In part 1 of our three-part checklist, we began our introduction to pump reliability practices 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 3 years to as much as 10 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 1,000 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 2 of this 3-part comprehensive checklist starts with lubrication issues (items 18 through 40) and concludes with mechanical seal issues (items 41 through 49).
LUBRICATION ISSUES
18. On sump-lubricated pumps, do not permit oil levels higher than through the center of rolling elements at the 6 o’clock position.
19. On pumps with dn-values (d = mean bearing diameter, mm; n = shaft rpm) in excess of 160,000, allowing lube oil to reach even this position may result in excessive heat generation. For these, the oil level may have to be lowered further and oil rings or flinger discs may have to be chosen by the user or pump manufacturer.
20. Except for moderate load applications in relatively cool ambient environments, ISO Grade 32 mineral oils and lighter lubricants are not suitable for rolling element bearings in centrifugal pumps. A thicker ISO Grade 68 lubricant will perform better in the majority of rolling element bearings in pumps. Unfortunately, it will not always work in pumps equipped with oil rings!
21. Bearing housing cooling may still be needed in pumps equipped with sleeve bearings. The majority of these pumps require ISO Grade 32 lubricants. Close viscosity control may have to be maintained for satisfactory long-term lubrication.
22. Observe required ISO viscosity grades in moderate climates (Europe, the Americas, Pacific Rim, Australia): rolling element bearings—ISO 68, synthesized hydrocarbon optional; sleeve bearings—ISO 32, synthesized hydrocarbon optional; combining both bearing types in the same housing—ISO 32 PAO or dibasic ester synthesized hydrocarbon mandatory for extended life.
23. In bearing housings with both rolling element bearings (preferred lube viscosity ISO 68) and sleeve bearings (which may require ISO Grade 32 lubricants), consider satisfying both needs by using ISO Grade 32 or 46 PAO or diester-based synthesized hydrocarbon oils.
24. Next to oil-jet lubrication, dry-sump oil mist applied in through-flow fashion per AP- 610 (8th and later editions) represents the most effective and technically viable lubrication and bearing protection method used by reliability-focused industry.
25. Be aware of upgrade and conversion options whereby a simple and economical inductive pump (a small pump with a free piston as its only moving part) can serve as the source of a continuous stream of pressurized lube oil. Used in conjunction with a spin-on filter, the resulting clean stream of lubricant can be directed at the bearing rolling elements for optimum effect.
26. Certain grease formulations cannot be mixed with other grease types. Incompatible greases often enter into a chemical reaction that renders them unserviceable in less than one year.
27. Over-greasing of electric motor bearings is responsible for more bearing failures than grease deprivation. Know where the spent grease ends up—hopefully not in the motor windings. Practicing proper re-greasing procedures is essential for long bearing life.
28. Lifetime-lubricated (sealed) bearings will last only as long as enough grease remains in serviceable condition within the sealed cavity. Whenever the product of bearing bore (mm) times shaft rotational speed (rpm) exceeds 80,000, reliability-focused plants consider it no longer economical to use lifetime-lubricated bearings in continuously operating industrial machinery.
29. Grease replenishing intervals depend on bearing inner ring bore dimension and shaft rotational speed. Reliability-focused user plants consider dn = 300,000 the maximum for grease lubrication of electric motors and other machines in continuous service. It has been reasoned that beyond this dn-value (d = bearing bore, mm; n = shaft rotational speed, rpm), grease replenishing intervals become excessively frequent and oil lubrication would be more economical.
30. Realize that oil ring lubrication very rarely represent state-of-art. Oil rings are alignment-sensitive and tend not to perform dependably if one or more of the following requirements are not observed:
- Unless the shaft system is absolutely horizontal, oil rings tend to “run downhill” and make contact with stationary components. Ring movement will be erratic and ring edges will undergo abrasive wear. The oil will become contaminated.
- The product of shaft diameter (inches) and shaft speed (rpm) should be kept below 8000. Thus, a 3-inch (75 millimeters) shaft operating at 3600 rpm (DN =~10,800) would not meet the low-risk criteria.
- Operation in lubricants that are either too viscous or not viscous enough will not give optimized ring performance and may jeopardize bearing life.
- The depth of immersion is closely controlled, the bore finish is 16 RMS or better, and the ring eccentricity does not exceed 0.002 inches (0.05 millimeters).
31. Solid metal flinger spools or flinger discs fastened to pump shafts often perform well for decades. When using retrofit flingers made with metal hubs/cores to which flexible elastomeric discs are firmly fused or attached, realize that the flex-parts have limited life.
32. If use of oil rings is unavoidable, be aware that a 30-degree angle between the contact point at the top of a shaft and points of entry into the oil represents proper depth of immersion. Too much immersion depth will cause rings to slow down, whereas insufficient depth tends to deprive bearings of lubricant.
33. Oil rings with circumferentially machined grooves will provide increased oil flow.
34. Consider buying true state-of-art bearing housing seals to preclude ingress of atmospheric contaminants and egress of lubricating oil. Install bulls-eye sight glasses on pump bearing housings so sealed.
35. On grease-lubricated couplings, verify that only approved coupling greases are used. Most motor bearing greases will centrifuge apart at the coupling’s peripheral speed!
36. Do not allow coupling greases to be used in electric motor bearings. Most motor bearings will fail prematurely unless a premium grade “EM” grease is used.
37. “All purpose” greases are not suitable for electric motor driver bearings in reliability-focused plants.
38. Fully consider vulnerabilities of unbalanced constant level lubricators. If you must use constant level lubricators, use pressure-balanced models (such as TRICO Optomatic Closed System II) only.
39. Mount constant level lubricators on the correct side of the bearing housing. Observe “up-arrow” provided by manufacturer of constant level lubricator. Incorrect mounting will lead to greater disturbances around the air/oil interface in the surge chamber of constant level lubricators. Correct mounting reduces the height difference between uppermost and lowermost oil levels. In other words, it ensures a more limited level variation.
40. Verify that re-lubrication and grease replenishment procedures take into account that:
- Mixing of incompatible greases will typically cause bearing failures within one year.
- Attempted re-lubrication without removing grease drain plugs will cause the grease cavity to be pressurized. Over-greasing will cause excessive temperatures.
- On shielded bearings, cavity pressurization tends to push the shield into contact with rolling elements or bearing cage, causing extreme heat and wear.
MECHANICAL SEAL ISSUES
It is generally acknowledged that most pump failure incidents involve mechanical seal distress. While this is true at many facilities, it is also true that a major refinery has documented an average mechanical seal life in excess of ten years (see reference 1). Using the right selection and installation procedures can markedly improve seal life and reduce pump failure incidents. Items 41 through 49 are of interest here.
41. Select mechanical seal types, configurations, materials, balance ratios, p-v values and flush plans certified to represent proven experience in identical services, or under verified-to-be-comparable service and operating conditions. Only these can guarantee to give extended seal life.
42. Except for gas seals (“dry seals”), mechanical seals must be operated so as to preclude liquid vaporization between faces. However, using cooling water in a jacketed seal chamber cannot effectively cool the seal environment. External flush cooling is far more effective.
43. Mechanical seals with quench steam provisions are prone to fail rapidly if quench steam flow rates or pressures are not kept sufficiently low. The installation of small diameter fixed orifices is very often needed to limit excessive steam quench rates.
44. Select the optimum seal housing geometry and dimensional envelope to improve seal life. Recognize that slurry pumps generally benefit from steeply tapered seal housing bores. The traditional concentrically bored stuffing box environment does not usually represent the optimum configuration for slurry pumps.
45. Avoid inefficient pumping rings on tandem seals. Consider using far more effective tapered pumping rings instead.
46. On hot service pumps, follow approved warm-up procedures. Verify that seal regions are seeing through-flow of warm-up fluid, i.e. are not dead-ended.
47. Understand the difference between conventional mechanical seals (seals where the flexing portion rotates) and stationary seals (where the flexing part is stationary). Maximum allowable speeds and permissible shaft run-outs are lower for conventional seals and higher for stationary seals.
48. Flush plans routing liquid from stuffing box back to suction are reliable only if the pressure difference exceeds 25 psi (172 kPa).
49. Fluid temperatures in seal cavity must be low enough to prevent fluid vaporization occurring in sea faces.
A LOOK AHEAD
Part 3 (next issue) will delve into hydraulic and installation issues. There, too, we will grapple with misunderstandings and oversights.
REFERENCES
1. Bloch, Heinz P., and Allan Budris. Pump User’s Handbook: Life Extension, Second Edition (2006), Fairmont Publishing Company: Lilburn, GA (ISBN 0-88173-517-5).
2. Bloch, Heinz P. Machinery Reliability Improvement, Third Edition (1998), Gulf Publishing Company: Houston, TX (ISBN 0-88415-661-3).
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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 heinzpbloch@gmail.com.
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MODERN PUMPING TODAY, September 2013
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