Vertical Turbine Pump Information

Section B -- Pump Application Data 1. DATUM OR GRADE - The elevation of the surface from which the pump is supported. 2. STATIC LIQUID LEVEL - The vertical distance from grade to the liquid level when no liquid is being drawn from the well or source. 3. DRAWDOWN - The distance between the static liquid level and the liquid level when pumping at required capacity. 4. PUMPING LIQUID LEVEL - The vertical distance from grade to liquid level when pumping at rated capacity. Pumping liquid level equals static water level plus drawdown. 5. SETTING - The distance from grade to the top of the pump bowl assembly. 6. TPL (TOTAL PUMP LENGTH) - The distance from grade to lowest point of pump. 7. RATED PUMP HEAD - Lift below discharge plus head above discharge plus friction losses in discharge line. This is the head for which the customer is responsible and does not include any losses within the pump. 8. COLUMN AND DISCHARGE HEAD FRICTION LOSS - Head loss in the pump due to friction in the column assembly and discharge head. Friction loss is measured in feet and is dependent upon column size, shaft size, setting, and discharge head size. Values given in appropriate charts in Data Section. 9. BOWL HEAD - Total head which the pump bowl assembly will deliver at the rated capacity. This is curve performance. 10. BOWL EFFICIENCY- The efficiency of the bowl unit only. This value is read directly from the performance curve.

11. BOWL HORSEPOWER- The horsepower - required by the bowls only to deliver a specified capacity against bowl head.

12. TOTAL PUMP HEAD - Rated pump head plus column and discharge head loss. Note: This is new or final bowl head. 13. SHAFT FRICTION LOSS - The horsepower required to turn the lineshaft in the bearings. These values are given in appropriate table in Data Section. 14. PUMP BRAKE HORSEPOWER - Sum of 'bowl horsepower plus shaft loss (and the driver thrust bearing loss under certain conditions). 15. TOTAL PUMP EFFICIENCY (WATER TO WATER) -The efficiency of the complete pump less.the driver, with all pump losses taken into account.

16. OVERALL EFFICIENCY (WIRE TO WATER)-The efficiency of the pump and motor complete. Overall efficiency = total pump efficiency X motor efficiency. 17. SUBMERGENCE-Distance from liquid level to suction bell.

Vertical Turbine Pumps Section 2 Section B -- Vertical Turbine Pumps Turbine Nomenclature Vertical Turbine Pumps Calculating Axial Thrust Under normal circumstances Vertical Turbine Pumps have a thrust load acting parallel to the pump shaft. This load is due to unbalanced pressure, dead weight and liquid direction change. Optimum selection of the motor bearing and correct determination of required bowl lateral for deep setting pumps require accurate knowledge of both the magnitude and direction (usually down) of the resultant of these forces. In addition, but with a less significant role, thrust influences shaft H.P. rating and shaft critical speeds. IMPELLER THRUST Impeller Thrust in the downward direction is due to the unbalanced discharge pressure across the eye area of the impeller. See diagram A.

Counteracting this load is an upward force primarily due to the change in direction of the liquid passing through the impeller. The resultant of these two forces constitutes impeller thrust. Calculating this thrust using a thrust constant (K) will often produce only an approximate thrust value because a single constant cannot express the upthrust component which varies with capacity.

To accurately determine impeller thrust, thrust-capacity curves based on actual tests are required. Such curves now exist for the "A" Line. To determine thrust, the thrust factor "K" is read from the thrust-capacity curve at the required capacity and given RPM. "K" is then multiplied by the Total Pump Head (Final Lab Head) times Specific Gravity of the pumped liquid. If impeller thrust is excessively high, the impeller can usually be hydraulically balanced. This reduces the value of "K". Balancing is achieved by reducing the discharge pressure above the impeller eye by use of balancing holes and rings. See diagram B. NOTE: Although hydraulic balancing reduces impeller thrust, it also decreases efficiency by 1 to 5 points by providing an addi-tional path for liquid recirculation. NOTE: Although hydraulic balancing reduces impeller thrust, it also decreas-es efficiency by one to five points by providing an additional path for liquid recirculation. Of even greater concern is that should the hydraulic balancing holes become clogged, (unclean fluids, fluids with solid content, intermittent services, etc.), the impeller thrust will increase and possibly cause the driver to fail. Hydraulically balanced impellers cannot be used in applications requiring rubber bowl bear-ings because the flutes on the inside diameter of the bearings pro-vide an additional path to the top side of the impeller, thus creating an additional down thrust. Hydraulically balanced impellers should be used as a "last resort" for those situations where the pump thrust exceeds the motor thrust bearing capabilities. Section B -- Vertical Turbine Pumps Turbine Nomenclat DEAD WEIGHT In addition to the impeller force, dead weight (shaft plus impeller weight less the weight of the liquid displaced) acts downward. On pumps with settings less than 50 feet, dead weight may be neglect-ed on all but the most critical applications as it represents only a small part of the total force. On deeper setting pumps, dead weight becomes significant and must be taken into account. NOTE: We normally only take shaft weight into consideration as dead weight, the reason being that impeller weight less its liquid displace-ment weight is usually a small part of the total. SHAFT SLEEVES Finally, there can be an upward force across a head shaft sleeve or mechanical seal sleeve. In the case of can pumps with suction pressure, there can be an additional upward force across the impeller shaft area. Again for most applications these forces are small and can be neglected; however, when there is a danger of upthrusts or when there is high discharge pressure (above 600 psi) or high suction pressure (above 400 psi) these forces should be considered. MOTOR BEARING SIZING Generally speaking a motor for a normal thrust application has as standard, a bearing adequate for shutoff thrust. When practical, motor bearings rated for shutoff conditions are preferred. For high thrust applications (when shutoff thrust exceeds the standard motor bearing rating) the motor bearing may be sized for the maximum anticipated operating range of the pump. Should the pump operate to the left of this range for a short period of time, anti-fraction bearings such as angular contact or spherical roller can handle the overload. It should be remembered, however, that bearing life is approximately inversely proportional to the cube of the load. Should the load double, motor bearing life will be cut to 1 ‫ ڸ‬of its original value. Although down thrust overloading is possible, the pump must never be allowed to operate in a continuous up thrust condition even for a short interval without a special motor bearing equipped to handle it. Such upthrust will tail the motor bearing. CALCULATING MOTOR BEARING LOAD As previously stated, for short setting non-hydraulic balanced pumps below 50 feet with discharge pressures below 600 psi and can pumps with Suction pressures below 100 psi only impeller thrust need be considered. Under these conditions:

Motor Bearing Load (lbs.) Timp= KHL, x SG Where: Impeller Thrust (lbs.) K=Thrust factors (lbs./ft.) HL = Lab Head (ft.) SG = Specific Gravity For more demanding applications, the forces which should be considered are impeller thrust plus dead weight minus any sleeve or shaft area force. In equation form: Motor Bearing Load = Timp + Wt(1) - sleeve force (2) -shaft area force(3) =Tt

(1) Wt.= Shaft Dead Wt. x Setting In Ft. (2) Sleeve Force=Sleeve area x Discharge pressure (3) Shaft Area Force = Shaft area x Suction pressure Oil Lube shaft does not displace liquid above the pumping water level and therefore has a greater net weight. THRUST BEARING LOSS Thrust bearing loss is the loss of horsepower delivered to the pump at the thrust bearings due to thrust. In equation form:

where: LTB = Thrust bearing loss (HP) BHP = Brake horsepower Tt = Motor Bearing Load (Lbs.) = Timp+ Wt(1) - sleeve force(2) - shaft area force(3)

Key Factors to Consider Before Applying a Vertical Turbine Pump Vertical Turbine Pumps (figure 1) can be an attractive choice for many water and wastewater applications because of their many advantages. The vertical construction takes up little floor space; priming problems can be avoided due to the impellers being submerged in the liquid; the first stage impeller can be lowered (by increasing the pit depth, if necessary) to provide the desired NPSH margin; the multistage construction offers higher efficiencies on high head, lower flow applications; and the modular construction allows the pumps to be customized for many applications.

Potential Critical Issues Care must, however, be taken when applying a vertical turbine pump (VTP) in short-set applications, due to a number of critical issues that can occur with this unique pump construction. Vertical Turbine pumps were originally developed for deep-water wells, but the design was found to have certain benefits in industrial and municipal applications. But the conditions that allow successful operation of VTP's in deep-water well applications are not always present in shorter set water and wastewater services. 1. Vertical turbine pumps do not balance the axial hydraulic pressure forces on the impeller, which are substantial and with a downward direction. Instead they rely on the thrust bearing in the motor to carry this high load. This motor thrust bearing must also carry the rotor weight, which can be quite large for deep well settings. Although high axial thrust is normally a bad thing in most pumps, in a VTP it is actually a good thing.

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This high axial thrust actually compensates for alignment and resonant frequency issues that often exist in the standard vertical pump construction, especially when applied on short set applications. The threaded line shaft couplings that are used in the standard VTP construction cannot hold accurate shaft alignment, because the diameter of the line shaft ends, which butt together, are too small compared to the section shaft lengths. In order to hold a reasonable concentricity (say .005 inches) of the line shaft at the end of a typical 10 feet span, each shaft end surface would have to be machined square to within a tolerance of ± .000025 inches. This is not possible with current machining technology. Vertical turbine pumps, instead, rely on the high axial thrust from the unbalanced pressure force on the impellers, plus the rotor weight to keep the shaft sections straight enough for successful operation. 2. Further, being so long and narrow, VTP components can be subject to high vibration. The typical line shaft bearing spacing is 10 feet, which means that most line shafts are operating above their first critical speed. Further, vertical turbine pumps have more then one "first critical speed", depending on the shaft end condition. First, there is the first "normal" (free-free end supports) critical speed, but instead of the next (normally second) critical speed being four times the first critical speed, it is actually 1.56 times the first critical speed (for free-fixed end supports), and 2.58 times the first critical speed for fixed-fixed end supports. Further, if you add reasonable safety margins around these various critical speeds, it becomes difficult to avoid operating very far away from a shaft critical speed, especially with a variable speed drive. Here again, it is the high axial thrust that helps typical vertical pumps in deep-well applications to achieve reasonable life spans. 3. So what happens when a VTP is applied in a short set application, and operated at high flow rates / low discharge pressures? Well, without the resulting high axial thrust, line shaft vibration and bearing loads can be greatly increased, reducing the life of the internal sleeve bearings. 4. Further, some or all of the internal sleeve bearings are lubricated / wetted by the liquid pumped, so they must be able to handle any corrosives and/or abrasives in the pumpage. Further, the bearings must be able to handle any air in the pumped liquid. The amount of entrained air can be substantially increased if there is any cavitation in the pump. Based on tests conducted by the writer for a Texas A&M Pump User's Symposium paper, the amount of dissolved air released (to become entrained air - the damaging kind), is greatly increased as the NPSH Margin (NPSHA / NPSHR) is reduced (see figure 1). While cavitation can be present in a centrifugal pump up to a NPSH Margin Ratios as high as 4 or more, the entrained air liberated by cavitation begins to dramatically increase as the NPSH Margin Ratio approaches around 1.4, near the best efficiency point (bep) flow rate, and 2.0 around the low flow suction recirculation condition. This means that as the NPSH Available from the sump (system) approaches the pump NPSH Required, that the internal sleeve bearings, especially those in the bowl assembly, will see more air (dry operation). Now if these bowl bearings are bronze (typical), and limited in lead content (which gives bronze bearings their lubricity), due to regulations in the drinking water industry, this can result in very short bearing life.

5. A vertical turbine pump can also experience excessive structural vibration of the discharge head and associated driver in the field, even though the driver has been shop tested with low vibration, and the rotating components were properly balanced. This could be the result of a "reed" (natural) frequency of the motor, head and foundation assembly. This is occasionally caused by the stiffness of the field foundation being different then anticipated. A "reed resonance" effect will result if the natural frequency of the assembly is at or near (within 15% of) the running frequency of the pump. Variable speed operation drastically increases this likelihood.

UPGRADES FOR SHORT SETTINGS: So does this mean that vertical turbine pumps should not be used for short settings, or variable speed operation? Not necessarily, if consideration is given to the following pump up grades: 1. Replace the standard threaded line shaft couplings with a clamp type (or equivalent) coupling that insures line shaft straightness. 2. Provide bowl line shaft bearings with better lubricity, such as those made of Vespel CR-6100. 3. Reduce the bearing spacing from 10 feet to 5 feet, to insure that the shaft first critical frequency is at least 15% above the maximum operating speed. 4. Change from threaded to flanged column pipe connections, for improved housing straightness. 5. Perform a Finite Element Analysis of the motor, head and foundation assembly. 6. Use a Suction Bell, instead of a Suction Case to minimize the required submergence and NPSH Requirement of the pump. 7. Insure that the NPSH Margin is at least "1.4", if the liquid end (bowl assembly) has "High Suction Energy" (see October 2007, Pump Tips Column).

Back to Basics: How to Improve Vertical Turbine Pump Reliability through Optimum Bearing Selection Vertical turbine pumps (VTPs) offer many unique advantages for many applications. For instance, the vertical construction takes up little floor space; priming problems can be avoided due to submersion of the impellers in liquid; the first stage impeller can be lowered (by increasing the pit depth, if necessary) to provide the desired NPSH margin; multistage construction and midrange specific speeds offer high efficiencies; and modular construction allows the pumps to be customized for many applications. VTPs are available in deep well, wet pit (short setting or close-coupled), canned, and submersible motor configurations. Accordingly, my September 2008 column on the advantages and cautions of using VTPs on water and wastewater applications concentrated primarily on cavitation, vibration and axial thrust, as well as how to avoid the associated field problems. An additional key issue that should be considered when applying a VTP on liquids that contain solids, abrasives and/or air is the selection of the bearing material and/or construction, given the fact that the bearings are immersed in and lubricated by the fluid pumped during most typical applications. As such, they are also generally the first component to deteriorate in a VTP.

Conventional Vertical Turbine Pump Bearings VTP bearings are found in the bowl assembly (in each bowl, suction case/bell and possibly the discharge case) and in the column assembly (unless the pump has a short setting or is driven by a submersible motor located below the bowl assembly). Bowl bearings are normally made of a low-lead bronze material, set against a 416 stainless steel bowl shaft without any replaceable sleeves (see Fig. 1).

Further, the column assembly connects the bowl assembly to the aboveground discharge head. Typical column bearings are either constructed of a cutless rubber (see Fig. 2), operating against a stainless steel shaft sleeve (lubricated by the fluid pumped) or bronze enclosed in a tube (lubricated by either an oil drip or water flush, introduced at the discharge head, and exiting into the well or sump at the top of the bowl assembly; see Fig. 3).

Open Lineshaft Open column line shaft bearing construction is recommended for ease of maintenance and/or whenever a special bearing material is required; it is not recommended for longer settings greater than about 100 feet. Renewable shaft sleeves or hard facing on the shaft are available for longer life, and typical bearing spacing is 10 feet for well applications. However, for shorter settings, the shaft size and spacing should be selected so that the shafting will operate below its first critical speed (see Fig. 4). For example, at 1,800 RPM, the maximum bearing spacing for a 1 11/16 inch (1.69 inch) shaft would be five feet.

Enclosed Column In this configuration, an enclosing tube provides the lineshaft with protection from the pumped liquid and ensures clean lubrication to the bearings prior to startup, which is especially important for deeper settings (over about 100 feet). The lineshaft bearings are typically spaced at five-foot intervals to support the lineshaft. An internal spiral groove allows the lubricant to flow between the shaft and the inner face of the bearing, while the outside of the bearing is threaded to connect the enclosing tube sections.

This construction minimizes maintenance of the column bearings in abrasive services. The oil (or water) lubrication for the enclosed construction is introduced at the surface. A tank attached to the discharge head provides oil through a solenoid valve to the tension bearing in the stuffing box. It then flows by gravity into the enclosing tube and through the bypass port in the bowl assembly discharge case. Alternate lubricants such as clean water or grease can also be used with enclosed lineshaft construction. Lubricating oils are available that are acceptable for discharge into the pumped liquid, even when it is intended for drinking water.

Bowl Assembly Bearings Line shaft bearings can be protected from abrasive wear by either constructing them of cutless rubber (which can tolerate fairly high levels of suspended solids) or by the use of an enclosed tube around the lineshaft; bowl bearings, however, must operate in the pumped liquid. This means that, in most cases, the choice of bearing material (see Table 1) is normally the only option that will allow the pump to handle higher levels of solids or air/vapor. Air can enter a VTP when the well or sump levels are low (vortexing), can be entrained in the pumpage due to mixing, or can be released from entrainment due cavitation in the first and/or second stage of the pump (as discussed in the September 2008 column). Once in the bowl assembly, the air and vapors (being lighter then water) can be centrifuged into the bearings. Some VTP manufacturers do, however, offer rifle-drilled bowl assembly shafts with an external water flush that can greatly improve the bowl bearing life when handling solids/abrasives and/or air/vapor, as shown in Figure 1.

Bearing Material Options Table 1 lists the classes of column and bowl assembly bearing materials generally available for VTPs, with each manufacturer typically offering their own specific alloys. The bronze alloys typically offered have very low (if any) lead due to environmental concerns, which reduces their dry-running ability. Carbon graphite bearings probably offer one of the most efficient dry-running capabilities but have very low solids/abrasive tolerance; they are also available with a variety of fillers. Teflon bearings also have excellent dry-running capability and poor abrasive tolerance, plus they are available with a variety of fillers. However, Teflon bearings cannot be retained with a press fit due to the potential of cold flow. Vespel bearings are also much like Teflon bearings but with improved dimensional stability, and they are also available with a variety of fillers, including Teflon and carbon. Rubber bearings are primarily used for open lineshaft column applications and are very proficient at handling solids, as long as they are not too sharp. However, they have poor lubricity and should be wetted prior to startup, which could pose a problem with deep settings (over about 100 feet), especially if the pump does not have a foot (check) valve and it takes too long for the pumped liquid to reach the upper bearings. Rubber bearings are seldom used for bowl bearings due to the larger required running clearance. Finally, hardened surface metal bearings are also available, such as chromium oxide and tungsten carbide, for abrasive/solid applications, but they are expensive and also have poor lubricity for handling air and/or vapors. About the Author: Allan R. Budris, P.E., is an independent consulting engineer who specializes in training, failure analysis, troubleshooting, reliability, efficiency audits, and litigation support on pumps and pumping systems. With offices in Washington, N.J., he can be contacted via email at [email protected].

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