Centrifugal Compressor Rotor Repair

Turbomachinery rotor repair is a complex business. It involves knowledge of design concepts, good machinery practices, and above all, patience and keeping track of details.

Repair work on the rotating elements of compressors and turbines has traditionally been the field of the original equipment manufacturer, not the equipment owner. However, a number of independently owned or original equipment manufacturer (OEM)-owned dedicated repair facilities are available to users worldwide. These should be considered for compressor rotor repair and reliability enhancements since their competence is often as good as, or even better than, the OEM's main factory.

FIG. 1 through 3 illustrate work in progress at one such non OEM location.

However, a reliability-focused user will seek involvement in the compressor repair process. Equipped with basic knowledge and comparison standards (checklists and procedures, etc.), the user is in a good position to ask relevant questions of those responsible for turbomachinery repair.

Compressor Rotor Repairs

There are two basic types of compressor rotors: the drum type and the built-up type. There are two variations of each style.

FIG. 1. Compressor repair and run-out verification at a major independent repair facility. (Source: Hickham Industries, La Porte, Texas.)

Built-Up Rotor

FIG. 2. Impeller construction in progress at a major independent repair facility. Clock wise from upper left: welding; brazing; machining; assembly. (Source: Hickham Industries, La Porte, Texas.)

FIG. 3. Turbomachine being reassembled at a major independent repair facility. (Source: Hickham Industries, La Porte, Texas.)

This style is used for virtually all centrifugal compressors and some axial designs.

1. Heavy-Shrink Style-The impellers are usually shrunk to the shaft with an interference fit of 3/4 to 1 1/2 mils/in. of shaft diameter. In other words, a six in. shaft may have an interference of four to nine mils total before the impellers are placed on the shaft. This requires heating of the wheels to 400°-600°F for assembly. A small axial gap of about four to six mils is necessary between elements, consisting of sleeves and impellers, to allow for expansion during a rub and the Flexing of the rotor during rotor mode shifts at critical speeds.

Sleeves can be used at the interstage seals to protect the shaft from heavy rubs. The sleeve material can be selected with excellent rubbing and heat dissipation characteristics. Sleeves can also be replaced readily.

2. "Stacked" Rotors-This design is similar to the shrunk rotor, except that light shrink or press fits are used on the impellers. A large nut on one or both ends hold the impellers and sleeves together axially.

This design permits very low tip speeds so it’s usually limited to blowers that operate below their first critical speed.

Drum Type Rotor

This type is used only for axial flow-type rotors.

Impeller Manufacture

There are several manufacturing techniques for impellers.

1. The oldest form of impeller, produced by riveting blades of simple curvature between a disc and a shroud of contour-machined, forged steel, is still in use on multistage compressors of moderate tip speeds-600-800 ft/sec. With rivets milled integrally out of blade stock, and with alloy steel discs of 120,000 psi tensile stress or higher, tip speeds up to 1,000 ft/sec have been obtained operationally.

2. Fabricated shrouded impellers, assembled with the aid of welding.

This permits greater aerodynamic refinement in blade curvature design and, with newly developed alloy steels and welding techniques, operational tip speeds up to 1,100 ft/sec.

3. Cast aluminum or cast steel impellers, both shrouded and open, are used where high production rates justify the pattern cost.

4. Aluminum alloy impellers produced by plaster mold or similar precision casting techniques are used on turbochargers up to 1,300 ft/sec tip speed.

5. For extreme tip speeds, 1,200-1,600 ft/sec, radially bladed, semi shrouded impellers are contour-milled out of aluminum, magnesium, or steel alloy forgings, sometimes combined with separately machined or cast inducers at the inlet end. Also, EDM (electric discharging machining) is used to fabricate impellers from a single forging.

A controlling design factor is the rotor velocity expressed in the form of tip speed or peripheral speed of the impeller. The head produced by a compressor of given geometry will be a square function of the tip speed alone, quite independent of the size of the machine. Tip speed rather than shaft speed in rpm controls the mechanical stresses of a given rotor configuration.

The tip speed of a conventional impeller is usually 800 to 900 ft/sec.

This means that an impeller will be able to develop approximately 9,500 ft of head (the resulting pressure depends on the gas being compressed).

Multistage compressors are needed if duties exceed this value. Heavy gases such as propane, propylene, or Freon (mol wt 60 or above) require a reduction in tip speeds due to lower sonic velocities of these gases when compared with air. For these gases, the relative Mach number at the inlet side of the impeller generally is limited to 0.8.

Overspeed Test of Impellers

Overspeed tests are carried out at 115 percent of maximum operating speed (132 percent of operating stress) on most milled or fabricated steel impellers. Cast impellers are sometimes spun at 120-140 percent of design speed (that is, 144-200 percent of design stress levels). Over speeding is done in a heavily shielded "spin pit" that can safely contain, if necessary, a bursting impeller. The pit is evacuated for each run to minimize the windage loss of the spinning impeller. Bursting speeds may range from 1,200 to 2,000 ft/sec (tip speed) on fabricated and cast impellers, and may be even higher on forged and milled wheels.

Impellers of a ductile material will deform long before reaching burst speed. If the operating tip speeds must be as high as 1,200-1,600 ft/sec, a 15 percent overspeed test will cause local yielding of the most highly stressed region near the hub bore. In a material of sufficient ductility- over about 8 percent-this local yield will redistribute the stresses and improve dimensional stability in later operation; the impeller will then have to be finish-machined after the spin test.

Impeller Materials

AISI 4140 is the most prevalent steel alloy used for impellers and shafts in petrochemical gas compression services. For hydrogen sulfide laden gases, corrosion engineers recommend that the heat treatment be held to 22 Rockwell "C," which results in a yield stress of 90,000 psi.

Impellers in relatively clean service such as in ammonia plants often exceed this value and 30-35 Rockwell "C" hardness is allowed here.

Working stresses of most impellers are about 50,000 psi.

Impeller Attachment

A half-sectional profile of an impeller disk is a right angle triangle shape having a peripheral angle of about 16°. The hub design is significant in that it constitutes the impeller suction eye configuration, transmits the driving torque from the shaft, provides the necessary disk and rigidity, and fixes the impeller balance position. Most U.S. manufacturers use keys staggered 90° to complement the balance of multi-stage machines. The impeller hubs are reamed to a shrink interference fit of 0.75mil/in. to 1.5mil/in. A sharp-cornered keyway with a snug fit can develop stress densities two to three times the nominal stress. A loose fit would compound this stress concentration. The use of two close fitted feather keys has merit.

The thickness of a feather key is one-fourth of the width (peripheral) dimension. All keyways should have well-rounded fillets.

The principal source of rotor internal friction is the interference shrink fit of rotor elements on the shaft. The friction effect on built-up rotors could be minimized by making the shrink fit as short and heavy as possible. This led to the universal and most important practice of under cutting the bore of all elements mounted on the shaft, as shown in FIG. 4.

The impellers are locked against axial movement by various methods: split rings with locking bands, threaded locknuts, or a combination of both designs. If threaded nuts or sleeves are used, a lock ring is needed. Special care is needed in setting the lock ring:

1. Most lock rings are AISI 410 stainless material.

2. Hardness should be dead soft, under Rockwell C-20 (255 Brinell).

3. The locking notches or grooves in the sleeve, impeller, or balancing drum should be deburred.

4. The locking ring should be installed, the sleeve tightened, and the location of the locking notches or grooves marked on the ring.

5. The ring should be removed and nicked or saw cut about one-third of the way through at the marks.

6. The ring and sleeve should be reinstalled with the cuts in the rings aligned with the locking notches or grooves.

7. A soft round-nosed tool should be used to bend the tabs (alternately) into the notches.

8. Don’t over-bend the tabs.

9. Locking rings should not be reused. Important! FIG. 4. Typical arrangement of impeller and spacer mounted on the shaft.

Compressor Impeller Design Problems

The high speed rotation of the impeller of a centrifugal compressor imparts the vital dynamic velocity to the flow within the gas path. The buffeting effects of the gas flow can cause fatigue failures in the conventional fabricated shrouded impeller due to vibration-induced alternating stresses. These may be of the following types:

1. Resonant vibration in a principal mode.

2. Forced-undamped vibration, associated with aerodynamic buffeting or high acoustic energy levels.

The vibratory mode most frequently encountered is of the plate type and involves either the shroud or disc. Fatigue failure generally originates at the impeller OD, adjacent to a vane. The fatigue crack propagates inward along the nodal line, and finally a section of the shroud or disk tears out.

To eliminate failures of the plate type, impellers operating at high gas density levels are frequently scalloped between vanes at the OD ( FIG. 5). The consequent reduction in disc friction may cause a small increase in stage efficiency.

Several rotors have been salvaged by scalloping the wheels after a partial failure has occurred. Scalloping describes a machining procedure which removes material from the impeller periphery between adjacent vanes. The maximum diameter from vane tip to vane tip 180° apart remains unchanged by scalloping. One eight-stage, 6,520 hp, 10,225 rpm compressor in low molecular weight service, had 54 scallops done to each wheel during an emergency shutdown. To accomplish this, the rotor was unstacked. Each wheel was set up in the milling machine and scalloped. Then each wheel was individually balanced on a mandrel.

The rotor was restacked and the machine returned to service in slightly over a week.

Impeller Balancing Procedure

The quality of any dynamic balancing operation depends upon the following:

1. The control of radial run-out.

2. The elimination of internal couples along the length of the rotor.

Individual balancing of impellers is vital. The following procedures are a must:

1. Prepare half-keys as required for the balancing of individual impellers on a mandrel. These keys must precisely fill the open keyways at the impeller bores.

2. An impeller precision balancing mandrel must be made; the actual geometry should match the minimum pedestal spacing and the roller configuration of the balancing machine.

FIG. 5. Compressor impeller "scalloping" via removal of material from both disk and cover.

Guidelines for designing the mandrel are:

1. The mandrel should preferably be made of low alloy steel, i.e., AISI 4140 or AISI 4340, which has been suitably stress-relieved.

2. The journal surfaces should preferably be hardened and ground, with a finish not poorer than 16 rms.

3. All diameters must be concentric within 0.0001 in. TIR.

4. The diameter of the section where the impeller is to be mounted, should be established on the basis of heating the impeller hub to a temperature of approximately 300°F for installation and removal.

5. Keyways are not incorporated in the mandrel.

6. The impeller balancing mandrel should be checked to assure that it’s in dynamic balance. Make corrections on the faces, if required.

7. The impeller balancing mandrel should be free of burrs and gouges.

8. Mount each individual impeller, together with its half-key, on the balancing mandrel. A light coating of molybdenum disulfide lubricant should be first applied at the fits. Such mounting requires careful and uniform heating of the impeller hub, using a rosebud-tip torch, to a temperature of approximately 300°F; a temperature-indicating Tempil® stick should be used to monitor the heating operation. Install the mandrel, with the mounted impeller, in the balancing machine.

Cool the impeller by directing a flow of shop air against the hub while slowly rotating the assembly by hand or with the balancing machine drive. When the impeller and mandrel have cooled to approximately room temperature, proceed to identify the required dynamic corrections with the balancing machine operating at the highest permissible speed.

9. Make the required dynamic corrections to the impeller by removing material over an extended area, with a relatively fine grade grinding disc, at or near the impeller tip on both the cover and disc surfaces.

Other Compressor Balancing

1. Inspect all impeller spacers to assure that they are of uniform radial thickness in any plane perpendicular to their axis; minor axial taper variations are of no concern. The spacers may be mounted on a machining mandrel for concentricity checking.

2. The balance piston is considered to be the equivalent of an impeller.

Marking of Impellers

Stamping of numbers on impellers should never be allowed. Use a high speed pencil grinder, also called rotating pencil, in low stress areas. Burn marks from the rotating pencil are not as harmful as stamping marks.

However, they do create stress raisers, which have occasionally caused fatigue failures. Therefore, even these marks should be used only where absolutely necessary, and in carefully selected locations.

Critical Areas

Sharp edges on impeller bore: These edges dig into the shaft, causing high stress concentrations and possible starting points for shaft cracks.

Polished radii should be provided.

Sharp corners in keyways: Keyway radii should be about 1/4 of the key way depth. Two keyways are preferable because single keyways will cause nonuniform heating of the shaft and also shaft warpage because there is no radial shrink stress at the keyway region, while full radial shrink stress acts on the shaft opposite to the keyway. This will cause a shaft bow which changes with shrink stress, i.e., with temperature and speed. The keyway in the impeller should be as shallow as possible, since this key is not meant to transmit torque, but to act only as a positioning and safety feature.

Two shallow keys instead of one deep key are expensive, because key ways must be positioned very accurately, and depth must be equal. But it eliminates a host of problems (cranking effect, contact-pressure effect, component balance problems, thermal bow), and when comparing cost of problems and balancing difficulties against the higher manufacturing cost of double keys, these will probably be more economical, even for initial cost, but certainly when field problems and production loss are included.

A makeshift solution can be applied in the field, by relieving the shaft surface opposite the single keyway to get equal contact pattern. The relief need not be deep (say 20mils or so). This will at least improve the contact pressure and thermal distortion.

Another possibility presently gaining favor is to eliminate the keys entirely, depending 100 percent on the shrink fit engagement. The parts are mounted with a very high shrink (2.5mils/in. or more). Special methods must be used if disassembly becomes necessary, or the parts may be severely damaged. The possibility of stress-corrosion comes to mind with these high shrink stresses (~70,000 psi) but actually the stress in a keyway corner is much higher, usually exceeding yield strength, even with a moderate shrink fit (1mil/in. or less).

Rotor Bows in Compressors and Steam Turbines

Rotors sometimes will operate very satisfactorily for years, then upon restarting after a shutdown, excessive vibration occurs at the first lateral critical speed. This vibration problem may originate from the following:

1. The unit is tripped at rated speed and the rotative speed abruptly drops.

2. The effective interference fit of the impellers or wheels shrunk on the shaft increases rapidly during this speed reduction. (The bore stress varies with the square of the rotative speed.)

3. At the first critical speed range, maximum vibration amplitudes result; the rotor is then in a simple bow configuration, with the maximum deflection at approximately mid-span.

4. The shrunk-on elements thus literally lock the rotor in this bow shape and the clamping action increases below the critical speed. The outermost fibers of the shaft are incapable of sliding axially at the clamp or interference fit areas.

5. At rest, the rotor exhibits a large residual shaft bow, and gross imbalance causes high vibration amplitudes upon restarting.

6. Usually, the shaft is bowed elastically; disassembly of the rotor generally results in restoration of the bare shaft to an acceptable run-out condition.

Multistage centrifugal compressors tend to be particularly susceptible to the foregoing if they incorporate the same interference fit at each end of the impeller bore. It has been said that rotative speed tends to reduce the interference fit to near zero with the first lateral critical speed at 50 percent of the maximum continuous speed. In some cases the impellers will actually move on the shaft. Susceptibility to this problem increases as the ratio between the maximum continuous speed and the lateral critical speed increases. A very successful method of avoiding this problem is:

1. Making the static interference fit at the impeller toe equal to approximately one-half that at the heel, while maintaining the same land axial length.

2. Making the land axial length of the spacer at the end adjacent to the impeller heel equal to approximately 15-20 percent of the length at the opposite end, while maintaining the same interference fit.

3. A 4-6mil axial gap between components (i.e., impellers, spacer sleeves, etc.) is provided.

Tightness of the impeller and spacer on the shaft at full operating speed is thus assured, while simultaneously providing for controlled axial sliding at the interference fits during deceleration. Some increase in rotor internal friction of course results, with a consequent minor effect on the non synchronous whirl threshold speed.

A transient bow condition similar to the above is sometimes produced in built-up rotors which are subject to rapid starting. This occurs if all shrunk-on elements are fitted tightly together axially during cold assembly. In effect, the shrunk-on elements heat up much faster than the bare shaft after startup; the resulting thermal growth of these elements, when combined with a lack of perpendicularity of the vertical mating faces, results in shaft bowing. An axial clearance is usually provided between an impeller or wheel and its adjacent spacer to avoid this. Four to six mils is usually adequate.

This discussion is paraphrased from some of the writings of Roy Greene, who has experience in centrifugal compressor design, manufacture, and operation for Clark, Cooper-Bessemer, and Ingersoll Rand, and who currently is a consultant.

Balancing

Since the details of rotor balancing have been fully covered in Section 6, we can confine our discussion to a brief recap of this important topic.

It’s simple to balance a rigid rotor which runs on rigid bearings, so long as neither change shape, or deflect at operating speed. For a high-speed rotor, this is no longer true, because both the rotor and the bearings will deflect as they are exposed to centrifugal forces, and these deflections will result in a very complex unbalance condition that was not there when the rotor was balanced at low speeds in a precision balancing machine. A low speed balancing machine, no matter how precise and/or sophisticated, cannot detect such conditions: They lie dormant in the machine until high speeds are reached, and only then will the beast show its true nature. There is absolutely no way to accurately predict the behavior of an existing high speed rotor, except to get it up to speed on its own bearings, in its own casing. Even then it’s impossible to determine the precise location of unbalance, and to make definitive and 100 percent effective corrections that will eliminate the vibration throughout the speed range. The best one can hope for is to get a correction, essentially by compromise, that permits smooth operation at one speed.

The only way to get a predictable rotor is to maintain perfect balance of each individual component during all operations of machining and assembly and never to disturb this balance by making indiscriminate corrections on the finished assembly. To do all this properly is exceedingly difficult, and the methods and accuracies required border on (and often exceed) the limits of available technology of manufacturing and measuring. These limits dictate how fast a rotor may run and how many impellers can safely be used at a given speed, both factors being of great economic importance (cost, efficiency).

Once a high-speed rotor is assembled and found out of balance at operating speed, the only way to reestablish predictable balance is to completely disassemble the rotor and to start from scratch.

Clean Up and Inspection of Rotor

Compressor rotors must be carefully inspected for any damage. To accomplish this, these guidelines should be followed:

1. The rotor should rest on the packing area that must be protected by soft packing, annealed copper, or lead to avoid any marring of polished surfaces. Don’t use Teflon® strips since Teflon® impregnation of the metal surfaces can alter the adhesion characteristics of the lubricant in contact with the journal. Lubrication problems could ensue.

2. The rotor should be given an initial inspection for the following:

a. Impeller hub, cover, and vane pitting or damage.

b. Are there any rubs or metal transfer on the hub or cover indicating a shifting of rotor position? All foreign metal should be ground off and the area inspected for heat checking as described in Item 8.

c. Journals

• Journal diameter-Roundness and taper are the two most critical dimensions associated with a bearing journal. These dimensions are established with a four-point check taken in the vertical and horizontal planes (at 90° to one another) at both the forward and aft edge of the journal. A micrometer is normally used for this purpose. The journal diameter must be subtracted from the liner bore diameter to determine the clearance. If the journal diameter is 0.002 in. or more outside of its drawing tolerance, it’s necessary to remachine the journal.

Another parameter that must be carefully watched is journal taper. Excessive taper produces an increase in the oil flow out one of the ends of the bearing, thereby starving the other end.

This can result in excessive babbitt temperatures. Journal tapers greater than about 0.0015 in. require remachining.

• Journal surface-Surfaces that have been scratched, pitted, or scraped to depths of 0.001 in. or less are acceptable for use.

Deeper imperfections in the range of 0.001 to 0.005 in. must be restored by strapping.

• Thrust collar-does it have good finish? Use same guidelines as for journals. Is the locking nut and key tight? If the collar is removed, is its fit proper? It should have 0.001 to 0.0005 in. interference minimum.

3. The journals, coupling fits, overspeed trip, and other highly polished areas should be tightly wrapped and sealed with protective cloth.

4. The rotor should be sandblasted using No. 5 grade, 80/120mesh, polishing compound, silica sand, or aluminum oxide.

5. When the rotor is clean, it should be again visually inspected.

6. Impellers and shaft sleeve rubs-rubs in excess of 5mils deep in labyrinth areas require reclaiming of that area.

7. Wheel location-have any wheels shifted out of position? Wheel location should be measured from a thrust collar locating shoulder.

There should be a 4-5mil gap between each component of the rotor; i.e., each impeller, each sleeve, etc.

8. On areas suspected of having heat checking or cracks, a dye penetrant check should be made using standard techniques or "Zyglo":

a. Preparation Cracks in forgings probably have breathed; that is, they have opened and closed during heat cycles, drawing in moist air that has condensed in the cracks, forming oxides and filling cracks with moisture. This prevents penetration by crack detection solutions. To overcome this condition, all areas to be tested should be heated by a gas torch to about 250°F and allowed to cool before application of the penetrant.

These tests require a smooth surface as any irregularities will trap penetrant and make it difficult to remove, thus giving a false indication or obscuring a real defect.

b. Application The penetrant is applied to the surface and allowed to seep into cracks for 15 to 20 minutes. The surface is then cleaned and a developer applied. The developer acts as a capillary agent (or blotter) and draws the dyed penetrant from surface defects so it’s visible, thus indicating the presence of a discontinuity of the surface. In "Zyglo" an ultraviolet light is used to view the surface.

9. A more precise method of checking for a forging defect would require magnetic particle check, "Magnaflux" or "Magnaglow." As these methods induce a magnetic field in the rotor, care must be taken to ensure that the rotor is degaussed and all residual magnetism removed.

10. The rotor should be indicated with shaft supported at the journals:

a. Shaft run out (packing areas) 0.002 in. TIR max.

b. Impeller wobble-0.010 in. TIR-measured near O.D.

c. Shroud band wobble-0.020 in. TIR.

d. Thrust collar-0.0005 in. TIR measured on vertical face.

e. Vibration probe surfaces 0.0005 in. TIR-no chrome plating, metalizing, etc., should be permitted in these areas.

f. Journal areas-0.0005 in. TIR, 20 micro in. rms or better.

g. Gaps between all adjacent shrink fit parts-should be 0.004 to 0.005 in.

11. If the shaft has a permanent bow in excess of the limit or if there is evidence of impeller distress, i.e., heavy rubs or wobble, the rotor must be disassembled. Similarly, if the journals or seal surfaces on the shaft are badly scored, disassembly in most cases is indicated as discussed below.

Disassembly of Rotor for Shaft Repair

If disassembly is required the following guidelines will be helpful.

1. The centrifugal rotor assembly is made with uniform shrink fit engagement (3/4 to 1 1/2 mil/in. of shaft diameter), and this requires an impeller heating process or, in extreme cases, a combination process of heating the impeller and cooling the shaft.

2. The shrinks are calculated to be released when the wheel is heated to 600°F maximum. To exceed this figure could result in metallurgical changes in the wheel. Tempil sticks should be used to ensure this is not exceeded. The entire diameter of an impeller must be uniformly heated using "Rosebud" tips-two or more at the same time.

3. Generally a turbine wheel must be heated so that it expands 0.006-0.008 in. more than the shaft so that it’s free to move on the shaft.

4. The important thing to remember when removing impellers is that the heat must be applied quickly to the rim section first. After the rim section has been heated, heat is applied to the hub section, starting at the outside. Never apply heat toward the bore with the remainder of the impeller cool.

5. To disassemble rotors, naturally the parts should be carefully marked as taken apart so that identical parts can be replaced in the proper position. A sketch of rotor component position should be made using the thrust collar as a reference point. Measure and record distance from the thrust collar or shoulder to first impeller hub edge. Make and record distance between all impellers.

6. When a multistage compressor is to be disassembled, each impeller should be stenciled. From thrust end, the first impeller should be stenciled T-1, second wheel T-2, and so on. If working from coupling end, stencil first wheel C-1, second wheel C-2, and so on.

7. The rotor should be suspended vertically above a sand box to soften the impact of the impeller as it falls from the shaft. It may be necessary to tap the heated impeller with a lead hammer in order to get it moving. The weight of the impeller should cause it to move when it’s hot enough.

Shaft Design

It’s not uncommon to design for short-term loads approaching 80 percent of the minimum yield strength at the coupling end of the shaft.

The shaft is not exposed to corrosive conditions of the compressed gas at this point. Inside of the casing, the shaft size is fixed by the critical speed rigidity requirements. The internal shaft stress is about 5,000-7,000 psi- very low compared to the impellers or at the coupling area. With drum type rotors there is no central portion of the shaft, there are only shaft stubs at each end of the rotor. The purpose of the shaft is to carry the impellers, to bridge the space between the bearings and to transmit the torque from the coupling to each impeller. Another function is to provide surfaces for the bearing journals, thrust collars, and seals.

The design of the shaft itself does not present a limiting factor in the turbomachinery design. The main problems are to maintain the shaft straight and in balance, to prevent whipping of overhangs, and to prevent failure which may be caused by lateral or torsional vibration, chafing of shrunk-on parts, or manufacturing inadequacies. The shaft must be accurately made, but the limits of technology are not approached as far as theory or manufacturing techniques are concerned. A thermally unstable shaft develops a bow as a function of temperature. To reduce this bow to acceptable limits requires forgings of a uniformity and quality that can only be obtained by the most careful manufacturing and metallurgical techniques.

Rotors made of annealed material are not adequate, because many materials, For example AISI 4140, have a high ductility transition temperature in the annealed condition. This has caused failures, especially of shaft ends. Therefore, it’s very important to make sure that the material has been properly heat-treated.

Most compressor shafts are made from AISI 4140 or 4340. AISI 4340 is preferred because the added nickel increases the ductility of the metal.

Most of the time the yield strength is over 90,000 psi and the hardness no greater than 22 Rockwell "C" in order to avoid sulfide stress cracking.

While selection of the material is fairly simple, quality control over the actual piece of stock is complicated. There are several points to consider.

1. Material Quality: Forgings of aircraft quality (= "Magnaflux quality") are required for all but the simplest machines. Bar-stock may not have sufficient thermal stability, and therefore must be inspected carefully. Note that shafts-as well as all other critical components-must be stress-relieved after rough machining, which usually leaves 1/16 in. of material for finishing.

2. Testing: Magnaglow of finished shaft. Ultrasonic test is desirable for large shafts. Heat indication test is required for critical equipment.

3. Shaft Ends: Should be designed to take a moderate amount of torsional vibration, not only the steady operating torque.

4. The shaft must be able to withstand the shrink stresses. Any medium strength steel will do this. After some service the impeller hubs coin distinct depressions into such shafts, squeezing the shaft, so to speak.

This squeezing process also causes shaft distortion and permanent elongation of the shaft, which can lead to vibration problems or internal rubbing. Since part of the initial shrink fit is lost, this may cause other types of problems, such as looseness of impellers, which then can lead to looseness-excited vibrations such as hysteresis whirl.

Rotor Assembly

1. Remove the balanced shaft from the balancing machine, and position it vertically in a holding fixture providing adequate lateral support; the stacking step on the shaft should be at the bottom.

2. Remove all of the half-keys.

3. Assembly of the impellers and spacers on the shaft requires heating, generally in accordance with the procedure previously outlined for mandrel balancing. The temperature that must be attained to permit assembly is determined by the micrometer measurement of the shaft and bore diameters, and calculation of the temperature differential needed.

4. Due to extreme temperatures, a micrometer cannot be used; there fore, a go-no go gauge, 0.006 in. to 0.008 in. larger than the shaft diameter at the impeller fit, should be available for checking the impeller bore before any assembly shrinking is attempted.

5. Shrink a ring (0.003 in. to 0.004 in. tight) on the shaft extending about 1/32 in. past the first impeller location. Machine the ring to the exact distance from the machined surface of the impeller to the thrust shoulder, and record it on a sketch. This gives a perfect location and helps make the impeller run true.

6. Heating the impeller for assembly is a critical step. The important thing to keep in mind is that the hub bore temperature must not get ahead of the rim temperature by more than 10°-15°F. The usual geometry of impellers is such that they will generally be heated so that the rim will expand slightly ahead of the hub section and tend to lift the hub section outward. With long and heavy hub sections, extreme care must be taken to not attempt too rapid a rate of heating because the bore of the hub can heat up ahead of the hub section and result in a permanent inward growth of the bore.

Heating of the wheel can be accomplished in three different ways:

a. Horizontal furnace: the preferred method of heating the wheel for assembly because the temperature can be carefully controlled.

b. Gas ring: The ring should be made with a diameter equal to the mass center of the impeller.

c. "Rosebuds": The use of two or more large diameter oxyacetylene torches can be used with good results. The impeller should be supported at three or more points. Play the torches over the impeller so that it’s heated evenly, remembering the 600°F limitation. Tempil sticks should be used to monitor the temperature.

7. The wheel fit of the shaft should be lightly coated with high temperature antiseize compound.

8. The heated wheel should be bore checked at about the center of the bore fits. As soon as a suitable go-no go gauge can be inserted freely into the impeller fit bore, the impeller should be quickly moved to the shaft. With the keys in place, the impeller bore should be quickly dropped on the shaft, using the ring added in step 5 as a locating guide.

9. Shim stock, of approximately 0.004-0.006 in. thickness, should be inserted at three equally-spaced radial locations adjacent to the impeller hubs to provide the axial clearance needed between adjacent impellers. This is necessary to avoid transient thermal bowing in service.

10. Artificial cooling of the impeller during assembly must be used in order to accurately locate the impeller at a given fixed axial position. Compressed air cooling must be immediately applied after the wheel is in place. The side of the impeller where air cooling is applied is nearest to the fixed locating ring and/or support point. The locating ring should be removed after the impeller is cooled.

11. Recheck axial position of the impeller. If an impeller goes on out of position and must be moved, thoroughly cool the entire impeller and shaft before starting the second attempt. This may take three to four hours.

12. After the impellers, with their spacers and full-keys, have been assembled and cooled, the shim stock adjacent to the impeller hubs should be removed.

13. If the rotor has no sleeves, another split ring is needed to locate the second impeller. This split ring is machined to equal the distance between the first and second wheels. Then, a split ring is required for the next impeller, etc. Any burrs raised by previously assembled impellers should be carefully removed and the surfaces smoothed out.

14. Check for shaft warpage and impeller runout as each impeller is mounted. It may be necessary to unstack the rotor to correct any deficiencies.

15. The mounting of sleeves and thrust collars requires special attention. Sleeves have a lighter shrink than wheels and because of their lighter cross section can be easily damaged by uneven heating or high temperature. Thrust collars can be easily warped by heat. The temperature of the thrust collar and sleeves should be limited to about 300°-400°F.

16. Mount the rotor, now containing all the impellers, in the balancing machine, and spin it at the highest possible speed for approximately five minutes.

17. Shut down and check the angular position of the high spots and runout at the three previously selected spacer locations between journals. The high spots must be within ±45°, and the radial run outs within 1/2 mil, of the values recorded during bare shaft checking. If these criteria are not satisfied, it indicates that one or more elements have been cocked during mounting, thus causing the shaft to be locked-up in a bow by the interference fits. It’s then necessary to remove the two impellers and spacers from the shaft, and to repeat the vertical assembly process.

18. Install the rotor locknut, being careful not to over tighten it; shaft bowing can otherwise result. If the rotor elements are instead positioned by a split ring and sleeve configuration, an adjacent spacer must be machined to a precise length determined by pin micrometer measurement after all impellers have been mounted.

19. Many compressors are designed to operate between the 1st and 2nd lateral critical speeds. Most experts agree that routine check balance of complete rotors with correction on the first and last wheels is wrong for rotors with more than two wheels. The best method is to balance the assembled rotor in three planes.

The residual dynamic couple imbalance should be corrected at the ends of the rotor, and the remaining residual static (force) imbalance should be corrected at about the middle of the rotor.

For compressors that operate below the first critical (stiff shaft machines), two plane balance is satisfactory.

20. Install the thrust disc on the rotor; this should require a small amount of heating. It’s most important that cold clearances not exist at the thrust disc bore, since it will permit radial throw out of a relatively large mass at operating speed. Install the thrust bearing spacer, and lightly tighten the thrust-bearing locknut.

21. Spin the rotor at the highest possible balancing speed, and identify the correction(s) required at the thrust-bearing location. Generally, a static correction is all that is necessary, and it should be made in the relief groove at the OD of the thrust disc. No correction is permitted at the opposite end of the rotor.

22. Check the radial runout of the shaft end where the coupling hub will mount. This runout must not exceed 0.0005 in. (TIR), as before.

Shaft Balancing

Despite its symmetrical nature the shaft must be balanced. Again, the reader may wish to refer to Section 6 for details on the following.

1. Prepare half-keys for the keyways of the bare shaft. These should be carefully taped in position, using high-strength fiber-impregnated tape; several turns are usually required.

Note: Tape sometimes fails during spinning in the balancing machine. It’s therefore important that adequate shields be erected on each side of the balancing machine for the protection of personnel against the hazard of flying half-keys.

2. Mount the bare shaft, with half-keys in place, in the balancing machine with the supports at the journal locations. Spin the bare shaft at a speed of 300-400 rpm for approximately ten minutes. Shut down, and check the radial runout (TIR) at mid-span using a 1/10 mil dial indicator; record the angular position of the high spot and run out valve. Spin the bare shaft at a speed of 200-300 rpm for an additional five minutes. Shut down, and again check the radial run out (TIR) at mid-span; record the angular position of the high spot and runout valve. Compare the results obtained after the ten minute and five minute runs; if they are the same, the bare shaft is ready for further checking and balancing. If the results are not repetitive, additional spinning is required; this should be continued until two consecutive five minute runs produce identical results.

3. Check the radial run-out (TIR) of the bare shaft in at least three spacer locations, approximately equidistant along the bearing span, and near the shaft ends. Record the angular position of the high spots and the runout values at each location. The shaft is generally considered to be satisfactory if both of these conditions are satisfied:

a. The radial runout (TIR) at the section of the shaft between journals does not exceed 0.001 in.

b. The radial runout (TIR) outboard of the journals does not exceed 0.0005 in.

4. With the balancing machine operating at its pre-determined rpm, make the required dynamic corrections to the bare shaft using wax.

When satisfactory balance is reached, start removing material at the face of the step at each end of the center cylindrical section of the shaft. Under no circumstances should material be removed from the sections of the shaft outboard of the journal bearings.

Rotor Thrust in Centrifugal Compressors

Thrust bearing failure has potentially catastrophic consequences in compressors. Almost invariably, failure is due to overloading because of the following:

1. Improper calculation of thrust in the design of the compressor.

2. Failure to calculate thrust over the entire range of operating conditions.

3. A large increase in thrust resulting from "wiping" of impeller and balance piston labyrinth seals.

4. Surging of machine so that rotor "slams" from one side of thrust bearing to the other, and the oil film is destroyed.

5. Thrust collar mounting design is inadequate.

Rotor Thrust Calculations

Thrust loads in compressors due to aerodynamic forces are affected by impeller geometry, pressure rise through the compressor, and internal leakage due to labyrinth clearances. The impeller thrust is calculated, using correction factors to account for internal leakage and a balance piston size selected to compensate for the impeller thrust load. The common assumptions made in the calculations are as follows:

1. Radial pressure distribution along the outside of disc cover is essentially balanced.

2. Only the "eye" area is effective in producing thrust.

3. Pressure differential applied to "eye" area is equal to the difference between the static pressure at the impeller tip, corrected for the "pumping action" of the disc, and the total pressure at inlet.

These "common assumptions" are grossly erroneous and can be disastrous when applied to high pressure barrel-type compressors where a large part of the impeller-generated thrust is compensated by a balance piston.

The actual thrust is about 50 percent more than the calculations indicate.

The error is less when the thrust is compensated by opposed impellers, because the mistaken assumptions offset each other.

Magnitude of the thrust is considerably affected by leakage at impeller labyrinth seals. Increased leakage here produces increased thrust independent of balancing piston labyrinth seal clearance or leakage. A very good discussion of thrust action is found in Reference 3.

The thrust errors are further compounded in the design of the balancing piston, labyrinths, and line. API-617, "Centrifugal Compressors," specifies that a separate pressure tap connection shall be provided to indicate the pressure in the balance chamber. It also specifies that the balance line shall be sized to handle balance piston labyrinth gas leakage at twice initial clearance without exceeding the load ratings of the thrust bearing, and that thrust bearings for compressors should be selected at no more than 50 percent of the bearing manufacturer's rating.

Many compressor manufacturers design for a balancing piston leakage rate of about 1 1/2-2 percent of the total compressor flow. Amoco and others feel that the average compressor, regardless of vendor, has a leakage rate of 3-4 percent of the total flow, and the balance line must be sized accordingly. This design philosophy would dictate a larger balance line to take care of the increased flow than normally provided. The balancing chamber in some machines is extremely small and probably highly susceptible to eductor type action inside the chamber which can increase leakage and increase thrust action. The labyrinth's leakage should not be permitted to exceed a velocity of 10 ft per second across the drum. The short balancing piston design of many designs results in a very high leakage velocity rate. Since the thrust-bearing load is represented by the difference between the impeller-generated thrust and the compensating balance piston thrust, small changes can produce overloading, particularly in high-pressure compressors.

Design Solutions

Many of these problems have been handled at Amoco by retrofitting 34 centrifugal compressors (57 percent of the total) with improved bearing designs. Most of the emphasis has been toward increased thrust capacity via adoption of a Kingsbury-type design, but journal bearings are always upgraded as part of the package. Design features include spray-lubed thrust bearings (about a dozen cases), copper alloy shoes, ball and socket tilting pad journals, pioneered by the Centritech Company of Houston, Texas, and many other advanced state-of-the art concepts.

Some of the balancing piston leakage problems have been solved by use of honeycomb labyrinths. The use of honeycomb labyrinths offers better control of leakage rates (up to 60 percent reduction of a straight pass-type labyrinth). Honeycomb seals operate at approximately 1/2 the radial clearance of conventional labyrinth seals. The honeycomb structure is composed of stainless steel foil about 10mils thick. Hexagonal-shaped cells make a reinforced structure that provides a larger number of effective throttling points.

Compressor shaft failures frequently occur because of loose fit of the thrust collar assembly. With no rotor positioning device left, the rotor shifts downstream and wrecks the machine. The practice of assembling thrust collars with a loose fit (1 to 5mils) is very widespread because it makes compressor end seal replacements easier. The collar is thin (some times less than 1 in. thick) and tends to wobble. The shaft diameter is small in order to maximize thrust bearing area. A nut clamps the thrust collar against a shoulder. Both the shoulder and the nut are points of high stress concentration. With a thrust action of several tons during surging, the collar can come loose. In addition, fretting corrosion between the collar and the shaft can occur.

The minimum thrust capacity of a standard 8-in. (32.0 square in.) Kingsbury-type bearing with flooded lubrication at 10,000 rpm is well in excess of 6 1/2 tons. The thrust collar and its attachment method must be designed to accommodate this load. In most designs the inboard bearing has a solid base ring and the thrust collar must be installed after this thrust bearing is in place. The collar can be checked by revolving the assembled rotor in a lathe. The collar is subsequently removed for seal installation and must be checked for true running, i.e., the face is normal to the axis of the bearing housing again after it’s finally fitted to the shaft.

This problem has been addressed at Amoco by redesigning the thrust collar to incorporate the spacer sleeve as an integral part and have a light shrink fit (0 to 1mil tight). A puller is used to remove the collar after a small amount of heat is applied.

Managing Rotor Repairs at Outside Shops

When it becomes necessary to have rotor repairs performed away from your own plant, the outside shop should be required to submit such procedures as are proposed for inspections, disassembly, repair, reassembly, balancing, and even crating and shipping. And, while it’s beyond the scope of this text to provide all possible variations of these procedures, two or three good sample procedures are given for the reader's information and review.

In the following section, the procedure proposed by a highly experienced repair shop for work to be performed on centrifugal compressor rotors is shown.

Procedures for Inspection, Disassembly, Stacking, and Balance of Centrifugal Compressor Rotors

Incoming Inspection

FIG. 6. Recording rotor imperfections.

FIG. 7. Typical record of axial distances for centrifugal compressor rotor.

FIG. 8. Dimensional record for compressor rotor sealing areas.

1. Prepare incoming documentation. Note any defects or other damage on rotor. Note any components shipped with rotor, such as coupling hub or thrust collar.

2. Clean rotor. Protect all bearing, seal, probe, and coupling surfaces.

Blast clean with 200 mesh grit. Glass bead, walnut shell, solvent, and aluminum oxide available if requested by equipment owner.

After cleaning, coat all surfaces with a light oil.

3. Perform non-destructive test. Use applicable NDT procedure to determine existence and location of defects on any components.

Record magnitude and location of any defects as indicated in FIG. 6.

4. Measure and record all pertinent dimensions of the rotor as shown in FIG. 7 and 8. Record on a sketch designed for the particular rotor. Record the following dimensions:

• Impeller diameter and suction eye

• Seal sleeves, spacers, and shaft

• Journal diameters

• Coupling fits and keyways

• Gaps between adjacent shrunk-on parts

5. Check and record pertinent runouts. Rotor is supported at the bearing journals on "V" blocks. Runouts should be phase-related using the coupling (driven end) keyway as the 0° phase reference. If the coupling area is double-keyed or has no keyway, the thrust collar keyway should be used as the zero reference. If this is not possible, an arrow should be stamped on the end of the shaft to indicate plane of zero-phase reference.

6. Check and record electrical runout probe area. Use an 8-mm diameter eddy probe. Probe should be calibrated to shaft material only.

Probe area tolerance should be 0.25mil maximum.

7. Check and record all pertinent axial stack-up dimensions. Referenced from thrust collar shaft shoulder or integral thrust collar.

If Disassembly Is Required

1. Visually inspect. Visually inspect each part removed. Measure and record all pertinent shaft and component dimensions as follows:

• Impeller bore sizes-key size where applicable

• Shaft sleeve bore sizes

• Balance piston bore sizes

• Thrust collar bore size-key size where applicable

2. Use of applicable non-destructive test procedures. Use NDT procedures to determine existence and location of any cracks on shaft and component parts. Maximum allowable residual magnetism 2.0 gauss.

3. Completion of inspection procedures. Upon completion of inspection procedures, customer is notified and the results evaluated and discussed. The repair scope most advantageous to the customer is confirmed and completed.

Assembly:

1. Check dynamic balance. Check dynamic balance of shaft. Balance tolerance, unless otherwise specified, is 4w/n per plane. Correction of unbalance is analyzed and made on an individual basis.

• Rotors that stack from the center out stack two wheels at a time

• Rotors that stack from one end stack one wheel at a time

• After each stacking step, allow components to cool to 120°F or less

• Runouts should not change more than 0.5mil between component stacks

• Maximum allowable runout on shaft is 1.0mil

• Maximum allowable eye face runout is 2.0mil

• If runouts exceed tolerance, de-stack (to problem point), check shaft runout, and restack

• Perform 12-point residual per plane after final trim balance is completed

Final Inspection

1. Document. Document final runouts and submit to customer.

2. Probe area. As required, check and record vibration probe area for electrical/mechanical runouts; correct as required; maximum allow able 0.25mil peak to peak.

3. Preserve. Preserve rotor as follows: coat rotor completely with Cesco 140, wrap rotor, and notify customer with shipping or storage information.

A sample specification or procedure that the responsible (and responsive) repair shop furnishes to its sub-vendors is shown in the following section.

Turbo Specification Chrome Plating and Finish Grinding

Repair Facility to Provide to Vendor

1. Clear, concise drawing detailing:

• Areas requiring plating or grinding

• Finish dimension required and tolerance to hold-unless specified OD tolerance + 0.0005 - 0.0000. Finish 16 RMS maximum

• Shop contact and job number

• Desired delivery date

• Hardness of area to be chromed

2. Proper support cradle for safe transportation (when the repair facility is providing transportation).

3. A calibrated standard-taper ring gauge on taper coupling chroming.

4. No chroming on a sleeve area under any circumstances.

Vendor to Provide to Repair Facility

1. Proper support cradle for safe transportation (when vendor furnishes transportation).

2. Incoming inspection, to note any areas of concern not covered in original scope.

3. Chrome plating. Prepare areas to be chromed by grinding all grooves, pits, scratches, and other blemishes in area to be plated.

4. Blending in of all sharp corners and edges, both internal and external, with adjacent surface. Chrome deposit should thus be blended into adjacent surfaces so as to prevent lack of deposit or build-up of deposit.

5. Anodic cleaning of surfaces to be coated to assure maximum adhesion of chrome.

6. Chrome plating deposited directly on the ground surface without the application of any undercoat.

7. Chrome plating free of any visible defects. It should be smooth, fine grained, and adherent. A dye penetrant inspection qualifies the above.

8. No chrome plating on top of chromium, unless specified by the repair facility.

Finish Grinding

1. Chrome coating should be ground to finish dimensions specified.

Tolerance on OD should be + 0.0005 - 0.0000 unless otherwise specified.

2. Grinding should be done with proper coolant and wheel speed to produce proper surface finish.

3. Desired surface finish should be 16 rms maximum unless otherwise specified.

4. Taper shaft fits-appropriate, calibrated, and approved; ring gauge should accompany to ensure standard taper. A blue check should be made prior to shipment.

5. Final inspection-dimensional and dye penetrant.

6. Prepare for shipment by wrapping finished areas with protective cloth to resist damage during handling and shipping, and notify the repair facility representative upon completion for shipping arrangements.

FIG. 9. Methods of determining and limited hub advance on tapered shaft.

Mounting of Hydraulically Fitted Hubs

Modern turbomachinery rotors are commonly fitted with coupling hubs.

For years, these hubs have incorporated keys. Lately, however, keyless hubs have gained favor.

Hubs that are not provided with a keyway receive (or transmit) the torque from the shaft through friction. Hence, the hubs must grip the shaft tightly.

This gripping is accomplished by advancing the hub on the tapered shaft a specified amount. To facilitate this advance one must expand the hub bore. Two methods are used most often: heating or hydraulic pressure.

When hydraulic pressure is used, a few specialized tools are needed.

Basically, they are an installation tool, a high pressure oil pump with pressure gauge, and a low pressure pump with pressure gauge.

To ensure satisfactory performance, the following procedure is recommended for proper installation when using the hydraulic pressure method.

It assumes that your installation employs O-rings, although experience shows that many modern coupling hubs can be installed without the use of O-rings. On these, please disregard any references to O-rings.

Check for Proper Contact. After the shaft and hub bore are thoroughly cleaned, spread a thin layer of mechanics blue on the shaft and push the hub snugly. A very slight rotation of the hub is permitted after it’s pushed all the way. Remove the hub and check the bore for blue color. At least 80 percent of the bore should have contact.

Improve the Contact. If less than 80 percent contact is found, the shaft and hub should be independently lapped using a ring and plug tool set.

Clean the Lapped Surfaces. Remove all traces of lapping compound using a solvent and lint-free towels. Immediately afterward, spread thin oil on the shaft and hub bore to prevent rusting. Recheck the hub to shaft contact.

Determine Zero Clearance (START) Position. Without O-rings in the shaft or hub, push the hub snugly on the shaft. This is the "start" position. With a depth gauge, measure the amount the hub overhangs the shaft end and record this value.

Prepare for Measuring the Hub Draw (Advance). The hub must be advanced on the shaft exactly the amount specified. Too little advance could result in the hub spinning loose; too much advance could result in the hub split ting at or shortly after installation. As the overhang cannot be measured during installation, other means to measure the advance must be found.

The best way is to install a split collar on the shaft, away from the hub by the amount of the specified advance. Use feeler gauges for accurate spacing. See FIG. 9.

Install O-Rings and Back-Up Rings. The oil is pumped between the hub and shaft through a shallow circular groove machined either in the hub or in the shaft. Install the O-rings toward this groove, the back-up rings away from this groove. Don’t twist either the O-rings or the back-up rings while installing. After they are installed, look again! The O-rings must be between the back-up rings and the oil groove! Spread a little bit of thin oil on the rubber surfaces.

Mount "Other" Components. Read the coupling installation procedure again. Must other components (such as a sleeve) be mounted on the shaft before hub? Now is the time to do it.

Mount the Hub on the Shaft. Avoid pinching the O-rings during mounting.

The O-rings will prevent the hub from advancing to the "start" position.

This is acceptable.

Mount the Installation Tool. Wet the threads with thin oil, and rotate the tool until it butts against the shaft shoulder. The last few turns will require the use of a spanner wrench.

Connect the Hydraulic Lines. Connect the installation tool to the low pres sure oil pump (5,000 psi minimum). Connect the high pressure oil pump (40,000 psi minimum) to the hole provided either in the center of the shaft or on the outside diameter of the hub, depending on design. Loosen the pipe plug of the installation tool and pump all the air out; retighten the plug. Both pumps must be equipped with pressure gauges. See Figure 9 10.

Advance the Hub to the Start Position, through pumping the low pressure oil pump. Continue pumping until the hub advances 0.005 to 0.010 in. beyond the start position.

FIG. 10. Hydraulic system for hub installation.

Expand the Hub. Pump the high pressure pump until you read 15,000 to 17,000 psi on the gauge. As the pressure increases, the hub will tend to move off the shaft. Prevent this movement by occasionally increasing the pressure at the installation tool.

Check for Oil Leaks. The hub should not be advanced on the shaft if leaks exist! The pressure at the high pressure oil pump will drop rapidly at first because the air in the system escapes past the O-rings. Continue pumping until pressure stabilizes. A pressure loss of no more than 1,000 psi per minute is acceptable. If the pressure drops faster than that, remove the hub and replace the O-rings. However, before removing the hub make sure that the leaks don’t occur at the hydraulic connections.

Advance the Hub. Increase the pressure at the installation tool and the hub will advance on the shaft. If all the previous steps were observed, the pressure at the high pressure gauge will gradually increase as the hub advances. If the pressure does not increase, then stop. Remove the hub and check O-rings. If the pressure increases, keep advancing the hub until it touches the split collar or until the specified advance is reached. Don’t allow the pressure to exceed 30,000 psi. If it does, open the pressure valve slowly and release some oil. If in doing this the pressure drops below 25,000 psi, pump the high pressure pump to 25,000 psi, and continue the hub advance.

Seat the Hub. Very slowly release all the pressure at the high pres sure pump. Don’t work on that hub for 1/2 hour, or one hour in cold weather. After that, release all the pressure at the installation tool and remove it.

Verify the Advance. Measure and then record the new overhang of the hub over the shaft. Subtract from the overhang measured in the start position and the result must be the specified advance.

Secure the Hub. Remove the split collar from the shaft and install the retaining nut, but don’t overtighten. Secure the nut with the setscrews provided.

Dismounting of Hydraulically Fitted Hubs

In current practice, when a hydraulically fitted hub is removed, it comes off the shaft with sudden movement. Lead washers or other damping means are used to absorb the energy of the moving hub.

Koppers Company, Inc., engineers have developed a dismounting procedure that eliminates the sudden movement of the hub. Without this sudden movement the dangers related to removing hydraulically fitted hubs are greatly reduced. However, normal safety procedures should continue to be used.

Koppers' dismounting method requires the use of the same tools used when mounting the hub. The following procedure is recommended:

1. Remove the shaft nut.

2. Mount the installation tool. Wet the shaft threads with thin oil and rotate the tool until it butts against the shaft shoulder. There should be a gap between the tool and the hub equal to or larger than the amount of advance when the hub was installed (check the records).

If the gap is less than required, the wrong installation tool is being used.

3. Connect the hydraulic lines. Connect the installation tool to the low pressure oil pump (5,000 psi minimum). Connect the high pressure oil pump (40,000 psi minimum) to the hole provided either in the center of the shaft or on the outside diameter of the hub, depending on the design. Loosen the pipe plug of the installation tool and pump all air out; retighten the plug. Both pumps must be equipped with pressure gauges.

4. Activate the installation tool. Pump oil into the installation tool. The piston will advance until it contacts the hub. Continue pumping until the pressure is between 100 to 200 psi. Check for leaks.

5. Expand the hub. Pump oil between the hub and the shaft by using the high pressure pump. While pumping watch both pressure gauges.

When the high pressure gauge reads about 20,000 psi the pressure at the low pressure gauge should start increasing rapidly. This pressure increase is caused by the force that the hub exerts on the installation tool, and is an indication that the hub is free to move. Continue to pump until pressure reaches 25,000 psi. In case the low pressure at the installation tool does not increase even if the high pres sure reaches 30,000 psi, wait for about 1/2 hour while maintaining the pressure. It takes time for the oil to penetrate in the very narrow space between the hub and the shaft. Don’t exceed 30,000 psi.

6. Allow the hub to move. Very slowly open valve at the low pressure pump. The oil from the installation tool will flow into the pump and allow the hub to move. The pressure at the high pressure gauge will also drop. Don’t allow it to fall below 5,000 psi. If it does, close the valve and pump more oil at the high pressure pump. Continue the process until the valve at the low pressure pump is completely open and the pressure is zero.

7. Remove the hub. Release the high pressure and back off the installation tool until only two or three threads are still engaged. Pump the high pressure pump and the hub will slide off the shaft. When the hub contacts the installation tool, release all the pressure and remove the tool. The hub should now come off the shaft by hand. Don’t remove the installation tool unless the pressure is zero.

8. Inspect O-rings. Reusing even slightly damaged rings invites trouble.

The safest procedure is to always use new seals and discard the old ones.

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