Improving Machinery Reliability 3 Episode 6 docx

Machinery Reliability Audits and Reviews 199 A-BtT, = 0 63,000 x HP rPm T = torque = .3 x torque coupling radius T, = axial thrust = in - Ib - - lbs - - diameter lbs thrust per - - tangential driving load x tan Y - - helix 2 A C = - tan Y A = C tan Y B B=DtanY D=- tan Y 200 Improving Machinery Reliability C+D=F A + B = F tan 9 A-B=-TA 2A = Ibs A= lbs B= lbs C= lbs D= Ibs Ibs F' Rated tangential load per helix = * - = 2 F' Service factor with maximum thrust = - largest force = 2 Evaluating Cooling Tower Fans and Their Drive Systems Over the past decades, cooling tower fans have been designed and put into service with diameters exceeding 30 ft. More often than not, the vendor is simply extrapolat- ing his past design by going from 26 ft to 28 ft, 30 ft, or even larger diameters. Extrapolation is generally synonymous with simple scaleup of representative dimen- sions. At other times, only the blade length is increased and the blade hub or blade internals are left untouched. Many of these extrapolations have resulted in costly failures risking extended downtime or injury to personnel. Detailed design reviews are appropriate and the following items represent a cross section of topics. 1. The dynamic natural frequency of cooling fan blades should be at least 20% away from the fan rpm and its multiples. Background: Force amplification resulting from coincidence or near-coinci- dence of blade natural frequency and forcing frequency has caused catastrophic blade failures in many cooling tower installations. Figure 3-81 shows one such event. 2. If urethane-foam filter material is used in constructing fan blades, the vendor should submit data showing dynamic natural frequency of blades after urethane filler material loses intimate bonding, Le., delaminates or separates from blade- skin interior surfaces. Alternatively, the vendor should submit proof that delam- ination or separation between filler and blade-skin interior will not occur with his design. Background: Loss of bonding has been experienced on a large number of blades. The resultant lowering of the blade dynamic natural frequency may cause coincidence or near-coincidence of blade natural frequency and forcing frequency. *F' = actual tangential driving load times lowest calculated service factor (service factor powbly per manufacturer's nameplate). [...]... 7c(4.5 )3 Machinery Reliability Audits and Reviews 2 13 Alternating torsional stress z = (0.2) (9 ,68 0) = 1,940psi , MF = (1 73, 250)(0 .3) = 52,000 lb-in MT = (1 73, 250) (sin 0.057) = 1 73 lb-in Mtotal= -,/(25,025)2+ (52,000 + 1 73) ’ Mtota! 57, 864 lb-in = 0, = (57, 864 ) (2.25) 64 = 6, 472 psi (n:) (4.514 0 = , n= (1 73, 250) (0. 03) = 770 psi (4.5) n: (2.25)2 0.94 1 770)2+ 1.95 6, 472 + 52,500 80,000 = 2.14 3 (2.9 ~ 1,940... of 17 ,60 0 HP Rather than engaging in a debate as to safe maximum absolute values of stress in turbine shafts, we merely investigated what factor of safety the shaft design embodied while equipped with a suitable gear coupling transmitting 17 ,60 0 HP at 6, 400 rpm: Shaft torque T = ( 63 , 000) (17 ,60 0) = 1 73, 250 lb -in 6, 400 16T Steady torsional stress 7, = -= ( 16) (1 73, 250) = 9 ,68 0 psi 7cd3 7c(4.5 )3 Machinery. . .Machinery Reliability Audits and Reviews 209 LE A CONTACT POINT Figure 3- 86 Shift in contact point experienced by gear-type coupling and a misalignment angle calculate ~1 = 0.057' (approximately 0.001 inhn parallel offset), we M, = (1 93, 000) (0 .3) = 57,900 lb - in M = (1 93, 000) (sin 0.057) = 1 93 lb - in , MT,,,~!= J(27,878)2 + (57,900 + 1 93) * = 64 , 435 lb - in This moment is... and I are the shaft radius and shaft area moments of inertia, respectively Thus 0, = (64 , 435 ) (2.25) nd4 /64 - (64 , 435 ) (2.25) (64 ) = 7,2 06 psi n (4 .34 - In addition, there is a mean tensile stress acting on the shaft cross-sectional area This mean stress equates to 0, = - TP ( D /~2 ) (d) COS e (1 93, 000) (0 .3) = 860 psi (4.5) n (2.25)' (0.94) where 8 = 20°, the pressure angle assumed for the gear teeth... all parties involved We have reproduced reliability review checklists listing generalized rotating machinery data review requirements for centrifugal compressors (Figure 3- 99), special-purpose steam turbines (Figure 3- IOO), centrifugal pumps (Figure 3- 101), cooling tower fan systems (Figure 3- 102), and forced and induced draft fans (Figure 3- 1 03) Machinery Reliability Audits for Existing Plants Experience... clearance C to be 0.001 2 0.0005 inch 3 Location H to accept 30 ,000 psi hydraulic fitting 4 O-rings to be Buna-N 90 durometer 5 Machine,wrench flat at 8 6 Outer cup A to have wall thickness consistent with 30 ,000 psi pressure retention requirement 7 Outer cup A to be AIS1 434 0 steel Figure 3- 95 Hydraulic mounting fixture for tapered-bore coupling hubs 222 Improving Machinery Reliability face and coupling-hub... machine reliability Machinery reliability audits should be conducted by mechanical engineers with extensive design and plant machinery experience These audits are generally structured to: 2 26 Improving Machinery Reliability _ _ Pro)ecl Title Project NO Purchase Order NO Inquiry No Dale Ordered ~- llem No ~ ~~ ~ ~ _ Date - ~~ ~ ~ Submit to Owner Before= P b::; ;:; Preliminary Review Figure 3- 100... removal in case rapid access to turbomachinery shaft seals should become necessary (1) Clamp hub t FLE~OR Coupling attaches here (2) Bore Tighten bolts /to “set”clamps sleeve Figure 3- 98 Anderson hub clamp for attaching coupling hubs to cylindrical shaft ends 224 Improving Machinery Reliability How to Keep Track of Reliability Review Tasks As mentioned earlier, machinery reliability reviews are aimed at... a maximum permissible torque of 125 ,66 4 Ib-in This torsional holding requirement is to be achieved with a coupling having a Ld2 = 56 According to Figure 3- 89, the required press-fit pressure is 12,000 psi Assuming further that the coupling d/D ratio is 0 .68 , Figure 3- 90 shows a Figure 3- 91 Temperature change above ambient required to increase hub bore Muchinery Reliability Audits and Reviews 219 required... operating plant will benefit from periodic reliability audits As can be expected, machinery reliability audits look for factors that could have, or in some cases already have had, an adverse impact on machinery reliability, and therefore plant profitability Machinery reliability audits differ from conventional safety and plant operability audits by emphasizing the machinery- related aspects of plant operation, . transmitting 17 ,60 0 HP at 6, 400 rpm: ( 63 , 000) (17 ,60 0) = 1 73, 250 lb -in. Shaft torque T = 6, 400 16T ( 16) (1 73, 250) = 9 ,68 0 psi Steady torsional stress 7, = - = 7cd3 7c(4.5 )3 Machinery. F): (3, +-=- Om 1 OEIkf Gyp where (3 - 4) 1 kf-+L Ga (J GYP n= 2 06 Improving Machinery Reliability "0 .01 .02 . 03 .04 .05 . 06 .07 .08 .09 .IO .I I -12 .I3 Figure. and shaft area moments of inertia, respectively. Thus (64 , 435 ) (2.25) nd4 /64 0, = = 7,2 06 psi - (64 , 435 ) (2.25) (64 ) - n (4 .34 In addition, there is a mean tensile stress acting