As we conclude this series, let us first review the major points raised in previous segments. We had introduced Parts 1 and 2 of our three-part checklist by making a few very important points:
- Centrifugal pumps in U.S. oil refineries and petrochemical plants typically reach mean-times-between failures (MTBFs) ranging from barely three years to as much as ten years. There’s therefore room for improvement at many plants.
- The average pump reliability improvement implementation costs only about 20 percent of the pump’s original cost.
- Pumps are probably responsible for one fire per 1000 pump repair events. Factoring the imputed value of fire avoidance into one’s upgrade cost justification makes sense.
- Our checklist of implementation items lists “things to consider” when pursuing reliability improvement. All are certainly known to best practices performers and further descriptions can be found in the listed reference books and other literature.
Part 3 of this 3-part comprehensive checklist starts with hydraulic issues, items 50 through 65, followed by mechanical improvement topics (66 through 76), installation issues (items 77 through 85), coupling topics, (items 86 through 91), and, finally, “more hydraulic issues” (items 92 through 100).
At the risk of repeating an earlier point: being familiar with all checklist items should be important to the reliability professional. Equally important should be the realization that a checklist is never intended to take the place of the rigorous explanations that can be found in books and articles. Interesting details and explanations are certainly available from both of the books listed in our references. Other publications should also be on the reading list, but you’ll have to start somewhere.
Since the first half of our pump reliability improvement checklist dealt with the mechanical side of process pumps, we continue here with important hydraulic issues, items 50 through 65.
50. Determine the suction energy and NPSH margin for the application. If dealing with a “high” or “very high” suction energy pump (per HI definition), then make sure that the pump is
- not operating in suction recirculation region.
- adequate NPSH margin has been provided.
- installed with no pipe stress, i.e. good piping practices have been followed.
51. Ascertain that pumps operating in parallel have closely matched operating points and share the load equally. Examine slopes of performance curves for each. Understand that differences in internal surface roughness may cause seemingly identical pumps to operate at different flow/pressure points.
52. Although centrifugal pump life can be greatly curtailed when operating in the low-flow range where impeller-internal flow recirculation is likely to exist, “last resort” help may be of value. A concentric “flow tube” inserted into spool piece adjacent to pump suction nozzle will reduce recirculation severity.
53. On pumps with power inputs over 230 kilowatts, verify that “gap A”, the radial distance between impeller disc tip and stationary parts is in the range of only 0.050 to 0.060 inches (1.2 to 1.5 millimeters). Pump-internal recirculation is thus kept to a minimum. On reduced diameter impellers, this would imply that trimming is done only on vane tips!
54. On pumps with power inputs over 230 kilowatts, ascertain that “gap B,” the radial distance between impeller vane tips and cutwater, is somewhere in the range of 6 percent of the impeller radius. This will reduce vibratory amplitudes occurring at blade passing frequency.
55. Ensure hydraulically-induced shaft deflections in single volute pumps are not excessive. This may require restricting the allowable flow range to an area close to BEP (best efficiency point).
56. Operate two-stage overhung pumps only at flows within 10 percent of BEP. Recognize severe shaft deflection and risk of shaft failure due to reverse bending fatigue when operating far away from BEP.
57. Check impeller specific speed versus efficiency at partial flow conditions. Consider installation of more suitable impellers for energy conservation.
58. Consider installing in the existing pump casing an impeller with different width, or with different impeller vane angle, or different number of vanes, or combinations of these. Observe resulting change trends in performance curves.
59. Consider changing slope of performance curve by inserting restriction bushing in the pump discharge nozzle.
60. Review if NPSH gain by cooling the pumpage is feasible and economically justified.
61. Consider extending the allowable flow range by using an impeller with higher NPSHr. Verify that NPSHa exceeds the NPSHr of the new impeller.
62. Use a ratio NPSHa/NPSHr of 3:1 or higher for carbamate and similar difficult services. There are many services where a 3-foot (~1 meter) difference between NPSHa and NPSHr is not sufficient to prevent caviatation.
63. Be aware of pre-rotation vortices and their NPSHr-raising effects on mixed flow pumps, i.e. pumps in certain specific speed ranges.
64. Consider use of a vertical column pump or placing pump below grade if NPSHa-gain is needed.
65. Consider inducer-type impellers where lower NPSH is needed, but be aware that to the right and left of BEP the new NPSH may now actually be higher than before.
MECHANICAL IMPROVEMENT OR UPGRADE OPTIONS
As we next remind the reader about mechanical improvement options, we note that these are often related to specifications, work procedures, and mechanical workforce training. Again, there may be some overlap because every job function in a process plant has potential impact on equipment reliability. Also note that a number of items relate to component upgrading. In reliability-focused facilities, every repair event is viewed as an opportunity to consider upgrade options.
66. Calculated shaft deflections should not exceed 0.002 inches (0.05 millimeters) over the entire operating range of the pump.
67. Implement suitable vortex breaker baffles on large vertical sump pumps.
68. Implement wear ring modifications (circumferential grooving) to reduce severity of rub in the event of contact due to excessive shaft deflection or run-out. Better yet, use Vespel® for all wear parts, but recheck new axial load!
69. Examine need for occasional measures to cure plate-mode or impeller cover vibration. Consider “scalloping,” if necessary to avoid impeller vibration other than unbalance-related vibratory action.
70. Use generous fillet radii (0.2 inch or 5 millimeter minimum) at shaft shoulders in contact with overhung impellers to avert reverse bending fatigue failures.
71. Verify shaft slenderness is not excessive. On API-610/5th Edition pumps, the stabilizing effect of packing may have been lost when converting to mechanical seals. Therefore, throat bushings may have to be replaced by minimum clearance (0.003 inch/inch or 0.003 millimeter/millimeter shaft diameter) shaft support bushings. These high performance plastic (Vespel® preferred) bushings should be wider than the customary open-clearance throat bushings originally installed.
72. Verify absence of shaft critical speeds on vertical pumps. Insist on conservative bearing spacing.
73. Verify acceptability of equipment spacing in pump pits and also ascertain conservatism of sump design. View Hydraulic Institute guidelines for spacing details.
74. Consider hollow-shaft motor drivers on vertical pumps and always use suitable reverse rotation prevention assemblies.
75. Beware of exceeding the rule-of-thumb maximum allowable impeller diameter for 3600 rpm overhung pumps: 15 inches (381 millimeters).
76. Consider in-between-bearing pump rotors whenever the product of power input and rotational speed (kilowatt times rpm) exceeds 675,000.
To survive or to reach long run lengths, centrifugal pumps must be properly installed. Installation checklist must be used and accountabilities defined. Items 77 through 85 address some (but not necessarily all) of the often overlooked issues.
77. Eliminate shaft misalignment by allowing foot-mounted pumps to reach equilibrium temperature. Only then secure or mount the bearing housing support bracket.
78. Do not allow shaft misalignment to exceed the limits plotted by competent alignment service providers. Use 0.5 mils per inch (0.0005 millimeter per millimeter) of shaft separation (DBSE, or distance between shaft ends) as the maximum allowable shaft centerline offset.
79. Use ultra-stiff, epoxy-filled formed steel base plates (“StayTru®” method)
- Proceed by first inverting and preparing.
- Use recommended grit blasting and primer paint techniques.
- Fill with suitable epoxy grout to become a monolithic block.
- After curing, turn over and machine all mounting pads flat and co-planar within 0.0005 inch per foot (0.04 millimeters per meter).
- Next, install complete baseplate on pump foundation, anchor it and level to within same accuracy.
- Finally, place epoxy grout between top of foundation and a reasonably wide perimeter beneath monolithic base plate.
80. On welded baseplates, make sure that welds are continuous and free of cracks.
81. On pump sets with larger than 75 kilowatt drivers, ascertain that baseplates are furnished with eight positioning screws per casing, i.e. two screws (“jacking bolts”) per mounting pad. Pad heights must be such that at least 1/8 inch (3 millimeter) stainless steel shims can be placed under driver feet.
82. Only now mount pump and driver. Use laser-optic or similarly accurate alignment device. Do not allow piping to be pulled into place by anything stronger than a pair of human hands. All the while, ascertain that no dial indicator moves more than 0.002 inches (0.05 millimeters) while tightening or loosening flange connections (see below).
83. While connecting piping to pump nozzle, observe dial indicators placed at nozzles in x-y-z directions and at bearing housing(s) in x-y directions. Dial indicator movements in excess of 0.002 inches (0.05 millimeters) are not acceptable.
84. Investigate magnitude and direction of thermal expansion of pipe and verify that pipe growth does not load up either pump suction or discharge nozzles. Sliding pipe supports must have two Teflon® plates. Steel-on-steel or single Teflon® plates are unacceptable.
85. Verify that vertical in-line pumps are free to respond to pipe movement. These pumps should not be bolted to the foundation.
Although not directly considered part of the pump, couplings must be selected with thought given to installation, safety, maintenance, and future performance parameters. Items 86 through 91 are of interest here.
86. Examine couplings for adequacy of puller holes or other means of hub removal. For parallel pump shafts with keyways, use 0.0-0.0005 inches (0.0 to 0.0012 millimeters) total interference. Use one of several available thermal dilation methods to mount hubs on shafts.
87. On grease-lubricated couplings, use only approved coupling grease. Realize that standard greases are being “centrifuged apart” at typical coupling peripheral speeds. Non-lubricated captured center member disk pack or diaphragm couplings are preferred.
88. Avoid elastomeric couplings for large pumps. Note that:
- Toroidal (“tire-type”) flexing elements will exert an axial pulling force on driving and driven bearings. Also, they are difficult to balance.
- Polyurethane flexing elements perform poorly in concentrated acid, benzol, toluene, steam, and certain other environments.
- Polyisoprene flexing elements fare poorly in gasoline, hydraulic fluids, sunlight (aging), silicate, and certain other environments.
89. Disallow loose-fitting keys for coupling hubs. Hand-fit keys to fit snugly in keyway. On all replacement shafts, machine radiused keyways and modify keys to match radius contour.
90. Use only limited end float couplings if drive motors are equipped with sleeve bearings.
91. On grease-lubricated couplings and after verifying that only approved coupling greases are used, review the regreasing procedure and interval. Understand what happens to the excess grease
92. Consider the effects on performance curve that could result from:
- vane underfiling
- vane overfiling
- “volute chipping”
93. Calculate axial thrust values and verify adequacy of thrust disc or balance piston geometry. Modify disc or piston diameter as required.
94. Consider opening of existing impeller balance holes if axial thrust must be reduced to extend bearing life. Realize that vendor-supplied balance holes are not always correctly dimensioned and may have to be opened.
95. Review and implement straight-run requirements for suction piping near pump inlet flanges entry. Aim for a straight pipe run of at least 5 pipe diameters between an elbow and the pump suction nozzle.
96. Realize that two elbows in suction piping at 90 degrees to each other tend to create swirling and pre-rotation. In this case, use ten pipe diameters of straight run piping between the pump suction nozzle and the next elbow.
97. On top suction pumps, maintain a ten-pipe diameter straight pipe length between suction block valve and pump suction nozzle.
98. Verify that eccentric reducers in suction lines are installed with flat side at top so as to avoid air or vapor pockets.
99. Use pumping vanes or suitably dimensioned impeller balance holes to reduce axial load acting on thrust bearings.
100. Oblique trimming of impeller vane exit tips, but retaining equal cover and disk diameters, will reduce the severity of vibratory amplitudes at blade passing frequencies and their harmonics. Oblique trimming of impeller vane exit tips, but retaining the full diameter only on the cover (suction side) will make the head-flow curve steeper, while still developing maximum head.
Again, this is “only” a checklist. But consulting it will avoid making errors and may lead the attentive reader to significant pump failure reductions.
1. Bloch, Heinz P., and Allan Budris. Pump User’s Handbook: Life Extension, Fourth Edition (2006), Fairmont Publishing Company: Lilburn, GA.
2. Bloch, Heinz P. Machinery Reliability Improvement, Third Edition (1998), Gulf Publishing: Houston, TX.
3. Bloch, Heinz P. Pump Wisdom (2011), Wiley and Sons: Hoboken, NJ.
ABOUT THE AUTHOR
Heinz P. Bloch, P.E., is one of the world’s most recognized experts in machine reliability and has served as a founding member of the board of the Texas A&M University’s International Pump Users’ Symposium. He is a Life Fellow of the ASME, in addition to having maintained his registration as a Professional Engineer in both New Jersey and Texas for several straight decades. As a consultant, Mr. Bloch is world-renowned and value-adding. He can be contacted at email@example.com.
MODERN PUMPING TODAY, October 2013
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