speed, manufacturing and testing of both the rotor and the thrust bearing were required. This report details the progress made in testing the high inertia rotor and bearing. A design change removed the radial bearing function from the high inertia rotor. The objective of the current testing was to measure the losses with the radial bearing pads removed and a cylindrical shroud installed to give an annular space with a radial gap of one-half Inch. The rotor, test bearing, and test facility, designed and built in Phase 1 of the ] project and modified in Phase 2, were used in the continuation. This report briefly j describes the further changes made in the test article and equipment. The results of I I bearing tests and loss measurements made to determine the effect of replacement of the 1-1
radial bearing pads by an annular space with a half-inch radial gap are presented and J discussed. 12
~, 7 l
1 l l
- 2. Summary l The seven radial bearing pads were removed from the test housing and were replaced by a continuous annular space having an average radial clearance of 0.512 in.
Dynamic analysis predicted that the high inertia tr rotor and shaft would continue to exhibit stable operation. The testing verified the prediction, the test facility remaining stable throughout the full speed range to 1761 rpm. Noncontacting displacement transducers were added to measure the relative radial positions of the rotor and housing. These transducers worked very well in providing information to enable the rotor to be kept well centered in the housing. The program was comp!stely successful in obtaining a large reduction in power losses with the removal of the radial bearing pads, as predicted prior to testing.
2.1 CONCLUSION
S a,c
- 1. With the radial bearing pads rerttoved and replaced by.a_ half inch radial gap, >
the powerloss at 1750 rpm was hp compared to hp with the pads.
- 2. The reduction la power loss when the radial pads were replaced by a half-inch a,c-radialgap was
- 3. The direction of rotation has no significant effect on the measured losses.
. ._ a,c
- 4. Increasing the thrust load from .
Ib increases the losses by less
~ ~
than two percent.
- 5. On startu with thrust loads of{p the coefficient _of friction of the thrust bearing is
- 6. On startup the water film began to fo_m1 at about rpm and was fully a,c developed by rpm with a
~
thrust load. 2-1 u
2.2 RECOMMENDATIONS Some of the following recommendations are directed toward optimizing the design for further reduction in power loss. These recommended tests can be performed as desired by the design organization. b E 1 e 22
- 3. Testing With Half-inch Radial Gap 3.1 MODIFICATIONS The stainless-steel-encased depleted uranium flywheel and the thrust bearing are shown in EMD Drawing No. 3D19195, Rev.1 and are fully described in Section 3 of the Phase 1 report.(1) A design change removed the radial bearing from the high inertia flywheel to an adjacent shaft location. To incorporate this change, the test rotor / housing assembly was removed from the test facility following the procedures described in Ref.
(1). Following disassembly and visual inspection, the seven radial bearing pads were removed from the test housing. All parts were carefully cleaned and reassembled. The flow plugs in the rotor (item 21 of EMD Drawing 19195) were reinstalled. The bumper plate (Item 8 of EMD Drawing 19195) was also reinstalled. These items had been removed as part of the testing performed in Phase 2.(2) A cylindrical shroud, provided by NATD, was installed in the radial housing to provide a continuous annular space with a radial gap of about a half inch. The selection of a half inch for the gap was based on EMD analysis that predicts an optimum gap.(3) This shroud is shown in Figures 1a and 1b. The inside diameter of the shroud is 29.00 in.,as shown in the assembly view in Figure 2a. Since the diameter of the high inertia rotor is 27.975 in., the radial gap with the rotor and levasig centered is 0.512 in. Various cutouts in the cylindrical shroud were needed to provide acccas for pressure taps, thermocouples, and the water flow outlet, drain, bypass, and relief valve. For comparison, Figure 2b shows the joumal and radial bearing assembly as used in Phases 1 and 2. Figure 3 shows a photograph of the completed shroud. The shroud was insta!!ad in the test housing as shown in Figure 4. It was held in both axial and circumferential alignment by the seven pivot pins (originally used for the radial bearing). These pins engaged the position pads shown as Detail B in Figure 2. Radial and angular alignment and support were provided by the pivot pins, a close fit 3-1
l between the shroud and the housing at the thrust bearing end, and machined shims l between the shroud and the housing at the bumper plate end, i The test facility is described and I!lustrated in Section 4 of the Phase 1 report.(1) Subsequent modifications made to the testiacility are described and illustrated in Sections 3.1,3.2, and 4.1 of the Phase 2 report.(2) Several further improvements were incorporated into the test rig in the present phase. The shaft was leveled by changing the support bearing pivot pins. This ensured that the rotor and housing axes were properly aligned. The support bearing preload was set to a nominal value of zero at ; room temperature to avoid overheating the support bearings at high running speed. l Extensive computer analyses were performed to ensure that the rotor and shaft would remain in stable operation at all speeds with the radial bearing removed and the support bearing preload reduced. The drive motor controller was tested and repaired to eliminate the oscillations noted at certain speeds during the earlier testing. In the prior portions of this program, the test housing had rested on and was positioned by the radial bearing pads during movement of the rotor / housing assembly from the assembly frame to the test rig. Since these pads were removed for the current testing, an altemate arrangement was needed. (The housing could not rest on the rotor because the water seal could not accommodate the necessary haC-inch radial motion and because of the possibility of denting the cylindrical shroud.) A Micarta spacer was built to fill the radial gap, supporting and positioning the housing during the moving of the rotor / housing assembly. This spacer is shown in Figure 5. The removal of the seven radial bearing pads removed the associated radial pad thermocouples, which were numbered 1 to 28. To avoid confusion, the remaining > test bearing thermocouples (shown schematically in Figure 6) were not renumbered. Thermocouples were installed at four axial locations, 1.00,8.25,9.25, and 16.50 in. from the thrust bearing end of the cylindrical shroud, as shown in Figure ia. Detail C. A thermocouple was installed to measure the temperature of the water that leaks past the water seal to enable this information to be used to correct the power calculations based on flow rate and temperature rise. Table 1 lists all the thermocouples connected to the data logging computer. 3-2
Eight noncontacting displacement transducers were installed to monitor the positions of the rotor and housing. Four were used on the shaft in horizontal and vertical orientations at two widely separated axiallocations. The other four were similarly disposed on the test housing. Figure 7 shows the displacement transducers used to measure shaft position between the drive motor coupling and the thrust loading cylinder. The top of the coup!!ng guard has been removed for the photograph. The other shaft-sensing transducers were mounted near the large support bearing housing as shown in Figure 8. The horizontal-direction transducers for the housing were mounted on a support framework straddling the right-side radial loader as shown in Figure 9. The vertical transducers for the housing were on separate mounts under the housing. Dynamic analysis was used extensively in the design of the transducer mounts to avoid any resonances throughout the range of running speeds. Two pressure transducers were replaced by improved models to increase their stability and accuracy. Other instrumentation on the housing remained the same es that used previously. Figure 10 shows the instrumentation on the test housing. An end view of the test housing is shown in Figure 11. The two radial loaders, located at 45 deg positions on the lower half of the housing, were used to position the housing both horizontally and vertically, keeping it centered on the rotor. A digital optical tachometer was installed on the drive motor as shown in Figure 12 to provkfe hurate speed measurements, independent from the motor drive controller. The data logging computer program was modified in accordance with all the changes in instnJmentation. From the eight displacement transducer values the computer calculated the relative positions of the rotor and housing and provided updated values of the horizontal and vertical gaps between the rotor and the cylindrical shroud. This information was used to adjust the housing position to keep the gap uniform throughout the testing. 3.2 BEARING TEST RESULTS , Following reassembly of the housing and rotor on the test facility all instrumentation was recalibrated. The noncontacting displacement transducers were adjusted to agree with physical measurements of the radial gap with the housing in 3-3 1
centered and off center positions. All subsystems were checked for proper operation. All computer calculations were checked and the program placed under configuration
'I control. Trials were made over the full range of operating speeds with fast Fourier transform equipment to ensure that the rotcr was stable.
A typical computer printout of test data is shown in Figure 13 for a forward direction test at 1754 rpm. The data file name or test number is a four or five digit test . Identification number. The first digit is the program phase. The second is the direction, O for forward and 1 for reverse. As before, the designations forward and reverse are arbitrary, with forward corresponding to clockwise rotation of the rotor when viewed from the bumper plate end. The third digit generally represents speed. The last one or two digits represent an index number for repeats of the test condition. Bearing test data printouts from all the tests in chronological order are given in the appendix. Figure 14 shows a similar printout of auxiliary data, mainly support bearing - temperatures and displaceinent transducer readings and calculated positions. This information was useful during the operation of the test facility. Figure 15 shows a schematic diagram of the thrust bearing temperatures along with other information. Four sets of test runs were made, each covering the full range of speeds, two sets in each rotation direction. Table 2 gives the power loss measurements for the first _ a,c set of runs at all speeds in the forward direction at about : lb thrust load. (This thrust
~
load is caused by the 50 psl water pressure acting on the 12In. diameter shaft anc represents the practical minimum thrust load safely attainable with the current configuration of the test facility.) At speeds above 600 rpm each test condition was held long enough for thermal equilibrium to be approached. This permitted the power losses to be determined independently from either torque and speed measurements or from water flow rate and temperature rise measurements. Table 3 gives corresponding results for tests in the reverse direction. Tables 4 and 5 give the results of two more series tests at three different thrust loads each at three different speeds in each rotation dimetion. A maximum speed of 1754 rpm was attained in the forward direction and 1761 rpm in reverse. 3-4
a,c The power losses from all the tests at about Ib thrust load are shown in Figure 16, plotted on log-log coordinates. Results are shown by different symbols for the forward and reverse directions and for power based on two different sets of measurements - torque and speed or water flow rate and temperature rise. The results for each direction and method of measurement form excellent straight lines when fitted by least squares regression. As shown in the figure, the losses measured by flow and temperature rise agreed very well in the two rotation directions. However, the losses measured by torque and speed were substantially higher in the forward direction than in reverse. The exact cause of the difference is unknown, perhaps some small pressure differential acting on the housing and affecting the torque measurement. This possibility was anticipated in the design of the test facility and is the reason for being able to rotate in each direction. The average of the two directions gives the correct loss measurement. Loss measurements based on torque and speed at about 600 rpm were excluded from the regression, because their low magnitude made them more subject to error in an unweighted regression. For measurements in the forward direction the remaining 31 points comprising the dashed line had an R2 og The points at a particular speed were so repeatable that they fall on top of each other in the pot. Interpolation to a speed of 1750 rpm resulted in a projected power loss of : hp, with loss varying with the { jpower of speed. For the reverse direction the 30 points (above 2 : Interpolation to 1750 rpm 612 rpm) making up t_he. dot-dash curve had an R of_ _ __ gave a power loss of hp, with loss varying as the power of speed. The average {_ _ __ loss for the two directions at 1750 rpm was hp. The equation for the average power
~~
loss vs. speed based on torque and speed measurements for the tests with a half-inch radial gap is: a,C , HP = _ (1) Loss measurements based on flow and temperature at about 600 rpm were excluded from the regression, because there was not enough time for them to come to a,c 2 thermal equilibrium. The remaining 31 points in the forward direction had an R og { 3-5
a,c Interpolation to a speed of 1750 rpm resulted in a projected power loss of hp, with ! loss varying with the : power cf speed. For the reverse direction the 30 points above
~
612 rpm had an R o{ 2 Interpolation to 1750 rpm gave a power loss of [ ~hp, with loss varying as theI __ t_
;power of speed. The average loss for the two directions at 1750 i rpm was hp with negligible effect of rotation direction. The equation for the average power loss vs. speed based on flow rate and temperature difference measurements for the tests with a half-inch radial gap is:
_ a,c , (2) HP=_ _ Table 6 summarizes these results and also includes power losses calculated from the drive motor armature voltage and current. These electrical power measurements include motor windage and support bearing losses as well as the losses in the test article, it is important to note that the electrical measurements confirm that ! rotation direction had no effect on losses. The average value of power loss in the two directions from the torque and spoed measurements agrees very closely with the average determined from flow and temperature rise. This agreement gives additional confidence in the accuracy of the averaged results. The results from the tests with a half-inch radial gap can be compared to the results from the Phase 2, Task 2, Part 2 testing.(2) The difference shows the effect on a,c power losses of replacing the seven radial bearing pads by a half inch radial gap. Table _ 7 shows this comparison at 1750 rpm. The power loss difference ranges from hp, showing very good agreement among the three methods of measurement. This agreement gives further confirmation of the accuracy of the averaged torque / speed measurements. Two of the sets of test runs were made specifically to determine the influence of thrust load on losses. -Tests were made at three speeds (about 600,1200, and 1750 rpm) and at three thrust loads in each direction of rotation. The results of these rlu s are listed in Tables 4 and 5. The losses based on l speed and torque are plotted in Figure 17. The nearly horizontal lines for loss vs. thrust f 3-6
load indicate the small influence found. The results for each speed and thrust load (six tests - three repeats in each direction) were averaged as shown in Table 8. Averaged results based on drive motor armature voltage and current were included. The increased power loss due to increasing thrust load fromi Ib is shown in Table 9. At L - _ an average speed of 1757 rpm these loss increases ranged from hp for the - different methods of measurement, with an average of _g 7 . This hp increbe amounts to less than two percent of the total power loss at this speed. Several additional starts were performed to measure the starting torque and the speed at which a lubricating film of water forms on the thrust bearing. The torque and speed signals were fed to an oscilloscope for these measurements. Typical starting traces are shown in the upper part of Figure 18. Beginning at zero torque and speed, as speed increases the torque rises to a maximum and then drops essentially to zero as the bearing friction decreases. Table 10 summarizes to results for tests with thrust loads of 1 ' !b. The coefficient of friction was calculated by dividing the torque by a,c
~
tlie thrust load and then dividing by 0.847 ft (the radius to the center of the thrust shoes). The average coefficient of friction was aboutI ]as expected. The speed where the torque first began to drop (indicating the begirining of water film build-up) was about seven rpm at each thrust load. The speed at minimum torque (indicating full film development) averaged { pm at Ib thrustload and rpm at
]Ib.
Trials were made to see how slowly the drive motor and controller could be
-'I operated without oscillations. Operation remained stable with Ib thrustload at speeds down to 80 rpm. Below that point oscillations occurred as shown in the lower part of Figure 18. Thus, slow speed testing below the speed of lubricating film formation will require a special low-speed drive. During the downward speed excursions, it was a,c interesting to note that the average speed at which torque began to increase was rpm, ,
~
indicating the point at which the water film began to break down. 3.3 DISCUSSION The effect of removing the seven radial bearing pads and replacing them with a half-inch radial gap is shown graphically in Figure 19. These curves illustrate the 3-7
equations based on water flow and temperature rise measurements for each case, Equation (4.4) from Ref (2) and Equation (2) from this report. A linear scale was used to emphasize the effect of speed on losses and to show the effect of pad removal consistently over the entire speed range, up to 1800 rpm. A summary of starts and operating time for the high inertia rotor and thrust bearing is given in Table 11. This list includes a!! testing in both directions of rotation in all three phases of the project. Each start is listed only at the approximate maximum speed attained. As noted, there were a total of 59 starts,21 of which attained at least 1700 rpm. Total running time was just over 110 hours,14.4 of which were at 1700 rpm or greater. The effect of rotation direction on pressure distribution with radial position on the housing shroud and bumper plate is shown in Figure 20. These pressures were measured at maximum operating speed for each direction. A sketch illustrates the gap or clearance between the rotor and the bumper plate or housing. In general, reverse direction rotation resulted in lower pressures at and inboard of the bumper plate. The difference in pressure drop between forward and reverse rotation was small, however. ; Figure 21 shows the effect on the pressure distribution of removing the radial pads and replacing them with a half-inch radial gap. The solid curve represents the average values for the two directions shown in Figure 20. The dashed curve shows the same information for the Phase 2 testing with the radial bearing pads insta!!ed. While there is a difference in the average speed attained in the two phases, the speed difference accounts for only a small part of the difference due to removal of the pads. The effect of speed on pressure is shown in Figure 22 at two radial positions. The influence of replacement of the radial bearing pads by a half inch radial gap is shown by comparison of the solid and dashed curves in the figure. The two transducer positions were selected to be near the center of rotation (R=1.5 in.) and at the point of greatest difference in the results (at the bumper plate, R=8.965 in.). Table 12 summarizes the average temperatures at maximum speed for tests with radial bearing pads and tests with the pads replaced by a half-inch radial gap. The difference in water inlet temperature was simply a result of the testing being conducted in 3-8
different seasons. The water outlet temperature was a result of the water flow rate used in each test. The average shroud temperature in the current tests was the average of Thermocouples 26 and 27 near the axial centerline of the shroud. This temperature was slightly higher than the water outlet temperature, whereas the average radial pad temperature was slightly lower. On the bumper plate end of the test chamber, removal of the radial pads resulted in higher temperature at inboard locations and lower temperature in the grooves of the bumper plate. At this point in the high inertia rotor testing program, it is useful to review some of the prior results, conclusions, and recommendations (1,2) to determine which are still valid and which should be modified. As discussed in Task 2 of the Phase 2 report (2), the speed measurements in Phase 1 were found to be inaccurate, especially in reverse. Thus, the losses reported (1) for Phase 1 are also not accurate. The testing at 47% preload was repeated in Phase 2 with greatly improved accuracy. In Phase 1 both radial load and thrust load were shown to have only a minor effect on losses. The small effect of thrust load has been confirmed by the current testing. The small effect of radial load is also believed to be correct.
~
From the Phase 1 report (1) Conclusions 1,2, and 3 are still valid. The high inertia rotor concept was verified by constructing and testing an encapsulated depleted-uranium rotor. Its operation as a combined joumal and thrust bearing was demonstrated successfully. And, the associated losses were higher than expected. Conclusion 4, relating to possible construction changes for reduction of losses, is still true, but the amount of improvement is not likely to be as large with the current configuration (without radial pads). Recommendations 2,3, and G from the Phase 1 report (1), relating to instrumentation improvements and more fundamental testing, have already been successfully completed. Recommendations 1 and 4 for smooth end caps were ! attempted in Phase 2, but are no longer expected to result in a large improvement. The portion of Recommendation 5 dealing with optimizing bearing performance is no longer i relevant to the current design, but will be essential for any future design that needs to use l i 3-9 i l 1
the flywheel as a journal bearing. The demonstration of 500 starts and stops in Recommendation 5 is still a valid recommendation. The results, conclusions, and recommendations found in Task 2 of the Phase 2 testing are all still valid. The current testing has shown that rotation direction can affect the torque measurement, requiring that such measurements be averaged for the two - rotation directions. i 3-10
- 4. Acknowledgements The authors wish to acknowledge the important contributions made by many individuals and organizations toward the successful completion of the testing of the high ,
inertia rotor. This project is part of the AP600 program sponsored by the U.S. Department of Energy under Contract No. DE-AC03-86SF10638. Technical direction of the project within Westinghouse was provided by D. E. Ekeroth, from the Westinghouse Nuclear and Advanced Technology Division. The continued interest and advice by L Veronesi and D. R. Nixon of the Westinghouse Electro Mechanical Division have been very valuable. They had designed and supervised the manufacturing of the high-inertia joumal and the test bearings for Phase 1. Conceptual design of the cylindrical shroud was performed by D. R. Nixon, based on A. E. Reed's analysis of optimization of the flywheel geometry to minimize losses.(3) Detailed design of the cylindrical shroud was performed by D. E. Ekeroth. Fabrication of the cylindrical shroud was done by Penn State Tool & Die Corp. L L Ross operated the computer and data logger during testing. L F. Wagner performed the rotor dynamics analyses and R. D. Holm, the vibration analyses of the noncontacting displacement transducer mounts. The technicians during modifications, assembly, and testing were E. J. Trax, F. A. Kramer, and R. L Conroy. R. L Johnson was in charge of the modifications, assembly, testing, and analysis of results, under the supervision of A. A. Raimondi. l i I i 4-1
- 5. References
- 1. R. L Johnson and A. A. Raimondi, *AP600 High inertia Rotor Testing - Phase .
1 Test Report," Westinghouse Electric Corporation Energy Systems Report . NSE 90-0093, March 1990.
- 2. R. L Johnson and A. A. Raimondi,"AP600 High inertia Rotor Testing - Phase l- 2 Report," Westinghouse Electric Corporation Energy Systems Report WCAP-13319, August 26,1991.
- 3. A. E. Raed, AP600 Flywheel Geometry Optimization for Minimum Losses, WEMD E.M. 6512, February 25,1991.
1 i 1 1 l 1 ! o 5-1
Table 1 - Them couples Connected to the Data Logger b 6-1
. ._i _ J l
Table 2 - Results of Loss Measurements for Forward Direction Tests with Half-Inch Radial Gap b 6-2 .
Table 3 - Results of Loss Measurements for Reverse Direction Tests with Half-Inch Radial Cap b e 6-3
- 3. .
i-Table 4 - Results of Loss Measurements for Reverse Direction Tests at Different Thrust Loads with Half-Inch Radial Gap i b i e t 6-4
Table 5 - Results of Loss Measurements for Forward Direction Tests at Different Thrust Loads with Half-Inch Radial Cap b me 6-5
i
.i i
i L , Table 6 - Summary o' Power Losses at 1750 rpm with Half-Inch Radial Cap'- - n 2 J i
't Table 7 - Effect of Replacement of Radial Bearing Pads by Half-Inch Radial cap on Averaged Power Losses at 1750 rps b
'i
+
r b 6-6
Table 8 - Summary of Power Losses at Different Speeds and Thrust Loads
.b l'
Table 9 _ Increase in Power _ Loss due to Increasing Thrust Load from lb. a,c b; 9 V f f 6-7
- g. v a
Fi : -l Y Table 10 '- Summary of Starting Torques and Speeds: - b .- .
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.t Table 11 - Susaary of Starts and' Operating Time ,
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Table 12 - Average Temperatures at Maximum Speed b e i 6-9
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N v EE i4.3. eDe 4# % ll.. .' 45* ngh Inerta Rotoe Half Inch Gep Shroud 17.500!l$ sheet I of 2 R030*(h- D.Cheroth (*29.00) 4/3/92~ DETAIL C Figure la - Cylindrical shroud for producing a half-luch radial gap around the high inertia test rotor. 7 _ ._- - .. _ _ .- -.. u .-.- _ u _ .a .. _ _-. -- - _ . _ . - - _. -. - ~.
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THERM 0 COUPLE' LOCATIONS a.c Figure 6 - Schematic of thermocouple locations on the eleven thrust shoes. 7-8
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h J 7 6L . . - Figure 8 - Non-contacting displacement transducers mounted near the large support bearing housing. 7-10
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b L Figure 13 - Test data at 1754 rpm with half-inch radial gap. 7-15
b, 1 Figure 14 - Support bearing temperatures and displacement gauge readings for~the test shown in Figure 13. 7-16
b-Figure 15 -- Schematic plot of temperatures in the thrust bearing at 1754 rpm with half-inch radial gap. , 7-17 e M+
b e Figure 16 - Variation of power losses from torque and speed and f rom flow and temperature rise with rotating speed and direction for tests with half-inch radial gap. 7-18
g I Figure 17 - Influence of thrust load on power loss for tests with half-- '1 inch radial gap. 7-19
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Figure 18 - Examples of torque an_d speed oscilloscope traces '.for a,c , testing with . thrust load. (Upper) Start No. 4. (Lower) Slow running after start No. 4. q 7-20 l
..s b-l 1 Figure 19 - Influence of replacement of radial bearing pads by a half-inch radial gap on power loss. 7-21
Lb i 9 4 Figure 20 - Influence of. rotation' direction on the pressure distribution with radial perit ion
' for tests with half-inch radini gap,1752 rpa forned and 1744 rpm reverse. ,
g - -- . . _ . - . . . _ _ .
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.s Figure 21 - Influence of- removal of radial bearing pads on the pressure distribution ~ with radial position,1724 rps with radial pads installed (Phase 2) and 1748 rps with radial pads replaced by a half-inch radial gap' (Phase 3).
7.-23 ' . . . . _ _ . . - . . _ . . . - - - - ._- . . ._ . . . ._ . ~ _ . - , .. . .- _ .
b i Figure 22 - Variation of pressure at two radial positions with speed showing influence of replacement of radial bearing pads by. a half-inch radial gap. 7-24 .j
L PAGES 8-1 TIIRU 8-115 CONTAIN
. APPENDIX BEARING TEST DATA FROM ALL TIIE TESTS IN CIIRONOLOGICAL ORDER.
TIIESE ARE PROPRIETARY TO WESTINGIlOUSE ELECTRIC CORPORATION l 1-8-1 _}}