ANSI Pump Impellers

In the ANSI pump world there are essentially 2 different choices for impellers; open (semi-open) and reverse vane. (I will hereafter refer to both the open and semi open designs simply as open.) True open impellers are not commonly used in industrial pumps because the demands of service require sufficient vane support (webbing or shrouds) to buttress the required torque loads and to maintain the relative geometric position of the vanes. Note: there is a closed impeller option for ANSI substituted in some rare cases.

In general, the reverse vane impeller clearance is established to the stuffing box and the open impeller style is matched to the casing. Each design has its pros and cons. Considering all the possible pump application variables neither style is the overall winner. The requirements of each application should dictate the design choice.

One Way or Another

• Less likely to clog when solids are present and have the advantage of passing soft, fibrous, and stringy solids.
• Lend themselves to inspection for damage/wear and to facilitate minor repairs or modifications.
• Less costly to manufacture when compared to either fully closed and/or reverse vane impellers.
• Rely on a close clearance with the casing for efficiency and performance.
• Utilize the adjacent casing area for the primary wear area.
• Clearance can ostensibly be set on the bench (without the casing) but it must be checked/adjusted again for accuracy when placed in the casing.
• When initially and properly set, can be just as efficient as a closed impeller.
• Typically utilize rear pump out vanes to reduce the axial hydraulic thrust, but in some cases will also feature hydraulic balance holes (ports).
• The stuffing box pressure will typically be higher than a pump with a reverse vane impeller.

• Used for aggressive fluids like acids and corrosives because of limited thread and crevice exposure.
• Uniquely designed with features of both closed and open impellers.
• Rely on the adjustable clearance with the back of the impeller and stuffing box face (mechanical seal chamber) for efficiency and performance.
• Utilize the rear cover (stuffing box/mechanical seal chamber) as the primary wear area.
• Facilitates accurately adjusting the impeller on the work bench. That is, the pump does not need to be installed in the casing to accurately set the clearance.
• Clearance can easily be set using the micrometer style rear cover design feature when compared to the open impeller techniques (Subjective).
• Utilize balance holes to reduce the axial hydraulic thrust.
• Achieve true repeatable performance. Each time you reset the impeller clearance the pump performance returns to like new which includes the stuffing box pressure and the magnitude of axial thrust.

Closing Remarks

Reverse Vane Impellers

There are claims that the reverse vane impeller requires less net positive suction head than open impellers for a given set of hydraulic conditions. We find this assertion unsubstantiated. Each individual application needs to be examined for its own merits.

Reverse vane designs typically prove more practical with double mechanical seal applications because the seal pot pressure will be lower.

Open Impellers

For open impellers, as each adjustment is made to re-establish clearance to the casing, the distance (gap) between the back of the impeller and the stuffing box will increase. A performance drop will start to occur when the gap begins to exceed 0.060 inch. The increased distance renders the pump out vanes less effective. Consequently both the stuffing box pressure and the axial thrust will increase.

A good practice is to limit re-establishing clearance on an open impeller design to three times before you disassemble the pump for wear inspection.

Dealing with Dynamic Forces

Pump out vanes are a design compromise consequently there is a small, but acceptable tradeoff with efficiency and power consumption. Some impellers will have both balance holes and pump out vanes, others just balance holes or just pump out vanes.

Do not use impellers with balance holes in lift situations as the potential for liquid product flashing in the seal chamber increases.

Summit Pump offers reverse vane, open and closed impellers for ANSI pumps.

Reduced shaft deflection results in greater shaft, bearing and mechanical seal longevity. This directly translates to increased reliability with a subsequent decrease in vibration and noise. All these factors will reduce pump total cost of ownership (TCO).

Background

When we first set course on the journey for reliable pump selection it is always prudent to thoroughly review how the pump will actually be operated and the liquid properties. I point this out because pumps are rarely operated at the desired/design coordinates. We all aspire for the pump to operate at the best efficiency point (BEP) or at least in the preferred (heart) of the allowable operating region (AOR), but real-world obstacles will frequently preclude us from reaching the goal.

Besides the pump operating point on the curve, shaft deflection is a function of many factors that include speed, shaft diameter, cantilevered length, and the geometry of the impeller. When coupled with additional negative factors like unbalanced impellers and bent shafts, deflection will be even worse.

Determining Deflection

In lieu of explaining the calculation of radial forces, determination of minimum flow and/or the Youngs Modulus of shaft materials; this month I simply want to remind our readers to consider a few general tips regarding shaft selection.

Most pump people are aware of the pump grading tool known as the “L over D” ratio (L³/D⁴). This tool is also known as the (shaft) slenderness ratio, deflection ratio or stiffness ratio. “L over D” is an index of shaft robustness and because it is a ratio there are no units. The ratio is a counterintuitive coefficient that describes the shaft’s ability to resist deflection due to the varying radial hydraulic forces during operation. The lower the number the better the shaft is able resist deflection. The lower ratio predicts a longer life for the mechanical seals, bearings, wear surfaces and overall increased pump reliability.

Note: “L over D” ratio is only relevant to single stage, overhung, end suction pumps. For example ANSI pumps (B73.1-2020) and other rotodynamic pumps with geometries designated as “OH” (refer to Hydraulic Institute).

Shaft Selection

Ideally you would select a solid shaft over a sleeved shaft. With today’s technical advancements in cartridge mechanical seals most applications can be handled successfully with solid shafts. The increase in shaft diameter will pay for itself in less downtime.

For a given pump/system issue you can’t always treat the root problem, but you can frequently address the symptoms and in some cases, you have viable choices. For example, mid-size ANSI pumps (Designations A-05 through A80) have the option to utilize an oversize shaft (LT versus an MT frame and shaft); The reduction in “L over D” ratio is significant.

Decision Factors:
It is more important to select a larger diameter shaft (lower ratio) when…

• The pump will be operated closer to minimum continuous stable flow (MCSF) than BEP
• The Suction Specific Speed (N_ss) is higher (as N_ss approaches 11,000 the allowable operating region reduces to a narrower envelope)
• The speed, flow, and head are high (a heftier shaft is exponentially more important on a 200 bhp pump than on a 5 bhp application)
• The impeller diameter is closer to maximum (radial force is a direct function of the impeller diameter)
• The impeller b₂ dimension is wider (radial force is a direct function of the width)
• The specific gravity (SG) is higher (radial force is a direct function of the SG)
• The viscosity is higher
• The duty cycle is continuous versus intermittent
• The impeller (rotor) balance is out of specification
• The casing is a single volute design in the larger sizes (The robust shaft requirement is reduced when the pump casing design is either a diffusor or dual volute)
• The casing is a standard nonconcentric (impeller center offset from casing center) design. If the pump is of a concentric design like a low flow-high head or a recessed impeller pump the need for a robust shaft is reduced.

Notes of Caution:

1. Do not overlook shaft material compatibility with the liquid
2. It is understood that pumps utilizing packing will require a sleeve, but deflection is less due to the “Lomakin effect” of the packing

When in doubt check with your RSM or engineering for further guidance. I have also written several technical columns on the subject, please see the links below:

Solid Shaft Designs and Cartridge Seals
Why Does My Pump Shaft Keep Breaking
Minimum Flow Redux
Casing vs. Impeller
What is the Root Cause of Shaft Failure

In the spirit of continuous improvement we like to occasionally remind our readers about important factors that are essential for a successful pump installation. Liquid properties such as viscosity has relatively little effect on positive displacement pump performance and in most cases makes them even more efficient, but viscosity is pure kryptonite for centrifugal (rotodynamic) pumps.

What is Viscosity?

Viscosity is that property of a fluid which tends to resist a shearing force. Simply stated; does it pour fast or slow …or how quick will a ball bearing fall through a vertical column of the liquid? You may also think of viscosity as the internal friction resulting when one layer of the fluid is made to move in relation to another layer. The viscosity of liquids will decrease as the temperature increases which is the opposite characteristic of gases.

Curve Corrections

Centrifugal pump published performance curves are based on water and so when moving viscous fluids the performance must be corrected. Technically the pump performance curve should be corrected for viscosities above 5 centipoise (cps) which most people do not do. Please be advised that at viscosities greater than 40 centipoise it is no longer an option and the curve must be corrected. For viscosity corrections we use centipoise units. See Hydraulic Institute/ANSI standard 9.6.7 for more details.

As viscosity increases:

• Pump efficiency will decrease significantly
• Flowrate will decrease
• Developed head will decrease
• Power (bhp) required will increase
• NPSHR will increase and the NPSHA will decrease
• System curve will become more restrictive
• Affinity laws will become inaccurate

Centipoise vs. Centistoke

A simple way to explain the difference between these two terms is that…

Centipoise: Dynamic viscosity is a measurement of the force required to overcome fluid resistance to flow through a tube (or capillary).

Centistoke: kinematic viscosities are timed flow rates through orifices where the driving force is typically gravity.

The relationship between Centipoise and Centistokes is proportional via the specific gravity.
Centistoke = Centipoise ÷ Specific Gravity

Pumps Have boundaries

All pumps have limits as to how much horsepower can be utilized. This boundary is based on power frame (shaft and bearing) speed versus torque limits and is normally expressed in BHP per 100 rpm terms.

Note: These stated limits are based on direct drive and must be reduced for belt or engine drive. See engineering for assistance.

Maximum Limit: Should you need to pump liquids at higher viscosities please contact us to review the application. Viscosities above 2000 – 2500 centipoise must be handled with positive displacement pumps.

Caution: Do not apply the Affinity Laws directly to viscosity performance corrections. You must first correct the viscous performance to its equivalent water performance before the Affinity Laws are applied. Once you have determined the new water performance then the viscous correction factor can be applied to estimate the new viscous performance.

Good news: All of these required viscosity calculations can be easily accomplished in a few keystrokes on Summit Select.

The first modern bearing isolator was patented in 1977. They were initially used in the food processing and chemical industries on pumps and mixers. Since the 1980’s they have been installed on most anything that rotates. Isolators protect the bearing housing from the introduction of liquids and debris (contamination ingress). By keeping the lube oil clean the bearing life is exponentially improved. Just as important the bearing isolators are also designed to keep the oil in the bearing housing of the pump.

We occasionally receive complaints from the field regarding bearing isolator leakage. 99% of these issues are easy to solve. We have complied a list of the most common remedies and some added tips.

Causes of Bearing Isolator Leakage:

• • 1.) Overfill of the bearing housing is the most common reason for oil leakage. The oil level should be at and no higher than the middle of the sight glass. This “sighted” level should correspond with the middle of the lowest ball in the ball bearing set, either radial or thrust.
    
    Field Tip: It is important to understand that once the housing is overfilled it will take some time for the excess oil to be dissipated from the saturated isolator. In the interim it is important to continue cleanup of the expelled oil for several days or even weeks of running post overfill incident.

• • 2.) Pump is not level. I recommend the use of a machinist level to check the pump and note the pump must be level in both directions (stem to stern and athwart-ships to describe it in nautical terms). Additionally, please make the distinction that “flat” is different from “level”.

• • 3.) Wrong isolator “Case A”. For example, the isolator for the inboard end of a reverse vane impeller pump has a different style isolator than other style pumps. Field Tip: Reverse vane pumps adjust the axial clearance of the impeller to the seal chamber (stuffing box) by rotating the threaded bearing carrier on the inboard end. The bearing carrier contains the isolator and so it could be in any radial position as a result of setting the impeller clearance. For this discussion the inboard is defined as being closest to the coupling and outboard being farthest from the coupling.

• • 4.) Wrong isolator “Case B”. For reasons other than #3 above, such as the wrong size or material. Please appreciate there are also different isolators for high/low temperature and or special applications such as a harsh chemical environment.

• • 5.) Installed improperly or damaged during installation. See Instruction and Operating Manual (IOM).

• • 6.) Expulsion port is in the wrong location. It must be at the 6 o’clock position.

• • 7.) Fit and tolerance. The housing bore dimension (ID) and / or the outside diameter (OD) of the stator is incorrect. Refer to the IOM and/or factory for proper dimensions and tolerances.

• • 8.) The isolator is installed properly but it is improperly moved at a later time. When adjusting the impeller clearance settings, or general maintenance and handling, care must be exercised so as not to damage the isolator. Sometimes the isolator stator separates from the rotor portion; in many cases you can simply push them back together.

• • 9.) Re/installed without the “O”-ring on the OD of stator. (Not applicable all styles)

• • 10.) Constant level oiler is set incorrectly (Cross nut adjustment is set to the wrong height… in this case too high) or it is installed in the wrong bearing housing penetration… or on the wrong side of the pump.Field Tip: Note constant level oilers are rotationally sensitive for proper operation. There is a correct and incorrect side of the pump to install them that depends on the direction of rotation of the pump.

• • 11.) Over-pressurization of the bearing housing. This issue can easily get complicated, but a simple explanation is that the pressure of contained air will rise and fall with temperature. Think “ideal gas” laws. Field Tip: With higher pressure in the bearing housing and the pump operating the resulting oil level can appear to be too low in the sight glass, which will erroneously invite the addition of oil and consequential overfill. Another issue is that the higher pressure can force oil level below the minimum level and the bearing and oil will overheat. The oil will turn black.

• • 12.) Excessive pipe strain. The resultant forces and moments of suction and discharge piping induced stress will cause the bearing bore and consequently the bore for the isolator device to be out of round, nonconcentric or noncongruent. The condition will likely create an offset centerline with the sister bearing and isolator. Field Tip: Check for pipe strain by placing dial indicators (or lasers) in two planes on the coupling hub. Release the bolts on one of the pump flanges…. if the dial indicator moves more than 0.002 inches then there is excessive pipe strain. Repeat for the other flange. Correct the cause of the stress before proceeding.

• • 13.) Excessive vibration. The vibration can be due to cavitation, air binding, impeller imbalance, surging and or operating outside of the allowable operating area.

• • 14.) Misalignment of the driver to the pump.

• • 15.) Drive train component such as a rotating coupling (spinning in the semi-enclosed space of an enclosed OHSA guard) is too close to the external face of the isolator. In rare cases and mostly on larger frame (higher horsepower) pumps, a low-pressure area can be created (by the air velocity also know in this case as windage). The low-pressure area creates a differential pressure outside of the bearing housing near the face of the bearing isolator. Consequently the oil can be picked up and moved with the air flow “windage”. To correct this situation you can distance the coupling and OSHA guard from each other and or from the isolator face. An alternate solution in many cases is to just use a bigger OSHA guard or incorporate perforated mesh metal on the ends of the guard.

• • 16.) Other miscellaneous reasons. Occasionally oil is leaking out of the pump, but it is not due to the bearing isolator. During the repair and maintenance cycles sometimes the bearing housing gaskets are omitted, manufactured in the incorrect material and/or of wrong dimension. In rare instances the bolts may not be properly torqued.

• • 17.) Oil sight glass is located at the incorrect height (manufacturing or repair defect). While not a leak issue, but of interest on this subject. We have witnessed the end user installing the (ANSI) pump next to a wall with the oil sight gage glass (indicator) on the wall side of the pump. To achieve direct line of sight to the indicator, the operator removed the sight gage from the port side of the pump and relocated it on the other side (starboard) of the bearing housing.On most ANSI pumps there is a normally plugged penetration where the finned oil cooler would be located, if used. Both the oil cooler and the sight glass penetrations are sized at 1″ NPT. However, the centerline of the cooler penetration is at a slightly lower level (in this case 2″) than the sight glass penetration to allow for shaft clearance. The result is that the new corresponding oil level will now be too low and consequently incorrect. The resulting bearing life will be severely compromised due to inadequate lubrication.
    
    
    Note that some pump manufacturers have penetrations on both sides of the pump bearing housing for the sight glass installation on either or both sides.

See my Pumps & System’s article for more detailed information on this subject.

Dylan Thomas is a young pump technician (operator-mechanic) at the local chemical plant. One day he noticed that the batch process on unit C was taking longer than normal. He suspected something was wrong with the main pump. The pump was a horizontal ANSI type driven by an induction motor. Dillion thought back to his recent pump fundamentals training and remembered something the instructor had stated numerous times… “The pump will operate on the published curve at the intersection of the system curve if it is able to do so”. If the pump is not operating properly, it could be due to one or several reasons.

Dylan pulled out his notes from the pump school and reviewed the list of possible causes for reduced pump performance (more on that later). After reviewing the list, Dylan remembered that he would need to measure the differential pressure across the pump and convert that result from pressure to head. Additionally he would need to check the maintenance records and pump nameplate for the correct pump size and impeller trim and download the pump curve from the manufacturer’s website.

Evaluation:

Dylan checked the current differential pressure across the pump as 77 psig (178 feet). The shop maintenance records documented the pump at 85 psig (196 feet) differential when it was first installed 3 years ago.

After converting the differential pressure to head, he looked over the pump curve for the hydraulic point where the pump would operate for the stated head based on the known diameter and speed. The head of 178 feet was further out and to the right on the operating curve than when the pump was new at 196 feet. The system’s flowmeter indication did not correspond to the flowrate per the pump curve. The process was taking longer so a larger flowrate didn’t make sense unless the pump was not operating on its published curve.

Diagnosis:

Utilizing the troubleshooting list he checked the following items one at a time:

1. Improper valve position… He checked the system valve lineup as correct.
2. Incorrect speed… Using a strobe tachometer he confirmed the motor driver was at the correct speed.
3. Net Positive Suction Head Available (Insufficient)… Dylan wasn’t sure how to calculate NPSHA but rechecked the tank level was at its normal height and he was confident that the other factors such as temperature (vapor pressure) and friction had not changed.
4. Clogged suction… because of the materials, the unchanged pump suction pressure and the overall system design Dylan was fairly confident that the suction line was not clogged. He conceded it may be a remote possibility but not likely.
5. Improper clearance setting… The pump clearance had been properly set years ago by the distributor’s technician. The pump had operated fine for several years.
6. Direction of rotation was checked as correct.
7. Change in the system curve… Based on the system materials, valve position and process cleanliness this factor was not totally eliminated but was considered a remote possibility.
8. Change in the liquid properties such as viscosity or specific gravity. The company lab tech confirmed through sampling that the liquid properties had not changed.
9. Pump wear… Dylan suspected this could be the most likely reason for the reduced performance and requested a pump shutdown.

Action:

At the earliest opportunity Dylan scheduled the pump shut down and tagged out. Without disassembling the pump he was able to check the total axial travel of the rotating element and noted it was 0.028 inches more than when the pump was new. Dylan reset the impeller clearance in accordance with the IOM (Instruction and Operating Manual) specifications for the size/model and product temperature.

The pump was returned to service and the pressure differential returned to 85 psig. The process time for unit C returned to normal with the improved flowrate.

Conclusion:

As the pump operated over the last few years the clearance between the impeller and the casing had worn by 0.028 inches. The wear occurred slowly over a long period of time and so the operators didn’t notice until the change was significant.

The wear (clearance) was enough to make the pump perform as if it was operating at a lower speed or the impeller had been trimmed. The pump was not operating on its original published curve due to the wear, however still intersected the system curve at a lower head.

In high school, I received a water-resistant watch for my birthday. During the same time, I was also taking SCUBA lessons at the local YMCA. I soon discovered that water resistant does not mean waterproof.

Fast forward 55 years and I witness pump owners applying standard pumps into harsh slurry services and expecting great results. These owner-operators are then both surprised and disappointed when the pump fails prematurely, because the pump was perhaps wear resistant but it was not wear proof.

Know Your Slurry:

In essence there are four classes of slurries covered in Hydraulic Institute/ANSI Standard 12.1-12.6 (2016). The ANSI standard explains how the classes are progressively defined by solid size, shape, type, population, and specific gravity. The determination and ranking of the slurry class is the important first step to help you pick the proper pump, speed, and materials of construction for the service. Slurries that are corrosive and viscous, especially when coupled with higher temperatures will add another factor in the decision process.

Normally for a slurry application, beside the flowrate and total head, we also need to know the following information about the liquid / slurry properties:

Impeller tip speed is another important factor and often misunderstood. You may look at an application and prefer the faster 12-inch impeller that can do the same hydraulics as a slower 24-inch impeller. The impeller tip speeds may be the same for each pump, but over time industry studies have shown that pump wear is more closely related to the actual pump speed (rpm) than it is to the impeller tip speed.

It is still important to consider the peripheral speed and the recommendations for the maximum impeller tip speed are as follows:

Maintain a Healthy Pump Life:

It is imperative to operate the pump as close to the best efficiency point as possible to avoid recirculation wear issues. Slurry recirculation is, in essence, “an effective and efficient sand blasting method for pump internals”.

Slurry pump impellers wear faster than those in other services; It is important to have them balanced both initially and often. There are many good reasons why slurry pumps have beefy bearings and unbalanced impellers is one of them.

Application data sheets are available on our webpage to assist in selection. Don’t let your next slurry pump application be a failure. Put the right pump in service.

Why a Vertical Pump?

Vertical pumps offer a smaller footprint than a horizontal pump of the same size making them an ideal solution for applications with limited space. If you could tour the engineering spaces of a nuclear submarine, you would notice that all of the major pumps are vertical. There is consideration for headroom (overhead) because the compartments must be accessible for operations and maintenance.

Using a vertical pump will greatly simplify the normally complicated piping arrangements associated with barge loading/unloading, tanker/terminal farms, computer cooling systems, HVAC and Hydronic applications, etc. Additionally, for those unplanned applications where a customer needs to add a pump in an already congested area such as the pump room in a high-rise and/or institutional building, the vertical inline pump can be the perfect solution.

One of the most common types of single stage vertical pumps is the inline vertical pump that conforms with the ANSI/Hydraulic Institute standard B73.2 for type OH-3 (VB) pumps. These pumps are designated inline because both the suction and discharge flanges have the same centerline. Similar to installing a common industrial valve, this type of pump will fit directly in line with the piping.

The vertical inline pump does not require a tedious precision alignment of the driver because it is initially positioned, self-aligned and maintained in place by the register (rabbet lock) machine fits to assure positive and permanent alignment. Further, there is no need for a foundation, baseplate, or grout which saves initial installation cost in both materials and labor.

Due to the robust design of vertical pump casings, the maximum allowable flange/nozzle load limits are higher than on a corresponding horizontal pump of similar class and size.

Consistent with a horizontal ANSI B73.1 pump, the vertical design is also a back pull-out design. This means the mechanical seal, pump shaft and impeller can be maintained without disturbing the piping or the driver.

Because the pump is vertically orientated, the stuffing box/mechanical seal chamber is the liquid high point in the pump. Air and other non-condensable gases coming out of solution can be trapped in the seal chamber causing damage to the seal. Since the pump is not self-venting it is recommended to be vigilant when venting and priming the pump and to apply the proper seal pipe plan as applicable.

Due to the vertical orientation, flooded oil lubrication is not an option. The common solution is to elect grease lubricated bearings.

Because the liquid flow into the pump suction must navigate a 90 degree turn before reaching the impeller, the net positive suction head required (NPSHR) will be slightly higher than similar size horizontal pumps. Additionally, this pump is not recommended for paper stock applications because the suction elbow can potentially become a bottleneck and block the paper stock.

Simply because the pump and driver fit neatly “in-line” with the piping does not mean you can ignore the additional weight with regard to pipe stress. Smaller pumps/motors can normally just be installed in the piping but larger units may require some sort of support. Please consult with a piping expert on how to properly support the pump and associated piping system for your application.

What Killed my Pump?

Ironically, after completing a list of best practices for your pump it will die a short time after you commission the system. Why? Because what you didn’t plan for was the elimination of simple pipe strain.

Pipe strain is the misalignment of suction and discharge piping in relation to the pump. Resulting forces are transferred into the pump, stressing internal components, degrading alignment, and ultimately causing premature equipment failure. Ideally, no external force should be required when aligning pipe flanges.

However, while you were busy working on another project, the group that installed the pump used a 10-foot pry bar, three come-alongs, one hoist motor, two mules and some heavy chains to persuade the already installed piping* to matchup and bolt to the pump flanges. If you could remove the bound flange bolts at this point (that are wedged in due to strain) the pipe would likely swing wildly two feet away from the pump flange.

*Note: You should always pipe away from the pump, not to the pump.

A 100-foot run of 6-inch schedule 40 steel pipe will expand over 1.50 inches when heated from ambient to around 200 degrees F. The force exerted on the pump flange would be close to 190,000 lbs. if left unrestrained.

Pump internals are not designed to operate within the environmental forces associated with pipe strain. Leaving out the formulas and stress calculations for simplification… with excess pipe strain the centerline of each bearing is no longer congruent to the other so the shaft does not run true. Further, the shape of each bearing is no longer round, the stress has made it eccentric.

Remember, the most expensive pipe anchor you could ever purchase is a pump.

Why Not?

1. There is NO OIL in the pump.
2. The impeller may or may NOT be set to the proper clearance.
3. The driver is NOT aligned to the pump.
4. The direction of rotation on the motor has NOT been determined.
5. The mechanical seal is NOT set.

If you already know these 5 things and fully understand the significance, then you can stop here. If you don’t know or would like a refresher please read on.

A pump shipped from the factory may or may not have the proper axial clearance when it arrives on site. The factory adjusts the clearance at a nominal setting for the pump type and size based on ambient temperature water specifications.

The factory does not know the liquid’s temperature or other properties for the operating system. Note: it is also very possible the settings could have been adjusted after it left the factory. Confirming the clearance in the field is both easy and a professional best practice. Why take the chance? Also, prior to running the pump is the perfect time to take the initial total axial movement readings for the maintenance records.

The driver will NOT be aligned precisely to a pump shipped from the factory. The factory utilizes laser manufactured templates for layout and performs a series of nominal checks to ascertain that the motor can be precisely aligned to the pump. Even if the factory did align the driver to the pump in accordance with the highest standards… as soon as the skid is picked up/transported the precision alignment will morph to unacceptable levels.

To learn more about about pump alignment, please check with your regional sales manager or refer to my articles on this subject:

• Does Your Pump Have an Alignment Problem?
• 19 Tips and Common Alignment Mistakes

A pump assembly shipped from the factory will NOT have the coupling spacer installed because you must first complete the driver rotational check with the coupling (spacer) removed. Additionally, with the coupling removed the process to set the impeller and mechanical seal is simplified.

The factory has a 50/50 chance of guessing the correct electrical phase rotation at your local site. If the rotation is wrong, the pump quickly converts to an expensive pile of useless scrap metal shortly after startup.

Factory installed mechanical seals will NOT be set. The pump comes with the seal clips in place (sleep position) to ostensibly preclude damage to the seal during shipping and handling. Plus prior to setting the seal the impeller clearance will need to be checked/set and the alignment completed.

A warning tag is attached to each pump to communicate these 5 key steps to the end user/installer. Of course these steps have always been stated in the Instruction and Operations Manual (IOM). The IOM is included with every pump and if misplaced can also be downloaded from our website.