GEAR DRIVES for TURBOMACHINERY by Peter Lynwander

GEAR DRIVES FOR TURBOMACHINERY by Peter Lynwander

INTRODUCTION driven machinery: Gearboxes for turbomachinery applications are Horsepower transmitted designed to achieve years of trouble-free oper- Ratio required (reduction or increasing) ation. As operating speeds increase, and relia- Speeds bility requirements become critical due to the Arrangement of Shafting (Fig. 1) high cost of downtime, gear manufacturers are When specifying a gear drive, the efficiency refining their analytical, design and manufac- requirement, noise generation, and space and turing techniques to keep pace with new tech- weight limitation must also be considered. The nology. It has become apparent that gear units, physical environment, dust, humidity, corrosive when incorporated into a system of rotating atmosphere, etc. must be addressed in the machinery, become susceptible to a variety of design stage. problems. The gear manufacturer and the user, Although the input and output shaft arrange- therefore, must take a systems approach to the ment can be concentric, parallel offset, right specification, installation, operation and maintenance of a gearbox. All characteristics of the drive system from the driver to the driven equip- Input Input ment including the lubrication system and acces- sories can influence gearbox operation. This chapter presents material concerning all phases of gearbox application with emphasis on information that will help the user operate and maintain the equipment. A discussion of selection and design procedures is included to familiarize Parallel Concentric the reader with gearbox principles. Offset Throughout the chapter standards and practices developed by the American Gear Manufac- turers Association are referred to. Successful Input selection, rating, installation and maintenance of gearboxes can be accomplished by the use of AGMA standards and practices. output TYPES OF GEAR DRIVES The choice of a gear drive depends on the appli- Right Angle cation, its environment and the physical con- straints of the system. The gearbox geometry is defined by four parameters which are determined Figure 1. Qpesof Gearbox Configurations by the characteristics of the driving and angle or skewed, the great majority of turbo- helical tooth forms. Figure 3 illustrates the dif' machinery gearboxes are used in systems that ference between spur and helical gearing. Spui require either parallel offset or concentric shaft gears have the advantage of not generating axiai configurations. Figure 2 illustrates a typical thrust loads but are limited in capability and the3 parallel shaft high speed gear unit. generate more noise and vibration than helicai Parallel offset or concentric drive shaft gear- gears. The reason for this is that helical gears boxes use either spur, single helical or double have an overlap in both the axial and transverse planes, Fig. 3. With a conventional spur gear, the load is transmitted by either one or two teeth at any time. Thus, the elastic flexibility is continuously changing as load is transferred. With helical gearing, the load is shared between a sufficient number of teeth to allow a smooth transfer- ence of load and a constant elastic flexibility. In addition, helical gears have larger load carrying capability. In order to take advantage of helical gearing, yet not generate axial thrust, double helical gearing can be used. Many industrial and marine applications utilize soft, approximately Rc 35, 2ouSle helical gearing. Double helical gearing offers low noise and vibration along with zero net axial thrust. Also, the ratio of face width to diameter in each half of the mesh can be held to reasonable limits, therefore, end loading of the tooth face due to 'tooth errors or deflection is less likely to occur. The advantage of zero thrust is offset by the fact that double helical gears must adjust them- selves axially. One gear of the set is free to move Figure 2. American Lohmann Parallel Shaft axially and continuously shifts to achieve equi- High Speed Gear Unit librium. This can lead to detrimental axial vibrations. Also, if the gear is hung up axially due to SPUR GEAR HELICAL GEAR external loads or internal friction overloading of one helix will occur. In order to achieve minimum envelope and maximum reliability the latest technology util-

Figure 4. Two Stage, Single Helical Generator Drive Manufactured by American Figure 3. Comparison of Spur and Helical Lohmann. 4500 HP, 14500 RPM In, Gear Teeth 1500 or 1800 RPM Out Table 1. High Speed Gear Unit Service Factors f(;)urlr.sy Amc7rrcnn Cvnr iManrr/ac.turc,rs As.sociot~r)n)

AGMA STANDARD PRACTICE FOR HIGH SPEED HELICAL AND HERRINGBONE GEAR UNITS Servlce Faclor Values Servlce Factar Pr~mcMover

Internal Cornbusl~on APPLICATION Eng~ne Motor Turb~ne (Multi-Cyl~nderl

BLOWERS Cenlr~fugal 14 16 17 Lobe 17 17 2 0

COMPRESSORS Centr~fugal-processgas excepl alr cond~t~on~ng 13 15 16 Cenlnfugal-a~rcond~ltonlng servlce 12 14 15 Centrifugal-a~ror plpe l~neservlce 14 16 17 Rolary-axtat flow-all lyoes 14 16 1 i Rotary-llquld plston (Nash) 17 17 2 0 Rolary-lobe-rad~alllow I7 17 ? 0 Rec~procal~ng-3or more cyl 17 17 2 0 Rec~procal~ng-2cyl 2 0 2 0 2 3

DYNAMOMETER - rest stand 11 11 13

FANS Cenlrlfuqai 14 I6 17 Forced draft 14 16 17 Induced draft 17 2 0 2 2 lndustrlal and mlne (large wlth trequent start cycles) 17 2 0 2 2

GENERATORS AND EXCITERS Base load or continuous 11 11 I3 Peak duty cycle 13 13 17

PUMPS Cenlr~fugal{all service except as 11stedbelow) 13 I5 17 Cenlrlfugal-boller feed 17 2 0 Centr~fugal-descal~ng(with surqe lank] 2 0 2 0 Centrifugal-hot 011 15 17 Cenlrifugal-p~pel~ne 15 17 2 0 CentrlfugaE-water works 15 17 2 0 Dredge 2 0 2 4 2 5 Rotary-axlal flow-all types I5 15 ! 8 Rotary gear 15 15 18 Rotary-llquld ptston 17 17 2 0 Rolary-lobe 17 17 2 0 Rolary-sl~d~ngvane 15 15 18 Rec~procat~ng-3cyl or more 17 17 2 0 Rec~procating-2cyl 2 0 2 0 2 3

PAPER INDUSTRY Jordan or ref~ner I5 15 Paper mactllne-line shaft 13 13 Paper machlne-sect~onaldrlve 15 Pulp beater 15

SUGAR INDUSTRY Cane knlfe 15 I5 18 Cenlr~fugal 15 17 2 0 M~ll 17 17 2 0 - izes single helical, hardened and precision ground gearing. With single helical gears the thrust load axially locates the gear shaft against the thrust LOW SPEED STATIONARY GEAR SHAFT bearing. Bearing design has progressed to the point where thrust loads are routinely handled either by hydrodynamic tapered land or tilting pad configurations or anti-friction thrust bearings. Because case hardened gears have maximum load carrying capacity, gear size can be minimized, therefore, the ratio of face width to diameter of a single helical gear can be held to reasonable limits. Also, the bearing span with single helical gears is shorter resulting in lower PLANET GEAR elastic deflection. Figure 4 illustrates a generator drive gearbox with two stages of single helical I I gearing. A single helical hardened and ground gear set can reduce by up to one half the envelope and weight of a through hardened double helical gear- fl ,PLANET GEAR box with equivalent capacity. The inherent prr- cision of the grinding process results in accurate LOW SPEED CARRIER SHAFT tooth geometry leading to minimum noise and vibration. It is possible to harden and grind double helical gear teeth, however, in order to grind a one piece double helical gear a large cen- tral gap is required between the two helices to allow runout of the wheel. Gears can be ground in halves and then assembled, but this presents serious alignment and attachment problems. Gearbox Rating 'ATIONARY RING GEAR Turbomachinery gear units are usually rated by a an established practice such as A.G.M.A. Stan- dard 421.06, "Standard Practice for High Speed Figure 5. Basic Planetary Gear Configurations Helical and Herringbone Gear Units" or API Standard 613, "Special Purpose Gear Units for Refinery Services." The two standards mentioned COMPRESSIVE STRESS above are basically for helical and double helical parallel shaft units. Gear horsepower rating is calculated on the basis of strength and durability. In addition, to gear rating, the standards cover other aspects of gearbox design such as bearings. shafting, etc. Standard parallel shaft helical or double helical gear units are available and described in pub- lished catalogs. Service Factor Gear catalog horsepower ratings are given with a service factor of one, and before a unit is selec~c~rl the operating conditions must be defined so 111at a suitable service factor can be chosen. Service factors are used to take into consideration intan- gible operating conditions such as misalign- ments, vibrations, transient loads and shocks. The actual horsepower is multiplied by the service Figure 6. Calculation of Compressive Stress factor to obtain an equivalent horsepower and the (Pitting Crit.erion) gear unit selected must have a rating equal to or greater than the equivalent horsepower. Table 1 presents A.G.M.A. service factors for high speed ' the lubricant film fails allowing metal to units. A high speed unit is defined as operating metal contact. Local welding is initiated with a pinion speed of 3600 revolutions per minute and the welded junctions are torn apart by and higher, or pitch line velocities of 5000 feet per the relative motion of the meshing gear minute and higher (PLV= RPM x Pitch Diameter teeth. The flash temperature index [l] xa/12). Standard catalog gear units are listed to appears to be the most reliable method of approximately 20,000 feet per minute. Applica- analysis used at present to predict scoring. tions exceeding this speed must be considered special and exceptional care must be taken in their design and manufacture. 8ENDlNG STRESS Planetary Gearing With planetary gearing the transmitted load is shared between several meshes. therefore, gearbox envelope and weight can be significantly reduced compared to parallel shaft designs. Figure 5 illustrates single stage planetary confieurations.- In addition to achieving minimum weight and envelope, the small, stiff components used in planetary gearing result in reduced noise and vibration and high efficiency. When confronted with a choice between parallel shaft and planetary gearing, it would appear that the planetary is more expensive and has more components, however, for high power, high speed applications, the reduced pitch line velocity and smaller components make planetary configurations very attractive and there is a trend toward Figure 7. Calculation of Bending Stress this type of gearing. (Breakage Criterion)

GEAR DESIGN FLASH TEMPERATURE RISE Modern design methods for turbomachinery drives make use of sophisticated analytical procedures. Computer programs enable the designer to optimize gearbox components with great precision. Detail gear tooth designs are defined using mathematical models describing the major failure modes of turbomachinery gearing; pitting, breakage and scoring. Pitting, Fig. 6, is a fatigue phenomenon and occurs as a result of repeated stress cycles which lead to surface and subsur- face cracks. Eventually particles detach and pits form. Pitting, although affected by the oil film, can occur even with an adequate film thickness. Breakage, Fig. 7, of gear teeth is caused by the root bending stress imposed by the transmitted load. Generally, tooth breakage is caused by bending fatigue rather than a transient overload exceeding the gear tooth fracture strength. In some cases pitting or wear may weaken the tooth to the extent that breakage occurs. Figure 8. Calculation of Flash Temperature Rise Scoring, Fig. 8, is a form of surface dam- (Scoring Criterion) age on the tooth flanks which occurs when DRIVES DRIVEN PINION EXTEWL GEAR

35.68@4888 169.088flfl44 TRflNSVERSE DIAflErRAL PITCH Q,?45388@ 5.7A27841 TRPNSVERSE PRESSIJRE ANGLE!DEt) 5.5659537 21.4341413 CENTER DISTeNCE

SEL4'I'JF 3OLLING SPEED !fiPfl) PITCH LINE VELOCITY (FPV! ?EES T??QIJE (IN-LES! MESH RATIO SENDING GEOMETRY FACTOR RELATIVE HlJSSEPO4ER pE? VESH a:LDiYG- - STTESS (Pfi) EFFECTIVE FbCE WIV6 :~';~I!NI:. LI'E [YJUfiS) STATIC TANCENiId LOAD (LPS! ~EK;;YG SPSETY Fal:;Ofi OYNANIC FACTOR lO9PPESSIVE STRESS [PSI) ALIGNMENT idCTOR CSEPaESSIVE LIFE !POURS) IODIFiED T4NGENTIBL LOAD iiBS) ISIP?SS;IVE SAFETY FACTOR SURFACE FINISH SLIDIYG VELOCITY AT TIP IFPV) FLASH TEMPERPTURE 9!.iE [DEC :I

;.C .y.A. ~bTE9IdL5RbDi 2.1688848 2.808flfl0fl PROFILE CONTACT RATIO Y-TERNRTING BENDING FACTOR I.08488@8 I .flBflBWfl FACE CONTKT RbTlO Vi.Y'BER jF PESkij PEF izV !.a848444 1.448@1BB #IN CONTACT LEYCTa 3PSE HELIX PkGLE (GI;) ¶.22,?743 9.8223743 RAlIMUn CONTACT LENGTH

IUTS :@E DIAMETER 4.928- 4.825 21.678-21.583 BACKLKH YI'CH DIAflETER 4,5552'37 i1,4?Pi413 CIRCULAR PITCH c3PY FiAflETER 4.353421i 21.2861243 LEAD EPROR (INIIN) EASE DIPVETER L.i?72171 1?.37499!5 BflSE PITCP R?37 DIAVETER 4.176- 4.216 21.654-21.874 DEPTH TO POINT OF #AX WEAR

%iL aHGLE-WAX OU;SIOE DIfl 34.6963974 28.787 138 1 3OLL 44G-E-ROUND EDGE DIA 34.2??713 23,631;347 - .. , ..--4NFLE-?16!4 SIr((LE TOOTS 27.2279037 ?7.5?2i?42 ?::: ?::: CNf:E-?!-CH DIfl 27,1894257 27.1884257 .?[I:; &NQE-LOU SINGLE TOFT6 24.1758991 25.5567683 ?:I;'- ?:I;'- AUXE-FORP DIAfliTER i?.2278847 i5.4919457

-07 14NF iYICKNESS 1,Bb56685 B,6550!63 'EX .:$,;I JE?;K 8.9566Sil 8.8526927 :?F#SVERSE CIR TOOTH THICKNESS 8.1'794667 4.18497?8 FGlRAL CIR TOOTH THICKNESS 8.1746- 8.1765 8.1813- 0.1833 h33WL 81ARETRbi PITCH 7.P999978 7.79979'?8 'W,;~*;L ?RE$fijRE fi&G!E 24.9799845 24.9979946 Lift11 93.5375887 392.1715817 aaam EDGE RIDIUS MAX B.BIBBB~B ~.PIBBBIB SOOT FILLS? SbEIUS MIN 8.fi548471 8.P497755 ESOLE MPTH COriSTANT 2.4814848 2.4484844 :,:aa~s:~ at TIP OF ~~oi,u ~.g:@391z 8.8583912

EALL Di4bETER 8.1258a88 8.!25886@ YE4 ?YREnENT OVER BALLS 4.5515- 4,556521.3779-21.4424 nr. I r*:~C?NTACT DIAOETER 4.3933- 4.339121.2231-21.2275 DIN OVER TOP LAND 8.1367593 8.1425351

Figure 9. American Lohmann Gear Analysis Computer Output Sheet GEAR DRIVES

The equations shown in Figs. 6 and 7 are in a and/or flank, Fig. 10. These tooth modifications basic form. In practice, when applied in the allow the gears to enter mesh smoothly and A.G.M.A. load rating system, various factors reduce dynamic loading, vibration and noise. are applied which account for detrimental oper- mpical tooth modifications are in the order of ating conditions. These factors are listed in Table .0004 - .0010 inches. 2 to illustrate the many parameters that have to be In order to reduce the effect of misalignment considered. References C11, C21 and C31 illustrate which causes end loading, gears are sometimes the use of the stress and scoring equations and crowned in the axial direction (lead modification). list allowable design values. Figure 10 illustrates this technique. The amount As mentioned earlier, detailed gear tooth geom- of crowning at the tooth end is generally in the etry is optimized through the use of computer order of .001 inches. analysis. Figure 9 presents a typical computer High speed turbomachinery gearing requires output sheet defining gear tooth stresses and excellent quality control. AGMA Standard 390.03, geimetry. Gear pitch diameters are generally January 1973, recommends gear specifications determined by compressive stress considerations. for quality. AGMA Quality Classes are numbered however, the detail-tooth design is defined by the from 3 to 15 with 15 being the most precise. The i flash temperature rise along the line of action. majority of turbomachinery gears will fall in the To optimize a given design, the maximum instan- 10-13 Quality ranges. Table 3, from AGMA 390.03, I taneous flash temperature rise during the arc of presents typical quality data. approach must equal the maximum instantane- Following is a definition of each tooth tolerance ous flash temperature rise during the arc of recess. element shown in Table 3: This is accomplished automatically by the use of Runout Tolerance - The amount of pitch iteration techniques built into the computer pro- line runout allowed. Checked by such gram. The computer program, starting with stan- methods as indicating over pins placed dard gear tooth addendums, will automatically in successive tooth spaces or rolling with vary the pinion and gear addendums in defined a master gear of known accuracy. increments until the optimum flash temperature Pitch Tolerance - The allowable amount condition is obtained. With the resulting "long of variation between corresponding points and short" addendum designs of this nature, on adjacent teeth (tooth to tooth). standard tooth thicknesses are no longer appli- Profile Tolerance - The amount of vari- cable. If standard tooth thicknesses are utilized, an unbalance of bending stresses between pinion and gear would result. To optimize the bending PROFILE MODIFICATION stresses, the program enters a second iteration procedure which varies tooth thickness until bending stresses are balanced. Tooth Modifications If gear teeth were manufactured perfectly and there were no deflections during operation, the tooth form would be a pure involute. Because there are always manufacturing errors and tooth deflections, gear teeth are relieved at the tip

LEAD MODIFICATION Table 2. Factors Used in AGMA Rating Equations

Factor Strength Durability Dynarn~cLoad K, C, Overload KO co Size K, C, Hardness Ratlo Cn L~fe KI CI Temperature Kt Ct Factor of Safety K, C. Load Distribution Km Cm Surface Condition G Figure 10. Illustration of Gear Tooth Modifications CHAPTER 11 GEAR DRIVES

ation allowed in the functional profile and can be detected in time for replacement to (involute). be effected. Lead Tolerance - Allowable variation of Journal bearing materials are usually steel the tooth surface along the face width. lined babbitt or a tri-metal constructon where a Gear Materials and Heat Treatment high strength, high temperature bearing material Proper choice of gear material and heat treatment is sandwiched between a thin babbitt layer and is probably the most important factor in the suc- the steel backing. cessful operation of a gear set. In choosing a gear Radial journal bearings may have round bores material the tooth hardness and type of heat treat- or various other shapes in cases where dynamic ment to achieve that hardness must be considered. instabilities are possible. Gears, case hardened by carburizing, provide Thrust bearings are either micheled for light the best metallurgical characteristics combining loads or have tapered lands or tilting pads. good case structure with reasonable ductility. Seals Carburizing produces the "strongest" gear pro- The most common method of sealing high speed viding bending and pitting fatigue resistance turbomachinery gearboxes is by the use of non- and an excellent wear surface. A disadvantage of contacting labyrinth seals. Two sets of labyrinths carburizing is that gear teeth distort during the are used with a drain in between which leads back heat treatment and, in order to obtain high pre- to the oil sump. If air pressure is available the cision, grinding after hardening is required. Sur- labyrinth can be internally pressurized with 1 face hardnesses attained by carburizing are in or 2 psig. the order of Rc 55-62. For critical applications Rc GEARBOX INSTALLATION 60 minimum should be specified. The best carburizing steel to achieve high load Installation of the transmission is critical for carrying capacity is SAE 9310 (AMS 6260 or AMS proper performance. The gearbox must be rigidly 6265). Other carburizing steels used include SAE connected to the foundation which must also be 8620, 4620, 4320 and 3310. rigid and have a flat mounting surface. If the unit The nitriding process is used to case harden is to be mounted on a surface other than hori- gears when distortion must be held to a minimum. zontal the manufacturer should be consulted to Often the gears are finish cut and then nitrided, determine if any lubrication system problems or eliminating the grinding requirement. Nitrided other difficulties will arise due to the gearbox gears do not have the bending and pitting fatigue attitude. resistance of carburized gears but do provide a When handling the gear unit at installation, hard, wear resistant case. With nitriding steel care must be taken not to stress parts which are such as AMS 6475 (nitralloy N) or AMS 6470 not meant to support the gearbox weight. Gear- (nitralloy 135) case hardnesses of RC 65-70 can be boxes should be lifted only by the means provided achieved. Steels such as SAE 4140 and 4340 are by the manufacturer such as lifting holes in the casing. nitrided to hardnesses of 320 to 380 BHN. If the transmission is to be mounted on a pedes- Through hardened gears in turbomachinery units are generally in the 300-400 BHN hardness tal or base plate the structure must be carefully range. '&pica1 through hardening steels are analyzed to determine if it will withstand oper- SAE 4140 and 4340. ating loads without excessive deflection. When mounting the unit on steel beams a base plate Often, combinations of pinion and gear materi- should be used. The base plate should extend als and heat treatments are used. For instance, if under the entire gearbox and be at least as thick the gear is very large it may not be practical to as the gearbox base. Both the unit and the base harden, and the gear set might consist of a carbu- plate should be rigidly bolted to the steel supports rized and ground pinion driving a through with proper shimming to achieve a level surface. hardened gear. If the gearbox is to be mounted on a concrete foun- Bearings dation, grout steel mounting pads into the con- High speed turbomachinery gearboxes generally crete base rather than grouting the transmission incorporate journal bearings. Journal bearings directly into the concrete. This will facilitate any offer the possibility of very long life whereas anti- shimming and alignment required. The concrete friction bearing fatigue life calculations generally should be set before bolting down the gearbox result in finite lives which may not be acceptable. and cured before loading is applied. There is also a feeling that journal bearings have The gearbox can be brought into alignment by far fewer components and therefore are more reli- placing broad flat shims under all mounting able. Also, journal bearing failure modes are less pads. A feeler gage is used to determine that the catastrophic than those of anti-friction bearings thickness of shims is correct under all mounting CHAPTER 11

pads. ated in the gearbox so that component tempera- The gearbox input and output shafts must be tures are not excessive. The majority of the oil correctly aligned with the driving and driven flow is required for the cooling function. For every machine shafts. Even if the total system is deliv- gear drive there is a mechanical rating; the load ered on a permanent mounting already aligned at the transmission can transmit based on stress the factory, couplings should be disconnected and wear considerations. In addition, there is a and alignment rechecked. The alignment may thermal rating which is the average power that have been disturbed during shipment. can be transmitted continuously without over- Alignment of the couplings can be performed heating and without using special cooling. AGMA in two phases. First the components are roughly thermal ratings are based on a maximum oil aligned by eye and the final adjustment is accom- sump temperature of 20VFC41. In turbomachin- plished with the use of measuring instruments. ery applications, the thermal rating is usually When accomplishing the rough alignment, the less than the mechanical rating and an external unit being moved into alignment is adjusted until cooling system is required. Figure 11 presents a the coupling half rims are close horizontally and typical gearbox lubrication system in schematic vertically. Also, the coupling halves axial dis- form including pump, cooler, filter and pressure tance should be as specified. When the rough relief valve. alignment is completed, the bolts around the When designing a lubrication system, the ini- base of the unit should be secured but not tial step is to estimate the oil flow to the com- tightened. ponents and the gearbox efficiency. The temper- The angular alignment is then checked by ature rise across the gearbox can then be mounting a dial indicator on the hub of one cou- calculated: pling with the indicator pin on the face of the A t= Q/M Cp mating coupling. The shaft on which the indi- Where: A t = temperature rise - "F cator is mounted is rotated and the indicator M = Flow - Lb./min. reading recorded. The unit is adjusted until the (Note - 1 GPM = 7.5 Lb/min.) specified indicator reading is achieved. A reason- Q = Heat - BTU/min. able value might be .003 inches TIR. The dial (Note - Q= HP x 42.4) indicator is then attached to the opposing cou- C, = Specific Heat = .5 BTU/Lb. "F pling hub and the procedure done again. For example, a gearbox transmitting 1000 HP The parallel alignment is then checked by with 98% efficiency will reject 20 HP or 848 BTU/ mounting a dial indicator on the hub of one cou- min. of heat to the oil. If the gearbox flow is 20 pling with the indicator pin on the outside diame- GPM or 150 Lb./min. the temperature rise across ter of the mating coupling. Again, the shaft is the gearbox will be 11°F'. Typical operating tem- rotated and the unit adjusted until the specified peratures for turbomachinery gearboxes are oil reading is achieved. The indicator is then "in" of 130°F with a rise of 30°F. These values are attached to the opposing coupling half and the for mineral oils; synthetic oils may operate at procedure done again. higher temperature levels. There may be several iterations until both The amount of oil flow supplied to a gear mesh angular and parallel misalignment requirements is generally based on experience and experimen- are satisfied. The mounting bolts should then be tal data. A rule of thumb sometimes used is .02 torqued and the unit operated until it attains Lbs./min./HP. Oil may be jetted going into or out normal temperatures. At this point, the alignment of mesh or both. For high speed gearing, it is gen- should be rechecked and adjusted if necessary. erally agreed the majority of oil should be jetted If couplings, sprockets, pulleys, or gears are to the outgoing side to achieve better cooling and installed in the field do not hammer them into reduce losses. position. This might damage bearings or gears. The quantity of flow passing through a gear- Instead heat the component and slide it on. box is regulated by the oil iet diameters. journal Special attention must be given to back-stops, bearing clearances and feed pressure. dearbox which should be checked before start-up. The feed pressures are generally in the order of 20-100 shaft should be turned by hand to determine if psig. To hold a constant feed pressure, a pressure the backstop is assembled correctly for the direc- regulating valve is incorporated at the gearbox tion of rotation. inlet as shown in Fig. 11. Oil pumps for gearboxes are positive displace- LUBRICATION SYSTEM ment types such as gear, vane or screw pumps. The purpose of the gearbox lubrication system is The pump capacity will be somewhat oversized to to provide an oil film at the contacting surfaces account for variations in gearbox flow and pump of working components and absorb heat gener- deterioration. For instance, a 23 GPM pump GEAR DRIVES

might be chosen for a 20 GPM application. The accepted as the standard of the petroleum output of positive displacement pumps is directly industry. proportional to pump speed and at any given AGMA Lubricant number recommendations speed is constant regardless of pressure. The for drives using all types of gearing except worm pump must develop sufficient pressure to supply gearing are given in Table 5. It can be seen from the gearbox feed pressure and make up the pres- Table 5 that most turbomachinery gearboxes will sure drops in the cooler, filter and oil lines and utilize an AGMA type 1 or 2 lubricant. fittings. Therefore, a gearbox with a 50 psig Lubricant maintenance, oil change intervals, design feed pressure may have an oil pump oper- flushing, etc., is discussed in the maintenance ating at 100 psig. Sometimes a high pressure section. relief valve is incorporated at the pump to protect GEAR TESTING AND OPERATION the pump in case a downstream line or valve is inadvertently closed. Tbrbomachinery gear units are generally tested The oil cooler is sized on the basis of the magni- by the manufacturer prior to delivery. The type of tude of cooling required and the pressure drop for test should be agreed to in advance. A typical the rated oil flow. Both air/oil and water/oil light load, full speed acceptance test is as coolers are used. If cold starts are anticipated a follows C51: thermal bypass can be incorporated in the cooler. 1. Operate the gearbox at maximum con- Dual oil coolers with a changeover valve are tinuous speed until bearing and lubri- sometimes specified for critical applications. cation oil temperature has stabilized. The oil filter is sized on the basis of dirt capac- 2. Increase the speed to 110 percent of maxi- ity and pressure drop for the rated oil flow. The mum continuous speed and run for a degree of filtration is indicated by the micron minimum of 15 minutes. rating of the filter element. Typical turbomachin- 3. Reduce speed to maximum continuous ery gearbox filters are in the order of 40 microns. and run for four hours at a minimum. Bearing in mind that a micron is ,000040 inches The following measurements should be made a case can be made for using fine filtration of 10 microns or better. For instance, typical gear tooth and bearing oil film thicknesses are in the order of .0005 inches, therefore, a 10 micron maximum filter is required if protection is desired against particles larger than the oil film thickness. As the oil filter collects debris, the pressure drop across the filter increases. The filter, therefore, should incorporate a bypass valve and an impending bypass warning. Dual oil filters with changeover valves are sometimes specified for critical applications. Lubrication system components can be supplied either by the gear manufacturer or the user. The purchase order should be specific as to who supplies what, which specifications apply and where the interfaces are in order to avoid compli- cations at installation. A.G.M.A. Standard 250.03, "Lubrication of Industrial Enclosed Drives", May 1972 gives recommendations for the grades of oils to be used in gearboxes. Lubricant viscosity recommendations are specified by AGMA Lubricant Numbers as shown in Table 4. The former AGMA viscosity system (AGMA 250.02, December 1955 and Oil pump 1 Suction pip@ Oil filter K Pzeaaure plpt AGMA 252.01, May 1959) has been abandoned in oil cooler L Return flow pl~e favor of a newer, more universal system as shown Oil tank H Oil level olllqe Prsssure qauqc N Primer and spare pomp unit in Table 4A. The AGMA Lubricant Numbers Thcrmowter o Temperature 94~9~ Check valve P connection for remte thermometer have been retained but the corresponding vis- Bypass Q Pressure requlator cosity ranges will now conform to the ASTM- ASLE system (ASTMD-2422 "Viscosity System Figure 11. Typical Gearbox Lubrication System for Industrial Fluid Lubricants") which is widely CHAPTER 11

Table 4. AGMA Lubricants (Courtesy American Gear Manufacturers Association)

AGMA STANDARD SPECIFICATION LUBRICATION OF INDUSTRIAL ENCLOSED GEAR DRIVES Viscosity Ranges for AGMA Lubricants

Rust and Oxidation Viscosity Range Extreme Pressure Inhibited Gear Oils ASTM System (2) Gear Lubricants (5)

AGMA Lubr~cantNo. SSU at 100°F AGMA Lubricant No.

1 193 to 235 2 284 to 347 2 EP 3 417 to510 3 EP

4 626 to 765 4 EP 5 918 to 1122 5EP 6 1335 to 1632 6EP

7 comp. (1) 1919 to 2346 7EP 6 comp. (1) 2837 to 3467 8 EP 8A comp. (1) 41 71 to 5098

NOTE: Viscosity ranges for AGMA lubricant numbers will henceforth be identical to those of the ASTM System (2).

Table 4A. AGMA Lubricants (Courtesy American Gear Manufacturers Association)

Equivalent Vlscositles of Other Systems (For Reference Only)

Viscosity Ranges from Equivalent Metric Equivalent Former AGMA System (3). (4) AGMA ASTM-ASLE Viscos~tyRanges . Lubricant No. Grade No (2) cST at 37.8"C (10O0F) SSU at 100°F SSU at210"F

1 S215 41.4 to 50.6 180 to 240 2.2EP S315 61.2 to 74.8 280 to 360 3.3EP S465 90 toll0 490 to 700

4.4EP S700 135 to 165 700 to 1000 5.5EP SlOOO 198 to 242 80 to 105. 5.6EP S1500 268 to 352 105 to 125

7 comp, 7EP S2150 414 to 506 125 to 150 8 comp. 8EP S3150 612 to 748 150 to 190 8A comp S4650 900 to 1100 190 to 250

(1) 011smarked comp. are compounded with 3% to 10% fatty or synthet~cfatty oils. (2) "Viscosity System for Industrial Fluid Lubricants". ASTM 2422. Also Brltish Standards Inst~tute.B.S. 4231. (3) AGMA 250.02 (December 1955). (4) AGMA 252.01 (May 1959). (5) Extreme pressure lubr~cantsshould be used only when recommended by the gear drive manufacturer. GEAR DRIVES

Table 5. AGMA Lubricant Recommendations (Courtesy American Gear Manufacturers Association)

AGMA STANDARD SPECIFICATION LUBRICATION OF INDUSTRIAL ENCLOSED GEAR DRIVES AGMA Lubricant Number Recommendations for Enclosed Helical, Herringbone, Straight Bevel, Spiral Bevel and Spur Gear Drives

Other Lubricants AGMA Lubricant Number Type of Unit Size of Un~t Ambient Temperature F

Main Gear Low Speed Centers - 40 - 20 to 0 to +25 15 to 60 50 to 125

Parallel shaft, (single reduction), up to 8 in 2- 3 3-4 Over 8 in. and up to 20 in. 2-3 4-5 Over20 in. 3-4 4-5

Parallel shaft. (double reduct~on),up to 8 in. 2-3 3-4 Over 8 in. and up to 20 in. - 3-4 4-5 .-as Over 20 in. I -m $ 3-4 4-5 ;: s0 zC Parallel shaft. (tr~plereduction), up to 8 in. .S I Z al 2-3 3-4 Over 8 in. and up to 20 in. u, u, 0 9 4-5 .? 1 - I 3-4 Over 20 in. 5 z 2 4-5 5-6 c 3 E 4 Planetary gear unlts z 6 2 O.D. Housing up to 16 in. o a 2- 3 3-4 -2mJ O.D. Hous~ngover 16 in. ij .z 3-4 4- 5 0E .-'E c c 3"' W Sptral orstraight bevel gear units 4 5 2 5 Cone distance up to 12 in. 2- 3 4-5 Cone distance over 12 in. 3-4 5-6

Gearmotorsand Shaft mounted units 2-3 4-5

High-speed Units (See Note 4) 1 2

NOTES

1. Pour point of lubrtcant selected should be at least 10°F lower AGMA 421.06. "Practice for High Speed Helical 8 Herring- than the expected minimum ambient starting temperature. bone Gear Units:' for detailed lubr~cationrecommendat~ons. If ambient starting temperature approaches lubricant pour 5. When they are available. good quality industrial oils hav~ng polnt, oil sump heaters may be required to facilitate starting s~rnilar properties are preferred over the automotive oils. and Insure proper lubr~cation.(See paragraph 5.)' The recommendation of automotive oils for use at ambient 2. Ranges are prov~dedto allow for var~ationsin operating temperatures below +15"F is intended only as a guide cond~t~onssuch as surface finish, temperature rise, loading, pending widespread development of satisfactory low temper- speed, etc. ature ~ndustrtaloils. Consult gear manufacturer before 3. AGMA viscosity number recommendations listed above proceeding. refer to R 8 0 gear oils shown in Table 2.' EP gear lubricants 6. Drives ~ncorporatingoverrunning clutches as backstopping In the corresponding viscosity grades may be substituted devices should be referred to the clutch manufacturer as where deemed necessary by the gear drive manufacturer. certain types of lubricants may adversely affect clutch 4. High speed un~tsare those operating at speeds above 3600 performance. rpm or pitch line velocit~esabove 5000 fprn. Refer to Standard 'Editors note: This refers to material outside this text.

AGMA 250.03-May. 1972 CHAPTER 11

method of full load and speed testing is back-to- Static torque npplied by rotatinq back operation. 'nvo identical units are coupled coupling flanges in opposite directions together, input shaft to input shaft, output shaft to output shaft. The desired torque is locked into the system by mechanical or hydraulic means. The unit is then operated at the desired speed by a prime mover which need only be large enough to d_"h~u> slave unit Test unit - make up the efficiency losses in the system. Figure 12 illustrates the simplest type of back-to- back test rig. On occasion full load, low speed or static torque tests are employed to demonstrate tooth contact, load carrying capability and casing rigidity.

Initial Field Start-up Figure 12. Back to Back Gear Test Rig Prior to starting the equipment the following pre- liminary checks should be performed: Check oil level and ensure that proper oil during the acceptance test: is being used. Oil inlet temperature Tighten all pipe connections. Scavenge oil temperature Check all electrical connections. Oil feed pressure Tighten all mounting and gearbox bolts Oil flow with proper torque. Shaft speed Check the mounting of all gauges, Other parameters that can be measured during switches, etc. gear testing are: Check all couplings for proper instal- Vibration lation and alignment. Shaft excursion Check inspection cover installations. Noise The following instructions pertain to the Bearing temperatures initial start-up: A detailed test log should be kept making 1. The unit should be pre-oiled to ensure entries of each measurement at 15 minute lubrication of the journal bearings at intervals. start-up. After completion of the mechanical running 2. The gearbox should be started slowly test, the gear unit should be opened for a visual under as light a load as possible. Observe inspection. Tooth meshes should be inspected for that the rotation is in the proper direction. surface damage and proper tooth contact. All Check system oil pressure. bearings and journals should be inspected for 3. After starting when the oil has been circu- signs of surface damage or overheating. lated the unit should be stopped and suf- With high speed gear drives it is common to ficient oil added to the sump to bring the conduct the full speed acceptance test driving sight gage oil level up to the specified the gearbox through a low speed shaft. This is amount. done if a high speed prime mover is not available. 4. As the unit is brought up to operating In this case, the gears are contacting on their speed it should be continuously monitored normally unloaded faces since the gearbox is for excessive noise, vibration, or tempera- being driven backwards. Such a test can still be ture. If any of these occur shut down useful to determine proper operation of the lubri- immediately, determine the cause and cation system and correct alignment of the gear take corrective action. Also check for oil shafts. Also, any gross machining or assembly leaks as the unit is initially operated. errors can be identified. 5. If possible, operate at half load for the The full speed "light" load test is widely used first 10 hours to allow final breaking in of since full load testing at a gear vendor's plant is the gear tooth surfaces. costly and sometimes a prime mover of sufficient 6. After the initial 50 hours of operation the capacity is not available. If a full load test is oil in a new unit should be drained and required by the customer, the load can be applied the case flushed with SAE 10 straight by mechanical or hydraulic methods such as mineral flushing oil containing no addi- dynamometers, prony brakes, etc. Another tives. Drain the flushing oil and refill GEAR DRIVES

with the recommended lubricant to the can be monitored by mounting two radial vibra- proper level. tion probes adjacent to each radial bearing and 7. After the initial 50 hours of operation one axial position probe for each shaft. Reference check all coupling alignments and re- C51 gives the following allowable level for the torque all bolts. Check all piping connec- double amplitude of unfiltered vibration in any tions and tighten if necessary. plane measured on the shaft adjacent to a radial If starts are made in a cold environment con- bearing: sideration should be given to pre-heating the 2.0 mils or the value defined by the following lubricant. Load should not be applied until the equation, whichever is less. lubricant has attained operating temperature.

Condition Monitoring Double amplitude including - 12000 '? The most efficient method of maintaining a piece runout in mils - (m) of equipment is to base repairs on the machine's I condition rather than overhaul at an arbitrary Temperature Measurement time period. This type of predictive maintenance Monitoring of the oil in and out temperatures can requires dependence upon instrumentation and provide information about the gearbox and lubri- the proper interpretation of the data it provides. cation system condition. Also, bearing scavenge Reference C61 provides a discussion of basic oil temperature and babbitt temperature can be dynamic motion, vibration and rotor position monitored. Drastic changes in temperature parameters, that should be measured and ana- parameters require investigation. lyzed. In addition to vibration and shaft position other parameters that are important to monitor Pressure Measurements are oil and bearing temperature, oil condition, Oil pressure into the gearbox should be moni- and oil pressure. tored to insure proper flow. The oil filter should have a differential pressure indicator which is Vibration Monitoring monitored to determine whether the filter is Placement of vibration pick-ups at 90" to one an- becoming clogged. other is common on gearbox casings particularly near the high speed shaft. The parameters moni- Oil Inspections tored are generally amplitude of vibration in peak Laboratory inspection of the lubricating oil can to peak mils displacement, velocity in inches per provide an indication of oil deterioration and second or peak G's acceleration. Casing measure- gearbox condition. An oil sample can be analyzed ments are utilized to attempt to lead to the early for proper viscosity and acid number. Also, a discovery of malfunctions. A normal operating spectographic analysis can be performed which machine will generally have a stable amplitude indicates the quantity and type of metallic par- reading of an acceptable low level. Any change ticles in the lubricant. This type of analysis, if in this amplitude reading indicates a change of performed periodically, can identify a component machine condition and should be investigated. In that is wearing. analyzing machine operation spectrum analysis of the vibration pick-up can be used. Figure 13 C61, illustrates a comparison between two identical Shut Down Switches gearboxes. The top spectrum plot depicts a gear- In some cases gearboxes are provided with instru- box in good mechanical condition with reason- mentation that automatically shuts down the ably low acceleration levels and a normal mix- equipment when predetermined levels are ex- ture of components. A similar measurement on a ceeded. Alternatively, warning lights can be acti- unit that had cracked pinion teeth is presented vated. Parameters that are generally monitored on the bottom plot. Note the high amplitudes at are: gear mesh frequency, the seventh harmonic of High oil "in" temperature pinion rotational speed, and even the pinion run- Low oil "in" pressure ning speed (Xhs)at 4.2 G's. Filter differential pressure Correlation of vibration and position data, Position Measurements temperatures, pressures and other external pa- Proximity probes, when mounted rigidly to a rameters which effect gearbox operation will bearing cap or casing of a gearbox, provide a lead to a good predictive maintenance program. vibration measurement of the relative motion Analysis of all the data will provide an indication between the shaft and the mounting of the prox- not only of equipment malfunction but also which imity probe. Bearing wear, both radial and axial, component is deteriorating. CHAPTER 11

FREOUENCY - Cyc!es/Minute x 10-3 Figure 13. Comparison Between Gearbox Casing Acceleration Characteristics GEAR DRIVES

MAINTENANCE every precaution must be taken to prevent foreign The objective of a maintenance program is to matter from entering the gearbox. The introduc- insure satisfactory gearbox performance at all tion of moisture, dirt or fumes can lead to sludge times and maintain the transmission in a state formation and deterioration of the lubrication of readiness if not in operation. A program should and cooling system. be planned which includes regular maintenance A recommended maintenance schedule is pre- actions and monitoring of operating character- sented in Table 6. Logs should be kept of instru- istics to determine whether the gearbox condition ment readings and maintenance action to keep a is deteriorating. running account of gearbox condition. There are several maintenance actions that The major causes of gear unit failure are im- should be effected during initial operation of the proper lubrication and overload. Care must be gearbox. After approximately 50 hours of opera- taken to check for proper oil level before operation check coupling alignments and retorque all tion. Excessive oil volume can be as detrimental bolts. Check all piping connections and tighten as lack of lubrication. Too much oil will result in if necessary. Also, after approximately 50 hours churning and overtemperature of components. of operation, the oil in a new unit should be Overload can be a result of vibration, shock, changed. This oil can be filtered through an ele- high torque at low speed, etc. If there is a possi- ment with the same micron rating as the gearbox bility that operating loads will exceed rated unit filter and reused. Particles may be found in the loads the manufacturer should be consulted. oil due to the wearing in process. At this point the In many gear units the teeth can be visually sump or reservoir should be thoroughly cleaned. inspected by removing inspection covers. When After draining the original oil it is recommended opening these inspection covers care must be that the gearbox and lubrication system be taken to insure that no foreign material enters flushed out with a flushing oil. If possible bring the gearbox. Gear teeth should be examined under the unit up to operating speed with light load good lighting and be wiped clean of oil to prevent after filling with flushing oil. Shut down immedi- a false diagnosis. The contact pattern should ately after achieving full speed. Then drain the cover approximately 80% of the tooth. Each gear flushing oil and refill with the recommended tooth should be examined for evidence of the con- lubricant to the proper level. ditions described in Table 7. When performing maintenance operations Gear teeth should be inspected for nicks, burrs

- -- Table 6. Scheduled Maintenance Actions

FREQUENCY 1 MAIN1ENANCE ITEM I CORRECTIVEACT ION Daily Check oil temperature If there IS a drast~c and pressure at oper- change from prevtous ating conditions. readings stop unit and determine cause. Check for noise, vibra- Add oil if necessary. tion and oil leaks. Check sump oil level. Weekly Check 011filter. Change filter element if necessary. Monthly Check lubricating oil Drain and refill lube for contamination. system if necessary. Change oil filter. Check all gauges. con- trols, alarm systems. Clean breather element. 'Every 2500 Change lubricating hours or 6 system oil. months 'If operating conditions are unusually severe such as high temperature or high molsture atmospheres, oil change requirements mlght be more frequent. Changes can be based on inspection of the oil for viscosity or acid number in such cases. CHAPTER 11

Table 7. Visual Inspection of Gears

CONDlTlON DESCRIPTION LIMITS ILLUSTRATION

Wear A general term describing loss Heavy wear to theextent that the of material from contacting working tooth surface is dis- surfaces of gear teeth. torted is unacceptable. Light wear is acceptable.

Pitting A failure of the material charac- Light pitting isacceptable. Any terized by the removal of metal other condition of pitting 1s and the formation of cavities. unacceptable. These cavit~esmay be small and not progress: they may be small initially and then combine or increase in slze by continued fatigue: or they may be of con- siderable size at the start.

Spalling Large cavities which may result Spalling is unacceptable. from a material or processing defect or the forming of a large pit from several smaller pits.

Frosting A surface distress condition Light frosting is acceptable. caused by welding of high spots during the break in period.

Scoring The removal of metal in direc- Light scoring isacceptable tion of sliding due to a break- down of the lubricant film.

End Loading A wear, pitting or scoring con- Light end wearing, scoring or dition that is predominantly On pitling is acceptable. one end of the toothcaused by misalignment of the mating tooth surface.

Breakage Bendingtatigue failure due to Breakage is unacceptable. overload, misalignment,or from inadvertent stress concentra- t~onssuch as notches in surface or sub-surface material defects.

Cracking Cracking is due to metallurgical Unacceptable or process problems

NOTE If condit~onssuch as p~ttlng,scoring or wear are noted, period~cinspectionsshould bescheduled to determine the progression of these conditions GEAR DRIVES

Table 8. Journal-Bearing Trouble-Shooting Chart

CAUSE REMEDY CAUSE REMEDY I WEAR Ill OVERHEATING SCORING 1 lnsulflclenl Increase boar~ngclearancedecrease011 GALLING 011film v~scosltyadd more 011-dlstr~but~ng 1 D~rt Change o~l~nstall new f~lterremove grooves lncrease supply pressure source of eontam~nantprovide Ilush- lncrease 011or~f~ce size lncrease Ing chamfers at end of Oilgr00WS use chamfer s~ze solter bear~ngmaterlat taembed small 2 Improper 011 Change lo proper v~scos~tyfor bearlng amounts of dirt or a harder lournal lo v~scos~ty load. speed and temperature reduce 11seffect 3 Tool~ttle Scrape or machlne out bear~ngbore 2 Mlsal~gn- Al~gnlournal wllh bearlng clearance men1 4 Heat from Prov~dec~rculat~ng-011 system wlth 3 Excesswe Provlde adequalelournal support. surroundings external coollng. Increase 011 flow use journal reduce load causlng dellecfcon more viscous 011 watercool~ng~nsulate deflect~on bear~ngreduce amblenl temperature 4 Boundary- Increase 011vlscoslty use antlwear or with bettervent~lal~on 11lm extreme-pressure 011or grease change 5 Hlgh load Reduce load and bell tension usecom- condltlon to bearing rnater~alwfth better compounded 011contalnang olllness or pat~b~l~lychange bearlng deslgn lor extreme pressureagents Change lo greater 011-lllmth~ckness stronger bearlng materlal and a bear~ng 5 TOOlow 011 Use heav~er-vlscos~ty011 Reduse bear- deslgned for h~gherloads v~scos~ty Ing operating temperature byeoollng 6 Hlgh speed Decrease 011v~scos~ly lncrease clear- 011 increase 011llow to lower the Bear- ance and 011flow Change to ell~ptlcal Ing temperature. rnodlty bearlng overshot or multlgroove bearlng deslgn to gkve higher load Capactly decrease journal size 6 Rough Pollsh or grind lournat IV FATIGUE lournal I Overload Use larger bearlng reduce load. use 7 Pmr Select compat~blemater~als, tncrease stronger bearlng mater~al bearbng- shall hardness or use hardened sleeve 2 Edge loadlng Correct shaft dellect~onscorrect bear- lournat on snaft Use add~tlve-type011s Ing al~gnmentcheck machln~ngol materlal bearlng and journal comblnat~on 3 Local h~gh Check lor corros~onfactors electr~cal 8 Loadm lncreale 011v~scos~ty Increase bearlng stress plts rough surface lln~shuse h~gher- hlgh size use stronger bearlng materlal concen- vlscoslty 011 with hardened shalt useantlwear cf tratlon extremepressure a11 4 Hlgh bearlng Reduce operating temperature use 9 V~bratlm Balance rotor Increase 011vls&OS~ly lemperature heav~ero~lchange bearlng deslgn lor use pressure bearlng three-lobeoc hlgher 011llow change to hlgher- ell~pt~caltype temperature bearlng malerlal 10 Journal Rwlace lournal 5 Rotor Balance rotor use stab~l~z~ngbear~ng eccenfrc~ly vlbrat~on des~gnof elllptlcal pressure or three- II Improper Take gra3ve4 Oul of loaded zone lob^ type Change rotorsupport todamp groovlng ~nbearlflg out vlbratlon 12 Efectr~cal Remove source of current flowing 6 Cav~tatlon flelocale 011holes and groovlng use plttlng thrqh bear~ngeleclr~cally lnsulale erosion h~gherv~scos~ty 011 and lower bear~ng bearlng ground journal temperature use latlgue-res~slant 13 lnsuff~c~ent Add a11 check pump filler cooler bearlng mater~al lubr~cat~on ptplng ~ncreaseo~l-~nlel.pnssure V SEIZURE I Tool~ffle Provlde adequate clearance taklng Into clearance account any lemperature effects and thermal expansion I1 CORROSION 2 lnsuff~c~ent Supply larger amounts of 011use h~gher 1 Colros~ve Remove contamlnatlon source atmos- lubr~cat~on v~scos~ly011 useextreme-pressure Contami phenc or olherwlse seal bearlng use lubr~cantuse proper deslgn ol 011-feed naflon corros~onres~slanl lubr~cat~on grooves and or~l~ces 2 Mo~sture Use rust ~nh~b~ted011 remove source 3 Too nlgh toad Increase bear~ngarea and modlfy 01 mo~?lture bear~ngdeslgn locarry load useh~gher- 3 ~orroslve Change 011per~odlcally loavold bulld up vlscoslty 011 use compounded 011 Cubr~cant olcorros~veoxldaf~on products avo~d 4 Overheating Provlde external cool~ngand ample possibly corros~veextremepressureo~ls 011Ilow 11not necessary change to non- 5 Unsuitable Use morecompat~blebearlng mater~al corros~blebearlng rnaterlal such as fln bear~ng wlth Satisfactory strength and thermal babbitt or alum~num mater~al propert~es CHAPTER 11

CAUSE REMEDY CAUSE REMEDY

1. OVERHEATED BEARING: j Improper mount~ng Correct d~rtyor off-squareshaft a. Inadequate lubr~cat~on Clean oil holes, filters.and vents. and housing shoulders and Use a fresh lubricant. seats. Avoid br~nelllngcausedby b. Excessive lubricant Use lower-viscos~tyoil, lower 011 pounding on bearing churning level tocenter of bottom ball or k. False brinelling Use vibration mounts for roller, use 011mist machine to isolate from platform c. Inadequate internal Use bearfngof greater loose- during idle periods clearance ness. allow for different~alther- I, sealrub Check for metal bearing seal or mal expansion, reduce ~nterfer- shield rubbtng on shaft.shaft enceof shaft and housing fits. shoulder. or housing correct any housing out of roundness or warping 3 LOSS OF LUBRICANT: d. H~ghseal fr~ction Use reduced spring tension with a. Oilleakagethroughseal Adjust oil level tocenteroflowest seals, lubr~cateseals, switch ball or roller. replace seal. use from rubbing seal to low- double-seal arrangement w~th clearance shield dra~nbetween, eliminate any e. Excess~vepreloading Use gasketsot sh~msto relieve unfavorable air flow by proper axial preload with opposed pair baffles and balancing channels or with two held bearings on a b. Leakageathousing split Use thin layer of gasket cement, shaft subjected to thermal replace housing expansion. Change design to c. Dry, caked residue Use silicone or other high- use only one held bearing temperature grease use 1. Spinning outer ring Use closer housing fit, use steel oxidation-inhibited or synthetic insert in soft aluminum housing, oil or grease, cool oil in external use garter sprdng or rubber cooler, cool bearing housing. holding ring increase oil flow to promote g. Misalignment Correct alignment by shimming cooling p~llowblocks, hous~ngs,or 4. LOOSE BEARING: machines shaftsand bear- a. Shaftdiametertoosmall Turn down shaft.chrome-plate ings in Itne. Check for misaiign- or metallize and regrind togive ment of bearing seats and shaft proper fit. Retighten adapter to and housing shoulders get firm grip on shaft 2. NOISY BEARING. b. Housing bore too large Build up bore with chrome plate VIBRATION: or metallizeand regrind. boreout a. Wrong type 011 Check recommendation of housing and press in sleeve to manufacturer give proper bearing fit (a slip fit b. Insufficient lubrication (See la) on OD is usually des~rableto c. Defective bearing Check lor brinefling, fatigue, allow for differential axial wear, groove wobble, poor cage. thermal expansion of a shaft Replace bearing between two bearings) HARD d. Dirt Clean bearing housing, replace OF SHAFT: worn seals. Improve seal arrangement, eliminate source a. Excessive bearing Use less interference 111on shaft of dirt preload or in housing,select bearingwith e. Corrosion Improve sealing to keep outcor- greater ~nternalclearance where rosive elements. use corrosion- heat conduction expands shaft reststing lubricant and inner bearing ring. Relieve axial preloading by housing 1. Toogreat internal Change to bear~ngw~th smaller shims with either two opposed clearance clearance bearings or two "held" bear~ngs g. Unbalance Balance rotor on one shaft h. Misal~gnment Align (see lg) b. Bearing pinching Scrape housing bore to relieve i. Too loose shaft or Build upshaft or bore with or cocking pinching. Replace or remachine hous~ngftt chrome plate or metallize warped housings,check bearing and regrind seats as source of cocking GEAR DRIVES

and scratches which may be repaired by blending major problems encountered in gear drives, their provided they are minor and not on the working causes and some corrective actions: surfaces of the tooth. The blend may be accom- Table 7 Visual Inspection of Gears plished using a small diesinker type file and an Table 8 Journal Bearing Troubleshooting India or carborundum stone. Crocus cloth should Chart C71 be used for the final polishing. All repairs must Table 9 Rolling Bearing Troubleshooting be finished smoothly. Power tools are not per- Chart L71 mitted for blend repairs. Table 10 Gearbox Troubleshooting Guide Troubleshooting Overhaul A step by step guide often proves useful in trouble- Generally, turbomachinery gearboxes do not shooting. The following group of tables lists the have a specific time period where the unit is dis-

Table 10. Gear Box Trouble-Shooting Chart

------

RECOMMENDED RECOMMENDED PROBLEM INSPECTION CORRECTIVE ACTION PROBLEM INSPECTION CORRECTIVE ACTION

Overheating 1. Oil cooler Check flow of coolant and Shaft Failure 4. High transient Applycouplingscapableof operation oil flow. Measure oil tem- (continued) loading absorbing shocks. Use perature into and out of coupl~ngswith shear plns. cooler. Check cooler 5. Torsional or lat- Adjust system masselastic internally for build up of era1 vibrations characteristics to control depos~tsfrom coolant critical speed location. water. Possibly coupling geome- 2. Is oil level too low Check oil level indicator. try can be modified. or too high 6. Cracksdue to Note cause of fretting and 3. Bearing Make sure bearingsare not fretting correct. Press fits between installation pinched and properly corrosions gear and shaft. adjusted. 4. Grade and condi- Check that oil is specified tion of 011 grade. Inspectoil toseeif it Oil Leakage 1. Exceed oil level Check otl level ind~cator. is Oxidized, dirty or with 2. Is breather open Check 011breather. high sludge content. 3. Are oil drains Check that all oil drain 5. Lubrication Check operation of oil open locations are free and System pump. Make sure suct~on clean. side is not sucking air. 4. Oil seals Check oil sealsand replace Measure flow. Check if oil if worn. Check condition of passages are free. Inspect shaft under seal and pol~sh oil line pressure regulator, I! necessary. nozzles and filters to be 5. Plugs at drains, Apply sealant and tighten sure they are free of levels, and pipe fittings. obstruction. fittings 6. Coupling float Check coupling alignment 6. Housings and Tighten cap screws or and alignment. and adjust end float. caps bolts. If not effective. Shaft Failure 1. Type of coupling Rigrd couplings between remove hous~ngcover and rigidly supported shafts caps. Clean mating sur- can cause shaft failure. faces and apply new seal- Replace with coupllng to tng compound. Reassem- provide required flexibility ble. Check compresston and lateral float. joints by tightening 2. Coupling Realign equipment as fasteners firmly. Alignment required. 3. Excessive Over- Reduce overhung load. hung Load Use outboard bearing or Unusual or 1. Check gears and See Tables 7.8 B 9 replace with higher increasing bearings capacity unit. noise CHAPTER 11

assembled and overhauled. It is more common to BIBLIOGRAPHY note deterioration of a bearing or gear and replace 1. AGMA Information Sheet 217.01, Gear Scoring the particular component at a convenient time. Design Guide for Aerospace Spur and Helical Usually the gearbox is supplied with an operating Power Gears, 1965. and maintenance manual which describes how to assemble and disassemble the unit. If the user 2. AGMA Information Sheet 225.01, Strength of Spur, is not completely familiar with the equipment, it Helical, Herringbone and Bevel Gear Teeth, would be prudent to have a factory representative 1967. accomplish any major component replacements. 3. AGMA Information Sheet 215.01, Surface Dura- bility (Pitting) of Spur, Helical, Herringbone and Bevel Gear Teeth, 1966. Spare Parts 4. AGMA Specification 250.03, Lubrication of Indus- Spare gears or bearings for a gear unit will not trial Enclosed Gear Drives, 1972. necessarily be readily available, particularly if the gearbox design was in some man- 5. American Petroleum Institute Standard 613, Second ner. When purchasing the drive the user should Edition, Special Purpose Gear Units for Refin- request a recommended spares list and determine ery Services, February 1977. what the availability of those parts will be. The 6. Application Notes, Bently Nevada Catalog. user and manufacturer can then arrive at some agreement over what spares will be available 7. Bearing Design and Application, Wilcock and and where they will be stored. Booser, McGraw Hill, 1957.

BIOGRAPHY Peter Lynwander Manager, Drive Systems American Lohmann Corporation At American Lohmann, Mr. Lynwander is respon- sible for the analysis, design and development of enclosed gear drives. Product lines of parallel shaft and planetary high speed transmissions are being introduced under his direction. From 1959 to 1978, he was with AVCO-Lycoming Division, a gas turbine manufacturer, where he headed up the Mechanical Components Group. His responsibilities included turbine engine gears, bearings, seals, clutches and lubrication systems. He established computerized procedures for the analysis and design of gear teeth and par- ticipated in numerous turbine engine power and accessory gearbox programs. Mr. Lynwander has a BSME (1959) and MSME (1964) from the University of Bridgeport and is a registered Professional Engineer. He is the Chair- man of the Wbology Division of the Aerospace Gearing Committee of the American Gear Manu- facturers Association. He has authored numerous publications concerning gearing, seals, and clutches.