Fluid Handling

Examine and break the cycle of pump repairs

Pump failures don’t have to be recurring problems.

By Heinz P. Bloch, P.E., Process Machinery Consulting

pump turbo

 Pump Symposium

Heinz P. Bloch, P.E., owner of Process Machinery Consulting, will present “Breaking the Cycle of Pump Repairs” at the 43rd Turbomachinery/30th Pump Symposia in Houston on Sept. 23 at 2 PM and on Sept. 24 at 10:30 AM. The tutorial will describe how achieving the lowest possible life cycle cost (LCC) or lowest cost of ownership has been an undisputed goal of most pump users. From his 52 years of solid engineering practice, Bloch knows that pump hydraulics and fluid sealing details have received due attention over the years. However, bearing protection and lubricant delivery components are now often decades old, and many no longer reflect the best available technology. Additionally, "lean and mean" has often morphed into "cheap and risky."  He will explain why a few truly reliability-focused users enjoy pump MTBRs that are four times greater than others’ and why their per-pump maintenance expenditures are often only one-fourth of those forced upon their struggling competition.  This tutorial will delineate at least 10 little-known but highly important details of major interest to pump users seeking to improve pump life and minimize maintenance cost.. Learn more about the 43rd Turbomachinery/30th Pump Symposia at http://pumpturbo.tamu.edu.

Of the numerous process centrifugal pumps undergoing repair right this very minute, an estimated 90% have failed randomly before. Some have run just fine until the very first repair two or three years after startup and were never quite the same since after the first repair. Other pumps failed frequently or randomly, perhaps once per year, from the time they were originally commissioned. That brings up some questions. Could it be we don’t really know why many process pumps are failing? Could it be we just don’t give pumps the attention they deserve? Is it because everybody’s priorities are elsewhere? Or are there perhaps elusive failure reasons — factors overlooked by all parties?

Fortunately, improvement is both possible and cost-justified. Allowing repeat failures on process pumps rarely makes economic sense. Simple benefit-to-cost or lifecycle analyses will easily demonstrate that the pursuit of remedial action greatly benefits users.

The cost of failures

One way of exploring the value of extending pump mean time between failures (MTBF) is to examine the likely savings if we could improve the MTBF from presently 4.5 years to a projected 5.5 years. Say a facility has 1,000 pumps; that’s 1,000/4.5 = 222 repairs before and 1000/5.5 = 182 repairs after understanding and solving the problem. Avoiding 40 repairs at $6,000 each is actually a very low estimate, but would be worth $240,000. Avoiding repairs frees up manpower for other tasks: At 20 man-hours times 40 incidents times $100/hr, reassigning these professionals to other repair avoidance tasks would be worth at least $80,000.

There is also about one $3 million fire per 1,000 pump failures. An engineer at a U.S. Gulf petrochemical plant thought it might more likely be 1 fire per 1,000 pump failures; then, out of 10 fires, he figures seven are less than $50,000, two are between $50,000 and $500,000 and one is more than $500,000. We use the number $3 million/1,000 failures based on 52 years of experience and two recent updates. Data obtained in 2012 at an oil refinery in Indiana considered my number low, whereas a facility in Minnesota in 2009 thought it was spot-on.

It should be noted that data and contributory details of catastrophic incidents are often closely guarded secrets. Virtually all consulting done today by qualified independent professional engineers is linked to a legally binding nondisclosure agreement. The client is often compelled to file reports with local and federal regulatory agencies. These reports might differ from the findings of consulting engineers who understand the true root causes of failures or whose sense of priorities is tuned to higher standards. Diverging statements or findings might feed a bureaucratic machine that will busy itself with issues of this type.

Meanwhile, we consider these estimates for the value of fire damage restoration as reasonable. Therefore, avoiding 40 repairs would be worth 40/1,000 x $3 million = $120,000. Together, the three items — $240,000, $80,000, and $120,000 — add up to $440,000.

Although $ 6,000 was used for repair cost avoidance calculations earlier, an average API pump repair at a Texas refinery costs slightly over $10,200; a refinery in Mississippi reported $11,000. If the incremental cost of upgrading during the next repair adds $2,000 to the repair bill and avoids even a single failure every 3 or 4 years over the 30-year total life of a pump, the payback will have been quite substantial. It would be reasonable to assume eight avoided repairs at $6,000 give payback of $48,000/$2,000 = 24:1.

The estimated repair cost numbers of close to $11,000 reflect what needs to be considered in a pump repair cost calculation: Direct labor, direct materials, employee benefits at roughly 50% of direct labor, refinery administration and services costs at close to 10% of direct labor, mechanical-technical service personnel overhead costs amounting to around 115% of direct labor, and materials procurement costs from 7% to 8% of materials outlay. Disregarding the true cost of failures or repairs is likely to deprive some users of seeing the true benefit-to-cost ratio associated with pump upgrades.

We could examine other ways to calculate, as well. It would be reasonable to assume that implementing a component upgrade, generally the elimination of a weak link, extends pump uptime by 10%. Implementing five upgrade items yields 1.1^5 = 1.61 — a 61% mean-time-between-repair (MTBR) increase. Or, say, we gave up 10% each by not implementing six reasonable improvement items. In that instance, 0.9^6 = 0.53, meaning that the MTBR is only 53% of what it might otherwise be. That might explain industry’s widely diverging MTBR. The MTBR gap is quite conservatively assumed to range from 3.6 years to 9.0 years in U.S. oil refineries, and, as of 2014, no well-informed pump professional has disagreed with this range of MTBR numbers.

Internal recirculation issues

In recent decades, pump manufacturers and users recognized that certain high-energy pumps will fail prematurely if operated at low flow. The central aim has been to determine at what fraction of the flow at best efficiency point (BEP) the pump would still perform satisfactorily. The pump manufacturer then publishes these minimum values as net positive suction head required (NPSHr). The user’s design contractor endeavors to provide a net positive suction head available (NPSHa) in excess of the stated NPSHr. However, the actual NPSHr needed for zero damage to impellers and other pump components may be many times the number published in the manufacturer’s literature. The manufacturers’ NPSHr plots are commonly based on test stand observations noting a 3% drop in discharge head or pressure. While there are hydrocarbon services where an NPSHa surplus of just 1 ft over NPSHr will be sufficient to avoid cavitation, there are services such as carbamate, where a 25 ft surplus is not nearly enough. To be safe, many pump users try to avoid pump damage by somewhat arbitrarily disallowing pump operation at flows below 60% of the flow at BEP. More exact numbers can be calculated on a case-by-case basis. A close pump-user-to-pump-manufacturer relationship is helpful to shed more light on long-term experience. Also, the importance of upgrading mechanical or “power” end of process pumps is only recently receiving more attention. The need for attention to detail should be evident from unresolved, unexplained, and recurring pump failures. Deviations from best available technology are often involved.

fluid handling
Figure 1. This is an angular contact bearing with massive bronze cage.

Beware of deviations from best available technology

In 2013 a refinery was experiencing serious pump distress. The refinery had just encountered another massive thrust bearing failure on an important 3,560 rpm process pump. The pump’s two oil mist lubricated 75 mm/40° angular contact thrust bearings were arranged back-to-back, as is customary. Also, in compliance with the current American Petroleum Institute standard, API 610, each of these bearings had massive bronze cages. Figure 1 shows a section of a similar bearing, albeit some pumps are not reaching lowest possible lifecycle cost with brass or bronze cage bearings. Depending on axial load and bearing speed, carefully selected bearings with non-identical load angles, or bearings with carefully selected high-performance plastic cages may actually be more suitable than the universal bearings listed in API 610. While the stipulations of API 610 fit the majority of applications, they never claim to represent superiority in all instances. More detailed guidelines are available in “Pump User’s Handbook — Life Extension,” which I co-wrote with Allen Budris.

The refinery’s reliability manager passed along some relevant observations. From these it became evident that the refinery was struggling with the effects of several deviations from best practices in its oil mist lubricated equipment:

  • fluid handling2
    Figure 2: An oil return slot is needed at “A” when using liquid oil lubrication. But this slot becomes an undesirable bypass route if pure oil mist is applied between bearing and housing end cap.
    The process pumps incorporated the typical oil return notch located at A in Figure 2. This notch is needed in equipment with conventional lubrication. Without the notch, oil could get trapped behind a bearing and overheat. However, these bearings     were lubricated with pure oil mist, an excellent choice whereby the mist is applied between the bearing and the housing end cap in Figure 2. With an open oil return notch at the 6 o'clock position of the housing bore, some oil mist bypassed the rolling elements and was no longer available for the dual purposes of lubrication and cooling — the first deviation from best practices.
  • The unusually wide bronze cage acted as a restriction orifice for the remaining oil mist. An unusually wide cage became the second deviation from best practices.
  • fluid handling3
    Figure 3. Attempts to apply lubricant in the direction of the arrow (oil flow from left to right) meet with windage (air flow right to left) from an inclined cage. The two directions often oppose each other.
    All angular contact bearings create windage (Figure 3). An inclined cage in angular contact bearings acts as a small fan. A fan either promotes oil mist flow or opposes oil mist flow. With high peripheral cage velocities, this windage must be taken into account. At around 3,560 rpm and with a shaft diameter of 75 mm, the peripheral velocity exceeded the safe value of 2,000 fpm. At higher than 2,000 fpm, one should use directed reclassifiers so as to overcome windage. Directed reclassifiers must be mounted with the mist opening no more than 3/8th in. (10 mm) from a bearing cage. However, the refinery used standard oil mist reclassifiers — the third deviation from best practices.
  • At this refinery the standard reclassifiers were mounted approximately 3 ft away from the bearing housing, near the oil mist manifold. Because high windage bearings were involved, a large distance from reclassifiers to bearings became the fourth deviation from best practices. And, while placing reclassifiers at the oil mist manifold is normally allowed, doing so will reduce the available factor of safety compared to mounting reclassifiers at the bearing housing. The closer a reclassifier is to the bearing, the easier it will be for the oil mist to overcome the windage or fan effect of an inclined bearing cage.

Most angular contact thrust bearings will incorporate cages (ball separators) that are angularly inclined, which means they are arranged at a slant. These cages often act as small impellers, and impellers promote fluid flow from the smaller toward the larger of the two diameters. This is evident from Figure 3 which illustrates why particular attention should be given to windage created by the impeller-like air flow action of an inclined bearing cage. In many cases, the pump is designed with an oil ring (slinger ring) to the left of this bearing. While the design intent is for oil to flow from left to right, windage from an inclined cage will act in the opposite direction.

Windage is often overlooked, and we must ask: How does one alleviate windage and/or its effects? The fact that windage may be generated by some of these bearings and is more likely found in particular bearing housing configurations requires thoughtful and sometimes purely precautionary abatement of unequal pressures inside a bearing housing.

Lubricant sump oil level

In slow speed pumps, oil levels are generally set to reach the center of the bearing rolling element at its lowermost (6 o’clock) position. However, at higher speeds the resulting “plowing through the oil” may cause the lubricant to heat up significantly and plowing should be avoided on susceptible process pumps. There’s excessive heat generation risk whenever dn (the mean distance from diametrically opposite rolling element centers as expressed in mm x rpm) exceeds a particular number. That six-digit number ranges from 150,000 to perhaps 300,000. It is predetermined by bearing manufacturers who estimate at what point churning and heat buildup will exceed desired limits. The manufacturers then advocate lowering the oil level so that it no longer contacts the rolling elements. In essence, as a certain dn threshold is exceeded, some other means of lifting oil into the bearing must be chosen.

Irrespective of base stock and oil formulation, the required lubricant viscosity is a function of bearing diameter and shaft speed. Technical reasons are described in numerous books and articles, including:

  • “Improving Machinery Reliability”
  • “TRW Engineer’s Handbook,”
  • “Ball and Roller Bearings: Theory, Design, and Application”
  • “Practical Lubrication for Industrial Facilities,”
  • “Bearings in Centrifugal Pumps.”

Cooling Water

Note the effect of cooling water jackets surrounding rolling element-type pump bearings. Bearing housing cooling is not needed on process pumps which incorporate rolling element bearings. Cooling is harmful if it promotes moisture condensation (water cooling coils) or restricts thermal expansion of the bearing outer ring (water cooling jacket). In 1967, these concerns were seen to influence pump reliability. The jacketed cooling water passages should be left open to the ambient air environment. The decision to delete cooling water from pumps with rolling element bearings was first implemented in 1967 at an oil refinery in Sicily. The owner’s engineers had recorded bearing lube oil in four identical pumps reaching an average of 176 °F with cooling water in the jacketed passages. Without cooling water, the lube oil averaged 158 °F, which is 18 °F cooler. The bearings now lasted much longer. These findings and experiences were shared with all those who were willing to read or listen. Don’t waste water, it’s a precious resource. And realize that cooling water is very often responsible for actually reducing the life of rolling element bearings in process pumps. However, regardless of lube application method on rolling element bearings, cooling will not be needed as long as high-grade mineral or synthetic lubricants are utilized. High-performance mineral oils developed after 2010 are deliberately mentioned here.

Most process pump bearings will reach long operating lives if the oil viscosity, at a particular operating temperature, is maintained in a range from 13 to 20 cSt, according to SKF’s “Bearings in Centrifugal Pumps.” It should be noted that whenever oil rings are used to lift the oil from sump to bearings, the need to maintain a narrow range of viscosities takes on added importance, according to “Bearing Design and Application,” by Donald F. Wilcock and E. Richard Booser. In the special case of the same bearing housing containing both rolling element and sliding bearings, it will be prudent to address the implications of some oil rings not being able to function optimally in the higher viscosity (ISO grade 68) lubricant that’s often chosen for rolling element bearings. The oil ring may have been designed to cater to sleeve bearings, which normally need a lower viscosity lubricant, but VG 32 mineral oils are rarely a best choice for rolling element bearings in pumps. A high-performance synthetic VG 32 will often succeed as the most suitable selection for different bearing styles sharing the same housing.

To restate the above: Oil overheating must be avoided, especially so on many pumps operating at 3,000 or 3,600 rpm with oil reaching the center of the lowermost bearing ball or roller. Because the plowing effect of rolling elements in a flooded sump produces frictional power loss and heat, an oil level is often chosen and provisions are made to lift the oil. This lowers oil levels, and lifting is chosen whenever DN is greater than 6,000 (in this expression, D = shaft diameter in inches, and N = shaft rpm). Another, separately derived empirical rule allows shaft peripheral velocities no higher than 2,000 fpm in bearing housings where the liquid oil sump level is set to reach the center of the lowermost rolling element.

In oil mist lubrication systems, it is generally understood that, with shaft surface velocities in excess of 2,000 fpm, windage effects are opposing the flow of oil mist. As this is being observed, uninformed or baffled oil mist users have, in some cases, reverted back to conventional oil sump lubrication. In sharp contrast, reliability-focused users have, for many decades, installed directed oil mist reclassifiers to overcome windage at speeds of greater than 2,000 fpm. The mist-dispensing opening in directed reclassifiers should be located about 0.2-0.4 in. from the rolling elements. Thousands of these have been supplied and used with total success. This information is available from dozens of texts and articles, including “Performance of Oil Rings” by Rene A. Baudry and Leonid M. Tichvinsky and “Experimental Study of Stable High-Speed Oil Rings” by Hooshang Heshmat and O. Pinkus.

Again, once the shaft peripheral velocity exceeds 2,000 fpm, the oil level should be no higher than a horizontal line tangent to the lowermost bearing periphery. This means there should be no contacting of the oil level with any part of a rolling element and oil lifting is needed.

fluid handling4
Figure 4. The oil ring on the left is in as-new (“wide and chamfered”) condition, while the oils ring on the right is abraded —  badly worn and now without chamfer. Reliability-focused users record both before and after widths.

Assume a situation in which DN is greater than 6,000. Therefore, and because initial cost was to be minimized, either shaft-mounted flinger discs or oil rings (Figure 4) were chosen. Both arrangements are available to lift the oil or to somehow get the oil into the bearing by creating a random spray. Shaft-mounted flinger discs are well-represented in many European-made pumps. If properly designed, their operating shaft peripheral speed range exceeds that of oil rings.

Either way, the vendor’s test stand experience is of academic interest at best. Pump manufacturers test under near-ideal conditions of shaft horizontality, oil ring concentricity and immersion, oil level and lubricant viscosity. As users we might ask ourselves how often we have seen non-round oil rings or rings that have shaft radius wear marks, from shaft fillet radii, on one side of the ring. Perhaps we have never looked.

The cartridge approach has been in use for an estimated 50 or 60 years on thousands of open-impeller ANSI pumps because it facilitates impeller position adjustment in the axial direction. The same cartridge approach may be needed to dimensionally accommodate flinger discs instead of vulnerable oil rings (Figure 4). Of course, cartridge-mounted bearings are a cost-adder and a pump vendor may claim that the benefit-to-cost ratio will not justify upgrading to cartridges. However, with the average API pump repair costing slightly more than $10,200 at a Texas oil refinery and $11,000 at an oil refinery in Mississippi, we might be surprised at the payback multiplier. Even a single avoided failure over the 30-year total life of a pump will probably pay for it many times over.

What makes oil rings vulnerable

Issues with oil rings are found in many works. On a website post in September 2012, the Malaysian equivalent of OSHA alerted us to catastrophic failures brought on by oil rings. Many sources observed problems with oil rings, although an industry source opined in 2011 that ring lubrication is an accepted practice and it would take user consensus to damn it. Of course, history shows us that innovations are rarely driven by consensus. If they were, the Wright Brothers would have worked on repeat pump and bicycle repairs instead of developing a powered flying machine.

Meanwhile, many illustrations of failed oil rings are available. Studies, observations, and measurements have shown their field reliability in process pumps out of harmony with the quest for higher reliability and availability. Work described in “Bearing Design and Application,” by Donald F. Wilcock and E. Richard Booser, recommends oil ring concentricity within 0.002 in. However, in 2009, I performed shop measurements at a pump user’s site in Texas. The oil rings measured in 2009 exceeded the 0.002-in. allowable out-of-roundness tolerances by a factor of 30, likely because they had not been manufactured with stress-relieving as an intermediate step.

Experience shows that oil rings are rarely the most dependable or least-risk means of lubricant application. They tend to skip around and even severely abrade unless the shaft system is truly horizontal, unless ring immersion in the lubricant is just right, and unless ring eccentricity, surface finish, and oil viscosity are within tolerance. Taken together, these parameters are not usually found within close limits in actual operating plants.

Reliability-focused purchasers thus often specify and select pumps with flinger discs. Although sometimes used in slow speed equipment to merely prevent temperature stratification of the oil, larger diameter flinger discs serve as efficient, non-pressurized oil distributors at moderate speeds. Of course, the proper flinger disc diameter must be chosen and solid steel flinger discs should be preferred over plastic materials. Insufficient lubricant application results if the diameter is too small to dip into the lubricant; conversely, high operating temperatures can be caused if the disc diameter is much too large or if no thought was given to its overall geometry.

The incremental cost — comprising material, labor, CNC production machining processes — of an average-size (30 hp) process pump with cartridge-mounted bearings has been estimated at $300. The value of even a single avoided failure was earlier shown to be more than $10,000 and the benefit-to-cost ratio would thus exceed 33-to-1.

In 1999 at least one major pump manufacturer saw fit to examine the situation more closely (“Investigations Into the Contamination of Lubricating Oils in Rolling Element Pump Bearing Assemblies,” by Simon Bradshaw, from Proceedings of the 17th International Pump User’s Symposium, Texas A&M University. However, the problem did not go away. Users in Canada reported that black oil persisted, and so did repeat failures, even after adopting non-metallic oil rings. Black oil can easily be traced to one of two origins. A simple analysis either will point to overheated oil or will detect slivers of elastomeric dynamic O-ring material from components that operate too close to sharp-edged O-ring grooves.

Constant level lubricator issues

At least one manufacturer, Trico, agrees with user specialists who have observed that caulking where transparent bottles meet die-cast metal bases will, over time, develop stress cracks, or fissures. Rain water can then reach the oil via capillary action. Accordingly, bottle-type constant level lubricators are a preventive maintenance item and should be replaced after four or five years of service.

fluid handling6
Figure 6. Drive arrangements similar to the highly reliable mechanical governor drive in this small steam turbine are suggested for bearing housing-internal lube oil pumps in process pumps.

Note also how, in Figure 6, the oil level in the bearing housing is no longer reaching the rolling elements. This constant level lubricator lacks pressure balance. Any pressure increase in the space above the liquid oil will drive the oil level down. For a while, the top layer of oil will overheat; carbon will form and black oil will appear in the glass bulb. Increasing temperature in the closed space causes a further pressure increase, and the oil level decreases even more. Oil then no longer reaches the rolling elements and another bearing failure is likely.

The lubricator is configured for a balance line which ensures that the oil levels in the die-cast lubricator support, or at the edge of the slanted tube shown in this illustration, and in the pump bearing housing are always exposed to the same pressure, according to Trico. Undersized balance lines can exist; either a generous diameter hard pipe or a suitably sized stainless steel hydraulic balance line is favored. If constant level lubricators cannot be avoided, a pressure-equalized model or arrangement is recommended.

Again, bearing distress is inevitable if a constant level lubricator fails to maintain the desired oil level. An incorrect level setting can be caused by a number of factors. It will be clear from Figure 6 that even small increases in the bearing housing-internal pressure can heighten the failure risk. Suppose there is heat generation and because of the addition of bearing protector seals the air no longer escapes and there’s a lack of housing-internal pressure balance. Perhaps the reasons why Worthington had included housing-internal balance holes have been forgotten. The result may well be that the housing-internal pressure goes up. As the housing-internal pressure rises ever so slightly, it will exceed the ambient pressure to which the oil level at the wing nut or slanted tube in the bulb holder portion of the constant level lubricator is exposed. According to the most basic laws of physics, a pressure increase in the bearing housing causes the oil level near the bottom of the bearing inner ring shoulder (Figure 6) to be pushed down. Lubricant will no longer reach the bearing’s rolling elements, oil turns black, and the bearing will fail quickly and seemingly randomly.

To re-state, at DN > 6,000 and to satisfy minimum requirements in a reliability-focused plant environment, a stainless steel flinger disc fastened to the shaft will often perform well. Such a disc will be far less prone to cause unforeseen outages than many other presently favored methods. Remember that traditional oil rings will abrade and slow down if they contact a housing-internal surface. They are sensitive to oil viscosity and depth of immersion, concentricity, and RMS surface roughness.

Bearing housing protector seals

At the risk of stating the obvious, let’s be sure the lube in a pump’s bearing housing is kept clean. Even the most outstanding lubricant cannot save a bearing unless the oil is kept clean. This is where bearing housing protector seals are of value.

Lubricant contamination originates from a number of possible sources and can also be a factor in “unexplained” repeat failures. Unless process pumps are provided with suitable bearing housing seals, an interchange of internal and external air, called “breathing,” takes place during alternating periods of operation and shutdown. Bearing housings “breathe” in the sense that rising temperatures during operation cause air volume expansion, and decreasing temperatures at night or after shutdown cause air volume contraction. Open or inadequately sealed bearing housings promote this back-and-forth movement of moisture-laden and dust-containing ambient air. But simply adding bearing protector seals could change windage or housing-internal pressure patterns in unforeseen ways. This, too, must be recognized as a potential source of “unexplained” failures in housings without internal balance holes.

fluid handling5
Figure 5. Note that the sketch does not replicate an actual product. The illustration merely highlights what can happen with bearing protector seal designs that incorporate sharp-edged grooves. It reminds us that we     should become familiar with how parts work and how they might fail.

Ideally, bearing housings should not invite breathing and the resulting contamination. There should be little or no interchange between the housing interior air and the surrounding ambient air. The conventional breather vents on top of bearing housings can often be removed and plugged. We should not be surprised by that statement. After all, many hundreds of millions of refrigerators and automotive air conditioning systems operate with neither vents nor breathers. Conceivably, old-style bearing housing seals allow an O-ring to contact an O-ring groove (Figure 5). Contact with sharp-edged grooves invites dynamic O-rings to scrape. That’s another disclosure which should not shock us; none of us would think that sliding our fingers over a knife is without risk.

Abraded elastomer shavings can contaminate the lubricant and cause oil to change color. Also, using only a single O-ring for clamping the rotor to the shaft makes the rotor less stable than if two rings are used for clamping duty. Two clamping O-rings will provide more stability than one single clamping O-ring. By picking bearing protector seals with O-rings that cannot contact sharp edges we will eliminate a possible contamination source.

In essence, the right bearing housing protector seals can greatly improve both life and reliability of rotating equipment by safeguarding the cleanliness of the lubricating oil. However, these protector seals add little value if oil contamination originates with oil ring wear, or if pressure-unbalanced constant level lubricators are used that allow air and moisture to intrude, or if the oil is not kept at the proper level, or if the bearing housing design disregards windage concerns, or if water enters into the oil.

One of the most straightforward ways to drive a housing-internal oil pump could be modeled on the right-angle worm drives typically found in small steam turbines. While the arrangement shown in Figure 6 is associated with a mechanical governor, it is shown here as but one of many highly reliable options that merit consideration for small oil pumps that take suction from the process pump’s oil sump and pressurize it.

Among the possibilities worthy of examination is reconfiguring the portion of the equipment shaft which is located between the radial bearing and the thrust bearing. It might be possible to re-contour this shaft section to become the rotor of a progressive cavity pump. After routing the pressurized lube oil exiting from such a housing-internal pump through a downstream spin-on filter, the pressurized oil would be sprayed into nozzles that direct the oil into the process pump bearings.

All bearing manufacturers consider spraying liquid oil into the rolling elements the best possible lubrication method. Every bearing manufacturer I’ve ever worked with since 1958 has ranked an oil spray (liquid “oil jet”) ahead of oil mist lubrication and far ahead of oil rings (slinger rings). That’s a compelling fact which should not be ignored.


Heinz P. Bloch, P.E., is owner of Process Machinery Consulting (www.heinzbloch.com) in Westminster, Colorado, and the author of more than 600 articles and books, including “Pump User’s Handbook — Life Extension” and “Practical Lubrication for Industrial Facilities.” His work has been translated into five foreign languages, and Bloch has seven machinery-related patents to his credit. He is an ASME Life Fellow and was one of the original Texas A&M University Pump Advisory board members.

As of 2014, some process pumps continue to experience costly repeat failures. Motivated reliability professionals and informed users can avoid these and will appreciate recommendations on failure risk reduction. For the truly reliability-focused pump users, four points stand out:

  • Small deviations and “shortcuts” add up, and allowing several risk-inducing practices to exist will cause unacceptably low safety factors.
  • Serious reliability-focused users will always use checklists. In particular, oil refineries should make conscious efforts to do things right the first time.
  • There is a cost to having uninformed workers in any particular job function at an oil refinery. Staffers may have to overcome a reluctance to read relevant books and articles. Learning never stops and goals such as becoming above-average performers will elude those for whom it’s always “business as usual.”
  • A series of repeat failures will cost much more than the few hours of reading, or listening to a competent consultant who, in this case, could have easily explained how repeat failures could be avoided.

Knowledgeable engineers can show that some widely accepted pump components tend to malfunction in the real world. Moreover, as industry often moves away from solid training and from taking the time needed to do things right, designing out risk and designing out maintenance become attractive propositions.