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Couplings and Alignment
18 Couplings and Alignment Couplings in most turbomachines attach the driver to the driven piece of machinery. High-performance flexible couplings used in turbomachines must perform three major functions: (1) efficiently transmit mechanical power directly from one shaft to another with constant velocity, (2) compensate for misalignment without inducing high stress and with minimum power loss, and (3) allow for axial movement of either shaft without creating excessive thrust on the other. There are three basic types of flexible couplings that satisfy these requirements. The first type is the mechanical-joint coupling. In this coupling, flexibility is accomplished by a sliding and rolling action. Mechanical-joint couplings include gear tooth couplings, chain and sprocket couplings, and slider or Oldham couplings. The second type is the resilient-material coupling. In resilient-material couplings flexibility is a function of flexing of material. Resilient-material couplings include those that use elastomer in compression (pin and bushing, block, spider, and elastomerannulus, metal-insert types); elastomer in shear (sandwich type, tire type), steel springs (radial leaf, peripheral coil types) and steel-disc and diaphragm couplings. The third type is the combined mechanical and material couplings where flexibility is provided by sliding, or rolling and flexing. Combination couplings include continuous and interrupted metallic-spring grid couplings, non-metallic gear couplings, nonmetallic chain couplings, and slider couplings that have non-metallic sliding elements. In choosing a coupling, the loading and speed must be known. Figure 18-1 shows the relation between coupling type, peripheral velocity coupling size, and speed. The loadings in these high-performance flexible couplings are as follows: 1. Centrifugal force. Varies in importance, depending on the system speed. 2. Steady transmitted torque. Smooth non-fluctuating torque in electric motors, turbines, and a variety of smooth torque-absorbing load (driven) machines. 3. Cyclically transmitted torque. Pulsating or cyclic torque in reciprocating prime movers and load machines such as reciprocating compressors, pumps, and marine propellers. 4. Additional cyclic torque. Caused by machining imperfections of drive components (particularly gearing) and imbalance of rotating drive components. 5. Peak torque (transience). Caused by starting conditions, momentary shock, or overload. 6. Impact torque. A function of system looseness or backlash. Generally, mechanical-joint flexible couplings have inherent backlash. 7. Misalignment loads. All flexible couplings generate cyclic or steady moments within themselves when misaligned. 8. Sliding velocity. A factor in mechanical-joint couplings only.
Gas Turbine Engineering Handbook, Fourth Edition. DOI: 10.1016/B978-0-12-383842-1.00018-4 c 2012 Elsevier Inc. All rights reserved. Copyright
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9. Resonant vibration. Any of the forced vibration loads, such as cyclic or misalignment loads, may have a frequency that coincides with a natural frequency of the rotating-shaft system, or any component of the complete power plant and its foundation, and may, thus, excite vibration resonance.
20,000
30,000
15,000
10,000
Speed in rpm 7,500
800 (244)
700
Maximum speed range for “high-speed” gear and metal-flexing couplings
600
Peripheral velocity at coupling flange O.D. – ft/s (mps)
5,000 500
400 (122)
3,600
300 Maximum speed range for standard gear and metal-flexing couplings 200
1,800
Balancing recommended for standard couplings
100
0 4 (102)
6
8
10
12
14 16 (356)
18
Coupling flange O.D. – inches (mm)
Figure 18-1 Flexible coupling operating spectrum.
20
22
24 (610)
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Table 18-1 Disc, Diaphragm, and Gear Couplings*
Speed capacity Power-to-weight ratios Lubrication required Misalignment capacity at high speed Inherent balance Overall diameter Normal failure mode Overhung moment on machine shafts Generated moment, misaligned, with torque Axial movement capacity Resistance to axial movement Suddenly applied Gradually applied
Disc
Diaphragm
Gear
High Moderate No Moderate
High Moderate No High
High High Yes Moderate
Good Low Abrupt (fatigue) Moderate
Very good High Abrupt (fatigue) Moderate
Good Low Progressive (wear) Very low
Moderate
Low
Moderate
Low
Moderate
High
High High
Moderate Moderate
High Low
*This table is intended as a rough guide only.
The gas turbine is a high-speed, high-torque drive and requires that its coupling has the following characteristics: 1. 2. 3. 4.
Low-weight, low-overhung moment High-speed, capacity-acceptable centrifugal stresses High balancing potential Misalignment capability
Gear couplings, disc couplings, and diaphragm-type couplings are best suited for this type of service. Table 18-1 shows some of the major characteristics of these types of couplings.
Gear Couplings A gear coupling consists of two sets of meshing gears. Each mesh has an internal and external gear with the same number of teeth. There are two major types of gear couplings that are used in turbomachinery. The first type of gear coupling has the male teeth integral with the hub as seen in Figure 18-2. In this coupling type the heat generated at the teeth flows in a different way into the shaft than it does through the sleeve to the surrounding air. The sleeve will therefore heat up and expand more than the hub. This expansion plus the centrifugal force acting on the sleeve will cause it to grow rapidly – as much as 3–4 mils more than the hub – causing an eccentricity, which can lead to a large, unbalanced force. Thus, this coupling type is more useful in low-horsepower units.
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Type HSIG grease-packed
Figure 18-2 Gear coupling (male teeth integral with the hub).
T
Figure 18-3 Gear coupling (male teeth integral with the spool).
The second type of coupling, shown in Figure 18-3, has the male teeth integral with the spool. In this coupling type the same amount of heat is produced, but the hollowbored spool will accept heat in a manner similar to the sleeve so that no differential growth occurs. Gear couplings have a pilot incorporated into the male tooth form to support the loose member of the coupling in a concentric manner at speed, as shown in Figure 18-4. The sliding friction coefficient is another area of evaluation in gear couplings. It produces a resistance to the necessary axial movement as rotors heat and expand. This relative sliding motion between the coupling elements takes care of the misalignment problem in gear couplings. Relative motion between meshing gears is oscillatory in the axial direction and has a low amplitude and a relatively high frequency. Some of the major advantages of the gear couplings are: 1. They can transmit more power per pound of steel, or per inch of diameter, than any other coupling.
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Major diameter fit
Sleeve
Hub
Figure 18-4 Schematic of gear used in coupling applications.
2. They are forgiving; they accept errors in installation and mistreatment more readily than other types of couplings. 3. They are reliable and safe; they do not throw around pieces of metal or rubber even when they fail, and they can work longer in corrosive conditions than many other couplings.
A major disadvantage in gear couplings is the misalignment problem. Tooth-sliding velocity is directly proportional to the tooth-mesh misalignment angle and the rotational speed. Therefore, misalignment of high-speed drives must be kept to a minimum to limit sliding velocity to an acceptable value. The coupling must be able to accommodate misalignment caused by cold startup. The physical misalignment capability of a gear-type coupling should never be considered an acceptable running condition for high-speed applications. The limits of misalignment versus operating speed are best stated on the basis of a constant, relative sliding velocity between the gear teeth. Figure 18-5 gives recommended limits of misalignment with the system at operating temperature. The graph is based on a maximum constant sliding velocity of 1.3 inches per second and includes coupling size, speed, and the axial distance between gear meshes. Gear couplings can be more tolerant of axial growth than other coupling types. In the disc-type couplings, the axial growth is limited by the disc deflection range, so the equipment must be adjusted with more axial accuracy than with gear couplings. High-speed couplings must be balanced very carefully and, with a low overhung moment. The effect of the coupling overhung moment is felt not only in the machine bearing load but in the shaft vibration. The advantage of a reduction in overhung moment is not only to reduce bearing loads, but also to minimize shaft deflection, which results in a reduction of the vibration amplitude. The reduction of the coupling overhung moment produces an upward
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0.026 Method 1
Method 2
0.024 0.022
Shaft B
Shaft A
Shaft B
Shaft A
0.020
Misalignment component
L
L
0.018
NOTE: Where “L” is very large, this method may prove more practical than Method 1. However, L must be equal to or greater than S (where S equals coupling size).
NOTE: Method 1 is valid for all cases, including when L = 0; that is, a standard coupling without spacer.
0.016 0.014
Misalignment measurements These MUST be taken from “Shaft A” to “Shaft B”, then from “Shaft B” to “Shaft A” (see illustrations). The LARGEST of these measurements is compared with the limit calculated from the graph.
0.012
Use of graph 1. Establish maximum coupling rpm. 2. With this value, enter graph, and where this value meets, the curve read off “Misalignment component” *3. Multiply this “Misalignment component” value by “L/2S + 1” (where S = coupling size; L as shown in illustrations). 4. The value thus obtained represents the maximum recommended T.I.R. diameter runout.
0.010 0.008 0.006
*NOTE: For marine spool-type coupling the max. recommended T.I.R. = (L/2S − 0.3) × misalignment component.
0.004
40,000
38,000
36,000
34,000
32,000
30,000
28,000
26,000
24,000
22,000
20,000
18,000
16,000
14,000
12,000
10,000
8,000
6,000
4,000
2,000
0.002
Operating speed
Figure 18-5 Recommended limits of misalignment vs. operating speed (Reference 3).
shift in shaft critical speeds. This change in natural frequencies results in an increase in the spread between natural frequencies. For many applications, reduced overhung moment is an absolute necessity to enable the system to operate satisfactorily at the required operating speed. The high-speed couplings have five components – usually two hubs, two sleeves, and a spacer. To obtain a proper balance, each hub should be balanced separately, then the spacer should be balanced, and finally the full coupling should be assembly balanced. The couplings should be carefully match-marked before removal from the balancing mandrels. Lubrication problems are a major consideration in the use of gear couplings. Relative sliding between the teeth of the hub and the sleeve requires proper lubrication to assure long component life. This sliding motion is alternative and is characterized by small amplitudes and relatively high frequencies. Gear couplings can be either packed with lubricant or continuously lubricated. Each system has advantages and disadvantages, and the choice depends on the conditions under which the coupling works.
Oil-Filled Couplings Very few high-performance couplings use this system because it requires large-volume couplings. It is, however, the best method of lubrication and, incidentally, the first
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used. Its major disadvantage is that it may leak lubricant from defective flange gaskets, etc.
Grease-Packed Couplings Besides enabling the user to select a good lubricant, grease-packing has the advantage of sealing the coupling from the environment. The high-performance coupling works under very small misalignment and usually generates very little heat. In most cases, the couplings receive more heat from the shafts than they generate. Very few greases can work in temperatures of more than 250 ◦ F (121 ◦ C), and for this reason greasepacked couplings cannot be installed within an enclosure that prevents the heat from dissipating. Greases also separate under large centrifugal forces. In many high-speed couplings forces exceed 8,000 g’s. New lubricants are appearing on the market that do not separate under high loadings. A second disadvantage of grease lubrication is the maintenance requirement. Coupling manufacturers generally recommend re-lubrication every six months. There are known cases, however, where grease-packed couplings were found to be in excellent condition after two years of maintenance-free service.
Continuously Lubricated Couplings Lubrication by continuous oil flow can represent an ideal method if there is: 1. Freedom to select the type of oil. 2. An independent lube circuit.
From the user’s point of view, neither condition is acceptable, not only because of the added cost of an independent lube circuit, but because it is almost impossible to prevent mixing of the oil from this circuit with the lube system for the rest of the equipment. In practice, continuously lubricated couplings are supplied with oil from the main lube system. The oil is not the best type for couplings, and also brings a large quantity of impurities to the coupling. The accumulated sludge shortens coupling life. Sludge accumulates within a coupling for two reasons: (1) because the lubricant is not pure, and (2) because the coupling centrifuges and retains the impurities. Very little can be done to prevent the coupling from retaining the impurities. The g forces in a coupling are very high, and the oil dam built in the sleeve configuration prevents the impurities from going over it. Some manufacturers now offer couplings without a dam, or with sleeves provided with radial holes. Experience has shown that such couplings accumulate no sludge. The dam has, however, two useful purposes: 1. It maintains an oil level high enough to submerge the teeth completely. 2. It retains a quantity of oil within the coupling even if the lube system fails.
Removing the oil dam defeats both these features. To maintain the same performance for a damless coupling, the oil flow to the coupling should be re-evaluated. Nothing can be done, however, to retain oil in the damless coupling, and some users
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will not accept them for this reason. A proper decision can only be made by weighing a possible coupling failure because of sludge accumulation against an accidental failure of the lube system.
Gear Coupling Failure Modes The main causes of failure in gear couplings are wear or surface fatigue caused by lack of lubricant, incorrect lubrication, or excessive surface stresses. Component fracture caused by overload or fatigue is generally of secondary importance. High speeds require relatively lightweight gear elements. All case-hardening procedures produce distortion – to keep this distortion to a minimum, nitriding is the preferred hardening method. This method is employed after all machinery operations are complete and no further corrections are to be made to the tooth geometry. Nitriding permits increased tooth loading. The amount of increased capacity is not exactly known, but a 20% increase in load at 10,000–12,000 rpm has proven reliable. A further advantage of the nitrided coupling is that the coefficient of friction is lower than that for through-hardened parts. The heat from friction in the coupling decreases. More important, the transmission of axial forces is decreased by the reduced friction. In many cases, gear shaving prior to nitriding has been used to correct or minimize small errors of tooth geometry caused by the shaping or hobbing processes. A method of assuring nearly perfect tooth contact is to match-lap the gear teeth after nitriding. Lapping eliminates the break-in period, which otherwise takes from 70 to 120 hours. It is during the break-in period, that tooth surface distress usually occurs. For maximum reliability, it is recommended that nitrided gear teeth be specified. Experience indicates that the extra cost of match lapping is justified. The major failure in gear couplings is the fretting on the gear teeth. Fretting can be caused by improper lubrication. Lubrication problems can be categorized by the type of lubrication system being used. The two types of lubrication systems are the batch type and the continuous lubrication type. Table 18-2 shows some of the common problems that affect gear couplings, depending on the lube system used. Misalignment is another problem with gear couplings. Excessive misalignment can lead to any of the following problems, such as: tooth breakage, scoring, cold flow, wear, and pitting. Fasteners are another problem source in couplings. Table 18-2 Types of Typical Gear Coupling Failures Standard or Sealed Lube
Continuous Lube
Wear Fretting corrosion Worm tracking Cold flow Lube separation
Wear Corrosive wear Coupling contamination Scoring and welding Worm tracking
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Table 18-3 Diagnostic Analysis of Gear Couplings Damage or Stress Signs
Cause
Gear tooth surface deterioration (high rate of wear, scoring, and worm tracking) Gear tooth surface deterioration and overheating Tooth breakage and wear Broken hub, keys sheared Lockup-worn and broken teeth
Low oil viscosity and/or excessive misalignment
Worm tracking Broken end or seal ring Galled bores
Discolored bores
Fracture of components Cold flow, wear, and fretting Bolt shearing, bolt hole elongation Separation of lubricant ingredients Retention of moisture impurities Lubricant deterioration
Misalignment, high sliding velocity High misalignment angle Too much shrink fit on shaft Contaminated lubrication system, excessive misalignment Misalignment, separation of lubricant, low oil viscosity Too much shaft-to-shaft spacing and misalignment Improper removing techniques, insufficient or incorrect heating, excessive interference fit Improper hydraulic fit, contamination between shaft and hub Overload or fatigue, shock loading High vibration Nut bottoming out on threads Centrifugal force Centrifugal force High ambient temperature
Coupling fasteners should be properly heat-treated to withstand the large forces they experience in high-speed coupling applications. Fasteners should be properly torqued and, after four-to-six disassemblies, the entire fastener set should be replaced. Bolt shearing or bolt-hole elongation results from the nut bottoming out on the threads before the coupling flanges are tight, thus transmitting force through the bolt rather than through the flange faces. Bolts and nuts should be weight-balanced to very close tolerances. Table 18-3 is a diagnostic analysis of gear coupling failures.
Metal Diaphragm Couplings The metal diaphragm coupling is relatively new in turbomachinery applications. Although the first recorded use of such a coupling dates back to 1922 on a condensing steam turbine locomotive, the contoured diaphragm did not come into wide use until the late 1950s. Diaphragm couplings accommodate system misalignment through flexing. Fatigue resistance is the main performance criterion. The life expectancy of a diaphragm
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coupling that operates within its design limits is theoretically infinite. Figure 18-6 is a photograph of a typical metal diaphragm coupling. Figure 18-7 shows a section through a diaphragm coupling. The coupling has only five parts: two rigid hubs, one spool piece, and two alignment rings. These five parts are solidly bolted together, and misalignment is accommodated through flexing of the two diaphragms of the spool. The spool piece is made up of three separate parts: two diaphragms and a spacer tube. These parts are welded together by an electron beam. The heart of these couplings is the flexing disc; it is manufactured from vacuumdegassed alloy steel, forged with a radial-grain orientation, and has a contoured profile machined on high-precision equipment. The contoured profile is shown in Figure 18-8. The diaphragm undergoes axial deflection. The forces acting on the disc that are generating the stresses are caused by the torque effects, centrifugal forces, and axial deflection. Standard methods for calculating centrifugal forces in a rotating disc show that both tangential and radial stresses increase rapidly with a decrease in the radius. The stresses imposed by axial deflection are much greater at the hub than at the rim, as seen in Figure 18-9. Therefore, to maintain uniform stresses in the diaphragm when all the various forces acting on the diaphragm are at their maximum, the diaphragm must be used to connect the contoured profile at both the hub and the rim to reduce stresses. Diaphragm couplings are more susceptible to axial movement problems than gear couplings, since the diaphragm has a maximum deflection that cannot be exceeded. Theoretically, a diaphragm coupling will have no problems or failures as long as it is operated within “design limits.” The diaphragm fails from excessive torque. Two distinct modes of failure can be found – one at a zero axial displacement and the other at a large axial displacement. Zero axial displacement is characterized by a circular crackline that goes through the thinnest portion of the diaphragm. The crack
Figure 18-6 Metal diaphragm coupling, one end shown (Courtesy of K´oppers Company, Inc.).
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Flex unit Removable assembly Flexible diaphragms Guards Tube
Flanged hub
High-strength body-bound bolts All metal locknuts
Figure 18-7 Schematic of atypical diaphragm coupling (Courtesy of K´oppers Company, Inc.).
Figure 18-8 Axial deflection in a disc.
is relatively smooth, and there is no buckling of the disc. The large axial movement and angular misalignment, which lead to disc failure, are characterized by a crackline that follows a random path from the thinnest to the thickest portion of the disc. The crackline is very irregular, and there is severe buckling of the unfailed part of the disc. Failure in this mode shows that the crackline propogates some 270◦ before disc buckling takes place, indicating that the torque load makes only a small contribution to the total stresses in the disc. Metal diaphragm couplings can also have problems due to corrosive action on the diaphragms. Thus, care must be taken to apply coating to protect against damage from a harsh environment.
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a
b
Figure 18-9 Stress distribution under axial deflection.
Hub
Spacer
Hub
Flexing element
Figure 18-10 Typical metal-flexing disc coupling.
Metal Disc Couplings The main difference between the metal diaphragm coupling and the typical metalflexing disc coupling is that a number of discs replace the single diaphragm between the hubs and the spacer. Figure 18-10 shows a schematic of this type of coupling. A typical metal-flexing disc coupling consists of two hubs rigidly attached by interference fit or flange bolting to the driving and driven shaft of the connected equipment. Laminated disc sets are attached to each hub to compensate for the misalignment. A spacer spans the gap between the shafts and is attached to the flexing elements at each end. The functional requirements and characteristics of the flexing elements are to transmit rated torque as well as any system overloads without buckling or permanent deformation. In other words, they must possess torsional rigidity. However, under conditions of parallel, angular, and axial misalignment, the flexing element must have sufficient flexibility to accommodate these conditions without imposing excessive forces and moments on equipment shafts and bearings. Both of the previous requirements must be met while maintaining stress levels that are safely within the fatigue
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f fx P
P
y
f
fy
x
Figure 18-11 Frictional damping in a metal-disc coupling.
limit of the flexing material. Metal-flexing couplings have been known to exhibit occasional large-amplitude vibrations in the axial direction when excited at the natural frequency of the coupling. The amount of damping present in a metal-flexing coupling is thought to be relatively small, although it is known to be greater for the laminated disc-type construction than for a coupling consisting of a single-piece membrane. The reason for the greater damping in the laminate disc configuration is that under conditions of axial movement, a microscopic amount of motion takes place between adjacent lamina, as shown in Figure 18-11. Since the element is clamped together under a bolt preload, there is a frictional force, which resists sliding. Field experience by manufacturers and users of turbomachinery has shown that resonant axial vibration of a metal-flexing coupling can cause problems at times that are reflected through the entire drive train. With laminated disc couplings, problems occurs only when an external forcing function exists. This condition could be a result of aerodynamic or hydraulic fluctuations in the machine train, out-of-square thrust collars, gearing inaccuracies, or electrical excitations of motor-driven equipment. It is usually possible to avoid operating the couplings at or near resonance if the condition is anticipated during the system design stage. However, such problems do not always occur until after a machine is in service. More information is needed on the nature and magnitude of external excitations.
Turbomachinery Uprates If an existing coupling is to be replaced with a new type of coupling because of a machinery uprate, or for any other reason, there is good justification to review, with the latest techniques, the nature of the rotating system to be coupled. Couplings, whether gear or disc-type, should not be simply picked from a catalog. Some installations are very old, and some have been revised in other ways in the field. Unfortunately, such engineering reviews are not easy to arrange with busy equipment suppliers. Therefore, the tendency is to match the obvious characteristics of the existing coupling and see what happens. Many older designs have relatively heavy and larger-diameter shafts, and retrofits have been very successful and trouble-free. Part of
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this success is due to the consideration given to the retrofit by cooperating engineers of the coupling manufacturer and the rotating equipment manufacturer. A large part of the success is due to the dedication and extra effort of the first companies offering the disc coupling to ensure success. If retrofits and new installations consume the available time of these engineers, the potential for omission increases. Therefore, more time should be allowed for the work. Coupling application is an engineering effort involving the coupling and rotating equipment designers. The user, by the purchasing technique he employs, can aid or hinder this effort, since he chooses the basic coupling style his operations and maintenance people will work with. In either case a good purchase specification should designate that the selection and design of the coupling must follow the rotor design work and exclude the coupling from becoming involved in competitive bids. It is simply too important an item to risk reliability for initial cost savings. Disc couplings are used as replacements for gear couplings for two reasons: (1) the disc couplings do not require lubrication, and (2) the machinery ratings can be uprated with disc couplings. Compressor and driver shafts often prove to be overstressed in equipment uprate situations; however, a change from conventional gear-type couplings to the more recent diaphragm coupling design can lower the shaft stress enough to avoid shaft replacement during power uprates of compressors or compressor drivers. A close examination of how the equipment vendor arrived at his maximum allowable stress levels may frequently show that such shaft replacements can be avoided without undue risk if the coupling selection is optimized. This situation is based on the fact that gear-type couplings have the potential of inducing both torsional stresses and bending stresses in a shaft, whereas diaphragm couplings tend to induce primarily torsional stresses and insignificant bending stresses at best. To determine if a machine’s performance can be uprated without installing a large shaft, the forces acting on the shaft must be computed. The forces acting on a shaft can be put into three separate categories: (1) torsional, (2) axial, and (3) bending forces. Torsional forces are a function of the shaft rotational speed and horsepower transmitted. They can be calculated from: T=
63,000 (hp) rpm
(18-1)
and the torsional stress τ T can be computed with: τT =
16T π d3
(18-2)
It is a generally accepted assumption that the axial stress will not exceed 20% of the torsional stress. τ a can therefore be obtained by τ a = 0.20 τ T . These two stresses
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will be the same for either type of coupling; however, the bending stress will vary depending on which type of coupling is used. There are three relevant bending moments caused by a gear coupling when transmitting torque with angular or parallel misalignment: 1. Moment caused by contact-point shift. This moment acts in the angular misalignment plane and tends to straighten the coupling. It can be expressed: Mc =
X T × Dp /2 2
(18-3)
where T = shaft torque Dp = gear coupling pitch diameter X = tooth face length (Figure 18-12) 2. Moment caused by coupling friction. This moment acts in a plane at a right angle relative to the angular misalignment. It has the magnitude: (18-4)
Mf = Tµ
where µ is the friction coefficient. 3. Moment caused by turning torque through a misalignment angle α. It acts in the same direction as the friction moment Mf and can be expressed as: MT = Tsin α
(18-5)
The total moment is the vector sum of the individual moments: q Mtotal = Mc2 + (Mf + MT )2
(18-6)
The contoured diaphragm coupling causes two bending moments: 1. Moment caused by angular misalignment. This results in bending of the diaphragm: MB = kBα
(18-7)
“X” th Too face
Lever arm
Mc Contact point
Figure 18-12 Shift in contact point.
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Auxiliary Components and Accessories
In this expression kB equals the angular spring rate of the diaphragm (lb-in/degree) and α is the misalignment angle. This moment acts in the angular misalignment plane, as did Mc in the gear-coupling analysis. 2. Moment caused by turning the torque through a misalignment angle α. It can be expressed: MT = Tsin α
(18-8)
The total moment is now: p Mtotal = MB 2 + MT 2
(18-9)
Comparing the bending moments caused by gear couplings with those resulting from contoured diaphragm couplings shows the former to be significant and the latter virtually negligible. The cyclic bending stress imposed on a gear coupling-equipped shaft can be computed from: σa =
Mtotal × C I
(18-10)
where C = shaft radius I = shaft area moment of inertia In addition, there is a mean tensile stress acting on the shaft cross-sectional area. This effect means stress equates to: σm =
Tµ (Dp /2)(π C2 ) cos 2
(18-11)
where 2 is the pressure angle assumed for the gear teeth. The cycle bending stress seen by the diaphragm coupling-equipped shaft can be obtained by a rapid ratio calculation: σ a (diaphragm coupling) Mtotal (diaphragm coupling) = σ a (gear coupling) Mtotal (gear coupling)
(18-12)
The mean tensile stress acting on the cross-sectional area of the diaphragm coupling-equipped shaft depends on how far the diaphragm is displaced axially from its neutral rest position and the axial spring rate of the diaphragm. For combined bending and torsion, the factor of safety can be calculated by the following relationships: n = s
1 σa σm kf + σ e σ y.p.
2
τa τm + 3 kf0 + σ e σ y.p.
2
(18-13)
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where σ e = endurance limit in tension σ yp = minimum yield strength in tension The stress concentration factor kf results from the keyway and must be used in torsional stress calculations. Factor kf0 takes into account the shaft step; it must be used in the bending stress calculation.
Curvic Couplings In essence, the curvic coupling is a ring of precision ground face splines that are meshed after index. The splines, or radial teeth, are ground in such a manner that on one member the sides of the teeth surfaces that meshed are convex and on the other member the sides are concave. The result is that after these members are clamped together, perfect index location is achieved, and further, the turret is perfectly on the center. The other important advantage of the curvic coupling is that its accuracy in both axes actually improves with use, rather than degrades. Care should be taken in the necessity of guarding against the entry of chips or other debris into the seating area. Figure 18-13 shows one-half of a typical curvic coupling, and Figure 18-14 shows a typical wheel in a gas turbine with curvic coupling machined into the disk. Figure 18-15 shows a typical cross section of a gas turbine. Many gas turbine’s rotors are of bolted construction with a positive torque incorporating such features as radial pins and curvic couplings, respectively. The rotor is supported by two-element tilting pad bearings and an upper-half fixed bearing. The thrust bearing is a double-acting type that uses the leading-edge groove lubrication system, as shown in Figure 18-15.
Figure 18-13 A typical curvic coupling.
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Figure 18-14 A typical disc with a curvic coupling.
Figure 18-15 Cross section of a gas turbine.
Figure 18-16 shows a schematic of how the axial compressor disks in the gas turbine are put together by the use of bolts and curvic couplings. These couplings transmit loads over 200 MW and are commonly used in many gas turbines.
Shaft Alignment The successful alignment of a gas turbine to the unit it is driving is of great importance. A major portion of operating problems experienced in the field can often be attributed to faulty misalignment. Operating problems caused by misalignment include excessive vibration, coupling overheat, wear, and bearing failures. Typically, misalignment problems will show up at two times rpm frequencies with axial vibrations at one and two times rpm. With diaphragm-type flexible couplings,
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Spindle bolt
Disc
Torque pin
Disc contact surface
Spindle bolt
Curvic coupling
Figure 18-16 Assembly mechanism of turbine disks using a curvic coupling.
vibrations may be somewhat suppressed, and consequently, trains using these couplings should be monitored periodically to ensure they are in alignment. Perfect alignment – exact shaft colinearity under operating conditions – is difficult and uneconomical to attain. The degree of tolerable misalignment is a function of coupling length, size, and speed. Some companies are now specifying a minimum coupling spacer length of 18 inches, since longer coupling lengths can tolerate more misalignment. The amount of misalignment that can be tolerated by the machine also depends on the types of journal and thrust bearings used. Tilting-pad-type bearings greatly reduce the misalignment problem. Figure 18-17 shows misalignment in both the journal and thrust bearings. The effect of misalignment on a journal bearing causes the shaft to contact the end of the bearing. Thus, journal length is a criterion in the amount of misalignment a bearing can tolerate; a shorter length obviously can tolerate more misalignment. The effect on the thrust bearing is to load up one segment of the thrust bearing arc and unload the opposite segment. This effect is more pronounced with higher loads and less flexible bearings.
The Shaft Alignment Procedure In essence, there are three steps in any alignment procedure. These are: (1) the prealignment survey, (2) cold alignment, and (3) the hot alignment check.
The Prealignment Survey This survey is carried out well ahead of the cold alignment. In this survey; piping, grouting, foundation bolts, shim packs, etc., are studied and ascertained to be appropriately done and of good quality. Again, casing distortion, piping strain, misalignment
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Brg
θ Shaft Line of centers
N
Hm Journal bearing
Thrust brg axis
Shaft Thrust bearing
Figure 18-17 Misalignment in both journal and thrust bearings.
of machine supports relative to the sole plate, etc. are determined, and corrections are made to ensure that these problems will not cause problems with the alignment. Piping strain is by far the greatest problem causer, and so piping should be carefully reviewed to ensure that it is properly done according to the code. Piping strains as high as 0.22 inches (0.5588 cm) have been observed. A typical cause of piping strain occurs when two flanges do not meet and pipefitters force them together. Pipe hangers that are poorly placed or tensioned can also cause significant piping stress problems.
Cold Alignment There are two predominant techniques used for cold alignment. These are: (1) the faceOD method, and (2) the reverse-dial indicator method. Both these techniques utilize dial indicators. For high-speed turbomachinery, the reverse-dial indicator method is the superior method and should be used.
Couplings and Alignment
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Figure 18-18 shows a face-OD indicator setup. As the name indicates, an alignment bracket is attached to one coupling hub, and face-OD readings are taken on the adjacent hub. The face and OD dial indicator readings give an indication of the angularity and offset of the shafts, respectively. The problems with this method are numerous. First, there is the problem of shaft axial float, which makes consistent readings difficult to obtain. Second, inaccuracies in the geometry of the coupling hub have to be taken into account. Third, the face diameter on which the readings are taken is relatively small, and errors are magnified over the length of the machine. The reverse-dial indicator method is shown in Figure 18-19. This method measures just the OD of the coupling hubs or shaft and eliminates the problem of shaft axial float. By spanning the entire coupling, angular misalignment is greatly magnified. For both the face-OD and reverse-dial indicator methods, it is important that sag in the alignment bracket be determined. Figure 18-20 shows a method for the determination of sag. Once the sag is determined, it must be permanently stamped on the bar. The alignment bracket should be considered an important precision tool and must be stored and handled with care so that it may be reused when realignment is required. Once the dial indicator readings are taken, a graphic plot of the two-shaft centerlines can be made on graph paper. It is at this stage that anticipated thermal growths are used in determining the shimming required to obtain shaft colinearity when the units are in the hot condition. Unfortunately, the values supplied by the manufacturers may not be accurate, and pipe strain and other external forces come into play. It is for this reason that the hot alignment check is conducted. A simple graphic plotting exercise for the reverse-dial indicator method shows the basic principles involved. A steam turbine compressor train is shown in Figure 18-21. Assume this train is a new installation and the manufacturer’s estimated thermal growths are as indicated in Figure 18-21. Reverse-dial indicator readings are taken to determine the relative shaft positions. Once readings are taken, the estimated thermal growths are incorporated by shimming, in the hope that a good, hot alignment can be achieved. Bore reading (O.D.) (Gives parallel offset)
Turbine shaft
Compr. shaft
Face reading (Gives angular offset)
Figure 18-18 Face-OD indicator setup.
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Auxiliary Components and Accessories
Dial indicator Bracket Read here Rotate Turbine shaft
Comp. shaft
First reading
Rotate
Turbine shaft
Comp. shaft
Second reading
Figure 18-19 Reverse-dial indicator setup.
The hot alignment check is used to determine the actual thermal growth, and then the final shim changes are made if needed. This example addresses only vertical movements. Horizontal movements are obtained in a similar fashion. The graphic plot uses an amplified scale on the vertical Y axis of one inch equals five mils vertical growth, while the X axis has a scale of one inch equals 10 inches (25 cm) of train length. In this example, it is assumed that Machine A is to be fixed, and all moves are to be conducted on Machine B. As shown in Figure 18-21, a “hot running line” is first drawn. This line is where the shafts should be when the machines are operating. Now, using estimated thermal growth of Machines A and B, a “cold target B” line is drawn. This line is where shaft B should lie so that when hot it will be colinear with shaft A on the hot running line. The next step is to use the dial indicator readings to determine where the shafts actually lie relative to each other. The B-to-A readings show that shaft B lies below shaft A by three mils (half-dial indicator readings) and the A-to-B reading shows that shaft A is above shaft B by five mils. Once these two points are located, shaft B can be plotted. This line is the “actual shaft B” line. Once this procedure is done, the shim changes needed can be easily found and “desired” indicator readings can be given to the millwrights.
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Indicator bar
Lathe center
SAG free
SAG free mandrel
3 mils True
Mandrel
Rotated through 180°
3 mils SAG −6
Figure 18-20 Method for determining sag.
A similar procedure is followed for horizontal movements. If the hot alignment check indicates a significant deviation from expected thermal growths and an unacceptable amount of misalignment, further shim changes can be achieved by similar plotting.
Hot Alignment Check This technique attempts to determine actual alignment status when the machines are hot. When the machines are running, it is impossible to use dial indicator techniques on the shafts.
716
Auxiliary Components and Accessories
5 mils
a
b 8″ 6″ 8″
24″
10″
24″ a nar y
io “Stat
0
0
B to A
A to B
−6
−10
8″
a cL”
6″
5.4
2.4
14.4
36″
14.4
Estimated thermal growths in mils
8″
36″ b
Hot running line
5 mils 3 mils
“Ac t
ual
7.5 mils (remove)
sha
ft
Lcb
”
“Cold target b”
Figure 18-21 Graphic plotting for reverse-dial indicator method.
The old concept of a “hot check” – in which the units were shut down and the coupling disassembled as quickly as possible to allow indicator readings to be taken – should not be used. Currently used, continuously lubricated couplings require significant time to disassemble during which considerable cooling occurs. Because of this factor, a number of hot alignment techniques have been developed. Optical and laser methods, proximity probe methods, and a purely mechanical means using dial indicators may be used for hot alignment checks. In all these methods, an attempt is made to use the cold position of the shaft as a benchmark and then to measure the shaft movement (or bearing housings) from the cold position to the hot position. The objective is to find the change in vertical and horizontal positions at each shaft end. Once this procedure is done along the train, the machines can be shut down and appropriate shim changes made to attain acceptable hot alignment. Basically, the optical method uses equipment such as alignment telescopes, jig transits, and sight levels. Instruments with built-in optical micrometers for measuring displacements from a referenced line of sight enable an accurate determination of target movements, which are mounted on the machine. Optical alignment reference points are located on the bearing housings of the units. A jig transit is then set up at some distance from the train, and readings are taken and recorded in the vertical plane for each reference point in the train. Then the transit is moved, and a similar set of readings are taken in the horizontal plane. This procedure should be done at the same time as the reverse-dial indicator readings are taken. Then, when the train is in its operating condition, another set of readings are taken. The two
Couplings and Alignment
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data sets and the cold alignment dial indicator readings enable the determination of vertical and horizontal growths of each point. The advantages of this system are that it is accurate and, once the reference marks are on the machine, there is no need to approach the machine. However, the equipment involved is expensive and delicate, and great care has to be taken during its use. Moreover, heat waves often cause some problems in taking readings. Alignment with laser techniques has also been used, but the equipment is expensive and can be applied only in certain situations such as for a bearing alignment check. It is used primarily by manufacturers of turbomachinery during fabrication and assembly of their units. Proximity probes have also been used to measure machine movements. Proximity probes are mounted in special water-cooled columns and aimed at “targets” mounted on bearing housings or on other parts of the unit. Changes in the gap distances are then displayed on electrical meters. The Dodd bar system utilizes proximity probes mounted on an air-cooled bar attached between the bearings of the two machines to be aligned. The Dodd bar system allows continuous monitoring of the relative positions of the two shafts. Another system uses proximity probes located within the coupling to continuously monitor the alignment. Digital readouts of misalignment angles, etc., are available from this system. A purely mechanical, hot alignment system utilizing dial indicators has also been developed. The system uses permanently mounted tooling balls made of stainless steel attached to the bearing housing and to the machine foundation. A spring-loaded device with a dial indicator is provided to determine accurately the distance between the two tooling balls. An inclinometer is also provided to give a measure of the angularity. Figure 18-22 shows a typical configuration. Cold readings are taken at the time when the reverse-dial indicator readings are taken, and hot readings are taken when the machine is on-line. These two sets of readings are enough to determine the vertical and horizontal movement of the shaft. The same procedure is followed at each end of the units in the train. Computations can be made either graphically or by a calculator with preprogrammed cards. Direct outputs are the degree of misalignment and the shim changes needed to correct the misalignment. Machine Bearing housing Benchmarks Gauge X
Benchmarks
Foundation
Figure 18-22 Hot alignment system with dial indicator.
718
Auxiliary Components and Accessories
It must be realized that correct alignment is of great importance in attaining high unit availability. Alignment procedures must be carefully planned, tools must be checked carefully, and, in general, great care must be taken during the alignment. The time, effort, and money spent on good alignment is well worth it.
Bibliography Bendix Fluid Power Corp, “Contoured Diaphragm Couplings,” Technical Bulletin. Bloch, H.P., “Less Costly Turboequipment Uprates through Optimized Coupling Section,” Proceedings of the 4th Turbomachinery Symposium, Texas A&M University, 1975, pp. 149–152. Calistrat, M.M., “Gear Coupling Lubrication,” American Society of Lubrication Engineers, 1974. Calistrat, M.M., “Grease Separation under Centrifugal Forces, American Society of Mechanical Engineers, Pub. 75-PTG-3 1975. Calistrat, M.M., “Metal Diaphragm Coupling Performance,” Proceedings of the 5th Turbomachinery Symposium, Texas A&M University, October 1976, pp. 117–123. Calistrat, M.M., and Leaseburge, G.G., “Torsional Stiffness of Interference Fit Connections,” American Society of Mechanical Engineers, Pub. 72-PTG-37, 1972. Calistrat, M.M., and Webb, S.G., “Sludge Accumulation in Continuously Lubricated Couplings,” American Society of Mechanical Engineers, 1972. Campbell, A.J., “Optical Alignment of Turbomachinery,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M Univ., 1973, pp. 8–12. Dodd, R.N., “Total Alignment Can Reduce Maintenance and Increase Reliability,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 123–126. Essinger, J.N., “Benchmark Gauges for Hot Alignment of Turbomachinery,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 127–133. Finn, A.E., “Instrumented Couplings: The What, the Why, and the How of the Indikon Hot Alignment Measuring System,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 135–136. Jackson, C.J., “Alignment Using Water Stands and Eddy-Current Proximity Probes,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 137–146. Jackson, C.J., “Cold and Hot Alignment Techniques of Turbomachinery,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M University, 1973, pp. 1–7. Kramer, K., “New Coupling Applications or Applications of New Coupling Designs,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M University, October 1973, pp. 103–115. Massey, C.R., and Campbell, A.J., “Reverse Alignment-Understanding Centerline Measurement,” Proceedings of the 21st Turbomachinery Symposium, Texas A&M University, 1992, p. 189. Peterson, R.E., Stress Concentration Factors, John Wiley & Son, 1953. Timoshenko, S., Strength of Materials: Advanced Theory & Problems, 3rd ed. Van Nostrand Reinhold Pub., 1956. Webb, S.G., and Calistrat, M.M., “Flexible Couplings,” 2nd Symposium on Compressor Train Reliability, Manufacturing Chemists Association, April 1972. Wilson, C.E., Jr., “Mechanisms – Design Oriented Kinematics,” American Technical Society, 1969.
Couplings and Alignment
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Wright, J., “A Practical Solution to Transient Torsional Vibration in Synchronous Motor Drive Systems,” American Society of Mechanical Engineers, Pub. 75-DE-15, 1975. Wright, J., “Which Flexible Coupling?” Power Transmission & Bearing Handbook, Industrial Publishing Co., 1971. Wright, J., “Which Shaft Coupling Is Best – Lubricated or Non-Lubricated?” Hydrocarbon Processing, April 1975, pp. 191–196.
Gas Turbine Engineering Handbook, Fourth Edition. DOI: 10.1016/B978-0-12-383842-1.00018-4 c 2012 Elsevier Inc. All rights reserved. Copyright
694
Auxiliary Components and Accessories
9. Resonant vibration. Any of the forced vibration loads, such as cyclic or misalignment loads, may have a frequency that coincides with a natural frequency of the rotating-shaft system, or any component of the complete power plant and its foundation, and may, thus, excite vibration resonance.
20,000
30,000
15,000
10,000
Speed in rpm 7,500
800 (244)
700
Maximum speed range for “high-speed” gear and metal-flexing couplings
600
Peripheral velocity at coupling flange O.D. – ft/s (mps)
5,000 500
400 (122)
3,600
300 Maximum speed range for standard gear and metal-flexing couplings 200
1,800
Balancing recommended for standard couplings
100
0 4 (102)
6
8
10
12
14 16 (356)
18
Coupling flange O.D. – inches (mm)
Figure 18-1 Flexible coupling operating spectrum.
20
22
24 (610)
Couplings and Alignment
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Table 18-1 Disc, Diaphragm, and Gear Couplings*
Speed capacity Power-to-weight ratios Lubrication required Misalignment capacity at high speed Inherent balance Overall diameter Normal failure mode Overhung moment on machine shafts Generated moment, misaligned, with torque Axial movement capacity Resistance to axial movement Suddenly applied Gradually applied
Disc
Diaphragm
Gear
High Moderate No Moderate
High Moderate No High
High High Yes Moderate
Good Low Abrupt (fatigue) Moderate
Very good High Abrupt (fatigue) Moderate
Good Low Progressive (wear) Very low
Moderate
Low
Moderate
Low
Moderate
High
High High
Moderate Moderate
High Low
*This table is intended as a rough guide only.
The gas turbine is a high-speed, high-torque drive and requires that its coupling has the following characteristics: 1. 2. 3. 4.
Low-weight, low-overhung moment High-speed, capacity-acceptable centrifugal stresses High balancing potential Misalignment capability
Gear couplings, disc couplings, and diaphragm-type couplings are best suited for this type of service. Table 18-1 shows some of the major characteristics of these types of couplings.
Gear Couplings A gear coupling consists of two sets of meshing gears. Each mesh has an internal and external gear with the same number of teeth. There are two major types of gear couplings that are used in turbomachinery. The first type of gear coupling has the male teeth integral with the hub as seen in Figure 18-2. In this coupling type the heat generated at the teeth flows in a different way into the shaft than it does through the sleeve to the surrounding air. The sleeve will therefore heat up and expand more than the hub. This expansion plus the centrifugal force acting on the sleeve will cause it to grow rapidly – as much as 3–4 mils more than the hub – causing an eccentricity, which can lead to a large, unbalanced force. Thus, this coupling type is more useful in low-horsepower units.
696
Auxiliary Components and Accessories
Type HSIG grease-packed
Figure 18-2 Gear coupling (male teeth integral with the hub).
T
Figure 18-3 Gear coupling (male teeth integral with the spool).
The second type of coupling, shown in Figure 18-3, has the male teeth integral with the spool. In this coupling type the same amount of heat is produced, but the hollowbored spool will accept heat in a manner similar to the sleeve so that no differential growth occurs. Gear couplings have a pilot incorporated into the male tooth form to support the loose member of the coupling in a concentric manner at speed, as shown in Figure 18-4. The sliding friction coefficient is another area of evaluation in gear couplings. It produces a resistance to the necessary axial movement as rotors heat and expand. This relative sliding motion between the coupling elements takes care of the misalignment problem in gear couplings. Relative motion between meshing gears is oscillatory in the axial direction and has a low amplitude and a relatively high frequency. Some of the major advantages of the gear couplings are: 1. They can transmit more power per pound of steel, or per inch of diameter, than any other coupling.
Couplings and Alignment
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Major diameter fit
Sleeve
Hub
Figure 18-4 Schematic of gear used in coupling applications.
2. They are forgiving; they accept errors in installation and mistreatment more readily than other types of couplings. 3. They are reliable and safe; they do not throw around pieces of metal or rubber even when they fail, and they can work longer in corrosive conditions than many other couplings.
A major disadvantage in gear couplings is the misalignment problem. Tooth-sliding velocity is directly proportional to the tooth-mesh misalignment angle and the rotational speed. Therefore, misalignment of high-speed drives must be kept to a minimum to limit sliding velocity to an acceptable value. The coupling must be able to accommodate misalignment caused by cold startup. The physical misalignment capability of a gear-type coupling should never be considered an acceptable running condition for high-speed applications. The limits of misalignment versus operating speed are best stated on the basis of a constant, relative sliding velocity between the gear teeth. Figure 18-5 gives recommended limits of misalignment with the system at operating temperature. The graph is based on a maximum constant sliding velocity of 1.3 inches per second and includes coupling size, speed, and the axial distance between gear meshes. Gear couplings can be more tolerant of axial growth than other coupling types. In the disc-type couplings, the axial growth is limited by the disc deflection range, so the equipment must be adjusted with more axial accuracy than with gear couplings. High-speed couplings must be balanced very carefully and, with a low overhung moment. The effect of the coupling overhung moment is felt not only in the machine bearing load but in the shaft vibration. The advantage of a reduction in overhung moment is not only to reduce bearing loads, but also to minimize shaft deflection, which results in a reduction of the vibration amplitude. The reduction of the coupling overhung moment produces an upward
698
Auxiliary Components and Accessories
0.026 Method 1
Method 2
0.024 0.022
Shaft B
Shaft A
Shaft B
Shaft A
0.020
Misalignment component
L
L
0.018
NOTE: Where “L” is very large, this method may prove more practical than Method 1. However, L must be equal to or greater than S (where S equals coupling size).
NOTE: Method 1 is valid for all cases, including when L = 0; that is, a standard coupling without spacer.
0.016 0.014
Misalignment measurements These MUST be taken from “Shaft A” to “Shaft B”, then from “Shaft B” to “Shaft A” (see illustrations). The LARGEST of these measurements is compared with the limit calculated from the graph.
0.012
Use of graph 1. Establish maximum coupling rpm. 2. With this value, enter graph, and where this value meets, the curve read off “Misalignment component” *3. Multiply this “Misalignment component” value by “L/2S + 1” (where S = coupling size; L as shown in illustrations). 4. The value thus obtained represents the maximum recommended T.I.R. diameter runout.
0.010 0.008 0.006
*NOTE: For marine spool-type coupling the max. recommended T.I.R. = (L/2S − 0.3) × misalignment component.
0.004
40,000
38,000
36,000
34,000
32,000
30,000
28,000
26,000
24,000
22,000
20,000
18,000
16,000
14,000
12,000
10,000
8,000
6,000
4,000
2,000
0.002
Operating speed
Figure 18-5 Recommended limits of misalignment vs. operating speed (Reference 3).
shift in shaft critical speeds. This change in natural frequencies results in an increase in the spread between natural frequencies. For many applications, reduced overhung moment is an absolute necessity to enable the system to operate satisfactorily at the required operating speed. The high-speed couplings have five components – usually two hubs, two sleeves, and a spacer. To obtain a proper balance, each hub should be balanced separately, then the spacer should be balanced, and finally the full coupling should be assembly balanced. The couplings should be carefully match-marked before removal from the balancing mandrels. Lubrication problems are a major consideration in the use of gear couplings. Relative sliding between the teeth of the hub and the sleeve requires proper lubrication to assure long component life. This sliding motion is alternative and is characterized by small amplitudes and relatively high frequencies. Gear couplings can be either packed with lubricant or continuously lubricated. Each system has advantages and disadvantages, and the choice depends on the conditions under which the coupling works.
Oil-Filled Couplings Very few high-performance couplings use this system because it requires large-volume couplings. It is, however, the best method of lubrication and, incidentally, the first
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used. Its major disadvantage is that it may leak lubricant from defective flange gaskets, etc.
Grease-Packed Couplings Besides enabling the user to select a good lubricant, grease-packing has the advantage of sealing the coupling from the environment. The high-performance coupling works under very small misalignment and usually generates very little heat. In most cases, the couplings receive more heat from the shafts than they generate. Very few greases can work in temperatures of more than 250 ◦ F (121 ◦ C), and for this reason greasepacked couplings cannot be installed within an enclosure that prevents the heat from dissipating. Greases also separate under large centrifugal forces. In many high-speed couplings forces exceed 8,000 g’s. New lubricants are appearing on the market that do not separate under high loadings. A second disadvantage of grease lubrication is the maintenance requirement. Coupling manufacturers generally recommend re-lubrication every six months. There are known cases, however, where grease-packed couplings were found to be in excellent condition after two years of maintenance-free service.
Continuously Lubricated Couplings Lubrication by continuous oil flow can represent an ideal method if there is: 1. Freedom to select the type of oil. 2. An independent lube circuit.
From the user’s point of view, neither condition is acceptable, not only because of the added cost of an independent lube circuit, but because it is almost impossible to prevent mixing of the oil from this circuit with the lube system for the rest of the equipment. In practice, continuously lubricated couplings are supplied with oil from the main lube system. The oil is not the best type for couplings, and also brings a large quantity of impurities to the coupling. The accumulated sludge shortens coupling life. Sludge accumulates within a coupling for two reasons: (1) because the lubricant is not pure, and (2) because the coupling centrifuges and retains the impurities. Very little can be done to prevent the coupling from retaining the impurities. The g forces in a coupling are very high, and the oil dam built in the sleeve configuration prevents the impurities from going over it. Some manufacturers now offer couplings without a dam, or with sleeves provided with radial holes. Experience has shown that such couplings accumulate no sludge. The dam has, however, two useful purposes: 1. It maintains an oil level high enough to submerge the teeth completely. 2. It retains a quantity of oil within the coupling even if the lube system fails.
Removing the oil dam defeats both these features. To maintain the same performance for a damless coupling, the oil flow to the coupling should be re-evaluated. Nothing can be done, however, to retain oil in the damless coupling, and some users
700
Auxiliary Components and Accessories
will not accept them for this reason. A proper decision can only be made by weighing a possible coupling failure because of sludge accumulation against an accidental failure of the lube system.
Gear Coupling Failure Modes The main causes of failure in gear couplings are wear or surface fatigue caused by lack of lubricant, incorrect lubrication, or excessive surface stresses. Component fracture caused by overload or fatigue is generally of secondary importance. High speeds require relatively lightweight gear elements. All case-hardening procedures produce distortion – to keep this distortion to a minimum, nitriding is the preferred hardening method. This method is employed after all machinery operations are complete and no further corrections are to be made to the tooth geometry. Nitriding permits increased tooth loading. The amount of increased capacity is not exactly known, but a 20% increase in load at 10,000–12,000 rpm has proven reliable. A further advantage of the nitrided coupling is that the coefficient of friction is lower than that for through-hardened parts. The heat from friction in the coupling decreases. More important, the transmission of axial forces is decreased by the reduced friction. In many cases, gear shaving prior to nitriding has been used to correct or minimize small errors of tooth geometry caused by the shaping or hobbing processes. A method of assuring nearly perfect tooth contact is to match-lap the gear teeth after nitriding. Lapping eliminates the break-in period, which otherwise takes from 70 to 120 hours. It is during the break-in period, that tooth surface distress usually occurs. For maximum reliability, it is recommended that nitrided gear teeth be specified. Experience indicates that the extra cost of match lapping is justified. The major failure in gear couplings is the fretting on the gear teeth. Fretting can be caused by improper lubrication. Lubrication problems can be categorized by the type of lubrication system being used. The two types of lubrication systems are the batch type and the continuous lubrication type. Table 18-2 shows some of the common problems that affect gear couplings, depending on the lube system used. Misalignment is another problem with gear couplings. Excessive misalignment can lead to any of the following problems, such as: tooth breakage, scoring, cold flow, wear, and pitting. Fasteners are another problem source in couplings. Table 18-2 Types of Typical Gear Coupling Failures Standard or Sealed Lube
Continuous Lube
Wear Fretting corrosion Worm tracking Cold flow Lube separation
Wear Corrosive wear Coupling contamination Scoring and welding Worm tracking
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Table 18-3 Diagnostic Analysis of Gear Couplings Damage or Stress Signs
Cause
Gear tooth surface deterioration (high rate of wear, scoring, and worm tracking) Gear tooth surface deterioration and overheating Tooth breakage and wear Broken hub, keys sheared Lockup-worn and broken teeth
Low oil viscosity and/or excessive misalignment
Worm tracking Broken end or seal ring Galled bores
Discolored bores
Fracture of components Cold flow, wear, and fretting Bolt shearing, bolt hole elongation Separation of lubricant ingredients Retention of moisture impurities Lubricant deterioration
Misalignment, high sliding velocity High misalignment angle Too much shrink fit on shaft Contaminated lubrication system, excessive misalignment Misalignment, separation of lubricant, low oil viscosity Too much shaft-to-shaft spacing and misalignment Improper removing techniques, insufficient or incorrect heating, excessive interference fit Improper hydraulic fit, contamination between shaft and hub Overload or fatigue, shock loading High vibration Nut bottoming out on threads Centrifugal force Centrifugal force High ambient temperature
Coupling fasteners should be properly heat-treated to withstand the large forces they experience in high-speed coupling applications. Fasteners should be properly torqued and, after four-to-six disassemblies, the entire fastener set should be replaced. Bolt shearing or bolt-hole elongation results from the nut bottoming out on the threads before the coupling flanges are tight, thus transmitting force through the bolt rather than through the flange faces. Bolts and nuts should be weight-balanced to very close tolerances. Table 18-3 is a diagnostic analysis of gear coupling failures.
Metal Diaphragm Couplings The metal diaphragm coupling is relatively new in turbomachinery applications. Although the first recorded use of such a coupling dates back to 1922 on a condensing steam turbine locomotive, the contoured diaphragm did not come into wide use until the late 1950s. Diaphragm couplings accommodate system misalignment through flexing. Fatigue resistance is the main performance criterion. The life expectancy of a diaphragm
702
Auxiliary Components and Accessories
coupling that operates within its design limits is theoretically infinite. Figure 18-6 is a photograph of a typical metal diaphragm coupling. Figure 18-7 shows a section through a diaphragm coupling. The coupling has only five parts: two rigid hubs, one spool piece, and two alignment rings. These five parts are solidly bolted together, and misalignment is accommodated through flexing of the two diaphragms of the spool. The spool piece is made up of three separate parts: two diaphragms and a spacer tube. These parts are welded together by an electron beam. The heart of these couplings is the flexing disc; it is manufactured from vacuumdegassed alloy steel, forged with a radial-grain orientation, and has a contoured profile machined on high-precision equipment. The contoured profile is shown in Figure 18-8. The diaphragm undergoes axial deflection. The forces acting on the disc that are generating the stresses are caused by the torque effects, centrifugal forces, and axial deflection. Standard methods for calculating centrifugal forces in a rotating disc show that both tangential and radial stresses increase rapidly with a decrease in the radius. The stresses imposed by axial deflection are much greater at the hub than at the rim, as seen in Figure 18-9. Therefore, to maintain uniform stresses in the diaphragm when all the various forces acting on the diaphragm are at their maximum, the diaphragm must be used to connect the contoured profile at both the hub and the rim to reduce stresses. Diaphragm couplings are more susceptible to axial movement problems than gear couplings, since the diaphragm has a maximum deflection that cannot be exceeded. Theoretically, a diaphragm coupling will have no problems or failures as long as it is operated within “design limits.” The diaphragm fails from excessive torque. Two distinct modes of failure can be found – one at a zero axial displacement and the other at a large axial displacement. Zero axial displacement is characterized by a circular crackline that goes through the thinnest portion of the diaphragm. The crack
Figure 18-6 Metal diaphragm coupling, one end shown (Courtesy of K´oppers Company, Inc.).
Couplings and Alignment
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Flex unit Removable assembly Flexible diaphragms Guards Tube
Flanged hub
High-strength body-bound bolts All metal locknuts
Figure 18-7 Schematic of atypical diaphragm coupling (Courtesy of K´oppers Company, Inc.).
Figure 18-8 Axial deflection in a disc.
is relatively smooth, and there is no buckling of the disc. The large axial movement and angular misalignment, which lead to disc failure, are characterized by a crackline that follows a random path from the thinnest to the thickest portion of the disc. The crackline is very irregular, and there is severe buckling of the unfailed part of the disc. Failure in this mode shows that the crackline propogates some 270◦ before disc buckling takes place, indicating that the torque load makes only a small contribution to the total stresses in the disc. Metal diaphragm couplings can also have problems due to corrosive action on the diaphragms. Thus, care must be taken to apply coating to protect against damage from a harsh environment.
704
Auxiliary Components and Accessories
a
b
Figure 18-9 Stress distribution under axial deflection.
Hub
Spacer
Hub
Flexing element
Figure 18-10 Typical metal-flexing disc coupling.
Metal Disc Couplings The main difference between the metal diaphragm coupling and the typical metalflexing disc coupling is that a number of discs replace the single diaphragm between the hubs and the spacer. Figure 18-10 shows a schematic of this type of coupling. A typical metal-flexing disc coupling consists of two hubs rigidly attached by interference fit or flange bolting to the driving and driven shaft of the connected equipment. Laminated disc sets are attached to each hub to compensate for the misalignment. A spacer spans the gap between the shafts and is attached to the flexing elements at each end. The functional requirements and characteristics of the flexing elements are to transmit rated torque as well as any system overloads without buckling or permanent deformation. In other words, they must possess torsional rigidity. However, under conditions of parallel, angular, and axial misalignment, the flexing element must have sufficient flexibility to accommodate these conditions without imposing excessive forces and moments on equipment shafts and bearings. Both of the previous requirements must be met while maintaining stress levels that are safely within the fatigue
Couplings and Alignment
705
f fx P
P
y
f
fy
x
Figure 18-11 Frictional damping in a metal-disc coupling.
limit of the flexing material. Metal-flexing couplings have been known to exhibit occasional large-amplitude vibrations in the axial direction when excited at the natural frequency of the coupling. The amount of damping present in a metal-flexing coupling is thought to be relatively small, although it is known to be greater for the laminated disc-type construction than for a coupling consisting of a single-piece membrane. The reason for the greater damping in the laminate disc configuration is that under conditions of axial movement, a microscopic amount of motion takes place between adjacent lamina, as shown in Figure 18-11. Since the element is clamped together under a bolt preload, there is a frictional force, which resists sliding. Field experience by manufacturers and users of turbomachinery has shown that resonant axial vibration of a metal-flexing coupling can cause problems at times that are reflected through the entire drive train. With laminated disc couplings, problems occurs only when an external forcing function exists. This condition could be a result of aerodynamic or hydraulic fluctuations in the machine train, out-of-square thrust collars, gearing inaccuracies, or electrical excitations of motor-driven equipment. It is usually possible to avoid operating the couplings at or near resonance if the condition is anticipated during the system design stage. However, such problems do not always occur until after a machine is in service. More information is needed on the nature and magnitude of external excitations.
Turbomachinery Uprates If an existing coupling is to be replaced with a new type of coupling because of a machinery uprate, or for any other reason, there is good justification to review, with the latest techniques, the nature of the rotating system to be coupled. Couplings, whether gear or disc-type, should not be simply picked from a catalog. Some installations are very old, and some have been revised in other ways in the field. Unfortunately, such engineering reviews are not easy to arrange with busy equipment suppliers. Therefore, the tendency is to match the obvious characteristics of the existing coupling and see what happens. Many older designs have relatively heavy and larger-diameter shafts, and retrofits have been very successful and trouble-free. Part of
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this success is due to the consideration given to the retrofit by cooperating engineers of the coupling manufacturer and the rotating equipment manufacturer. A large part of the success is due to the dedication and extra effort of the first companies offering the disc coupling to ensure success. If retrofits and new installations consume the available time of these engineers, the potential for omission increases. Therefore, more time should be allowed for the work. Coupling application is an engineering effort involving the coupling and rotating equipment designers. The user, by the purchasing technique he employs, can aid or hinder this effort, since he chooses the basic coupling style his operations and maintenance people will work with. In either case a good purchase specification should designate that the selection and design of the coupling must follow the rotor design work and exclude the coupling from becoming involved in competitive bids. It is simply too important an item to risk reliability for initial cost savings. Disc couplings are used as replacements for gear couplings for two reasons: (1) the disc couplings do not require lubrication, and (2) the machinery ratings can be uprated with disc couplings. Compressor and driver shafts often prove to be overstressed in equipment uprate situations; however, a change from conventional gear-type couplings to the more recent diaphragm coupling design can lower the shaft stress enough to avoid shaft replacement during power uprates of compressors or compressor drivers. A close examination of how the equipment vendor arrived at his maximum allowable stress levels may frequently show that such shaft replacements can be avoided without undue risk if the coupling selection is optimized. This situation is based on the fact that gear-type couplings have the potential of inducing both torsional stresses and bending stresses in a shaft, whereas diaphragm couplings tend to induce primarily torsional stresses and insignificant bending stresses at best. To determine if a machine’s performance can be uprated without installing a large shaft, the forces acting on the shaft must be computed. The forces acting on a shaft can be put into three separate categories: (1) torsional, (2) axial, and (3) bending forces. Torsional forces are a function of the shaft rotational speed and horsepower transmitted. They can be calculated from: T=
63,000 (hp) rpm
(18-1)
and the torsional stress τ T can be computed with: τT =
16T π d3
(18-2)
It is a generally accepted assumption that the axial stress will not exceed 20% of the torsional stress. τ a can therefore be obtained by τ a = 0.20 τ T . These two stresses
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will be the same for either type of coupling; however, the bending stress will vary depending on which type of coupling is used. There are three relevant bending moments caused by a gear coupling when transmitting torque with angular or parallel misalignment: 1. Moment caused by contact-point shift. This moment acts in the angular misalignment plane and tends to straighten the coupling. It can be expressed: Mc =
X T × Dp /2 2
(18-3)
where T = shaft torque Dp = gear coupling pitch diameter X = tooth face length (Figure 18-12) 2. Moment caused by coupling friction. This moment acts in a plane at a right angle relative to the angular misalignment. It has the magnitude: (18-4)
Mf = Tµ
where µ is the friction coefficient. 3. Moment caused by turning torque through a misalignment angle α. It acts in the same direction as the friction moment Mf and can be expressed as: MT = Tsin α
(18-5)
The total moment is the vector sum of the individual moments: q Mtotal = Mc2 + (Mf + MT )2
(18-6)
The contoured diaphragm coupling causes two bending moments: 1. Moment caused by angular misalignment. This results in bending of the diaphragm: MB = kBα
(18-7)
“X” th Too face
Lever arm
Mc Contact point
Figure 18-12 Shift in contact point.
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Auxiliary Components and Accessories
In this expression kB equals the angular spring rate of the diaphragm (lb-in/degree) and α is the misalignment angle. This moment acts in the angular misalignment plane, as did Mc in the gear-coupling analysis. 2. Moment caused by turning the torque through a misalignment angle α. It can be expressed: MT = Tsin α
(18-8)
The total moment is now: p Mtotal = MB 2 + MT 2
(18-9)
Comparing the bending moments caused by gear couplings with those resulting from contoured diaphragm couplings shows the former to be significant and the latter virtually negligible. The cyclic bending stress imposed on a gear coupling-equipped shaft can be computed from: σa =
Mtotal × C I
(18-10)
where C = shaft radius I = shaft area moment of inertia In addition, there is a mean tensile stress acting on the shaft cross-sectional area. This effect means stress equates to: σm =
Tµ (Dp /2)(π C2 ) cos 2
(18-11)
where 2 is the pressure angle assumed for the gear teeth. The cycle bending stress seen by the diaphragm coupling-equipped shaft can be obtained by a rapid ratio calculation: σ a (diaphragm coupling) Mtotal (diaphragm coupling) = σ a (gear coupling) Mtotal (gear coupling)
(18-12)
The mean tensile stress acting on the cross-sectional area of the diaphragm coupling-equipped shaft depends on how far the diaphragm is displaced axially from its neutral rest position and the axial spring rate of the diaphragm. For combined bending and torsion, the factor of safety can be calculated by the following relationships: n = s
1 σa σm kf + σ e σ y.p.
2
τa τm + 3 kf0 + σ e σ y.p.
2
(18-13)
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where σ e = endurance limit in tension σ yp = minimum yield strength in tension The stress concentration factor kf results from the keyway and must be used in torsional stress calculations. Factor kf0 takes into account the shaft step; it must be used in the bending stress calculation.
Curvic Couplings In essence, the curvic coupling is a ring of precision ground face splines that are meshed after index. The splines, or radial teeth, are ground in such a manner that on one member the sides of the teeth surfaces that meshed are convex and on the other member the sides are concave. The result is that after these members are clamped together, perfect index location is achieved, and further, the turret is perfectly on the center. The other important advantage of the curvic coupling is that its accuracy in both axes actually improves with use, rather than degrades. Care should be taken in the necessity of guarding against the entry of chips or other debris into the seating area. Figure 18-13 shows one-half of a typical curvic coupling, and Figure 18-14 shows a typical wheel in a gas turbine with curvic coupling machined into the disk. Figure 18-15 shows a typical cross section of a gas turbine. Many gas turbine’s rotors are of bolted construction with a positive torque incorporating such features as radial pins and curvic couplings, respectively. The rotor is supported by two-element tilting pad bearings and an upper-half fixed bearing. The thrust bearing is a double-acting type that uses the leading-edge groove lubrication system, as shown in Figure 18-15.
Figure 18-13 A typical curvic coupling.
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Auxiliary Components and Accessories
Figure 18-14 A typical disc with a curvic coupling.
Figure 18-15 Cross section of a gas turbine.
Figure 18-16 shows a schematic of how the axial compressor disks in the gas turbine are put together by the use of bolts and curvic couplings. These couplings transmit loads over 200 MW and are commonly used in many gas turbines.
Shaft Alignment The successful alignment of a gas turbine to the unit it is driving is of great importance. A major portion of operating problems experienced in the field can often be attributed to faulty misalignment. Operating problems caused by misalignment include excessive vibration, coupling overheat, wear, and bearing failures. Typically, misalignment problems will show up at two times rpm frequencies with axial vibrations at one and two times rpm. With diaphragm-type flexible couplings,
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711
Spindle bolt
Disc
Torque pin
Disc contact surface
Spindle bolt
Curvic coupling
Figure 18-16 Assembly mechanism of turbine disks using a curvic coupling.
vibrations may be somewhat suppressed, and consequently, trains using these couplings should be monitored periodically to ensure they are in alignment. Perfect alignment – exact shaft colinearity under operating conditions – is difficult and uneconomical to attain. The degree of tolerable misalignment is a function of coupling length, size, and speed. Some companies are now specifying a minimum coupling spacer length of 18 inches, since longer coupling lengths can tolerate more misalignment. The amount of misalignment that can be tolerated by the machine also depends on the types of journal and thrust bearings used. Tilting-pad-type bearings greatly reduce the misalignment problem. Figure 18-17 shows misalignment in both the journal and thrust bearings. The effect of misalignment on a journal bearing causes the shaft to contact the end of the bearing. Thus, journal length is a criterion in the amount of misalignment a bearing can tolerate; a shorter length obviously can tolerate more misalignment. The effect on the thrust bearing is to load up one segment of the thrust bearing arc and unload the opposite segment. This effect is more pronounced with higher loads and less flexible bearings.
The Shaft Alignment Procedure In essence, there are three steps in any alignment procedure. These are: (1) the prealignment survey, (2) cold alignment, and (3) the hot alignment check.
The Prealignment Survey This survey is carried out well ahead of the cold alignment. In this survey; piping, grouting, foundation bolts, shim packs, etc., are studied and ascertained to be appropriately done and of good quality. Again, casing distortion, piping strain, misalignment
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Auxiliary Components and Accessories
Brg
θ Shaft Line of centers
N
Hm Journal bearing
Thrust brg axis
Shaft Thrust bearing
Figure 18-17 Misalignment in both journal and thrust bearings.
of machine supports relative to the sole plate, etc. are determined, and corrections are made to ensure that these problems will not cause problems with the alignment. Piping strain is by far the greatest problem causer, and so piping should be carefully reviewed to ensure that it is properly done according to the code. Piping strains as high as 0.22 inches (0.5588 cm) have been observed. A typical cause of piping strain occurs when two flanges do not meet and pipefitters force them together. Pipe hangers that are poorly placed or tensioned can also cause significant piping stress problems.
Cold Alignment There are two predominant techniques used for cold alignment. These are: (1) the faceOD method, and (2) the reverse-dial indicator method. Both these techniques utilize dial indicators. For high-speed turbomachinery, the reverse-dial indicator method is the superior method and should be used.
Couplings and Alignment
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Figure 18-18 shows a face-OD indicator setup. As the name indicates, an alignment bracket is attached to one coupling hub, and face-OD readings are taken on the adjacent hub. The face and OD dial indicator readings give an indication of the angularity and offset of the shafts, respectively. The problems with this method are numerous. First, there is the problem of shaft axial float, which makes consistent readings difficult to obtain. Second, inaccuracies in the geometry of the coupling hub have to be taken into account. Third, the face diameter on which the readings are taken is relatively small, and errors are magnified over the length of the machine. The reverse-dial indicator method is shown in Figure 18-19. This method measures just the OD of the coupling hubs or shaft and eliminates the problem of shaft axial float. By spanning the entire coupling, angular misalignment is greatly magnified. For both the face-OD and reverse-dial indicator methods, it is important that sag in the alignment bracket be determined. Figure 18-20 shows a method for the determination of sag. Once the sag is determined, it must be permanently stamped on the bar. The alignment bracket should be considered an important precision tool and must be stored and handled with care so that it may be reused when realignment is required. Once the dial indicator readings are taken, a graphic plot of the two-shaft centerlines can be made on graph paper. It is at this stage that anticipated thermal growths are used in determining the shimming required to obtain shaft colinearity when the units are in the hot condition. Unfortunately, the values supplied by the manufacturers may not be accurate, and pipe strain and other external forces come into play. It is for this reason that the hot alignment check is conducted. A simple graphic plotting exercise for the reverse-dial indicator method shows the basic principles involved. A steam turbine compressor train is shown in Figure 18-21. Assume this train is a new installation and the manufacturer’s estimated thermal growths are as indicated in Figure 18-21. Reverse-dial indicator readings are taken to determine the relative shaft positions. Once readings are taken, the estimated thermal growths are incorporated by shimming, in the hope that a good, hot alignment can be achieved. Bore reading (O.D.) (Gives parallel offset)
Turbine shaft
Compr. shaft
Face reading (Gives angular offset)
Figure 18-18 Face-OD indicator setup.
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Auxiliary Components and Accessories
Dial indicator Bracket Read here Rotate Turbine shaft
Comp. shaft
First reading
Rotate
Turbine shaft
Comp. shaft
Second reading
Figure 18-19 Reverse-dial indicator setup.
The hot alignment check is used to determine the actual thermal growth, and then the final shim changes are made if needed. This example addresses only vertical movements. Horizontal movements are obtained in a similar fashion. The graphic plot uses an amplified scale on the vertical Y axis of one inch equals five mils vertical growth, while the X axis has a scale of one inch equals 10 inches (25 cm) of train length. In this example, it is assumed that Machine A is to be fixed, and all moves are to be conducted on Machine B. As shown in Figure 18-21, a “hot running line” is first drawn. This line is where the shafts should be when the machines are operating. Now, using estimated thermal growth of Machines A and B, a “cold target B” line is drawn. This line is where shaft B should lie so that when hot it will be colinear with shaft A on the hot running line. The next step is to use the dial indicator readings to determine where the shafts actually lie relative to each other. The B-to-A readings show that shaft B lies below shaft A by three mils (half-dial indicator readings) and the A-to-B reading shows that shaft A is above shaft B by five mils. Once these two points are located, shaft B can be plotted. This line is the “actual shaft B” line. Once this procedure is done, the shim changes needed can be easily found and “desired” indicator readings can be given to the millwrights.
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Indicator bar
Lathe center
SAG free
SAG free mandrel
3 mils True
Mandrel
Rotated through 180°
3 mils SAG −6
Figure 18-20 Method for determining sag.
A similar procedure is followed for horizontal movements. If the hot alignment check indicates a significant deviation from expected thermal growths and an unacceptable amount of misalignment, further shim changes can be achieved by similar plotting.
Hot Alignment Check This technique attempts to determine actual alignment status when the machines are hot. When the machines are running, it is impossible to use dial indicator techniques on the shafts.
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Auxiliary Components and Accessories
5 mils
a
b 8″ 6″ 8″
24″
10″
24″ a nar y
io “Stat
0
0
B to A
A to B
−6
−10
8″
a cL”
6″
5.4
2.4
14.4
36″
14.4
Estimated thermal growths in mils
8″
36″ b
Hot running line
5 mils 3 mils
“Ac t
ual
7.5 mils (remove)
sha
ft
Lcb
”
“Cold target b”
Figure 18-21 Graphic plotting for reverse-dial indicator method.
The old concept of a “hot check” – in which the units were shut down and the coupling disassembled as quickly as possible to allow indicator readings to be taken – should not be used. Currently used, continuously lubricated couplings require significant time to disassemble during which considerable cooling occurs. Because of this factor, a number of hot alignment techniques have been developed. Optical and laser methods, proximity probe methods, and a purely mechanical means using dial indicators may be used for hot alignment checks. In all these methods, an attempt is made to use the cold position of the shaft as a benchmark and then to measure the shaft movement (or bearing housings) from the cold position to the hot position. The objective is to find the change in vertical and horizontal positions at each shaft end. Once this procedure is done along the train, the machines can be shut down and appropriate shim changes made to attain acceptable hot alignment. Basically, the optical method uses equipment such as alignment telescopes, jig transits, and sight levels. Instruments with built-in optical micrometers for measuring displacements from a referenced line of sight enable an accurate determination of target movements, which are mounted on the machine. Optical alignment reference points are located on the bearing housings of the units. A jig transit is then set up at some distance from the train, and readings are taken and recorded in the vertical plane for each reference point in the train. Then the transit is moved, and a similar set of readings are taken in the horizontal plane. This procedure should be done at the same time as the reverse-dial indicator readings are taken. Then, when the train is in its operating condition, another set of readings are taken. The two
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717
data sets and the cold alignment dial indicator readings enable the determination of vertical and horizontal growths of each point. The advantages of this system are that it is accurate and, once the reference marks are on the machine, there is no need to approach the machine. However, the equipment involved is expensive and delicate, and great care has to be taken during its use. Moreover, heat waves often cause some problems in taking readings. Alignment with laser techniques has also been used, but the equipment is expensive and can be applied only in certain situations such as for a bearing alignment check. It is used primarily by manufacturers of turbomachinery during fabrication and assembly of their units. Proximity probes have also been used to measure machine movements. Proximity probes are mounted in special water-cooled columns and aimed at “targets” mounted on bearing housings or on other parts of the unit. Changes in the gap distances are then displayed on electrical meters. The Dodd bar system utilizes proximity probes mounted on an air-cooled bar attached between the bearings of the two machines to be aligned. The Dodd bar system allows continuous monitoring of the relative positions of the two shafts. Another system uses proximity probes located within the coupling to continuously monitor the alignment. Digital readouts of misalignment angles, etc., are available from this system. A purely mechanical, hot alignment system utilizing dial indicators has also been developed. The system uses permanently mounted tooling balls made of stainless steel attached to the bearing housing and to the machine foundation. A spring-loaded device with a dial indicator is provided to determine accurately the distance between the two tooling balls. An inclinometer is also provided to give a measure of the angularity. Figure 18-22 shows a typical configuration. Cold readings are taken at the time when the reverse-dial indicator readings are taken, and hot readings are taken when the machine is on-line. These two sets of readings are enough to determine the vertical and horizontal movement of the shaft. The same procedure is followed at each end of the units in the train. Computations can be made either graphically or by a calculator with preprogrammed cards. Direct outputs are the degree of misalignment and the shim changes needed to correct the misalignment. Machine Bearing housing Benchmarks Gauge X
Benchmarks
Foundation
Figure 18-22 Hot alignment system with dial indicator.
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Auxiliary Components and Accessories
It must be realized that correct alignment is of great importance in attaining high unit availability. Alignment procedures must be carefully planned, tools must be checked carefully, and, in general, great care must be taken during the alignment. The time, effort, and money spent on good alignment is well worth it.
Bibliography Bendix Fluid Power Corp, “Contoured Diaphragm Couplings,” Technical Bulletin. Bloch, H.P., “Less Costly Turboequipment Uprates through Optimized Coupling Section,” Proceedings of the 4th Turbomachinery Symposium, Texas A&M University, 1975, pp. 149–152. Calistrat, M.M., “Gear Coupling Lubrication,” American Society of Lubrication Engineers, 1974. Calistrat, M.M., “Grease Separation under Centrifugal Forces, American Society of Mechanical Engineers, Pub. 75-PTG-3 1975. Calistrat, M.M., “Metal Diaphragm Coupling Performance,” Proceedings of the 5th Turbomachinery Symposium, Texas A&M University, October 1976, pp. 117–123. Calistrat, M.M., and Leaseburge, G.G., “Torsional Stiffness of Interference Fit Connections,” American Society of Mechanical Engineers, Pub. 72-PTG-37, 1972. Calistrat, M.M., and Webb, S.G., “Sludge Accumulation in Continuously Lubricated Couplings,” American Society of Mechanical Engineers, 1972. Campbell, A.J., “Optical Alignment of Turbomachinery,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M Univ., 1973, pp. 8–12. Dodd, R.N., “Total Alignment Can Reduce Maintenance and Increase Reliability,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 123–126. Essinger, J.N., “Benchmark Gauges for Hot Alignment of Turbomachinery,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 127–133. Finn, A.E., “Instrumented Couplings: The What, the Why, and the How of the Indikon Hot Alignment Measuring System,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 135–136. Jackson, C.J., “Alignment Using Water Stands and Eddy-Current Proximity Probes,” Proceedings of the 9th Turbomachinery Symposium, Texas A&M University, 1980, pp. 137–146. Jackson, C.J., “Cold and Hot Alignment Techniques of Turbomachinery,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M University, 1973, pp. 1–7. Kramer, K., “New Coupling Applications or Applications of New Coupling Designs,” Proceedings of the 2nd Turbomachinery Symposium, Texas A&M University, October 1973, pp. 103–115. Massey, C.R., and Campbell, A.J., “Reverse Alignment-Understanding Centerline Measurement,” Proceedings of the 21st Turbomachinery Symposium, Texas A&M University, 1992, p. 189. Peterson, R.E., Stress Concentration Factors, John Wiley & Son, 1953. Timoshenko, S., Strength of Materials: Advanced Theory & Problems, 3rd ed. Van Nostrand Reinhold Pub., 1956. Webb, S.G., and Calistrat, M.M., “Flexible Couplings,” 2nd Symposium on Compressor Train Reliability, Manufacturing Chemists Association, April 1972. Wilson, C.E., Jr., “Mechanisms – Design Oriented Kinematics,” American Technical Society, 1969.
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Wright, J., “A Practical Solution to Transient Torsional Vibration in Synchronous Motor Drive Systems,” American Society of Mechanical Engineers, Pub. 75-DE-15, 1975. Wright, J., “Which Flexible Coupling?” Power Transmission & Bearing Handbook, Industrial Publishing Co., 1971. Wright, J., “Which Shaft Coupling Is Best – Lubricated or Non-Lubricated?” Hydrocarbon Processing, April 1975, pp. 191–196.