AGMA 95FTM6-1995 Development of an Epicyclic Gearbox for Reduced Sub-Synchronous Vibrations in Gas Turbo-Generator Sets《要发展成为一个行星齿轮箱 为降低次同步振动的天然气涡轮发电机组》.pdf

《AGMA 95FTM6-1995 Development of an Epicyclic Gearbox for Reduced Sub-Synchronous Vibrations in Gas Turbo-Generator Sets《要发展成为一个行星齿轮箱 为降低次同步振动的天然气涡轮发电机组》.pdf》由会员分享，可在线阅读，更多相关《AGMA 95FTM6-1995 Development of an Epicyclic Gearbox for Reduced Sub-Synchronous Vibrations in Gas Turbo-Generator Sets《要发展成为一个行星齿轮箱 为降低次同步振动的天然气涡轮发电机组》.pdf（9页珍藏版）》请在麦多课文档分享上搜索。

1、95FTM6Development of an Epicyclic Gearboxfor Reduced Sub-SynchronousVibrations in Gas Turbo-Generator Setsby: A. Rakhit, Solar TurbinesAmerican Gear Manufacturers AssociationTECHNICAL PAPERDevelopment of an Epicyclic Gearbox for ReducedSub-Synchronous Vibrations in Gas Turbo-Generator SetsA. Rakhit,

2、 Solar TurbinesThe statementsandopinions contained hereinare thoseof theauthor and should notbe construed as an official action oropinion of theAmerican Gear Manufacturers Association.AbstractVibrations are present practically in all types of gearboxes. In general, vibrations at gear tooth mesh freq

3、uencies and theirharmonics areof concern toa gear designer. In starepicyclicgearboxes, lowfrequencysub-synchronous (below turbinespeed) vibrations are also quite predominant. Regular turbo-generator package testing shows that these low frequencyvibrations may arise even when the rotating components

4、are balanced precisely. This is attributed to the dynamiceccentricity of rotating parts such as ring gear and splined coupling. In star epicyclic gearboxes of stoeckicht design,splined couplings between the ring gears and outputshaft usually are preferredto make the ring gears fully floating forequa

5、l load sharing. The sub-synchronous vibrationsthat arisefrom suchcomponents finally gettransmitted along withthe mesh frequency vibrations to the turbine rotor bearings.Although low frequency sub-synchronous vibrations do not cause any damage to gear life, they seem to deteriorate thelife of the tur

6、bine rotor bearings which preferably are babbitted sleeve bearings of tilt-pad design. When the amplitudelevels of these vibrations and their harmonics exceed a certain limit, the life of these bearings, which are designed for aminimum servicelifeof 100,000hoursis significantlyreduced. Toreduce thel

7、evel of such vibrations,an improvement ismade to the design of star epicyclic gearbox that limits the dynamic eccentricity of ring gears and splined couplings.Gearboxes manufactured and tested to the new design were found to induce a very low level of sub-synchronousvibrations. The improved gearboxe

8、s not only allowed the service life of the turbine rotor bearings to be maintained, butalso increased the success rate of package testing the fnst time.Copyright 1995American Gear Manufacturers Association1500 King Street, Suite 201Alexandria, Virginia, 22314October, 1995ISBN: 1-55589-655-3DEVELOPME

9、NT OF AN EPICYCLIC GEARBOX FOR REDUCEDSUBSYNCHRONOUS VIBRATIONS IN GAS TURBOGENERATOR SETSA.K. RakhitSolar Turbines Incorporated, San Diego, CaliforniaIntroduction : between an ideal Stoeckicht and the alternate design by Allen Gears. Testingcarded out I showed the Allen design offers lower face loa

10、d distributinn in theIn gas turbine driven generators, the power turbine speed is reduced to the ring gearteeth. Also, conlrary to the Stoeckicht design, the face load distributiondesired generator rotor speed with a suitable gear speed reduce_. For medium factor was not found to be dependent on the

11、 Uansmitted torque. At full load, therange (5000 to 25,000 hp) gas turbines, power turbiue speeds usually vary from face load distribution factor of 1.13 was obtained with the Allen design of star8000 to 20,000 rpnL On the other hand,file most common 4-pole generator rotor epicyclic gearboxes compar

12、ed to 1.47 with the Stoeckicht design.speed is either 1800 rpm (60 cps) or 1500 rpm (50 cps). Thus, for 4-polegenerators the speed reduction ratios may vary from 5.33 to 13.33. Nevertheless, in gas t1_begenemtor applications, gearboxes with either Allenor Stoeckicht design seem to induce, besides me

13、sh frequency vibrations, someOf the various gear speed reducer arrangements used in gas turbine low frequency vibrations known as subsynchronons vibrations (SSV), the applications, a simple star epieyelic gearbox of Stoecldcht design offers the synchronous ones being at the turbine rotor frequency.

14、In turbomachinery,optimum design for reductionratios as high as 10:1. For higher ratios, compound when these low frequency vibrations get transmitted to the turbine rotor bearings,star epicyclics are preferred. Figure 1 illustrates the principle components of a theycause severe damage that reduces t

15、heir expected life of I00,000 hours. Thesimple and a_ star arrangements. In the present investigation, a simple present investigation was directed, first, to determine the causes of thesestar epieyclic gearbox was used to reduce the turbine speed of 8625 rpm to the undesirable SSV and, then to devel

16、op an improved gearbox for reduced levels ofgenerator rotor speed of 1500 rpm. Figure 2 shows a sectional view of this these vibrations at the turbine rotor bearings.gearbox.Gearbox Vibrations and Life of Turbine RotorIn contrast to a Stoeckicht design, the arrangement as depicted in Figure 2 Bearin

17、gs :was developed by Allen Gears Ltd 1 of England in which the radial and axialforces generated at the gemteeth are balanced by the reaction forces. The toroidal In agearbox, vibrations can occur at many different frequencies. The mostcouple is balanced by the tooth separatingforces acting on each i

18、nternal ring gear, common of these vibrations are foundat the gear teeth mesh frequencies and theirthus eliminating any deformation of teeth. Fignre 3 illustrates the basic difference harmonics. These are known to be caused by the transmission error of gear teeth.CARRIERSUN SUN- - mB-SIMPLE EPICYCLC

19、 (Star) COMPOUND EPICYCLIC (Star) RE95034MFigure 1: Epicyclie Gearbox - Star ArrangementsRING COUPLING CGULPLINGGEAR(XiJ ?_hcxizxmtally(x 1(full load). For each load, vibration records were taken at the turbine bearings, , i i ,ring gear, and coupling (fore andaf_locations). Among the turbine bearin

20、gs, No. |5 (being closest to the gearbox) showed the highest level of low frequency SSV. _ i ! _ li iThus, vibration levels at this beating were of primary concern. A typical i , vibration spectrum ofNo. 5 bearing taken at full load with an amplitude of 0.779rail P-P at 25 Hz is shown in Figure g. A

21、lso recorded were the dynamiceccentricities of the ring gears and coupling: 21.27 mils for the ring gear and _s%_e,z ,v .r_7 ,P. ?s9EU ExP1010to19.83 mils for the coupling. Figure 9 illusWates some typical orbital movementsof the turbine rotor at the No. 5 bearing, ring gear, and coupling. 10B._3 _.

22、 ,_ _G 101010_ _ _ Lv_t_Mz,PEAX FIE10 LIETEu._s eel0 .779 IB 5VAnalysis of Vibration Records : 103 As mentioned earlier,the minimum power required to lift the geometric axis L10_-,-_-_-:of the ring gears and coupling to the gearbox rotational axis is 338 hp 1500 _ 1_ I _rpm. This is significantly lo

23、wer than the power applied during the performancetest. Thus, it can be assumed that there were no low frequency SSV induced by PRESS _ _O EXITthe ring gears and conpling net being lifted to the gearbox rotational axis. Ontbe , T._sus_ o_s_other hand, the analysis of No. 5 bearing vibration spectrum

24、indicated that the 1010_1_sB_,z _ L“_q4_s_ _ _us_10_z s_c_10_ _ _vibration levels that exceeded 0.4 mil P-P were still at the rotational frequency ofthe ring gear and coupling. Also, phase measurements showed them was no _ _o_: _ _v_ 0_ 10u_ ,T_ _ _: _VL_“_10phase difference between the vibrations a

25、t the bearings and the vibrations at the _RYPa0H-2VaRG.5H _C3502_83 LOgO FULL LO_D 8 5V5._.t93 OIL TEHP 140ring gear andcoupling. This further identified the bearing vibrations to be of the _o._x_u_ -forced nature and could only be excited by the gearbox components (ring gearand coupling) which obvi

26、ously did not self-center completely with the gearbox . _rotational a_is. LOg -_t_-i L I_-S-vv_ “ “ I1_ Ill _ !*= I “ iAn investigation into the gearbox manufacturers manufacturing _t_,._ =_ “ i,_,_:_i-!individually with a residual unbalance less than the acceptable level of (4 W)/Noz. in. per API s

27、pecification, where W is the weight of a component (Ib) aa_ “ * _“ _ aandN is thespeed (rpm). The problem is, even in a precisely balanced assembly, _ u ta certain amount of unbalance is bound to exist, particularly if the componentshave splincs or gear teeth. Furthermore, to ensure the required bal

28、ancing Figure 8: Vibration Spectrums at No, 5 Bearing With Splined Couplingofa splined component, it is essential that balancing is done in a torque in the Gearboxload condition. Also, the balancing accuracy of assembled splined componentsis dependent on the orientation of the mating splines. Thus,

29、in practice, wheretorque and spline orientation of assembled components could vary, there wilt gearredneer the sun pinion is supported at the mesh points by the planets at onealways be some residual unbalance that causes dynamic eccentricity of these end, while the other is kept free, so that the pi

30、nion can adjust its position axiallycomponents. This was quite evident from the orbital excursion records of the and radially. Also, in a Stoeckicht or the alternate Allen Gears design of starcoupling and ring gears. The frequency of this orbital motion was therotational epicy_lics, the two ring gea

31、rs with opposite handed helices are connected to thefrequency of the coupling and ring gears, output shaft by a splined coupling. In this way, the ring gears are free to floatrelative to the planets, ensuring equal load sharing.An analysis of vibrationrecords at the No. 5 bearing indicates that ther

32、e wasa relationship between the P-P amplitude of these vibrations with the dynamic Gearboxes designed and built to these principles were used fiequently in eccentricity of the coupling and ring gears. The higher this eccenuicity, the tarbogenerator sets. The equal load sharing was verified by measur

33、ing the loadhigher were the vibration amplitudes at the turbine beatings. From the records on each planet and ring gears by several manufacturers. The test results wereof various packages tested, an approximate vibration transmission factor of 1/25 then correlated with the planet and ring gear teeth

34、 contact patterns. A gear toothwas calculated. This means that for 25 mils of dynamic eccentricity of ring gears contact pattern is considered good with a contact mark of about 80% along theand coupling, SSV amplitudes at the No. 5 bearing may be as high as I rail P-P. face width of teeth, 90% along

35、 the tooth height, and well centered. In the presentIt is thus apparent from this relationship that the dynamic eccentricity of ring investigation, gearboxes that did and did not meet the acceptable vibration levelgears and coupling cannot exceed 10 mii P-P for an acceptable SSV amplitude at the tur

36、bine bearings were all found to have good tooth contact patternsof O.4rail P-P at the beatings. Unfortunately, it was quite difficult to control this indicating there was equal load sharing among the planets and ring gears. Thisec.centricity to the desired limit with the current design of the gearbo

37、x. Thus, a iusfigated the improveraent of the gearbox design to limit the dynamic ecceatricitynm_nberof gearboxes were rejected during package testing because of high SSV of coupling and ring gear and would still ensure equal load sharing among theat the turbine bearings, planets and ring gears.Gear

38、boxes that did not pass package testing standards were removed and Of the various concepts considered, the one with limited floating featuredisassembled for possible rework of coupling and ring gears. Sometimes, when instead of the present fully floating coupling offered the best solution. This wast

39、hese components were assembled with phasing of splines at 90 degrees to the achieved by eliminating the splines at the aft end with a 12-bolt connection to theprevious orientation, vibration levels were reduced to the acceptable imiL In output shaft as illustrated in Figure 10. In addition, the clas

40、s and fit of all themany instances, it took several teardowns and testings before an acceptable remaining splines (coupling fore end and ring gears) were changed for reducedvibration level at the bearings was achieved. To eliminate such costly tcardowns clearances which also contributed in limiting

41、the dynamic eccentricity of theseand ensure allowable SSV at the turbine bearings, it was obvious that the best parts. The quality of planet and ring gear teeth was improved from AGMA C“lasspossible way to solve this vibration problem was to improve the design of the 10 to Class 12 to reduce ring ge

42、ar pitch diameter variability and mn out.epicyclic gearbox 6. Attention was also paid to proper selection of hardware (uniform weight) tomaintain the balancing accuracyof the assembly.Desima of an Improved Eoievelic Gearbox :With all the proposed design and quality improvements, two gearboxes ofOne

43、of theimportantfeatures of epicyclic gearbox design is to ensure equal Allen Gears design were built for test. Figure 11 shows one of these assembledload distribution among the planet gears. To accomplish this in a starepicyclic gearboxes with the author at the gearbox manufacturers plant in England

44、.z_-I I zs.o. - (a) iI !._.)_.I I It.-L_- ._,“:“-_-_ -“_ .i“,! _IMEN i.?.ZDIMENRBch : : : : “ I. .E . “ “ l ._i.i“ :.“I .TMIL : : “ _ :“ I MILi_.z.“i _._10-OOm -._-_w,. lO._m I !1 : _ I._. S i. IIZB. -ZB.0, MIL. ZB.Om _ -ZB.Om-_ Nit _._ZB.Om II 8-84-I .-.i.i )(d)i .T.i.i.“ i . :_. i it3i “:- BBch !

45、! . : : : I. _. - . !-T.FMIL I .: “ , : : : I: : HIL b.“._._._ _. :i . :,i lI I._i “ i :“ .:“- . . 1 . . . I :“ “: I I “ i “ : : : : :_ .= ) ! .: : .-zS.o.( : . -8.84.i ?“.“.“_ -zB.Om MIL ZS.0- -8.84m NIL 8.84=(a)Ring Gear;(b)Coupling(ForeEnd);(c)Coupling(aftend);(d)TurbineRotorL_Ne.5BearingFigure9.

46、TypicalVibrationOrbitsCOUPLING OUTPUT “ _)_ ._RE950_7MaFigure 10.“ Outer Coupling - Bolted DesignTest Results :At themanufacturers piant, both gearboxes were first tested at no load andfull speed. Then the load was gradually increased to about 1200 hp - themaximum capacity of the dynamometer availab

47、le. This power was above theminimum that is required to lift the geometric axes of rotating parts to therotational axis of the gearbox. A pictorial view of the set up is shown in Figure Figure 11: Improved Gearbox Assembly12. The results indicated lower dynamic eccentricity for the new designcompare

48、d to gearboxes with splines at both ends of the coupling. 13 depicts a typical vibration spectrum at the No. 5 bearing with an amplitudelevel of only 0.117 railP-P for one of the gearboxes. The level of vibration withThe two gearboxes were thenbroughtto the turbine manufacturers plant for the second

49、 gearbox was 0.3 rail PP. The dynamic eccenlricity of the ring gearfull-load testing with turbogenerator packages. The levels of vibration at the was found to be about 7.5 mil in both gearboxes.turbine bearings were found to be significantly lower in these packages. FigureLOG “,11 1 “ :-AP LNXI BRSE287.58 HZ V $8112 _P 352 EU SPEC EXP H 18Sunwheel R.H. Working Flank 188 54 25 I:PJ_IIRI_ AVG , DG 8OB NTG FT FR IXtlZLVL A!PEh_ FIH_ tlS_PLRCER3 5_/H 8233H _ PJ-d_ HZ EU _c I 12.588 8_47