pump clinic _ centrifugal troubleshooting chapters 1 - 43

Kelair Pumps Australia Pty Ltd ABN 28 001 308 381 215 Walters Road Arndell Park NSW 2148Ph: 1300 789 466 Fax: 02 9678 9455 Email: [email protected] www.kelairpumps.com.auQLD Fax: 07 3808 8758 VIC Fax: 03 9569 7866 TAS Fax: 03 6331 9102 WA Fax: 08 9248 2255

Pump Clinic 1 Centrifugal Troubleshooting 28/03/06 of 4

PUMP CLINIC 1

Centrifugal Troubleshooting

TABLE 1 - IDENTIFICATION CHART

SYMPTOMSPOSSIBLE CAUSES

(See definitions Table 2)

1. Pump does not deliver liquid 1/2/3/510/12/13/14/16/21/22/25/30/32/38/40

2. Insufficient capacity delivered2/3/4/5/6/7/7a/10/11/12/13/14/15/16/17/18/21/22/23/24/25/31/32/40/41/44/63/64

3. Insufficient pressure developed4/6/7/7a/10/11/12/13/14/15/16/18/21/22/23/24/25/34/39/40/41/44/63/64

4. Pump loses prime after starting 2/4/6/7/7a/8/9/10/11

5. Pump requires excessive power20/22/23/24/26/32/33/34/35/39/40/41/44/45/61/69/70/71

6. Pump vibrates or is noisy at all flows2/16/37/43/44/45/46/47/48/49/50/51/52/53/54/55/56/57/58/59/60/61/67/78/79/80/81/82/83/84/85

7. Pump vibrates or is noisy at low flows 2/3/17/19/27/28/29/35/38/77

8. Pump vibrates or is noisy at high flows 2/3/10/11/12/13/14/15/16/17/18/33/34/41

9. Shaft oscillates axially 17/18/19/27/29/35/38

10. Impeller vanes are eroded on visible side 3/12/13/14/15/17/41

11. Impeller vanes are eroded on invisible side 12/17/19/29

12. Impeller vanes are eroded at discharge near centre 37

13. Impeller vanes are eroded at discharge near shrouds or atshroud/vane fillets

27/29

14. Impeller shrouds bowed out or fractured 27/29

15. Pump overheats and seizes1/3/12/28/29/38/42/43/45/50/51/52/53/54/55/57/58/59/60/61/62/77/78/82

16. Internal parts are corroded prematurely 66

17. Internal clearances wear too rapidly 3/28/29/45/50/51/52/53/54/55/57/59/61/62/66/77

18. Axially split casing is cut through wire-drawing 63/64/65

19. Internal stationary joints are cut through wire-drawing 53/63/64/65

20. Packed box leaks excessively or packing has short life 8/9/45/54/55/57/68/69/70/71/72/73/74

21. Packed box sleeve is scored 8/9

22. Mechanical seal leaks excessively 45/54/55/57/58/62/75/76

23. Mechanical seal has damaged faces, sleeve bellows 45/54/55/57/58/62/75/76

24. Bearings have short life3/29/41/42/45/50/51/54/55/58/77/78/79/80/81/82/83/84/85

25. Coupling fails 45/50/51/54/67



TABLE 2 - DEFINITIONS

Suction Troubles

1. Pump not primed2. Pump suction pipe not completely filled with liquid3. Insufficient available NPSH4. Excessive amount of air of gas in liquid5. Air pocket in suction line6. Air leaks into suction line7. Air leaks into pump through stuffing boxes or through mechanical seal7a. Air in source of sealing liquid8. Water seal pipe plugged9. Seal cage improperly mounted in stuffing box10. Inlet of suction pipe insufficiently submerged11. Vortex formation at suction12. Pump operated with closed or partially closed suction valve13. Clogged suction strainer14. Obstruction in suction line15. Excessive friction losses in suction line16. Clogged impeller17. Suction elbow in plane parallel to the shaft (for double-suction pumps)18. Two elbows in suction piping at 90º to each other, creating swirl and pre-rotation19. Selection of pump with too high a suction specific speed

Other Hydraulic Problems

20. Speed of pump too high21. Speed of pump too low22. Wrong direction of rotation23. Reverse mounting of double-suction impeller24. Uncalibrated instruments25. Impeller diameter smaller than specified26. Impeller diameter larger than specified27. Impeller selection with abnormally high head coefficient28. Running the pump against a closed discharge valve without opening a bypass29. Operating pump below recommended minimum flow30. Static head higher than shut-off head31. Friction losses in discharge higher than calculated32. Total head of system higher than design of pump33. Total head of system lower than design of pump34. Running of pump at too high a flow (for low specific speed pumps)35. Running pump at too low a flow (for high specific speed pumps)36. ----37. Too close a gap between impeller vanes and volute tongue or diffuser vanes38. Parallel operation of pumps unsuitable for the purpose39. Specific gravity of liquid differs from design conditions40. Viscosity of liquid differs from design conditions41. Excessive wear at internal running clearances42. Obstruction in balancing device leak-off line43. Transients at suction source (imbalance between pressure at surface of liquid and vapour pressure

at suction flange)



Mechanical Troubles - General

44. Foreign matter in impellers45. Misalignment46. Foundation insufficiently rigid47. Loose foundation bolts48. Loose pump or motor bolts49. Inadequate grouting of baseplate50. Excessive piping forces and movements on pump nozzles51. Improperly mounted expansion joints52. Starting the pump without proper warm-up53. Mounting surfaces of internal fits (at wearing rings, impellers, shaft sleeves, shaft nuts, bearing

housings, etc) not perpendicular to shaft axis54. Bent shaft55. Rotor out of balance56. Parts loose on the shaft57. Shaft running off-centre because of worn bearings58. Pump running at or near critical speed59. Too long a shaft span or too small a shaft diameter60. Resonance between operating speed and natural frequency of foundation, baseplate or piping61. Rotating part rubbing on stationary part62. Incursion of hard solid particles into running clearances63. Improper casing gasket material64. Inadequate installation of gasket65. Inadequate tightening of casing bolts66. Pump materials not suitable for liquid handled67. Certain couplings lack lubrication

Mechanical Troubles - Sealing Area

68. Shaft or sleeves worn or scored at packing69. Incorrect type of packing for operating conditions70. Packing improperly installed71. Gland too tight, prevents flow of liquid to lubricate packing72. Excessive clearance at bottom of stuffing box allows packing to be forced into pump interior73. Dirt or grit in sealing liquid74. Failure to provide adequate cooling liquid to water-cooled stuffing boxes75. Incorrect type of mechanical seal for prevailing conditions76. Mechanical seal improperly installed

Mechanical Troubles - Bearings

77. Excessive radial thrust in single-volute pumps78. Excessive axial thrust caused by excessive wear at internal clearances or by failure or, if used,

excessive wear of balancing device79. Wrong grade of grease or oil80. Excessive grease or oil in anti-friction bearing houses81. Lack of lubrication82. Improper installation of anti-friction bearings such as damage during installation, incorrect assembly

of stacked bearings, use of unmatched bearings as a pair, etc83. Dirt getting into bearings84. Moisture contaminating lubricant85. Excessive cooling of water-cooled bearings



TABLE 3 - DIAGNOSIS FROM APPEARANCE OF STUFFING BOX PACKING

SYMPTOMS CAUSES

� Wear on one or two rings next to packing gland (other rings OK)

Improper packing installation

� Wear on outside diameter of packing ringsPacking rings rotating with shaft sleeve or leakagebetween rings and inside diameter of box. Wrongpacking size or incorrectly cut rings

� Charring or glazing of inner circumference of rings

Excessive heating. Insufficient leakage to lubricatepacking or unsuitable packing

� Inside diameter of rings excessively increased or heavily worn on part of inner circumference

Rotation eccentric

TABLE 4 - VIBRATION

VIBRATION FREQUENCY CAUSES

� Several times pump r/min Bad anti-friction bearings

� Twice pump r/minLoose parts on rotor, axial misalignment ofcoupling, influence of twin-volute when gap isinsufficient

� Running speedImbalance of rotor, clogged impeller, couplingmisalignment

� Running speed times number of impeller vanes

Vane passing syndrome – insufficient gap betweenimpeller vanes and collector vanes. This is alsosometimes seen during operation with suctionrecirculation

� One-half running speed Oil whirl in bearing

� Random low frequency Internal recirculation in impeller or cavitation

� Random high frequency Usually resonance

� Sub synchronous frequency at 70% to 90% Hydraulic excitation of resonance

Kelair Pumps Australia Pty Ltd ABN 28 001 308 381 215 Walters Road Arndell Park NSW 2148 Ph: 1300 789 466 Fax: 02 9678 9455 Email: [email protected] www.kelairpumps.com.au QLD Fax: 07 3808 8758 VIC Fax: 03 9569 7866 TAS Fax: 03 6331 9102 WA Fax: 08 9248 2255 PUMPS STEAM TURBINES BUILDING & FIRE WASTEWATER SERVICE

Pump Clinic 2 Introduction to Cavitation & Net Positive Suction Head 03/05/06 Page 1 of 6

PUMP CLINIC 2

Introduction to Cavitation and Net Positive Suction Head Cavitation in Centrifugal Pumps When the pressure of flowing liquids drops to, or below the liquid’s vapour pressure, the liquid boils and vapour cavities (bubbles) form locally inside the liquid. If the pressure within the flow path subsequently increases above the vapour pressure, the vapour cavities implode, releasing energy. The formation and sudden collapse of these bubbles is called Cavitation. The generation of head in a centrifugal pump does not commence until the liquid enters the vane area and is accelerated towards pump discharge. As the liquid flows between the pump inlet flange and vanes, several points of head loss occur due to: a) Friction in the suction nozzle.

b) Acceleration losses as the liquid velocity increases from the suction nozzle to the impeller eye.

c) Shock losses as the liquid contacts the leading edges of the impeller vanes.

The sum of these losses is known as the entry loss. If the suction head minus the entry loss reduce the liquid pressure to or below the vapour pressure, then a condition for cavitation exists. Figure 1 Page 2 illustrates the above. Net Positive Suction Head (NPSH) The net positive suction head is a statement of the minimum suction conditions required to prevent cavitation. The required NPSH (referred to as NPSHR) is the minimum value of NPSH required at the pump inlet for satisfactory pump operation and must be determined by test and is stated by manufacturers (appears on the pump performance curve as an NPSHR curve). The NPSHR is equivalent to the entry loss as shown in Figure 1.



The available NPSH (referred to as NPSHA) is a function of the suction side pumping system and is defined as the absolute pressure head on the liquid surface, plus the static liquid level above the pump centre line (negative for suction lift) minus the friction loss in the piping system leading to the pump minus the vapour pressure head at the pumping temperature. The discharge pumping system has no effect on NPSHA. Figure 2 Page 4 shows four typical suction systems with the NPSHA formulae applicable to each. Please note that the units of the terms in the formulae are metres absolute of the liquid being pumped. To avoid cavitation, NPSHA must always be greater than NPSHR at the design flow. Problems Caused By Cavitation The presence of cavitation due to inadequate NPSH can be diagnosed during pump operation by a steady crackling noise in and around the pump suction. This should not be confused with a random crackling noise with high intensity knocks which indicates another condition termed suction recirculation (not covered in these notes). If the problem was one of noise alone, it is likely that most situations would call for no remedial action. However, continual cavitation causes mechanical and operational problems as follows: 1) Erosion of impeller, particularly at the leading edges of the impeller vanes. In some

cases, the casing itself will show signs of erosion. The extent of damage experienced is significantly affected by product-related factors such as corrosion and abrasion. Apart from the damage to the parts, the erosion can cause loss of pump efficiency and out-of-balance problems with the impeller.

2) The vibrations caused by cavitation and unbalanced loads significantly accelerate the rate of bearing and mechanical seal failures.

3) The vapour cavities will impede the flow of liquid through the impeller. In some cases, the flow may be completely blocked. This will result in reduced capacity plus reduced and/or unstable developed head.



What Can be Done? The obvious answer is to ensure proper pump selection at the initial stage. Most users would agree that the majority of pump vendors are sufficiently competent in giving customers what was asked for in the specifications. Having said this, it is imperative that the issue of NPSHA vs NPSHR is properly understood and considered by both user and supplier. The solutions to existing cavitation conditions can be determined by considering both NPSHA (the system) and NPSHR (the pump). On the system side, the NPSHA can be increased by one or more of the following: a) Increase the static liquid level above the pump or reduce the suction lift. This can

be done in the case of a flooded suction by raising the liquid level in the suction tank, raising the suction tank to a higher level or lowering the pump, e.g. one floor down.

In the suction lift situation, the liquid level in the sump or suction tank can be raised or the pump can be lowered, e.g. mounting the pump off the sump side or building a dry sump beside the existing sump.

b) Reduce the friction losses by increasing pipe sizes and reducing the length of pipe runs and the number of fittings, e.g. tees, bends, valves. Selection of fittings with lower friction loss, e.g. long radius elbows and full flow ball valves should also be considered. In particular, resist the use of suction strainers that can clog.

c) Reduce the vapour pressure by reducing the temperature of the product. This can

be done by reducing the operational temperature of the process (if feasible) or cooling the temperature in the suction line, e.g. cooling annulus on the suction pipework. It must be noted the reduction of vapour pressure by reducing the temperature is rarely possible.

The remedies detailed below can be applied to the pump: a) Reduce the flow rate by throttling on the pump discharge. This will generally

reduce NPSHR (always check the pump curve) and increase NPSHA (due to reduced friction losses). Care must be taken to ensure that the flow rate is not reduced below the minimum flow rate recommended by the manufacturer.

b) Reduce the pump speed as this reduces NPSHR. This will require the user to

accept reduced pump performance c) Reduce the pump speed and install a larger diameter impeller. This will have a

two-fold effect as lower speed means lower NPSHR and in many cases the larger impeller diameter has lower NPSHR characteristics.

d) Install a different pump. This would normally mean installation of a larger pump as

they generally have a better NPSHR value for the same flow rate. The selection of a larger pump is sometimes required with speed reduction.



e) Change the impeller material to one that is more resistant to erosion, e.g. from cast

iron to stainless steel. This does not eliminate cavitation but will reduce the impact of cavitation.

What Types of Installations Are More Likely To Encounter Cavitation Problems? The presence of any of the conditions detailed below significantly increases the possibility of low NPSHA values and cavitation. Please consider this in your pump selections and if in doubt, discuss the matter with the pump supplier. 1. High temperature or boiling liquids: This will increase the vapour pressure

head. 2. Volatile liquids: These have a high vapour pressure head. 3. Suction tank under vacuum: This will reduce the absolute pressure head on the

liquid surface. 4. High suction lift applications 5. Circuitous suction pipework: This will lead to increased friction loss. 6. A high number of fittings in suction pipework: This will increase friction losses. Conclusion Cavitation and NPSH seem to be some of the least understood topics associated with pump applications. To some engineers, these topics appear mysterious or, at best, are only partially understood. Many highly technical research papers have been written on this subject. These notes are an attempt to give a simple introduction to cavitation and highlight areas where some critical thought should be applied. It should also be stated that although these notes are based on application to centrifugal pumps, the majority of the principles apply equally to all other types of pumps.


Pump Clinic 3 What is Recirculation & Separation 24/05/06 Page 1 of 4

PUMP CLINIC 3

What is Recirculation and Separation? There is a small flow from impeller discharge to suction through the wearing rings and any hydraulic balance device present. This takes place at all capacities but does not usually contribute to raising the liquid temperature very much unless operation is near shut-off. When the capacity has been reduced by throttling (or as a result of an increase in system head), a secondary flow called recirculation begins. Recirculation is a flow reversal at the suction and/or at the discharge tips of the impeller vanes. All impellers have a critical capacity at which recirculation occurs. The capacities at which suction and discharge recirculation begin can be controlled to some extent by design, but recirculation cannot be eliminated. Suction recirculation is the reversal of flow at the impeller eye. A portion of the flow is directed out of the eye at the eye diameter, as shown in Figure 1 Page 1 and travels upstream with a rotational velocity approaching the peripheral velocity of the diameter. A rotating annulus of liquid is produced upstream from the impeller inlet and through the core of this annulus passes an axial flow corresponding to the output capacity of the pump. The high shear rate between the rotating annulus and the axial flow through the core produces vortices that form and collapse, producing noise and cavitation in the suction of the pump.

Discharge recirculation is the reversal of flow at the discharge tips of the impeller vanes, as shown in Figure 2 Page 1. The shear rate between the inward and outward relative velocities produces vortices that cavitate and usually attack the pressure side of the vanes.



The capacity at which suction recirculation occurs is directly related to the design suction-specific speed S of the pump. The higher the suction-specific speed, the closer will be the beginning of recirculation to the capacity at best efficiency. Figure 3 Page 2 shows the relation between the suction-specific speed and suction recirculation for pumps up to 2500 (1530) specific speed and Figure 4 Page 2 shows the same relation for pumps up to 10,000 (6123) specific speed. For water pumps, the minimum operating flows can be as low as 50% of the suction recirculation values shown for continuous operation and as low as 25% for intermittent operation. For hydrocarbons, the minimum operating flows can be as low as 60% of the suction recirculation values shown for continuous operation and as low as 25% for intermittent operation.

Figure 3: Suction-specific speed S at best efficiency flow, single suction or one side of double suction (to obtain S in SI units, multiply by 0.6123)

Figure 4: Suction-specific speed S at best efficiency flow, single suction or one side of double suction (to obtain S in SI units, multiply by 0.6123)



The high turbulence produced by recirculation and separation accounts for most of the high power consumed at shut-off. This may vary from about 30% of the normal power for pumps of very low-specific speed to nearly three times the normal power for propeller pumps. Separation and, possibly, cavitation may take place on the casing tongue or diffusion vanes at very low capacities. Operation near shut-off causes not only excessive heating, but also vibration and cavitation, which may cause serious mechanical damage. Diagnosis of Suction and Discharge Recirculation Cause and Effect: Recirculation occurs at reduced flows and is the reversal of a portion of the flow-back through the impeller. Recirculation at the inlet of the impeller is known as suction recirculation. Recirculation at the outlet of the impeller is discharge recirculation. Suction and discharge recirculation can be very damaging to pump operation and should be avoided for continuous operation. Diagnosis From Pump Operation: Suction recirculation will produce a loud crackling noise in and around the suction of the pump. Recirculation noise is of greater intensity than the noise from low NPSH cavitation and is a random knocking sound. Discharge recirculation will produce the same characteristic sound as suction recirculation except that the highest intensity is in the discharge volute or diffuser. Diagnosis From Visual Examination: Suction and discharge recirculation produce cavitation damage to the pressure side of the impeller vanes. Viewed from the suction of the impeller, the pressure side would be the invisible, or underside, of the vane. Figure 5 Page 3 shows how a mirror can be used to examine the pressure side of the inlet vane for cavitation damage from suction recirculation. Damage to the pressure side of the vane from discharge recirculation is shown in Figure 6 Page 3. Guide vanes in the suction may show cavitation damage from impingement of the back-flow from the impeller eye during suction recirculation. Similarly, the tongue or diffuser vanes may show cavitation damage on the impeller side from operation in discharge recirculation.

Figure 5: Examining the pressure side of the inlet vanes for suction recirculation.

Figure 6: Damage to the pressure side of the vane from discharge recirculation.



Corrective Procedures: Every impeller design has specific recirculation characteristics. These characteristics are inherent in the design and cannot be changed without modifying the design. An analysis of the symptoms associated with recirculation should consider the following as possible corrective procedures.

1. Increase the output capacity of the pump.

2. Install a bypass between the discharge and the suction of the pump.

3. Bleed air into the suction of the pump to reduce the intensity of the noise, vibration and cavitation damage.

4. Substitute a harder material for the impeller to reduce the rate of cavitation damage.


Pump Clinic 4 Viscosity-How does it affect centrifugal pump performance 27/02/08 Page 1 of 4

PUMP CLINIC 4

VISCOSITY IMPACT ON CENTRIFUGAL PUMP PERFORMANCE

The performance of centrifugal pumps will vary when viscous liquids are pumped. For medium and high viscosities, the power requirement increases considerably, whilst the head, and to a lesser extent the flowrate, is reduced. With the aid of Table 4 (Page 3), the characteristics of centrifugal pumps pumping viscous liquids can be calculated, providing the characteristics for pumping water are known. (This diagram may also be used as an aid in the selection of a pump for required duty). The correction factors established from the diagram are sufficiently accurate for general application within the limits given below. If more accurate values are required, a test should be performed with the particular liquid. When pumping highly-viscous liquids, it is recommended that the running costs are investigated to establish whether other types of pump (eg rotary positive-displacement-type) could be more economic due to the steep drop in efficiency of centrifugal pumps under these conditions.. The limits of centrifugal pumps are:

For nominal discharge pipe diameters; ≤ 50mm - approx 120 – 130 mm²/s ≤ 150mm - approx 300 – 500 mm²/s > 150mm - approx 800 mm²/s

Limitations and notes on the use of Table 4

∗ The diagram should only be used for centrifugal pumps with radial impellers within the normal Q-H range. The diagram must not be used for pumps with mixed flow and axial flow impellers or for special pumps for viscous or heterogeneous liquids. Table 4 is not applicable to side-channel pumps.

∗ The diagram may only be used if sufficient NPSH (NPSHavail ) is available to prevent cavitation.

∗ The diagram may only be used for homogeneous Newtonian fluids. For gelatinous liquids, widely-scattering results are obtained in practice, depending upon the special properties of the liquid.

∗ With multistage pumps, the head per stage must be used in the calculation.

∗ When pumps have double-entry impellers, one half of the flowrate must be used in the calculations.



Determining the size of a pump for a viscous liquid (Approximate determination of an equivalent operating point for water)

Subscripts: vis = viscous liquid w = water opt = best efficiency point Given: Qvis in m³/h, kinematic viscosity v in mm²/s Hvis in m, ρvis in kg/dm³ Required: Qw in m³/h } to determine a suitable pump for which only Hw in m } performance data relating to water are known

Pvis in kW to determine the driver power

To establish the correction factors from the diagram, the following procedure is used: Starting with the flowrate Q on the horizontal axis, move vertically up to the intersection with the required head H, then proceed horizontally (to the right-hand side or to the left-hand side) to the intersection with the viscosity v of the liquid, thence vertically up to the intersection with the lines of the different correction factors. To determine the correction factory CH for total head, the curve l.0 x Qopt is to be used.

This gives: Qw ≈ Qvis , Hw ≈ Hvis , �vis , ≈ C� x �w

CQ CH Example: Qvis = 100m³ v = 100mm²/s Hvis = 29.5m ρvis = 0.90kg/dm²

From the diagram the correction factors are found as follows:

CH = 0.94 CQ = 0.98 C� = 0.70

With this data the water values can be calculated:

Qw ≈ 100m³/h ≈ 102m³/h, Hw ≈ 29.5m = 31.4 0.98 0.94

A pump with �w = 75% is used. Therefore, �vis = 0.70 · 75% = 53% ρvis ≈ Qvis · Hvis · ρvis = 100 · 29.5 · 0.90 kW = 13.6 kW 367 · �vis 367 · 0.53



This procedure is to be considered as an approximation only, as the numerical values for rate of flow and total head shown in the diagram apply to water. However, in most cases, this procedure has sufficient accuracy for preliminary pump selections.

If the flowrate is: Qw < 0.9 x Qopt or > 1.1 x Qopt

Respectively, the selection should be checked by the more accurate method given in the following section.


Pump Clinic 5 Centrifugal or Positive Displacement 04/05/06 Page 1 of 3

PUMP CLINIC 5

CENTRIFUGAL OR POSITIVE DISPLACEMENT Which pump to choose?

Firstly, let us take a look at the two classifications in question and define both classes discussing the merits of individual types of pumps in each class.

1. Rotodynamic (which the main sub-classification is centrifugal)

2. Positive displacement

Rotodynamic Rotodynamic pumps are rotary machines in which energy is continuously imparted to the pumped liquid by a rotating impeller, propeller or rotor.

For this discussion we will not consider the special category as they are rarely used and only under very specific conditions. Before going on to review centrifugal pumps which account for probably well over 95% of Rotodynamic applications, when do we use a peripheral pump (sometimes called side-channel or regenerative turbine pumps)? They are most definitely not suited to handling solids because for efficient operation, they depend on close clearances between their impellers and guide plates which also limits their viscosity-handling capabilities to under 20mm²/sec. However, they are ideal for low capacities limited to 10 l/sec at quite high heads up to 310 metres through multi-staging, plus they have a built-in self-priming capability. Finally, many peripheral flow pumps have the ability to handle quantities of vapour mixed with liquid for substantial periods. Back to what we would all consider true centrifugal pumps, we can, for the purpose of this discussion, consider the following classes:

1. Closed impeller 2. Open Impellers 3. Slurry pumps



In considering these classes we must look at how they handle solids and viscous liquids. Closed impeller pumps below 80mm for example, should not be used for liquids of viscosity greater than, say, 50mm²/sec because the viscous nature of the liquid creates too many internal losses to operate efficiently, and likewise ‘lightly muddy’ water is about the worst solids they can handle. However, the larger the centrifugal pump, the higher the viscosity it can efficiently pump, such that at over 150mm for example, it can handle up to 800mm²/sec. Open impellers can handle solids up to 5% provided that the individual solids can fit through the impeller passage-ways and remembering that high velocities within a centrifugal pump encourage high abrasive wear, if that is a characteristic of the solids. In these cases, a positive displacement pump with its lower internal speeds could be a more economic option as appropriate materials must be selected to accommodate the wear and this can be expensive. For serious solid volumes, firstly these can be handled in sewage style pumps with single and two-vane impellers that have passages able to pass solids the size of the suction connection, generally beginning at 80mm. Then we come to true slurry centrifugal pumps which can handle high volume solids, say to 50-60% generally at under 100 metres heads. These are built with large clearances with internal adjustment for wear, plus wear plates and components which are readily replaceable, including rubber-lined parts. Positive Displacement Positive displacement pumps are rotary or reciprocating machines in which energy is periodically added by application of force to movable boundaries of enclosed fluid containing volumes, resulting in a direct increase in pressure.

All positive displacement pumps can handle viscous liquids generally to very high viscosities and most are capable of handling substantial solids with the exception of vane, gear, multiple-screw and some forms of lobe pumps. For clean liquids of low viscosity, again many positive-displacement pumps can handle these liquids. However, many rotary types do not do it economically because of slippage of thin liquids (low viscosities) through their clearances i.e. gear, lobe and multiple-screw pumps.



For positive displacement pumps it should be remembered that in sizes up to say 50mm discharge, the capital costs generally are similar to centrifugal pumps. However, after that, the positive displacement pump rapidly increases in cost, such that a 150mm discharge-type pump can cost many times that of a standard centrifugal water pump. Conclusion The above supports, in general, the basic conclusion that:

1. Centrifugal pumps are for low viscous clean fluids

2. Positive displacement pumps are for slurries and viscous liquids

However, there are some important instances which do not follow these basic conclusions:

a) Slurry applications, e.g. 80mm discharge and above and generally below 100m head should be centrifugal.

b) For clean liquid duties below 3 l/sec and above, 180 metres total head positive displacement pumps should be considered.

c) For viscous liquid applications with up to 800mm²/sec viscosity and capacities above 70 l/sec a centrifugal pump should be considered.

d) Generally, raw sewage applications should use centrifugal pumps

e) All applications below 0.5 l/sec should be positive displacement pumps

To conclude, every pump application should be individually considered as to the type of pump most suitable. If in doubt, consult your pump supplier.

Kelair Pumps Australia Pty Ltd ABN 28 001 308 381 215 Walters Road Arndell Park NSW 2148 Ph: 1300 789 466 Fax: 02 9678 9455 Email: [email protected] www.kelairpumps.com.au QLD Fax: 07 3808 8758 VIC Fax: 03 9569 7866 TAS Fax: 03 6331 9102 WA Fax: 08 9248 2255

PUMPS STEAM TURBINES BUILDING & FIRE WASTEWATER SERVICE

Pump Clinic 6 Speed Variation with Centrifugal Pumps 08/06/06 Page 1 of 11

PUMP CLINIC 6

SPEED VARIATION WITH CENTRIFUGAL PUMPS

Affinity Laws There are fundamental laws which can be used to predict changes in pump performance with variations in speed. It is important in pump applications to be able to develop performance curves corresponding to various speeds from standard performance curves. The mathematical relationships between flowrate, head, power and speed which enable this are known as the Affinity Laws. For variation in speed with constant impeller diameter, the following laws apply:

a) Pump flow rate (Q) varies directly with the speed (N) i.e. Q1/Q2 = N1/N2

b) Pump head (H) varies with the square of the speed (N) i.e. H1/H2 = (N1/N2)²

c) Power absorbed varies with the cube of the speed (N) i.e. P1/P2 = (N1/N2)³

In using the above formulae, it is assumed that efficiency remains constant. In practice, the efficiency if slightly less at lower speeds since friction and drag constitute a larger proportion of hydraulic power. It is important to note that these laws do not apply to NPSH. Example: If you have a pump performance at a speed of 1300rpm, what is the performance at 880rpm (refer to following performance curve Page 2). The first step is to select 5 operating points on the 1300rpm performance and tabulate as shown below: Flow (m³/hr) 0 800 1500 2200 2800Head (m) 75.5 75 73 67 56Efficiency - - 77 87 84kW * 230 313 386 460 507 * kW may be read from the power curve where efficiencies are not detailed or may be calculated using the formula.

kW = Flow x Head x SG k x Efficiency Where: Flow is in l/sec or m³/hr Head is in metres SG is specific gravity Efficiency is to a decimal point i.e. 84% efficiency becomes 0.84 k = 102.2 if flow is in l/sec = 368 if flow is in m³/hr




Using the Affinity Laws, the following factors can be applied to the 1300 rpm performance figures when operating at 880 rpm:

Flow - (880/1300) i.e. 0.677 - (880/1300)² i.e. 0.458 - (880/1300)³ i.e. 0.310

The new performance for the pump at 880 rpm is tabulated below: Flow (m³/hr) 0 542 1015 1490 1896 Head (m) 34.6 34.4 33.4 30.7 25.7 Efficiency - - 77 87 84 kW 71.3 97 119.7 142.6 157




Pumps should be selected for maximum flow The pump and the associated plant equipment such as pipes, valves and tanks must always be designed to cover for the maximum pumped volume. The following must be taken into consideration to determine the maximum capacity of the plant:

· Provision for increasing demand · Excess demand for pumping capacity in exceptional circumstances, eg when the

tanks are being emptied or refilled. · In the event of emergencies, such as fire, heavy rainfall etc.

Forms of Control Since pumps are selected for the maximum plant capacity, a form of control must be provided to regulate the volume of flow for variation in pumping demands. An average pumped quantity Qm may be only a fraction of the maximum pump capacity Qp. The duration curve in Figure 1.2 below illustrates for example, how for the most part of over a one-year period, the pump may operate at reduced capacity.




Speed control is more economical than other forms of control The pump flow can be controlled by using the following control methods:

· Throttle (or choke) control by means of a valve · Start-Stop control of the pump · Speed control eg, by means of a frequency converter

Throttle control is, even today, the most commonly used control form in industrial applications. Its efficiency is, however, very low when compared with speed control, which in many cases gives more than a 50% saving in energy. Pumps at waterworks and sewage water treatment plants are normally controlled by means of start-stop control. Its efficiency is often also poor (Figure 1.3) and besides, stress due to frequent starting and stopping may cause damage to the pipes and other plant equipment.

Energy Efficient Throttle control means that the flow of liquid in the pipes is restricted by means of a valve. This results in a waste of energy because the pump is continuously working against the high pressure imposed by the valve. The power consumed by the pump can be calculated from the formula:

P = Q x H x ρ 368 x η Where : P = Power (kW) Q = Pumped quantity (m³/hr) H = Pump head (m) ρ = Specific gravity of the liquid η = Pump efficiency




The above formula shows that the power requirement P is directly dependent on the product of the pumped quantity Q and the pump head H. Figures 2.1 and 2.2 illustrate the power requirements which are represented by the hatched areas in both figures. It can be seen that in this example, the power requirement with speed control is less than half of that with throttle control. The saving obtained in energy depends essentially on the average pumped quantity. Figure 3 shows how much energy saving can be at different pumped volumes. When the power saving is known, the saving in energy can always be calculated by multiplying the power saving by the time factor. The methods of calculation are described in more detail (Page 1).




Fewer pumps required Control of flow is often arranged by means of two or more parallel-connected pumps which are of different sizes. Step by step control is thus achieved by running the pumps in turn. Improved control with a lesser capital cost is achieved if a single large pump is provided with the control as shown in Figure 4. Parallel-connected pumps and motors require additional valves and piping which again, increases costs.

Reduced need for tanks Pressure tanks and upper water tanks are used for keeping a uniform pressure in the pipes in applications where the pump runs on intermittent duty as, for example, in waterworks. If the pump is provided with a frequency converter, the tanks can be made smaller or may be totally dispensed with. In addition to the lower investment costs, a better control result is achieved, which means a more uniform pressure at the consumer end.




Savings in electrical distribution costs The starting current which a pump provided with a frequency converter takes from the electrical supply line is but a fraction of the starting current required by direct starting. On account of this, the electrical distribution equipment can be made smaller and be purchased at a lower price. A typical objective for saving purposes may be a standby generator for critical pumps. When a frequency converter is used for the speed control of the pump, the generator size need only be 30 to 50% of that previously required.

Compensating capacitors can be dispensed with Squirrel cage motors need reactive power which somehow needs to be generated. To avoid loading the distribution network unnecessarily with reactive power, compensation is normally effected by means of capacitors near to the motor. The frequency converter generates the reactive power required by the motor and no compensating capacitors are needed. The cost of investment is reduced and an optimum compensation effect achieved.




Improved control effect An improved control effect is more easily achieved with speed control than with other non-linear forms of control. A drawback of intermittent duty, for example, is the discontinuity of the control. The controlled parameters, the flow or pressure, for example, keeps varying. An accurate and linear control is achieved with the converter. Figure 8 shows the control graph of a plant with three parallel-connected pumps P1, P2 and P3. When one of the pumps (P1) is provided with the control, a linear control curve (R) is obtained, whilst the curve (T) of intermittent control is stepped, which can lead to abrupt variations in the pumped volume of liquid.

Reduced maintenance costs When the control is used, the pump, pipes and valves experience less wear, which means increased service life and reduced maintenance (particularly with plastic pipes). · Static stress is reduced because the system need not operate with a high pumping pressure

all the time as with choke control. The pressure is as high or low as required.

· Dynamic stresses are far lower with a smooth control than with an intermittent start-stop control. Pressure strokes (Figure 9) which wear the pipes and other plant equipment can thus be avoided, and the service life may even be doubled.




Calculate the savings yourself 1. Power requirement is determined:

(a) Either by means of the pump characteristics The power requirement with choke control P1 and with speed control P2 can best be determined by means of the pump characteristic curves, provided that complete curves are available. P1 and P2 are determined according to Figure 10. When QM and the system characteristics are known, the pump curves η1 and η2 as well as the corresponding power curves P1 and P2 are found. The power requirements can be read at the point where QM and the power curves intersect.

(b) Or by means of calculation If the complete power characteristics are not available, P1 and P2 are calculated from the following formulae. Choke control P1:

P1 = QM[m³/h] x H1 [m] x ρ kW

368 x η1 Speed control P2:

P2 = QM[m³/h] x H2 [m] x ρ kW 368 x η2

2. Calculate the power saving The power saving obtained by means of control is:

PS = P1 – P2 0.9

Where the divisor 0.9 is the approximate efficiency of the motor.

3. Calculate the energy saving The saving in energy per year is obtained when the power saving is multiplied by the operating hours, that is:

WS = PS x ta = P1 – P2 x ta 0.9

4. Saving in money The saving in money per year is obtained when the energy saving is multiplied by the unit price of energy k.

Ks [money saved per annum] = WS [kWh/a] - k [price/kWh]

5. Cost pay-off time The cost pay-off time is obtained by comparing the cost difference Kp between the speed control and the choke control with the achieved saving per year Ks.

tt = Kp [cost]

Ks [money saved per annum]




Example The pump (as per following curve Page 11) is designed for a nominal duty of 2200m³/hr at 67.5 metres head. If a secondary flowrate of 1500m³/h is required for 50% of the time, what are the comparative power costs for throttling versus speed control.

By drawing system characteristics for the pump curves, we obtain:

H1 = 73m η1 = 0.765 (throttling) H2 = 31m η2 = 0.87 (speed control to 880 rpm)

Power with throttling control:

P1 = 1500 x 73 x 1.0 = 389kW 368 x 0.765

Power with speed control:

P2 = 1500 x 31 x 1.0 = 145kW 368 x 0.87

Power saving:

PS = 389 - 145 = 271kW 0.9




Energy Saving: (assuming 8000 hrs/annum total operating time)

WS = 271kW x 4000 hr/a = 1084000 kW/a

The saving in money and the cost pay-off time can be calculated by inserting the right values for the unit price of energy k and the capital cost difference Kp in the following formulae: Saving in money: (based on power cost of 8¢ / kW hr)

Ks = 1084000 kW hr/a x 8¢ [cost/kW hr] = $86720 per annum

Cost pay-off time:

tt = Kp [cost] Ks [money saved per annum]



Pump Clinic 7 Installation and Piping 29/06/06 Page 1 of 11

PUMP CLINIC 7

INSTALLATION AND PIPING

INSTALLATION Instruction Books Instruction books are intended to help keep the pumps in an efficient and reliable condition at all times. It is necessary, therefore, that instruction books be available to all personnel involved in this function. Preparation for Shipment After a pump has been assembled in the manufacturer’s shop, all flanges and exposed machined metal surfaces are cleaned of foreign matter and treated with an anti-corrosion compound, such as grease, Vaseline, or heavy oil. For protection during shipment and erection, all pipe flanges, pipe openings and nozzles are protected by wooden flange covers or by screwed-in plugs. Usually the driver is delivered to the pump manufacturer, where it is assembled and aligned with the pump on a common baseplate. The baseplate is drilled for driver mounting, but the final dowelling is performed in the field after final alignment. When size and weight permit, the unit is shipped assembled with pump and driver on the baseplate. If drivers are shipped directly to be mounted in the field, the baseplate should be drilled at the job site. Care of Equipment in the Field If the pumping equipment is received before it can be used, it should be stored in a dry location. The protective flange covers and coatings should remain on the pumps. The bearings and couplings must be carefully protected against sand, grit and other foreign matter. More thorough precautions are required if a pump must be stored for an extended period of time. It should be carefully dried internally with hot air or by a vacuum-producing device to avoid rusting of internal parts. Once free of moisture, the pump internals should be coated with protective liquid such as light oil, kerosene or antifreeze. Preferably, all accessible parts, such as bearings and couplings, should be dismantled, dried and coated with Vaseline or acid-free heavy oil and then properly tagged and stored. If rust preventive has been used on stored parts, it should be removed completely before final installation and the bearings should be relubricated. Pump Location Working space must be checked to assure adequate accessibility for maintenance. Axially-split casing horizontal pumps require sufficient headroom to lift the upper half of the casing free of the rotor. The inner assembly of radially split multistage centrifugal pumps is removed axially. Space must be provided so that the assembly can be pulled out without canting it. For large pumps with heaving casings and rotors, a travelling crane or other facility for attaching a hoist should be provided over the pump location. Pumps should be located as close as practicable to the source of liquid supply. Wherever possible, the pump centreline should be placed below the level of the liquid in the section reservoir.

(Figure 1 - Foundation bolt)




Foundations Foundations may consist of any structure heavy enough to afford permanent rigid support to the full area of the baseplate and to absorb any normal strains or shocks. Concrete foundations built up from solid ground are the most satisfactory. Although most pumping units are mounted directly on baseplates, very large equipment may be mounted directly on the foundations by using sole plates under the pump and driver feet. Misalignment is corrected with shims. The space required by the pumping unit and the location of the foundation bolts is determined from the drawings supplied by the manufacturer. Each foundation bolt should be surrounded by a pipe sleeve three or four diameters larger than the bolt. After the concrete foundations are poured, the pipe is held solidly in place but the bolt may be moved to conform to the corresponding hole in the baseplate. When a unit is mounted on steelwork or some other structure, it should be placed directly over, or as near as possible to, the main members, beams or walls and should be supported so that the baseplate cannot be distorted or the alignment disturbed by any yielding or springing of the structure, or of the baseplate. Mounting of vertical wet-pit pumps A curve ring or soleplate must be used as a bearing surface for the support flange of a vertical wet-pit pump. The mounting face must be machined because the curb ring or soleplate will be used in aligning the pump. If the discharge pipe is located below the support flange of the pump (belowground discharge), the curb ring or soleplate must be large enough to pass the discharge elbow during assembly. A rectangular ring should be used. If the discharge pipe is located above the support flange (aboveground discharge), a round curb ring or soleplate should be provided with clearance on its inner diameter to pass all sections of the pump below the support flange. A typical method of arranging a grouted soleplate for vertical pumps is shown below. If the discharge is belowground and an expansion joint is used, it is necessary to determine the movement that may be imposed on the structure. The pump casing should be attached securely to some rigid structural members with tie rods. If vertical wet-pit pumps are very long, some steadying device is required irrespective of the location of the discharge or of the type of pipe connection. Tie rods can be used to connect the unit to a wall, or a small clearance around a flange can be used to prevent excessive displacement of the pump in the horizontal plane. Alignment When a complete unit is assembled at the factory, the baseplate is placed on a flat, even surface. The pump and driver are mounted on the baseplate and the coupling halves are accurately aligned using shims under the pump and driver mounting surfaces where necessary. The pump is usually dowelled to the baseplate at the factory, but the driver is left to be dowelled after installation at the site. The unit should be supported over the foundation by short strips of steel plate or shim stock close to the foundation bolts, allowing a space of ¾ to 2 inches (2 to 5 cm) between the bottom of the baseplate and the top of the foundation for grouting. The shim stock should extend fully across the supporting edge of the baseplate. The coupling bolts should be removed before the unit is levelled and the coupling halves are aligned. Where possible, it is preferable to place the level on some exposed part of the pump shaft, sleeve, or planed surface of the pump casing. The steel supporting strips or shim stock under the baseplate should be adjusted until the pump shaft is level, the suction and discharge flanges are vertical or horizontal as required, and the pump is at the specified height and location. When the baseplate has been levelled, the nuts on the foundation bolts should be made handtight.




During this levelling operation, accurate alignment of the unbolted coupling halves must be maintained. A straightedge should be placed across the top and sides of the coupling, and at the same time, the faces of the coupling halves should be checked with a tapered thickness gauge or with feeler gauges to see that they are parallel. For all alignment checks, including parallelism of the coupling faces, both shafts should be pressed over to one side when taking readings. When the peripheries of the coupling halves are true circles of equal diameter and the faces are flat and perpendicular to the shaft axes, exact alignment exists when the distance between the faces is the same at all points and when a straightedge lies squarely across the rims at any point. If the faces are not parallel, the thickness gauge or feelers will show variation at different points. If one coupling is higher than the other, the amount may be determined by the straightedge and feeler gauges. Sometimes coupling halves are not true circles or are not of identical diameter because of manufacturing tolerances. To check the trueness of either coupling half, rotate it while holding the other coupling half stationary and check the alignment at each quarter turn of the half being rotated. Then the half previously held stationary should be revolved and the alignment checked. A variation within manufacturing limits may be found in either of the half-couplings and proper allowance for this must be made when aligning the unit. A more exact method for checking alignment that is recommended requires the use of a dial indicator. With the indicator bolted to the pump half of the coupling, both radial and axial alignment can be checked. This method is called face-and-rim alignment. With the button resting on the periphery of the other coupling half, the dial should be set at zero and a mark chalked on the coupling half at the point where the button rests. For any check (top, bottom or sides) both shafts should be rotated the same amount, that is all readings on the dial should be made with the button on the chalk mark. The dial readings will indicate whether the driver must be raised, lowered, or moved to either side. After any movement, it is necessary to check that the coupling faces remain parallel to one another. For example, if the dial reading at the starting point is set to zero and the diametrically opposite reading at the bottom or sides shows + 0.202 in (+ 0.508 mm) the driver must be raised or lowered by shimming or moved to one side or the other by half of this reading. The same procedure is used to align gear couplings but the coupling covers must first be moved back out of the way and all measurements should be made on the coupling hubs.




When an extension coupling connects the pump to its driver, a dial indicator should be used to check the alignment. The extension piece between the coupling halves should be removed to expose the coupling hubs. The coupling nut on the end of the shaft should be used to clamp a suitable extension arm or bracket long enough to extend across the space between the coupling hubs. The dial indicator is mounted on this arm, and alignment is checked for both concentricity of the hub diameters and parallelism of the hub faces. Changing the arm from one hub to the other provides an additional check. The dial extension bracket must be checked for sag, and readings must be corrected accordingly.

The clearance between the faces of the coupling hubs and the ends of the shafts should be such that they cannot touch, rub, or exert a pull on either pump or driver. The amount of this clearance may vary with the size and type of coupling used. Sufficient clearance will allow unhampered endwise movement of the shaft of the driving element to the limit of its bearing clearance. On motor-driven units, the magnetic centre of the motor will determine the running position of the motor half-coupling. This position should be checked by running the motor uncoupled. This will also permit checking the directon of rotation of the motor. If current is not available at the time of installation, move the motor shaft in both directions as far as the bearings will permit and adjust the shaft centrally between these limits. The unit should then be assembled with the correct gap between the coupling halves. Large horizontal sleeve-bearing motors are not generally equipped with thrust bearings. The motor rotor is permitted to float, and as it will seek its magnetic centre, an axial force of rather small magnitude can cause it to move off this centre. Sometimes it will move enough to cause the shaft collar to contact and possibly damage the bearing. To avoid this, a limited-end float coupling is used between the pump and the motor on all large units to restrict the motor rotor. The setting of axial clearances for such units should be given by the manufacturer in the instruction books and elevation drawings. When the pump handles a liquid at other than ambient temperature or when it is driven by a steam turbine, the expansion of the pump or turbine at operating temperature will alter the vertical alignment. Alignment should be made at ambient temperature with suitable allowances for the changes in pump and driver centrelines after expansion. The final alignment must be made with the pump and driver at their normal temperatures and adjusted as required before the pump is placed into permanent service. For large installations, particularly with steam-turbine-driven pumps, more sophisticated alignment methods are sometimes employed using proximity probes and optical instruments. Such procedures permit checking the effect of temperature changes and machine strains caused by piping stresses while the unit is in operation. When such procedures are recommended, they are included with the manufacturers’ instructions.




When the unit has been accurately levelled and aligned, the hold-down bolts should be gently and evenly tightened before grouting. The alignment must be rechecked after the suction and discharge piping has been bolted to the pump to test the effect of piping strains. This can be done by loosening the bolts and reading the movement of the pump, if any, with dial indicators. The pump and driver alignment should be occasionally rechecked, for misalignment may develop from piping strains after a unit has been operating for some time. This is especially true when the pump handles hot liquids as there may be a growth or change in the shape of the piping. Pipe flanges at the pump should be disconnected after a period of operation to check the effect of the expansion of the piping and the adjustments should be made to compensate for this. Grouting Ordinarily, the baseplate is grouted before the piping connections are made and before the alignment of the coupling halves is finally rechecked. The purpose of grouting is to prevent lateral shifting of the baseplate to increase its mass to reduce vibration and to fill in irregularities in the foundation. The usual mixture for grouting a pump baseplate is composed of one part pure Portland cement and two parts building sand, with sufficient water to cause the mixture to flow freely under the base (heavy cream consistency). To reduce settling, it is best to mix the grout and let it stand for a short period then remix it thoroughly before use without adding any more water. The top of the rough concrete foundation should be well saturated with water before grouting. A wooden form is built around the outside of the baseplate to retain the grout. Grout is added until the entire space under the baseplate is filled to the top of the underside. The grout holes in the baseplate serve as vents to allow the air to escape. A stiff wire should be used through the grout holes to work the grout and release any air pockets.

The exposed surfaces of the grout should be covered with wet burlap to prevent cracking from too-rapid drying. When the grout is sufficiently set so that the forms can be removed, the exposed surfaces of the grout and foundations are finished smooth. When the grout is hard (72h or longer), the hold-down bolts should be finally tightened and the coupling halves rechecked for alignment. There is considerable controversy over whether the levelling strips or wedges should be removed after grouting. The best practice is to remove these in all cases for reciprocating machinery because pounding action or vibration will ultimately loosen the unit from the foundation. The space formerly occupied by shims or wedges must be re-grouted. There is less danger is not removing the strips or wedges with rotating machinery, provided care is used in mixing the grout material and there is no shrinkage or drying. The strips or wedges can also be removed from a rotating unit. Erectors can follow their own preference in this matter. The pump and driver alignment must be rechecked thoroughly after the grout has hardened permanently and at reasonable intervals thereafter.




Dowelling of pump and driver When the pump handles hot liquids, dowelling of both the pump and its driver should be delayed until the unit has been operated. A final re-check of alignment with the coupling bolts removed and with the pump and driver at operating temperature is advisable before dowelling. Large pumps handling hot liquids are usually dowelled near the coupling end, allowing the pump to expand from that end out. Sometimes the other end is provided with a key and a keyway in the casing foot and the baseplate. PIPING Suction piping The suction piping should be as direct and short as possible. If a long suction line is required the pipe size should be increased to reduce frictional losses. (The exception to this recommendation is in the case of boiler-feed pumps where difficulties may arise during transient conditions of load change if the suction piping volume is excessive. This is a special and complex subject and the manufacturer should be consulted). Where the pump must lift the liquid from a lower level, the suction piping should be laid out with a continual rise toward the pump avoiding high spots in the line to prevent the formation of air pockets. Where a static suction head will exist, the pump suction piping should slope continuously downward to the pump. Flow velocity should be less than 2m per second and the pipe diameter should be calculated accordingly. Generally, the suction line is larger than the pump suction nozzle and eccentric reducers should be used. If the source of supply is above the pump, the straight side of the reducer should be at the bottom. Installing eccentric reducers with a change in diameters greater than 4 inches (10 cm) could disturb the suction flow. If such a change is necessary, it is advisable to use properly vented concentric reducers.




Elbows and other fittings next to the pump suction should be carefully arranged, or the flow into the pump impeller will be disturbed. Long-radius elbows are preferred for suction lines because they create less friction and provide a more uniform flow distribution than standard elbows. Pumps should not be connected near any fitting. There should be a straight length of pipe minimum 10 x pipe diameter before the pump inlet. If a common suction head for two or more pumps is used, branches should be designed so that flow disturbance before the pump inlets is not caused. T-pieces should not be used - a Y-piece is recommended.

Discharge Piping Generally, both a check valve and a gate valve are installed in the discharge line. The check valve is placed between the pump and the gate valve and protects the pump from reverse flow in the event of unexpected driver failure, or from reverse flow from another operating pump. The gate valve is used when priming the pump or when shutting it down for inspection and repairs. Manually operated valves that are difficult to reach should be fitted with a sprocket rim wheel and chain. In some cases, discharge gate valves are motorised and can be operated by remote control. Piping Strains Cast iron pumps are never provided with raised face flanges. If steel suction or discharge piping is used, the pipe flanges should be of the flat-face and not the raised-face type. Full-face gaskets must be used with cast iron pumps. Piping should not impose excessive forces and movements on the pump to which it is connected, since these might spring the pump or pull it out of position. Piping flanges must be brought squarely together before the bolts are tightened. The suction and discharge piping and all associated valves, strainers, etc should be supported and anchored near to, but independent of, the pump so that strain will be transmitted to the pump casing. There are four factors to be considered in determining the effect of nozzle loads; material stress in pump nozzles resulting from forces and bending movements, distortion of internal moving parts affecting critical clearances, stresses in pump hold-down bolts, and distortion in pump supports and baseplates resulting in driver coupling misalignment. With large pumps, or when major temperature changes are expected, the pump manufacturer generally indicates to the user the maximum movements and forces that can be imposed on the pump by the piping. Expansion Joints Expansion joints are sometimes used in the discharge and suction piping to avoid transmitting any piping strains caused by misalignment, or by expansion when hot liquids are handled. On occasion, expansion joints are formed by looping the pipe. More often, they are of the slip-joint or corrugated-




diaphragm-type. However, they transmit to the pump a force equal to the area of the expansion joint, times the pressure in the pipe. These forces can be of very significant magnitude, and it is impractical to design the pump casings, baseplates, etc to withstand them. Consequently, when expansion joints are used, a suitable pipe anchor must be installed between them and the pump proper. Alternately, tie rods can be used to prevent the forces from being transmitted to the pump. Suction Strainers Except for certain special designs, pumps are not intended to handle liquid containing foreign matter. If the particles are sufficiently large, such foreign matter can clog the pump, reduce its capacity, or even render it altogether incapable of pumping. Small particles of foreign matter may cause damage by lodging between close-running clearances. Therefore, proper suction strainers may be required in the suction lines of pumps not specially designed to handle foreign matter. In such an installation, the piping must first be thoroughly cleaned and flushed. The recommended practice is to flush all piping to waste before connecting it to the pump. Then a temporary strainer of appropriate size should be installed in the suction line as close to the pump as possible. This temporary strainer may have a finer mesh than the permanent strainer installed after the piping has been thoroughly cleaned of all possible mill scale or other foreign matter. The size of the mesh is generally recommended by the pump manufacturer. Venting and Draining Vent valves are generally installed at one or more high points of the pump casing waterways to provide a means of escape for air or vapour trapped in the casing. These valves are used during the priming of the pump or during operation if the pump should become air or vapour-bound. In most cases, these valves need not be piped up away from the pump because their use is infrequent and the vented air or vapours can be allowed to escape into the surrounding atmosphere. On the other hand, vents from pumps handling flammable, toxic or corrosive fluids must be connected in such a way that they endanger neither the operating personnel, nor the installation. The suction vents of pumps taking liquids from closed vessels under vacuum must be piped to the source of the pump suction above the liquid level. All drain and drip connections should be piped to a point where the leakage can be disposed of or collected for re-use if worth reclaiming. Warm-Up Piping When it is necessary for a pump to come up to operating temperature before it is started, or to keep it ready to start at rated temperature, provision should be made for a warm-up flow to pass through the pump. There are a number of arrangements used to accomplish this. If the pump is operated under positive pressure on the suction, the pumped liquid can be permitted to drain out through the pump casing drain connection to some point at a pressure lower than the suction pressure. Alternately, some liquid can be made to flow back from the discharge header through a jumper line around the check valve, into the pump and out into the suction header. An orifice is provided in this jumper line to regulate the amount of the warm-up flow. Care must be exercised in such an installation to maintain the suction valve open (unless the warm-up line valve is closed, as when the pump is to be dismantled), lest the entire pump, suction valve, and suction piping be subjected to full discharge pressure. The manufacturer’s recommendations should be sought in all cases as to the best means of providing an adequate warm-up procedure.




Relief Valves Positive displacement pumps, such as rotary and reciprocating pumps can develop discharge pressures much in excess of their maximum design pressures. To protect these pumps against excessive pressures when the discharge is throttled or shut off, a pressure relief valve must be used. Some pumps are provided with internal integral relief valves, but unless operation against a closed discharge is both infrequent and of very short duration, a relief valve with an external return connection must be used and the liquid from the relief valve must be piped back to the source of supply.




Surge Chambers Generally, centrifugal pumps do not require surge chambers in their suction or discharge piping. Reciprocating pumps may have a suction and discharge piping layout that does not require compensation for variations in the flow velocity in the piping system. In many cases however, reciprocating pump installations require surge chambers when the suction or discharge lines are of considerable length, when there is an appreciable static head on the discharge, when the liquid pumped is hot, or when it is desirable to smooth out variations in the discharge flow. The type, size and arrangement of the surge chamber should be chosen on the basis of the manufacturer’s recommendations. Instrumentation There are a number of instruments which are essential to maintaining a close check on the performance and condition of a pump. A compound pressure gauge should be connected to the suction of the pump and a pressure gauge should be connected to its discharge at the pressure taps which may be provided in the suction and discharge flanges. The gauges should be mounted in a convenient location so that they can be easily observed. In addition, it is advisable to provide a flow-metering device. Depending upon the importance of the installation, indicating meters may be supplemented by recording attachments. Whenever pumps incorporate various leak-off arrangements, such as a balancing device or pressure-reducing labyrinths, a check should be maintained on the quantity of these leak-offs by measuring orifices and differential gauges installed in the leak-off lines. Pumps operating in important or complex services, or operating completely unattended by remote control, may have additional instrumentation such as speed indicators, vibration monitors, and bearing or casing temperature indicators.

Kelair Pumps Australia Pty Ltd A Mace Group Company ABN 28 001 308 381 215 Walters Road Arndell Park NSW 2148

Call Direct: 1300 789 466 Facsimile: 02 9678 9455 Email: [email protected] www.kelairpumps.com.au QLD Fax: 07 3808 8758 VIC Fax: 03 9569 7866 TAS Fax: 03 6331 9102 WA Fax: 08 9248 2255

Pump Clinic 8 Startup & Operation of Centrifugal Pumps 29/06/06 Page 1 of 5

PUMP CLINIC 8

START-UP AND OPERATION OF CENTRIFUGAL PUMPS OPERATION Pumps are generally selected for a given capacity and total head when operating at rated speed. These characteristics are referred to as ‘rated conditions of service’ and, with few exceptions, represent those conditions at or near which the pump will operate the greatest part of the time. Positive displacement pumps can not operate at any greater flows than rated, except by increasing the speed; nor can they operate at lower flows except by reducing their operating speed or bypassing some of the flow back to the source of supply. On the other hand, centrifugal pumps can operate over a wide range of capacities, from near zero flow to well beyond the rated capacity. Because a centrifugal pump will always operate at the intersection of its head-capacity and system-head curves, the pump operating capacity may be altered either by throttling the pump discharge (hence altering the system-head curve, or by varying the pump speed (changing the pump head-capacity curve). This makes the centrifugal pump very flexible in a wide range of services and applications which require the pump to operate at capacities and heads differing considerably from the rated conditions. There are, however, some limitations imposed upon such operation by hydraulic, mechanical, or thermodynamic considerations. Operation of Centrifugal Pumps at Reduced Flows There are certain minimum operating flows which must be imposed on centrifugal pumps for either hydraulic or mechanical reasons. Four limiting factors must be considered:

· Radial thrust · Temperature rise · Internal recirculation · Shape of the power curve

For sustained operation, it is imperative to adhere to the minimum flow limits recommended by the pump manufacturer The thermodynamic problem that arises when a centrifugal pump is operated at extremely reduced flows is caused by the heating up of the liquid handled. The difference between the power consumed and the water power developed represents the power losses in the pump, except for a small amount lost in the pump bearings. These power losses are converted to heat and transferred to the liquid passing through the pump. If the pump were to operate against a completely closed valve, the power losses would be equal to the shutoff power and since there would be no flow through the pump, all this power would go into heating the small quantity of liquid contained in the pump casing. The pump casing would heat up, and a certain amount of heat would be dissipated by radiation and convection to the atmosphere.




However, because the temperature rise in the liquid pumped could be quite rapid, it is generally safer to ignore the dissipation of heat through radiation and the absorption of heat by the casing. Calculations for determining the temperature rise in the liquid are available from Kelair. The maximum permissible temperature rise in a centrifugal pump varies over a wide range, depending on the type of service and installation. The minimum capacity based on thermodynamic considerations is then established as that capacity at which the temperature rise is the maximum permitted. There are also hydraulic considerations which may affect the minimum flow at which a centrifugal pump can operate. In recent years, correlation has been developed between operation at low flows and the appearance of hydraulic pulsations both in the suction and in the discharge of centrifugal impellers. It has been proven that these pulsations are caused by the development of an internal recirculation at the inlet and discharge of an impeller at certain flows below the best efficiency capacity. The pump manufacturer’s recommendations on minimum flows dictated by these considerations should always be followed. The NPSHR curve becomes increasingly unstable at low flows. As a rule of thumb, do not operate pumps at flowrates lower than that equivalent to the left-hand end of the NPSHR curve. This rule has to be considered in conjunction with other issues detailed in this section. Priming With the exception of self-priming pumps, no centrifugal pump should ever be started until it is fully primed. That is, until it has been filled with the liquid pumped and all the air contained in the pump has been expelled. Reciprocating pumps of the piston or plunger type are, in principle, self-priming. However, if quick starting is required, priming connections should be piped to a supply above the pump. Positive displacement pumps of the rotating type, such as rotary or screw pumps, have clearances that allow the liquid in the pump to drain back to the suction. When pumping low viscosity liquids, the pump may completely dry out when it is idle. In such cases, a foot valve should be used to help keep the pump primed. Alternately, a vacuum device may be used to prime the pump. When handling liquids of higher viscosity, foot valves are usually not required because liquid is retained in the clearances and acts as a seal when the pump is restarted. However, before the initial start of a rotating positive displacement pump, some of the liquid to be pumped should be introduced through the discharge side of the pump to wet the rotating element. The various methods and arrangements used for priming pumps are available from Kelair. FINAL CHECKS BEFORE START-UP A few last-minute checks are recommended before a pump is placed into service for its initial start.

· Lubricate the bearings with fresh grease if the pump has been standing for a long time since delivery. With oil-lubricated bearings, fill the bearing housing with the correct quantity and quality of oil (as per manufacturer’s recommendations).

· Open any sealing and cooling liquid valves where applicable. A mechanical seal should not normally show a leak, but a soft-packed gland should drip during the start-up period and be adjusted as required.




· Check that the cooling/sealing liquid outlet pipes are not blocked, where applicable.

· Fill the pump casing by opening the suction valve, or through a filling pipe. Vent casings by opening the discharge valve, or ventcocks when fitted.

· Ensure the pump shaft turns freely by hand. If it does not, check for foreign matter in the pump, an over-tightened gland or pipe stresses causing the impeller to seize in the casing.

· Check the direction of rotation by running the motor for a few seconds.

It is recommended when checking the direction of rotation, that coupling halves are disconnected to eliminate the risk of loosening the impeller, which could cause damage to the pump.

Starting and Stopping Procedures The steps necessary to start a centrifugal pump depend upon its type and upon the service on which it is installed. For example, standby pumps are generally held ready for immediate starting. The suction and discharge gate valves are held open and reverse flow through the pump is prevented by the check valve in the discharge line. The methods followed in starting are greatly influenced by the shape of the power-capacity curve of the pump. High and medium head pumps (low and medium specific speeds) have power curves that rise from zero flow to the normal capacity condition. Such pumps should be started against a closed discharge valve to reduce the starting load on the driver. A check valve is equivalent to a closed valve for this purpose, as long as another pump is already on the line. The check valve will not lift until the pump being started comes up to a speed sufficient to generate a head high enough the lift the check valve from its seat. If a pump is started with a closed discharge valve, the recirculation bypass line must be open to prevent overheating. If a bypass line is not installed, start the pump with the discharge valve cracked open. Low head pumps (high specific speed) of the mixed flow and propeller type, have power curves that rise sharply with a reduction in capacity; they should be started with the discharge valve wide open against a check valve, if required to prevent backflow. Assuming that the pump in question is motor-driven, that its shutoff power does not exceed the safe motor power, and that it is to be started against a closed or cracked open gate valve, the starting procedure is as follows:

1. Prime the pump, opening the suction valve, closing the drains, etc to prepare the pump for operation.

2. Open the valve in the cooling water supply to the bearings, where applicable.

3. Open the valve in the cooling water supply if the stuffing boxes are water-cooled.

4. Open the valve in the sealing liquid supply if the pump is so fitted.

5. Open the warn-up valve of a pump handling hot liquids if the pump is not normally kept at operating temperature. When the pump is warmed up, close the valve.

6. Open the valve in the recirculating line if the pump should not be operated against dead shutoff (if fitted) or slightly crack open the discharge valve.

7. Start the motor.




8. Open the discharge valve slowly.

9. Observe the leakage from the stuffing boxes and adjust the sealing liquid valve for proper flow to ensure the lubrication of the packing. If the packing is new, do not tighten up on the gland immediately, but let the packing run in before reducing the leakage through the stuffing boxes.

10. Check the general mechanical operation of pump and motor eg; bearing temperature, noise vibration.

11. Close the valve in the recirculating line once there is sufficient flow through the pump to prevent overheating.

If the pump is to be started against a closed check valve with the discharge gate valve open, the steps are the same, except the discharge gate valve is opened prior to the motor being started. In certain cases, the cooling water to the bearings and the sealing water to the seal cages are provided by the pump. This, of course, eliminates the need for the steps listed for the cooling and sealing supply. Just as in starting a pump, the stopping procedure depends upon the type and service of the pump. Generally the steps followed to stop a pump which can operate against a closed gate valve are:

1) Open the valve in the recirculating line.

2) Stop the motor.

3) Close the gate valve.

4) Open the warm-up valve if the pump is to be kept at operating temperature.

5) Close the valve in the cooling water supply to the bearings and to the water-cooled stuffing boxes.

6) If the sealing liquid supply is not required while the pump is idle, close the valve in this supply line.

7) Close the suction valve, open the drain valves, etc as required by the particular installation or if the pump is to be opened up for inspection.

In general, the starting and stopping of steam-turbine-driven pumps require the same steps and sequence prescribed for a motor-driven pump. As a rule, steam turbines have various drains and seals which must be opened or closed before and after operation. Similarly, many turbines require warming up before starting. Finally, some turbines require turning gear operation if they are kept on the line ready to start up. The operator should therefore follow the steps outlined by the turbine manufacturer in starting and stopping the turbine. Most of the steps listed for starting and stopping centrifugal pumps are equally applicable to positive displacement pumps. There is, however, a notable exception and that is:

Never operate a positive displacement pump against a closed discharge. If the gate valve on the discharge must be closed, always start the pump with the recirculation bypass valve open.




Auxiliary Services on Standby Pumps Standby pumps are frequently started up from a remote location, and several methods of operation are available for the auxiliary services, such as the cooling water supply to the bearings or to water-cooled stuffing boxes:

a) A constant flow may be kept through the bearing jackets, seal plates, oil coolers and through the stuffing box lantern rings, whether the pump is running or on standby service.

b) The service connections may be opened automatically whenever the pump is started up eg via solenoid valves.

c) The service connections may be kept closed while the pump is idle, and the operator may be instructed to open them shortly after the pump has been put on line automatically.

The choice among these methods must be dictated by the specific circumstances surrounding each case. There are, however, certain cases where sealing liquid supply to the pump stuffing boxes must be maintained, whether the pump is running or not. This is the case when the pump handles a liquid which is corrosive to the packing or which may crystallise and deposit on the shaft sleeves. It is also the case when the sealing supply is used to prevent air infiltration into a pump when it is operating under a vacuum.

Restarting Motor-Driven Pumps after Power Failure Assuming that power failure will not cause the pump to go into reverse rotation, that is, that a check valve will protect the pump against reverse flow, there is generally no reason why the pump would be permitted to restart once current has been re-established. Whether the pump will start again automatically when power is restored will depend on the type of motor control logic used. Because pumps operating on a suction lift may lose their prime during the time that power is off, it is preferable to use starters with low load protection for such installations to prevent an automatic restart. This does not apply, of course, if the pumps are automatically primed, or if some protection device is incorporated so that the pump can not run unless it is primed.



Pump Clinic 9 Noise Estimation (Sound Pressure Levels) for General Pumping Equipment 14/12/06 Page 1 of 6

PUMP CLINIC 9

NOISE ESTIMATION (SOUND PRESSURE LEVELS) FOR GENERAL PUMPING EQUIPMENT

INTRODUCTION Maximum noise levels for pumping machinery are being increasingly specified by customers as part of the overall performance requirements. Alternatively, queries are being raised by users on the noise levels that can be expected from their existing pumps. These requirements are as a direct result of increasing awareness of the damage and nuisance to health of operators that can result in prolonged exposure to high noise levels which in extreme cases can lead to permanent loss of hearing. Additionally, there is the necessity to comply with statutory Occupational Health and Safety requirements.

IMPORTANT: It should be recognised that noise levels emitted from similar sized machines and drivers can vary between quite wide limits. For this reason, the noise levels that can be estimated from the use of this data should be used with caution and never form part of a contractual agreement. If guaranteed figures are required, consult your pump supplier.

There is difficulty sometimes in relating a sound pressure level (SPL) to a real practical situation. Additionally, the overall SPL is a function of high intensity sound producing vibrations over a wide range of frequencies audible to the human ear. The graph below indicates typical noise levels by octave band frequencies for various real situations:




Estimation of Overall Noise Level The information that follows may be used to determine the expected sound pressure levels (SPL) expressed in decibels (dB) for conventional single-stage and multi-stage centrifugal pumps. These cannot be used for solids-handling, vortex or vertical mixed flow and axial flow pumps. The values are based on test data and the following conditions:

a) Background noise 10dB (minimum) below all noise levels in each octave band.

b) Values are at a distance of 1 metre horizontally from pump surfaces and 1.5 metres above the floor.

c) Overall noise level in dBA (“A“ scale) and octave mid band frequencies are basis “C” scale. Table A below shows overall SPL for single, two-stage and three or more stage pumps based on best efficiency point (BEP) power consumption unit:

Table A - Overall Noise Level - DBA

Absorbed kW @ BEP Absorbed kW @ BEP

dBa 1 & 2 Stage Multistage (3 & over) dBA 1 & 2 stage Multistage

(3 & over) 67 1.8-2.2 81 42.7-57.5 104.5-134.368 2.2-2.8 82 57.5-70.9 134.3-164.169 2.8-3.5 83 70.9-85.6 164.1-208.970 3.5-4.5 84 85.6-108.2 208.9-268.071 4.5-5.7 85 108.2-141.8 268.0-335.072 5.7-7.0 86 141.8-186.5 335.0-418.073 7.0-9.0 16.4-20.9 87 186.5-261.0 418.0-537.074 9.0-11.2 20.9-26.1 88 261.0-373.0 537.0-670.075 11.2-14.2 26.1-33.6 89 373.0-522.0 670.0-840.076 14.2-17.9 33.6-41.0 90 522.0-700.0 840.0-1045.077 17.9-23.1 41.0-52.2 91 700.0-820.0 1045.0-1345.078 23.1-29.1 52.2-67.1 92 820.0-970.0 1345.0-1680.079 29.1-36.6 67.1-82.1 93 970.0-1120.0 1680.0-2100.080 36.6-42.7 82.1-104.5

When calculating absorbed kW, use the following specific gravity figures:

a) For S G < 1.0 use S G = 1.0 b) For S G > 1.0 use actual S G

The values given in Table A are valid for flow rates that fall within the range of 75 - 125% of BEP flow. For flow rates outside this range, a correction is required as detailed below: a) For flows in the ranges 62%-75% and 126%-136% of BEP, add 1 to the dBA values on Table A.

b) For flows in the ranges 50%-61% and 137%- 50% of BEP, add 2 to the dBA values in Table A.

c) For flows in the range 38%-49% of BEP, add 3 to the dBA values in Table A.

d) For flows in the range 25%-37% of BEP, add 4 to the dBA values in Table A.




Estimation of Noise Levels at Octave Mid-Band Frequencies To estimate the noise levels at mid-band frequencies, subtract the appropriate value as shown in Table B from the overall dBA level.

Table B

Octave Mid-Band Frequency Hz

Pump Type 31.5 63 125 250 500 1K 2K 4K 8K Single & 2-stage 4 6 4 3 4 5 6 9 12 Deduct

from dBA level, dB Multi-stage

(3 & over) 11 9 7 5 3 3 3 7 12

Combined Noise Levels for Pump and Driver Components When two or more sources produce the total combined noise level (mid-bands) may be obtained by a simple addition of dB values according to Graph II:




Noise Reduction at Various Distances To obtain noise levels (either dBA or any mid-band frequency) for distances greater than 1 metre from equipment, reduce the levels using the values shown in Graph III:

Calculation of Overall SPL from Known Mid-Band Frequency dB Values Noise at constant level of emittance does not have the same irritation at all frequencies. For a given noise level, the higher the frequency the more objectionable the noise becomes.




The impact of this can be seen by the A scale weighting factor shown on the example below:

Figure

Conversion of sound pressure levels at octave band mid frequencies to an overall SPL in dB(A) Octave Band mid frequency (Hertz) 31.5 63 125 250 500 1000 2000 4000 8000 1000

SPL in dB for frequency 84 82 84 85 84 83 82 79 76 -

A scale weighting factor dB -39 -26 -10 -9 -3 0 +1 +1 +1 +6

SPL dB(A) 45 56 74 76 81 83 83 80 77 -

dB correction for addition of SPLs Note 1 .5 2.0 2.0 1.7 0

SPL dB(A) Note 2 56.5 78 85 84.7 77

dB correction for addition of SPLs Note 1 0 3.0 0

SPL dB(A) Note 2 78 88 77

dB correction for addition of SPLs Note 1 0 0

SPL dB(A) Note 2 88 0

dB correction for addition of SPLs Note 1 0

SPL dB(A) Note 2

Note 1: Obtained from Graph II Note 2: Obtained by adding dB correction to higher of noise levels

Worked Example Find the maximum expected noise levels of a single stage pump at the following design conditions:

Flow 490 l/sec (1768m³/hr) Head 38 metres S G 0.94

For the above duty, the impeller diameter is 540mm and pump efficiency is 83% Best efficiency point (BEP) conditions are:

Flow 722 l/sec (2600M³/hr) Head 30.5 metres Efficiency 88% Using the formula below, power absorbed at BEP kW = l/sec x metres x S G 102.2 x efficiency = 722 x 30.5 x 1.0 102.2 x .88 = 244.88 kW




Note for S G value below 1, we use 1.0 as the S G.

Using Table A the expected overall SPL at a BEP power of 244.88kW is 87 dBA. A correction needs to be applied to this value as the duty flow rate (490 l/s) is 68% of the BEP flow rate. The correction is +1 dB. This will give an overall SPL at the duty point of 88 dBA. The chart below summarises the estimates for dB values at mid band frequencies utilising Table B:

Mid Band Frequency - Hz dBA 31.5 63 125 250 500 1K 2K 4K 8K Noise @ BEP 87 83 81 83 84 83 82 81 78 75 PL Correction +1 +1 +1 +1 +1 +1 +1 +1 +1 +1 Noise @ Cond. Pt.

dB 88 84 82 84 85 84 83 82 79 76

Assuming that the pump is driven by an 82 dBA turbine and a 91 dBA gear box, find the combined noise level of the equipment at 1 metre and also at a distance of 15 metres. Pump - 88 dBA Turbine - 82 dBA Gear - 91 dBA Using Graph II: 88-82 = 6 - Add 1 to 88 = 89 dBA 91-89 = 2 - Add 2 to 91 = 93 dBA 93 dBA Total System Noise @ 1 metre Using Graph III: For a distance of 15 metres from the equipment - subtract 23dB for: 93-23 = 70 dBA @ 15 metres away



Pump Clinic 10 Pump Operation and Maintenance 05/07/06 Page 1 of 11

PUMP CLINIC 10

PUMP OPERATION AND MAINTENANCE

Index

Page

2 Centrifugal Pumps Fault-finding Table 1

3 Reciprocating Pumps Fault-finding Table 2

4 Reciprocating (Piston) Pumps Fault-finding Table 3

5 Reciprocating Pumps (Direct Acting

5 Rotary Pumps

6 Lobe-Rotor Pumps

6 Condition Monitoring

7 Rotary Pumps Fault-finding Table 4

8-9 Lobe-Rotor Pumps Fault-finding Table 5

10 Mean Time Between Repair (MTBR)

10-11 Pump Protection




TABLE 1 - Centrifugal Pumps - Fault-finding Fault Cause Remedy or Action

Driver not running Check fuses, circuit breakers Keys sheared Replace Drive belt slip Check and adjust Coupling fault Check if slipping or broken, replace if necessary

Pump not turning

Shaft or gears sheared Check, replace if necessary

Inlet valve closed Open valve Inlet clogged or restricted Check and clear Air leaks on suction side Replace seals, check line(s) for leaks Liquid drained or syphoned from system Fit check or foot valve to prevent draining Pump not priming

Worn pump impeller Inspect, increasing pump speed might help, also fitting foot valve

Lack of prime (see also above) Open all vent cocks to release trapped air and fill pump and suction pipe completely with fluid

Excessive suction lift Check pump inlet for clogging etc. causing excessive friction head Check that valves are open Check piping for obstructions or blockage Excessive discharge head Check total head

Speed too low Check that pump rev/min is consistent with manufacturer’s recommendations

Pump clogged Check that impeller is not clogged Wrong direction of rotation Check that pump is running in the correct direction

Bleed suction pipe to clear air lock Vapour lock Check that suction pipe is properly submerged Relief valve not properly adjusted Check adjustment, check for dirt on valve seat

No discharge

Air leak Check seals, check line(s) for air leaks

Check suction piping and pump for air leaks Air leaks Check pump gaskets

Vapour lock Check NPSH and fluid temperature to ensure that liquid in suction line is not ‘flashing’

Low NPSH or damage As above & also check suction pipe, foot valve etc Clogged strainer(s) Check and clean if necessary Excessive inlet friction Suction line too small, or too many fittings adding fluid friction Relief valve incorrectly set or jammed Check and adjust as necessary

Excessive system back pressure Reduce system friction by re-design Worn impeller Inspect and replace if necessary Worn wear rings As above Wrong direction of rotation Check against manufacturer’s specification

Check that foot valve size is adequate (if fitted) Construction in suction line Check for other possible obstruction Wrong pump size Check that pump is adequate for the job

Low delivery

Poor suction Check suction pipe is properly submerged and in best position




TABLE 2 - Reciprocating Pumps - Fault-finding

Fault Cause Remedy or Action High fluid viscosity Check that fluid viscosity is consistent with anticipated

performance Excessive fluid temperature Reduce speed and/or delivery, decrease suction head Low delivery (Cont.)

Speed too low Check operating rev/min against specification for pump

performance

Check that packing is not too tightly or badly fitted Check packing lubricant (where possible) Check that packing is consistent with manufacturer’s specification

Stuffing box over-heats

Check cooling flow (where applicable) Check oil level or lubricant condition Check if correct lubricant is being used Check bearing for misalignment or excessive tightness Check fitting and condition of oil seals

Bearings over-heating

Check that operating speed is not excessive Fluid too viscous Reduce fluid viscosity (eg by heating)

Over-heating

Excessive pressure Reduce pump speed, increase delivery line size(s)

Cavitation Check pump operating conditions Excessive fluid viscosity Check product suitability Entrained air Check for air leaks High vapour pressure fluid Check product/pump suitability Improper pump assembly Check and rectify Unbalanced impeller Check that impeller is not damaged or clogged Misalignment Check alignment with driver Non-rigid mount Check mounting for rigidity Bent shaft, faulty bearings Check and replace if necessary Pump wear Strip down and check for wear

Vibration and noise

Relief valve chattering Re-adjust, repair or replace as necessary

Misalignment As above Out-of-balance As above Non-rigid mount As above Bent shaft As above Lack of vibration Check quantity and quality of lubricant Dirt in pump Use filter to remove

Corrosion Check that pump materials are compatible with fluid being handled

Too high operating speed Check against manufacturer’s recommendations for fluid viscosity

Operating pressure too high Reduce speed or pressure, eg change in system

Excessive wear

Abrasives present in fluid Check product/pump suitability

Speed too high Check against recommended rev/min Misalignment Check alignment of pump and driver, also foundations Internal friction Check for rubbing contact, clogging etc

Tight bearings Check bearings and packings (bearing temperature will be a clue)

Lack of lubrication Check quantity and quality of lubricants

Pump requires excessive power

High fluid viscosity Check that fluid viscosity is not too high for economic handling




TABLE 3 - Reciprocating (Piston) Pumps - Fault-finding

Fault Cause Remedy or Action Prime

Not primed Open vents on discharge side to release trapped air and leave open until all air is discharged

Excessive suction lift Reduce suction lift or reduce suction friction with larger diameter pipe or eliminate bends etc

Air leaks Check system and eliminate air leaks by sealing etc

Vapour-bound Check fluid temperature and vapour pressure - fluid temperature may be too high, or suction lift excessive for fluid temperature Check for blockage in suction pipe, foot valve or strainer Blockage Check pump suction valves

No discharge

Deterioration Check suction valves, piston packing, piston rod packing, worn valves or badly scored cylinder

Low steam pressure Check for obstruction, leak or partially closed valve in steam system

Tight packing Loosen gland until leakage is apparent

Excessive backpressure Check that pump is not operating against excessive system head

Low discharge pressure

Deterioration Check cylinder bore for wear, also condition of piston packings and valves

Intermittent steam supply Check for blockage etc Check for excessive valve wear and leakage Valve trouble Check that valve timing is correct

Pump stops or hesitates

Excessive back pressure Check that system head is not excessive

Air leaks Check system for air leaks

Misalignment Check alignment of pump and possible distortion due to unsupported piping connected to pump cylinder

Excessive suction lift As above Vapour-bound Reduce fluid temperature, or reduce suction lift Tight packing Loosen gland

Variable delivery

Excessive speed Check that operating speed is consistent with specification

Excessive cushioning Adjust cushioning valves to obtain stroke Worn valve Check valve for leakage, re-face as necessary Entrained gas Modify suction intake as necessary Incorrect valve timing Check against specification and adjust as necessary

Pump short strokes

Worn bore Replace liner, or re-bore and fit oversize piston

Excessive cushioning Adjust cushioning valves to obtain correct stroke Worn valve Check valve for leakage, re-face as necessary Entrained gas Modify suction intake as necessary Incorrect valve timing Check against specification and adjust as necessary

Piston short-strokes

Worn bore Replace liner, or re-bore and fit oversize piston

Cushion valves Close down as necessary to reduce stroke

Piston leakage Replace piston packing, replace worn liner or re-bore cylinder Piston over-strokes

(hits head) Valve leakage Check valves on liquid head for leakage, re-grind or re-set if

necessary

Misalignment Check alignment Bent piston rod Check for straightness Worn bore Replace liner or re-bore Excessive wear

Fluid Check that pump is compatible with fluid being handled




In many designs of pump intended for severe service conditions, provision is made for easy inspection, removal and replacement of worn parts without completely stripping down the pump, thus simplifying maintenance and making if both logical and economic to attend to maintenance at frequent intervals. The only reason that really justifies shut-down and complete dismantling of the pump for inspection and overhaul is a marked loss of performance. If this is suspected, the pump can first be tested to establish if head, capacity or power input figures have deteriorated and whether remedial action is needed. Reciprocating Pumps (Direct-Acting) Steam pumps are generally trouble-free in operation and require little attention other than routine lubrication. Gaskets and packings will, however, require periodic replacement and maintenance will be required to correct for mechanical wear etc. Disassembly is usually straightforward and a specific procedure is usually detailed in the manufacturer’s instruction manual. Power-driven reciprocating pump maintenance is largely confined to checking the valve condition together with periodic attention to seals or packings. Valve seats may be re-ground, re-faced and ground, or replaced and ground in, depending on the severity of the wear. Operating time before such attention is required will vary widely with the type and design of pump and the service conditions. It is good practice to check the valves on a new pump after three months’ service unless there is some indication that earlier attention is needed. After that, maintenance periods can be based on experience. Rotary Pumps Rotary pumps in general have a minimum of moving parts and porting rather than valve systems, hence maintenance is normally held to a minimum and is largely confined to inspection for wear, corrosion or material defects, together with keeping the seals in good condition. Disassembly normally follows a logical sequence when individual parts may be inspected for condition and wear, and clearance is checked against permitted tolerances. Excessively worn or defective parts should be replaced. Where corrosion is also present, it is advisable to check the material specifications against the liquid being handled as a replacement part of a better material may be available. There may be specific instructions for disassembly, eg a spindle should be withdrawn in one direction only to avoid damage to a seal. The manufacturer’s instruction book will, therefore, be the primary guide for maximum permissible wear on individual components. The performance of many rotary pumps, particularly of the lobe rotor and multi-screw types, depends on accurate timing and the maintenance of prescribed clearances between the rotating elements. Any deviation from such settings normally results in a marked reduction in pump life and loss of efficiency. The importance of periodic timing and clearance checks is also emphasised. In vane pumps, the principal wear is on the vanes themselves and the liner or casing on which they rub. With sliding vanes replacement is usually a straightforward matter. In some designs, it is possible to reassemble a worn vane the other way round and so double its effective life. With swinging vanes, the shape of the pivoted portion of the rotor is such that a substantial degree of wear can take place without interfering with the sealing efficiency. There will, however, be a limit to the amount of wear which can be taken up purely by geometry and once this is exceeded, it will be necessary to replace the vane. With both types, vanes may usually be replaced merely by removing the cover plate or head plate, leaving the rotor in position.




Lobe-Rotor Pumps Lobe-rotor pumps are precisely made machines; undue force must never be used during servicing. Be especially careful not to damage the means of alignment between rotor case and gear case. Ensure that sealing surfaces of mechanical seal rings are not scratched or damaged in any way. If in doubt, obtain correct fitting length from manufacturer. For shut-down and maintenance, isolate the pump from electrical and hydraulic supplies. Do not allow the product to solidify in the pumping chamber or on shaft sealing surfaces. Product wetted parts can often be cleaned with hot water and detergent. It is advisable to hold components of the following types in stock. Quantities depend on pump design, service conditions and stocking policy:

a) O-ring seals - pump head, shaft sleeve, front valves, mechanical seals, port connections, gear case.

b) Oil seals and joints - gear case.

c) Glands - packing sets, preferably pre-formed rings.

d) Mechanical seals - rotating and stationary sealing rings, springs.

e) Transmission - flexible coupling parts, belts, gearbox parts.

f) Miscellaneous - lubricating oil, relief valve springs, fasteners.

Condition Monitoring The terms predictive maintenance, or machine diagnostics, are often used to anticipate pump failure and determine a probable cause. Incorrectly stating the problem is one of the most common errors in troubleshooting pumps. The primary advantages of predictive maintenance are:

· Reduced maintenance

· Increased machinery availability

· Improved plant safety Condition monitoring plays an important role in avoiding pump failure. In the case of centrifugal pumps, three areas of monitoring can be distinguished:

1) Monitoring the mechanical components

2) Monitoring the shaft seal

3) Monitoring the hydraulic components Automatic monitoring has become more widely used, particularly in fields where safety is a major factor. Comprehensive condition monitoring in the form of documental reports and data should be kept to help identify recurring problems and determine optimum timing for scheduling preventative maintenance. Tracking the mean time between repair (MTBR) for all pumps serviced can improve reliability and identify troublesome units.




TABLE 4 - Rotary Pumps - Fault-finding

Fault Cause Remedy or Action

Not primed Prime to fill pump

Excessive suction lift Reduce suction lift or reduce friction in suction side with large pipe

Air leaks Check and rectify, check gaskets Blockage Check adjustment and setting

Excessive wear Check components for wear against manufacturer’s permitted tolerances

Wrong rotation Check that pump is being rotated in the correct direction

No discharge

Insufficient speed Check that pump is running at rated speed

Insufficient speed As above Wrong rotation As above Excessive suction lift As above Air leaks As above, check gasket particularly Air entrainment Re-position suction inlet Relief valve or bypass valve Setting may be too low. Check and re-set

Low discharge pressure or reduced capacity

Excessive wear Check as above

Misalignment (where applicable) Check alignment of driver and pump and drive connection Internal damage Bent or broken rotor Unbalance If suspected, check rotor for static and dynamic balance Air entrainment Re-position suction inlet Air leaks Check and rectify Cavitation Check against causes of cavitation

Excessive pressure Relief valve set too high, adjust to correct setting consistent with pump rating

Excessive noise

Deterioration Check for excessive wear or clearances on components

System pressure If system pressure is too high for pump rating, a larger pump will have to be used, when some relief may be possible

Relief valve or bypass valve Check and re-set relief valve for correct pressure Excessive discharge pressure

System throttled Discharge valve may be partially closed or system partially blocked

Abrasive liquid Check that pump is suitable for handling liquid if abrasive solids are present, or check that filter or strainer used is adequate

Distortion Pipework loads transmitted directly to the casing may cause distortion

Excessive pressure developed See above

Excessive wear

Excessive speed Check that speed is consistent with pump specification for viscosity of liquid handled

Damage Check for bent or damaged shaft etc Excessive pressure As above

Excessive fluid viscosity Check speed rating against actual viscosity of fluid, reduce speed for higher viscosities

Excessive input power required

Excessive speed Check against pump rating for fluid viscosity handled

Relief or bypass valves Check that settings are correct Excessive speed for fluid handled Check that speed is consistent with rating for fluid viscosity Excessive pressure As above Pump overheats

Discharge throttle Looped flow through relief valve will cause heating, may be relieved by separate relief valve discharging to tank




TABLE 5 - Lobe-Rotor Pumps - Fault-finding

No

disc

harg

e

Und

er c

apac

ity

Irreg

ular

dis

char

ge

Prim

e lo

st a

fter s

tarti

ng

Pum

p st

alls

whe

n st

artin

g

Pum

p ov

er-h

eats

Mot

or o

ver-h

eats

Exc

essi

ve p

ower

abs

orbe

d

Noi

se a

nd v

ibra

tion

Pum

p el

emen

t wea

r

Exc

essi

ve g

land

/sea

l wea

r

Pro

duct

loss

thro

ugh

glan

d

Sei

zure

Cause Remedy or Action

x Incorrect direction of rotation Reverse motor.

Expel gas from supply line and pumping chamber and x Pump unprimed Introduce liquid. Increase supply line diameter and increase static suction x x x x x Insufficient NPSH available head. Simplify supply line. Configuration & reduce length. Reduce speed. Decrease product temp., check effect of increased viscosity on available

x x x x Product vapourising in supply line

permitted power inputs. Remake pipework joints. Adjust x x x x Air entering supply line or repack gland. Expel gas from supply line and pumping chamber and x x x x x Gas in supply line introduce liquid. Raise product level. Lower outlet position, increase sub- x x x x Insufficient head above supply

vessel outlet mergence of supply pipe.

x x x x x Foot valve/strainer obstructed or blocked Service fittings

Decrease pump speed. x x x x x x Product viscosity above rated figure Increase product temperature

Increase pump speed. x Product viscosity below rated figure Decrease product temp.

Cool the product/pumping x x x x x Product temperature above rated figure chamber.

Heat the product/pumping chamber. (Check with pump X X X Product temperature below rated

figure maker). Clean the system. Fit x x x x x Unexpected solids in product strainer to supply line. Check for obstructions. Service system and revise to prevent problem recurring.

x x x x x x x x Delivery pressure above rated figure

Simplify delivery time.

x x x x x Gland over-tightened Slacken and readjust gland.

x x x x x Gland under-tightened Adjust gland.




(TABLE 5 - Lobe-Rotor Pumps - Fault-finding (Cont.)

No

disc

harg

e

Und

er c

apac

ity

Irreg

ular

dis

char

ge

Prim

e lo

st a

fter s

tarti

ng

Pum

p st

alls

whe

n st

artin

g

Pum

p ov

er-h

eats

Mot

or o

ver-h

eats

Exc

essi

ve p

ower

abs

orbe

d

Noi

se a

nd v

ibra

tion

Pum

p el

emen

t wea

r

Exc

essi

ve g

land

/sea

l wea

r

Pro

duct

loss

thro

ugh

glan

d

Sei

zure

Cause Remedy or Action

Check that fluid flows freely into x x Gland flushing inadequate

gland. Increase flow rate.

x x x x x Pump speed above rated figure Decrease pump speed.

x Pump speed below rated figure Increase pump speed

Check alignment of pipes. Fit flexible pipes or expansion x x x x x x Rotor case strained by pipework fittings. Support pipework. Re-tension to maker’s

x Belt drive slipping recommendations. Check flange alignment and

x x x x Flexible coupling misaligned adjust mountings accordingly. Fit lock-washers to slack

x Insecure pump/driver mountings fasteners and re-tighten. Refer to pump maker for

x x x x x x Shaft bearing wear or failure advice and replacement parts. Refer to pump maker for

x x x x x x x Worn unsynchronised timing gears advice and replacement parts.

Refer to pump maker’s x x x x x Gear-case oil quantity/quality

incorrect instructions. Check rated and duty pressure.

x x x x x x x Metal-to-metal contact of pumping element Refer to pump maker.

x Worn pumping element Fit new components

Check pressure setting and readjust if necessary. Examine and clean seating surfaces.

x x x Front cover relief valve leakage

Replace worn parts. Check wear of sealing surfaces

x x Relief valve chatter guides etc, replace as necessary. Readjust spring compression. Valve should lift about 10% x x Relief valve incorrectly set above duty pressure.

DIAGNOSIS WILL BE GREATLY ASSISTED BY TAKING ON-STREAM PRESSURE READINGS AT THE PUMP’S INLET & OUTLET PORTS.




Mean Time Between Repair (MTBR) Mean time between repair is calculated using the following formula:

MTBR = (P-NR) x number of months data Number of repairs Where: P = Total number of pumps in population NR = Number of pumps that have installed spares such that one is

normally not running Repair = An event that makes the pump unavailable for pumping

Average pump life can be quite an accurate measure of pump reliability. Before a pump is taken out of service, as much hydraulic performance data as possible should be obtained. This can be achieved by accurately measuring flows, temperature, pressure, specific gravity, viscosity etc around the pump and by consulting historical data and other useful information about its operating condition. Simple vibration monitoring and analysis can be an accurate and rapid method for detecting mechanical problems in a pump. Three basic measurements involve:

1. Measuring the bearing cap or casing vibration. 2. Measuring the shaft vibration relative to the bearing. 3. Measuring the absolute shaft vibration.

The objective in monitoring rotating machinery vibrations is to determine when the rate of change in vibration level begins to change. Sealless pumps generally require more monitoring than mechanical seal pumps. Temperature measurement, bearing condition monitoring, low current trips (for cavitation protection) are all useful for magnetic drive pumps. Canned motor pumps also require liquid level monitoring. Some of the most common problems with pumps include misalignment, oil contamination, incomplete priming and the absence of detailed up-to-date spare parts records. Pump Protection In many cases, the reasons for pumping system failure are other than the pumping system itself. These include improper selection of the pump, improper handling of the system and improper selection of protective devices. Protection for a pump falls into six broad categories:

1. Against phase failure and unbalanced supply 2. Against over-loading 3. Against dry-running 4. Against over-heating 5. Against moisture 6. Against under/over voltage




Typically, the selection of appropriate protective devices should take into account the following points:

a) It operates on the current sensing principle. b) It senses the negative sequence components of the supply. c) It offers protection against phase failure faults. d) It offers protection against failure even at No-load conditions. e) It offers protection against overloading according to the thermal-withstanding

characteristics of the motor. f) Dry running protection is not based on sensing the water level directly by a sensor but by means of indirect methods like sensing current, rpm or pressure.

Protective devices incorporating these characteristics will almost certainly lead to longer pump life. The development of sophisticated electronics, computers, chemical analysis, lasers, vibration pick-ups, sonic measurements, ultra-sonics, and radio graphics have permitted the analysis and forecast of the life of any pump component so that opening up maintenance can be carried out at a reasonable period before expected failure, rather than at routine times. There is really no point in stripping of a machine if it is certain that it is operating satisfactorily. It is probable that the cost of monitoring will amount to less than the costs of routine maintenance disassembly, particularly if predictive maintenance is employed.



Pump Clinic 11 Specifying and Purchasing Pumps 06/07/06 Page 1 of 11

PUMP CLINIC 11

SPECIFYING AND PURCHASING PUMPS

Page 1 Pump Purchasing Sequence

2-4 Engineering of System Requirements -Fluid Type -System-Head Curves -Alternate Modes of Operation -Margins -Wear -Future System Changes

5-7 Selection of Pump and Driver -Pump Characteristics -Code Requirements -Fluid Characteristics -Pump Materials -Driver Type

7-10 Pump Specifications -Specification Types -Codes and Standards -Alternates -Bidding Documents -Technical Specification -Commercial Terms

10-11 Special Considerations -Performance Testing -Pump Drivers -Intake -Drawing & Date Requirement Forms




Pump Purchasing Sequence The sequence involved in obtaining a pumping system, following the initial decision that pumping equipment is required for a system, and culminating with the purchase of the equipment, can be divided into the following general steps:

· Engineering of system requirements · Selection of pump and driver · Specification of pump · Bidding and negotiation · Evaluation of bids · Purchasing of selected pump

In the process of specifying pumping equipment, the engineer is required to determine system requirements, select the pump type, write the pump specification, and develop all information and data necessary to define the required equipment for the supplier. Having completed this phase of work, the engineer is then ready to take the necessary steps to purchase the equipment. These steps include issuance of the specifications for bids or negotiations, evaluation of pump bids, analysis of purchasing conditions, selection of supplier, and release of all data necessary for purchase order issuance. It is imperative that a pragmatic approach be taken when specifying requirements. Too often, purchasers provide the same degree of specification, whether a pumpset is valued at $5000 or $500,000. It is possible that documentation and testing requirements can be greater than the cost of the equipment. The message is, ‘resist using a sledgehammer to crack a walnut.’ Engineering of System Requirements The first decision the engineer must make is to determine the requirements and conditions under which the equipment will operate. Fluid Type One of the initial steps in the defining of the pumping equipment is the development of physical and chemical data on the fluid handled, such as viscosity, corrosiveness, lubricating properties, chemical stability, volatility and amount of suspended particles. Depending upon the process and the system, some or all of these properties may have an important influence on pump and system design. For example, the degree of corrosiveness of the fluid will influence the engineer’s choice of materials of construction. If the fluid contains solids in suspension suitable types of pump seal designs and abrasion-resistant pump construction materials may have to be considered. Fluid physical and chemical data of interest to the engineer should cover the entire expected operating range of the pumping equipment. The influence of such parameters as temperature, pressure and time upon the fluid properties, should also be considered. System-Head Curves The engineer should have a clear concept of the system in which the pumping equipment is to operate. A preliminary design of the system should be made and include an equipment layout and a piping and instrumentation diagram (or other suitable diagram), showing the various flow paths, their preliminary size and length, elevation of system components and all valves, equipment, piping specialties etc. which establish the system-head losses.




System Head Curves (Cont.) The engineer should then determine the flow paths, flow rates, pressures and temperatures for various system operating conditions and calculate line, and estimate pipe runs. With this information, the engineer can develop system-head curves which show the graphic relationship between flow and hydraulic losses in a pipe system. In calculating the hydraulic losses, the engineer may need to include adequate allowances for corrosion and scale deposits etc. in the system over the life of the plant. Since hydraulic losses are a function of flow rate, piping sizes, and layout, each flow path in a system will have its own characteristic curve. Care must be taken when specifying pump characteristics to take into account the characteristic curve of each possible flow path served by the equipment. In specifying pump equipment, it is convenient to add the effect of static pressure and elevation differences to the system-head curve to form a combined system-head curve. The resultant curve shows the total head required of the pumping equipment to overcome system resistance. The pump head must be at, or above, the combined system curve at all expected operating points, and for all flow paths the pumping equipment is expected to serve. Alternate Modes of Operation The various modes of operation of the system are important considerations when specifying pumping equipment: · Is operation of the pumping equipment to be continuous or intermittent? · Is the flow head to be fluctuating or constant? · Will there be a great difference in flow and head requirements for different flow paths? These and other questions arising from different modes of operation greatly influence such decisions as to the number of pumps, their capacities, and whether booster pumps are needed in some flow paths. The engineer should also consider the continuity of service expected of the pumping system. This factor will influence the decision on number, type and capacity of installed spares and the quality expected of the equipment. Frequently, reliability considerations will dictate the use of multiple pumps, such as 2 full-size pumps and 3 half-size pumps, or, where continuity is more important than full capacity, 2 half-size pumps. Where 2 half-size, 3 third-size etc. pumps are used, loss of 1 pump will cause the others to run-out on their system-head curves. This run-out case should be evaluated when engineering the system and specifying the pump characteristics. This loss of a pump can occur no only by pump malfunction, but by motor failure, external damage, loss of power supply, loss of control power etc. The likelihood of these causes should be evaluated as part of the pump selection process. Margins Pumps are frequently specified with margins over and above the normal rating. Over any long period of time, it is possible that a system can operate at transient conditions, such as may be caused by changes in modes of operation, malfunction in system components, or electric system disturbances. It is necessary for the engineer specifying the pumping equipment to examine the probability and duration of such transients and to specify adequate margins to allow the equipment to undergo such transients without damaging effects.




Margins (Cont.) This also involves an evaluation of the combined effects of equipment cost, degree of criticality of the system, inconvenience due to unavailability of the equipment and other economic and technological factors. Some transients often considered in design are pressure and temperature fluctuations, electric voltage and frequency dips. If the maintenance of continuous flow is important, then adequate margins must be allowed in the pump rating. For example, margins are added to the pump head and capacity to allow the pumping equipment to maintain rated flow in case of small electric frequency dips. In addition, certain design features may be included to allow the pumps to operate without damage through such transients as suction pressure dips which can cause cavitation. Pumps should not be purchased for capacities greatly in excess of requirements. An over-sized pump could operate at capacities less than those recommended by the manufacturer which could present mechanical and hydraulic problems. Wear Wear is an ever-present factor in equipment and system design. No material that is handling fluids or used in contacting moving surfaces is free from wear. Thus, operating characteristics of both the pumping equipment and the system can be expected to change due to wear as time goes by. The engineer should assess the extent of such wear over the life of the plant and provide adequate margins in the system parameters so that the pumps can provide the expected flow, even at the end of equipment life. Where abrasive or suspended materials are handled, pumps with replaceable liners are frequently specified. These liners are usually made of either resilient material such as rubber compounds, or extremely hard alloys of cast iron. In addition, plastic linings (including impellers) are also frequently chosen for these types of services. In some applications, especially in power plants, the expected pump life is specified as the same as plant life. However, the design life of a pump is a decision based on an evaluation of economic factors. The wear margin to be added is a function of such factors as mode of operation (continuous or intermittent) and fluid properties (abrasiveness, corrosiveness). Future System Changes A final factor to be considered in the engineering system requirements for pumping equipment is the possibility of providing for future system changes. If the system changes can be predicted with any degree of certainty, then the system can be designed to enable the changes to be effected with minimum disturbance to operation. Thus it is important to review the possibilities and effects of such future system changes as well as provide pumping equipment to satisfy the immediate system requirements. The engineer should attempt to present future requirements based on projection of available data and then evaluate the possibility and desirability of designing the equipment to allow for the changes (such as providing extra flow or head margins or specifying a pump with impeller less than the maximum for a given casing size) versus the desirability of modifying the whole system, including the pumps, when the changes are made in the future. In any event, it must always be kept in mind that the equipment must operate satisfactorily in the present system and this should be a factor in whatever evaluation is being made.




Selection of Pump and Driver The selection of the pump class and type for a particular application is influenced by such factors as system requirements, system layout, fluid characteristics, intended life, energy cost, code requirements and materials of construction. Basically, a pump is expected to fulfil the following functions:

· Pump a given capacity in a given length of time · Overcome the resistance in the form of head or pressure imposed by the systems while

providing the required capacity The behaviour of the system has a very important bearing on the choice of pump:

· What are the required heads and capacities at different loads? · Does the required head increase or decrease with changes in capacity? · Does the required head remain substantially constant?

These are some of the questions the engineer must answer. Pump Characteristics Constant-speed reciprocating pumps are suitable for applications where the required capacity is expected to be constant over a wide range of system head variations. This type of pump is available in a wide range of design pressures, from low to the highest produced. However, the capacity is relatively small for the size of the equipment required. That the output from a reciprocating pump will be pulsating is a factor to be considered. Where this is objectionable, rotary pumps may be required. However, the application of rotary pumps is limited to low to medium-pressure ranges. Centrifugal pumps are often used in variable-head, variable-capacity applications. Straight centrifugal pumps are generally used in low to medium, to high-pressure applications, while low-head, high-flow conditions suggest that an axial-flow pump may be more suitable. Mixed-flow impellers are used in intermediate situations. It should be noted that some reciprocating and rotary pumps may be self-priming, but centrifugal pumps, unless specifically designed as such, are not. This may be an important consideration in certain applications. In some cases, the system layout can influence the decision on the choice of pump type. In general, centrifugal pumps will require less floor space than reciprocating pumps, and vertical pumps less floor space than horizontal pumps. However, more head room may be required for handling the vertical pumps during maintenance and installation. Where the available NPSH is limited, such as when a saturated liquid is being handled, and the application calls for a centrifugal pump, the engineer may have to investigate the use of a vertical canned-suction centrifugal pump to gain adequate NPSH. In other cases, the design may call for the installation of a pump immersed in the liquid handled, and here a vertical turbine pump may be advantageous. Code Requirements The construction ratings and testing of most pumps normally used in industry are governed by codes such as the ISO, ANSI, API or the Standards of the Hydraulic Institute. However, other codes of regulatory bodies may impose additional requirements which can affect both pump rating and construction. For example, the Boiler and Pressure Vessel Code requires feed pumps to be capable of feeding the boiler when the highest set safety valves are discharging.




Fluid Characteristics Fluid characteristics such as viscosity, density, volatility, chemical stability and solids content are also important factors for consideration. Sometimes, exceptionally severe service may rule out some classes of pumps at once. For example, the handling of fluids having solids content will exclude the use of reciprocating piston pumps, or pumps with close clearances. Rotary pumps are suitable for use with viscous fluids, such as oil or grease, whereas centrifugal pumps can be used for both clean, clear fluids, and fluids with high solids content. On the other hand, if it is undesirable for the process liquid to come into contact with the moving parts, diaphragm pumps may have to be used. Pump Materials Materials are affected by both the pumped fluid and the environment. Resistance to corrosion and wear are two of the more important material properties in this regard, and the engineer should evaluate materials to determine which are most suitable and economical for the purpose intended. Often this becomes an evaluation for the desirability of specifying expensive long-life materials versus specifying cheaper materials which must be frequently replaced. Operating factors, such as type of service (continuous or intermittent, critical or non-critical), running speed preferences (high or low) and intended life, will also influence the engineer’s decision. For example, equipment used in continuous and/or critical service will generally demand heavier duty design and construction than equipment for intermittent and/or non-critical service. High-speed operation, if allowed, will permit the use of smaller, usually less expensive equipment. The life of the equipment cannot be predicted with certainty. For a given life expectancy, the engineer must evaluate the effects of materials of construction, design, severity of service etc. before making a choice. Driver Type The choice of driver type for the pumping equipment is as important as choosing the pump, for frequently the driver can cost more than the pump. Depending on the available energy sources, pumps may be driven by electric motors, steam turbines, steam engines, gas turbines, or internal combustion engines. Also, pumps may be driven at constant speed or at variable speed. Variable speed can enable centrifugal pumps to operate along the system-characteristic curve and thus save on power for part-load operations. Electric motor drives are usually used in constant-speed service unless a hydraulic coupling or other speed-varying device is introduced into the system. Internal combustion engine drives are usually chosen because of location (no electric power available), portability, or redundancy (loss of power back-up) requirements. They can operate as either constant-speed for variable-speed drivers. Steam turbines, eddy-current couplings, adjustable-speed motors, fluid couplings, and gears and belts are frequently used where variable-speed operation is required. In large, complex installations where the equipment is to be operated continuously, the decision as to type of driver and variability of pump speed should be based on a comparison to the total operating and capital costs for the pump system over the intended plant life for the several alternatives. Variable-speed operation would usually result in lower operating costs. However, the total first cost of the driving equipment to accomplish this would frequently be higher than for constant-speed equipment. The first cost should include cost of equipment, building space etc and the operating cost should include such factors as energy costs, maintenance, and replacement costs. This comparison usually results in the choosing of a pumping unit that provides the lowest cost per gallon pumped over the useful life of the plant.




Driver Type (Cont.) It should be recognised that these are general guidelines and that there may be overriding non-technical and non-economic factors, such as prior satisfactory experience or excellent technical or service assistance which may dictate the final choice of pumping equipment. Pump Specifications Specification Types When selecting a pump, the first decision to be made is whether the procurement will be based upon a formal specification, or whether some abbreviated form of requirements will be suitable. For relatively simple or inexpensive pumps, or for replacement pumps where duplication is desired, a specification is frequently not used. For inexpensive pumps, the time and cost required to write a specification and obtain and analyse competitive bids, frequently exceed the potential cost-savings. In this case, and where the pump supplier is already established (replacement/duplication), a direct quotation is frequently requested from the supplier. It is important, when requesting this quotation, to have the principal requirements well defined and known to the supplier so that they can be properly included in the technical, and priced, offering. Thus, while a formal specification may not be appropriate, the purchaser should have the requirements well established. Attached are four (4) data sheets that can be utilised for enquiry purposes. As a minimum, details under Operating Conditions and the Quantity required are to be defined by the purchaser. Where a formal specification is indicated, the type to be written is of fundamental importance. In most cases, the specification will be of the performance type rather than the construction type. The performance specification basically establishes the performance which the pump must achieve and does not attempt to dictate pump design or construction methods, although certain details of construction are frequently established, particularly where choices may exist. For example, where either leak-off or mechanical shaft seals may be offered, the performance specification usually states a preference. The performance specification, however, basically establishes ‘what’ not ‘how’. The construction specification establishes in some detail the type of design, construction, and methods to be employed in designing the pump and certain other features which, if the performance specification is utilised, are left to the manufacturer’s discretion. From the standpoint of legal responsibility, if a construction specification is used, manufacturers may respond and advise that since the purchaser has established certain design features of the pump, the manufacturer cannot be responsible for the performance. It is therefore, important that care be taken when writing a construction specification not to relieve the manufacturer of responsibility for applicability, suitability, and performance and that care also be exercised by the purchaser to avoid any unnecessary assumption of responsibility for the proper performance of the pump. In short, unless there are unusual circumstances, it is far more appropriate to specify the pump on the basis of performance required, rather than construction, unless the purchaser has a high degree of assurance that the requirements called out in the specification can be met and that the pump supplier will not be relieved of responsibility.




Codes and Standards When specifying a pump, the codes and standards that apply are of major importance. Standards relating to quality of materials should be referenced ANSI, ISO, API or other industrial standards which establish such factors as metallurgy, dimensions, tolerances, and flange facing and drilling should be referenced where appropriate. Similarly, if a pump is to meet certain critical service requirements, there are, in many cases, industrial codes which apply to design, construction and application. An example of this is the ASME Boiler and Pressure Vessel Code, Section III. These codes, in some cases, are extremely detailed in specifying pump construction and are a rather well-defined specification in themselves. It is, of course, essential to establish the dimensional standards which apply, such as SI or English, the codes which may apply to the construction and fittings of the pump and the industrial codes that apply to the application of the pump for the service intended. In all cases however, the engineer must review each reference to ensure that it does not introduce conflict. Some codes and standards include alternate choices of material or inspection methods requiring selection by the engineer. Others may include cross references to additional codes which the engineer may wish to exclude. Alternates It is extremely difficult for a specification to cover all possible pumps offered by various manufacturers. Coupled with that is the problem faced by a potential user in remaining up-to-date with the changing state-of-the-art, and the development work being performed by manufacturers. It is good practice to allow manufacturers to offer alternatives. This gives them an opportunity to present their best offer and also gives the buyer the advantage of obtaining potentially attractive, alternate offerings. However, the choice of whether or not to accept the alternates is fully up to the purchaser who may choose to reject any, and all bids, including alternates. Bidding Documents The bidding documents for pumps normally consist of two major parts:

1. Technical specifications 2. Commercial terms

The technical specification establishes the performance requirements, materials of construction and major technical features. The commercial terms include the contract language and cover such items as the location of the work, requirements for guarantees/warranties, shipping method, time of delivery, method of payment, normal inspection and expediting requirements. Frequently, the commercial terms and conditions are relegated to second place, especially when standard inexpensive pumps are being bought, but in many cases, the commercial terms can assume more significance than many of the performance requirements. Technical Specification The technical specification should consist of a series of carefully defined and distinct sections. The more complete and specific the specification, the more competitive will be the bid prices. A typical specification might contain the following:

1. Scope of work: Pump, baseplate, driver (if included), interconnecting piping, lubricating oil pump and piping, spare parts, instrumentation (pump-mounted), erection supervision.

2. Work not included: Foundations, installation labour, anchor bolts, external piping, external wiring, motor starter




3. Rating and service conditions: Fluid pumped, chemical composition, temperature, flow, head, speed range, preference, load conditions, overpressure, run-out, off-standard operating requirements, transients.

4. Design and construction: Care should be taken to provide latitude in this section, as this borders on dictating construction requirements. Codes, standards, materials, type of casing, stage arrangement, balancing, nozzle, orientation, special requirements for nozzle forces and movement (if known), weld-end standards, supports, vents and drains, bearing type, shaft seals, baseplates, interconnecting piping, resistance temperature detectors, instruments, insulation.

5. Lubricating oil system (if applicable): System type, components, piping, mode of operation, interlocks, instrumentation.

6. Driver: Motor voltage standards, power supply and regulation, local panel requirements, wiring standards, terminal boxes, electric devices. For internal combustion drivers, fuel type preferred (or required), number of cylinders, cooling system, speed governing, self-starting or manual, couplings or clutches, exhaust muffler.

7. Cleaning: Cleaning, painting, preparation for shipment, allowable primers and finish coats, flange and nozzle protection, integral piping protection, storage requirements.

8. Performance testing: Satisfactory for service, smooth-running, free of cavitation and vibration, shop tests for pump and spare rotating elements, hydrostatic tests, test curves, field testing.

9. Drawing and data: Drawings and data to be furnished, outline, speed versus torque curves, WK² data, instruction manuals, completed data sheets, recommended spare parts.

10. Tools: One set of any special tools.

11. Evaluation basis: Power, efficiency, proven design. Supplementing these may be technical specifications relating to other requirements of the order, such as specifications for the electric motor, steam turbine, or other type of driver, a specification on marking for shipment, a specification on painting, and requirements for any supplementary quality control testing. In addition, it is important that any unusual requirements be listed in the technical specification so that the manufacturers are aware of them. Examples of these are special requirements for repair of defects in pump castings, a sketch of the intake arrangement for wet-pit applications and special requirements regarding unique testing, for example, metallurgical testings which may be required during manufacture, apart from performance testing. It is helpful to the pump supplier to provide system-head curves, sketches of the piping system (dimensioned if this is significant), listing of piping and accessories required etc. Pump data sheets are extremely useful in providing a summary of information to the bidder and also in allowing the ready comparison of bids by various manufacturers. As can be noted, by inspecting these sheets, some of the items are filled in by the purchaser and the balance by the bidder to provide a complete summary of the characteristics of the pump, the materials to be furnished, accessories, weight etc. The data sheets should be included with the technical specification. Commercial Terms The commercial terms included with the bidding documents should cover the following information:

1. General: Name of buyer, place to which proposals are to be sent, information on ownership of documents, time allowed to bid, governing laws and regulations.

2. Location of plant site: This establishes the geographic area in which the equipment is to perform and in a broad way the scope of the work. It should also state maximum temperatures, humidity, storage provisions (indoor or outdoor), and altitude (so that the motor drivers can be selected for the proper cooling).




3. Proposal: This establishes the format of the proposal, number of copies, owner’s right to accept or reject any bids, status of alternates.

4. Schedule: Including requirements for all drawings and design data submittals,

manufacturing schedule, and equipment delivery.

5. Acceptable terms of payment, retention, liquidated damages

6. Transportation: Transportation to and from point of use (or installation) is frequently a consideration, since with very large equipment, it may not be possible to ship by truck and it may be necessary to either barge, rail or ship the material. In addition, it is important to establish the method of shipment which may be used so that the bidder can include the proper allowance for freight and to establish responsibility for the risk of loss.

Thus a bid could either include a freight allowance, that is, be FOB manufacturer’s plant WFA (with freight allowed) to point of use, or be FOB point of use, in which case freight is included. In either case, the risk of loss remaining with the seller and that assumed by the purchaser should be clearly stated.

Special Considerations Performance Testing An important part of any specification is the requirement for testing. Normally, small commercial pumps that are routinely produced by a manufacturer, up to about 6in (150mm) are tested on a sample selection, quality control basis and from that, standardised curves of pump performance are available. Thus for pumps of this size, it is not necessary to require certified tests unless the pumps are to be used in critical service, such as fire protection or boiler feed. However, for larger pumps or pumps with more critical service requirements, a certified performance test should be required. This test requires the manufacturer to test the pump at several points on its performance curve to establish its exact head curve. Since it is necessary to assure the pump driver is of the proper size, power curve must be furnished with the head curve. Occasionally, pumps for special services or extremely large pumps cannot be tested in the manufacturer’s shop. Examples of this are very large low-head pumps for circulating water service, low-lift irrigation pumps and pumps for liquid-metal service. The actual performance testing in this case takes place following installation of the pump. It is important that the purchaser and the supplier agree upon a proper (field) test method in some detail. This method should include the number of points at which the head curve will be determined, the applicable code, the specific method of traversing the pump discharge characteristics across the cross-section of the discharge pipe, and the manner in which the head will be varied. Care should be taken in establishing this procedure to set forth the characteristics of the fluid and other variables which can affect the performance test. The specification should establish the performance testing requirements for the pump and whether or not it is necessary the testing of the pump be witnessed. Witnessing and furnishing of certified test data (including the test work done) are frequently priced separately and if not specified, can be a source of dispute between purchaser and supplier. Pump Drivers Pump drivers (motors, turbines, engines etc) can be purchased either with the pumps or separately. With small pumps, pumps using ‘monobloc’ construction (where the pump is mounted on, and supported by, the motor) and pumps built to special codes, such as ‘Underwriters’ engine-powered fire pumps, there is not usually any cost-advantage to buying the driver separately.




Pump Drivers (Cont.) Where the driver is excluded from the pump scope of supply, the specification should require the pump supplier to determine the proper characteristics of the driver. This includes establishing the proper motor speed, sizing the driver for both accelerating and running loads, assuring end-float compatibility and/or thrust-bearing loadings (including direction), and selecting and fitting the couplings. If the driver is purchased separately and can be economically and conveniently shipped to the pump supplier’s plant, the pump supplier should be required to mount the driver half of the coupling, as well as align and mount the driver (for common baseplate installations). For very large drivers, or where it is costly or impractical to ship the driver to the pump supplier, it will be necessary to perform this work at the point of installation. To assure compatibility with the other drivers in the plant, it is important to specify the driver enclosure type, insulation standards, and special features required, such as heaters and oversize junction boxes. For steam turbine drivers, speed range, throttle pressure, steam quality, exhaust pressure and control method should also be specified. Intake Vertical wet-pit pumps are sensitive to the geometry of their suction pit. Factors to consider include clearance beneath the bottom of the suction bell and the floor of the pit, spacing between pumps or between the pump and the pit walls (both side and rear), the approach angle of the floor of the pit (including surging and surcharge), submergence and lack of uniform approach flow. The standards of the Hydraulic Institute include recommendations on the geometry of the intakes. These, as well as the recommendations of the pump manufacturer, should be carefully reviewed. Suction piping, where complex or unusual, should be treated in a similar manner When specifying vertical wet-pit pumps, a layout of the installation should be furnished to the bidders for their information and comment. In many cases, if the geometry of the installation is not fixed, bidders can recommend small changes to improve pump performance. Where the geometry is fixed, it may be necessary to add anti-vortexing baffles, surge walls, or flow-directing vanes (or walls) to avoid pump operating problems. For moderate or large installations where any design question exists, model testing may be considered. Several pump manufacturers offer this as a service, as do a number of universities and commercial testing laboratories. Responsibility for proper pump performance will rarely be assumed by the bidder when the intake pit is of non-optimum size or shape. The use of model testing is usually resorted to in these cases also. Drawing and Date Requirement Forms The purchase should define the type of drawings and data required both for preliminary design purposes and for final information, that is, the as-built dimensions and the certifications which demonstrate that the pump meets the specified requirements. Attached: 4 x Data Sheets 1. Positive Displacement Pump 2. Seal/less Centrifugal Pump 3. Centrifugal Pump 4. Vertical Extended Spindle Pump



Pump Clinic 12 Radial & Axial Thrust in Centrifugal Pumps 12/09/06 Page 1 of 12

PUMP CLINIC 12

RADIAL & AXIAL THRUST IN CENTRIFUGAL PUMPS Radial Thrust A centrifugal pump consists of a set of rotating vanes, enclosed within a housing or casing and used to impart energy to a fluid through centrifugal force. Thus, stripped of all refinements, a centrifugal pump has two main parts:

1. A rotating element including an impeller and shaft 2. A stationary element made up of a casing, stuffing box and bearings

In a centrifugal pump the liquid is forced by atmospheric or other pressure into a set of rotating vanes. These vanes constitute an impeller which discharges the liquid at its periphery at a higher velocity. This velocity is converted to pressure energy by means of a volute (Fig 1.1) or by a set of stationary diffusion vanes (Fig 1.2) surrounding the impeller periphery. Pumps with volute casings are generally called volute pumps, while those with diffusion vanes are called diffuser pumps. Diffuser pumps were once quite commonly called turbine pumps, but this term has recently been more selectively applied to the vertical deep-well centrifugal diffuser pumps usually referred to as vertical turbine pumps.

The diffuser is seldom applied to a single-stage radial-flow pump. Except for certain high-pressure multi-stage pump designs, the major application of diffusion vane pumps is in vertical turbine pumps and in single-stage low-head propeller pumps.




In a single-volute pump casing design (Fig 1.3) uniform or near-uniform pressures act on the impeller when the pump is operated at design capacity (which coincides with the best efficiency). At other capacities, the pressures around the impeller are not uniform and there is a resultant radial reaction. A graphical representation of the typical change in this force with pump capacity is shown in (Fig 1.4) - NOTE that the force is greatest at shut-off.

For any percentage of capacity, radial reaction is a function of total head and of the width and diameter of the impeller. Thus a high-head pump with a large-diameter impeller will have a much greater radial reaction force at partial capacities than a low-head pump with a small-diameter impeller. A zero radial reaction is not often realised; the minimum reaction occurs at design capacity. In a diffuser-type pump which has the same tendency for over-capacity unbalance as a single-volute pump, the reaction is limited to a small arc repeated all around the impeller with the individual forces cancelling each other. In a centrifugal pump design, shaft diameter and bearing size can be affected by allowable deflection as determined by shaft span, impeller weight, radial reaction forces and the torque to be transmitted. Formerly, standard designs compensated for reaction forces if maximum-diameter pump impellers were used only for operations exceeding 50% of design capacity. For sustained operations at lower capacities, the pump manufacturer, if properly advised, would supply a heavier shaft, usually at a much higher cost. Sustained operation at extremely low flows, without informing the manufacturer at the time of purchase, is a much more common practice today. The result is broken shafts, especially on high-head units. Because of the increasing operation of pumps at reduced capacities, it has become desirable to design standard units to accommodate such conditions. One solution is to use heavier shafts and bearings. Except for low-head pumps in which only a small additional load is involved, this solution is not economical. The only practical answer is a casing design that develops a much smaller radial reaction force at partial capacities. One of these is the double-volute casing design, also called twin-volute or dual-volute. The application of the double-volute design principle to neutralise reaction forces at reduced capacity is illustrated in (Fig 1.5). Basically, this design consists of two 180º volutes; a passage external to the second, joins the two into a common discharge. Although a pressure unbalance




exists at partial capacity through each 180º arc, forces F1 and F2 are approximately equal and opposite, thereby producing little, if any, radial force on the shaft and bearings.

Axial Thrust in Single-Stage Pumps The pressures generated by a centrifugal pump exert forces on both its stationary and rotating parts. The design of these parts balances some of these forces, but separate means may be required to counter-balance others. Axial hydraulic thrust is the summation of unbalanced impeller forces acting in the axial direction. As reliable large-capacity thrust bearings are not readily available, axial thrust in single-stage pumps remains a problem only in larger units. Theoretically, a double-suction impeller is in hydraulic axial balance with the pressures on one side equal to, and counter-balancing the pressures on, the other (Fig 2.1). In practice, this balance may not be achieved for the following reasons:




The suction passages to the two suction eyes may not provide equal or uniform flows to the two sides.

1. External conditions such as an elbow being too close to the pump suction nozzle may cause unequal flows to the suction eyes.

2. The two sides of the discharge casing may not be symmetrical, or the impeller may be located off-centre. These conditions will alter the flow characteristics between the impeller shrouds and casing, causing unequal pressures on the shrouds.

3. Unequal leakage through the two leakage joints will tend to upset the balance. Combined, these factors create definite axial unbalance. To compensate for this, all centrifugal pumps, even those with double-suction impellers, incorporate thrust bearings. The ordinary single-suction radial-flow impeller with the shaft passing through the impeller eye (Fig 2.1) is subject to axial thrust because a portion of the front wall is exposed to suction pressure, thus exposing relatively more back wall surface to discharge pressure. If the discharge chamber pressure were uniform over the entire impeller surface, the axial force acting towards the suction would be equal to the product of the net pressure generated by the impeller and the unbalanced annular area. Actually, pressure on the two single-suction impeller walls is not uniform. The liquid trapped between the impeller shrouds and casing walls is in rotation and the pressure at the impeller periphery is appreciably higher than at the impeller hub. Although we need not be concerned with the theoretical calculations for this pressure variation, (Fig 2.2) describes it qualitatively. Generally speaking, axial thrust towards the impeller suction is about 20% to 30% less than the product of the net pressure and the unbalanced area.

To eliminate the axial thrust of a single-suction impeller, a pump can be provided with both front and back wearing rings. To equalise thrust area, the inner diameter of both rings is made the same (Fig 2.3). Pressure approximately equal to the suction pressure is maintained in a chamber located on the impeller side of the back wearing ring by drilling so-called balancing holes through the impeller. Leakage past the back wearing ring is returned into the suction area through these holes. However, with large single-stage suction pumps, balancing holes are considered undesirable because leakage back to the impeller suction opposes the main flow, creating disturbances. In such pumps, a piped connection to the pump suction replaces the balancing holes.




Another way to eliminate or reduce axial thrust in single-suction is by use of pump-out vanes on the back shroud. The effect of these vanes is to reduce the pressure acting on the back shroud of the impeller (Fig 2.4). This design, however, is generally used only in pumps handling gritty liquids where it keeps the clearance space between the impeller back shroud and the casing free of foreign matter. So far, the discussion of the axial thrust has been limited to single-suction impellers with a shaft passing through the impeller eye and located in pumps with two stuffing boxes, one on either side of the impeller. In these pumps, suction pressure magnitude does not affect the resulting axial thrust. On the other hand, axial forces acting on an overhung impeller with a single stuffing box (Fig 2.5) are definitely affected by suction pressure. In addition to the unbalanced force found in a single-suction, two-box design (Fig 2.2) there is an axial force equivalent to the product of the shaft area through the stuffing box and the difference between suction and atmospheric pressure. This force acts towards the impeller suction when the suction pressure is less than the atmospheric, or in the opposite direction, when it is higher than the atmospheric.

When an overhung impeller pump handles a suction lift, the additional axial force is very low. For example; if the shaft diameter through the stuffing box is 2” (area = 3.14 sq.in) and if the suction lift is 20ft of water (absolute pressure – 6.06 psia), the axial force caused by the overhung impeller and acting towards the suction will be only 27lb. On the other hand, if the suction pressure is 100 psi, the force will be 314lb and acts in the opposite direction. Therefore, as the same pump may be applied for many conditions of service over a wide range of suction pressures, the thrust bearing of pumps with single-suction overhung impellers must be arranged to take thrust in either direction. They must also be selected with sufficient thrust capacity to counteract forces set up under the maximum suction pressure established as a limit for that particular pump.




Axial Thrust in Multistage Pumps It might seem that the advantages of balanced axial thrust and greater available suction area in a double-suction impeller would warrant applying such impellers to multistage pumps. But there are definite shortcomings to this practice. The average multistage pump has relatively low capacity when compared to the entire range covered by modern centrifugal pumps. It is seldom necessary, therefore, to use double-suction impellers just to reduce the net positive suction head (NPSH) required for a given capacity. Even if a double-suction impeller is desirable for the first stage of a large capacity multistage pump, it is hardly necessary for the remaining stages. As to the advantage of the axial balance it provides, it must be considered that a certain amount of axial thrust is actually present in all centrifugal pumps and the necessity of a thrust bearing is therefore not eliminated. Most important, the use of double-suction impellers in a multistage pump adds needless length to the pump shaft span. Additional space is required for the extra passage leading to the second inlet of each successive stage. In a pump with four or more stages (Fig 2.6) this increase becomes quite appreciable and causes additional casting difficulties. If shaft diameter is increased to compensate for the longer span so as to maintain reasonable shaft deflection, the impeller inlet areas are correspondingly reduced. The result is that the advantage of superior suction conditions usually offered by double-suction impellers is considerably reduced. Finally, as it is impractical to arrange the various double-suction impellers in any but the ascending order of the stages, the impeller at one end of the casing becomes the last stage impeller and the pressure acting on the adjacent stuffing box becomes the discharge pressure on the next-to-last stage. To reduce this pressure, a pressure-reducing bushing must be interposed between the last-stage impeller and the stuffing box and this bushing further increases the overall length. The result of all these considerations is that most multistage pumps are built with single-suction impellers.




Two obvious single-suction impeller arrangements for a multistage pump are as follows: Several single-suction impellers may be mounted on one shaft, each having its suction inlet facing in the same direction and its stages following one another in ascending order of pressure (Fig 2.7). The axial thrust is then balanced by a hydraulic balancing device.

An even number of single-suction impellers can be mounted on one shaft, one half of these facing in an opposite direction to the second half. With this arrangement, axial thrust on the one half is compensated by the thrust in the opposite direction on the other half (Fig 2.8). This mounting of single-suction impellers back-to-back is frequently called ‘opposed impellers’.




An uneven number of single-suction impellers may be used with this arrangement, provided the correct shaft and interstage bushing diameters are used to give the effect of an hydraulic balancing device that will compensate for the hydraulic thrust on one of the stages. It is important to note that the opposed impeller arrangement completely balances axial thrust only under the following conditions:

1. The pump must be provided with two stuffing boxes.

2. The shaft must have a constant diameter.

3. The impeller hubs must not extend through the interstage portion of the casing separating adjacent stages.

Except for some special pumps that have an internal and enclosed bearing at one end, and therefore only one stuffing box, most multistage pumps fulfil the first condition. But because of structural requirements, the last two conditions are not practical. A slight residual thrust is usually present in multistage opposed-impeller pumps, unless impeller hubs or wearing rings are located on different diameters for various stages. Because such a construction would eliminate axial thrust only at the expense of reduced inter-changeability and increased manufacturing costs, this residual thrust, being relatively small, is usually carried on the thrust bearing.

Hydraulic Balancing Devices If all the single-suction impellers of a multistage pump face in the same direction, the total theoretical hydraulic axial thrust acting towards the suction end of the pump will be the sum of the individual impeller thrusts. The thrust magnitude (in pounds) will be approximately equal to the product of the net pump pressure (in pounds per square inch) and the annular unbalanced area (in square inches). Actually, the axial thrust turns out to be about 70% to 80% of this theoretical value. Some form of hydraulic balancing device must be used to balance this axial thrust and to reduce the pressure on the stuffing box adjacent to the last-stage impeller. This hydraulic balancing device may be a balancing drum, a balancing disk or a combination of the two. Balancing Drums The balancing drum is illustrated in (Fig 3.1). The balancing chamber at the back of the last stage impeller is separated from the pump interior by a drum that is either keyed or screwed to the shaft and rotates with it. The drum is separated by a small radial clearance from the stationary portion of the balancing device, called the ‘balancing drum head’ which is fixed to the pump casing.




The balancing chamber is connected either to the pump suction or to the vessel from which the pump takes its suction. Thus the back pressure in the balancing chamber is only slightly higher than the suction pressure, the difference between the two being equal to the friction losses between this chamber and the point of return. The leakage between the drum and the drum head is, of course, a function of the differential pressure across the drum and of the clearance area.

The forces acting on the balancing drum in (Fig 3.1) are the following:

1. Toward the discharge end; the discharge pressure multiplied by the front balancing area (Area B0) of the drum.

2. Toward the suction end; the back pressure in the balancing chamber multiplied by the back balancing area (Area C) of the drum.

The first force is greater than the second thereby counter-balancing the axial thrust exerted upon the single-suction impellers. The drum diameter can be selected to balance axial thrust completely or within 90% to 95% depending on the desirability of carrying any thrust-bearing loads. It has been assumed in the preceding simplified description that the pressure acting on the impeller walls is constant over their entire surface and that the axial thrust is equal to the product of the total net pressure generated and the unbalanced area. Actually, this pressure varies somewhat in the radial direction because of the centrifugal force exerted upon the water by the outer impeller shroud (Fig 2.4). Furthermore, the pressures at two corresponding points on the opposite impeller faces (D and E Fig 3.1) may not be equal because of variation in clearance between the impeller wall and the casing section separating successive stages. Finally, pressure distribution over the impeller wall surface may vary with head and capacity operating conditions. This pressure distribution and design data can be determined by test quite accurately for any one fixed operating condition and an effective balancing drum could be designed on the basis of the forces, resulting from this pressure distribution. Unfortunately, varying head and capacity conditions change the pressure distribution, and as the area of the balancing drum is necessarily fixed, the equilibrium of the axial forces can be destroyed. The objection to this is not primarily the amount of the thrust but rather that the direction of the thrust cannot be pre-determined because of the uncertainty about internal pressures. Still, it is advisable to pre-determine normal thrust direction as this can influence external mechanical thrust-bearing design. Because 100% balance is unattainable in practice, and because the slight but predictable unbalance can be carried on a thrust bearing, the balancing drum is often designed to balance only 90% to 95% of total impeller thrust. The balancing drum satisfactorily balances the axial thrust of single-suction impellers and reduces pressure on the discharge side stuffing box. It lacks, however, the virtue of automatic compensation for any changes in axial thrust caused by varying impeller reaction characteristics. In effect, if the axial thrust and balancing drum forces become unequal, the rotating element will tend to move in the direction of the greater force. The thrust bearing must then prevent excessive movement of the rotating element. The balancing drum performs no restoring function until such time as the drum force again equals the axial thrust. This automatic compensation is the major feature that differentiates the balancing disk from the balancing drum.




Balancing Disks The operation of the simple balancing disk is illustrated in (Fig 3.2). The disk is fixed to, and rotates with, the shaft. It is separated by a small axial clearance from the balancing disk head which is fixed to the casing. The leakage through this clearance flows into the balancing chamber and from there, either to the pump suction or to the vessel from which the pump takes its suction. The back of the balancing disk is subject to the balancing chamber back pressure, whereas the disk face experiences a range of pressures. These vary from discharge pressure at its smallest diameter to back pressure at its periphery. The inner and outer disk diameters are chosen so that the difference between the total force acting on the disk face and that acting on its back will balance the impeller axial thrust.

If the axial thrust of the impellers should exceed the thrust acting on the disk during operation, the latter is moved towards the disk head, reducing the axial clearance between the disk and the disk head. The amount of leakage through the clearance is reduced so that the friction losses in the leakage return line are also reduced lowering the back pressure in the balancing chamber. This lowering of pressure automatically increases the pressure difference acting on the disk and moves it away from the disk head, increasing the clearance. Now the pressure builds up in the balancing chamber and the disk is again moved towards the disk head until an equilibrium is reached. To assure proper balancing disk operation, the change in back pressure in the balancing chamber must be of an appreciable magnitude. Thus, with the balancing disk wide open with respect to the disk head, the back pressure must be substantially higher than the suction pressure to give a resultant force that restores the normal disk position. This can be accomplished by introducing a restricting orifice in the leakage return line that increases back pressure when leakage past disk increases beyond normal. The disadvantage of this arrangement is that the pressure on the stuffing box packing is variable - a condition that is injurious to the life of the packing and therefore to be avoided. The higher pressure that can occur at the packing is also undesirable.




Combination Balancing Disk and Drum For the reasons just described, the simple balancing disk is seldom used. The combination balancing disk and drum (Fig 3.3) was developed to obviate the shortcomings of the disk while retaining the advantage of automatic compensation for axial thrust changes.

The rotating portion of this balancing device consists of a long cylindrical body that turns within a drum portion of the disk head. This rotating part incorporates a disk similar to the one previously described. In this design, radial clearance remains constant regardless of disk position, whereas the axial clearance varies with the pump rotor position. The following forces act on this device:

1. Towards the discharge end; the sum of the discharge pressure multiplied by Area A, plus the average intermediate pressure multiplied by Area B.

2. Towards the suction end; the back pressure multiplied by Area C. Whereas the position-restoring feature of the simple balancing disk required an undesirably wide variation of the back pressure, it is now possible to depend upon a variation of the intermediate pressure to achieve the same effect. Here is how it works. When the pump rotor moves towards the suction end (to the left in Fig 3.3) because of increased axial thrust, the axial clearance is reduced and pressure builds up in the intermediate relief chamber, increasing the average value of the intermediate pressure acting on Area B. In other words, with reduced leakage, the pressure drop across the radial clearance decreases, increasing the pressure drop across the axial clearance. The increase in intermediate pressure forces the balancing disk towards the discharge end until equilibrium is reached. Movement of the pump rotor towards the discharge end would have the opposite effect, increasing the axial clearance and the leakage, and decreasing the intermediate pressure acting on Area B.




There are now in use numerous hydraulic balancing device modifications. One typical design separates the drum portion of a combination device into two halves, one preceding and the second following the disk (Fig 3.4). The virtue of this arrangement is a definite cushioning effect at the intermediate relief chamber thus avoiding too positing a restoring action which might result in the contacting and scoring of the disk faces.

Pump Handbook ed.Igor J. Karassik, William C. Krutzsch, Warren H. Fraser, Joseph P. Messina (USA:1986) Igor J. Karassik, Roy Carter, Centrifugal Pumps (USA:1960)


Information provided by Warren Rupp a unit of Idex Corporation Pump Clinic 13 Sandpiper Pumps Principle of Operation 28/03/07 Page 1 of 2

PUMP CLINIC 13

Sandpiper Air-operated diaphragm Pumps -Principle of Operation-

Basic Design Features Most Warren Rupp (Sandpiper) diaphragm pumps are driven by compressed air. The directional air distribution valve and pilot valve, referred to as the "air end", are located in the center section of the pump. Liquid moves through two manifolds and outer chambers of the pump, referred to as the "wet end". Generally, check valves are located at the top and bottom of each outer chamber or on a common manifold. The two outer chambers are connected by suction and discharge manifolds. The pumps are self-priming.

No-Lube Air Distribution Valve During operation, the Air Distribution Valve controls alternate pressurising of one diaphragm, then the other. The Valve automatically transfers air pressure to the opposite chamber after each stroke. This provides alternating suction and discharge strokes, as the diaphragms move in parallel paths. Warren Rupp air valves require no lubrication. This is the preferred mode of operation. Clean, dry air will enhance pump performance.

Diaphragms Flexible diaphragms are clamped at their outer perimeters, between the inner and outer chambers. The diaphragms are connected at their movable centers by a rod.

Check Valves As fluid moves through the pump, check valves open and close. This allows each outer chamber to alternately fill and discharge. The check valves respond to differential pressures. Ball-type check valves can pass very small particles.


Information provided by Warren Rupp a unit of Idex Corporation Pump Clinic 13 Sandpiper Pumps Principle of Operation 28/03/07 Page 2 of 2

The Pumping Cycle As the Air Distribution Valve directs pressurised air to the left diaphragm, the diaphragm is pushed outward. This is a discharge stroke, which forces liquid from the left outer chamber. Discharged liquid moves from the chamber, through an open discharge check valve, and exits the pump at the discharge manifold. The position of the discharge port can be top, bottom or side. As the left diaphragm is pressurised outward, the connecting rod pulls the right diaphragm inward on a suction stroke, which fills the left chamber with fluid. Liquid enters the pump at the suction manifold, moves through an open suction check valve and fills the chamber. At the end of the cycle, the Air Distribution Valve automatically shifts the air pressure to the opposite diaphragm, initiating another pumping cycle.


Information kindly reproduced with permission of the McNally Institute Pump Clinic 14 24/05/07 Page 1 of 5

PUMP CLINIC 14

PUMP AND DRIVER ALIGNMENT

In the pump business alignment means that the centerline of the pump is aligned with the centerline of the driver. Although this alignment was always a consideration with packed pumps, it is critical with sealed pumps especially if you are using rotating seal designs where the springs or bellows rotate with the shaft.

A little misalignment at the power end of the pump is a lot of misalignment at the wet end, and unfortunately that is where the seal is located in most pump applications.

Misalignment will cause many problems:

· It can cause rotating mechanical seals to move back and forth axially two times per revolution. The more the seals move the more opportunity for the lapped faces to open.

· Packing could support a misaligned shaft. A mechanical seal cannot.

· Misalignment will cause severe shaft or sleeve fretting if you use spring loaded Teflon® as a secondary seal in your mechanical seal design.

· The pump bearings can become overloaded.

· The misalignment could be severe enough to cause contact between stationary and rotating seal components.

· The wear rings can contact.

· The shaft can contact the restriction bushing often found at the end of the stuffing box.

· The shaft or sleeve can contact the stationery face of the mechanical seal.

· The shaft can contact the disaster bushing in an API (American Petroleum Institute) gland.

· The impeller could contact the volute or back plate.

Regardless of the alignment method you select, you must start with a pump and driver in good repair. A perfectly aligned piece of junk is still a piece of junk. You should also check the following:

· A straight shaft that has been dynamically balanced.

· Good wear rings with the proper clearance.

· The correct impeller to volute, or backplate clearance.

· The elimination of "soft foot".

· Eliminate all pipe strain.

· Good bearings installed on a shaft with the proper finish and tolerances.

· A good mechanical seal set at the proper face load. The closer the seal is to the pump bearings the better off you are going to be.



All pump to driver alignments consist of four parts:

· You must level the pump and driver. If the pump is aligned without being level, the oil level will be incorrect and you will develop bearing problems.

· You then take a series of radial and axial measurements to see where the pump is located in respect to its driver (motor).

· You make calculations to see how far the driver must be moved to align the centerline of the pump to the centerline of the driver. These calculations must consider that the pump and driver operating temperature will probably be very different than the ambient temperature when you are taking the readings.

· Most pump manufacturers should be able to supply you with the proper readings for a hot alignment. They are the only people that know how their unit expands and contracts with a change in temperature.

· You must now shim and move the driver to get the alignment. Most of the small pump designs are not equipped with "jack bolts" so this will be the most difficult part of the alignment procedure. You cannot move the pump because it is connected to the piping.

I see lots of pumps that have never been aligned properly. When you talk to the people that should be concerned, you get the following comments:

· Alignment is not important. I have been working with pumps for years and we never do it at this facility. And we do not do dynamic balancing of the rotating assembly either!

· There is no time to do an alignment. Production wants the unit back on line, and they will not allow me the time to do it properly.

· We purchase good couplings. The coupling manufacturer states that their coupling can take a reasonable amount of misalignment.

It turns out that there are at least three methods of getting a good pump to driver alignment, and a good coupling is not one of them. The coupling is used to transmit torque to the shaft and compensate for axial thermal growth, nothing else. You install a good coupling after you have made the pump to driver alignment, not instead of making the alignment.

Here are some acceptable methods:

The reverse indicator method is an acceptable method, but it does take a great deal of time. There are plenty of schools that teach this method if you are interested in learning how to do it:



· Very accurate especially for small diameter flanges

· Not affected by axial float.

· Can be used with a flexible coupling in place.

· You have to rotate both shafts

The laser is the latest method. It is also the most popular. There are lots of people that can teach you to use the equipment, once you have made the purchase.

The "C or D" frame adapter is probably the easiest method of all and available from most quality pump manufacturers. It solves most of the problems with thermal expansion.

You use a machined, registered fit to insure the alignment.

The shaft to coupling spool method:

· The best method when there are big distances between the shaft ends.

· A simple method to use.

· Most people rotate both shafts

Face and rim method:

· Use this method if one of the shafts cannot be rotated.

· An excellent method for large shaft diameters (8 inches or 200mm or greater) or if the diameters are equal to, or greater than the span from the bracket location to the face and rim location where the readings are to be taken.

· Not too good a method if there is axial float from sleeve or journal bearings.



Given a choice I would select the C or D frame every time:

· The "C frame" is for inch sizes The "D frame" for metric sizes.

· Automotive people use the same concept to align an automobile transmission to the engine. They call the adapter a "bell housing".

· The concept was originally developed for the marine industry where it would be impossible to bolt the motor and pump to the deck of the ship, and then do an alignment. The hull flexes making any conventional alignment ineffective. The same logic applies to off shore drilling rigs.

· The adapter does a better job of equalizing the heat transfer between the pump and the driver. It does not all have to conduct through the shaft.

· The adapter is available for all quality end suction centrifugal pumps. Check with your supplier for the availability of one for your pump.

· When given a choice, select a ductile rather than a cast adapter.

· Up to about thirty-horse power (22 KW) you hang the motor on the pump. Above thirty-horse power (22 KW) you hang the pump on the motor.

· The adapter solves the problem of "there is no time to do an alignment".

· If your motor does not have a "C or D" end bell, one can be installed when the motor is rewound. Some, but not all explosion-proof motors are available with a C or D frame end bell. Check with your supplier.

If you do not have a C or D frame adapter you will be involved in the last three steps of the four-step procedure.

Moving the pump driver:

Once you have made all the measurements, put in the recommended compensation for thermal expansion, and figured out all the calculations for how much to move the driver, and in which direction; now comes the fun part; moving the driver.

You can hit the motor with a big hammer, but small dimensions are hard to get with this method.

Some people use an adjusting wheel that attaches to shims. This will give you a very precise movement that is necessary for a proper alignment



Another method is to use an adjusting wheel that slips over the motor hold down bolts. Many mechanics make there own tools and these units also work very well for precise motor movement.

How concerned should you be about alignment? You do it on your automobile when you notice uneven tyre wear or the car drifts to one side of the road when you loosen your grip on the wheel, and have no problem justifying the cost and time involved. It is the same logic you use towards the added cost and time spent balancing the tyres and wheels of your car.

We do not always apply the same logic to our very expensive rotating equipment in the shop, but we should. A mechanical seal should run trouble-free until the carbon sacrificial face has worn away.

When we inspect the seals we remove from leaking pumps we find that in better than 85% of the cases there is plenty of carbon face left on the seals. The seals are leaking prematurely and the seal movement caused by pump to motor misalignment is a major contributing factor.


Information reproduced with permission of author, Bill McNally, the McNally Institute Pump Clinic 15 20/06/07 Page 1 of 5

PUMP CLINIC 15

MECHANICAL SEAL DESIGN, OPERATION AND MAINTENANCE PROBLEMS

“In my seminars I teach that mechanical seals fail prematurely because:

· The lapped faces open

· A seal component becomes damaged

In the following paragraphs we will learn how these failures can be separated into: · Design problems

· Operation problems

· Maintenance problems

The purpose of this paper is to give you an overview of the subject, and assist you in your troubleshooting function.

MECHANICAL SEAL DESIGN PROBLEMS

Problems with the Seal Faces: · Wrong carbon or hard face selected. The material is not compatible with the fluid you are sealing, and

the cleaner or solvent used to clean or flush the system

Face flatness problems:

· The face cross-section is too narrow causing temperature or pressure distortion problems

· The material modulus of elasticity is too low

· The face is not hard enough

· All clamping forces must be "equal and opposite" to prevent face distortion. In many designs they are not

· The differential expansion between the seal face and its holder can cause the face to go out of flat

· The faces were not lapped at a cryogenic temperature and the seal is being specified for cryogenic service

· Bad packaging

Poor heat conductivity

· Carbon is a poor conductor of heat compared to most hard faces

· Many ceramics are not good conductors of heat

· Plated or coated faces can "heat check" due to a differential expansion rate between the coating and the base material

· The seal face is sometimes insulated by a gasket or elastomer

· Low expansion steel face holders are not usually corrosion resistant

· No vibration damping has been provided to prevent "slip stick" vibration problems. This is a major problem with metal bellows seals

· Unbalanced seal designs require excessive flushing or cooling to remove unwanted heat



· The carbon must be dense enough to prevent entrained air pockets from expanding and causing pits in the carbon face. An "unfilled carbon" with four impregnates is the best

The Springs or bellows: · Springs in the fluid can clog easily, especially the small springs

· Stainless steel springs and bellows are sensitive to chloride stress corrosion problems

· A single spring can be wound in the wrong direction

· Thin bellows plates and small cross section springs are sensitive to abrasive wear

· Rubber bellows experience a catastrophic failure mode when the bellows ruptures

· Stressed metal corrodes faster. Springs and metal bellows are subjected to high stress

· Too much spring or bellows movement will cause an early fatigue of the metal

The Dynamic Elastomer (the one that moves): · Some elastomers do not move to a clean surface as the face wears

· Spring loaded elastomers stick to the shaft or sleeve and are sensitive to the shaft diameter and finish

· Elastomers positioned in the seal face are subject to the heat generated between the seal faces

· Dynamic elastomers are very sensitive to the shaft tolerance and finish

Operating conditions too severe for the design: · Elastomers and some seal faces are sensitive to temperature extremes

· Excessive pressure can distort seal faces causing them to go out of flat

· Excessive pressure can cause elastomer extrusion

· High speed can separate the seal faces in rotating seal designs

· High speed can cause excessive heat at the seal faces

· Excessive shaft movement separates faces also

· Hard vacuum can "out gas" an elastomer causing it to leak

Dual seals: · Rotating "back to back" designs

· Centrifugal force throws solids into the inner faces

· Inner seal blows open if barrier fluid pressure is lost

· Inner stationary face is not positively retained to prevent movement if the pressure is lost between the faces

· When the outboard seal fails the inboard will fail also due to the pressure drop between the faces

· The inner seal has to move into the sealing fluid as the face wears. This is a major problem if the fluid contains solids

· Failure to use "two way" hydraulic balance causes the inner faces to open with a reversal in barrier fluid pressure

Design problems that cause excessive shaft movement: · An elbow is installed too close to the pump suction inlet

· The mass of the foundation is not five times the mass of the pump and its driver

· Wrong size pump was specified because of safety factors and, as a result, the pump is operating off the B.E.P



· The pump was selected oversize in anticipation of a future need

· A "centerline" design should have been selected when the operating temperate exceeded 200°F (100°C)

· The shaft L3/D4 is too high

The pump is cavitating due to a design problem: · Too high a N.P.S.H. is required. You need a double suction pump

· The suction specific speed number is too high

· You are using too low a specific speed impeller

· A reducer has been installed up side down, letting an air pocket into the suction

· The impeller to cutwater clearance is too low

· There is too much suction resistance due to excessive piping

· Too much suction lift for the fluid temperature

Other design problems: · Some seal designs cannot compensate for thermal shaft growth or impeller adjustment. Cartridge

versions are needed for this feature

· The pumping fluid is located at the inside diameter of the seal faces

· Solids will be thrown into the lapped faces destroying some face materials

· Solids will pile up in front of the movable faces, preventing them from compensating for wear

· Most seal faces are weak in tension

· Hysteresis (delay) problems caused by the seal mass and sliding elastomers

· Poor packaging that allows face damage during shipment and storage

· Designs that frett (damage or groove) the shaft or sleeve

· High speed requires the use of stationary seal designs. Centrifugal force can open rotating designs above 5000 fpm. (25 m/sec.)

· The seal is positioned too far from the bearing housing

· Lack of a self-aligning feature is causing excessive face movement

· A tapered stuffing box can cause face damage

· No vent has been provided to vent the stuffing box in a vertical application

· Hardened shafts and sleeves can cause the seal set screws to slip

· A discharge recirculation line is aimed at the lapped faces, causing them to wear, and interfering with the seal movement

Problems caused by the product you are sealing: · The fluid can flash or vaporize between the faces

· Viscous fluids open seal faces as they restrict seal movement

· Products that solidify will open and damage seal faces

· Crystallizing products restrict seal movement and open the faces

· Film building products cause the faces to open. Hot oil is typical

· The fluid can attack one of the seal components, especially the elastomer

· All chemicals have the potential for corroding a seal component. It is just a mater of time

Some fluids are poor lubricants:



· This can cause excessive wear

· Color contamination problems as the carbon wears

· "Slip stick" vibration problems

· Slurries clog up the sliding seal components and open the faces

· Cryogenic fluids can attack some carbon faces and most elastomers

· High temperature fluids attack elastomers and change the state of the fluid you are sealing

· Some fluids can cause the formation of ice outboard the seal, restricting seal movement as the face wears

· Agitation can cause some fluids to change their viscosity

· Cleaners or solvents are attacking a seal component

OPERATION PROBLEMS

Operations that cause excessive shaft movement that will open or damage the seal faces: · Opening and closing valves in the suction and/or discharge causing the pump to operate off the B.E.P,

and the shaft to deflect

· Pumping the supply tank dry, causing excessive vibration and heat

· Series or parallel pump operation can cause shaft deflection

· Running at a critical speed will cause the shaft to defect

Cavitation problems:

· Low N.P.S.H

· Air getting into the system through packing

· A stuffing box, suction recirculation line is heating the incoming fluid

· A discharge bypass line is heating the suction fluid

· A discharge recirculation line is aimed at the seal face restricting its movement

· Water hammer is opening or damaging the lapped faces

· The piping system has been altered since the pump was installed

· The pump is being started with the discharge valve shut or severely throttled

· Starting a pump with the discharge valve open is just as bad

Operations that cause excessive heat and corrosion problems: · Cleaners or solvents used in the lines can attack a seal component, especially the elastomer

· A product concentration change will affect corrosion

· A change of product

· Either a temperature or pressure change in the system will affect both

Operations that cause the seal faces to open: · The seal is seeing frequent reversing pressures

· Loss or lack of an environmental control

· Flush not working

· Quench is shut off

· Barrier fluid not circulating



· Loss of heating or cooling

· Heating jacket clogged

· Pressure drop in the stuffing box

· Flushing with a dirty product

· Quenching with shop water leaves solids outboard of the seal that will cause a hang-up as the seal moves forward to compensate for wear

· The quenching steam pressure is too high. It is getting into the bearings

MAINTENANCE PROBLEMS · The pump and driver are not aligned & emdash; causing excessive seal movement

· Pipe strain

· Thermal growth

· Bad installation techniques that can injure a seal component

· The wrong lubricant was put on the dynamic elastomer

· The impeller clearance was set after the seal installation

· The face is inserted backwards, only one side is lapped

· The seal is set at the wrong installation length

· The sleeve moved when the impeller was tightened to the shaft

· A lubricant was put on the seal face that froze when the product evaporated across the lapped faces

· The rotating assembly is not dynamically balanced

· The shaft is bent

· The sleeve is not concentric to the shaft

· Impeller clearance is not being maintained, causing vibration problems

· The impeller is positioned too close to the cutwater

· The seal has been set screwed to a hardened shaft

· No seal or gasket between the shaft sleeve and the solid shaft. This is a big problem with double ended pumps

· The seal environmental control is not being maintained

· Flushing fluid is being restricted or shut off

· Quenching steam is shut off

· The barrier fluid tank level is too low

· The convection tank is running backwards

· The cooling jacket is restricted due to a calcium build up

· You are running both a discharge recirculation line and a cooling jacket

· Out of tolerance shaft dimensions will restrict seal movement

· The impeller clearance was made without re-adjusting the seal face load

· The shaft sleeve was removed to accommodate a smaller diameter seal. The sleeve was providing corrosion resistance

· A gasket is protruding into the stuffing box restricting the seal movement.“



PUMP CLINIC 16

AN OVERVIEW OF SEAL TROUBLESHOOTING

“Seal problems are almost always associated with face leakage, but as we will soon learn, there are other leak paths in addition to the obvious one between the lapped seal faces.

In the following paragraphs, we'll be looking at all these leak paths. Keep in mind that seals are classified into many categories: stationary, rotary, balanced, unbalanced, inside, outside, metallic, non-metallic, single, dual, elastomer, metal bellows, rubber bellows, cartridge, split, solid, etc. Try to keep these classifications in mind as we investigate the cause of seal failure.

I will be presenting the troubleshooting hints in an outline form. You should not find these terms confusing because I've assumed you have a pretty good knowledge of mechanical seals or otherwise you wouldn't be attempting to troubleshoot them.

LEAKAGE AT THE SEAL FACES

The seal face is not flat. (Flatness should be measured within three helium light bands, (0,000033" or 1 micron)

• The face was damaged by mishandling.

• Poor packaging. The seal should be able to survive a 39" (1 metre) drop. To ensure this, the seal must be shipped in a reusable box insulated with plenty of foam or any other adequate insulation.

• The face was distorted by high pressure or surges in pressure. "Water hammer" would be an example.

• The face was distorted when you tightened it against an uneven surface.

• The clamping is not "equal and opposite" across the stationary hard face. This is a common problem with "L" shaped and "T" shaped stationary faces.

• The "hard" seal face has been installed backwards. You're running on a non-lapped seal face. It is common practice to lap only one side of a hard face.

• The face is being distorted by a change in temperature. This happens when you forget to vent a vertical pump.

• The face never was flat. You have a bad part.

• The carbon metal composite was not stress-relieved after the carbon was "pressed in".

The face has been chemically attacked.

• Oxidizing agents attack all forms and grades of carbon graphite.

• Some de-ionized water will attack any form of carbon.

• Corrosion increases with a temperature increase. A 10°Centigrade (18°F) rise in temperature will double the corrosion rate of most corrosives.

• A cleaner or solvent is being flushed through the lines and is attacking the carbon.



• You are using a poor grade of Carbon. Go to an unfilled grade such as Pure Carbon Company grade 658 RC. This is a common occurrence if the seal is being repaired by someone other than the original manufacturer.

The plating or hard coating is coming off the hard face.

• All coatings are porous. The chemical is penetrating this porous coating and attacking the bond between the coating and the base material, or the base material itself.

• An inferior plating was originally put on the base material.

• Differential expansion of the dissimilar materials is causing them to separate.

The seal face is cracked, pitted or damaged.

• High temperature is heat checking (cracking) the plated face. This is a common problem with cobalt based tungsten carbide. The nickel base version is less likely to crack.

• The product is solidifying between the faces and they're breaking at start up. Most face materials have high compressive strength, but tend to be weak in tension.

• Excessive vibration is causing the drive pins to crack the face. Low cost seals experience this problem quite often.

• There is a high temperature differential across the ceramic, 7 to 10 cycles can break even good ceramics in hot water or hot petroleum products.

• Air is trapped in the carbon face. Heat is causing it to expand and blow out pieces of the carbon face. The carbon usually blisters prior to blowing out. The solution is to go to a more dense carbon.

• The product is vaporizing and allowing solid material to blow across the lapped face. This is a common occurrence in boiler feed water applications.

• The seal faces have opened, solids penetrated and imbedded into the soft carbon are causing rapid wear in the hard face. The same problem occurs if the carbon was re-lapped using lapping powder.

• Lubricant on the faces is freezing in cryogenic (cold) applications.

• The elastomer is being chemically attacked and swelling up. This can break the face in those seal applications where the elastomer is positioned in the seal inside diameter. In some instances the swelling elastomer will open up the two faces, allowing the solids to penetrate. This can be a problem with boot mounted faces

• The rotating shaft, or sleeve, is hitting the stationary face. This can happen if the pump is running off of its B.E.P. which almost always occurs at start up.

• The seal is being mishandled during installation. Good packaging and proper training can solve many of these problems.

• The crack may have occurred during disassembly. Check to see if there is discoloration deep in the crack. Discoloration means that it occurred during, or before, operation.

• Petroleum products can "coke" at the face causing pieces of carbon to be pulled out as the face rotates. You will have to select two hard faces for this application.

• The rotating face is not centered in the stationary face and is running off the edge of the stationary face. Look for rubbing marks around the O.D. of the rotary unit. A bent shaft or out of balance rotating assembly is the most common cause.

o You will notice a much wider wear track if you are experiencing this problem.



o The seal will appear to "spit" as lubricant is dragged across the face and off the seal outside diameter.

o Dirt can be dragged across the faces as they separate.

The movable face is not free to follow whip, wobble or run out.

• The rotating face is hitting the I.D. of the stuffing box.

• The recirculation line from the pump discharge is aimed at the seal faces and interfering with their free movement.

• Dirt or solids are clogging the movable components. Magnetite is a very big problem in most hot water applications.

• The product is interfering with the free movement of the components. It is:

o Crystallizing ( like sugar)

o Solidifying (like glue)

o Viscous (molasses)

o Building a film on the sliding components (hard water or paint)

o Coking (oil or any other petroleum product)

• The elastomer has been chemically attacked causing it to swell up and interfere with free movement of the face.

• Temperature growth of the shaft is interfering with the free movement of the movable face.

• The shaft or sleeve is the problem.

o It is over size - + 0.00" - 0.002" (0,00-0,05 mm.) is ideal.

o It is too rough; it should be at least 32 R.M.S. (0,8 microns)

o It is fretted, corroded or damaged in some way.

o Solids have attached themselves to that portion of the shaft where the dynamic elastomer is located.

• A gasket or fitting is protruding into the stuffing box.

• Solids from outside the stuffing box are getting under the faces. This is a common problem with vertical pumps.

• The elastomer is spring loaded and the interference on the shaft is restricting the face movement.

• The elastomer has extruded because of high pressure or excessive clearance.

• A foreign object has passed into the seal chamber and is interfering with the free movement of the seal.

The product has plated, or formed on the face and a piece of it has broken off.

• This problem occurs with products that are sensitive to temperature and/ or pressure changes.

The set screws have come loose.



• The shaft has been hardened.

• They have worked loose in a sleeve that is too soft.

• The hardened set screws have corroded.

• They were not replaced when the seal was rebuilt and as a result are not "digging" into the shaft.

The face has lost its spring load.

• The initial setting was wrong.

• Temperature growth of the shaft has altered the original setting.

• The impeller has been adjusted towards the wet end of the pump.

• The sleeve moved when the impeller was tightened to the shaft.

• The cartridge seal was pushed on the shaft by pushing on the gland and the seal is now over compressed.

o In a dual seal application this will over compress the inner seal and open up, or unload the outer seal.

The product is vaporizing and blowing the faces open. This happens in hot applications if there is water in the product.

• It can also occur if the pump/seal was hydrostatically tested with a water base fluid.

The inner seal, of a dual seal application was not balanced in both directions and is opening up with reversing pressure. This is a common problem in unbalanced seals that are subject to both vacuum and pressure or if the barrier fluid pressure varies.

The single spring, found in some seal designs, was wound in the wrong direction for the shaft rotation.

The Bellows seal has lost cooling and the anti vibration lugs are engaging the shaft. Shaft movement will cause the faces to open.

LEAKAGE AT THE ELASTOMER LOCATION

Compression set (the elastomer has changed shape).

• Either the product is too hot or there is too much heat being generated at the seal faces. You must vent vertical pumps to prevent this problem.

The elastomer is cracked.

• The shelf life has been exceeded. Buna N (Nitrile) has a shelf life of only twelve months because of its sensitivity to ozone attack.

• High heat is the main cause.

• Chemical attack. In most cases the elastomer swells, but cracking and shrinking does occur in isolated cases.

• Cryogenic (cold) temperatures freeze the elastomer and it will crack when hit.

• The rubber bellows did not stick to the shaft because the wrong lubricant was used. The shaft turned inside the bellows causing high heat.



• The seal faces stuck together. The shaft was turning inside the rubber bellows causing excessive heat.

The elastomer is cut or damaged.

• Mishandling.

• The elastomer was slid over a rough spot on the shaft or sleeve. Be careful of old set screw marks, splined shafts, key ways, etc.

• It was extruded by high pressure. You may need a backup ring.

• The product is penetrating into the elastomer and blowing out the other side. This problem is a common occurrence when you are trying to seal ethylene oxide.

• Teflon jacketed o-rings can split in the presence of halogenated fluids. The halogen will cause the elastomer to swell up, inside of the teflon jacket. Halogens can be recognized because most of them end in the letters "ine", such as bromine, astintine, chlorine, fluorine, iodine, etc..

The elastomer is not seated properly.

• It was twisted during installation. O-rings groove.

• Solids have "built up" or penetrated between the elastomer and the shaft.

• The shaft is corroded, damaged, or fretted.

• The shaft is oversized.

• Excessive travel can cause the elastomer to "snake". Most o-rings can roll up to one half of their diameter.

• The o-rings groove is damaged or coated with a solid material.

The elastomer has swollen or changed color.

• Product attack. This is the most common cause and usually occurs within five to ten days

• The wrong lubricant was used at installation. As an example, you should never put petroleum grease on EPR o-rings.

• Solvents or chemicals used to clean the lines are not compatible with the elastomer.

• Steam can harm many elastomers including most grades of Viton®.

• Oxidizers can attack the carbon black in o-rings and other elastomers.

The elastomer leaks when pressurized in the opposite direction.

• A common problem with unbalanced, dual seal applications. Two-way balanced seals are recommended for these applications.

• Remember that o-rings are the only elastomers that seal in both directions. Wedges, U cups, and chevrons do not have this ability.

OTHER LEAK PATHS TO CONSIDER

Between the carbon and its metal holder.



• Some seal companies, and most seal repair facilities, glue the carbon in place. The glue may not be compatible with the product you're sealing.

• "Pressed in" carbons can leak in a high temperature application because of the differential expansion between the carbon and its metal retainer. Low expansion metal is available for these applications

Between the shaft and the sleeve.

• Damaged gasket or gasket surface.

• Distorted sleeve or shaft. Many packed, double ended pumps have this problem because there's no gasket between the impeller and the sleeve that's holding it in place.

Stationary face gasket or elastomer leaking.

• This leak path isn't always visible. It often looks like face leakage.

Gland gasket or gasket surface leakage.

• This leak path should always be visible.

Pipe flange leaking above the seal and dripping into the seal area.

• I found this one after every other troubleshooting avenue was exhausted.

At the weld location if a seal face holder is welded to the cartridge sleeve.

At the pipe connections, ancillary hardware, A.P.I. Gland fittings, and recirculation lines.

A scratch or nick in the o-ring groove. Remember that up to 100 psi (6 bar) o-rings seal on the O.D. and the I.D. not the sides.

Seal faces will not leak visibly if they are lapped flat and we keep them in total contact. Shaft movement is the main contributor to the opening of the seal faces and allowing solids to penetrate. Shaft movement is caused by many factors. In the following paragraphs we'll be looking at most of them.

CAUSES OF EXCESSIVE SHAFT MOVEMENT, INCLUDING VIBRATION

Cavitation

• Vaporization caused by too high a product temperature, or too low a suction head.

• Air is entering the stuffing box. A common problem with pumps that run in a vacuum or taking a suction from an evaporator or condenser.

• Internal recirculation. Occurs when the Suction Specific Speed is too high, or when either the impeller or wear ring clearance becomes excessive.

• The vane passing syndrome form of cavitation occurs if the O.D. of the impeller is too close to the pump cutwater. This clearance should be at least 4% of the impeller diameter in the smaller size impellers and at least 6% in the larger diameter impellers (greater than 14 inch or 355 mm.)

• Turbulence. Occurs if there's not laminar flow in the lines.

The bearings are worn excessively.

• Contamination of the lubricant is the biggest cause. Grease or lip seals have a useful life of only 2000 hours (84 days).



• Poor fit or installation.

• Serious misalignment. The misalignment can be the result of pipe strain, or misalignment between the pump and its driver.

The shaft is bent.

• Usually occurs during sleeve removal, or if the bearing was installed with an arbor press.

• Improper storage with the long shaft supported only on the ends causing it to sag.

• Heating the shaft to remove the sleeve is another common cause

The impeller is out of balance.

• The impeller was damaged by either wear or corrosion or cavitation.

• Product has built up on the vanes or in the balance holes.

• The impeller diameter was reduced and the impeller was not re-balanced

• The impeller never was balanced.

An unbalanced rotating assembly.

Pressure surges or water hammer.

Worn coupling.

The pump is operating off its best efficiency point.

Rubbing of a rotating component.

• The shaft is hitting the wear ring, or a stationary wear ring is contacting a rotating wear ring.

• The shaft is hitting the seal gland or stationary face.

• A seal rotating component is hitting the stuffing box I.D..

• A gasket or fitting is protruding into the stuffing box.

The stationary seal face is not perpendicular to the rotating shaft. This causes the spring loaded, rotating face to move back and forth twice per revolution.

• The stuffing box face is not square to the shaft. The stuffing box face is often a rough casting.

• Tightening the gland bolts through a gasket is cocking the stationary face.

• Pipe strain.

• Temperature growth.

• A convection tank, or some other heavy device is hanging off the gland distorting it.

• Bearing fit or wear.

• Coupling alignment.

• Shaft deflection. The deflection can be caused by operating the pump off its best efficiency point, the rotating assembly is out of balance, or the shaft is bent.



• Poor installation technique.

VIBRATION AT THE SEAL FACES

Harmonic vibration.

• The seal is vibrating in harmony with some rotating component. The same thing that causes a rear view mirror to vibrate in an automobile. Most harmonic vibration can be stopped by changing the speed of the equipment or "damping" the vibrating component.

Slipstick (an alternating slipping and sticking of the seal faces,) caused by:

• Poor lubricating fluids.

• Hot water.

• Solvents.

• Some detergents.

• Gases

• Dry running applications.

• Too high a face load.

o You are using unbalanced seals.

• Poor installation technique.

• Face load has changed because of temperature growth, or impeller adjustment.

• You are using a high friction face combination. Often occurs if you use two hard faces.

A discharge recirculation line aimed at the seal faces.

• Each time the impeller passes the recirculation connection it causes a pulse of fluid at the seal face.

Vaporization of the product at the seal face.

• Happens with products that contain water, and are operated at elevated temperature.

• Can occur at the seal face because of high face load caused by using unbalanced seals.

EXCESSIVE AXIAL MOVEMENT OF THE SEAL

• Temperature growth.

• The impeller was adjusted, after the seal was installed, to compensate for wear.

• The rotor motor, moved to its magnetic centre at start up.

• The equipment is equipped with sleeve or babbitted bearings and has excessive end play.

• Shaft thrust.



o There is a thrust towards the bearings caused by the combination of the fluid changing direction in the impeller and acting on the shaft and/or impeller surfaces. This thrust is offset by a thrust towards the wet end caused by the impeller shape.

o In centrifugal pumps the resulting force can be in either direction, depending upon how close the pump is operating to its best efficiency point. Above 65% of its best efficiency, the thrust is towards the wet end. Below 65% of the best efficiency the thrust is towards the power or bearing end. There is little to no movement at 65% of the pump’s best efficiency. This means that at start up the shaft moves in both directions accounting for a higher percentage of seal failure at start up.

• Vertical mixer shafts often lift vertically when solids are mixed with liquid.

THE SHAFT IS NOT CONCENTRIC WITH THE STUFFING BOX, this will cause a wiping action in stationary seals.

• The shaft is bending as you move away from the pump B.E.P.

o It bends at 240 degrees, from the cutwater, at low flow and high head.

o It bends at 60 degrees, from the cutwater, at high flow and low head.

• Coupling misalignment.

• Poor bearing fit.

• Pipe strain.

• Temperature growth causes the stuffing box to move relative to the shaft.

• The sleeve is not concentric with the shaft.

• The seal is not concentric with the sleeve / shaft.

• A bolted on stuffing box has slipped.

• The back plate is not machined concentric to the stuffing box.

Heat is always an indication of wasted energy, but it can also have a disastrous affect on seal life and performance. Let's take a look at what's causing this heat.

CAUSES OF HIGH HEAT AT THE SEAL FACES.

Too much spring compression.

• Installation error.

• No print was used, or the mechanic cannot read the print he was given.

• The shaft installation reference was marked in the wrong location.

• The mechanic used the wrong marking tool. The mark is too wide.

• The sleeve moved when the impeller was tightened.

• The impeller was adjusted after the seal was installed.



• A cartridge seal was installed on the shaft, by pushing on the gland. Interference from the sleeve elastomer has caused an over compression of the seal. In some dual seal applications the outer seal will become under compressed.

• The shaft moved because of thrust.

• Thermal growth of the shaft.

Problems with some seal designs.

• Unbalanced seals are supplied by original equipment companies. They generate more heat than balanced seals.

• The elastomer is located too close to the seal faces. The heat generated at the faces is affecting both the elastomer and the seal face.

• The carbon face is insulated by an elastomer.

• The face is too wide causing the hydraulic force to generate excessive heat.

• The carbon seal face is too narrow causing excessive heat from the spring pressure.

• A vertical seal installation is not being vented. The faces are running dry in a bubble.

• Speeds above 5000 F.P.M. (25 m/sec) require a special hydraulic balance and less spring load. A 60/40 balance and a face load of 8 psi to 15 psi (0,07 to 0,2 n/mm2) would be normal.

• An outside metal or elastomer bellows seal is almost impossible to vent.

• Spring loaded elastomers cause varying seal face loads. The actual load depends upon shaft tolerance and installation dimension.

• Some seal faces are glued in. The glue acts as an insulator preventing the face heat from conducting to the metal holder.

• Many single spring designs are uni-directional requiring both right handed and left handed seals on a double ended pump.

• Many metal bellows designs lack effective vibration damping.

• Stationary seal designs require clean flushing if solids are present. Centrifugal force does not throw the solids away from the moveable (spring loaded) components.

Problems with face materials.

• Heat conductivity is low in some materials. (Ceramic, carbon, Teflon)

• The coefficient of friction varies with face combinations and various sealing products.

• Carbon / metal composite faces conduct heat better than plain carbon / graphite, as long as there is a true interference fit and they're not glued together to hold them in place.

Problems with the pump operation that causes high heat at the faces.

• Operating off the B.E.P

• The degree of the problem is determined by the L3/D4 ratio.

• Operating too close to the vapor point, causing cavitation.



• Running dry.

• Gases.

• Dry solids.

• Pumping a tank dry.

• Losing barrier fluid in a dual seal application.

• Shutting off the flushing water.

• Vacuum applications.

• Vertical pumps not vented in the stuffing box.

• The liquid is not a lubricant.

• Pump out rings on the back of the impeller, running too close to the pump back plate.

Other causes of high heat.

• The shaft, or sleeve is rubbing a stationary component.

o The gland.

o The bushing in the bottom of the stuffing box.

o The bushing in the A.P.I. gland.

o A pump wear ring.

o A protruding gasket.

o A fitting.

o The stationary portion of a mechanical seal.

• The shaft, or sleeve, is not straight.

o It is bending, because the pump is operating off its best efficiency point.

o It is bent. This often happens when the sleeve is removed.

o The rotating assembly is not balanced.

o The shaft never was straight.

• There is not enough circulation around the seal.

o Install a large diameter stuffing box. You should be able to get at least 1" (25 mm) all around the rotating unit.

o Connect a recirculation line from the bottom of the stuffing box to the suction side of the pump. You can do this in almost every case except when you're pumping a product at its vapor point, or if the solids have a specific gravity lower than the fluid.

• The cooling jacket is clogged.

• There is no carbon restriction bushing in the bottom of the stuffing box and you are using the cooling jacket. The restriction bushing slows down the heat transfer.



• Loss of an environmental control.

o The flush is not constant. The pressure is changing.

o Quenching steam or water has been shut off during pump shut down.

o The double seal barrier fluid is not circulating.

o The cooling jacket has become clogged by the calcium in the hard water. Try condensate instead.

• The filter, or separator, is clogged.

• Either the suction or discharge recirculation line is clogged.

• If you are using double seals, remember that two seals generate twice as much heat and conventional cooling may not be sufficient. Contact the manufacturer for the rules when using convection tanks and dual seals. You may need a "built in" pumping ring.

• Solids in the stuffing box are interfering with a rotating component.”



PUMP CLINIC 17

UNDERSTANDING THE PUMP SYSTEM CURVE

“All pump manufacturers would like to recommend the perfect pump for your application. To do this they would like you to provide an accurate system curve that would describe the capacity and head needed for your various operating conditions. Once they have your system curve, they can plot pump curves on top of the system curve and hopefully select something that will come close to your needs. Without this system curve, neither of you has much chance of coming up with the right pump.

To create a system curve we plot the desired capacities against the required head over the total anticipated operating range of the pump. The head will be measured in feet or meters and the capacity will be measured in gallons per minute or cubic meters per hour.

Some of the confusion begins when we realise that there are three different kinds of head:

STATIC HEAD This is the vertical distance measured from the centre line of the pump to the height of the piping discharge inside the tank. Look at figure "A" and note that the piping discharge is below the maximum elevation of the piping system. We do not use the maximum elevation in our calculations because the siphoning action will carry the fluid over this point once the piping is full of liquid. This is the same action that lets you siphon petrol out of a vehicle into a storage can.

The pump will have to develop enough head to fill the pipe and then the siphoning action will take over. The pump operating point should move back towards the best efficiency point (B.E.P.) if the pump was selected correctly.

FIGURE "A"

DYNAMIC OR SYSTEM HEAD As the liquid flows through the piping and fittings, it is subject to the friction caused by the piping inside finish, restricted passages in the fittings and hardware that has been installed in the system. The resulting "pressure drop" is described as a "loss of head" in the system, and can be calculated from graphs and charts provided by the pump and piping manufacturers. These charts are not included with this article, you can find them in the Hydraulic Institute Manuals. This "head" loss is related to the condition of the system and makes the calculations difficult when you realise that older systems may have



"product build up" on the piping walls, filters, strainers, valves, elbows, heat exchangers, etc., making the published numbers somewhat inaccurate. A general "rule of thumb" says that the friction loss in clean piping will vary approximately with 90% of the square of the change in flow in the piping, and 100% of the square with the change of flow in the fittings and accessories. You calculate the change in flow by dividing the new flow by the old flow and then square the number. As an example:

At 45m³/hr the piping resistance, calculated from published charts (not included) is twenty-three metres (23m). What will it be at 67.5m³/hr? 67.5 = (1.5)² = 2.25 x 23m = 51.75 x 90% of the change = 46.58m of resistance head 45 In other words, when we went from 45m³/hr to 67.5m³/hr the piping resistance increased from 23m to 46.58m The loss through the fittings and hardware was calculated at 7.6m. What will the new loss be? 67.5 = (1.5) ² = 2.25 x 7.6m = 17.1 x 100% of the change = 17.1 new metres of head 45

In the original application system, loss was a combination of the loss through the piping and the loss through the fittings for a total of 30.6 metres at 45m³/hr. When we increased the flow to 67.5m³/hr our system head changed to a total of 63.68m (46.58 + 17.1). This change would have to be added to the static and pressure heads to calculate the total head required for the new pump.

Please note that the pump is pumping the difference between the suction head and the discharge head, so if you fail to consider that the suction head will be either added to or subtracted from the discharge head, you will make an error in your calculations. The suction head will be negative if you are lifting liquid from below ground or if you are pumping from a vacuum. It will be positive if you are pumping from a tank located above ground. If the suction head is pressurized, this pressure must be converted to head and subtracted from the total head required by the pump.

A centrifugal pump will create a head/capacity curve that will generally resemble one of the curves described in figure "B" The shape of the curve is determined by the Specific Speed number of the impeller.

Centrifugal pumps always pump somewhere on their curve, but should be selected to pump as close to the best efficiency point (B.E.P.) as possible. The B.E.P. will fall some where between 80% and 85% of the shut off head (maximum head).

The manufacturer generated these curves at a specific R.P.M.. Unless you are using synchronous motors (you probably are using induction motors on your pumps) you will have to adjust the curves to match your actual pump speed. Put a tachometer on the running motor and record the rpm. difference between your pump and the speed shown on the pump manufacturer's published curve. You can use the pump affinity laws to approximate the change.



POSITIVE DISPLACEMENT PUMPS have a different shaped curve. They look something like Figure "C".

In this system, the head remains a constant as the capacity varies. This is a typical application for:

• A boiler feed pump that is supplying a constant pressure boiler with a varying steam demand. This is a very common application in many process systems or aboard a ship that is frequently changing speeds (answering bells).

• Filling a tank from the top and varying the amount of liquid being pumped, is the normal routine in most process plants. The curve will look like this if the majority of the head is either static or pressure head.



The second system is the ideal one, Figure "E" describes it:

In this system the entire head is system head so it will vary with the capacity. Look for this type of curve in the following applications:

• A circulating hot or cold water heating/ cooling system.

• Pumping to a non pressurized tank, a long distance from the source with little to no elevation involved. Filling tank cars is a typical application.



System curve "G" is a common one. It is a combination of static, pressure and system heads.

Once the pump manufacturer has a clear idea as to the shape of your system curve, and the head and capacity numbers needed he can then select the proper centrifugal pump. The shape of his curve will be pretty much determined by the specific speed number of the impeller.

In addition to specific speed he can select impeller diameter, impeller width, pump rpm., and he also has the option of series or parallel operation along with the possibility of using a multi-stage pump to satisfy your needs.

The sad fact is that most pumps are selected poorly because of the desire to offer the customer the lowest possible price. A robust pump, with a low L3/D4, is still your best protection against seal and bearing premature failure when the pump is operating off of its best efficiency point. Keep the following in mind as you select your pump:

• A centrifugal pump will pump where the pump curve intersects the system curve. This may bear no relationship to the best efficiency point (B.E.P.), or your desire for the pump to perform a specific task.

• The further off the B.E.P. you go, the more robust the pump you will need. This is especially true if you have replaced the packing with a mechanical seal and no longer have the packing to act as a support bearing when the shaft deflects. Shaft deflection is always a major problem at start up.

• When you connect pumps in parallel, you add the capacities together. The capacity of a pump is determined by the impeller width and r.p.m. The head of a centrifugal pump is determined by the impeller diameter and rpm. If the heads are different, the stronger pump will throttle the weaker one, so the impeller diameters and rpm's must be the same if you connect pumps in parallel. Check the rpm's on these pumps if you are experiencing any difficulties.



• If you connect the pumps in series, the heads will add together, so the capacities must be the same or one of them will probably cavitate. You could also have a problem operating too far to the right of the best efficiency point with a possible motor "burn out".

• When you vary the speed of a centrifugal pump, the best efficiency point comes down at an angle. The affect is almost the same as changing the diameter of the impeller. This means that the variable speed motor will work best on a system curve that is exponential (Figure "F"). Unfortunately most process and boiler feed pump system curves are not exponential.

• Pump curves are based on a speed of 1450, 1750, 2900, 3500, rpm. Electric induction motors seldom run at these speeds because of "slip". You can estimate that a 2% to a 5% slip is normal in these pumps with the "slip" directly related to the price of the motor.

• You should also keep in mind that if the motor is running at its best efficiency point that does not mean that the pump is running at its B.E.P.

Since you will be using pumps that were supplied at the lowest cost, you can do the following to resist some of the shaft displacement:

• Use a solid shaft. Sleeves often raise the L3/D4 number to over 60 (2 in the metric system), and this is too high a number for reliable seal performance.

• Try to keep the mechanical seal as close to the bearings as possible. It is the mechanical seal that is the most sensitive to shaft deflection and vibration.

• Once the seal has been moved closer to the bearings, you can install a sleeve bearing in the packing space to support the shaft when the pump is operated off of its B.E.P. This is especially important at start up, or any time a pump discharge valve is operated.

• Stop the cavitation if you are experiencing any.

• Balance the rotating assembly.

• Check that the shaft is not bent or the rotating assembly is not out of dynamic balance.

• Use a "C" or "D" frame adapter to solve pump motor alignment difficulties.

• A centre line design wet end can be used if pipe strain, due to temperature expansion, is causing an alignment problem.

Do not trust the system prints to make your calculations. The actual system always differs from that shown on the print, because people tap into the lines, using the pumped fluid for a variety of purposes and after having done so forget to change or "mark up" the original system print. You are going to have to "walk down" the system and note the pipe length, the number of fittings, etc., to make an accurate system head calculation.

Do not be surprised to find that the discharge of your pump is hooked up to the discharge of another pump further down the line. In other words, the pumps are connected in parallel and no body knows it. Pressure recorders (not gauges) installed at the pump suction and discharge is another technique you can use to get a better picture of the system or dynamic head. They will show you how the head is varying with changes in flow.”



PUMP CLINIC 18

PUMP AFFINITY LAWS

There are occasions when you might want to permanently change the amount of liquid you are pumping, or change the discharge head of a centrifugal pump. There are four ways you could do this:

• Regulate the discharge of the pump by using a valve or orifice.

• Change the speed of the pump by changing the motor or using a variable speed drive

• Change the diameter of the impeller.

• Purchase a new pump

Of the four methods the middle two are generally the most sensible ones. In the following paragraphs we'll learn what happens when we change either the pump speed or impeller diameter, and as you would guess, we will see what other characteristics of the pump are going to change along with these values.

To determine what is going to happen we begin by taking the new speed or impeller diameter and divide it by the old speed or impeller diameter. Since changing either one will have approximately the same affect we will refer to only the speed in this part of the discussion.

As an example: NEW SPEED = A VALUE, or 1500 RPM = 0.5 OLD SPEED 3000 RPM

The capacity, or amount of fluid you're pumping, varies directly with this number.

• Example: 50 metres³ per hour x 0.5 = 25 Cubic metres per hour

The head varies by the square of the number.

• Example : 20 metre head x 0.25 ( 0.52) = 5 metre head

The Power required changes by the cube of the number.

• Example: A 9 kW motor was required to drive the pump at 1500 rpm. How much is required if you go to 3000 rpm?

• We would get: 9 x 8 (23) = 72 kW is now required.

• Likewise if a 12 kW motor was required at 3000 rpm and you decreased the speed to 1500 the new kilowatts required would be: 12 x 0.125 (0.53) = 1.5 kW required for the lower rpm.

The following relationships are not exact, but they give you an idea of how speed and impeller diameter affects other pump functions.

The net positive suction head required by the pump (NPSHR) varies by the square of the number.

• Example: If the NPSHR at 1450rpm is 3m, what would be the NPSHR for the given pump if its speed was increased to 2900rpm

• A 3 metre NPSHR x 4 (22) = 12 metre NPSHR at 2900rpm



The amount of shaft run out (deflection) varies by the square of the number

• As an example you had 0.07 mm run out at 2900 rpm and you slowed that shaft down to 1450 rpm the run out would decrease to 0.07 mm x 0.25 (0.52) or 0.018 mm.

The amount of friction loss in the piping varies by approximately 90% of the square of the number. Friction loss through fittings and accessories varies by almost the square of the number.

• As an example: If the system head loss was calculated or measured at 65 metres at 1450 rpm, the loss at 2900 rpm would be: 65 metres x 4 (22) = 260 x 0.9 = 234 metres

The wear rate of the components varies by the cube also

• Example: At 1450 rpm the impeller material is wearing at the rate of 0.5 mm per month. At 2900 rpm the rate would increase to: 0.5 x 8 (23) or 4.0 mm per month. Likewise a decrease in speed would decrease the wear rate eight times as much.

We started this discussion by stating that a change in impeller speed or a change in impeller diameter has approximately the same effect. This is true only if you decrease the impeller diameter to a maximum of 10% This is true because as you cut down the impeller diameter, the housing is not coming down in size correspondingly so the affinity laws do not remain accurate below this 10% maximum number.

The affinity laws do however remain accurate for speed changes and this is important to remember when we convert from gland packing to a balanced mechanical seal. We sometimes experience an increase in motor speed rather than a drop in amperage during these conversions and the affinity laws will help you to predict the final outcome of the change.



PUMP CLINIC 19

ABOUT NET POSITIVE SUCTION HEAD (NPSH)

In past Pump Clinic Articles we have talked about cavitation which most people who have any involvement with pumps will come across at some stage. In this article however we want to talk only about NPSH and what it means as it is a term that is misunderstood by many people.

We do not want bubbles in our process fluid for a lot of reasons:

• Bubbles take up space, causing the pumping capacity to diminish. The head also reduces because energy has to be expended to increase the velocity of the liquid used to fill up the cavities, as the bubbles collapse. As the velocity goes up, the head or pressure goes down.

• Excessive vibration can occur when part of the impeller is handling a liquid and another part is handling a vapour. This vibration can lead to pump failure.

• Air is a poor heat transfer medium, meaning that the fluid we are pumping will get hotter and in almost no cases is there any advantage in heating up the process fluid.

• A bubble is a hole or cavity in the liquid. It is these cavities that are going to cause a cavitation problem that will damage both the impeller and volute.

Bubbles or cavities form in a liquid when the fluid temperature gets too high, or the fluid pressure gets too low. This is called vapourisation, or sometimes boiling although the word boiling tends to imply that the liquid is hot which need not be the case. We all know that if you throw dry ice into cold water it will bubble and vapourise, but it is not hot.

For the purpose of this article we will use vapourise and further state that a fluid will vapourise any time the pressure falls below its vapourisation point.

Since temperature is a variable with different fluids, there are charts that will give you the vapour pressure for any fluid at its various temperatures.

Take a look at the following chart. For the purpose of this article we will use a chart in imperial units. You will note that the vapour pressure for 60˚F chlorine is 80 psi (540kPa), and the vapour pressure for 68˚F fresh water is about 0.3 psi (2kPa). These numbers are required to calculate our NPSH available.



A fluid pressure can be lowered in several ways:

• Put the fluid in a container, and then pull a vacuum on the container. This happens in the hot well of condensers. This can be referred to as a loss of "pressure head"

• Lift the liquid out of a hole. This will diminish the position of the liquid level in respect to the pump centre line. This can be referred to as a loss of "static head"

• Accelerate the fluid. As its velocity increases its pressure will decrease. This is referred to as "velocity head"

• As the fluid moves through piping, fittings, restrictions and valving, some friction losses occur that will drop the fluid pressure. This is referred to as an increase in friction head, resulting in some loss of "positive suction head."



Heating of the incoming fluid is not usually a problem, but it can occur several ways:

• Internal recirculation in the pump because of worn wear rings or failure to make an impeller adjustment.

• Piping, exposed to the elements, can heat up the liquid on hot and sunny days.

We do not know how much pressure a centrifugal pump will develop, but we do know the head it can produce. The head is a function of the shaft speed and the impeller diameter. The faster the speed the higher the head.

The larger the diameter, the bigger the head. To determine the pressure we have to know the weight or "specific gravity" of the fluid we are pumping, and since any given centrifugal pump can move a lot of different fluids, with different specific gravities, it is simpler to discuss the pump's head and forget about the pressure.

Here are the formulas you can use to convert from one to the other:

Head= Pressure x 0.1Specific gravity

Pressure= Head x specific gravity

0.1

In the above formula:

• Head is measured in metres (m)

• Pressure is measured in kilopascals (kPa)

The pump manufacturer has decided how much head the pump needs to prevent cold water from vapourising at different capacities and these values are published on his pump curve. The values have been obtained by testing the pump at different capacities, throttling the suction side and waiting for the first signs of cavitation. The pressure was noted, converted to head, and transferred to the pump curve.

This observed number is called the "Net Positive Suction Head Required (NPSHR).

The attached pump curve shows the numbers. On the chart they are located at the bottom of the dotted lines and they run from 2 to 16. According to this graph a 13-inch impeller, running at its best efficiency point (60+%), would need a NPSH required of 9 feet. An 11-inch impeller running at its best efficiency point would need 7 feet of NPSH required. Remember this requirement is for cold water only.



Be sure to keep in mind that any discussion of NPSH or cavitation is only concerned about the suction side of the pump. There is almost always plenty of pressure on the discharge side of the pump to prevent the fluid from vapourising.

• If we go back to our formula and put the 0.3 psi/ 2kPa vapour pressure for 68˚ water into the numbers, it comes out to 0.7 feet or 0.2 metres of head is required to stop the water from vapourising and forming cavities. So why does the NPSH required increase as the capacity is increasing? It's because the velocity of the liquid is increasing, and anytime the velocity of a liquid goes up, the pressure or head comes down.

Now that we know what head is required, we can calculate the head we have available, and remember we are only interested in the suction side of the pump. Generally we will be looking at three kinds of head.

• The static head measured from the liquid level to the centre line of the pump. If the liquid level is above the pump centre line you will have a positive number. If the level is below the centre line you will have a negative number.

• The pressure head. Here we will be using only absolute numbers. In other words atmospheric pressure is 101kPa at sea level so you will add that number (converted to metres, using the above formula) to the static head if you have an open tank. If the fluid is under vacuum we will convert the absolute pressure reading to head and use that number, instead of atmospheric pressure. The friction loss in the piping will be a minus number. You get the number from charts showing pipes size vs flow, and flow through fittings and valves.



• The next thing we have to do is subtract the vapour pressure of our fluid (converted to feet of liquid) using the first formula I gave you. All of the above, added together is the NPSH available. If this number is equal to, or more than the NPSH required by the pump manufacturer, the liquid will not form bubbles or cavities on the suction side and the pump will not cavitate.

In summary, NPSH available is defined as:

NPSHA = Atmospheric pressure + static head + pressure head - the vapor pressure of your product - loss in the piping, valves and fittings.

NPSHA will always need to be greater than the NPSHR for the pump to operate without cavitation. Most people involved in pumping system design demand an additional safety factor of 1m.



PUMP CLINIC 20

PUMP SELECTION ‘HOW TO PICK THE CORRECT SIZE AND TYPE OF PUMP

FOR YOUR APPLICATION’

Have you ever wondered why sometimes after giving two or more suppliers the same information they come back with quite different pump selections or, why sometimes pump suppliers ask a lot of questions? Many readers of these articles will know what goes into a good pump selection but here we will look at a few of the basics so anyone who is not as familiar with the pump selection process will understand where all the questions are coming from.

We need to begin by deciding what operating conditions the pump has to meet. Generally when you approach a pump supplier you will be armed with this information. To clearly define the capacity and pressure needs of your system sometimes you may construct a system curve. This system curve will then be given to the pump suppliers and they will try to match it with a pump curve that satisfies these needs as closely as possible.

• Decide the capacity you'll need. This means the flow rate usually in cubic meters per hour or litres per second. You must also consider if this capacity will change with the operation of your process. A boiler feed pump is an example of an application that needs a constant pressure with varying capacities to meet a changing steam demand The demand for boiler water is regulated by opening and closing a control valve on the discharge side of the pump with a discharge re-circulation line returning the unneeded portion back to a convenient storage place, or the suction side of the pump. Remember that with a centrifugal pump if you change its capacity you change the head also. A positive displacement pump is different. It puts out a constant capacity regardless of the pressure.

• For other centrifugal pump applications, you're going to have to calculate how much pressure will be needed to deliver different capacities to the location where you'll need them. You'll need enough pressure to :

o Reach the maximum static head or height the fluid will have to attain.

o Overcome any pressure that might be in the vessel where the fluid is discharging, such as the boiler we just discussed. This is called the pressure head.

o Overcome friction resistance in the lines, fittings and any valves or hardware that might be in the system. As an example: high-pressure nozzles can be tricky, especially if they clog up. This resistance is called the friction head.

These heads need to be calculated for both the suction and discharge side of the pump. To get the total head you'll subtract the suction head from the discharge head. This is the head that the pump must produce to satisfy the application.

The total head of a pump seldom remains static. There are a number of factors that can change the head of a pump while it is operating and this is the point where most people start adding in safety factors to compensate for some of the unknowns such as aging or clogged pipework or fittings with unknown resistance. These safety factors will almost always guarantee the selection of an oversized pump that will run off its best efficiency point (BEP) most of the time.

The pump itself requires a certain amount of net positive suction head (NPSHR) to prevent cavitation. This value is shown on the pump curve. When you look at the curve you'll also note that the net positive suction head required (NPSHR) increases with any increase in the pump's capacity. Remember that the net positive suction head required (NPSHR) number shown on the pump curve is for fresh water at 20°C and not the fluid or combinations of fluids you'll be pumping

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Either you or the pump supplier will be calculating the net positive suction head available (NPSHA) to be sure that the pump selected will not cavitate. Cavitation is caused by cavities or bubbles in the fluid collapsing on the impeller, and volute and has been covered in previous Pump Clinic articles so we will not expand on it here other than to say its presence has harmful effects on most pump styles over time and it should be avoided wherever possible.

• You may have to install an inducer on the pump, add a booster pump, or go to a double suction or side channel pump design if you don't have enough net positive suction head available (NPSHA)

• Once the duty information has been established you need to look into materials of construction. Will you need any special materials for the pump components?

o The pump supplier will try to choose pump metal components that are chemically compatible with what you're pumping, as well as any cleaners or solvents that might be flushed through the lines. If the temperature of the pumpage changes, the corrosion rate can change also. Will the material selection have an impact on your stock levels or values if the materials are exotic. This can sometimes be a bit of a balancing act.

o If the product you're pumping is explosive, or a fire hazard, you should be looking at non-sparking materials for the pump components. Do not depend totally upon the pump manufacturer to make this decision for you. If you're not sure what materials are compatible with your product, how will the pump supplier know? Also, keep in mind that some of the fluids you'll be pumping could be proprietary products known only by their trade name.

o Dangerous and radioactive materials will dictate special materials.

o Food products require high-density seal and pump materials that are easy to clean and sterilize. They also sometimes affect the design of the pump internal and external components

o If there are abrasive solids in the pumpage, you'll need materials with good wearing capabilities. Hard surfaces and chemically resistant materials are often incompatible. You may have to go to some type of coating on the pump wetted parts or select an expensive duplex metal.

• Occasionally you'll find an application where metal is not practical. There are many monomer and polymer materials available for these applications, but their cost is generally higher than comparable metal parts. Be aware that if you're using a mechanical seal in a non-metallic pump, the seal can't have metal parts in contact with the fluid for the same reasons the pump was manufactured from non-metallic materials. Use a non-metallic seal or perhaps magnetic drive in these applications

When the pump supplier has all of this information in his possession he can then hopefully select the correct size pump and driver for the job. Since we all want to quote a competitive price we are now going to make some critical decisions:

First we need to look at the style of pump we would recommend:

• If the capacity were going to be very low we would recommend a positive displacement (PD) pump. This could be a gear or diaphragm type pump for small capacities.

• Between about 5 m3 /hr and 120 m3/hr, depending on the differential head we would probably select a single stage end suction centrifugal pump. At higher capacities we may go to a split case or double suction design with a wide impeller or even use two pumps in parallel.

• You might need a high head, low capacity pump in which case we may look at a positive displacement pump again or a multi stage centrifugal pump

• Does the pump need to be self-priming? A self priming pump removes air from the impeller eye and suction side of the pump. Some operating conditions dictate the need for a self-priming design. If you



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do not have a self-priming pump and you're on intermittent service, will priming become a problem the next time you start the pump?

• How will the pump be operated?

o If the pump is going to run twenty-four hours a day, seven days a week and you're not going to open and close valves; you will not need a heavy-duty pump. It's easy to select a pump that'll run at its best efficiency point and at the best efficiency point (BEP) there's very little shaft displacement or vibration.

o Intermittent service is sometimes the more difficult application because of changing temperatures, vibration levels, thrust direction, etc.

• How important is efficiency in your application? High efficiency is desirable, but you pay a price for efficiency in higher maintenance costs and a limited operating window. You should be looking for performance, reliability, and efficiency in that order. Too often the engineer specifies efficiency and loses the other two. The following designs solve some operation and maintenance problems, but their efficiency is lower than conventional centrifugal pumps.

o A magnetic drive or canned pump may be your best option if you can live with the limitations they impose. The main ones being their intolerance of dry running and solids

o A vortex or slurry pump design may be needed if there is a concentration of solids or "stringy" material in the pumpage.

o A double volute centrifugal pump can eliminate many of the seal problems experienced when we operate off the pump's best efficiency point. The problem is trying to find a supplier that will supply one for your application. Although readily available for impellers larger than 14 inches (355 mm) in diameter they have become very scarce in the smaller diameters.

• The supplier should recommend a centerline design to avoid the problems caused by thermal expansion of the wet end if you're operating at temperatures over 200°F (100°C)?

• Will you need a volute or circular casing? Volute casings build a higher head; circular casing are used for low head and high capacity.

• Do you need a pump that meets a standard? ANSI, API, DIN or ISO are some of the current standards.

• The decision to use either a single or multistage pump will be determined by the head the pump must produce to meet the capacities you need. Some suppliers like to recommend a high speed small pump to be competitive, other suppliers might recommend a more expensive low speed large pump to lessen NPSH and wear problems.

There are additional decisions that have to be made about the type of pump the supplier will recommend:

• Will the pump be supplied with a mechanical seal or packing? If the stuffing box is at negative pressure (vacuum) a seal will be necessary to prevent air ingestion.

• If fitted with a mechanical seal will it also have an oversized stuffing box and any environmental controls that might be needed?

• Will the pump have a jacketed stuffing box so that the temperature of the seal fluid can be regulated? How do you intend to control the stuffing box temperature? Will you be using water, steam or maybe a combination of both? Electric heating is sometimes an option.

• How will the open or semi-open impeller be adjusted to the volute casing or back plate? Can the mechanical seal face loading be adjusted at the same time? If not, the seal face load will change and the seal life will be shortened.




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• If the pump is going to be supplied with a closed impeller you should have some means of knowing when the wear rings have to be replaced. If the wear ring clearance becomes too large the pumps efficiency will be lowered causing heat and vibration problems. Most manufacturers require that you disassemble the pump to check the wear ring clearance and replace the rings when this clearance doubles.

• Is the pump fitted with a metric motor frame adapter, or will the pump to motor alignment have to be done manually using dual indicators or a laser aligner to get the readings? A close-coupled design can eliminate the need for an alignment between the pump and driver.

• What type of coupling will you use to connect the pump to its driver? Couplings can compensate for axial growth of the shaft and transmit torque to the impeller. They cannot compensate for pump to driver misalignment as much as we would like them to although this is much better than it was a few years ago. Universal joints and ‘Hardi Spicer’ shafts are a potential problem for those not experienced with their use because they have to be misaligned to be lubricated properly

• Belt drives are an option that allows a reasonable amount of fine tuning of pump speed to be achieved as well as changes of speed later in the pumps life. They are often used on slurry applications.

• The supplier may decide to run two pumps in parallel operation if you need a high capacity, or two pumps in series operation if you need a high head. Pumps that run in parallel or series require that they are running at the same speed. This can be a problem for some induction motors or installations where identical location of the pumps is not possible.

• An inline pump design can solve many pipe strain and thermal growth problems.

• The pump supplier must ensure that the pump will not be operating at a critical speed or passing through a critical speed at start up. If he has decided to use a variable speed drive or motor this becomes a possibility.

• We all want pumps with a low net positive suction head required to prevent cavitation problems but sometimes it's not practical. The manufacturer sometimes has the option of installing an inducer or altering the pump design/application to lower the net positive suction head required, but if he goes too far all of the internal clearances will have to be perfect to prevent cavitation problems.

• Shaft speed is an important decision. Speed affects pump component wear and NPSH requirements, along with the head, capacity, and the pump size. High speed pumps cost less initially, but the maintenance costs can be much higher depending on the pump duty. Speed is especially critical if you're going to be specifying a slurry pump.

There are multiple decisions to be made about the impeller selection and not all pump suppliers are qualified to make them:

• The impeller material must be chosen for both chemical compatibility and wear resistance. You should consider one of the duplex metals because most corrosion resistant materials are too soft for the demands of a pump impeller.

• The decision to use a closed impeller, open impeller, semi-open, or vortex design is another decision to be made.

• Closed impellers require wear rings and these wear rings present another maintenance problem.

• Open and semi-open impellers are less likely to clog, but need manual adjustment to the volute or back-plate to get the proper impeller setting and prevent internal recirculation.

• Vortex pump impellers are great for solids and "stringy" materials but they are up to 50% less efficient than conventional designs.






















• Investment cast impellers are usually superior to sand cast versions because you can cast compound curves with the investment casting process. The compound curve allows the impeller to pump abrasive fluids with less vane wear.

• If you're going to pump low specific gravity fluids with an open impeller, a non-sparking type metal may be needed to prevent a fire or explosion. You'll be better off choosing a closed impeller design with soft wear rings in these applications.

• The affinity laws will predict the affect of changing the impeller speed or diameter.

• Either you or the supplier must select the correct size electric motor, or some other type of driver for the pump. The decision will be dictated by the specific gravity of the liquid you'll be pumping, along with the specific gravity of any cleaners or solvents that might be flushed through the lines. The selection will also be influenced by how far you'll venture off the best efficiency point (BEP) on the capacity side of the pump curve. If this number is under-estimated there is a danger of burning out some electric motors.

• How are you going to vary the pump's capacity? Are you going to open and close a valve or maybe you'll be using a variable speed drive, or maybe a petrol or diesel engine. Will the regulating valve open and close automatically like a boiler feed valve, or will it be operated manually? A variable speed motor or a frequency inverter might be an alternative if the major part of the system head is friction head rather than static or pressure head.

• The viscosity of the fluid is another consideration because it'll affect the head, capacity, efficiency and power requirement of the pump. You should know about viscosity and how the viscosity of the pumpage will affect the performance of the pump. There are some viscosity corrections you can make to the pump curve when you pump viscous fluids.

• After carefully considering all of the above, the pump supplier will select a pump type and size, present the quote and give you a copy of the pump curve. Hopefully you'll be getting the supplier’s best pump technology.

If all of the above decisions were made correctly, the pump supplier will place his pump curve on top of your system curve and the required operating window will fall within the pump's operating window on either side of the best efficiency point (BEP). Additionally, the motor will not overheat and the pump should not cavitate.

If the decisions were made incorrectly, the pump will operate where the pump and system curves intersect and that will not be close to, or at the best efficiency point, producing radial impeller loading problems that will cause shaft deflection, resulting in premature seal and bearing failures. Needless to say the motor or driver will be adversely affected also.











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Pump Clinic 21 28/02/08 Page 1 of 3

PUMP CLINIC 21

STALLING IN SOME AIR-POWERED DOUBLE DIAPHRAGM PUMPS

One of the more serious operating issues for some air powered double diaphragm pumps is referred to in the Industry as "Stalled” or “Stalling”. This occurs when a pump would not restart after it was dead headed or was simply found to have stalled. The condition under which Stalling occurs almost always includes where the discharge side is air or vapour bound and where operation is at a low speed or dead headed. It can also be evident where critical air valve components are worn allowing internal air by-pass. The results are the same. The pump may fail to restart or will stop without warning. The only means to restart are to remove and reapply the air supply or with some air powered double diaphragm pumps to use hammer adjustment by hitting the main air valve housing with a hammer or similar object.

HAMMER ADJUSTMENT

A Stalling condition should not be confused with a Sticking condition; both of which may exist. The conditions that can cause both are outlined as follows: A pump can be considered Stalled when both the main air valve and the pilot valve spools are centered in their respective travel. This is when air pressure is equalised on each side of the air valve or pilot valve. When the air valve is in this position it will divert or block off the supply air (power source) preventing the unit from reciprocating. If either valve spool is off centre, the unit will again function. Diaphragm assemblies can float when the discharge system is compressible (air or vapor bound). In these instances, if the pump were dead headed by closing a valve in the discharge piping, they would float back and forth in their travel as they compress and decompress the fluid mixture in their respective chambers. As the pump continues to cycle, i t can repeatedly pulse the pilot valve, hence building up balanced air pressures causing the valve spools to centre. This phenomenon is regularly encountered in unbalanced or bias designed air valves and this problem presents itself as a rapid fire rattle or machine gun type noise in the pump. A similar condition can also exist with low flow / low speed situations where the same events will occur. Two applications where both conditions can occur are Filter Press service and On-Demand Spray service. Both have low speed / low flow coupled with dead head and can have air / vapour build-up on the discharge side of the pump. Eliminating the condition of air / vapour build-up fixes the problem but this fix is not always possible. A pump can be considered to be Sticking if i t fails to restart or stops and inspection reveals that the main air valve spool is seized or stuck in the valve sleeve. This can come about through the air valve being contaminated by a poor quality air supply or pumped product that may have migrated into the air valve through a failed diaphragm.

New Design Enhancement



This enhancement with Sandpiper pumps shifting mechanism has made a dramatic improvement in the performance of the pumps in so called trouble applications. This has been proved in many field trials in very tough and not so tough field applications. The not so tough applications were to prove that Sandpiper did not overcome one problem only to create another one. The design has been applied across the complete range of Sandpiper pumps. The enhancement has been referred to as "cross-drilled shifters" for lack of a better term. This term describes the most apparent part of the change, the cross-drilled porting. This change has been in the main air valve sleeve and/or the pilot valve body. The concept remains the same across all Sandpiper pump sizes but the means vary according to the unit designs.

The enhancement provides for what can be called an Air Detent that positively locks the main air valve spool on one end of its travel or the other. There is no way for the valve spool to float allowing i t to centre. Only after the pilot valve has made a complete, positive shift can the main air valve spool shift to the opposite end of its travel. The Sandpiper air valve design has proven tolerant to poor quality air supplies through its design and selection of materials. Several components were also specifically developed for use in mine service where poor quality air is not uncommon because of the application environments in mines. The newer design compensates for these conditions and 'strokes right through them' with an audible, positive stroke. The improvement can be heard. The components have proven themselves in this application by extending service life and reliability.

This can only be taken so far in anyone's design until the air valve assembly becomes so fouled that it must be removed to be cleaned and or serviced. In these instances, the Sandpiper ESADS Plus feature will prove its worth through its design for ease of access and serviceability. This provides for servicing of the main valve and pilot valve assemblies without needing to disconnect suction and discharge pipework.



How is this Air Detent feature accomplished? Simply, the air is ported from the inner chamber that is on a discharge stroke (under power) to the end of the main air valve that has the pilot air signal applied. This air is directed through the 'cross-drilled' porting detailed earlier. This supplemental or Air Detent holds the main air valve spool in place even if the pilot signal is lost. Only after the pilot signal has been positively applied to the opposite side of the main air valve spool can it shift. At the same instant, the opposite side of the valve has its pilot signal air pressure and the Air Detent pressure vented through the pilot valve exhaust and the main air valve exhaust. This causes a very rapid and greater than standard, differential pressure across the main air valve spool causing a very quick, positive shift. This assures consistent, full length pump strokes. As detailed earlier, you can hear the difference.

All the air is ported to where it needs to be through a series of engineered orifices or cross-drilled ports. It is more than just connecting one part to another, with all the modifications undergoing rigorous testing both in-house and in the field.


Pump Clinic 22 Viscosity 27/03/08 Page 1 of 5

PUMP CLINIC 22

VISCOSITY

The viscosity of a fluid is that property which tends to resist a shearing force. It can be thought of as the internal friction resulting when one layer of fluid is made to move in relation to another layer. A detailed discussion on viscosity is a major undertaking and this article serves to provide a basic understanding of viscosity and how it impacts on pumping.

Consider the model shown in Fig. 1, which was used by Isaac Newton in first defining viscosity. It shows two parallel planes of fluid of area A separated by a distance dx and moving in the same direction at different velocities V1 and V2.

Fig. 1

The velocity distribution will be linear over the distance dx, and experiments show that the velocity

gradient, is directly proportional to the force per unit area,

Where n constant for a given liquid and is called its viscosity.

The velocity gradient, describes the shearing experienced by the intermediate layers as they move with respect to each other. Therefore, it can be called the "rate of shear", S. Also, the force per

unit area can be simplified and called the "shear force" or "shear stress," F. With these simplified terms, viscosity can be defined as follows'.

Newtonian Liquids Isaac Newton made the assumption that all materials have, at a given temperature, a viscosity that is independent of the rate of shear. In other words, a force twice as large would be required to move a liquid twice as fast. Fluids which behave this way are called Newtonian fluids. There are, of course, fluids which do not behave this way, in other words their viscosity is dependent on the rate of shear. These are known as Non-Newtonian fluids.



Fig. 2 shows graphically the relationships between shear Stress (F,) rate of shear (S,) and viscosity (n) for a Newtonian liquid. The viscosity remains constant as shown in sketch 2, and in absolute units, the viscosity is the inverse slope of the line in sketch 1. Water and light oils are good examples of Newtonian liquids.

Fig. 2 Newtonian Liquid

Non-Newtonian Liquids Fig. 3 shows graphically the three most common types of Non-Newtonian liquids. These liquids can present problems to the pump suppliers. Group A shows a decreasing viscosity with an increasing rate of shear. This is known as a pseudo-plastic material. Examples of this type are grease, molasses, paint, soap, starch, and most emulsions. They present no serious mechanical pumping problems since they tend to thin out with the high rates of shear present in a pump. They can present problems in positive displacement pump selection because slippage through clearances may increase due to the drop in viscosity and pump speeds may need to be increased to compensate. Group B shows a dilatant material or one in which the viscosity increases with an increasing rate of shear. Clay slurries and candy compounds are examples of dilatant liquids. Pumps must be selected with extreme care since these liquids can become almost solid if the shear rate is high enough. The normal procedure would be to oversize the pump somewhat and open up the internal clearances in an effort to reduce the shear rate.

Group C shows a plastic material. The viscosity decreases with increasing rate of shear. However, a certain force must be applied before any movement is produced. This force is called the yield value of the material. Tomato sauce is a good example of this type of material. It behaves similar to a pseudo-plastic material from a pumping standpoint.

Fig. 3 Non-Newtonian Liquids



The viscosity of some Non-Newtonian liquids is dependent upon time as well as shear rate. In other words, the viscosity at any particular time depends upon the amount of previous agitation or shearing of the liquid. A liquid whose viscosity decreases with time at a given shear rate is called a thixotropic liquid. Examples are asphalts, glues, molasses, paint, soap, starch, and grease. Liquids whose viscosity increases with time are called rheopectic liquids, but they are seldom encountered in pumping applications.

Units of Viscosity

There are two basic viscosity parameters: dynamic (or absolute) viscosity and kinematic viscosity. Dynamic viscosities are given in terms of force required to move a unit area a unit distance. This is usually expressed in pound-seconds per square foot in the English system which is equal to slugs per foot-second. The Metric system is more commonly used, however, in which the unit is the dyne-second per square centimetre called the Poise. This is numerically equal to the gram per centimetre-second. For convenience, numerical values are normally expressed in centipoise, which are equal to one-hundredth of a poise.

Most pipe friction charts and pump correction charts list kinematic viscosity. The basic unit of kinematic viscosity is the stoke which is equal to a square centimetre per second in the Metric system. The corresponding English unit is square foot per second. The centistoke which is one-hundredth of a stoke is normally used in the charts. The following formula is used to obtain the kinematic viscosity when the dynamic or absolute viscosity is known:

There are various units used for viscosity and these are determined by the type of viscometers utilised for determining liquid viscosities, most of which are designed for specific liquids or viscosity ranges. The Saybolt viscometers are probably the most widely used in the United States. The corresponding units are the SSU (Seconds Saybolt Universal).

These units are found on most pipe friction and pump correction charts in addition to centistokes. Conversion charts for various units of viscosity are attached.

Viscosity and Pumping

1. Centrifugal pumps. Centrifugal pump performance curves are primarily based on the viscosity of water; namely 1cst. Higher viscosities affect the capacity-head performance and more significantly the pump efficiency and therefore power requirements. The water performance of pumps may be adjusted for any viscosity and this is covered in a separate Pump Clinic titled Viscosity Impact on Centrifugal Pump Performance.

Because of the significant impact of viscosity on power requirements, there are general viscosity limits for centrifugal pumps. These are simply arbitrary figures. The PIA Handbook defines the limits based on the dimension in millimetres of the pump discharge connections and these are

< 50 mm maximum 300 cst

>50mm but <150 mm maximum 500 cst

>150 mm maximum 800 cst

Our experience has indicated that these viscosities may be a little high and better limits are;

< 50 mm maximum 100 cst

>50mm but <150 mm maximum 250 cst

>150 mm maximum 400 cst

This only shows the arbitrary nature of these limits. It is important to understand the viscosity characteristics as the liquid is sheared in particular where the viscosity increases with shear and the pump manufacturer should be consulted in these instances.



2. Positive Displacement Pumps. The application of positive displacement (PD) pumps is easier as the majority of PD pump selection procedures and software programs use viscosity as one of the determining parameters for pump size, speed and motor selection. The other parameters are flow, pressure and other liquid conditions e.g. solids content.

The change in viscosity as the product is sheared is more important with PD pump selection irrespective of whether viscosity increases or decreases. With decreasing viscosity, the impact of liquid slippage through pump clearances from pump discharge to suction may increase significantly (dependent on differential pressure across the pump) and this needs to be considered in pump size and speed selection. With increasing viscosities, mechanical issues as well a speed reduction is a major consideration. Contact the pump supplier with viscosity change information in these instances.

Section D -- Properties of Liquids

Viscosity Conversion Table 1



Viscosity Conversion Table 2


Pump Clinic 23 Controlling Surge and Pulsation Problems 23/04/08 Page 1 of 1

PUMP CLINIC 23

CONTROLLING SURGE AND PULSATION PROBLEMS

Reprinted with kind permission of Blacoh Fluid Control, Inc


Pump Clinic 24 Sealing in Pumps (Reprinted from ITT Goulds Website) 29/05/08 Page 1 of 10

PUMP CLINIC 24

SEALING IN PUMPS

(Reprinted from ITT Goulds Pumps Website)

The proper selection of a seal is critical to the success of every pump application. For maximum pump reliability, choices must be made between the type of seal and the seal environment. In addition, a sealless pump is an alternative, which would eliminate the need for a dynamic type seal entirely.

Sealing Basics There are two basic kinds of seals: static and dynamic. Static seals are employed where no movement occurs at the Juncture to be sealed. Gaskets and O-rings are typical static seals.

Dynamic seals are used where surfaces move relative to one another. Dynamic seals are used, for example, where a rotating shaft transmits power through the wall of a tank (Fig. 1), through the casing of a pump (Fig. 2), or through the housing of other rotating equipment such as a filter or screen.

Fig. 1 Cross Section of Tank and Mixer

Fig. 2 Typical Centrifugal Pump A common application of sealing devices is to seal the rotating shaft of a centrifugal pump. To best understand how such a seal functions a quick review of pump fundamentals is in order.



In a centrifugal pump, the liquid enters the suction of the pump at the center (eye) of the rotating impeller (Figures 3 and 4).

Fig. 3 Centrifugal Pump, Liguid End

Fig. 4 Fluid Flow in Centrifugal Pump As the impeller vanes rotate, they transmit motion to the incoming product, which then leaves the impeller, collects in the pump casing, and leaves the pump under pressure through the pump discharge.

Discharge pressure will force some product down behind the impeller to the drive shaft, where it attempts to escape along the rotating drive shaft. Pump manufacturers use various design techniques to reduce the pressure of the product trying to escape. Such techniques include: 1) the addition of balance holes through the impeller to permit most of the pressure to escape into the suction side of the impeller, or 2) the addition of back pump-out vanes on the back side of the impeller.

However, as there is no way to eliminate this pressure completely, sealing devices are necessary to limit the escape of the product to the atmosphere. Such sealing devices are typically either compression packing or end-face mechanical seals.

Stuffing Box Packing A typical packed stuffing box arrangement is shown in Fig. 5. It consists of: A) Five rings of packing, B) A lantern ring used for the injection of a lubricating and/or flushing liquid, and C) A gland to hold the packing and maintain the desired compression for a proper seal.



Fig. 5 Typical Stuffing Arrangement (description of parts) The function of packing is to control leakage and not to eliminate it completely. The packing must be lubricated, and a flow from 40 to 60 drops per minute out of the stuffing box must be maintained for proper lubrication.

The method of lubricating the packing depends on the nature of the liquid being pumped as well as on the pressure in the stuffing box. When the pump stuffing box pressure is above atmospheric pressure and the liquid is clean and nonabrasive, the pumped liquid itself will lubricate the packing (Fig. 6).

Fig. 6 Typical Stuffing Arrangement when Stuffing Box Pressure is Above Atmospheric Pressure When the stuffing box pressure is below atmospheric pressure, a lantern ring is employed and lubrication is injected into the stuffing box (Fig. 7). A bypass line from the pump discharge to the lantern ring connection is normally used providing the pumped liquid is dean.



Fig. 7 Typical Stuffing Box Arrangement when Stuffing Box Pressure is Below Atmospheric Pressure When pumping slurries or abrasive liquids, it is necessary to inject a dean lubricating liquid from an external source into the lantern ring (Fig. 8). A flow of from .05 to .12 m3/hr is desirable and a valve and flowmeter should be used for accurate control. The seal water pressure should be from .7 to 1.0 bars above the stuffing box pressure, and anything above this will only add to packing wear. The lantern ring Is normally located In the center of the stuffing box. However, for extremely thick slurries like paper stock, it is recommended that the lantern ring be located at the stuffing box throat to prevent stock from contaminating the packing.

Fig. 8 Typical Stuffing Box Arrangement when Pumping Slurries The gland shown in Figures 5 through 8 is a quench type gland. Water, oil, or other fluids can be injected into the gland to remove heat from the shaft, thus limiting heat transfer to the bearing frame. This permits the operating temperature of the pump to be higher than the limits of the bearing and lubricant design. The same quench gland can be used to prevent the escape of a toxic or volatile liquid into the air around the pump. This is called a smothering gland, with an external liquid simply flushing away the undesirable leakage to a sewer or waste receiver.

Today, however, stringent emission standards limit use of packing to non-hazardous water based liquids. This, plus a desire to reduce maintenance costs, has increased preference for mechanical seals.

Mechanical Seals A mechanical seal is a sealing device which forms a running seal between rotating and stationary parts. They were developed to overcome the disadvantages of compression packing. Leakage can be reduced to a level meeting environmental standards of government regulating agencies and maintenance costs can be lower. Advantages of mechanical seals over conventional packing are as follows:



1. Zero or limited leakage of product (meet emission regulations.)

2. Reduced friction and power loss.

3. Elimination of shaft or sleeve wear.

4. Reduced maintenance costs.

5. Ability to seal higher pressures and more corrosive environments.

6. The wide variety of designs allows use of mechanical seals in almost all pump applications. The Basic Mechanical Seal All mechanical seals are constructed of three basic sets of parts as shown in Fig. 9:

1. A set of primary seal faces: one rotary and one stationary shown in Fig. 9 as seal ring and insert.

2. A set of secondary seals known as shaft packings and insert mountings such as 0-rings, wedges and V-rings.

3. Mechanical seal hardware including gland rings, collars, compression rings, pins, springs and bellows.

Fig. 9 A Simple Mechcanical Seal How a Mechanical Seal Works

The primary seal is achieved by two very flat, lapped faces which create a difficult leakage path perpendicular to the shaft. Rubbing contact between these two flat mating surfaces minimizes leakage. As in all seals, one face is held stationary in a housing and the other face is fixed to, and rotates with, the shaft. One of the faces is usually a non-galling material such as carbon-graphite. The other is usually a relatively hard material like silicon-carbide. Dissimilar materials are usually used for the stationary insert and the rotating seal ring face in order to prevent adhesion of the two faces. The softer face usually has the smaller mating surface and is commonly called the wear nose.

There are four main sealing points within an end face mechanical seal (Fig. 10). The primary seal is at the seal face, Point A. The leakage path at Point B is blocked by either an 0-ring, a V-ring or a wedge. Leakage paths at Points C and D are blocked by gaskets or 0-rings.



Fig. 10 Sealing Points for Mechanical Seal The faces in a typical mechanical seal are lubricated with a boundary layer of gas or liquid between the faces. In designing seals for the desired leakage, seal life, and energy consumption, the designer must consider how the faces are to be lubricated and select from a number of modes of seal face lubrication.

To select the best seal design, it's necessary to know as much as possible about the operating conditions and the product to be sealed. Complete information about the product and environment will allow selection of the best seal for the application.

Mechanical Seal Types Mechanical seals can be classified into several tvpes and arrangements:

PUSHER: Incorporate secondary seals that move axially along a shaft or sleeve to maintain contact at the seal faces. This feature compensates for seal face wear and wobble due to misalignment. The pusher seals' advantage is that it's inexpensive and commercially available in a wide range of sizes and configurations. Its disadvantage is that ft's prone to secondary seal hang-up and fretting of the shaft or sleeve. Examples are Dura RO and Crane Type 9T.



UNBALANCED: They are inexpensive, leak less, and are more stable when subjected to vibration, misalignment, and cavitation. The disadvantage is their relative low pressure limit. If the closing force exerted on the seal faces exceeds the pressure limit, the lubricating film between the faces is squeezed out and the highly loaded dry running seal fails. Examples are the Dura RO and Crane 9T.

CONVENTIONAL: Examples are the Dura RO and Crane Type 1 which require setting and alignment of the seal (single, double, tandem) on the shaft or sleeve of the pump. Although setting a mechanical seal is relatively simple, today's emphasis on reducing maintenance costs has increased preference for cartridge seals.

NON-PUSHER: The non-pusher or bellows seal does not have to move along the shaft or sleeve to maintain seal face contact, The main advantages are its ability to handle high and low temperature applications, and does not require a secondary seal (not prone to secondary seal hang-up). A disadvantage of this style seal is that its thin bellows cross sections must be upgraded for use in corrosive environments Examples are Dura CBR and Crane 215, and Sealol 680.



BALANCED: Balancing a mechanical seal involves a simple design change, which reduces the hydraulic forces acting to close the seal faces. Balanced seals have higher-pressure limits, lower seal face loading, and generate less heat. This makes them well suited to handle liquids with poor lubricity and high vapor pressures such as light hydrocarbons. Examples are Dura CBR and PBR and Crane 98T and 215.

CARTRIDGE: Examples are Dura P-SO and Crane 1100 which have the mechanical seal premounted on a sleeve including the gland and fit directly over the Model 3196 shaft or shaft sleeve (available single, double, tandem). The major benefit, of course is no requirement for the usual seal setting measurements for their installation. Cartridge seals lower maintenance costs and reduce seal setting errors

Mechanical Seal Arrangements SINGLE INSIDE: This is the most common type of mechanical seal. These seals are easily modified to accommodate seal flush plans and can be balanced to withstand high seal environment pressures. Recommended for relatively clear non-corrosive and corrosive liquids with satisfactory' lubricating properties where cost of operation does not exceed that of a double seal. Examples are Dura RO and CBR and Crane 9T and 215. Reference Conventional Seal.

SINGLE OUTSIDE: If an extremely corrosive liquid has good lubricating properties, an outside seal offers an economical alternative to the expensive metal required for an inside seal to resist corrosion. The disadvantage is that it is exposed outside of the pump which makes it vulnerable to damage from impact and hydraulic pressure works to open the seal faces so they have low pressure limits (balanced or unbalanced).



DOUBLE (DUAL PRESSURIZED): This arrangement is recommended for liquids that are not compatible with a single mechanical seal (i.e. liquids that are toxic, hazardous [regulated by the EPA], have suspended abrasives, or corrosives which require costly materials). The advantages of the double seal are that it can have five times the life of a single seal in severe environments. Also, the metal inner seal parts are never exposed to the liquid product being pumped, so viscous, abrasive, or thermosetting liquids are easily sealed without a need for expensive metallurgy. In addition, recent testing has shown that double seal life is virtually unaffected by process upset conditions during pump operation; a significant advantage of using a double seal over a single seal. The final decision between choosing a double or single seal comes down to the initial cost to purchase the seal, cost of operation of the seal, and environmental and user plant emission standards for leakage from seals. Examples are Dura double RO and X-200 and Crane double 811T.

DOUBLE GAS BARRIER (PRESSURIZED DUAL GAS): Very similar to cartridge double seals ... sealing involves an inert gas, like nitrogen, to act as a surface lubricant and coolant in place of a liquid barrier system or external flush required with conventional or cartridge double seals. This concept was developed because many barrier fluids commonly used with double seals can no longer be used due to new emission regulations. The gas barrier seal uses nitrogen or air as a harmless and inexpensive barrier fluid that helps prevent product emissions to the atmosphere and fully complies with emission regulations. The double gas barrier seal should be considered for use on toxic or hazardous liquids that are regulated or in situations where increased reliability is the required on an application. Examples are Dura GB2OO, GF2OO, and Crane 2800.



TANDEM (DUAL UNPRESSURIZED): Due to health, safety, and environmental considerations, tandem seals have been used for products such as vinyl chloride, carbon monoxide, light hydrocarbons, and a wide range of other volatile, toxic, carcinogenic, or hazardous liquids. Tandem seals eliminate icing and freezing of light hydrocarbons and other liquids which could fall below the atmospheric freezing point of water in air (32? F or 0? C). {Typical buffer liquids in these applications are ethylene glycol, methanol, and propanol.) A tandem also increases online reliability. If the primary seal fails, the outboard seal can take over and function until maintenance of the equipment can be scheduled. Examples are Dura TMB-73 and tandem PTO.


Pump Clinic 25 Pump Reliability (Reprinted from ITT Goulds Website) 23/06/08

PUMP CLINIC 25

PUMP RELIABILITY

(Reprinted from ITT Goulds Pumps Website)

Allan R. Budris, Director Product DevelopmentEugene P. Sabini, Director TechnologyR. Barry Erickson, Vice President TechnologyITT Industrial Pump Group

Recently, significant attention has been given to thelife cycle cost of owning a pump. Major componentsof the cost of ownership are initial cost, installationcost, operating cost, and maintenance cost. Inprocess plants it has been found that under manycircumstances the cost of unscheduled maintenanceis the most significant cost of ownership. Althoughnumerous papers have been presented on thesubject of pump reliability, that literature primarilyaddresses mechanical means of improving reliability.The results of this attention to the mechanical issuesas been a marked increase in the “Mean TimeBetween Repair” (MTBR) for process plants. This hasbeen achieved largely through improved installationpractices, and increased attention to operatingprocedures.

Efforts such as these will continue to yieldimprovements in MTBF, but will be limited inpotential unless a holistic approach is used. Such an approach would give more attention to the besthydraulic fit to optimize reliability. There are fourbasic hydraulic selection factors which can have asignificant affect on pump reliability. They are PumpSpeed, Percent of Best Efficiency Flow, Suction Energyand NPSH Margin Ratio. These last two factors havefurther been combined into an NPSH MarginReliability Factor (NPSH-RF), which has been shownto be reasonably effective in predicting the reliabilityof High Suction Energy pumps.

The Laboratory reliability factors presented here (1)

are based on correlation of the Block and Geitner (2)

reliability factors with laboratory pump bearingframe oil temperature, and vane pass vibration testson 3 API (end suction) pumps, plus publishedmechanical seal face and abrasive wear rates.

The field test reliability factors presented are derivedfrom curve fits (trend lines) of Mean Time BetweenRepair data, on 71 ANSI and 48 split case pumps, intwo process plants. There was much scatter of thedata, due to the fact that the records were notcleansed of failures caused by factors other thanhydraulic selection, such as human error, difficult tohandle liquids, system interactions, or themechanical design of the pumps. The duty cycles(operating times) varied between pumps, especiallywhere pumps were on standby service. Also, thepumps were not always operated at the conditionsof service analyzed. Despite the resulting largescatter in the data, definite trend lines could be and were developed, on the strength of the largenumber of pumps evaluated.

OPERATING SPEED:

Operating Speed affects reliability through rubbingcontact, such as seal faces, reduced bearing lifethrough increased cycling, lubricant degradation andreduced viscosity due to increased temperature, andwetted component wear due to abrasives in thepumpage. Operating Speed also increases theenergy level of the pump, which can lead tocavitation damage.

Figure 1 compares the API-610 pump laboratoryreliability predictor test results with the reliabilitytrend line from actual MTBR data on 119 actualprocess pumps, as a function of the ratio of theactual to maximum rated pump speed. TheReliability factor for the field test data was based onzero pump repairs in a 48 month period, which wasassumed to be equal to a MTBR of 72 months. Bothcurves show a marked increase in reliability withreduced speed.

1

Pump Reliability - Correct Hydraulic Selection Minimizes Unscheduled Maintenance

IN THIS ISSUE:Feature:Pump Reliability - Correct HydraulicSelection Minimizes UnscheduledMaintenance ........................Page 1

Tech Talk:Business Environment Changes Drive Product CostReduction Initiative to GainCompetitive Advantage ........Page 4

New Products:New PumpSmart®

Model PS100 Offers Cost Effective Pump Protection Plus Up to 70% Energy Savings ..................................Page 7

Goulds Model 3355 ..............Page 8

ANSI Combo Units Expand the Polyshield® Foundation Offering.................................Page 9

Material Matters:NACE MR0175 ......................Page 9

Personnel Moves:Manna Named Vice PresidentGlobal Marketing ...............Page 11

New Positions Support After SalesServices................................Page 11

Service Solutions:Think Big!The CPC Internalift Screw Pump Story...............Page 12

View this issue and previous issues of PumpLines

on our website at www.gouldspumps.com

Innovation...Technology...Leadership

FALL 2001

continued on page 2

Figure 1.

2

PERCENT BEST EFFICIENCY FLOW RATE (Flow Ratio):

The Flow Ratio affects reliability through theturbulence that is created in the casing andimpeller as the pump is operated away fromthe best efficiency flow rate. As a result,hydraulic loads, which are transmitted to theshaft and bearings, increase and becomeunsteady. Also, the severity of these unsteadyloads can reduce mechanical seal life.Operation at reduced flow rates that put thepump into its recirculation mode can also leadto cavitation damage in High Suction Energypumps. Refer to ANSI/HI 9.6.3 (5) for moreguidance on the allowable operating regionfor centrifugal and vertical pumps.The field data to laboratory reliabilitycomparison for the Flow Ratio is presented infigure 2. The field data is, however, onlybased on the 48 split case pumps, since nodefinitive trend line could be established fromthe ANSI plant data. Also, for trend purposes,the 1.00 Field Reliability Factor is based on aMTBR of 52 months. Correlation between thefield and laboratory data is good in thenormal operating range, with the maximumreliability occurring around 90 percent of thebest efficiency flow rate.

SUCTION ENERGY:

Suction Energy is another term for the liquidmomentum in the suction eye of a pumpimpeller, which means that it is a function ofthe mass and velocity of the liquid in the inlet.Suction Energy, as originally approximated byBudris and Mayleben (3), is defined as follows:

Suction Energy (S.E.) = De x n x S x s.g. Equation (1)

Where:De = Impeller Eye Diameter (inches)N = Pump Speed (RPM)S = Suction Specific Speed

(RPM x (GPM).5 / (NPSHR).75

s.g. = Specific Gravity of Liquid pumped

Since the suction energy numbers are quitelarge, the last six digits are normally dropped(S.E. x E6). It should be noted that, if notreadily available, the Impeller Eye Diametercan be approximated as follows:

End Suction Pump: De = 0.9 x Suction Nozzle Size

Split Case/Radial Inlet Pumps: De = 0.75 x Suction

Nozzle Size

Budris and Mayleben (3) have also proposeddistinct gating values for High and Very HighSuction Energy, for End Suction and RadialSuction (also know as split case or doublesuction) pumps, based on the analysis ofhundreds of pumps from several manufacturers.

Start of High Suction Energy:End Suction Pumps: S.E. = 160 x 106

Split Case/Radial Inlet Pumps: S.E. = 120 x 106

Start of Very High Suction Energy:End Suction Pumps: S.E. = 240 x 106

Split Case/Radial Inlet Pumps: S.E. = 180 x 106

The above definition of Suction Energy(Equation (1)), and “High” and “Very High”gating values are consistent with valuespresented in ANSI/HI 9.6.1 (4).

Pumps with values of suction energy belowthese values are considered to have lowsuction energy. Generally speaking, LowSuction Energy pumps are not prone to noise,vibration or damage from cavitation. However,there could be detrimental effects onmechanical seals from the air or vapors whichmay be liberated from the liquid during theformation of the cavitation bubbles, under lowNPSH Margin conditions (below 1.1 – 1.3NPSH Margin Ratio).

NPSH MARGIN:

NPSH Margin Ratio is defined as the NPSHAvailable to the pump by the application,divided by the NPSH Required by the pump. By Hydraulic Institute definition, the NPSHR of apump is the NPSH that will cause the totalhead to be reduced by 3%, due to flowblockage from cavitation vapor in the impellervanes. NPSHR is by no means the point atwhich cavitation starts. That level is referred toas incipient cavitation. It can take an NPSHA offrom 2 to 20 times NPSHR to fully suppresscavitation within a pump, depending on pumpdesign and Flow Ratio (percent bep). Thehigher values are normally associated with highsuction energy, high specific speed, pumps withlarge impeller inlet areas, or reduced flowoperation in the region of suction recirculation.This means that a high percentage of pumpsare operating with some degree of cavitation.It is the amount of Energy associated with thecollapse of the cavitation bubbles thatdetermines the degree of noise, vibration ordamage from cavitation, if any.

Figure 4 shows the affect of the NPSH MarginRatio on pump reliability, based on the 77 fieldpumps. Again, the Low suction Energy failures(below 48 months) were deleted, because it isunlikely that these failures were caused by

Pump Reliability...continued from page 1

Figure 2.

Figure 3.

Figure 3 is based strictly on thefield data for 77 ANSI and SplitCase pumps, with the 42 Lowsuction Energy failures (below48 months) being deleted,because it is unlikely that thesefailures were caused by factorsrelated to Suction Energy,mainly cavitation. Here also, a1.00 Reliability Factor equatesto no failures in 48 months, ora MTBR rate of 72 months. The trend is unquestionable,with higher suction energypumps requiring the mostfrequent repairs.

factors related to Suction Energy, mainlycavitation. Based on this data, the NPSHMargin Ratio does have a definite influence onpump reliability, especially for High and VeryHigh Suction Energy pumps, due to the factthat some cavitation usually exists below aRatio of 4.0.

NPSH MARGIN RELIABILITY FACTOR:The NPSH Margin Reliability Factor (Fig. 5) was developed to quantify the relationshipbetween NPSH Margin and Suction Energy onpump reliability. The NPSH Margin ReliabilityFactors are based on the fact that, above thegating suction energy values (start of HighSuction Energy), the greater the suction energy

the more important it is to suppress theresidual cavitation that exists above the NPSHR,to prevent damage. This reliability factor is onlyapplicable within the allowable operating flowregion, above the start of suction recirculation(see ref. 5). Much higher NPSH Margin valuesare required in the region of suctionrecirculation, for High and Very High SuctionEnergy pump applications.

The diagonal lines (in figure 5) are lines ofconstant relative Suction Energy (x 106).Therefore, (for example) the line marked“180/240” (Double Suction Suction Energy level/ End Suction Suction Enegy level) representsthe start of Very High Suction Energy. Pumps ofthis suction energy level require a minimum

3

Pump Reliability...continued from page 2

Figure 5.

Figure 4.

Figure 6.

NPSH Margin Ratio of 2.5 formaximum reliability.

To validate the NPSH MarginReliability Factors in figure 5,NPSH R.F. values were plottedagainst the field reliability ofthe 77 ANSI and Split Casepumps (without the 42 LowSuction Energy failures / below48 months), as shown in figure6. Although not perfect, theagreement is quite good. Itmust be remembered that theNPSH R.F. only applies to “HighSuction Energy” and “Very HighSuction Energy” pumps.

CONCLUSIONS:

The speed,flow ratio,suction energy and NPSH margin reliability propositions andmethodologies were confirmed by fieldexperience.

The “Mean Time Between Repair” (MTBR) andLife Cycle Cost of most centrifugal pumps canbe improved if slower pump speeds are used,and pumps are selected to operate in theirpreferred operating range (70% - 120% of bepflow rate – ref. 5).

Further, the Mean Time Between Repair ofHigh and Very High Suction Energy pumps can be increased by keeping the NPSH MarginRatio above the values recommended in figure 5, and/or by reducing the SuctionEnergy Level. The easiest way to lower theSuction Energy and increase the NPSH Marginof a pump application is by lowering thespeed of the pump. ■

REFERENCES:

1. Erickson, R. B., Sabini E. P. and Stavale, A. E., October2000, “Hydraulic Selection to Minimize the UnscheduledMaintenance Portion of Life Cycle Cost,” Pump UsersInternational Forum 2000, Karlsruhe, Germany.

2. Bloch, H.P. and Geitner, F. K., 1994, “An Introduction toMachinery Reliability Assessment,” Gulf PublishingCompany, Houston, TX.

3. Budris, A. R. and Mayleben, P. A., 1998, “Effects ofEntrained Air, NPSH Margin, and Suction Piping on Cavitationin Centrifugal Pumps,” International Pump UsersSymposium proceedings, Texas A&M University, Houston, TX.

4. ANSI/HI 9.6.1 – 2000, “Centrifugal and Vertical Pumpsfor NPSH Margin,” Hydraulic Institute, Parsippany, NJ.

5. ANSI/HI 9.6.3 – 2000, “Centrifugal and Vertical Pumpsfor Allowable Operating Region,” Hydraulic Institute,Parsippany, NJ


Pump Clinic 26 Origins of Sealless Pump Technology Page 1 of 13 29.07.2008

PUMP CLINIC 26

ORIGINS OF SEALLESS PUMP TECHNOLOGY

INTRODUCTION The origins of seal/less magnetically driven pumps date back to 1933 when the first known patent was granted in the UK.

Early commercial development of the magnetic drive pump was pioneered by Geoffrey Howard of HMD Pumps Limited, UK in the late 1940's and a few years later by Franz Klaus in West Germany.

This development was in response to a need for 100 percent containment of diphyl heating fluids. At that time, development of mechanical seals had barely started, and all dynamic seals were prone to leakage, especially at elevated temperatures. Two companies pioneered the use of magnet-drive pumps; Imperial Chemical Industries in the UK and Bayer in West Germany.

For the first 30 years, their application was limited essentially to pumping life threatening or extremely hazardous fluids. Because of the higher cost of the equipment, and possibly the stigma of some unsuccessful or unreliable products that came on the market, seal less pumps tended to be considered categorically by many as the solution of the last resort.

However, by the 1970's enough experience had been gained in the chemical processing industries to bring some engineers to the conclusion that the magnet drive pump had been developed to the point that it had become the most economical solution in many process systems.

ORIGINS OF SEALLESS PUMP TECHNOLOGY 1933 (UK): First known dated patent for design of a seal less magnetically-driven pump. (Only low power chrome and cobalt steel magnets were available; not commercially viable)

1939 (UK): First known patent for design of a seal less motor-driven pump (Used wet stator motor)

1941 (UK): First known production of seal less motor drive pump. (Wet stator design; proved to be commercially viable)

1947 (UK): First commercially viable magnetically driven seal less pump. (AINiCo magnets used; synchronous; required a soft start due to low magnetic powers)

1950-1953 (UK, USA, Europe): First known patents for seal less motor pumps with dry motor windings. (Dry stator design; tagged as "canned motor" commercially viable product)

1958 (UK): First known "induction" magnetically driven pump (Higher power magnets available; also known as "torque ring")

MAGNET DRIVE PUMPS THEORY OF OPERATION

The magnet drive pump, while unique when compared with conventionally designed centrifugal pumps, is a simple combination of standard components and proven concepts. Figure 1 depicts a typical magnet drive pump with a separately mounted electric motor drive. In this installation, the base, the electric motor and the motor coupling are identical to parts used in conventional pumps The differences between magnet drive pumps and conventional pumps occur in two areas:



1) Driving torque is transmitted magnetically rather than mechanically.

2) The impeller shaft rides in bushings housed within the pump enclosure rather than be bearings mounted externally.

In Figure 1, the drive motor is coupled directly to the outer magnet ring (OMR) by the motor coupling. The overhung load of the OMR is carried by bearings in the bearing housing. Figure 1 also shows that the pump impeller is mounted on the same shaft as is the inner magnet ring (IMR). (NOTE: In models with a non-synchronous drive, this is called a torque ring rather than an IMR because it does not contain magnets as we shall see later in this discussion). The impeller drive shaft is carried by two bushings which are within the pumping enclosure. You will note that the pump enclosure is formed by the pump casing and the containment shell. The driving torque of the electric motor is transmitted to the pump impeller by the magnetic coupling of the OMR and the IMR (or torque ring) without breaching the pumping enclosure. It is this magnetic coupling which replaces the mechanical seals of conventional centrifugal pumps.

MAGNETIC COUPLING

Before the theory of magnetic couplings as applied to sealless pumps can be examined, review of the fundamentals of magnets and electromagnetism is needed. Recall these basic principles:

1. Magnets have a north pole and a south pole. When two unlike poles are near each other, they are attracted. When two like poles are near each other, they are repelled.

2. When a magnetic field is moved past an electrical conductor which is in a closed loop, an electric current will flow in that loop.

3. When an electric current flows in a closed loop, an electromagnet is created with a north pole at one end of the loop(s) and a south pole at the other end of the loop(s).

4. An electromagnet is formed even if the "core" material is air. The electromagnet can be made stronger by inserting grain-oriented silicon steel into the "core" area.

5. Electromagnets behave just like permanent magnets with respect to the laws of attraction and repulsion.



TORQUE RING DRIVE

With these principles in mind, examine how the torque ring drive operates. The power supply shaft carries magnets secured to a rotating cylinder (also called OMR). The driven shaft has a steel hub or core in which a series of copper bars or rods are imbedded around its periphery, much like an induction motor rotor, except sheathed with a high alloy metal for corrosion resistance.

When the power supply shaft is rotated, the magnets sweep past the copper bars and induce electrical currents in the core. Remember, however, that our electromagnetic currents are induced or generated by the magnetic field sweeping past the conducting circuit. Therefore, there must be a slightly slower speed in the driven shaft (torque ring) than in the power supply shaft (OMR). The difference in speeds is called "slip".

Figure 2 is a schematic representation of the torque ring coupling. The outer magnet assembly is driven by a separately mounted motor. The OMR consists of a number of permanent magnets securely attached to a cylindrical frame, evenly distributed to provide a uniform magnetic field. The torque ring is made of a mild steel core with an outer facing of stainless steel or other metal compatible with the pumpage. Beneath this layer, a conductive metal (copper bars) is placed to provide an electromagnetic coupling circuit.

The containment shell is an extension of the pump casing (pressure casing) and therefore is a pressure containing component which completes the sealing off of the pumpage from the external environment. Thus the torque ring operates in the process pumpage while the OMR operates in the ambient atmosphere surrounding the pump. When the outer magnet ring rotates, the magnetic field passes through the containment shell, through the copper of the torque ring, through the mild steel beneath the torque ring and then returns to the OMR to complete the circle. The rotating magnetic field produces eddy currents in the copper and these eddy currents create electromagnets which tend to follow the rotating magnetic field which created them.

In the torque ring shown in Figure 2 a series of parallel copper strips are laid parallel to the pump shaft. In practice, these strips are separated and tied together at the ends, much like the "squirrel cage" of an a.c. motor. This is called a rodded torque ring. In any case, the copper conducting path(s) in the torque ring are firmly connected to the mild steel cylinder which, in turn, is solidly attached to the pump shaft.



Incidentally, the greater the "slip" in torque ring speed, the greater the eddy current flowing and the greater the torque. If a pumpage has high viscosity when it is cold, the eddy-current drive will provide high starting torques and will also provide greater heating of the pumpage (copper losses are higher at higher "slip" levels). This heating of the pumpage will be an advantage in liquids with high viscosity at cooler temperatures, getting the pumping operation under way more quickly than would be the case for a synchronous drive.

SYNCHRONOUS DRIVE Figure 3 shows a schematic representation of the synchronous drive coupling. Just as with the eddy-current coupling, the outer magnet assembly is driven by a separately mounted motor. The differences between torque ring and synchronous coupling occur within the inner ring. In the synchronous coupling, the IMR contains the same number of magnets as are mounted in the OMR.

The number of magnets is determined by the torque which must be transmitted. Thus, when the outer magnet ring rotates, the inner magnet ring rotates in synchronism with the outer magnet ring. The absence of "slip" means that the magnet to magnet coupling drive has a higher speed than the torque ring coupling drive.

The inner magnet ring is mounted on the same shaft as the pump's impeller. The containment shell is an extension of the pump casing and thus the IMR operated in the process pumpage while the OMR operates in the ambient atmosphere surrounding the pump. The OMR is enclosed by a coupling housing to protect it from dirt and to shield operating personnel from the high speed OMR.


Pump Clinic 26 Origins of Sealless Pump Technology Page 5 of 13 29/07/08

MAGNETIC COUPLING DESIGN

TORQUE RING SYNCHRONOUS Internal Composition Steel / Copper Rods Magnets Principal of Operation Slip Synchronous Limitation Power Torque

ADVANTAGES OF TORQUE RING DRIVE

• Temperature capabilities to 450°C • Excellent abuse factor ("slip" provides high resistance to decoupling) • High starting torque characteristics • Ferrite particles in pumpage do not build up • Well suited to cold (viscous) start • Inherently provides a "soft start"

ADVANTAGES OF SYNCHRONOUS COUPLING

• Output speed = Input speed (no slip) • Compact design: Greater power capability in smaller envelope than torque ring drive • Maximum efficiency at design flow (no slip) • Allows compliance with dimensional/performance standards (due to no slip)

MAGNETIC PERMANENCE There are three conditions which can alter the magnetic strength of permanent magnets. They are undue physical abuse, excessive temperatures and powerful extraneous magnetic fields.

PHYSICAL ABUSE Mechanical stress or shock has long been known to demagnetise steel bars. Modern high coercive force permanent magnets such as is used in magnet drive pumps, however, are generally insensitive to these mechanical degradations. Generally, a mechanical stress large enough to demagnetise modern permanent magnets would have to be so great that it would physically damage the magnet. For practically all applications, mechanical stresses can be ignored as contributing to instability.

HIGH TEMPERATURES In early magnetic materials, changes in magnetic structure could occur at room temperatures. Today's magnetic materials vary in sensitivity to very high temperatures. AINiCo magnets such as used in torque ring drives have Curie Temperatures (level above which they are useless as magnets) from 800°C to 900°C. Samarium Cobalt magnets begin permanently losing strength at around 350°C (depending on grade). Neodymium-Iron-Boron magnets begin permanently losing strength at around 120°C (again, depending on grade). In general, sintered "rare earth" magnetic materials' flux density is inversely related to their ability to withstand temperature.

HIGH MAGNETIC FIELDS To effect a magnet, a magnetic field must be stronger than the field used in the initial formation of the magnet. Since the fields used to create the magnet drive pump magnets are stronger than any fields found in plant environments, this mode of demagnetisation is all but eliminated. In other words, the magnets used today in magnet drive pumps are PERMANENT magnets. CONTAINMENT SHELL DESIGN CONSIDERATIONS IN MAGNET DRIVE PUMPS DESIGN CRITERIA • Containment shell design is influenced by: • Pump working pressure/temperature • Internal bearing support system • Minimising magnetic coupling losses



Engineers responsible for preparing specifications for sealless pumps sometimes critically review the design and manufacture of the containment shell to ensure a level of integrity in the pressure vessel boundary. Specifications generally dictate; • Conformance to relevant pressure vessel codes (eg. ASME VIII, AS) • Minimum containment shell tube thickness • Corrosion allowance • Number of welds • Avoidance of externally applied loads that would cause undue stress on the containment shell EDDY CURRENTS As the inner and outer components rotate, eddy currents are created by the rotating magnetic field cutting through the stationary containment shell, resulting in losses and reducing the overall drive efficiency. These eddy currents are a function of; • Speed² • Length of magnets in rotor • Diameter² of containment shell • Thickness of containment shell • Field strength² • Resistivity of shell material • Number of magnets The resistivity of containment shell materials commonly used is 80 x 10-3 ohm cm for 316 Stainless Steel and 130-3 ohm cm for Hastelloy C. This variation in resistivity between materials will result in a 62.5% increase in losses for the stainless steel assembly over the Hastelloy C design. Temperature losses similarly increase. CONSTRUCTION MATERIALS Several materials in addition to stainless steel and Hastelloy C are readily available for containment shell design to match the pump application. They include: Hastelloy B, Alloy 20, Inconel 718 and various space age alloys such as Nimonic 75 and Nimonic 90 for extraordinarily severe pressures. Recent advances in the manufacture of ceramic and plastic materials for pressure vessel containment have introduced these materials into magnetic drive pump design. Use of these materials effectively eliminates eddy current loss. However, the thickness required for the containment shell to contain design pressure, along with manufacturing tolerances, can result in drive size increases up to 50%. Corresponding increases in viscous frictional losses and material performance limitations have limited their application in magnet drive pumps. WELDED CONTAINMENT SHELL DESIGN Historically, manufacturers of magnet drive pumps have manufactured containment shells using plate for the containment shell flange/shells and rolled tube for the shell tube. Three welds are required to manufacture the containment shell. All welds are performed by qualified welders and thoroughly inspected with qualified procedures from an independent body. Extensive fixturing and further non destructive testing (NDT) procedures are carried out to ensure the integrity of the containment shell. However, one disadvantage of using welds in the manufacturing process is that it may be considered a weak point in the component and a source for corrosion susceptibility. HYDROFORMED (STAMPED) CONTAINMENT SHELL DESIGN An alternative is the hydroformed containment shell. During the forming process, a plate of the material specified is shaped by an external load, and several forms are required to achieve the final design. Also,



because the material work hardens during the process, it is necessary to heat treat the component prior to final sizing. The final fabrication is completed using only one weld. Potential problem areas in this design include uniformity of thickness and accommodation of the increased length due to rounded end shape.

Metallic Non-Metallic Typical Materials of Construction Hast-C / 316SS Ceramic, ETFE, Polypropylene,

PTFE

Eddy Current Losses Yes No

Temperature Sensitivity Low High

Pressure Sensitivity Low High

Abrasive Sensitivity Low High

OUTER MAGNET RING CONSTRUCTION CHALLENGES OF OMR CONSTRUCTION Outer magnet rings are subject to a variety of stresses including rapid rotation, torque loading, and attraction to the inner magnet ring. The two primary challenges in OMR construction involve (1) retention of the magnet in the outer assembly and (2) fragility of the magnet material - rare earth magnetic materials widely used today are sintered powders with little inherent strength.



DISADVANTAGE OF ADHESIVE BONDING For rare earth assemblies, the easiest and cheapest way to secure magnets is with an adhesive. While acceptable for many applications, severe problems can result if repeated temperature cycling occurs or if the adhesive is contaminated by exposure to certain process liquids or solvents. If the adhesive bond fails, magnets can drop from the assembly and act as a cutting tool, attacking the containment shell. The result will be dramatic: vibration from loss of balance; heat rise from mechanical friction; trepanning and ultimate breaching of the containment shell; costly shutdown and repairs. Since there is no method for monitoring the integrity of an adhesive bond, failure cannot be predicted. MAGNET MATERIAL BRITTLENESS CAUSES FAILURE Another potential source of trouble is the inherent brittleness of the sintered, low- strength rare earth magnet material. If roughly handled during pump or OMR assembly, this material easily chips and flakes. This will almost certainly happen if OMR removal occurs prior to the removal of the inner magnet ring. An undetected chip could become dislodged during use and cut the containment shell, with similar results to those described for adhesive bonding failure. MAGNET RETENTION Some manufacturers with rare earth OMR assemblies have the magnets mechanically retained in the body and totally enclosed. The magnets cannot be seen, and it is not possible to damage them. Should the magnet become damaged while the OMR is being assembled, all flakes and chips will be fully enclosed. There are no components that can degrade, and any mechanical damage to the assembly can easily be seen by the assembler. Other manufacturers use epoxy or other compounds as means of retention. A potential issue with this process is that epoxy degradation and loss of strength may cause magnets to dislodge form the OMR. BUSHINGS AND BEARINGS Possibly one of the reasons that magnet drive pumps have not been more fully utilised is the concern of some engineers over having bearings (bushings) in the process fluid. Recognising that the application of these pumps is presently on relatively clean liquids of low viscosity, a category into which the majority of pumping applications fall, it may be reasonable to make comparisons between "bushing and shaft life" in a magnet drive pump and "shaft, mechanical seal and antifriction bearing life" in conventional sealed pumps in order to quantify this concern. This would seem to be functionally correct way to make one comparison between sealed and sealless pumps. The minimum rated life of ANSI B73.1 pump bearings is 17,500 hours at maximum load. Mechancial seal life varies widely, but two to three years of life would generally be considered excellent. Feedback from many maintenance engineers has painted a picture of the combination of bearings and seals seldom approaching two to three years of life in the real world. Compare this experience with the operating experience of magnet drive pumps. Spare parts order records of one manufacturer and field reports support the conclusion that average bushing life is three to five years in typical magnet drive pump service. Ten years operation with original bushings has been achieved in a number of instances. Perhaps some of the concern over internal bushing life comes from over 25 years of experience with canned motor pumps where inherent close clearances cannot provide the longer term wearing capability which is common in magnet drive pumps. By far the most common bushing material in use in magnet drive pumps today is special plain carbon. Bushings are sometimes pressed but most often are an interference fit to handle higher temperatures, pinned for the highest temperature services. Carbon is consistent in performance with a fluid and has good lubricating qualities that will normally enable a pump to survive a period of dry running, provided it is short enough so that the pump itself is not damaged. (Dry running is a fact of life with most pumps at one time or another). With essentially no binder to be attacked, carbon is suitable for all but a few services throughout the range of magnet drive pumps. Bushings are provided with spiral grooves of about 1/8 in diameter on the ID to permit solids to pass through the bushings. They are combined with hardened shaft journals, except where bushings are lightly loaded and the hardening is not required to obtain satisfactory shaft life.



Acids such as sulphuric acid and nitric acid do not attack plain carbon, but electrolytic damage makes many carbons unsuitable. Filled PTFE versus 316 SS has proven to be satisfactory in these services (although not as consistent as carbon), has good lubricating properties and will survive short duration dry running. Shorter maintenance intervals must be planned when this material is initiated because of the inconsistency cited. PTFE bushings may be carbon filled, glass filled or mica filled, each having different chemical compatibility. The temperature range of PTFE as bushing bearing is limited to 120°C whereas carbon is suitable for the full range of magnet drive service, -40 to +450°C. For abrasive service, unusually high bearing loads, and corrosive services where PTFE is inadequate, silicon carbide bushings and journals are recommended. They have the advantages of extremely high load capacity compared with carbon or PTFE, and they are more tolerant of solids and abrasive materials. The negative aspects of their use are: 1. Higher cost 2. Complexity of detail design, because of the need to provide for different coefficients of thermal

expansion with this very brittle material

and most significantly, 3. The inability to tolerate even a short period of dry running, which makes it much less abuse resistant

than other materials described. Fragments of a failed silicon carbide bushing can cause extensive damage in a pump.

However, it has generally been accepted that silicon carbide offers the best long term bearing material for applications other than high temperature and has become the most widely used bearing material in HMD pumps. INTERNAL BEARING MATERIAL Carbon Graphite Teflon SIC Dry running capability Good Good Poor Solids handling capability Fair Poor Excellent Thrust handling capability Fair Poor Excellent Complexity of design Low Low High Cost Low Low High

COUPLING RECIRCULATION SYSTEM INTRODUCTION The recirculation flow for any magnet drive pump has three basic functions. These are: 1. Removal of the heat generated from magnetic losses 2. Lubrication of the internal radial and thrust bearings 3. Thrust balancing of the free floating rotating assembly If the liquid being pumped is "dirty", recirculation flow must perform a fourth function. 4. Removal or flushing of solids through magnetic coupling and internal bearings. To successfully meet these requirements, the recirculated pumpage must remain in the liquid phase at all points within the magnetic coupling area. The determining factors for preventing phase change are the mass flow rate, heat capacity, and localised pressure at any point within the coupling. The type of recirculation flow design will have impact on the variables of mass flow rate and localised pressure.



AVAILABLE DESIGNS There are two standard recirculation flow designs commercially available in magnet drive pumps (more specialised recirculation flow paths are available as an option); a) Discharge-to-suction Fluid flow enters the magnetic coupling area at a high pressure discharge point and returns to the bulk flow at the suction eye of the impeller. b) Discharge-to-discharge Fluid flow enters the magnetic coupling area at a high pressure discharge point and returns to the bulk flow at a point behind the rear shroud of the impeller. Each design has its advantages and must be considered on an individual basis. DISCHARGE-TO-SUCTION RECIRCULATION The discharge-to-suction design (Fig. 5) involves pulling a slip stream from the high pressure point of the casing and returning it to the bulk flow at the suction eye of the impeller. The flow is routed to the suction through either the thrust balance holes in the impeller or, in certain designs, through a hole gun drilled along the axis of the pump shaft. The recirculated fluid is driven by the differential pressure between the recirculation inlet and return locations. Additional pumping action is provided by the rotation of the internal magnetic coupling components. The total dynamic head (TDH) generated by the pump provides the differential pressure that drives the recirculation flow. As the differential pressure increases, internal flow rate increases but at a decreasing rate. The internal flow will reach a maximum beyond which any additional increase in differential pressure will have negligible impact. This occurs when the friction losses begin to become the dominant factor effecting flow. The observed internal pumping effects are primarily caused by the action of the inner magnet ring and thrust washers. These components operate at high speeds within tight clearances and behave as rudimentary impellers. Discharge-to-suction recirculation yields a flow pattern that is characterised by high fluid velocities. This high velocity profile provides advantages that will be discussed later in this analysis.



DISCHARGE-TO-DISCHARGE RECIRCULATION The discharge-to-discharge design (Fig. 6) involves pulling a slip stream from the high pressure point of the casing (discharge pressure) and returning this recirculation flow to the bulk flow at a point behind the rear impeller shroud adjacent to the impeller boss. The recirculated fluid is driven by the differential pressure between the recirculation inlet and return locations, however, this pressure differential is not as great as that in the discharge-to-suction system. To insure proper cooling of the magnetic coupling, larger flow passages are provided. The rotation of the internal magnetic coupling components also provides additional pumping action. Discharge-to-discharge recirculation yields a flow pattern that is characterised by high localised pressure and little interference with suction flow. These characteristics also provide advantages that will be discussed in the following section. COMPARATIVE ANALYSIS In comparing the benefits of discharge-to-suction and discharge-to-discharge recirculation, both systems remove the heat generated from the magnetic coupling and lubricate the internal radial and thrust bearings equally well. While discharge-to-suction recirculation creates a higher recirculation velocity due to its high "driving" differential pressure, its mass flow is comparable to that of discharge-to-discharge recirculation due to the latter's large flow path. Internal pressures in both systems are such that "flashing" at the magnetic coupling interface or internal bearings are avoided with most liquids. Discharge-to-suction recirculation tends to have better impeller thrust balancing characteristics than the discharge-to-discharge system due to its routing of the flow through the impeller eye balance holes. This advantage is, however, minimal, and with the growing use of silicon carbide thrust bearings it is not a significant issue. Discharge-to-suction recirculation also tends to flush solids better due to its higher velocities. In general, however, the handling of solids can be increased through the use of silicon carbide bearings, since these bearings are less affected by the abrasive nature of most commonly encountered solids. A pump equipped with discharge-to-discharge recirculation typically requires less NPSH. Lower NPSH characteristics are achieved by routing the recirculation return flow to the rear of the impeller so that fluid flow through the suction eye is not interfered with. Discharge-to-discharge recirculation also eliminates the chance of "flashing" at the impeller eye. Routing the recirculation return flow which has been heated by magnetic coupling losses and internal bearing friction to the low pressure, suction eye of the pump may cause flashing with certain fluids. By routing return flow to the higher pressure, behind the impeller, the potential for flashing is eliminated. FUNCTIONS OF RECIRCULATION

• Removes heat from magnetic coupling • Lubricates internal bearings • Assists in balancing thrust loads

ALSO • Must pass solids contained within liquid

COMPARISON OF RECIRCULATION DESIGN Discharge to Discharge Discharge to Suction NPSHR Low High Internal flow rate Medium High Potential for flashing Low Medium



SOLIDS HANDLING CAPABILITY The advent of silicon carbide bearings which will grind up particles of a lesser hardness has resulted in wild claims for the solids handling capabilities of sealless pumps. It is true that the presence of solids can be less problematic with silicon carbide bearings than with other bearing options such as carbon and PTFE. However, the upper percentage and size limits have changed little since these limits are dictated by non bearing-related parameters. SOLIDS HANDLING Many years ago, solids limit for the magnet drive pumps was uncertain and generally set at 1.5 percent up to 150 microns, based on two important considerations: 1. The appreciation that although solids and seal less do not mix, there is always some solids in the

pumpage. The 1.5 percent figure was felt to be low enough to suggest caution but high enough to allow the presence of pipe scale and occasional pull over of filtrate or crystallate.

2. The result of some lab testing and tests carried out by customers on various pumps, indicate that wear to carbon bearings increases rapidly above 2 percent of solids.

In instances where solids larger than 150 microns were present in small quantities, a rough in line filter would be fitted on the feed (from pump discharge) to the magnetic drive. PAST CONVENTIONS BEING RE-EVALUATED Because of increased competitive pressure and a greater understanding of internal flow in the magnetic drive, re-evaluations of past conventions have been undertaken. Comprehensive test programs to establish definitive numbers for allowable percentage and size of solids have been carried out by various manufacturers. These programs take into account the nature of solids; (ie: sticky or fibrous solids may block flow channels). As much information as possible on the type of solids should be obtained from the customers before pump selection is made. Other considerations are that very abrasive solids will eventually wear metallic parts, that wear-resistant materials may require selection to give maximum pump life, and that wear in the impeller will be at a minimum when the pump operates near to its best efficiency (design) point. At low flows recirculation will accelerate wear, and at higher flow rates solids will increase wear because of velocity and incorrect incidence angles to impeller blades. GENERAL GUIDELINES The following guidelines are generally recognized but specific pump limits must always be confirmed with the manufacturer: 1. Between 3-5% solids A standard pump fitted with silicon carbide bearings will handle between 3 and 5 percent solids to 150 microns. The limit on particle size is dictated by bearing clearances. Higher particle size may be screened out with a filter between the pump discharge and the magnetic drive. 2. Up to 30% solids Up to 30 percent solids may be pumped up to 750 microns (wear ring clearances) if a clean flush is provided to the magnetic drive or if a closed-loop system is fitted (available only from a small number of manufacturers) to separate the magnetic drive from the pump head. For a clean flush, 10 to 25 l/m may be required for cooling. With a closed-loop system, 7 to 15 1/hr is required from a pressurised supply. Pumping solids of 30 percent concentration would dictate the use of hardened steel or iron if useful pump life is to be obtained.



3. Magnetic solids For highly magnetic solids, the torque ring drive which does not have magnets immersed in the liquid is recommended. Modern magnets are extremely powerful and will attract ferrite particles which will build up between the inner magnet and containment shell. If the torque ring drive is not available for the particular model selected, then a magnetic filter may be fitted between the pump discharge and magnetic coupling. 4. Definitive data not yet available All pump manufacturers would like to publish definitive data on the solids handling capabilities of their products. However, there is considerable variety in size, hardness, chemistry, solubility, abrasiveness and flocculation of solid particles. Each solids handling application requires evaluation based in testing on experience. Conventional pumps have generally been developed for specific applications on the basis of experience (ie: coal slurry pumps, paper stock etc). The same is true for the sealless market with customer/supplier cooperation to mutually resolve pumping problems in this difficult area.


Pump Clinic 27 Positive Displacement Pumps Page 1 of 15 24/09/08

PUMP CLINIC 27

POSITIVE DISPLACEMENT PUMPS

This article has been developed from a variety of sources including manufacturers, industry trade organisations, internet articles and common PD industry knowledge.

By definition, positive-displacement (PD) pumps displace a known quantity of liquid with each revolution of the pumping elements. This is done by trapping liquid between the pumping elements and a stationary casing. Pumping element designs include gears, lobes, rotary pistons, vanes, screws and hoses.

PD pumps are found in a wide range of applications -- chemical-processing; liquid delivery; marine; biotechnology; pharmaceutical; as well as food, dairy, and beverage processing. Their versatility and popularity is due in part to their relatively compact design, high-viscosity performance, continuous flow regardless of differential pressure, and ability to handle high differential pressure.

Positive displacement (PD) pumps are divided into two broad classifications, reciprocating and rotary (Figure 1). This article covers rotary pumping principles.

Figure 1

By definition, PD pumps displace a known quantity of liquid with each revolution of the pumping elements (i.e., gears, rotors, screws, vanes). PD pumps displace liquid by creating a space between the pumping elements and trapping liquid in the space. The rotation of the pumping elements then



000,000cP.

reduces the size of the space and moves the liquid out of the pump. The broad category of PD pumps is able to handle fluids of all viscosities up to 1,320,000 cSt / 6,000,000 SSU, capacities up to 1,150 M3/Hr (delete) and pressures up to 700 BAR (delete). Rotary pumps are self-priming and deliver a constant, (delete) flowrate, regardless of pressure variations.

Selection of a positive displacement (PD) rotary pump is not always an easy choice. There are many types of PD pumps available. In this article, we cover the more common ones: internal gear, external gear, timed lobe, vane, screw and peristaltic. Most PD pumps can be adapted to handle a wide range of applications, but some types are better suited than others for a given set of circumstances.

The first consideration in any application is pumping conditions. Usually the need for a PD pump is already determined, such as a requirement for a given amount of flow regardless of differential pressure, viscosity too high for a centrifugal pump, need for high differential pressure, or other factors.

Inlet conditions, required flow rate, differential pressure, temperature, particle size in the liquid, abrasive characteristics, and corrosiveness of the liquid must be determined before a pump selection is made.

A pump needs proper suction conditions to work well. PD pumps are often self-priming, and it is often assumed that suction conditions are not important. But they are. Each PD pump has a minimum inlet pressure requirement to fill individual pump cavities. If these cavities are not completely filled, total pump flow is diminished. Pump manufacturers supply information on minimum inlet conditions required. If high lift or high vacuum inlet conditions exist, special attention must be paid to the suction side of the pump. INTERNAL GEAR PUMP OVERVIEW

Internal gear pumps are exceptionally versatile. While they are often used on thin liquids such as solvents and fuel oil, they excel at efficiently pumping highly viscous liquids such as asphalt, chocolate, and adhesives. The useful viscosity range of an internal gear pump is from 1cPs to over 1,

In addition to their wide viscosity range, internal gear pumps have a wide temperature range as well, handling liquids up to (delete) 400°C. This is due to the single point of end clearance (the distance between the ends of the rotor gear teeth

and the head of the pump). This clearance is adjustable to accommodate high temperature, maximize efficiency for handling high viscosity liquids, and to accommodate for wear.

The internal gear pump is non-pulsing, has some self-priming capability, and can run dry for short periods. They're also usually bi-rotational, meaning that the same pump can be used to load and unload vessels. Because internal gear pumps have only two moving parts, they are reliable, simple to operate, and easy to maintain.

How Internal Gear Pumps Work

1. Liquid enters the suction port between the rotor (large exterior gear) and idler (small interior gear) teeth. The arrows indicate the direction of the pump and liquid.

2. Liquid travels through the pump between the teeth of the "gear-within-a-gear" principle. The crescent shape divides the liquid and acts as a seal between the suction and discharge ports.

3. The pump head is now nearly flooded, just prior to forcing the liquid out of the discharge port. Intermeshing gears of the idler and rotor form locked pockets for the liquid which assures volume control.

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4. Rotor and idler teeth mesh completely to form a seal equidistant from the discharge and suction ports. This seal forces the liquid out of the discharge port.

The crescent internal gear pump has an outer or rotor gear that is generally used to drive the inner or idler gear (Figure 1).

The idler gear, which is smaller than the rotor gear, rotates on a stationary pin and operates inside the rotor gear. The gears create voids as they come out of mesh and liquid flows into the pump. As the gears come back into mesh, volumes are reduced and liquid is forced out the discharge port. Liquid can enter the expanding cavities through the rotor teeth or recessed areas on the head, alongside the teeth. The crescent is integral with the pump head and prevents liquids from flowing to the suction port from the discharge port.

The rotor gear is driven by a shaft supported by journal or antifriction bearings. The idler gear contains a journal bearing rotating on a stationary pin in the pumped liquid. Depending on shaft sealing arrangements, the rotor shaft support bearings may run in pumped liquid. This is an important consideration when handling an abrasive liquid as it can wear out a support bearing.

Figure 1. Internal gear pumps are ideal for high-viscosity liquids, but they are damaged when pumping large solids.

The speed of internal gear pumps is considered relatively slow compared to centrifugal types. Speeds up to 1450 rpm are considered common, although some small designs operate up to 3,450 rpm. Because of their ability to operate at low speeds, internal gear pumps are well suited for high-viscosity applications and where suction conditions call for a pump with minimal inlet pressure requirements.

For each revolution of an internal gear pump, the gears have a fairly long time to come out of mesh allowing the spaces between gear teeth to completely fill and not cavitate. Internal gear pumps have successfully pumped liquids with viscosities above 1,320,000 cSt / 6,000,000 SSU and very low viscosity liquids, such as liquid propane and ammonia.

Internal gear pumps are made to close tolerances and are damaged when pumping large solids. These pumps can handle small suspended particulate in abrasive applications, but gradually wear and lose performance. Some performance loss is restored by adjusting the pump end clearance. End clearance is the closeness of the rotor gear to the head of the pump

Advantages• Only two moving parts • Only one stuffing box • Non-pulsating discharge • Excellent for high-viscosity liquids • Constant and even discharge regardless

of pressure conditions • Operates well in either direction • Can be made to operate with one direction

of flow with either rotation • Low NPSH required • Single adjustable end clearance • Easy to maintain • Flexible design offers application

customization

Disadvantages• Usually requires moderate speeds • Medium pressure limitations • One bearing runs in the product pumped • Overhung load on shaft bearing

Applications



GEAR PUMP OVERVIEW

ines.

r extreme high temperature applications.

nd for precise transfer and

metering applications involving polymers, fuels, and chemical additives.

Common internal gear pump applications include, but are not limited to:

• All varieties of fuel oil and lube oil

• Resins and Polymers

• Alcohols and solvents

• Asphalt, Bitumen, and Tar

• Polyurethane foam (Isocyanate and polyol)

• Food products such as corn syrup, chocolate, and peanut butter

• Paint, inks, and pigments

• Soaps and surfactants

• Glycol

Materials Of Construction / Configuration Options

• Externals (head, casing, bracket) - Cast iron, ductile iron, steel, stainless steel, Alloy 20, and higher alloys.

• Internals (rotor, idler) - Cast iron, ductile iron, steel, stainless steel, Alloy 20, and higher alloys.

• Bushing - Carbon graphite, bronze, silicon carbide, tungsten carbide, ceramic, colmonoy, and other specials materials as needed.

• Shaft Seal - Lip seals, component mechanical seals, industry-standard cartridge mechanical seals, gas barrier seals, magnetically-driven pumps.

• Packing - Impregnated packing, if seal not required.

EXTERNAL

External gear pumps are a popular style of pump and are often used as lubrication pumps in machine tools, in fluid power transfer units, and as oil pumps in eng

External gear pumps can come in single or double (two sets of gears) pump configurations with spur (shown), helical, and herringbone gears. Helical and herringbone gears typically offer a smoother flow than spur gears, although all gear types are relatively smooth. Large-capacity external gear pumps typically use helical or herringbone gears. Small external gear pumps usually operate at speeds up to 3000 rpm and larger models operate at speeds up to 640 rpm. External gear pumps have close tolerances and shaft support on both sides of the gears. This allows them to run to

pressures beyond (delete) 200 BAR, making them well suited for use in hydraulics. With four bearings in the liquid and tight tolerances, they are not well suited to handling abrasive o

Tighter internal clearances provide for a more reliable measure of liquid passing through a pump afor greater flow control. Because of this, external gear pumps are popular



o

ne

drives the other gear. Each gear is supported by a shaft with bearings on both sides of the gear.

let side of the pump. Liquid flows into the cavity and is trapped by the gear teeth as they rotate.

sing in the pockets between the teeth and the casing -- it does not pass between the gears.

3. Finally, the meshing of the gears forces liquid through the outlet port under pressure.

sually, small external gear pumps 3000 rpm and larger versions operate at

speeds up to 640 rpm.

very

. When an external gear pump wears, it must be rebuilt or replaced.

must be slowed down considerably when pumping viscous liquids.

strength may not be ade cturers supply torque en it is a factor.

ages

• modates wide variety s

uid area

• Fixed End Clearances

Common external gear pump applications include, but are not limited to:

• Various fuel oils and lube oils

How External Gear Pumps W rk

External gear pumps are similar in pumping action to internal gear pumps in that two gears come into and out of mesh to produce flow. However, the external gear pump uses two identical gears rotating against each other -- ogear is driven by a motor and it in turn

1. As the gears come out of mesh, they create expanding volume on the in

2. Liquid travels around the interior of the ca

Because the gears are supported on both sides, external gear pumps are quiet-running and are routinely used for high-pressure duties such as hydraulic applications. With no overhung bearing loads, the rotor shaft can not deflect and cause premature wear. Uoperate at speeds up to

The design of external gear pumps allows them to be made to closer tolerances than internal gear pumps. The pump is not forgiving of particulate in the pumped liquid. Since there are clearances at both ends of the gears, there is no end clearance adjustment for wear

External gear pumps handle viscous and watery-type liquids, but speed must be properly set for thick liquids. Gear teeth come out ofmesh a short time, and viscous liquids need time to fill the spaces between gear teeth. As a result, pump speed

The pump does not perform well under critical suction conditions. Volatile liquids tend to vaporize locally as gear teeth spaces expand rapidly. When the viscosity of pumped liquids rises, torque requirements also rise, and pump shaft

Figure 2. External gear pumps(shown is a double pump) are typically used for high-pressure

applications such as hydraulics.

quate. Pump manufa limit information wh

Advant• High speed • High pressure • No overhung bearing loads• Relatively quiet operation

Design accom

Disadvantages• Four bushings in liq• No solids allowed

of material

Applications

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n-

How Lobe Pumps Work

ox, aft

1. As the lobes come out of mesh, they create expanding volume on the inlet side of the pump.

2. Liquid travels around the interior of the casing in the pockets between the lobes and the casing -- it

3. Finally, the meshing of the lobes forces liquid through the outlet port under pressure.

• Chemical additive and polymer metering

• Chemical mixing and blending (double pump)

• Industrial and mobile hydraulic applications (log splitters, lifts, etc.)

• Acids and caustic (stainless steel or composite construction)

• Low volume transfer or application

LOBE PUMP OVERVIEW

Lobe pumps are used in a variety of industries including, pulp and paper, chemical, food, beverage, pharmaceutical, and biotechnology. They are popular in these diverse industries because they offer superb sanitary qualities, high efficiency, reliability, corrosion resistance, and good clean-in-place and sterilize-in-place (CIP/SIP) characteristics.

These pumps offer a variety of lobe options including single, bi-wing, tri-lobe (shown), and multi-lobe. Rotary lobe pumps are nocontacting and have large pumping chambers, allowing them to handle solids such as cherries or olives without damage. They are also used to handle slurries, pastes, and a wide variety of other liquids. If wetted, they offer self-priming performance. A gentle

pumping action minimizes product degradation. They also offer reversible flows and can operate dry for long periods of time. Flow is relatively independent of changes in process pressure, so output is constant and continuous.

Rotary lobe pumps range from industrial designs to sanitary designs. The sanitary designs break down further depending on the service and specific sanitary requirements. These requirements include 3-A, EHEDG, and USDA. The manufacturer can tell you which certifications, if any, their rotary lobe pump meets.

Lobe pumps are similar to external gear pumps in operationin that fluid flows around the interior of the casing. Unlike external gear pumps, however, the lobes do not make contact. Lobe contact is prevented by external timing gears located in the gearbox. Pump shaft support bearings are located in the gearband since the bearings are out of the pumped liquid, pressure is limited by bearing location and shdeflection.

Liquid flows into the cavity and is trapped by the lobes as they rotate.

does not pass between the lobes.

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Lobe pumps (Figure 3) are similar to external gear pumps in operation, except the pumping elements or lobes do not make contact.

Pump shaft support bearings are located in the timing gear case. Since the bearings are out of the pumped liquid, pressure is limited by bearing location and shaft deflection. There is not metal-to-metal contact and wear in abrasive applications is minimal. Use of multiple mechanical seals makes seal construction important.

Lobe pumps are frequently used in food applications, because they handle solids without damaging the pump. Particle size pumped can be much larger in lobe pumps than in other PD types. Since the lobes do not make contact, and clearances are not as close as in other PD pumps, this design handles low viscosity liquids with diminished performance. Loading characteristics are not as good as other designs, and suction ability is low. High-viscosity liquids require considerably reduced speeds to achieve satisfactory performance. Reductions of 25% of rated speed and lower are common with high-viscosity liquids.

Figure 3. Lobes in lobe pumps do not make contact, because they are driven by external timing gears. This design handles low-viscosity liquids.

Lobe pumps are cleaned by circulating a fluid through them. Cleaning is important when the product cannot remain in the pumps for sanitary reasons or when products of different colors or properties are batched.

Advantages• Pass medium solids • No metal-to-metal contact • Superior CIP/SIP capabilities • Long term dry run (with lubrication to

seals) • Non-pulsating discharge

Disadvantages• Requires timing gears • Requires two seals • Reduced lift with thin liquids

Applications

Primary applications for rotary lobe pump applications are with food, beverage, pharmaceutical and personal care products i.e in industries deemed to be ‘clean” industries.

Lobe pumps may also be used in other industrial applications as detailed below however specific care needs to be taken in these applications and manufacturers recommendations should be sought.

• Polymers

• Paper coatings

• Soaps and surfactants

• Paints and dyes

• Rubber and adhesives

• Pharmaceuticals

• Food applications (a sample of these is referenced below)


• Externals (head, casing) - Typically 316 or 316L stainless steel head and casing



nal ance,

60°C

vane pumps).

apor.

nd

d ics common to the vane pumping principle.

• Externals (gearbox) - Cast iron, stainless steel

• Internals (rotors, shaft) - Typically 316 or 316L stainless steel, non-galling stainless steel

• Shaft Seal - O-rings, component single or double mechanical seals, industry-standard cartridge mechanical seals

VANE PUMP OVERVIEW

While vane pumps can handle moderate viscosity liquids, they excel at handling low viscosity liquids such as LP gas (propane), ammonia, solvents, alcohol, fuel oils, gasoline, and refrigerants. Vane pumps have no internal metal-to-metal contact and self-compensate for wear, enabling them to maintain peak performance on these non-lubricating liquids. Though efficiency drops quickly, they can be used up to 500 cPs / 2,300 SSU.

Vane pumps are available in a number of vane configurations including sliding vane (left), flexible vane, swinging vane, rolling vane, and extervane. Vane pumps are noted for their dry priming, ease of maintenand good suction characteristics over the life of the pump. Moreover, vanes can usually handle fluid temperatures from -32°C (delete) to 2(delete) and differential pressures to 15 BAR (delete) (higher for hydraulic

Each type of vane pump offers unique advantages. For example, external vane pumps can handle large solids. Flexible vane pumps, on the other hand, can only handle small solids but create good vacuum. Sliding vane pumps can run dry for short periods of time and handle small amounts of v

How Vane Pumps Work

Despite the different configurations, most vane pumps operate under the same general principle described below.

1. A slotted rotor is eccentrically supported in a cycloidal cam. The rotor is located close to the wall of the cam so a crescent-shaped cavity is formed. The rotor is sealed into the cam by two sideplates. Vanes or blades fit within the slots of the impeller. As the rotor rotates (yellow arrow) afluid enters the pump, centrifugal force, hydraulic pressure, and/or pushrods push the vanes to thewalls of the housing. The tight seal among the vanes, rotor, cam, and sideplate is the key to the goosuction characterist

2. The housing and cam force fluid into the pumping chamber through holes in the cam (small red arrow on the bottom of the pump). Fluid enters the pockets created by the vanes, rotor, cam, and sideplate.

3. As the rotor continues around, the vanes sweep the fluid to the opposite side of the crescent where it is squeezed through discharge holes of the cam as the vane approaches the point of the crescent (small red arrow on the side of the pump). Fluid then exits the discharge port.

Vane pumps (Figure 4) operate quite differently from gear and lobe types.



A rotor with radial slots, is positioned off-center in a housing bore. Vanes that fit closely in rotor slots slide in and out as the rotor turns. Vane action is aided by centrifugal force, hydraulic pressure, or pushrods. Pumping action is caused by the expanding and contracting volumes contained by the rotor, vanes, and housing.

Figure 4. Vane pumps have better dry priming capability than other positive displacement pumps.

Vanes are the main sealing element between the suction and discharge ports and are usually made of a nonmetallic composite material. Rotor bushings run in the pumped liquid or are isolated by seals.

Vane pumps usually operate at 1,000 rpm, but also run at 1,450 rpm. The pumps work well with low-viscosity liquids that easily fill the cavities and provide good suction characteristics. Speeds must be reduced dramatically for high-viscosity applications to load the area underneath the vanes. These applications require stronger-than-normal vane material.

Because there is no metal-to-metal contact, these pumps are frequently used with low-viscosity non-lubricating liquids such as propane or solvent. This type of pump has better dry priming capability than other PD pumps. Vane pumps can run dry, but are subject to vane wear.

Vane pumps are not well suited to handling abrasive applications. Vane pumps have fixed end clearances on both sides of the rotor and vanes similar to external gear pumps. Once wear occurs, this clearance cannot be adjusted, but some manufacturers supply replaceable or reversible end plates.

Advantages• Handles thin liquids at relatively higher

pressures • Compensates for wear through vane

extension • Sometimes preferred for solvents, LPG • Can run dry for short periods • Can have one seal or stuffing box • Develops good vacuum

Disadvantages• Can have two stuffing boxes • Complex housing and many parts • Not suitable for high pressures • Not suitable for high viscosity • Not good with abrasives


• Externals (head, casing) - Cast iron, ductile iron, steel, and stainless steel.

• Vane, Pushrods - Carbon graphite, PEEK®.

• End Plates - Carbon graphite

• Shaft Seal - Component mechanical seals, industry-standard cartridge mechanical seals, and magnetically-driven pumps.

• Packing - Available from some vendors, but not usually recommended for thin liquid service

SCREW PUMP OVERVIEW



Within the rotary pumps family, single, twin and three screw pumps have earned reputations for some specific and significant applications.. Single-screw pumps are also known as progressive cavity or helical rotor pumps have some limitations and require some care in application. Multiple screw pumps can handle high pressure, temperature, speed and power combined. While other pump types can handle these variables well individually, their combined force is a challenge.

Progressive Cavity (PC) Pumps

The PC pump is made of three major sections:

A) The pumping element B) The suction housing C) The drive train

The pumping element is made from the rotor and stator elements. Normally, the rotor is made of steel or other metal and has the shape of a single helix (external shape). The stator is normally made from an elastomer and has the shape of a double helix (internal shape). The rotor is manufactured slightly larger than the stator so an interference fit exists when the rotor is inserted into the stator.

Some design enhancements could include features to make maintenance simpler and reliability better. A close-coupled, or so-called "block" design, results in a smaller pump package, less upfront cost and no drive alignment issues. Sealed pivot style universal joints (as in the above illustration) keep the joints lubricated.

Easy access to the mechanical seal is important to simplify seal service and reduce downtime with quick changeover. However, packings are still more accepted for sealing fluids in typical PC applications (as they rarely pump tough enough chemicals to require mechanical seals). When equipped with augers, PC pumps produce better NPSHR values with higher volumetric efficiencies and higher percent solids capabilities. Oversized open hopper inlets handle thicker liquids and eliminate bridging (for example, filter cake up to 55 percent solids can be handled by the auger-augmented PC pumps).

Improvements in the pumping element go beyond more traditional 1:2 geometry (single rotor lobe) to multi-lobe configurations, such as 2:3 (rotor/stator lobes) geometry. A multi-lobe design can increase flow per revolution and reduce initial pump cost. Equal wall stator doubles pressure capability per stage as compared to standard designs with constant stator outside wall thickness.

Tie rod construction makes the entire assembly much easier to service, as compared to more conventional threaded stator designs. Hollow cast rotors reduce inertia and result in lower vibrations. Coated chromed rotors resist wear and last longer.



A temperature probe, installed at the stator wall, can prevent rapid temperature rise and failure during dry running. In practice, few installations take advantage of this feature, since many maintenance departments tend to prefer simpler designs with fewer "gadgets."

Progressive cavity pumps have smooth output flow and good self-priming ability. Capable of pumping both thick and thin fluids, they are successful pumping liquids with high solids and abrasives content.

These capabilities have made progressive cavity pumps a choice for many tough applications. They work well in the wastewater treatment industry, but they perform equally well at the "opposite end of the spectrum" (the food industry) due to their minimal impact on shear sensitive fluids, such as sauces, cream products and similar fluids.

Through the years, advances in pump design, electronic monitoring and materials of construction have improved PC pump energy efficiency, decreased maintenance requirements and allowed them to handle more severe application conditions.

However, some designs actually have a clearance between these, referred to as a "single undersize rotor," "double-undersized" and even "triple-undersized." The designs with clearance between the rotating elements are rare (interference fits are probably common in 99 percent of applications). When designed with a clearance, it is limited to low differential pressures and relatively thick (viscous) product-otherwise the "slip" (loss of flow) would be substantial. As the rotor turns inside the stator, a cavity is formed between the two shapes and "progresses" (hence the name) axially from one end of the element to the other.

PC pumps, like other pump types, have limitations. Typically, they are not useable in high temperature applications because of the limitations of elastomeric stators. They require significant floor space and should not be run at speeds much higher than 300- to 400-rpm due to the unbalanced nature of the rotating element. However, when used correctly, the PC pump's benefits can be substantial. A PC pump will provide long-lasting service to any plant with proactive maintenance practices including vibration trending programs, root cause analysis and similar equipment-caring techniques.

Advantages• Handles thin and highly

viscous liquids • Low shear pumping action • Self priming capability (must

have Liquid in the pump chamber)

• Can handle abrasive liquids • Have one seal or stuffing

box • Non pulsing flow

Materials Of Construction / Configuration Options • Casings - Cast iron, steel, 316 stainless steel Alloy C. • Rotor – Steel, 316 stainless steel Alloy C. Hard chrome

coatings available • Stators – Neoprene, natural rubber, Perbunan, Viton,

PTFE, Hypalon and many others • Shaft Seal - packed glands and component mechanical

seals, industry-standard cartridge mechanical seals,

Disadvantages• Serious

damage if run dry

• Temperature limitations due to elastomeric stator

• Long floor space requirement

Multiple Screw Pumps



This section compares the twin and triple screw pumps. Unlike gear and lobe pumps, screw pumps are axial flow rather than radial flow machines; flow moves along the axis of rotation rather than perpendicular to it. This axial flow allows multiple screw pumps to operate at relatively high direct drive speeds while still maintaining low fluid inlet velocities and low NPSH requirements. Figure 1 illustrates the flow path within these pumps.

Each wrap of screw thread forms a cavity that moves axially from suction to discharge. The wrap, or cavity, acts as a pressure stage. Low pressure pumps have only one or two wraps (stages), while high pressure pumps may have 12 or more wraps. The staging effect allows each stage to handle a moderate pressure rise, resulting in low stress levels within the pump even at high pressure operation.

Triple screw pumps have one driven screw and two idler screws. There is contact between the driven and idler screws. In twin screw pumps, external timing gears and bearings keep the screws from contacting each other or their casing bores and do not rely on pumped fluid characteristics.

The majority of twin screw pumps are double suction designs, which effectively puts two single suction pumps in parallel in one casing. The double suction designs, both three-screw and two-screw, are inherently in hydraulic balance in the axial direction due to their symmetry.

Ph: 1300 789 466 Fax: 02 9678 9455 Email: [email protected] www.kelairpumps.com.au QLD Fax: 07 3805 7309 VIC Fax: 03 9569 7866 TAS Fax: 03 6331 9102 WA Fax: 08 9248 2255 PUMPS STEAM TURBINES BUILDING & FIRE WASTEWATER SERVICE

Kelair Pumps Australia Pty Ltd ABN 28 001 308 381 215 Walters Road Arndell Park NSW 2148


In a radial direction, twin screw pumps are not hydraulically balanced and require radial bearings at each end of each shaft.

Triple screw pumps are fitted with one seal whereas twin screw pumps require four seals; one ate ach end of the two shafts.

Both twin screw and three screw pumps share many applications including hydraulic services, machinery lubrication, compressor and expander gas sealing and some refinery (heavy fuel, asphalt, vacuum tower bottoms, bitumen) and chemical processing (synthetic fibers, explosives, polyol, isocyanate) applications.

Examples of critical applications include modified twin screw pumps used to handle multiphase flow, i.e., oil well head flows ranging from nearly 100 percent gas to 100 percent liquid including liquid slugs. Three screw pumps find use aboard combat ships for hydraulic services where extremely quiet operation is necessary to avoid acoustic detection. Both twin and three screw pumps are used in medium and heavy crude oil pipelines operating at efficiencies far above centrifugal pumps.


• Externals (head, casing, bracket) - Cast iron, ductile iron, steel, stainless steel,

• Internals (rotor, idler) - steel, stainless steel.

• Bushing - Carbon graphite, bronze, silicon carbide, tungsten carbide, ceramic, colmonoy, and other specials materials as needed.

• Shaft Seal - Primarily component mechanical seals, industry-standard cartridge mechanical seals

Peristaltic Pump Overview

Peristaltic pumps (also known as hose pumps) have been around for many years, with some designs dating back more than 75 years. The pumping action is provided by rotating shoes squeezing a hose and hence enclosing a volume of liquid which is progressed to discharge by the rotating shoes.

The shoe design hose pump uses two or more fixed shoes to compress the hose twice per revolution by grinding against the hose.

They are excellent devices for pumping slurries, due to their ability to handle very abrasive slurries. Hose pumps are also very good at dosing chemicals, since they are a positive placement device that can very accurately control the flow rate desired.

Hose pumps can run dry for long periods of time without damage to the pump. Typically the only wearing part is the rubber hose, which is also the only part in contact with the pumped medium.



NPSHA is not a concern for hose pumps because they create their own suction on the inlet side. Finally, hose pumps will never cavitate.

The biggest challenge is manufacturing the hose itself, which is the main element and only repair part. As such, the hose is the source of the greatest MTBF.

The slower a hose pump is run, the better, because it places fewer revolutions on the hose. One school of thought suggests that the abrasiveness of the slurry is what destroys the hose in a hose pump. This is not the reality, because the number one factor in determining hose life in a hose pump is how many compressions are placed on the rubber hose. The number two factor that contributes to hose wear is the amount of stress being placed on the hose during a compression and how much heat is generated from that compression force.

In other words, the best way to maximize hose life and eliminate pump downtime is to reduce the number of compressions on the hose and compress the hose in the less damaging manor.

There are basically two types of pumps that can be considered a hose pump, but there is a clear distinction between hose pumps and tube pumps. The difference between these two types of pumps are that tube pumps typically do not have a glycerin bath, pump at very low pressures and also are very small and low flow rate devices.

A hose pump is more of an industrial piece of equipment than a piece of lab equipment. Hose pumps typically range is size from 12 mm to 150 mm and typically have a maximum pressure capability of ~15 bars, depending on manufacturer. This discussion is focused on industrial hose pumps, rather than tube pumps.

There are many designs of hose pumps, but there are essentially only three means employed by all of these designs to compress the hose. The first is the shoe design, where two or more fixed shoes compress the hose twice per revolution by grinding against the hose. This type of design damages the hose the most because it generates a lot of heat and creates a lot of stress/damage to the hose on each revolution.

(delete)

Shoe designed pumps have a significant limitation regarding the speed at which the pump can be operated. Because of the high drag/friction across the rubber hose, the pump heats up significantly. Due to this heat and friction, these types of pumps cannot run at very high speeds.

For example, a 80 mm pump may be capable of running at only 40-rpm continuously, which, in turn, limits the amount of flow that can be produced continuously. Manufacturers of this type of pump tend to push the user to the next larger size pump so the rpm is kept lower. Though this strategy is correct, the user of a rolling design peristaltic pump can typically work with a unit that is one size smaller.

Also, a limiting factor on the shoe type of pump involves running a very low rpm. The high drag created from a very low rpm may frequently trip the variable frequency drive (VFD).



Advantages• No pumping moving parts • Seal/less configuration • Non-pulsating discharge • Good for high-viscosity liquids • Can handle abrasive • Self priming capability • Low shear pumping action • Easy to maintain

Disadvantages• Requires moderate speeds • Pulsing flow • Low hose life • Large space requirements • Application may be limited by hose

material availability

Materials Of Construction / Configuration Options As the liquid contacts only the hose and the connections, the limitation is the hose material. Many hose materials are available from various manufacturers. Connections are available in steel, stainless steel and higher alloys.

PumpClinic……… Issue 28

Pump Clinic 28 Page 1 of 1 29/10/08

CORROSION SOLUTIONS

(Reprinted with the kind permission of Sterling Fluid Systems Group)

STERLING FLUID SYSTEMS GROUP

CORROSION SOLUTIONS WORLDWIDE

1

INTRODUCTION

The annual cost of corrosion and of protection against corrosion inthe world is staggering. A plant may spend considerable amountsof money each year in painting steel to prevent rusting. Corrosionin radiators, exhaust systems and water heaters, increases thiscost further.

The world’s economy would be entirely different if it were not forcorrosion. Even though corrosion is here to stay, its cost can beconsiderably reduced in industry through proper selection ofmaterials, correct design of products, and preventativemaintenance. Corrosion contributes to the depletion of our naturalresources and the recent concern over this is becomingincreasingly influential in inducing people to be ”corrosion-cost“conscious.

In an effort to apply corrosion principles (from fluid flow tounderground soil and atmospheric) to contained chemical systems,and in particular to centrifugal pumps, the following schematicdiagrams have been used (Page 2). These indicate the differentforms of corrosion that may take place in a pump. The diagramsdepict an impeller in a casing. The seals, bearing brackets, etc. areto be imagined.

To view corrosion engineering in its proper perspective, it isnecessary to remember that the choice of material depends onseveral factors:

a. Cost

b. Corrosion resistance

c. Availability

d. Strength

e. Fabrication

f. Appearance

In dealing with pump systems, some other factorsneed to be considered:

a. Suction and operation conditions

b. Continuous/intermittent service

c. Are there several pumpsinvolved, in series, or parallel?

d. Type of seal

e. Flushing fluid

f. Temperature change

The above factors influence the corrosion rates ofany given material, and each is dealt with in detailwhen considering a pump’s hydraulics, but thechoice of a proper material depends on howaccurately these factors have been calculated.Additionally, any changes subsequently made inoperation or processing are critical and may makere-evaluation of the sizing and choice of materialnecessary.

2

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��

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EIGHT FORMS OF CORROSION AND TWO FORMS OF CRACKING PHENOMENAFIG. 1A MECHANICS OF CORROSION

Fig. 1B CRACKING PHENOMENA

General Corrosion Erosion

Localised Corrosion Cavitation

Pitting Crevice/Deposit Corrosion

Galvanic Intergranular (Corrosion)

Stress Corrosion Cracking Corrosion Fatigue

FLOW

BUBBLE

HEAT AFFECTEDZONE

HEAT AFFECTEDZONE

ALLOY R53NOBLE

ALUMINIUMBASE

DYNAMICSTRESS

WELD

WELD

CRACK

STATICSTRESS

3

GENERAL CORROSION

General corrosion leads to relatively uniform thinning. For round bars and wires,corrosion proceeds radially inward at an essentially uniform rate around the entirecircumference. Castings suffer corrosion starting at the wall exposed to the fluid (forexample, the impeller face of the casting) and proceeding gradually and uniformly to theouter wall. Methods of reducing or eliminating general corrosion are the use of coatings,the selection of a more corrosion-resistant material (a general rule is to select an alloywith a higher chrome and/or nickel content), the use of inhibitors, or cathodic protection.General corrosion proceeds by many different means:

A. The corrosion reaction product may be protective; it may form a passivating barrierthat stifles further corrosion. In this case, the material is not inert, but continues tocorrode at a low rate, and to continually repair the passive film. Most corrosion-resistantaustenitic materials, such as stainless steels, show this type of passivating behaviourwith the help of a surface film of oxides. This type of protection is very sensitiveto solids in the pumped fluids. These solids may continually scour away theprotective film which would otherwise form, thus leading to erosion corrosion.

B. The corrosion product may be soluble in the pumped fluid at a ratedetermined by the electrode potential of the metal. This is illustratedvery well when steel is used in oxygenated water.

C. A special case of an artificially controlled uniform dissolutionprocess may be attempted by controlling the pH or current density of agiven solution. This principle is utilised in chemical machining andelectropolishing of stainless steels to improve either the corrosionresistance or frictional characteristics.

In pumps, the recognition of general corrosion is compounded by velocity andpressure variations. The surface casing may show whorls and pockets wherevelocity variations have influenced the rate of corrosion. These variations may appear tobe caused by solids or erosive products in the fluid. However, close examination willalways reveal the fact that corrosion has left an etched appearance on the surface.

CASE HISTORYILLUSTRATINGGENERALATTACK

Alloy: Sterling R52

Environment:30% Hydrochloric Acid withminor impurities.Temperature 66oC.

Description: Notice theattack is all over the surfacewithout an exaggerated effecton the outer periphery. This pump was in service fortwo years.

Remedy: Review cost andevaluate possibility of usinghigher grade alloy, such aszirconium, which is resistantup to 37% HCI to 70 ºC.

CASE HISTORY ILLUSTRATING SEVEREGENERAL ATTACK

Alloy: 316 S.S.

Environment:70% Sulphuric Acid.Temperature 60 oC.

Description: A 316 S.S.impeller was substituted foran R-55 impeller.

Remedy: Do notinterchange castingswithout checkingsuitability ofapplication.

4

Erosion consists of two types of damage modes:

A. Mechanical-chemical or erosion-corrosion

In the erosion-corrosion mode, the flowingliquid may be free of abrasive particles.However, the velocity of the liquid may causeflow aberrations and turbulence due to surfacediscontinuities. The surface discontinuity maybe a weld bead, or flashing in the volute of acasing. This loosely adhered particle isremoved by the liquid velocity, making thedamage look erosive. There is a defined”breakaway velocity“ at which erosion-corrosion begins and is characteristic of agiven alloy/pumped-fluid system.

If this mode is identified and there is noparticle impingement, then reducing the flowrate will help reduce the erosion-corrosion. In

contrast, the removal of the discontinuity may produce the same result.

Impingement of abrasive particles carried by a fluid can affect the surface of the casing, impeller,etc. by causing mechanical damage. The particles are now capable of destroying the protectiveoxide film continually by fluid shear, thereby increasing the rate of damage. In this case, reducingthe flow rate will not help. It is the angle of incidence of the abrasive particle that is of primeimportance. Filtration of these particles wherever possible may be the best solution.

A good example of discontinuity, such as a scratch or a small pit, causing damage is often seen inmechanical seals. In this case, high pressure fluid in the constricted scratch zone causes thedevelopment of a channel. A number of closely knit channels causes the material to “wire-draw”.

This usually reveals itself in a characteristic ripple pattern. The impingement of hard particlescauses multiple cratering. Here the surface undergoes deformation and eventually extrusion. It isthe angle of impingement, velocity, hardness and angularity of the particles that affect this type oferosion. It is extremely sensitive to the flow paths and thus may appear to be localised.

EROSION - CORROSION

CASE HISTORYILLUSTRATING

EROSION

FOR (A) AND (B)

Alloy: 316 S.S.Environment: SodiumNitrate with other solids

SP. GR. 1.24.

Description: Very littledamage around shaft

opening. Outer peripheryshows damage increasing

due to velocity and pressure.By-pass throat area shows

severe damage. Erosionmarks indicate direction

of liquid flow.

Remedy: a) Check rotationon impeller and b)

Substitute with Sterling R48or other more abrasion-

resistant material. This willoften correct the situation.

FLOW

(A)

5

CAVITATION

B. Purely mechanical or particle erosion

In the case of mechanical erosion, the action is limited to the outer periphery of the casing. It ishere that the velocity and pressure of the liquid are the greatest. The central portion around theshaft opening is generally untouched.

Methods of preventing or reducing Erosion-Corrosion can be accomplished by use of one ormore of the following methods. Use materials with improved corrosion resistance to provide astronger protective oxide film. Improve design of system to reduce turbulence. alteration of theenvironment such as filtering to remove solids or reducing the temperature. Use coatings, suchas hard facing, if the coating has the required corrosion resistance, as well, as hardness.Cathodic Protection has been found to reduce Erosion-Corrosion in some applications.

This form of erosion is attributed to the following:

A . Formation of bubbles: At the eye of the impeller the pressure on the liquid is sufficiently

reduced to cause the liquid to vapourise or form bubbles.

B. Collapse of bubbles: As the liquid is now pumped to the outer periphery of the impeller, the

pressure is increased, causing the bubble to implode. Repetition of this process at high speed

causes the bubbles to form and collapse rapidly.

(B)

6

These rapidly imploding bubbles may produce shockwaves with pressure as high as 4400 bar.This is well beyond the yield strength of anumber of materials. Such forces causeplastic deformation in metals, which isindicated by the presence of slip lines on theimpeller and casing.

An imploding bubble causes the metal to beroughened. This roughened area in turn acts as anucleating site for a new bubble to form. Thecollapsing bubbles appear to cause closely-spacedpitted areas and considerable roughening of thesurface. Some measures which can be taken toalleviate this problem are:

A. Change of the design to minimise the hydro-dynamic pressure differences in the process fluid.

B. Use of a more corrosion-resistant material.


CAVITATION TYPEDAMAGE

Alloy: Sterling R55.

Environment: Water with50 ppm chloride at 44oC.

Description: Theimploding of bubbles

during re-absorption on thepressure side of the blade

causes damage in the formof closely spaced pits.

Remedy: a) Check theconditions under which the

pump is operating.b) Use a more cavitation-

resistant material.


CAVITATION TYPEDAMAGE


Environment: Water with50 ppm chloride at 44oC.

Description: Theimploding of bubbles

during re-absorption on thepressure side of the blade

causes damage in the formof closely spaced pits.

Remedy: a) Check theconditions under which the

pump is operating.b) Use a more cavitation-

resistant material.


CAVITATION TYPE - GAS

CONCENTRATIONCORROSION

Alloy: 316 S.S.

Environment: Unknown.

Description: A restrictedsuction condition caused aninsufficient amount of liquid

to fill the discharge throat,yet a large amount of gas

was present. This caused avacuum on the low velocity

side as shown.

Remedy: a) Review designof suction.

b) Determine if gas in liquidcan be reduced.

C. Smoothing of the finish on the impeller to reducenucleating sites for the bubbles.

D. Use of a rubber or plastic coating that inherentlypossesses a strong metal-coating interface.

E. Cathodic protection can reduce cavitation byforming bubbles on the metal surface, therebycushioning the shock waves produced by thecollapsing bubbles and thus preventing damage tothe metal surface.

Cavitation damage is located anywhere between theinlet eye of the impeller and the tip of the blades.The closely spaced pits are usually seen on thelagging side of the blades. In certain violentinstances, damage is noticed on the leading side ofthe blades. The extent and location of the damage isdependent on the fluid being handled, thetemperature, partial pressures and the degree of re-circulation flow inherent in the design.

7

PITTING

This form of attack is extremely localised. It usually results in a cavity that has

approximately the same dimensions in breadth and in depth. As the breadth

increases, the depth increases, causing a hole through the wall of the casing.

Pits have the following characteristics:

CASE HISTORYILLUSTRATING PITTINGCORROSION

Alloy: 304L S.S.

Environment:Hydrochloric and nitric acidmixtures. Temperature andconcentration unknown.

Description: Severaldamage marks (causedduring material handling)served as nucleating sitesfor an autocatalytic reactionto occur. This resultedin the pits shown.

Remedy: Use alloy withmolybdenum such as CF3M. Methods for combatingCrevice Corrosion generallyapply for pitting. Theaddition of Molybdenum of2% or greater in stainlesssteels contribute greatly inincreasing resistance topitting.

A. They are difficult to detect because they are often covered withcorrosion products.

B. Pits usually grow in the direction of gravity. This is substantiated by thefact that they require a dense concentrated solution for continuing activity.

C. Pitting usually requires an extended initiation period before visible pitsappear. This period ranges from months to years, depending on both thespecific metal and the corrosive liquid.

D. Pitting is autocatalytic. That is, the corrosion processes within a pitproduce conditions which are both stimulating and necessary for thecontinuing activity of the pit.

E. Pitting is usually associated with stagnant conditions. For example, atype 304 stainless steel pump would give good service handling seawater if the pump ran continuously, but would pit if shut down forextended periods of time.

F. Most pitting is associated with halide ions, such as chlorides, bromides,and hypochlorites. Fluorides and iodides have comparatively lesser pittingtendencies. Oxidizing metal ions, such as cupric, ferric, and mercuric incombination with chlorides are considered to be the most aggressive.Non-oxidizing metal ions such as sodium chlorides and calcium chloridesare much less aggressive. This type of corrosion differs from crevicecorrosion in that it creates its own crevice. Materials that are susceptibleto crevice corrosion do not necessarily become susceptible to pittingcorrosion, whereas the reverse may be considered to be true.

8

CREVICE/DEPOSIT CORROSION

This type of corrosion occurs in restricted areas, either metal to metal, (threaded drain plug),or metal to non-metal, (gasketed joints), where free access to the pumped fluid is restricted. Itis aided by the presence of deposits such as sand, dirt, and carbonaceous material that shieldand create a stagnant condition. In certain cases corrosion products will deposit and form acrevice.

As with pitting corrosion, an autocatalytic reaction fosters the growth of crevice corrosion.Thus, the initial driving force is often an oxygen or metal ion concentration cell, but continuedgrowth by accumulation of acidic hydrolyzed salts within the crevice. The external surfacesare protected cathodically.

This kind of attack occurs in many media, however, it is very common in chloride-containingenvironments. It is slow to start, but grows at an ever-increasing rate.There are a number ofactions that can be taken to prevent crevice corrosion:

A. Use of welded joints instead of threaded joints.B. Weld on both sides of a flange to pipe joint, thus avoiding penetration from either side.C. Ensuring that the pump is completely drained.D. Use of gaskets that are non-absorbent, such as teflon, wherever possible.E. Use of flushing in seal areas to avoid stagnant conditions in the bore of the stuffing box cover.


CREVICE CORROSION


Environment: Dilutesulphuric acid with smallamounts of hydrochloric

acid and sodium chloride.Temperature and

concentration unknown.

Description: Thegasketed shielded areahas limited diffusion of

oxidizing ions, thuscreating an imbalance andthe initiation of corrosion.

Note the pitting that isusually associated with

this type of corrosion.

Remedy: a) Review gasketmaterial.

b) Consider welding thejoint between impeller

and shaft

9

GALVANIC

TABLE 1 - GALVANIC SERIESIN PURE SEA WATER

Corroded End (anodic, or least noble)Magnesium

Magnesium Alloys↓

Galvanised Steel orGalvanised Wrought Iron

↓Aluminium

(5052, 3004, 3003, 1100, 6053 in this order)↓

Cadmium↓

Aluminium(2117, 2017, 2024 in this order)

↓Mild Steel

Wrought IronCast Iron

↓NI-RESIST

↓Type 410, Stainless Steel (active)

↓50-50 Lead Tin Solder

↓Type 304, Stainless Steel (active)Type 316, Stainless Steel (active)

↓LeadTin↓

Muntz MetalManganese Bronze

Naval Brass↓

Nickel 200 (active)INCONEL alloy 600 (active)

↓Yellow Brass

Admiralty BrassAluminium Bronze

Red BrassCopper

Silicon Bronze70-30 Copper Nickel

Comp. G-BronzeComp. M-Bronze

↓Nickel 200 (passive)

INCONEL alloy 600 (passive)↓

MONEL alloy 400↓

Type 304, Stainless Steel (passive)Type 316, Stainless Steel (passive)

INCOLOY alloy 825↓

INCONEL alloy 625HASTELLOY alloy C

TitaniumProtected End (cathodic, or most noble)

A potential difference exists between two dissimilar metals whenthey are immersed in a corrosive and conductive solution. If thesemetals are now connected electrically conductively on the outside,an electron flow is produced. One of the two metals will corrodefaster than the other. The metal which is corroding at a faster ratebecomes anodic, while the other metal is cathodic.

The most commonly used series, based on electrical potentialmeasurements and galvanic corrosion tests in unpolluted seawater is shown alongside. (See Table 1).

A similar series is needed for all of the various situations. Thenumber of tests required would be almost infinite. Thus the seriesshould be used only for predicting galvanic relationships. Theseparation between the two metals or alloys in the series is anindication of the probable magnitude of corrosive effects. Effectssuch as polarisation (potential shifts as the alloys tend to approacheach other), area, distance and geometry play a definite role ingalvanic corrosion.

There are several ways in which to combat galvanic corrosion:

A. Material selection is extremely important. Substitution ofimpellers of different alloys in an existing system must be donecarefully. Care should be taken to avoid wide separation in therelevant galvanic series.

B. The pumped fluid may be controlled by the use of a corrosion inhibitor.

C. Barrier coatings and electrical isolation by means of insulatorsto break the electrical continuity are sometimes employed.

D. Cathodic protection by way of using sacrificial metals may be introduced.

E. Design changes involving the avoidance of the unfavourable arearatios, using bolts and other fasteners of a more noble metal thanthe material to be fastened, avoiding dissimilar metal crevices, (as at threaded connections), and using replaceable sections withlarge corrosion allowances of the more active member.

ALUMINIUMBASE

STERLING R53NOBLE

10

INTERGRANULAR CORROSION

In a survey conducted by the Material Technology Institute of theChemical Process Industries, it was shown that sensitisation andresulting intergranular corrosion were the cause of over half of thereported incidents of unsatisfactory performance.

Metals and alloys consist of individually oriented crystals whichform from the molten state (castings or weld metal). These crystalsdevelop into specific atomic arrangements known as crystalstructures (e.g. cubic, hexagonal etc.). These crystal structures canbe manipulated by varying chemistry and heat treatment. Each typeof atomic arrangement has specific physical and mechanicalproperties of its own. Two important arrangements in stainlesssteels are called Ferrite and Austenite. During the solidificationprocess, the development of facets indicative of their crystalstructure is prevented when the growing crystals impinge on eachother. Where the crystals come in contact with each other, theirfacets form a boundary that takes the form of a lattice. Theseboundaries have a much greater degree of structural imperfectionthan within the grains. The resulting energy states at theboundaries can promote the concentration of alloying elements,and of metallic and non-metallic impurities, and of greatestimportance - precipitates.

Sensitisation may be referred to as carbide precipitation in the grainboundaries. The structure of low carbon austenitic stainless steelsconsists of three crystallographic phases: Ferrite, austenite andcarbide under equilibrium condition. Rapid cooling of these steelswill ensure the retention of austenite (a high temperature phase),but if they are heated to around 800oC for any appreciable length oftime, the carbide will precipitate in the grain boundaries. The effectof sensitisation on the chromium and carbon concentration isshown in Figure 3. The figure depicts a transient state only. Figure 2shows the variation in carbon content in passing from one grain,through the grain boundary, to another grain. The chromiumcontent varies in the manner shown in the lower portion of Figure3. There is a narrow region at the grain boundary which containsless than 12% chromium. This is below the 12% chromiumminimum required for corrosion-resistance. If there was no carbonpresent in the alloy, this condition would not be possible.

The sensitisation of an austenitic alloy permits corrosive attack tostart at the grain boundary, (lowest energy level), where there is adeficiency of free chromium. Since the grains (high energy level),are more resistant than the boundaries (low energy level),corrosion follows the boundaries, which is typical of intergranularcorrosion.

In translating this phenomenon to the macro-scale, the attack is firstrecognised as ditching along the grain boundaries, on the surfaceof the casting. As the attack progresses, it permeates the completecasting wall and results in leakage and possible grain dropping. Itshould be noted that cracking, however, does occur in austeniticalloys, and is mainly due to stress corrosion.

EFFECT OF SENSITISATION ON THE CARBON & CHROMIUM CONCENTRATION

GROWING CRYSTAL FACETS FORMING A GRAIN BOUNDARY

AVERAGE CONTENT 0.08%

SATURATION VALUE 0.02%

LEVEL FORRESISTANCE12%

GRAIN BOUNDARY

GRAINBOUNDARY

CA

RB

IDE

PAR

TIC

LEC

AR

BID

E PA

RT

ICLE

18%

% CARBON

% CHROMIUM

HEAT AFFECTEDZONE

HEAT AFFECTEDZONE

WELD

WELD

Fig 3 - EFFECT OF SENSITISATION ON CARBON

AND CHROMIUM CONCENTRATIONS

Fig 2 - LATTICE MISMATCH

11

In practice at the foundry level, it must be mentioned that sensitisation will occur during slowcooling in the mould. Immediate cooling of castings from the mould has been tried. However,the hostility of this production step has limited its applicability. Quench cracks have beennoted in stress concentration areas also. Sensitisation is prevented by providing a solutionanneal above the sensitising range temperature.

CASE HISTORYILLUSTRATINGINTERGRANULARCORROSION (a)


Environment:25% Hydrochloric acid with100 ppm Chlorine. Temperature: Ambient.a) The impeller on the leftwas solution annealed andthen put in service.b) The impeller on the rightis in the sensitisedcondition. This impeller hasundergone severecorrosion with graindropping occurring at thetip of the blades.

CASE HISTORYILLUSTRATINGINTERGRANULARCORROSION (b)


Environment:25% Hydrochloric acid with100 ppm Chlorine. Temperature: Ambient.Close-up of the impeller onthe right in (a) above.

12

STRESS CORROSION CRACKING

CASE HISTORYILLUSTRATING STRESS

CORROSIONCRACKING

Alloy: 304L S.S.

Environment:50% Caustic (Sodium

Hydroxide) with trace

amounts of Sodium

Chlorides.

Temperature 66OC.

Description: A drastic

brittle fracture in a ductile

material in a period

of four months.

Remedy: Use another

alloy type.

Two characteristics are necessary for environment cracking to occur: a tensile stress and acorrosion reaction. There are several forms of environmental cracking such as stress corrosioncracking, hydrogen induced cracking, liquid metal cracking, and corrosion fatigue. Generally,cracks produced by this method are unexpected and sometimes are dangerous. They are oftenwrongly interpreted. For example, intergranular corrosion does not require a tensile stress,however, the morphology of the cracks may be very similar to stress corrosion cracking.Welding sometimes produces hot-short cracks which may be identified as stress corrosioncracking.

The stresses that exist in a given situation are usually very complex. The surface net stress incontact with the pumped fluid will be the controlling parameter. The cracks produced areperpendicular to the stress vector. These cracks may be single (as in corrosion fatigue) ormultiple (as in stress corrosion cracking). They may be intergranular or transgranular.

Several of the parameters and the controlling methods used are discussed below:

A. Stress. For a given alloy/fluid system, a threshold stress for cracking exists. In such instances,stresses below the threshold will not cause cracking, but as the stress is increased above thethreshold, cracking is immediately evident. Lowering the residual and thermal stresses by heat-treatment and shot-peening is carried out to decrease the stress levels below the threshold. Thelatter generates compressive stresses in the material which often offset the tensile stressnecessary for cracking to occur. Compatibility of various materials in contact with respect topolarisation of potentials along with geometrics that increase salt ion concentrations (likecrevices) should be considered in detail.

B. Metallurgical: The list of specific environments that aid in stress corrosion cracking is differentfor each major alloy classification. For example, caustics being handled by a carbon steel,chlorides pumped by stainless steel pumps, and copper alloys in an ammonia environment. Themost common method is to utilise another alloy that is not susceptible to this attack.

STATICSTRESS

13

CASE HISTORY

ILLUSTRATING

FATIGUE CORROSION

Alloy: 316 S.S.

Description: A crack

propagated at an inclusion

present in the material after

several months of operation.

CORROSION FATIGUE

This attack results from the cyclic tensile and

corrosive fluid in contact. There is some mystery

in this type of attack due to the fact that failure in

this mode may occur in the absence of corrosive

action. Most of the discussion in stress corrosion

cracking is applicable to corrosion fatigue.

Pump shafts have often failed due to mechanical

fatigue with no contribution by the corrosive. On

the other hand, stress corrosion cracking under

static tensile stress has also

been known to occur. The

phenomenon then covers a

broad spectrum and is

difficult to define clearly.

The importance of microstructure may be illustrated by placing:

1. a sensitised 316 material in a nitric acid solution. Stress corrosion cracking is noticed.2. a solution-annealed 316 material in the same solution. This will not induce cracks.

C. The corrosive liquid in certain cases can be made less effective in causing stress corrosion cracking by the use of an inhibitor such as chromates in a caustic solution. Elimination of the critical chemicals from the liquid is probably the most desirable. A review of the entire system is usually necessary if this attack has been identified.

D. An increase in temperature generally has a detrimental effect, that is, it tends to induce stress corrosion cracking. If, however, the temperature is high enough to remove the criticalchemicals, then the tendency reduces.

E. Coatings and electro-chemical techniques are also used. The coatings normally act as a barrier between the metal and pumped fluid. Electro-chemical techniques are generally used to polarise an alloy to an oxidizing potential out of the range that will cause stress corrosion cracking.

DYNAMICSTRESS

14

CORROSION BY ACIDS

The acids most generally used by industry are

sulphuric, nitric, phosphoric and hydrochloric

acids; and these cause some of the most severe

corrosion problems. The widespread use of these

acids places them in an important position with

regard to costs and destruction by corrosion. In

some cases, corrosion increases with the

concentration of the acid, in others it decreases.

Oxidizing and reducing mixtures of acids and

salts also causes different reactions to different

materials. Velocity and aeration are factors

that must be taken into consideration. Finally,

impurities in the system can cause

severe problems.

SULPHURIC ACIDSelection of a metal for this service depends

primarily on the reducing or oxidizing nature

of the solutions. Below 85% at room

temperature and about 65% up to 66oC, the

acid is reducing and is better handled by

materials resistant to reducing conditions. In

higher concentrations, the acid is oxidizing

and materials resistant to oxidizing media

are essential.

Cast Iron:

Cast irons show good resistance in very strong

sulphuric acids. In a number of instances, it is

more economical to use cast irons, although

the corrosion rates are higher. The resistance

in these alloys is attributed to the graphite

network interfering with the reaction between

the acid and the metallic matrix. In oleum,

however, the acid is known to penetrate the

metal along the graphite flakes, and a little

corrosion in these confined areas builds up

enough pressure to split the iron. This

wedging action is confined to cast irons and is

not apparent in ductile iron, which may be

satisfactory for oleum service.

Types 304 and 316 Stainless Steels:

These stainless alloys are occasionally utilised

for cold, very dilute sulphuric acid and under

conditions that are not strongly reducing

in nature.

Sterling Alloy K26:

This is a very widely used alloy for

applications involving sulphuric acids. It

provides resistance over the entire range of

concentration, including oleum. The

isocorrosion chart reveals the dip in the curve

15

at around 60-80% concentration range at

approximately 66oC. This will maintain the

corrosion rate within 20 mm per year. If the

pump, for example, is to be used

intermittently, then the temperature limitations

may be increased to 82oC. Ferric sulphate and

copper sulphate in the acid act as inhibitors

and decrease attack. Ferric chloride and cupric

chloride in appreciable concentrations are

known to cause pitting.

Sterling Alloy R55:

R55 is a nickel-chromium-molybdenum-copper

alloy that shows outstanding resistance to

sulphuric acid and many other media. It will

withstand the corrosion of both oxidizing and

reducing agents to moderately high

temperatures, (80oC). It is not recommended

for halogen acids or halogen salt solutions in

contact with the material, but provides

resistance over a wide range of oxidizing and

reducing conditions. The 4% copper in the

alloy is kept in solid solution which is essential

for sulphuric acid service. The R55 alloy has

numerous advantages over K26 and is the

most widely used Sterling alloy for sulphuric

acid and most sulphur compounds, such as

sulphur dioxide and hydrogen sulphide gases.

The copper in these alloys does not discolour

the product.

Sterling Alloys R52 and R53:

Alloy R53 is a nickel-chromium-molybdenum

alloy that shows a great deal of thermal

stability at high temperatures. It is useful over

the entire concentration range and oxidizing

conditions. The chromium content in the alloy

provides excellent resistance to oxidizing

conditions. This alloy is very suitable for

chlorides, up to 220 ppm, at a maximum

temperature of 70oC and over the entire

concentration range.

Alloy R52 on the other hand, is a nickel-

molybdenum alloy. This alloy is also known to

possess good corrosion resistance in the

intermediate and strong concentration range

of sulphuric acid. It is better suited to reducing

conditions, and is particularly susceptible to

oxidizing contaminants such as nitric acid,

chlorine, cupric and ferric chlorides, ferric

sulphates, and even aeration.

NITRIC ACIDOne of the most important ingredients for

resistance to nitric acid is chromium. As the

chromium content increases, the corrosion

rate decreases. The minimum amount of

chromium generally accepted is 18%. This

makes the austenitic stainless steels very well

suited for practically all concentrations and

temperatures. The addition of molybdenum to

stainless steels, as in type 316, as opposed to

304, does not improve corrosion resistance to

nitric acid.


Type 304 stainless steel exhibits excellent

resistance to nitric acid at room temperatures

up to 30oC, and also to boiling acids up to

50% strength. The corrosion resistance

decreases as the concentration and

temperature are increased beyond 50% and

30oC. Type 304 does, however, show excellent

resistance to red and white fuming nitric acids

at room temperature.

Because of the susceptibility of sensitised Type

304 (when exposed in the 430oC to 870oC

range) to intergranular attack in nitric acid,

boiling 65% nitric acid (Huey test) is often

used to detect the existence of this condition

prior to fabrication. This is only an indicative

test and is not a prediction of definite behaviour.

16

Nitric acid, when mixed with sulphuric,

phosphoric, or acetic acids, shows reduced

corrosivity to stainless steels and K26 alloys.

On the other hand, mixtures of nitric acid with

hydrochloric or hydrofluoric acid are corrosive

to stainless steels and their rates depend on

concentration and temperatures.

Sterling Alloys R52 and R53:

These alloys are not suited for nitric acid

services as they are readily corroded.

Sterling Alloy R55:

This alloy shows good resistance, but does

not increase resistance sufficiently to justify

the additional cost.

Sterling Alloy R48:

This shows excellent resistance and is

probably the only age-hardenable stainless-

type alloy that shows good resistance, even in

the hardened condition.

Titanium:

This alloy has outstanding resistance at all

concentrations and at temperatures well above

the atmospheric boiling points. It shows less

than 5 mpy in 65% nitric acid at 180oC. It is an

expensive material, but in some cases is the

only material that will do the job. The

presence of oxidizing ions in nitric acid tends

to decrease the corrosion resistance of

titanium - maybe its only drawback.

PHOSPHORIC ACIDPhosphoric acid obtained by the wet process,

is used in the production of fertilisers. The acid

obtained by the electric furnace process is

purer in form and is used in the manufacture

of soap, detergent, food, plasticisers and

insecticides.

Because of the impurities, such as sulphates,

fluorides, and fluosilicates, present in acid

made by the wet process, and the nearly pure

acid from the electric furnace process, the

corrosion behaviour and alloy selection are

based on the manufacturing process. Other

variables include the concentration,

temperature, aeration, etc.

In a number of studies, it has been found that

the cupric and ferric ions in solution inhibit the

corrosion of stainless steels in phosphoric

acid. The cupric ion may be provided by the

initial corrosion of an alloy such as R48 or K26.

On the other hand, the addition of chloride or

fluoride ions to the phosphoric and

hydrochloric or hydrofluoric acid, increases

the corrosion rates by breaking down the

passive film.

There is some debate about stainless steels

that have been affected by sensitising

processes:- corroding intergranularly. Until

there is more evidence, it would be prudent to

use extra low carbon or stabilised alloys for

severe services.

Sterling Alloy K26:

K26 is widely used in phosphoric acid service.

Great care is taken to make sure that the

castings are properly solution-annealed.

CORROSION BY ACIDS :- Continued

17

Sterling Alloy R55:

This is usefully resistant to all concentrations

of phosphoric at temperatures up to 90oC.

Sterling Alloy R52:

This is excellent in hot concentrated pure

phosphoric acid. However, copper ions (an

impurity) behave somewhat differently in

solution. Copper ions at first decrease the

corrosion rate, but beyond about 10 ppm, they

tend to increase the corrosion rate. If this alloy

must be used, the copper content must be

controlled to extremely low values.

Sterling Alloy R48:

This alloy works very well at the same

temperatures and concentrations as K26, is

less expensive and has the added advantage

of handling abrasives better than K26.

Wrought materials used in conjunction with

R48 castings, such as shafts and piping, are

now available.

HYDROCHLORIC ACIDThis is the most difficult of acids to handle

from a standpoint of corrosion. Hydrochloric is

corrosive to most common metals and alloys.

Oxidizing agents and minor impurities such as

ferric chloride (or cupric chloride) and nitric

acid present a very rugged corrosive

condition.

Great care and good judgement is required to

obtain a balance between service life and cost

of the equipment.

Sterling Alloy R53:

This alloy shows good resistance to all

concentrations of hydrochloric acid at room

temperature and has been used successfully

up to 50oC. Due to its high chromium content,

it provides better resistance to oxidizing

environments. It must be kept in mind,

however, that dissolved oxygen is not strong

enough to passivate the material.

Sterling Alloy R52:

R52 is widely used to handle hydrochloric acid

at all concentrations and temperatures up to

the boiling point. Due to the absence of

chromium in this alloy, its resistance to

aeration and oxidizing impurities such as nitric

acid or ferric chloride (when present even in

small quantities) is often destructive.

Types 316SS and Sterling Alloy K26:

The austenitic stainless steels, including K26

are to be used only at very low concentrations

at room temperature. Increasing the

temperature decreases the critical

concentrations at which the stainless steels

start to corrode. Rapid corrosion occurs at pH

4 or 5, or below. Pickling solutions which are

sometimes handled by these materials require

inhibitors if the pump is to be handling the

liquid on a continuous basis.

Nickel and Nickel Irons:

Aeration affects these alloys to a great extent.

They are generally not considered to be

suitable for hydrochloric acid service because

they are susceptible to influences other than

the acid itself and must be used with caution,

only when specific conditions are

definitely known.

Titanium:

This alloy is good up to 10% at room

temperature. The presence of ferric and cupric

chlorides actually decreases the corrosion rate

of titanium.

18

ACETIC ACIDAcetic acid is an intermediate chemical used in

the production of cellulose acetate for paints

and acetic anhydride for artificial fibres.

Various processes are used to produce the

acids, which include:

A. Acetaldehyde Process: The most widely

used process, it has specific problems with

regard to material selection due to catalyst

carry-overs, formation of peroxides and acetic

anhydride.

B. Butane Oxidation: In selection of materials

of construction you must consider formation

of peroxides, formic acid and other solvents

which are typical of this process.

The effect of contaminants is twofold:

1. Contaminants such as sulphur dioxide and

sulphur trioxide increase the corrosion rate.

Formic acid also increases the corrosion rate.

This increased corrosivity can normally be

tolerated by type 316 stainless steel.

2. The presence of aldehydes, ketones and

esters in the process stream have been known

to greatly reduce the corrosion rate.

Acetic anhydride usually will increase the

corrosive attack, especially when the acetic

acid is at 100% concentration. In this instance,

K26 is well-utilised.

Chlorides in the stream have been known to

cause stress corrosion cracking of the

austenitic stainless steels: R53 is usually

recommended if this is the case.

If there is a possibility of temperature increase,

the low carbon grades such as 316L and 304L

should be evaluated to prevent excessive

corrosion and contamination of the acid. The

iron contamination is greatly increased as the

temperature and time of exposure is

increased, especially in the case of 304.

Ferroxyl testing of pumps servicing a

meticulous grade of acetic acid is often done

before shipment. Practice “A” of ASTM A262

(Oxalic Acid Etch), is often recommended as a

qualification test to determine the sensitivity of

the alloy to attack by an acetic acid

environment.

R52 and R53 both provide excellent resistance

to acetic acid at all concentrations and

temperatures. These alloys are more

expensive than type 316 stainless steel and

K26, thus a service life-cost justification should

be done before Sterling types R52 and R53

are utilised.

CORROSION BY ACIDS :- Continued

19

CORROSION BY ALKALIS

Of all the available alkaline materials, caustic

soda (sodium hydroxide) is the most widely

used. It is produced along with chlorine by the

electrolysis of sodium chloride. The type of

electrolytic cell used for production

determines the purity that is obtainable.

Mercuric cells produce 50% grade caustic,

whereas the diaphragm cells produce 9% to

15% grade caustic, which is further purified

before sale.

The major users of caustic soda are the

chemical, pulp and paper, and aluminium

industries.

Iron and steel are widely used at low

temperatures (if iron contamination is not

detrimental), whereas nickel and nickel alloys are

used at higher temperatures.

In concentrations above 75%, and including

molten caustic soda, cast nickel does an

excellent job. When temperatures above

320oC are to be considered, the cast nickel

pump castings should be solution-annealed to

minimise the possibility of graphite

precipitation at grain boundaries and a

resultant loss in ductility.

Velocity and aeration have little effect, except

at high temperatures such as above 540oC.

The thermal decomposition at 260oC of

impurities such as chlorates (present in caustic

soda produced by the diaphragm cell method)

increases the corrosion rate of cast nickel. In

such instances, caustic soda produced by

other methods should be utilised, or reducing

agents such as sucrose or dextrim may be

added to minimise corrosion and product

contamination.

Oxidizable sulphur compounds also tend to

increase the corrosion rate of cast nickel at

elevated temperatures.


The cast versions of austenitic stainless steels

such as types 304 and 316 are used up to 10%

concentration and up to the boiling point of

caustic soda. However, at concentrations

above 10% the critical temperature decreases.

Chlorides in the process stream have been

known to contribute to stress corrosion

cracking of these alloys and consideration

must be given to the stress and temperature

limitations of these alloys.

Sterling Alloy K26:

Pumps made of this material have been used

for handling caustic soda up to 70% and

120oC. Galvanic effects must be considered if

this alloy is to be used with nickel and nickel-

based alloys.

Sterling Alloy R55:

This alloy is used in similar situations to K26. It

provides better stress corrosion cracking

resistance than K26.

Sterling Alloy R52 and R53:

The data available for these alloys is not

sufficient to make any indicative statements.

They have been used up to 50% at the

boiling point.

Ni-Resist Type 2 Cast Iron:

Ni-Resist Type 2 cast iron and spheroidal nickel

iron are both used where minimum

contamination of the product by copper is

desired. These alloys may be used up to 70%

caustic soda concentrations. Stress relief of

these alloys may help minimise stress corrosion

cracking of these alloys.

20

CORROSION BY LIQUID METALS

Heat-transfer characteristics of low boiling

point metals make them particularly attractive

for use in the power plant industries. The

efficiency of a power plant is increased by

operating at higher temperatures. Water and

steam require high pressure equipment which,

besides being hazardous, is also expensive.

The liquid metals and fused salts employed

are usually high thermal conductors. Due to

their low melting points, they save a great deal

in initial heat-up during the start-up of a power

plant. In addition, they require lower pumping

power due to their lower density. For example,

mercury requires very high pumping power,

whereas sodium requires low pumping power.

Liquid metals cause different types of

corrosive attack. The most salient feature is

the lack of electrochemical reactions. In the

simplest type of attack, the solid metal

dissolves in the liquid metal, resulting in

uniform thinning or preferential leaching of a

selective constituent from the solid metal. This

dissolution may result in the formation of

brittle alloy phases.

This kind of uniform thinning also occurs when

solid metal may dissolve at a hot zone of a

pump and precipitate on the walls of a cool

zone, where its solubility is less.

Contact of dissimilar metals with the same

liquid metal can cause transfer of one solid

metal through the liquid metal to the other

solid metal. This causes rapid dissolution

without saturation, leading to destruction.

Finally, impurities such as dissolved gases can

change the solubility limits, the wetting

tendencies, and the activity of the solid

metal ions.

The surface area to volume ratio is of utmost

importance. The greater the ratio, the lower is

the corrosion rate. This is because the greater

the liquid volume, the greater is the amount of

solid metal that can be held in solution.


These alloys may be used to pump sodium

and sodium-potassium mixtures. They have a

temperature limitation of 540oC. If used

intermittently, care should be taken to prevent

carburisation by carbonaceous material. Both

of these alloys can handle lithium, thallium,

mercury, bismuth and bismuth-lead alloys up

to various temperatures.

Grey cast iron: is also good for some of

these liquid metals, such as cadmium and

Bi-Pb-Sn alloys.

Cast nickel: possesses the greatest resistance

to stress cracking in lead, bismuth, tin, and

their alloys. They do not undergo as many

rupture failures as do the nickel chromium

steels.

21

Neither Sterling Fluid Systems, nor any of its officers, directors or employees accept any responsibility for the use of the methods and materials discussed herein.

The information is advisory only and the use of the materials and methods is solely at the risk of the user. Reproduction of the contents in whole or part or transfer

into electronic or photographic storage without permission of copyright owner is expressly forbidden.

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PumpClinic…Issue 29

Pump Clinic 29 Minimum Flow Due To Thermal Considerations Page 1 of 3 25/11/08

MINIMUM FLOW DUE TO THERMAL CONSIDERATIONS

The factors which determine minimum allowable rate of flow include the following:

• Temperature rise of the liquid as it passes through the pump

• Radial hydraulic thrust on impellers -- This is most serious with single volute pumps and, even at flow rates as high as 50% of BEP could cause reduced bearing life, excessive shaft deflection, seal failures, impeller rubbing and shaft breakage. This has been covered in Pump Clinic 12.

• Flow re-circulation in the pump impeller -- This includes suction and discharge recirculation when operating at flows other than the best efficiency point. This has been covered in Pump Clinic 3.

• Total head characteristic curve -- Some pump curves droop toward shut off, and some curves show a dip in the curve. Operation in such regions should be avoided.

This article considers minimum allowable flows based on temperature rise considerations only. To avoid thermal problems during low flow operation and to prevent a potential hazardous or mechanically damaging temperature rise within the pump, the temperature rise at shut off (i.e. fully closed discharge) and the minimum flow required for thermal protection should be calculated and the required flow be bypassed to dissipate heat generated due to pump inefficiency.

In the majority of cases, considerations other than thermal issues will dictate minimum allowable flow. Thermal considerations are important where liquids are at, or close to, the boiling point e.g. boiler feed pumps.

TEMPERATURE RISE AT SHUT OFF

The rate of temperature rise in the pump at shut off can be calculated by: TR = P x 14.4 Q x SH x SG where:

TR= temperature rise per minute, in degrees Centigrade P = power at shut off in kilowatts Q = volume of liquid in the pump in litres SH= specific heat of the liquid in calories/gm. C

SG= specific gravity of the liquid This calculation disregards any allowance for heat dissipated by radiation from pump casings. MAXIMUM ALLOWABLE TEMPERATURE RISE The maximum allowable temperature rise can be determined by T2 – T1 where:

T2 = saturation temperature corresponding to the absolute pressure of the pumped liquid at the pump suction flange

T1 = temperature of the liquid at the pump suction flange



TEMPERATURE RISE AT VARIOUS FLOWRATES

The temperature at the discharge will exceed the suction temperature and this can be calculated for any given flow. This rise is determined by the following formula:

TR = H

432.4 x SH where:

TR = temperature rise, in degrees Centigrade H = total head in metres SH= specific heat in calories/gm. C E = pump efficiency in % at the flow involved.

Values of total head at various flows can be read from the pump performance curve and the temperature rise at the various flows can be calculated. The figure below gives a graphical representation of this formula and allows determination of the minimum allowable operating flow once the maximum allowable temperature rise has been selected.

As shown, the temperature rise increases very rapidly with a reduction in flow. This is caused by the fact that the losses at low deliveries are greater when the flow of liquid that must absorb the heat developed in the pump is low.



If the pump is fitted with a balancing device for axial thrust e.g. balance drum or balance disc, a certain portion of the suction capacity known as leakoff is returned either to the pump suction or to the suction supply vessel. In this case, the discharge capacity does not represent the true flow through the pump. The formula for the temperature rise can still be used, provided a correction is made to take care of the increase in pump flow representing the balancing device leakoff. Balance device leak off information is provided by pump manufacturers.

The formula is modified to:

TR = H X Qd

432.4 x SH Qd + Qb where:

TR = temperature rise, in degrees Centigrade H = total head in metres SH= specific heat in calories/gm. C E = pump efficiency in % at the flow involved.

Qd = flow through pump discharge in litres/sec

Qb = leak off from balance device in litres/sec


Pump Clinic 30 Designing a Trouble-free Installation - Diaphragm Metering Pumps Page 1 of 5 28/01/09

DESIGNING A TROUBLE-FREE INSTALLATION DIAPHRAGM METERING PUMPS


PC31 Inducers for Centrifugal Pumps (Reprinted with kind permission of Lawrence Pumps Inc USA) Page 1 of 3 04/02/09

INDUCERS FOR CENTRIFUGAL PUMPS

(Reprinted with kind permission of Lawrence Pumps Inc USA) Providing sufficient net positive suction head (NPSH) to a centrifugal pump may be costly if it involves increasing the height of a vessel or maintaining higher than normal inventory levels.

Inducers are a low cost alternative that reduce the NPSH required by pumps, but their application is not without peril. In this issue we'll discuss inducers as well as the advantages and disadvantages of using them.

If you rapidly sweep your open palm through the air you'll notice a higher pressure on the side of your hand that is pushing against the air than you do on your hand's trailing side. A spinning impeller blade undergoes the same effect as it sweeps through liquid. The back side of the blade (the part that you can't see when looking into an impeller eye) pushes against the liquid, trying to accelerate it within the impeller passage, as the front side of the blade experiences a localized reduction in pressure. This localized low pressure area is where cavitation develops.

Vapor bubbles form in a pump inlet whenever the local absolute pressure of the liquid falls below its vapor pressure. The bubbles collapse rapidly and violently, resulting in noise, vibration, erosion of material from the impeller surface, and most importantly, reduced pump service life due to resultant mechanical problems. This rapid formation and collapse of vapor bubbles is cavitation. The severity of the effects of cavitation varies as a function of a machine's horsepower.

An Inducer is an axial flow impeller with blades that wrap in a helix around a central hub. An Inducer serves as a small booster pump for the main impeller. Usually inducers have between 2 and 4 vanes, although there may be more. The inducer imparts sufficient head to the liquid so that the NPSH requirement of the adjacent main impeller is satisfied. Although the inducer usually has a lower NPSH requirement than the main impeller, it can, and often does, cavitate during normal operation. The key is that there is so little horsepower involved with an inducer that there is virtually no noise, vibration, or resulting mechanical problems. Meanwhile, the higher horsepower main impeller sees sufficient head to operate without cavitation.



An inducer invariably has a higher suction specific speed (S) than the adjacent impeller. S is a dimensionless term that describes the inlet characteristics of a pump. For a constant RPM and flow, a lower NPSH requirement means a higher suction specific speed. Inducers commonly have suction specific speeds of between 15,000 and 25,000. A pump equipped with an inducer may operate at 1/2 to 1/3 the NPSHR levels of a non-inducer version of the same pump.

So what's the downside? There are a few that the user should be aware of.

1. The pump should be mechanically compatible with the addition of an inducer

Inducers add mass, cantilevered away from the bearings. This will increase shaft deflection and reduce the 1st critical speed of the unit. This is not a problem if the pump is designed for the added mass, but it may be a problem otherwise.

2. The NPSH requirements of the inducer need to be compatible with the entire operating range of the pump

Many inducers have a steeply rising NPSHR characteristic on either side of the design flow rate. Often the NPSHR will exceed that of a non-inducer pump when operating off design. Under these conditions, it is possible to make problems worse by applying an inducer.

3. High suction Specific speed (S) limits the allowable operating range of the pump.

High suction specific speed pumps become unstable when operated off design. Below is a general chart showing operating range vs. S. The width of the operating range also varies with the horsepower. Lower power pumps generally have a broader operating range; higher power pumps have a narrower operating range. Specific pump geometry will also affect the operating range. But it's good to remember that while a 10 hp pump can use up to 10 hp to create damage, an 800 hp pump can use 800 hp. It's unreasonable to treat them the same.



With a compatible pump design and operating conditions, an inducer can effectively reduce the initial capital cost related to system construction and can increase equipment reliability through improved inlet conditions.


PC32 Pumping Liquids with Entrained Gas Page 1 of 2 March 25, 2009

Pumping Liquids with Entrained Gas

Pump applications may involve handling liquid and gas mixtures either as part of the process or as a result of unwanted application conditions or upsets. Although symptoms may be similar to cavitation, the process of is a little different.

Cavitation is the formation and subsequent collapse of vapour bubbles that are formed when pressure within the pump drops below the vapour pressure of the pumped product. As pressure increases within the pump, the vapour bubbles re-condense (implode) resulting in shock waves that cause damage to pump components.

The pumping of entrained gas is different in that the gas is non condensable. Centrifugal action tends to cause a separation between the liquid and gas due to differences in density. This centrifugal action moves the liquid to the outside and concentrates the gas in the eye of the impeller. This restricts the flow area for the liquid causing significant pressure drop and hence cavitation even though NPSH calculations will indicate that cavitation should not occur.

Some industries introduce air into products as part of the process. An example is the Pulp and Paper industry where air in the range of 4-10% is introduced into pulp slurry as part of the ink removal process in paper recycling plants. Some processes also involve pumping a two-phase flow. Excess agitation in suction vessels or vortex formation due to inadequate submergence or incorrect sump designs may result in introduction of undesirable gas entrainment.

The proper selection of a centrifugal pump for liquid and gas (two-phase) mixtures is highly dependent on the amount of gas and the characteristics of the liquid. The presence of entrained gases will reduce hydraulic performance of centrifugal pumps and can potentially cause loss of prime. Standard centrifugal pumps’ pump designs can be used for entrained gases up to 4% by volume.

For gas entrainment values above 4%, specifically modified impellers can be used effectively. Pump performance corrections are required in all cases with gas content from around 2%. Gas concentrations above 10% can be handled, however pumps designed with specific gas handling characteristics are required e.g. vortex impeller pumps, side channel pumps and regenerative turbine pumps.

Most centrifugal pump can handle low concentrations of entrained gas. As detailed earlier, the gas will accumulate in the eye of the impeller restricting flow and head generation. Continued gas accumulation may cause the pump to vapour lock and lose prime. Fig. 1 below shows performance variation as gas concentrations vary form 0% to 10%. At concentrations up to 2%, the impact is relatively insignificant. Impact on performance is still reasonably acceptable up to 4%.


PC32 Pumping Liquids with Entrained Gas Page 2 of 2 March 25, 2009

As the percentage of gas increases further, the performance begins to quickly deteriorate (Fig. 1) until the pump becomes unstable and loses prime. Increasing the running clearances between the impeller and pump casing (.004 to .008 mm) allows for additional leakage and this may assist in preventing loss of prime at the higher gas concentrations.

Much testing has been done by various manufacturers, however the many variables impacting on the effect of gas entrainment does not allow for presenting specific selection information.

Manufacturers should be contacted for detailed selections on applications where gas entrainment is an issue.


PC33 Suction System Design Page 1 of 8 April 21, 2009

Suction System Design

The design of a piping system can have an important effect on the successful operation of a centrifugal pump.

Selection of the discharge pipe size is primarily a matter of economics. The cost of the various pipe sizes must be compared to the pump size and power cost required to overcome the resulting friction head.

The suction system design is far more important. Many centrifugal pump troubles are caused by poor suction conditions. Such items as sump design, suction piping design, suction and discharge pipe size, and pipe supports must all be carefully considered.

The information detailed below is to be taken as guidelines only and all readers should have specific designs undertaken for individual applications.

Suction Piping Design

The function of suction piping is to supply an evenly distributed flow of liquid to the pump suction, with sufficient pressure to the pump to avoid excessive turbulence in the pump impeller.

The suction pipe should never be smaller than the suction connection of the pump and in most cases should be at least one size larger. Suction pipes should be as short and as straight as possible. Suction pipe velocities should be in 1.5 to 2.5 metres per second range unless suction conditions are unusually good.

Higher velocities will increase the friction loss and can result in troublesome air or vapour separation. This is further complicated when elbows or tees are located adjacent to the pump suction nozzle, in that uneven flow patterns or vapour separation keeps the liquid from evenly filling the impeller. This upsets hydraulic balance leading to vibration, possible cavitation and excessive shaft deflection, especially on high and very high suction energy pumps. Shaft breakage or premature bearing failure may result.

Ideally, a straight length of pipe of an equivalent length of five times the pump inlet size (5D) should be installed before any fitting or valve. Please refer to individual pump instruction books for individual manufacturer’s recommendations.

On pump installations involving suction lift, air pockets in the suction line can be a source of trouble. The suction pipe should be exactly horizontal, or with a uniform slope upward from the sump to the pump as shown in Fig. 1. There should be no high spots where air can collect and cause the pump to lose its prime. If high spots are unavoidable, automatic vent valves should be installed at the high points on the piping.

Eccentric rather than concentric reducers should always be used, on horizontal installations, with the flat side located on top.

If an elbow is required at the suction of a double suction pump, it should be in a vertical position if at all possible. Where it is necessary for some reason to use a horizontal elbow, it should be a long radius elbow and there should be a minimum of five diameters of straight pipe between the elbow and the pump as shown in Fig 2.

Fig 3 shows the effect of an elbow directly on the suction. The liquid will flow toward the outside of the elbow and result in an uneven flow distribution into the two inlets of the double suction impeller. Noise and excessive axial thrust will result.



Fig 1 Air pockets in suction piping



Fig. 2 Elbows at pump suction

Fig. 3 Effect of elbow directly on suction Supply Tank and Sump Design There are several important considerations in the design of a suction supply tank or sump. These are: Turbulence It is imperative that the amount of turbulence and entrained air be kept to a minimum. Entrained air will cause reduced capacity and efficiency as well as vibration, noise, shaft breakage, loss of prime, and/or accelerated corrosion. The free discharge of liquid above the surface of the supply tank at or near the pump suction can cause entrained air to enter the pump. All lines should be submerged in the tank, and baffles should be used in extreme cases as shown in Fig. 4.



Fig. 4 Keeping Air out of pump

Submergence

Improper submergence of the pump suction line can cause a vortex, which is a swirling funnel of air from the surface directly into the pump suction pipe. In addition to submergence, the location of the pipe in the sump and the actual dimensions of the sump are also important in preventing vortexing and/or excess turbulence.

The amount of submergence required depends upon the size and capacity of the individual pumps as well as on the sump design. Past experience is the best guide for determining the submergence. The pump manufacturer should be consulted for recommendations in the absence of other reliable data.

1. Pump Flowrates below 315 lites/sec

For horizontal pumps, Fig. 5 can be used as a guide for minimum submergence and sump dimensions for flows up to approximately 5000 US gallons/min (315 litres/sec). For larger flowrates , refer to item 2 below.



Fig. 5 Minimum suction pipe submergence and sump dimensions



Baffles can be used to help prevent vortexing in cases where it is impractical or impossible to maintain the required submergence. Fig. 6 below shows three such baffling arrangements.

Fig. 6 Baffle arrangements for vortex prevention On horizontal pumps, a bell should be used on the end of the suction pipe to limit the entrance velocity to 1 - 2.5 metres per second. Also, a reducer at the pump suction flange to smoothly accelerate and stabilize the flow into the pump is desirable.

2. Pump Flowrates above 315 litres/sec

For larger units (over 315 litres/sec) taking their suction supply for an intake sump (especially vertically submerged pumps), requires special attention.

The function of the intake structure, whether it is an open channel, a fully wetted tunnel, a sump or a tank, is to supply an evenly distributed flow to the pump suction. An uneven distribution of flow, characterised by strong local currents, can result in formation of surface or submerged vortices and with certain low values of submergence, may introduce air into the pump causing a reduction of capacity, an increase in vibration and additional noise. Uneven flow distribution can also increase or decrease the power consumption with a change in total developed head. The ideal approach is a straight channel coming directly to the pump or suction pipe. Turns and obstructions are detrimental, since they may cause eddy currents and tend to initiate deep-cored vortices.

The amount of submergence available is only one factor affecting vortex-free operation. It is possible to have adequate submergence and still have submerged vortices that may have an adverse effect on



pump operation. Successful, vortex-free operation will depend greatly on the approach upstream of the sump.

Complete analysis of intake structures can only be accurately accomplished by scale model tests. Model testing is especially recommended for larger pumping units.

Subject to the qualifications of the foregoing statements, Figures 7 through 10 have been constructed for single and multiple intake arrangements to provide guidelines for basic sump dimensions.

D = (760 x Q)0.5 S = D + (29560 x Q)/ D1.5

W = 2D Where

Y > 4D S – in mm

A > 5D D – in mm

C = .3D to .5D Q – flow in litres/sec

Fig. 7 Sump dimensions calculations

Fig. 8 Sump dimensions, Plain view, wet pit type pumps

Fig. 9 Sump dimensions, elevation view, wet pit type pumps



Optimum Sump Volume

During start-up of an electric motor, motors encounter high starting currents. These effects put a constraint on the maximum number of starts in a given time and hence there is an optimum sump volume to minimise the possibility exceeding the maximum allowable starts for motors.

Pumps start most frequently when the flowrate into the sump is exactly half the pumping rate (Q litres/sec) and the cycle time (t secs) is determined by the following formula:

t = 240 x V /Q

where V= volume in litres

Therefore for a maximum allowable start frequency of 10 starts/hour (i.e. t =360 secs), V =1.5 x the pumping rate (Q litres/sec) or 6 minutes


PC34 10 Misconceptions on Rotary PD Pumps May 27, 2009

10 Misconceptions on Rotary PD Pumps

Reprinted with kind permission of Viking Pump Inc, a unit of IDEX Corporation

(Authors John Petersen, Technical Customer Service and CH Tan, Regional Director for Asia Pacific Region)

PUMPS & SYSTEMS www.pump-zone.com NOVEMBER 2007 75

There is considerable resident knowledge about cen-trifugal pumps, which comprise the majority of all pump installations. However, rotary positive dis-

placement (PD) pumps are not well understood by many specifying engineers and users.

Rotary PD pumps account for 15 percent of all installa-tions, with sales divided into many different pumping tech-nologies. As a result, PD pumps are sometimes misapplied, incorrectly specifi ed or simply not used where they could or should be. Though PD pumps move liquids through many different approaches, much commonality is shared in design and operation.

Centrifugal pumps move liquid in a much different fash-ion than do PD pumps, and their resulting performance dif-fers as well. Centrifugal, or kinetic, pumps impart rotational energy to the liquid and convert it to potential energy (pres-sure) through the design of the volute. PD pumps, on the other hand, move liquids by transferring confi ned amounts of liquid (defi ned by pumping element geometry) from the inlet to the outlet of the pump.

The key here is that centrifugal pumps generate pres-sure and fl ow results, while PD pumps generate fl ow and pressure results. In other words, a PD pump generates just enough pressure to overcome system resistance created by the fl ow of liquid through it. Because of this, fl ow output from a centrifugal pump varies with differential pressure, whereas fl ow from a PD pump is essentially constant with varying differential pressure.

All of the many different types of PD pumps use geom-etry of the parts to expand and contract volumes of liquid. Volume expansion draws liquid into the pump and volume contraction moves liquid out of the pump. Though a number of geometries are involved, most rotary PD pumps share common design and operating characteristics:

All try to displace the same amount of liquid with each • rotation of the shaft.Flow is directly proportional to speed.• All can pull liquid from below the pump or self prime.• Most have close fi tting internal parts.• Most have one pumping element driving another (gears, • rotor/vanes, etc.).All have a small amount of liquid that goes from dis-• charge back to suction. This is called slip, and it varies with liquid viscosity and differential pressure.All require some form of overpressure protection.•

Conventional wisdom says that a PD pump must be used over a centrifugal pump when one or more of the fol-lowing application conditions exist:

Liquid viscosity is too high (generally anything over • 150-cps requires a PD pump).Constant fl ow is needed over varying differential pressure.• Suction lift or self priming ability is needed.•

Although these criteria are all sound, PD pumps may still not be used, or there may be rationale to use a PD pump when none of these conditions exist. At this juncture, many of the common misconceptions on PD pumps come into play. Let’s explore ten of them.

1. PD Pumps Are Not Well Suited for Thin LiquidsAsk almost any pump user what type of pumps to use for a thin liquid application and the response is generally a cen-trifugal pump. This is often the right answer, but sometimes not. Some believe that PD pumps cannot be used on thin liquids at all due to their design characteristics. For exam-ple, how can a gear pump be used on a thin non-lubricating

Ten Misconceptions on Rotary PD PumpsJohn Petersen and C.H. Tan, Viking Pump, Inc.

Rotary positive displacement (PD) pumps are not well understood by many specifying engineers and users and are sometimes misapplied, incorrectly specifi ed or simply not used where they could or should be. This article explores the areas of commonality between centrifugal and PD pumps and ten common PD pump misconceptions and the reality behind each.

Practice + Operations

76 NOVEMBER 2007 www.pump-zone.com PUMPS & SYSTEMS

Practice + Operations

liquid when one gear drives the other? The fact is most PD pumps – including gear pumps – can

be used on thin liquids. Water is the most common thin liquid, and the internal gear pump was actually invented to handle it. Liquefi ed petroleum gas, refrigerants, solvents, fuel oils, gaso-line and even liquid carbon dioxide are some of the other thin liquids handled very successfully with PD pumps. Selecting proper pump materials is important when moving a thin liquid, and most manufacturers offer many choices to handle the low lubricity and viscosity typically associated with thin liquids.

2. PD Pumps Do Not Need Overpressure ProtectionA centrifugal pump that is dead headed, either accidentally or on purpose, can develop a predictable maximum shut off pres-sure. It will be above the normal operating pressure, but gener-ally not much more.

A PD pump that is dead headed tries to displace the same amount of liquid for each revolution of the shaft. Because of this, pressure continues to increase until something breaks in the system, the pump is damaged or the driver runs out of power. None of these are safe or desirable conditions. To pre-vent this, having some form of overpressure protection either on the pump or in the system is important.

Many manufacturers offer pump mounted relief valves, but other ways can accomplish the same purpose. System relief valves, rupture disks, torque limiting couplings and motor power load monitors can all be used to limit maximum attainable pressure in the pump or system. Though a system is designed to be continually open, inadvertent valve closing, plugged fi lters or other system upsets may cause enough block-age to signifi cantly increase pressure. Overpressure protection is a must with all PD pumps – and this is often overlooked.

3. PD Pumps Damage Shear Sensitive LiquidsThis may well be the most common misconception on PD pumps, particularly with gear type pumps. With close fi tting internal parts, the liquid is often thought to be simply sheared or damaged by close running components. While it is true that some of the liquid is sheared within the close internal running clearances, only a small amount of liquid is actually being sheared within the pump. The vast majority of liquid going through the pump comes through in “chunks” and is not sheared at all.

Numerous applications and actual testing have busted this myth. Most notable is wastewater polymers, where excessive shearing changes the viscosity. Gear pumps are frequently used to transfer these polymers at actual manufacturing facilities and do no damage to the liquid. Many installations of gear pumps at wastewater treatment facilities operate without affecting liquid properties.

Field testing of other sensitive liquids proves these pumps

can be used very successfully. One particular customer would not allow use of gear type PD pumps because of perceived liquid damage. Actual testing showed this perception to be wrong. Most important, PD pump manufacturers typically reduce speed and slightly increase internal clearances to minimize the effects of shear. The takeaway is PD pumps can handle shear sensitive liquids without damage if properly applied.

4. PD Pumps Are Not Suitable for Abrasives or SolidsTrue, some PD pump principles do not handle solids, but most can handle some form of abrasives. Lobe pumps and progressing cavity pumps do a good job on solids and can handle abrasives too. Gear pumps handle abrasives quite well, with a few mate-rial changes to retard wear, but they will not tolerate solids.

A large majority of shingle manufacturing relies on gear pumps to handle a mixture of asphalt and up to 60 percent to 70 percent fi nely ground stone. This mixture is quite abrasive, but reasonable life can be attained by reducing pump speed and using hard parts in key wear areas.

Abrasive wear is caused by forces acting on the relative motion surfaces within the pump in the presence of an abrasive media. Pressure and speed of the pump create these forces, so minimizing them can extend pump life. Reducing pump speed is the logical way to do this. Most manufacturers recommend this in varying degrees, depending upon the abrasive nature of the product being pumped. Pressure is a function of the system, so proper design helps extend pump life by reducing the load the pump is required to handle.

5. PD Pumps Cannot Handle Non-Lubricating LiquidsNon-lubricating liquids range from thin to thick viscosity. Many thin liquids have poor lubricating properties, but so do some thick liquids. For example, number 6 fuel oil (sometimes used for heating or diesel fuel) can be as thick as 15,000-cps, but it is not very lubricating.

Manufacturers tend to limit the load from pumping ele-ment contact in PD pumps by limiting maximum pressure with non-lubricating liquids. Doing this minimizes pumping element wear on most non-lubricating liquid applications.

The critical area in PD pumps is pumping element sup-port, which can be either a journal bearing operating in the liquid pumped or an external antifriction bearing. PD pump manufacturers have a wide array of journal bearing and shaft materials to handle low lubrication situations. The key is choos-ing the right materials for the particular liquid characteristics to get the best pump life.

External antifriction bearing support takes away the mate-rial problem, but in most designs this moves the support far-ther away from internal pump loads, resulting in higher bearing loads that may also limit maximum pressure on a particular design.

PUMPS & SYSTEMS www.pump-zone.com NOVEMBER 2007 77

6. PD Pumps Do Not Run FastBrowsing product information for a number of different pump designs and manufacturers reveals this is not true. Granted, PD pumps must run more slowly as viscosity increases and also when fl ow is above 200-gpm, but direct motor speed operation is not uncommon.

Many smaller displacement pumps run at direct motor speeds of 1800-rpm, with some even going to 3600-rpm. Motor speed PD pumps provide an economical solution to liquid transfer for any number of applica-tions where liquid properties like viscosity, abrasiveness and shear sensitivity are not a problem.

7. PD Pumps Are Expensive to OwnThe initial cost of most PD pumps is more than centrifugal pumps, but one must consider the total cost of ownership. Using a simplistic approach to cost of ownership, the main con-tributors are fi rst cost, repair costs and energy to operate.

Assuming an average pump life of seven years, a couple of repairs during its life span and rather modest pressure require-ment energy to operate the pump equals one half of the total cost of ownership. First cost is the lowest, with repair parts coming in second. Figure 1 shows the detail. PD pumps can be quite effi cient. A small increase here can save considerable money over the life of the pump.

8. PD Pump Repairs Are ExpensiveRepair parts costs vary between manufacturers. Since these costs differ, understanding them before the pump purchase is always a good idea. Misconception 7 above assumed an aver-age pump life of seven years with two lower level repairs and a major overhaul (replacing all critical wear parts) during the life of the pump. Cost for the parts required for a major overhaul is only two-thirds the price of a new pump. PD pumps are usually easy to work on too, so labor costs are not high.

9. PD Pumps Have Pulsing FlowThis is true for reciprocating PD pumps, but not for most rotary PD pumps. This may seem counterintuitive since these pumps move liquid by delivering confi ned volumes of liquid from the suction to the discharge port. Although it would seem that confi ned volumes or buckets of liquid would result in fl ow pulsations, they actually don’t.

In theory, gear, vane, some lobe pumps and all types of screw pumps deliver a continuum of liquid resulting in very little, if any, fl ow pulsations. There may be slight differences in slip within the pump, depending on rotational position of the pumping elements that result in minimal pressure pulsa-

tions, but theoretical output remains constant. Traditional lobe pumps with three lobed rotors do have theoretical fl ow pulsa-tion, but again this is minimal.

10. PD Pumps Cannot Run DryPD pumps self prime, meaning they are capable of pulling a liquid from a level below the pump into the pump port. This means the pump must run without liquid for the time it takes to get liquid up the elevation. In other situations, PD pumps are asked to empty tanks, which many times results in periods of running with little, if any, liquid inside the pump.

Most PD pumps can run dry for short periods of time without damage. In many cases there is a small amount of liquid in the pump, which keeps the parts wetted to the point damage does not occur. Obviously, extended periods of run dry are not recommended and some designs are more tolerant than others. In any case, run dry situations are more common than anyone likes to admit and PD pumps can usually handle them.

SummaryRotary PD pumps are (for the most part) simple devices, but more complex in application. Understanding the basic operat-ing characteristics of PD pumps and system requirements is a great place to start in correctly applying a pump.

Recognizing these common misconceptions opens up a whole new application area for these pumps to improve perfor-mance, extend service life or operate more effi ciently. Consider all the alternatives to a pumping problem before making a deci-sion – you’ll be surprised by how many choices you have.

P&S

John Petersen is vice president, technical customer service and CH Tan is regional director for the Asia Pacifi c Region of Viking Pump, Inc., a unit of IDEX Corporation, 406 State Street, P.O. Box 8, Cedar Falls, IA 50613-0008, 319-266-1741, Fax: 319-273-8157, www.vikingpump.com.

FIRST COST19%

ENERGY COSTS57%

TOTAL REPAIR24%

PUMP LIFE CYCLE COST

Cast Iron Pump: 140 gpm, 7500 ssu, 100 psi

Figure 1


PC35 Single vs Parallel Operation of Centrifugal Pu8mps June 29, 2009 Page 1 of 4

Single Versus Parallel Operation of Centrifugal Pumps

Parallel operation may be required to meet variable demands, such as flood control, or to satisfy a temporary condition that occurs such as when changing over pumps in an uninterruptible process. Figure 2 illustrates the characteristic of single vs. parallel operation. Two identical pumps operating in parallel are capable of producing twice the flow of a single pump at any given Total Dynamic Head (TDH). However, the actual flow rate realised in the system is dictated by the intersection of the system curve with the pump curve. Unless the system curve is variable, the flow increase may not be that significant. For example, assume that there is a set of fixed spray nozzles, where the system resistance is purely frictional, and varies only as a result of flow change. When a second pump is introduced, resistance in the system increases as the flow increases. The flow will increase only to where the system curve intersects the two-pump curve, as shown in figure 3. The amount of flow increase is dictated not only by the system curve, but also by the steepness of the pump curves. Pumps with flat curves will have less TDH separation than pumps with steep curves and therefore will have less flow rate change.



A fire pump installation would be an example of a system that has a variable system curve. Figure 4 illustrates this. Each time an additional fire nozzle is activated, the system resistance is decreased. This causes the system curve to move to a higher flow on the pump curve, increasing the kW load on the pump and decreasing the amount of TDH available. Eventually, another fire pump may need to be activated to maintain the system pressure as more nozzles come on line.

CONTINUOUSLY RISING TDH CURVES

Pump specifications often dictate that pumps have a continuously rising head curve to shut-off. It is a characteristic of certain pumps to have a head curve that droops as the flow approaches shut-off. This characteristic is quite prevalent in pumps with specific speeds under 30 (Ns~1550 US units). Curves with a drooping shut-off characteristic may experience load sharing problems when operating in parallel with other pumps.

In figure 5, the system and TDH curves intersect at a TDH that is greater than the shut-off TDH value and at a lower TDH than the max TDH of the curve. When the second pump starts, the flowrate will increase only to the first point where the TDH matches the pressure in the system. Because the first pump’s



operating TDH point is mirrored by another at a lower flow, the second pump’s flowrate will not increase beyond the low flow point. At best, the pumps will share the load unevenly, with the second pump operating at a lower efficiency. Worse, the second pump might operate at less than its minimum allowable design flow, resulting in damage to the equipment and possible injury to personnel.

A similar condition may exist when one pump is worn. As a pump wears, the amount of TDH produced at any given flow rate diminishes. If the new pump’s intersection with the system curve is above the shut-off TDH of the worn pump, it will force the worn pump into a shut-off condition. Load sharing problems, between pumps operating in parallel, may increase wear, reduce seal and bearing life, lower operating efficiencies and limit process operations. In the absence of any flow measurement capability, an uneven performance distribution, between pumps operating in parallel, is easier to avoid than to detect. Proper pump selection for parallel operation and pump performance monitoring are the best tools in avoiding load sharing problems and maintaining a well operating parallel pump installation. Parallel Operation of Unmatched Pumps

When pumps operate in parallel, the flow rate at any given TDH point is additive. In the case of pumps that have identical operating characteristics, the flow would double. For example, two pumps that each had a capacity of 100 M3/hr at 50 M TDH would have a combined capacity of 200 M3/hr at 50 M TDH. Again, the system curve does not change, so the actual change in flow that occurs with bringing a second pump on line, in parallel, is determined by the characteristic curve intersection with the system curve. This is shown graphically in Fig. 3. When a pump is operated in parallel with another pump that has a different operating characteristic, the same rule applies as for identical pumps: for any TDH, common to both pumps, the flow characteristic will be additive. If one pump exhibits a lower shut-off TDH characteristic, it will operate at shut-off until the dominant pump moves far enough out on its curve so that its TDH falls below the shut-off TDH of the pump with lower head (Fig 6). The danger here is in the system-pump interaction. If the system curve intersection with the characteristic curve is at a higher TDH than the shut-off flow of the weak pump, the weak pump will be forced to run at shut-off and a serious failure could occur. In Fig. 6, a zero-flow condition for the weak pump will exist when the system curve intersection is to the left of the vertical dashed line.



As a general rule it is a good idea to have flow measurement installed for any pumps designed to operate in parallel. Without flow measurement, it is very difficult to determine what the load sharing is between two pumps.

Motor power is often a questionable indicator of flow, as many power curves are quite flat and show small changes in load over relatively large changes in flow. Also, when wear does occur, the power draw may remain relatively constant even though performance is falling off. This is due to a decrease in pump efficiency which is not visible to the pump operator.


PC36 Metering Pumps Page 1 of 14 12-Aug-09

METERING PUMPS

The metering pump is a positive displacement chemical dosing device with the ability to vary capacity manually or automatically as process conditions require. It features a high level of repetitive accuracy and is capable of pumping a wide range of chemicals including acids, bases, corrosive or viscous liquids and slurries. The pumping action is developed by a reciprocating piston, plunger or diaphragm which is either in direct contact with the process fluid, or is shielded from the fluid by a diaphragm. Diaphragms may be activated by direct mechanical link or by hydraulic fluid.

Metering pumps are generally used in applications where one or more of the following conditions exist:

• Low flow rates are required

• High system pressure exists

• High accuracy feed rate is demanded

• Dosing is controlled by computer, microprocessor, DCS, PLC, or flow proportioning

• Corrosive, hazardous, or high temperature fluids are handled

• Viscous fluids or slurries need to be pumped

METERING PUMP CHARACTERISTICS 1. The pumping action is developed by the reciprocating action. This reciprocating motion develops a

flow sine wave. Actual flow rate is determined by the following formula:

Figure 1



2. Unlike centrifugal pumps, flow rate is not greatly affected by changes in discharge pressure.

Figure 2

3. The metering pump flow vs. stroke characteristic curve is linear. It is not however, necessarily proportional in that 50% stroke setting may not equal 50% flow. This is due to the fact that the calibration line may not pass through 0 on both axes simultaneously. By measuring flow at 2 stroke settings, plotting both points and drawing a straight line through them, other flow rates vs. stroke can be accurately predicted. The steady state accuracy of a correctly installed industrial grade metering pump is generally +/- 1.0% or better. Although a metering pump can generally be adjusted to pump at any flow rate between 0 and its maximum capacity, its accuracy is measured over a range determined by the pump's turndown ratio. Most metering pumps have a turndown ratio of 10:1 which simply means that the pump is within its accuracy rating anywhere between 10% and 100% of capacity. Some newer designs of metering pumps feature higher accuracy, and a greater turndown ratio of 100:1. Therefore, this design will accurately dose anywhere between 1% and 100% of capacity.

Figure 3



METERING PUMP DESIGN

LIQUID END The liquid end design and materials of construction are determined by the service conditions, and the nature of the fluid to be handled. Temperature, flow rate, fluid viscosity, corrosiveness and other factors are considered. DRIVE MECHANISM The drive mechanism translates the rotary motion of the driver into reciprocating movement. FLOW ADJUSTMENT Pump flow rate is adjustable by varying stroke length, effective stroke length or stroking speed. Most metering pumps are supplied with a micrometer screw adjustment similar to the one shown here. The micrometer can also be replaced by an electronic or pneumatic actuator to adjust pump flow rate in response to a process signal

DRIVERS The pump is usually driven by an AC constant speed motor. Electromagnetic drive is available in small flow pumps.

Figure 4

DRIVE MECHANISMS Metering pumps can be powered by a variety of drivers however, the almost universal driver is an electric motor. The motor speed is normally reduced to pump design speed by the use of gearing built into the pump power end. This rotary power is converted to a linear motion through one of three methods:

- a crank mechanism with either fixed (Figure 5) or variable (Figure 6) stroke length

- an eccentric or cam arrangement (Figure 7)

Depending on the type of adjustable output flow mechanism used, the power can be utilised on both the forward thrust of the crank and the back thrust of the crank. The eccentric or cam arrangement, however, can provide power in only one direction.

Metering pumps with solenoid power ends (Figure 8) are another type of drive and create linear, reciprocating motion using electromagnets. The solenoid type could be considered the ideal power end as it does not require any type of transmission to convert motion from rotary to linear. Another advantage is that it has in built overload protection as the pump simply stops at excessive load. A disadvantage is that availability is limited to very small powers.



Figure 5

Figure 6



Figure 7

ADJUSTMENT SCREW

Figure 8



LIQUID END Metering pumps generally fall into five basic types:

- Piston with packed seal - Plunger with gland packed seal - Mechanically actuated diaphragm - Hydraulically actuated diaphragm - Hydraulically actuated tube Both the piston packed pump and the plunger packed pump allow some degree of leakage past their dynamic seals. In some cases this is not an objectionable shortcoming; in other cases, it can be very objectionable and, in most instances, costly as well. Diaphragm and tube pumps are seal/less and therefore overcome this potential problem. It is often possible to multiplex pumps i.e. coupling or ganging of two or more liquid ends and use only one driver. The main purposes of multiplexing are: - provide greater flow or pressure without significantly increasing the driver size - provide exact proportioning or synchronization between heads - provide greater turndown ratio i.e. flow control range - reduce pulsations PISTON WITH GLAND PACKED SEAL This type of pump is rarely used these days. The piston is driven by either a crank, a connecting rod, or a crosshead driven by the crank. The piston provides the liquid flow and is designed to displace a measured volume of liquid with a high degree of accuracy as it reciprocates within the pump (Figure 9). Rings or packing located on the piston move back and forth with the piston to effect a dynamic seal with the inside diameter of the cylinder and a static seal with the outside diameter of the piston. Packing on the connecting rod provides at static seal so that it becomes double acting and displacement may occur on both directions. This provides double the flow of single acting pumps.

Figure 9. Piston packed pump

The forward travel of the piston reduces the internal volume of the liquid chamber, displacing the metered liquid out the discharge check valve. The pressure required to move the liquid through the discharge check valve is also applied to the suction check valve, forcing it into a closed position, ensuring correct flow direction



The reverse travel of the piston decreases the pressure within the liquid chamber by enlarging the internal volume of the chamber. This change of pressure results in a rapid closing of the discharge check valve caused by the external pressure acting on the valve and allows the suction check valve to open because of an external pressure under the check valve that can be either above or below atmospheric pressure. The accuracy of the reciprocating metering pump is achieved by the previously described predetermined controlled piston travel of the pump, the control of the strokes per minute, and the precise opening and closing of the check valves. The inaccuracy, on the other hand, is caused by leakage past the piston packing and the check valves. PLUNGER WITH GLAND PACKED SEAL The plunger packed pump is very similar to the piston packed pump except for the packing design and location. The packed plunger, unlike the packed piston, has the packing installed in a stationary gland in the inside diameter of the cylinder. There is no moving or dynamic seal located on the plunger. As the plunger reciprocates within the pump, a dynamic seal is made between the outside diameter of the plunger and the inside diameter of the packing, and a static seal is made between the outside diameter of the packing and the inside diameter of the stuffing box (Figure 10). There is a strong trend to leak free (diaphragm and tube) metering pumps, however plunger pumps continue to be used. The main reasons are that plunger pumps are cheaper, can develop higher discharge pressures and are more suitable for extreme temperature conditions.

Figure 10 Plunger with packed gland

MECHANICAL ACTUATED DIAPHRAGM To overcome the leakage problem, a diaphragm pump can be used. The power side of the pump and the capacity control are the same as was previously described for other types of reciprocating pumps. However, in place of a piston rod or plunger, the mechanical diaphragm pump uses a connecting rod fastened to the centre of a diaphragm. The mechanical diaphragm pump's principle of positive displacement output is similar to that of the piston plunger pump except that the diaphragm becomes the displacement measuring element, as it moves back and forth in the fluid chamber (Figure 8). HYDRAULIC DIAPHRAGM PUMP The hydraulically balanced diaphragm pump is a hybrid design that provides the principal advantages of the other three pump types. Like the other pumps, its power end and capacity control are common. This, however, is where the similarity ends, since the piston or plunger does not come into contact with the



pumped fluid, and the actuation of the diaphragm is by hydraulic power instead of mechanical power (Figure6). The measuring piston or plunger reciprocates within a precisely sized cylinder at an established stroke length, displacing a volume of hydraulic liquid, not the product liquid. The hydraulic liquid is stable and has excellent lubricating qualities. The piston uses the hydraulic oil to move the diaphragm forward and backward, causing a displacement that expels the product liquid through the discharge check valve and, on the suction stroke, takes in an equal amount through the suction check valve. The diaphragm isolates the liquid product being contained within the liquid chamber and check valves. These are the only parts that must be made of chemically compatible material. The diaphragm's only job is to separate two liquids. It normally does no work, carries no load, and pumps no liquid; rather it serves as a moving barrier between liquids during periods of pressure imbalance. It is simply a moving partition with pressure hydraulically balanced on both its sides; on one side is the liquid product and on the other side is the hydraulic oil. At full deflection, the diaphragm undergoes total combined stresses well within the endurance limit of the diaphragm material. Contoured support plates are provided on either side of the diaphragm to ensure that stresses are kept within limits. When properly installed and working within the recommended temperature range and not affected by corrosion or abrasion, the diaphragm has an unlimited life. As previously stated, the piston or plunger handles only hydraulic oil. Conventional seals are used on the piston or plunger, which does not require power flushing and complicated drain systems as are found on conventional piston or plunger pumps handling corrosive or hazardous liquids. Even the slightest leakage past the piston is replaced on the suction stroke through the automatic functioning of a compensation system, which draws in replacement oil from the oil reservoir (Figure 11).

Figure 11 Function of oil make-up valve

Any excess pressure within the hydraulic system or the liquid product chamber is relieved through the automatic action of a pressure relief valve. This valve blows off oil, under excess pressure ahead of the piston, back into the oil reservoir. This valve blows off oil, under excess pressure ahead of the piston, back into the oil reservoir (Figure 12).



Figure 12 Function of pressure relief valve The vacuum and pressure compensator systems actually perform three important functions that the other described types of metering pumps cannot do unless auxiliary equipment is added to their piping systems. As described previously, they compensate for any leakage occurring within the hydraulic system of the pump, ensuring a balanced diaphragm movement. In addition, they serve to protect the process system from an over-pressure condition produced by the pump. For instance, the positive displacement pump, because of its design, must over pressure the system to the point of damaging the pump, bursting pipes, or damaging other downstream equipment should an operator mistakenly close a shut-off valve downstream from the pump. The hydraulic diaphragm pump will, however, relieve any pump-produced pressure beyond the set pressure of the pressure relief valve, thus avoiding the dangerous build up of pressure. The compensation system also serves to protect the pump from a closed suction line or a partially clogged strainer in the suction line. Should this occur, the backward movement of the diaphragm is prevented and the vacuum relief system would automatically open to relieve the starved suction condition within the pump. In doing so, however, a surplus of hydraulic oil enters into the system between the diaphragm and piston. As the piston starts forward on its discharge stroke, the diaphragm is displaced forward and will come into contact with the contoured dish support plate in the process liquid chamber, because of the surplus oil drawn into the hydraulic chamber. At the moment of diaphragm contact with its support plate, an over-pressure condition starts to develop within the hydraulic system. The pressure relief valve now opens to relieve the surplus oil back into the hydraulic reservoir, preventing a dangerous build up of pressure. The interaction of the two compensation systems continue stroke after stroke to activate a fluid-clutch-type action to prevent overloading of the pump's power end until the condition plugging the suction or discharge lines is found and corrected.



FLOW ADJUSTMENT Flow control can be achieved in a number of ways, however the most common are as follows. 1. Changing the stroke rate within the pump by changing internal gearing in the drive mechanism. This

is done at time of pump manufacture and further changes cannot be made after pump installation.

2. Changing the stroke rate by changing the driver speed. When the driver is an electric motor, using a frequency inverter will achieve this result. With solenoid operated pumps, this can be done by setting the switching on the solenoid.

3. Changing the stroke length. This can be achieved in a number of ways and is dependent on the pump construction. Details are given below.

4. Adjustable cranks and hydraulic bypass have also been used in the past but are rarely seen these days.

STROKE LENGTH ADJUSTMENT There are two main categories of stroke length adjustment mechanisms - lost motion and full motion. Each of these designs changes the way in which the internal piston travels within the piston cylinder. Adjustments can be made manually as shown in all the diagrams below or adjustments can be automatic with the use of electric or pneumatic actuators. A 4-20 mA electric or 3-15 psig pneumatic process signal would be required for the actuators. LOST MOTION (Figures 13, 14 and 15) Smaller pumps (low flow) are typically of the lost motion stroke length mechanism design. The pump motor turns the worm shaft which, in turn, rotates the eccentric gear within the pump gearbox. The cam rotates with the eccentric gear and actuates the piston via the cam follower. On each discharge stroke of the pump the cam follower pushes the piston towards the pump reagent head displacing the pump diaphragm. After the piston reaches its full forward position the piston is retracted via spring force. Displacement per stroke is controlled through limiting the rearward travel of the piston. Adjusting the stroke length mechanism extends and / or retracts the internal adjustment screw. When a stroke length of less than 100% is desired the internal adjustment screw is rotated and the rearward piston travel is limited based on the stroke length setting. As a result there is no contact with the cam for a portion of the cam rotation and the piston stops moving until the cam rotates to a position in which contact is re-established with the cam follower.

Figure 13



Figure 14

Figure 15

The flow characteristics produced from a lost motion style of pump are shown below. As indicated, the flow at 100% stroke length can be represented as a sine wave. When the stroke length is decreased the maximum amplitude of each stroke is maintained while the full potential volume per stroke is decreased.



LOST MOTION FLOW CURVE

FULL MOTION (Figures 16,17 and 18) Pumps with larger flow requirements are typically handled by full motion pumps. Full motion pumps rely on internal linkages for the adjustment of stroke length. The pump motor turns the worm shaft which, in turn, rotates the eccentric gear within the pump gearbox. The eccentric gear transmits motion to a connecting rod which is attached to an oscillating housing. The oscillating housing is stationary at its top — the resulting motion is similar to that of a pendulum. Within the oscillating housing exists a housing block which is, in turn, connected to a connecting rod. The connecting rod is attached to the piston. It can be seen from Figure 16 that when the housing block is at its full bottom position (100% stroke length) the piston will maximize its horizontal movement. As a result the pump will produce its greatest potential displacement per stroke. When the housing block is adjusted to its full top position (0% stroke length) the piston will be stationary and, as a result, the pump will produce no flow. Adjustment of stroke length is actually an adjustment of the housing block position within the oscillating housing — this adjustment will determine how far back and forth the piston can travel and, as a result, the volume per stroke that the pump can produce.

DISCHARGE

270

— 60% STROKE LENGTH

— 100% STROKE LENGTH

360

SUCTION

- - (


PC36 Metering Pumps August 12, 2009 Page 13 of 14

Figure 16

Figure 17

Figure 18


PC36 Metering Pumps August 12, 2009 Page 14 of 14

FULL MOTION FLOW CURVE Full motion flow characteristics are detailed below. As with lost motion pumps the flow characteristic at 100% stroke length can be characterized as a sine wave. Adjustment of stroke length decreases the sine wave amplitude (displacement per stroke).

With the use of variable speed motors and drives many metering pumps are controlled through variable speed in lieu of stroke length. The flow characteristics of full motion and lost motion pumps are identical if stroke length is maintained at 100% and motor speed is used to adjust flow. Instead of changing volume (amplitude) per stroke the adjustable motor speed will modify strokes per minute (frequency). The resulting output will be identical regardless of stroke length type. Finally, a discharge pulsation dampener is a typical recommendation for all styles of metering pumps. The discharge pulsation dampener transforms a diaphragm metering pump's reciprocating flow to laminar flow. As a result, the flow characteristics downstream of a metering pump, regardless if it is lost motion or full motion, will be identical when a discharge pulsation dampener is installed. Acknowledgements: Metering Pump Handbook (Pulsafeeder Inc) www.pulsa.com www.miltonroy.com

− 100% STROKE LENGTH

− 60% STROKE LENGTH


PC37 Solids Handling with Centrifugal Pumps October 21, 2009 Page 1 of 5

SOLIDS-HANDLING WITH CENTRIFUGAL PUMPS The following paper considers only the types of impellers available for solids handling and not at the metallurgy or abrasion considerations associated with pumping solids. All impellers can handle solids with different limitations on size and type. The information detailed is for general guidance only and every application should be checked with pump manufacturers or suppliers.

Enclosed Impeller

Enclosed impellers An enclosed impeller incorporates a full front and back shroud. Fluid flows through the internal impeller passages without hydraulic interaction with the stationary casing walls. In a well designed enclosed impeller, the relative velocity between the impeller and the fluid at any given radius is quite small. This results in less wear than other impeller styles.

A portion of the fluid exiting an enclosed impeller leaks back to the pump suction by traveling through the gap between the front impeller shroud and the casing. An enclosed impeller typically has wear rings or radial pump-out vanes to control this leakage. A centrifugal pump with an enclosed impeller is usually not dependent on tight axial clearances to manage leakage. Therefore an enclosed impeller pump can tolerate moderate wear with little adverse effect on overall performance and efficiency.

Wear rings provide an adequate solution for applications that occasionally handle light solids and for the practical design and manufacture of multi-stage pumps where liberal axial clearance precludes tolerance stack-up problems during assembly. Wear rings control recirculation through flow restriction, and are used in conjunction with impeller balance holes to control axial thrust. The flow restriction created by the tight clearances between the stationary and rotating wear ring faces causes very high local velocities hence a high wear rate. Wear rings, because they are subject to a very high flow velocity, will have an unacceptably short life span in an abrasive environment, even when hard materials or treated surfaces are used.

wear rings

Pump-out vanes offer a better alternative for handling abrasive solids. Pump-out vanes control both leakage and axial thrust through a pumping action. Local flow velocities with pump-out vanes are much lower and spread over a greater area. The lower local velocity results in a much lower wear rate. It is not uncommon for pump-out vane life to equal or exceed the life of the main impeller blades.



Impeller w/pump-out vanes

The disadvantage of pump-out vanes is that they consume power while controlling leakage and thrust. When new, a pump impeller equipped with pump-out vanes will likely have a lower efficiency than its wear ring counterpart. However, it will come close to maintaining its "as installed" efficiency throughout its operational life.

An impeller with wear rings loses efficiency rapidly as the rings wear. If ring wear is severe, the high velocity zone can shift from the wear rings to the impeller thrust balance holes resulting in an expensive and premature repair or replacement of the impeller. It is not uncommon to have several outages to replace wear rings over the life of a single impeller when wear rings are used in an aggressive solids application. A disadvantage of enclosed impellers is that the front and back shrouds, rotating in close proximity to the casing walls, generate disc friction that lowers the efficiency of the pump relative to that found in open impeller designs. Another disadvantage is that the enclosed impeller is more easily plugged. Large solids that might otherwise be broken up by the grinding action generated by a rotating open impeller and the stationary casing wall can easily become lodged in the eye of an enclosed impeller. This may create a mechanical or hydraulic imbalance that has the potential to damage the pump, or at the least causes a pre-mature outage to remove the blockage.

Open Impellers An open impeller is characterized by impeller blades that are supported almost entirely by the impeller hub. This is the simplest impeller style and it is primarily applied to clean, non-abrasive, low horsepower applications. An open impeller is lighter in weight than its shrouded counterpart. Less impeller weight reduces shaft deflection and enables the use of a smaller diameter shaft, at a lower cost, than an equivalent shrouded impeller.

Open Impeller

An open impeller typically operates at a higher efficiency than a shrouded impeller of the same specific speed. The largest contributor to efficiency loss in an enclosed radial impeller is disc friction caused by the front and back impeller shrouds turning in close proximity to the stationary casing walls. Removing the shrouds eliminates the disc friction.

One drawback of the open impeller is that it is more susceptible to abrasive wear than a shrouded impeller. High velocity fluid on the impeller blades in close proximity to the casing walls establishes rotating vortices that accelerate wear when abrasives are present.

A tight clearance between the impeller and the front and back casing walls is necessary to maximise



efficiency. As the impeller wears, these clearances open and efficiency drops rapidly. The tight operating clearances required on both sides of an open impeller for efficient operation precludes adjustment of the impeller axial position to compensate for wear.

Semi-Open Impellers A semi-open impeller is a compromise between an open and an enclosed impeller. It incorporates a single shroud, usually located on the back of the impeller. A semi-open impeller has a solids passing capability similar to that found in an open impeller. With only a single shroud a semi-open impeller is easy to manufacture and completely accessible for applying surface hardening treatments. For moderately abrasive slurries, especially if plugging is a concern, a semi-open impeller is a good choice.

semi-open impeller

A semi-open impeller operates more efficiently than an enclosed impeller because of lower disc friction and tighter axial clearances. It has an advantage over an open impeller in that it can be adjusted axially to compensate for casing wear.

High axial thrust is the primary drawback of a semi-open impeller design. Axial thrust balance is manageable through design for both open and enclosed impellers. On a semi-open impeller, the entire backside surface of the shroud is subject to the full impeller discharge pressure. The front side of the shroud is at suction pressure at the eye of the impeller and increases along the impeller radius due to centrifugal action. The differential between the pressure profiles along the two sides of the shroud creates the axial thrust imbalance. This can be managed somewhat through the use of pump-out vanes on the back side of the shroud, but the vanes will start to lose effectiveness if the impeller is moved forward in the casing to compensate for wear. Some manufacturers have integrated an adjustable wear-plate into the casing design so that clearance adjustments can be made. Combined with hard materials or surface hardening treatments, this option provides a good design in lightly to moderately abrasive applications.

An obvious question is why use a semi-open impeller in a solids application if an open impeller with an adjustable wear plate could be used instead? It might seem logical that an open impeller of hard metal construction, used in conjunction with an adjustable wear liner, would combine good solids handling characteristics, with low thrust imbalance, light weight, and adjustability for wear. Unfortunately, true open impellers lack the structural support to prevent blade collapse or deformation under the demands of most industrial applications. A semi-open impeller is well suited for handling solids in applications where the blades might encounter high impact loads from rocks and the like, or in higher power applications. In both situations the shroud provides additional structural support and reinforcement to protect against blade collapse or deformation.

One improvement that has been made to the semi-open impeller is the use of a partial shroud. Most of the pressure developed by the impeller, and most of the shroud surface area, is in the outer diameter region of the impeller. Elimination of the shroud in this area reduces the axial thrust in a semi-open impeller without compromise to the structure support provided by the full back shroud.

semi-open impeller with a partial shroud



Recessed Impellers Recessed impeller pumps are characterized by impeller vanes that either do not extend into the pump casing or extend only partially into the casing, essentially leaving the casing as an open flow passage. Recessed impeller pumps are well suited for handling large or stringy solids. The maximum solid size is usually limited by the pump suction opening such that any solid that enters the pump will pass through.

In operation, some of the fluid is drawn into the rotating impeller and discharged back into the casing through centrifugal action. Through fluid dynamics, the partial flow through the impeller imparts a centrifugal rotating motion to the entire fluid body within the casing. Large solids entering the casing are transported by the rotating fluid body from inlet to outlet without necessarily making contact with the impeller.

In addition to handling large solids, a recessed impeller pump will handle a higher concentration of entrained gas than a traditional pump with the impeller centered in the casing. In a traditional centrifugal pump, gas accumulating at the impeller eye prevents fluid from reaching any down stream part of the impeller. This results in a breakdown of the pumping action. In a recessed impeller pump, gas entering the pump does not have to pass through the impeller to exit the pump. Additionally, gas present in one area of the impeller does not prevent other parts of the impeller from pumping. Gas handling capabilities in excess of 30% by volume have been reported with this style of impeller. However, the actual concentration of entrained gas that can be handled for any specific application is dependent on the phase

characteristics of the fluid and should be determined by test.

The efficiency of a recessed impeller pump will be less than the efficiency of a traditional centrifugal pump. Efficiency losses result from flow recirculation around the impeller passages, and from the inefficiency of a flow pattern where fluid rotates around the casing numerous times prior to exiting the discharge. Efficiencies in the 40%-50% range are common for recessed impeller pumps.

A recessed impeller pump is sometimes promoted for gentle handling, but caution should be the rule before investing. Much of the energy being imparted to the fluid is lost to turbulence and friction, both of which conflict with gentle handling. A better style of pump for gentle handling applications is the screw centrifugal pump described below.

Screw Centrifugal Impeller The screw centrifugal pump impeller is shaped like a tapered Archimedes screw. Originally developed for pumping live fish, the screw centrifugal pump has become popular for many solids handling applications, especially those where gentle handling is an important consideration.

Most screw centrifugal impellers have a single helical vane wrapping around an expanding hub from inlet to outlet. The single passage allows for an easy transition of fluid and solids from the pump inlet onto the impeller. Its inducer-like design exhibits good NPSHR characteristics. Liquid entering the impeller is accelerated more gradually along the smoothly expanding hub to the pump outlet than with a traditional impeller design.

The screw centrifugal pump is a popular choice for handling delicate products such as food and crystals. Its low shear characteristic reduces emulsification when pumping mixtures. The pump's ability to pass long fibrous materials such

as rope without clogging makes it a frequent choice for municipal waste water applications.



A screw centrifugal pump typically has an operating efficiency of 70% to 85%. It has a relatively steeply rising head/capacity curve shape giving it good flow control capability over its allowable operating range.

The relatively large size of the screw impeller is a primary disadvantage of this style pump. The heavy impeller mandates a large shaft and power frame to limit shaft deflection; which increases the unit cost. In addition, the single vane impeller is prone to high side thrust when operating off-design. A 1x rpm vane

passing frequency vibration, that sets up as the single vane outlet passes the casing cutwater during each rotation, is not uncommon. Multi-vane screw impellers that provide smoother operation are available. The primary trade-off is solid size capability.

Disc Impellers Disc impellers incorporate two or more parallel discs and do not have traditional impeller vanes. Instead this design relies upon fluid friction and viscosity to generate a pumping action. As liquid enters the disc impeller, friction between the fluid boundary layer and the disc's surface accelerates the boundary layer to about the same speed as the impeller. Resistance to sheer (or viscosity) between the boundary layer and the adjacent fluid layer creates motion in the adjacent layer as well. Each layer in turn is set in motion by the viscous drag from the adjacent layer.

Slip, or the difference in speed between the disc and each layer, increases with distance from the impeller. The effectiveness of a disc impeller is related to the spacing of the discs and the

viscosity of the fluid. Close disc spacing and higher viscosity produce better performance than low viscosities with wider disc spacing.

A disc impeller is well suited for gentle handling of delicate materials. It also performs well in abrasive services as there is little relative motion between the fluid contacting the impeller and the impeller itself. Disc impellers have entrained gas capabilities superior to standard impeller pumps, as gas can enter the impeller and move through the boundary layers without impediment.

The efficiency of a disc impeller is less than that of a standard centrifugal pump. The efficiency of a disc impeller pump is commonly in the 35%-50% range. Because it relies on close disc spacing for effective performance, a disc pump is typically not well suited for large diameter solids.

Non-Clog Impellers In reality, all impellers are able to be clogged in some way so the term “non-clogging” is, in reality, a misnomer. Non-clog impellers are designed to accommodate large, soft solids and at the same time limiting clogging the pump. The impeller generally has a small number of vanes (maximum of three) and large vane width. The impellers can be either of the closed or semi open type. Leading edges of vanes tend to be rounded so that solids are not caught up and get tangled. As the vane width is large to allow passing of solids, these pumps tend to be applied on applications requiring relatively high flowrates.

Acknowledgment: Lawrence Pump Company


PC38 Improving Centrifugal Pump Performance November 26, 2009 Page 1 of 4

IMPROVING CENTRIFUGAL PUMP PERFORMANCE

Production facilities often require increased output and consider means of improving the performance of installed equipment. For a pump, this is generally the need for increased flow rate.

The applicable concept is the fact that all centrifugal pumps operate at a flow rate that corresponds to the intersection of the pump curve and the system curve.

To increase the flow rate from an existing pump, either the pump operating characteristics, or the system resistance characteristics need to change. If other modifications to the pumping system are being considered such as new piping, heat exchangers, etc, a new system curve needs to be developed as a baseline reference before considering any of the following modifications Options for increased flow from an existing pump:

Increase the impeller diameter

Increase the pump speed

Modify the impeller blades

Install a different impeller

Install a suction inlet splitter vane

Decrease the system resistance

Impeller diameter/speed change In systems where the primary flow resistance originates from friction (pipe, valves, heat exchangers, etc) flow will increase in direct proportion to a change in impeller diameter or speed (a 10% increase in speed or impeller diameter will yield a 10% change in flow). The pump operating point as a relative percentage of the pump best efficiency point will remain unchanged In systems where the primary resistance comes from a pressure vessel, any increase in flow will be primarily dependent on the shape of the pump characteristic curve. Flatter curves will experience a higher percentage of flow increase than steep curves. The pump operating point relative to the pump best



efficiency point may change. The only way to predict the performance is to plot the intersection of the pump and system curves.



Impeller blade modifications In some cases flow may be increased through modifications to an existing impeller or volute. Impeller modifications might include blade shape modifications through machining or grinding and typically result in a flattening of the pump curve, resulting in higher head as the pump reaches higher flows. Head increases are usually less than 10% using these methods. Modifications of this type are usually expensive to perform and are unpredictable in their result or repeatability. They are mentioned here by means of acknowledgement - not suggestion

Installation of a different impeller The available head from an impeller is largely a function of peripheral speed which is set by diameter and RPM. The available flow from an impeller operating at a constant peripheral speed is controlled by the blade angles at the impeller inlet, and the volume of the impeller vane passages. Many pump manufacturers offer multiple impeller designs to fit within a single pump body that, along with interchangeable diffusers or volute liners, offer a significant change in flow characteristic for a given TDH.

Installation of a suction splitter vane Installation of one or more stator vanes adjacent to the pump inlet will have an effect similar to that described under impeller blade modifications. The amount of performance increase achievable is dependant on impeller design and operating point, but a 10-15% increase of TDH at BEP is not uncommon. Unlike the impeller modifications, inlet vane installation is often relatively simple and does not require customization of spare parts. The impact on NPSHR should be determined through testing.



Decrease system resistance Upgrading valves, replacing heat exchangers, or modifying piping or other system components to lower the resistance head at any given flow will result in a pump that will operate at an increased flow. The pump will also have a reduced horsepower requirement as compared to the previous methods that increase pump TDH to achieve the desired increase in flow.

Summary

Although several methods of increasing flow rate have been discussed, it is our opinion that the modification of impellers and installation of suction splitter vanes are not really viable. The reasons are the uncertainty of the improvement and the repeatability of the process particularly in the case of blade modification. The better means of achieving increased flow are

- increase speed

- increase impeller diameter

- reduce system resistance

Other considerations There are a number of associated mechanical and hydraulic issues that should also be evaluated when investigating modifications to increase a pump’s flow rate. It is a good idea to consult the OEM’s service department to assist in evaluating the pump requirements. Some of the items are as follows:

NPSHR: It is likely that the NPSHR will increase if either flow or pump RPM is increased.

Seal flush requirements: The pressure at the seal chamber will increase in proportion to the pump discharge pressure. Make sure that there is adequate seal flush pressure to give the required flow rate through the seal.

Power: Increasing flow rate will result in increased power draw unless it is accompanied by an adequate decrease in system head requirements. When changing speed or impeller diameter, power will vary approximately as the cube of the ratio of the change. Couplings, base plate dimensions, and electrical components will all require re-evaluation.

Mechanical: Planned changes to the operating speed should be evaluated with respect to rotor critical speed, pressure containment, and mechanical seal limitations.

Operating Limits: Changes that move the pump further away from the best efficiency point may decrease pump reliability due to hydraulic instability.

Courtesy of Lawrence Pump Inc

Pump Clinic 39 Peristaltic Pumps Principles of Operation Page 1 of 7

PERISTALTIC PUMPS Principles of Operation The term peristaltic pump is applied to various ranges known as hose or tube pumps. Although there are specific differences between hoses and tubing, the terms hose and tube are interchangeable in this article.

Generally hose pumps have a large wall thickness, polyamide reinforcing layers and are capable of handling differential pressure to 15 bars. Tubing is smaller wall thickness and tube pumps are generally capable of differential pressures to 4 bars. Peristaltic pumps are self-priming rotary positive displacement pumps that operate on the peristaltic principle. The pump consists of three major parts: hose or tubing, housing and rotor. The hose is placed in the tubing bed between the rotor and the housing. The rotor has a number of "rollers” or "shoes" attached to the external circumference. These move across the hose where it is occluded (squeezed) pushing the fluid. The hose behind the shoe or roller recovers its shape, creating a vacuum and drawing fluid in behind it. Liquid is trapped between the rollers specific to the ID of the hose and the geometry of the rotor. Flow rate is determined by multiplying speed (rpm) by the volume of the trapped liquid. The volume moved is consistent, even under a wide range of viscosities or density. The flow rate is directly proportional to the gearbox speed (rpm).

Dry-Running Design

This design incorporates a unique tube bed that always ensures one roller is occluding the hose. This is termed dry-running because rollers that do not operate in a lubricated bath occlude the hose. Dry running pumps generally have lower flow and pressure capabilities than lubricated pumps.

Dry-running pump design

PumpClinic…….Issue 39 Feb 2010


Lubricated Bath Design

This design uses two sliding shoes on the rotor to occlude the hose. The rotor and hose operate in a lubricated bath that reduces friction on the hose and provides the long life typical of this type of hose pump design. Benefits of Peristaltic Pumps

• Available in 15 different sizes that provide flow rates of 0 to 155 m3/h

• Will handle fragile fluids with little wear

• Since only the inside of the hose and hose inserts are in contact with the fluid, they can be processed without the devastating damage experienced by other pumping technologies.

• Hoses are constructed of natural rubber, NBR, EPDM, Hypalon, FKM

• Provide high levels of volumetric accuracy for sampling and metering applications

• Ability to pass solids in the material being pumped, 20% of hose ID

• Dry-run capability allows tank and line stripping.

• Seal-less design eliminates leaks, contamination and wear problems associated with difficult to seal products.

• Self-priming up to 9,8 meters at sea level on water

• Reversible operation allows pumping in both directions

• Durable construction of ductile iron and steel construction allows higher discharge pressures up to 15 bars.

• Low maintenance requirements of the hose and shoes

Description of Hose versus Tube Pumps

Hose Pumps

Higher pressure peristaltic hose pumps which can operate at pressures up to 15 bars, typically use shoes and have casings filled with lubricant to prevent abrasion of the exterior of the pump tube and to aid in the dissipation of heat, and use reinforced tubes, often called "hoses". This class of pump is often called a "hose pump".

The hoses in a hose pump are typically reinforced, with a very thick wall. For a given ID the hoses have much bigger OD than tubing for the roller pump. This thicker wall, combined with a stiffer material typically used in the hoses make the forces necessary to occlude the hose much greater than for the tubing. This results in a bigger pump and motor for a given flow rate with the hose pump than the roller pump resulting in more energy to run.

Lubricated-bath pump design


The biggest advantage with the hose pumps over the roller pumps is the high operating pressure of up to 15 bars. If the high operating pressure is not required, a tubing pump is a better option than a hose pump. With recent advances made in the tubing technology for pressure, life and chemical compatibility, as well as the higher flow rate ranges, the advantages that hose pumps had over roller pumps are diminishing.

Tube pumps

Lower pressure peristaltic pumps, typically have dry casings and use rollers and use non-reinforced, extruded tubing. This class of pump is sometimes called a "tube pump" or "tubing pump".

These pumps employ rollers to squeeze the tube. These pumps have a minimum of 2 rollers 180 degrees apart, and may have as many as 8, or even 12 rollers. Increasing the number of rollers increases the frequency of the pumped fluid at the outlet, thereby decreasing the amplitude of pulsing. The downside to increasing number of rollers it that it proportionately increases the number of squeezes, or occlusions, on the tubing for a given cumulative flow through that tube, thereby reducing the tubing life.

There are two kinds of roller design in peristaltic pumps:

1. Fixed occlusion - the rollers have a fixed locus as it turns, keeping the occlusion constant as it squeezes the tube. This is a simple, yet effective design. The only downside to this design is that the occlusion as a percent on the tube varies with the variation of the tube wall thickness. Typically, the wall thickness of the extruded tubes vary enough that the % occlusion can vary with the wall thickness (see above). Therefore, tubing at the high end of the wall thickness, but within the accepted tolerance, will have higher % occlusion, increasing the wear on it, thereby decreasing the tube life. The tube wall thickness tolerances today are kept pretty tight so that this is not much of a practical issue.

2. Spring loaded rollers - As the name indicates, the rollers are mounted on a spring. This design is a bit more elaborate than the fixed occlusion, but helps overcome the variations in the tube wall thickness over a broader range. Irrespective of the variations, the roller imparts the same amount of stress on the tubing that is proportional to the spring constant, making this a constant stress operation. The spring is selected to overcome not only the hoop strength of the tubing, but also the pressure of the pumped fluid.

The operating pressure of these pumps is determined by the tubing, and the motor's ability to overcome the hoop strength of the tubing and the pressure.

Key Design Parameters

The key design parameter in hose pumps is hose life. Surveys have shown that 95% of spares value used on peristaltic pumps is on replacement hoses.

The major issues that affect hose life are as follows

Occlusion/Shimming Philosophy

The minimum gap between the roller and the housing determines the maximum squeeze applied on the tubing. The amount of squeeze applied to the tubing affects pumping performance and the tube life - more squeezing decreases the tubing life dramatically, while less squeezing decreases the pumping efficiency, especially in high pressure pumping. Therefore, this amount of squeeze becomes an important design parameter.


The term "occlusion" is used to measure the amount of squeeze. It is either expressed as a percentage of twice the wall thickness, or as an absolute amount of the wall that is squeezed. The amount of squeeze is influenced by the amount of shimming underneath the shoes. The means the greater the shimming depth, the greater the squeeze and the lower the hose life. As discharge pressure increases, additional shimming is required so the squeeze increases.

The occlusion is typically 10 to 20%, with a higher occlusion for a softer tube material and a lower occlusion for a harder tube material. Rubber composition therefore becomes important so that efficiency is maintained without “over squeezing” the hose and affecting hose life.

Thus for a given pump, the most critical issues are rubber composition, the amount of shimming and wall thickness. An interesting point here is that the inside diameter of the tubing is not an important design parameter for the suitability of the tubing for the pump. Therefore, it is common for more than one ID be used with a pump, as long as the wall thickness remains the same.

Mechanical Capability of Hose Material Mechanical capability is simply a function of the hose material and the number of times the hose is squeezed. The factors that influence the number of squeezes is the speed of rotation and the number of shoes or rollers. Therefore pump speed is a significant factor in hose life.

The hose needs to be elastomeric to maintain the circular cross section after millions of cycles of squeezing in the pump. This requirement eliminates a variety of non-elastomeric polymers such as PTFE, PVDF etc. from consideration as material for pump tubing.

Different hose materials have different mechanical flex life. The most common hose material is natural rubber. Natural rubber has double the mechanical life of EPDM and three times the life of Buna/Nitrile and Hypalon.

Fluoroelastomers have been used but have very poor mechanical life that tends to make this material impractical. Some manufacturers are currently testing various compositions with the aim of improving fatigue life.

Chemical Compatibility

The pumped fluid contacts only the inside surface of the tubing. There are no other valves, O-rings, seals or packings to worry about in a peristaltic pump. Therefore, the only compatibility to worry about in a peristaltic pump is the hoses for the fluid being pumped.

The most popular hose materials (as distinct from tube) are: • natural rubber • EPDM (ethylene propylene diene monomer) • Buna also known as Nitrile • Hypalon • Viton flouroelatomer

Hoses are made in different ways by manufacturers. Some use full width in the particular material. Others use a common outside layer of natural rubber with an inside layer of other material selected for superior chemical compatibility. Some problems have been encountered with dual material hoses with the integrity being compromised due to pin holes in the inside layer and delamination between the two layers.

Typical elastomers for pump tubing (as distinct from hose) are silicone, PVC, EPDM+ polypropylene (as in santoprene), polyurethane, neoprene and a number of proprietary materials. Extruded fluoroplymer tubes such as FKM (viton, fluorel, etc.) have good compatibility with acids, hydrocarbons, and petroleum fuels. However, the material has poor fatigue resistance to get meaningful tube life that can be practical.

There are a couple of newer pump tubing developments that offer a broad chemical compatibility - a lined tubing approach and the use of fluoroelastomer approach.


With the lined tubing, the thin inside liner is made of a chemically resistant material such as poly-olefin and PTFE that form a barrier for the rest of the tubing wall from coming in contact with the pumped fluid. These liners are materials that are not elastomeric therefore the entire tube wall cannot be made with this material for peristaltic pump applications. These tubing provide adequate chemical compatibility and life to use them in chemically challenging applications. There are a few things to keep in mind when using these tubes - any pin holes in the liner during manufacturing could render the tubing vulnerable to chemical attack. In the case of stiff plastic liners like the polyolefins, with repeated flexing in the peristaltic pump they can develop cracks, rendering the bulk material again vulnerable to chemical attack. A common issue with all lined tubing is delamination of the liner with repeated flexing that signal the end of that tube life. For those with need for chemically compatible tubing, these lined tubing offer a good solution.

There are many online sites for checking the chemical compatibility of the tubing material with the pumped fluid. The manufacturers of these tubing may also have compatibility charts specific to their tubing.

While these charts cover a list of commonly encountered fluids, they may not have all the fluids. If there is a fluid whose compatibility is not listed anywhere, then a common test of compatibility is the immersion testing. A 1 to 2 inch sample of the tubing is immersed in the fluid to be pumped for anywhere from 24 to 48 hours and the amount of weight change from before and after the immersion is measured. If the weight change is greater than 10% of the initial weight, then that tube is not compatible with the fluid, and should not be used in that application. This test is still a one way test, in the sense that there is still a remote chance that the tubing that passes this test can still be incompatible for the application since the combination of borderline compatibility and mechanical flexing can push the tube over the edge, resulting in premature tube failure.

NOTE. The natural starting point for selection of hose materials is chemical compatibility. This happens particularly with pump selection software linked to chemical compatibility charts. This can lead to serious errors as mechanical life is just as important If not more important than chemical compatibility. The correct selection philosophy is to consider both chemical and mechanical implications.

There have been many examples of where natural rubber, because if its superior mechanical life, has provided the best hose life notwithstanding chemical compatibility charts giving natural rubber a “poor” rating.

The best selection process is to start with natural rubber and then eliminate it if deemed to be totally unfit chemically. Natural rubber is also the cheapest material so it is often worth trialling and then eliminating it.

Liquid Temperature

The implications of temperature must be considered. The normal rated temperature for hose pumps is 40°C. Normal maximum temperature is 80°C subject to adjustments to either maximum pump speed or maximum discharge pressure. Manufacturer’s selection charts would normally show derating requirements at higher temperatures. If not, contact the manufacturer.

Discharge Pressure As discharge pressure increases, additional shimming is required to maintain pump efficiency. This decreases hose life. As discharge pressure increases, the maximum allowable pump speed and therefore flow rate decreases to compensate for the reduction in hose life due to additional “squeeze”. Care should be taken when moving pumps between applications to ensure that those design parameters are not exceeded. Possible Results of Hose Failure If a hose ruptures, the following may occur dependent on where the rupture occurs:

• the pump housing fills with the product • product drains from the suction line and suction tank into the pump housing and then leaks from

the pump to the floor. • product drains from the discharge line and discharge tank into the pump housing and then leaks

from the pump to the floor.


Manufacturers do not provide any guarantee on hose life, product loss or consequential damage due to ruptured hoses. It is the customer’s responsibility to prevent pumped liquid loss or any other loss with additional hose rupture detectors and/or non return valves and automatic shut down valves. Various rupture detection devices are available from manufacturers and these include:

• float type magnetic reed switch which detects a change in liquid level in the hose housing • conductivity probes however these are limited to use with products that are conductive • pressure transmitters fitted to the hose housing which detect a change in liquid level.

The best method of mitigating these losses is by preventative maintenance. A hose replacement regime needs to be established so that hoses are changed prior to rupture occurring. This can be instigated after hose life is determined. This does not mean that rupture detection systems can be forgotten.

Recognising Cause of Hose Failure

Hose Destruction Areas

Over Shimming

• This failure occurs in the cheek of the hose as the hose starts to re-vulcanise • Rubber is built up under the shoe, leading to severe internal friction and a heat build up, resulting in

re-vulcanisation • The failure will occur over the rotor shoe’s full contact path

Normal end of life Over shimming

Chemical attack Under shimming

Over shimming


Chemical Compatibility • The inner rubber has softened due to chemical action • Parts of the rubber breakaway when they are pulled and may stick together

Under Shimming or Pulsation Damage

• Damage occurs where the shoe leaves the hose • Failure is due to backflow of an abrasive product

from severe pulsation • Remedy - remove the source of pulsation but do not

increase number of shims • Shows as very high discharge side impulse loss

Pump Clinic 40 Glossary of Pump Terms Page 1 of 10

Glossary of Pump Terms

A

Absorbed Power: The actual amount of power being consumed by the pump at a specific flow and head.

Adapter (or Support Head): Connects and aligns the power (bearing) end of an ANSI pump to the wet end.

A.N.S.I. B73.1 Standard: American National Standards Institute. A set of specifications (envelope dimensions) for centrifugal pumps.

Absolute pressure (normally in kPaA) : Atmospheric pressure added to gauge pressure.

Affinity laws: They are used to determine changes in capacity, head and power due to changes in shaft speed. The same laws can be used to estimate changes in capacity, head and power due to changes in impeller diameter.

Air ingestion: Air is coming into the stuffing box because of a negative suction pressure.

Alignment: The centreline of the pump is perfectly aligned with the centreline of the driver (usually an electric motor).

Ambient temperature or pressure: The environmental temperature or pressure in the area where the equipment is located.

Atmospheric pressure: At sea level, atmospheric pressure 101.3 kPaA.

API Standards: Standards produced by the American Petroleum Institute for various pieces of equipment applied in the oil and gas industries. Various API codes are applicable for pumps dependent on the type of pump. Examples are API610, API685 etc.

Axial Thrust: The resultant of all axial forces (i.e. in direction of the pump shaft) acting in the pump rotor.

Axial Thrust Balancing: Methods by which the axial thrust is balanced to minimize bearing loads.

B

Back to Back Impellers: In multistage pumps, some impellers are located on the shaft facing opposite directions. This is one method of achieving axial thrust balancing.

Back plate: Used in some centrifugal pumps to position the stuffing box and provide an impeller wear surface.

Back Vane: A radial narrow vane located on the back shroud of an impeller and is designed to balance axial thrust.

Back pull out pump: A design that allows the wet end of the pump to be left on the piping yet allows the power end and adapter to be removed. A.N.S.I. pumps are designed this way.

Back to back double seal: Two mechanical seals located in a pump with the rotating seal faces in opposite directions.

PumpClinic…….Issue 40 April 2010


Balance Holes: Holes in the back shroud of an impeller designed to balance axial thrust.

Balanced seal: A design in which the seal face closing area is reduced to lower the closing force, and reduce the heat generation between the faces.

Ball bearing: Consists of an inner race, an outer race, and a series of balls between them. Often called a precision or anti friction bearing.

Bar: Metric unit for pressure and is equal to 100 kPa.

Barrier fluid: The high-pressure fluid that is circulated between the two seals in a double mechanical seal. The fluid should enter the bottom and leave the top to prevent air pockets.

Base plate: The pump and motor are mounted on this unit.

Bearing: Supports the rotating shaft and allows it to turn with a minimum amount of friction. Could be either sleeve or anti-friction type.

Bellows: A component of a mechanical seal that can be manufactured from metal or non-metallic materials to eliminate flexing, rolling or sliding.

Belt Drive: A combination of belts and pulleys that transfers torque from the driver to the pump.

B.E.P (Best Efficiency Point): The best efficiency point on a pump curve for a specific impeller diameter. It is the point where the power coming out of the pump (water power) is the closest to the power used by the pump (absorbed power) from the driver. This is also the point where there is no radial deflection of the shaft cause by unequal hydraulic forces acting on the impeller.

Buffer fluid: The low pressure fluid that is circulated between the two mechanical seal in a tandem seal.

Buna N: Sometimes called Nitrite. A common elastomer used in the sealing of oil or water.

Bushing: A close fitting support device used to restrict flow between two liquids, thermally isolate a hot liquid, support the rotating shaft or break down pressure.

Bypass line: Used to either re-circulate fluid from the pump discharge to the stuffing box, the stuffing box to the pump suction, or the pump discharge to a lower pressure point in the system. This can also be used as a manual means of flow control.

C

Canned pump: A seal/less pump with the shaft, bearings and rotor contained in a can to prevent product leakage. These are generally limited to pumping clean lubricating liquids.

Capacity: Volumetric flow of liquid measured in, litres/sec, m3/hr etc.

Cartridge seal: A self-contained assembly containing the seal, gland, sleeve, and both stationary and rotating seal faces. Usually needs no installation measurement. Must be used in a pump with impeller clearance adjustments are made.

Cavitation: Cavities or bubbles form in the fluid low-pressure area and collapse in a higher-pressure area of the pump, causing noise, damage and a loss of capacity.

Centreline design: The pump is mounted to the base plate by feet attached to the sides of the volute instead of the bottom. These are used in higher temperature pumping applications e.g. API applications and allows thermal expansion to occur in without the necessity of re alignment.

Centipoise: One unit for dynamic viscosity of a liquid.

Centistoke: One unit for the kinematic viscosity of a liquid. Dynamic viscosity divided by the liquid specific gravity at the same temperature gives kinematic viscosity.

Centrifugal pump: A pump that imparts energy to a liquid with centrifugal force.


Centrifugal separator (or cyclone separator): Equipment that utilises centrifugal force to separate solids out of the fluid. In pumping, they are used in seal flush lines.

Change of state: This defines a change in phase for any material e.g. liquid to vapour, liquid to solid, solid to vapour.

Close-coupled: A close-coupled pump is characterised by a common or rigidly-coupled motor and pump shaft. These pumps do not have a flexible coupling. There are no pump bearings and all thrust is carried by the motor bearings.

Concentricity: When the parts share the same centerline they are concentric to each other.

Condensate: This defines the change on phase of a vapour to liquid e.g. steam to water.

Constant Level Oiler: This is used to maintain the oil level in a bearing housing to the correct level as oil is used.

Cooling jacket: Cooling jackets can be located on the pump casings and /or the stuffing box of the pump to control the temperature of the fluid. Cooling jackets can be part of the component casting or separate bolt on items.

Corrosion: This is a chemical or electrochemical reaction on material surfaces that changes the profile and/or composition of the surface.

Corrosion resistant: An arbitrary term that indicates a corrosion rate of less than 0.05 mm per year.

Coupling: This is used to connect the pump to the driver. It transmits torque between the driver and pump. It can be a flexible coupling that allows some axial and radial misalignment or rigid which does not allow for any misalignment. Allowable misalignment is determined by manufacturers of the coupling.

Critical speed: Any object made of an elastic material has a natural period of vibration. When a pump rotor or shaft rotates at any speed corresponding to its natural frequency, minor unbalances will be magnified. These speeds are called the critical speeds.

Cryogenic Pumping: This generally refers to pumping liquid gases at very cold temperatures.

Cutwater: A part of the pump casing that directs the pumped liquid to the pump discharge.

D

D.I.N. standard: This is a German standard that defines various industrial products.

Deflection: Movement or displacement of the shaft in a radial direction.

Density: Measured in kg/m3 or lb/in 2. This is the ratio of the mass of a liquid in a given volume to the magnitude of that volume. Refer to Specific Gravity.

Design Duty Point: This is generally defined for a pump as a capacity at a head or pressure of the liquid being pumped, ideally the design duty point on a centrifugal pump as at BEP.

Dilatant liquid: A liquid whose viscosity increases with increased shear rate e.g. agitation.

Differential Pressure: The difference between the outlet pressure and the inlet pressure. Differential pressure is sometimes called the Pump Total Differential Pressure.

Discharge Head: The outlet pressure of the pump converted to head of liquid.

Double seal: An outdated term describing two seals in a pump. The latest terminology is "dual seals", back-to-back double seals or tandem seal. In the past the term was used to describe a higher-pressure barrier fluid between dual seals.

Double suction pump: A pump with an impeller where liquid enters the impeller on both sides. The rotor is generally (but not always) suspended between two bearings. These pumps are generally of higher capacities.


Double volute: A centrifugal pump design that incorporates two cut waters to decrease radial loads and minimise shaft deflection when the pump is operating away from the B.E.P. Lowers the efficiency of the pump and therefore seldom used on smaller size impellers.

Dry running: Operating a pump without liquid in the pump and therefore in the seal area.

Dual Seal: Two seals running in various configurations: back to back, tandem, face to face, or concentric.

Ductility: The property of a metal that allows a great deal of mechanical deformation without cracking.

E

Efficiency: In centrifugal pumps, this is the useful power in the liquid to the expended power. That is power out of the pump divided by power into the pump.

Elastomer: A rubber-like material that, when compressed and then released will basically return to its original shape in less than five seconds.

Electrolysis: A process involving chemical change caused by the passage of an electric current through a liquid.

E.P.D.M or E.P.R: Ethylene propylene rubber. This is a common elastomer used in the sealing of water based and higher pH materials. Cannot be used in petroleum products.

Erosion: This is wear caused by mechanical action of the liquid on the surface of the materials. It is obviously more prevalent if the liquid contains solids.

Eye of the impeller: The center of the impeller where the fluid enters.

F

Face combination: The materials chosen for the two lapped seal faces. An example is carbon graphite running on silicon carbide.

Face-to-face seals: Two seals running against a common seal face. The barrier fluid pressure is always lower than stuffing box pressure.

Face lubrication: The fluid or vapor that exists between lapped mechanical seal faces.

Face pressure: The sum of all the loads on the seal face including the spring load, hydraulic load and shaft axial thrust, divided by the area of the seal face. This face load is reduced by friction between the sliding elastomer and the shaft or sleeve.

Flashing: A rapid change in phase from a liquid to a gas.

Flooded suction: This refers to a situation where the liquid level on the suction side of a pump is higher than the pump centerline and the liquid flows to the pump by gravity.

Fluorocarbon: This is a genetic term for an elastomer of which Viton (a Dupont product) is a typical example. The material has good compatibility with hydrocarbons, has high temperature capability but poor mechanical life.

Flush: Putting an outside liquid into the stuffing box of the pump at a pressure higher than stuffing box pressure. All of this liquid mixes with and dilutes the pumped fluid.

Foot Valve: A type of check valve with a built-in strainer. Used at the point of the liquid intake to retain liquid in the system, preventing the loss of prime when the liquid level is below the pump centreline.

Free length: The uncompressed axial length of a seal.

Friction Head: This is the head loss due to friction as liquid flows in pipes and fittings.


G

Gasket: This is used between two static surfaces to provide a seal. Made from a variety of deformable materials.

Gland: The part that holds one half of the mechanical seal and attaches to the stuffing box.

H

Hard face: A seal face either rotating or stationary. The most common materials are silicone carbide, ceramic, tungsten carbide, Stellite, Ni-resist. The hard face must be the wider seal face.

Hastelloy "C” (also known as Alloy C): A nickel-rich, corrosion-resistant and very hard alloy.

Head: The equivalent height of the liquid that will produce a particular pressure. Can be calculated from H (metres)= pressure in kPa/(9.8 x specific gravity).

Horizontally Split Pump: This is a pump where the casing is split into two sections in the axial plane. This means there is a top and bottom-half casing. Connections are normally on the bottom-half casing to allow removal of the top casing for pump inspection without needing to disconnect pipework.

Hydraulic balance: A method of reducing mechanical seal face loading by reducing the seal face closing area.

Hydrocarbon: A petroleum product consisting of hydrogen and carbon.

I

I.D.: Inside diameter.

I.S.O.: International Standards Organization. Sets pump and seal standards for the metric community.

Impeller: The rotating component of a centrifugal pump that imparts energy to the fluid being pumped. Available in open, semi-open and closed design.

Impeller eye: The centre of the impeller or the point where fluid enters the impeller.

Impeller shroud: The plates located on one or both sides of the impeller vanes. Prevents solids from penetrating behind the vanes.

Impeller vane: Located between the eye and the discharge side of the impeller. Directs the flow of the liquid to the outside diameter of the impeller.

Inducer: A small axial flow vane that attaches to the impeller of a centrifugal pump that reduces the N.P.S.H. required by a pump. This improvement occurs across a narrow capacity range and the impact can be detrimental outside of this range.

Induction motor: The most common type used in industry. Has a slippage of 2 to 5 percent compared to synchronous motors.

Inline pump: Mounted in the piping generally between two flanges. No base plate or alignment required.

J

Jacket: Usually refers to the heating/cooling jacket surrounding the stuffing box on some pumps.

K

Kalrez®: An "elastomer-like" material manufactured by E.I. Dupont that is used to seal most solvents and other aggressive fluids. It is available in several different grades.

Kilowatt: One thousand watts. The normal unit for power in the metric system.


Kpa (kilopascals): A metric unit of pressure.

L

Labyrinth seal: A non-contacting seal utilising a tortured path for the escape of the fluid. Utilises a series of pressure drops to reduce the leakage.

Lantern Ring: A device used to supply lubricant to packing. Usually located in the middle of the packing ring set.

Line bearings: These position the rotor or shaft radially and are normally of the sleeve type. Generally used in vertical pumps.

Liquefied Natural Gas LNG: This is liquefied gas from natural sources

Liquefied Petroleum Gas LPG: This is liquefied petroleum gas which is a by-product of the refining of crude petroleum oil.

M

Magnetic drive: A type of seal less pump that utilises permanent magnet technology to provide the rotation of the impeller.

Mating ring: Another name for the hard face in a mechanical seal. It can be either rotating or stationary.

Mechanical seal: A positive sealing device used to seal all fluids (liquids and gases). The primary seal is a set of lapped seal faces that are installed perpendicular to the shaft.

Metal bellows: Used in mechanical seal designs to eliminate the need for a dynamic elastomer and springs.

Metal fatigue: A breakage of the metal caused by the bending and flexing of a metal part beyond its endurance limit.

Minimum flow: The minimum capacity of a pump to prevent thermal and/or mechanical damage.

Moment of inertia: This represents a magnitude of the inertia in respect of the rotation around the axis of the pump and drive rotor.

Multistage Pump: This defines a pump that has more than one impeller on the shaft.

N

Negative pressure: A pressure below atmospheric pressure.

Newtonian Fluid: A Newtonian liquid is one whose viscosity does not change with increasing shear rate e.g. when agitated.

Non Overloading Power: This refers to the maximum power absorbed by a pump with a specific impeller diameter and liquid. Motors are generally sized at the next size above this power.

N.P.S.H.A: The net positive suction head available to prevent cavitation of the pump. It refers to the suction side of a pump installation and is defined as the head acting on top of the liquid + static head -vapor pressure head - friction head loss in the suction piping.

N.P.S.H.R.: Net positive suction head required to prevent cavitation of a pump and is dependent on impeller and pump design. The pump manufacturer determines the NPSHR by testing. In all cases, it is imperative that NPSHA >NPSHR to prevent cavitation

O

O.D.: Outside diameter.


Operating length: This measurement is set by manufacturers to provide the correct closing pressure on the two mechanical seal faces. The measurement can be made in a number of ways dependent on manufacturer. One measurement is from the face of the stationary face to the location screws for the rotating part of the seal.

Orifice Plate: A plate with a hole smaller than the pipe diameter in which it is located

Over hung impeller: Not supported with bearings on either side of the impeller.

P

Packing: The soft rings that a mechanical seal replaces to stop leakage. Packing must have a small leak because it works on the theory of a series of pressure drops to reduce the stuffing box pressure to the point where the leakage is acceptable. Generally, a minimum of five rings of packing is required to do this.

Parallel operation: This refers to two or more pumps that are discharging to a common header. It is important that the impeller speed and outside diameters are the same or one of the pumps may cause other pumps to operate at shut off.

pH: A measure of the acidity or the alkalinity of a fluid. The scale ranges from 0 (acid) to 14 (alkali) with 7 considered neutral.

Pipe strain: The strain on the pump volute caused by the piping. It will cause excessive mechanical seal movement and can cause contact between rotating and stationary pump and seal components. It can also cause serious misalignment with resultant damage to bearings and couplings.

Pitting: Surface voids caused by corrosion, erosion or cavitation. It is possible for the three to occur at the same time.

Positive Displacement Pumps: This is a collective definition of all pumps that operate according to the positive displacement principle. That is, the liquid being pumped is displaced by a body which periodically increases and decreases the working volume.

Power end: The end of the pump that attaches to the power source and is not wetted by the liquid. The bearings are in this part.

Precision bearing: Ball or roller bearing as opposed to a sleeve bearing.

Pressure gradient: The pressure drop between the seal faces.

Priming: This refers to the filling of a pump with liquid prior to operation.

Q

Quench: The introduction of a fluid outside the seal to cool the product, dilute any leakage across the seal faces or isolate seal faces from atmosphere.

R

Radial Bearing: This bearing handles most of the radial loads put on the impeller. In an end-suction centrifugal pump it is the bearing located closest to the stuffing box.

Radial Thrust: This is the thrust produced in the radial direction i.e. at 90 degrees to the centerline of the shaft, by forces acting on the impeller when operating at points other than BEP.

Radially Split Casing: A pump casing with the casing joint at 90 degrees to the shaft axis.

Rated Operation: This is the basis of selection of both pump and driver. When rated operation is specified, it generally exceeds the requirement of the design operation.

Ring Section Pumps: These are multistage pumps with several identical stage casings arranged in tandem behind each other. The stage casings are radially split.


Rockwell "C": The scale used to define hardness of materials.

Run out: Twice the distance that the centre of the shaft is displaced from the axis of rotation.

S

Saturation Pressure: Same as vapour pressure.

Seal faces: The lapped faces that provide the primary seal in a mechanical seal.

Self Priming Pump: These pumps are capable of pumping liquids on a suction lift by being able to remove air from the suction line of its own accord. Most self-priming pumps need to be filled with liquid (primed) before the initial start, however their design means that subsequent starts do not require priming. Pumps that can self-prime from dry are limited to diaphragm and peristaltic type pumps only (both are positive displacement pumps).

Series operation: Two or more pumps connected with the discharge of the first pump discharging to the suction of the other etc. Rarely used these days.

Shaft packing: The soft packing located in the stuffing box to provide a shaft seal for pumps.

Shaft Power: The mechanical power absorbed at the pump shaft.

Shut off head: This is the maximum head that the pump can generate with a given impeller outside diameter. It is normally at zero capacity.

Sleeve bearing: A non-precision or anti-friction bearing. It is usually manufactured from carbon, teflon, brass, white metal, other synthetic bearing materials.

Slurry: A slurry is a liquid in which solids are present in suspension.

Solubility: This defines the ability of a liquid to dissolve with another liquid. For example, ethanol will fully dissolve in water whereas oil will float on water.

Specific Gravity (SG): This is the ratio of the mass of a liquid for a known volume to the weight of water for the same volume. The reference is water at 4 degrees C with an SG of 1.0. If the liquid you are questioning will float on water the specific gravity is less than one. If it sinks, it is higher than one. Note that this is based on the liquid not being soluble in water.

Specific Speed: Specific speed of a pump is determined by the geometry of a pump impeller. The higher the specific speed the less N.P.S.H. required.

Stainless steel: Alloy steels containing a high percentage of chromium and/or nickel.

Static head: The height of a liquid above a reference point e.g. pump centerline.

Stationary face: The seal face that does not rotate with the shaft.

Stuffing box: The portion of the pump that held the packing and now holds the mechanical seal.

Stuffing box pressure: The pressure in the stuffing box and generally between suction and discharge pressure but closer to suction pressure.

Submersible pump: A pump/motor pumpset that operates only when totally submersed in the fluid which is being pumped.

Suction lift: Pumping application where the liquid level on the suction side of the pump is below the pump centerline.

System resistance Curve: A graphed representation of how total dynamic head varies with capacity. A pump will operate where the system resistance curve intersects the pump performance curve.

System head: The head caused by friction in the piping valves and fittings.


T

Tandem seals: The seals are facing in the same direction with a low-pressure barrier fluid circulating between them.

Thermal conductivity: A measure of the material's ability to conduct heat. This is a very important factor in the selection of mechanical seal faces.

Thixotropic fluid: The viscosity of the fluid decreases with agitation. Non-drip paint is an example of such a fluid.

Throttling: This means closing of a valve on the discharge side of a pump to increase friction loss. This steepens the system resistance curve with a resultant decrease in flowrate. Opening the valve results in an increase in flowrate.

Thrust bearing: This locates the rotor or shaft axially and is designed to handle any excess axial thrust load. In an end-suction pump, it is normally located close to the coupling.

Thrust: In a centrifugal pump it refers to the axial movement of the shaft. The thrust can be towards the wet or power end of the pump and at start up it thrusts in both directions.

Total Discharge Head: This is equal to the pressure at the pump discharge connection converted to head of liquid.

Total Dynamic Head: Total dynamic head is equal to total discharge head minus total suction head

Total Suction Head: This is equal to the pressure at the pump suction connection converted to head of liquid.

Tungsten carbide: A common hard face seal material available in several grades depending upon hardness and corrosion resistance. Cobalt and nickel are the two most common types.

Turbulence: This refers to disturbance of fluid as it enters the suction connection and /or the impeller. This can cause cavitation problems in a centrifugal pump. This is often caused by an insufficient length of straight pipe before the pump suction inlet.

U

Unbalanced seal: A mechanical seal not designed to balance the closing force between seals. Refer to Balanced Seal.

V

Variable speed motor: This is used to control flow in a system by varying the frequency of the motor. A better system than throttling as it reduces power consumption significantly.

Vacuum: This is a pressure less than atmospheric.

Vapor pressure: Below this pressure, the liquid at this temperature will vaporise.

Vaporisation: The fluid passes from a liquid to a gaseous state. If this happens at the seal faces the seal faces will be blown open.

Velocity: A measurement of the speed of the liquid in the system. This is measured in metres per second.

Velocity head: This is part of the total head calculation. This is calculated from the formula H = v2/2g.

Vent: This removes air or gas from the system. It is important to vent the stuffing box in vertical pumps to prevent the seal faces from running dry.


Viscosity: This is defined as the property of a liquid that offers resistance due to the existence of internal friction between layers within the liquid.

Viton®: An E.I. Dupont Dow manufactured fluorocarbon elastomer widely used in the sealing industry. Refer to Fluorocarbon.

Volute casing: This derives its name from a spiral-shaped casing surrounding the pump impeller. It converts velocity energy to pressure energy.

Vortex Pump: A type of pump used to handle liquids with entrained solids, particularly stringy solids. The impeller is recessed into the volute. A very low efficiency design, but practical in many applications.

Vortexing liquid: Creating a "whirlpool affect" that can draw air into the suction of the pump. Vortecies can form both from the surface of the liquid and in vertical pumps, from the floor of a pit or channel in which they are located.

W

Water hammer: This occurs in a closed piping system as a result of the pressure being rapidly increased when the liquid velocity is suddenly changed. This damaging effect is usually the result of sudden starting, stopping, change in pump speed, or the sudden opening or closing of a valve. Water hammer can usually be controlled by regulating the valve closure time, surge chambers, relief valves or other means.

Water Power: The calculated power coming on water at an efficiency of 100%.

Watt: A measure of power.

Wear ring: This is used with closed impeller pumps to seal leakage from the high-pressure side of the pump to the low-pressure side. This may need to be replaced as it wears when the recommended clearance is doubled or when reduction in pump performance can no longer be tolerated.

Welded metal bellows: A seal design used to eliminate the use of elastomers. Excellent for cryogenic and hot applications. Not as effective for hot petroleum applications because of "coking" problems.

Wet end: The part of the pump that gets wet from the pumping fluid. Includes the volute, stuffing box, impeller wear rings, and shaft or sleeve.

PumpClinic…….Issue 41 May 2010

When to use a Positive Displacement Pump

The choice between a centrifugal and positive displace-ment pump is not always apparent and cannot be made without a full understanding of the differences. The fun-damental difference is that positive displacement pumps, with the exception of air operated diaphragm pumps are basically constant volume machines where flowrates are independent of pressure. Flow is depend-ent only on speed.

The performance chart to the left illustrates this differ-ence. The centrifugal has varying flow depending on pressure or head, whereas the PD pump has more or less constant flow regardless of pressure.

Viscosity impacts on pump performance in very different ways. The graph on the left shows how centrifugal pump performance reduces markedly as viscosity in-creases. It also shows that with a positive displace-ment pump, higher viscosity increases the flowrate. The reason for this is that slippage (flow back towards suction because of differential pressure) reduces with increased viscosity. The volumetric efficiency of posi-tive displacement pumps increases with increased vis-cosity.

m3/hr

met

res

20

40

60

80

20 40 60

The pumps behave very differently when considering mechanical efficiency as well.

By looking at the efficiency chart to the left you can see the impact of pressure changes on the pump’s effi-ciency. Changes in pressure have little effect on the PD pump but a dramatic one on the centrifugal.

Efficiency - Head

20 30 40

Metres of Head

Another consideration is NPSHR. In a centrifugal pump the NPSHR varies as a function of flow and flow is determined by the system resistance (total dynamic head). If the total head varies for any reason, the flow will change and NPSHR will also change as a consequence. In a PD pump NPSHR varies as a function of flow which is determined by speed. At a fixed speed the flow is constant irrespective of pressure and there-fore NPSHR is constant.

……..2

When comparing the two types of pumps, it is important to understand that a centrifugal pump has an ideal operating flowrate i.e. the flowrate at the best efficiency point. At flowrates other than best effi-ciency flowrate, other factors need to be considered. Radial loads (the load that applies a bending moment to the shaft) increases at all flowrates other than that at best efficiency.

With a PD pump you can operate the pump on any point of the curve. In fact the volumetric efficiency actually improves at the high speed part of the curve. This is because the volumetric efficiency is af-fected by slip, which is essentially constant regardless of speed. At low speed, the percentage of slip in relation to volumetric displacement is higher than at high speed.

The data presented in these charts is the actual data for a specific application. The centrifugal was picked at its Best Efficiency Point (BEP) and the PD pump (Internal Gear) selected to match the flow, viscosity, and pressure. Different applications will have different curves and efficiencies. These curves are presented as an example of the type of performance behaviour between the two different principles.

The most obvious reason to use a PD pump is when you have a high viscosity application. It is com-mon knowledge that a centrifugal becomes very inefficient at even modest viscosity. The acceptable viscosity ranges for centrifugal pumps tends to be dependent on pump size. Published data is also variable in this area. We believe that acceptable viscosity limits for centrifugal pumps are as follows.

For nominal discharge pipe diameters;

≤ 50mm – up to 100 mm²/s

≤ 150mm – up to 250 mm²/s

> 150mm – up to 400 mm²/s

However, there are other reasons to select a PD pump over a centrifugal other than high viscosity. These can be summarised as follows:

A simple rule of thumb is that a PD pump should be selected where the smallest available centrifugal pump needs to operate at a flow less than 50% of best efficiency flow.

Additionally, PD pumps tend to produce higher heads or pressures at a more economical price.

PD pumps may be a more appropriate selection at low flow, high head applications.

A PD pump would be used on applications that have variable pressure conditions. The flowrate from centrifugal pumps will vary up and down the performance curve which can cause process problems. A PD pump will give near constant flow that makes it possible to match the flow to the process re-quirements. The desire to have constant flow is a reason that a PD pump is normally applied for me-tering applications.

Generally speaking centrifugal pumps tend to shear liquids more as speed is increased and the cen-trifugal is a high speed pump. This makes the PD pump better able to handle shear sensitive liquids. Shear rates in PD pumps vary by design but they are generally low shear devices, especially at low speeds. It is important to contact the manufacturer for specific information on shear rates and appli-cation recommendations.

By their nature, many PD pumps are self priming i.e. will operate on suction lift without the necessity of a foot valve. This capability can vary from pump to pump so manufacturers’ recommendations must be followed.

Pump Clinic 41 When to use a Positive Displacement Pump Page 2 of 2

Pump Clinic 42 Pump Standards Page 1 of 7

Pump Standards

There are many pump standards used in the pump industry today. These standards may be national, international, industry specific, company specific or project specific. This paper attempts to cover the more commonly used pump specifications within Australia. Undoubtedly, there are other pump specifications being utilised and the writer would be pleased to hear about them. Other associated standards such as those defining test standards are also available however we have not attempted to cover these in this paper. This paper is broken up into the following sections: • Definitions: This area details the organisations that either prepare or certify pump specifications.

• The various standards are covered by application as follows: - Fire - Oil and Gas - Process - General Purpose

There are many national specifications e.g. Japanese, British, Australian etc, however the vast majority are based on ISO standards or reference other standards such as API. Project specific specifications would generally reference other standards. The following should be noted

1. Many pumps available on the market today do not comply with any documented standard. This does not mean that these pumps are not of excellent quality or fit for purpose.

2. If the intention of a buyer is to define compliance with a specific standard, ensure that it is

relevant to the application. For example, specifying API compliance for a building services application does not make any sense.

3. Any pump specifier should take care that defining compliance with a particular specification does

not disqualify other suitable pumps. For example, specifying ANSI B73.1 may disqualify offers of very suitable pumps complying with ISO 5199.

4. There have been a number of instances where compliance with more than one pump

specification is detailed. This may appear to be a safe process however this often causes confusion for a pump supplier as there may be contradictions between the two standards.

DEFINITIONS Standards Australia – AS Standards Australia is the nation’s peak non-government Standards organisation. It is charged by the Commonwealth Government to meet Australia’s need for contemporary, internationally aligned Standards and related services. It leads and promotes a respected and unbiased Standards development process ensuring all competing interests are heard, their points of view considered and consensus reached.

PumpClinic…….Issue 42 June 2010


International Organisation for Standardisation - ISO The International Organisation for Standardisation, widely known as ISO, is an international-standard setting body composed of representatives from various national standards organisations. Founded on 23 February 1947, the organisation promulgates worldwide proprietary industrial and commercial standards. It has its headquarters in Geneva, Switzerland. While ISO defines itself as a non-governmental organisation, its ability to set standards that often become law, either through treaties or national standards, makes it more powerful than most non-governmental organisations. In practice, ISO acts as a consortium with strong links to governments. ISO classifies pumps as Class I, II and III with Class I having the most stringent requirements. The selection of class is determined by the application and the intention is that it is agreed between purchaser and supplier. It is impossible to standardise the class of technical requirements, however, the criteria for class determination may include;

- reliability - required operating life - operating conditions - environmental conditions - local ambient conditions

It is possible that pumps built in accordance with Classes I, II and III may work beside one another in the same plant.

American Petroleum Institute - API The American Petroleum Institute, commonly referred to as API, is the main U.S trade association for the oil and natural gas industry, representing about 400 corporations involved in production, refinement, distribution, and many other aspects of the petroleum industry. The association’s chief functions on behalf of the industry include advocacy and negotiation with governmental, legal, and regulatory agencies; research into economic, toxicological, and environmental effects; establishment and certification of industry standards; and education outreach. API both funds and conducts research related to many aspects of the petroleum industry. American National Standards Institute- ANSI The American National Standards Institute or ANSI is a private non-profit organisation that oversees the development of voluntary consensus standards for products, services, processes, systems, and personnel in the United States. The organisation also coordinates U.S. standards with international standards so that American products can be used worldwide. For example, standards make sure that people who own cameras can find the film they need for that camera anywhere around the globe. ANSI accredits standards that are developed by representatives of standards developing organisations, government agencies, consumer groups, companies, and others. These standards ensure that the characteristics and performance of products are consistent, that people use the same definitions and terms, and that products are tested the same way. ANSI also accredits organisations that carry out product or personnel certification in accordance with requirements defined in international standards.

Hydraulics Institute- HI The Hydraulic Institute is a non-profit industry (trade) association established in 1917. HI and its members are dedicated to excellence in the engineering, manufacture, and application of pumping equipment. The Institute plays a leading role in the development of pump standards in North America and worldwide. HI standards are developed within guidelines established by the American National Standards Institute (ANSI). HI members work through a number of technical committees to develop draft standards. The Institute involves pump users and other interested parties to reach consensus on published standards. HI standards are developed to define pump products, installation, operation, performance, testing, and pump life and quality. National Fire Protection Association - NFPA The National Fire Protection Association (NFPA) is a U.S. organisation (albeit with some international members) charged with creating and maintaining minimum standards and requirements for fire prevention and suppression activities, training, and equipment, as well as other life-safety codes and standards. This includes everything from building codes to the personal protective equipment utilised by firefighters while extinguishing a blaze.


FM Global FM Global is a U.S.-based insurance company, with offices worldwide, that specialises in loss prevention services primarily to large corporations throughout the world in the Highly Protected Risk (HPR) property insurance market sector. "FM Global" is the communicative name of the company, whereas the legal name is "Factory Mutual Insurance Company". The company employs a non-traditional business model whereby risk and premiums are determined by engineering analysis as opposed to historically based actuarial calculations. This business approach is centered on the belief that property losses can be prevented or mitigated. FM Global engineering personnel regularly visit insured locations to evaluate hazards and recommend improvements to their property or work practices to reduce physical and financial risks if a loss occurs. FM Approvals certifies industrial and commercial products and services for companies worldwide. When a product or service meets the standards of FM Approvals, it is issued the "FM APPROVED" mark to signify it will perform as expected and support property loss prevention. Deutsches Institut für Normung (DIN) DIN, headquartered in Berlin, is the German national organisation for standardisation and is that country's ISO member body. There are currently around thirty thousand DIN Standards covering nearly every field of technology. One of the earliest, and probably the most well-known, is DIN 476 — the standard that introduced the A-series paper sizes in 1922 — adopted in 1975 as International Standard ISO 216.

It was founded in 1917 as the Normenausschuss der deutschen Industrie (NADI, "Standardisation Committee of German Industry"). In 1926, the NADI was renamed Deutscher Normenausschuss (DNA, "German Standardisation Committee"), to reflect the fact that the organisation now dealt with standardisation issues in many fields; viz., not just for industrial products. Since 1975, the DNA is known as 'DIN' and is recognised by the German government as the official national-standards body, representing German interests at the international and European levels.

STANDARDS AND CERTIFICATIONS FIRE AS2941 Fixed fire protection installations - Pumpset systems This Standard specifies requirements for pumpset systems for use with fixed fire protection installations such as sprinkler, hydrant, water spray, and hose reel systems. It covers water supplies and specific requirements for pumps, drivers, fire pump controllers, and auxiliary equipment. Requirements for installation and acceptance testing for electrical and compression-ignition drivers are also included. AS2419-2005 Fire hydrant installations This Standard sets out requirements for the design, installation, and commissioning of fire hydrant systems to protect properties. It applies to fire hydrant systems installed to protect buildings, structures, storage yards, marinas and associated moored vessels, wharves, and plant. This Standard also applies to street fire hydrants used in lieu of on-site fire hydrants or to supplement the coverage afforded by street fire hydrants. This standard has relevance as fire pump packages regularly include piping and valving. AS2118 Fire Sprinkler Systems AS2118.1-06 (Sprinkler) and AS2118.6-95 (Combined Sprinkler/Hydrants) define the general requirements for automatic sprinkler systems. This has relevance as fire pump packages regularly include piping and valving. NFPA20 Standard for the Installation of Stationary Fire Pumps for Fire Protection. This standard deals with the selection and installation of pumps supplying liquid for private fire protection. The scope of this document shall include liquid supplies; suction, discharge, and auxiliary equipment; power supplies, including power supply arrangements; electric drive and control; diesel engine drive and control; steam turbine drive and control; and acceptance tests and operation. FM Certification FM Approvals not only evaluates sprinklers and sprinkler system components for compliance with existing standards, but also work closely with manufacturers to evaluate new products and develop appropriate standards. FM provides certification for fire pump sets to be utilized in sprinkler systems. It approves both fire pump package components e.g. pump, driver, panel and also the package assembler.


OIL AND GAS API 610 Centrifugal Pumps for General Refinery Service This International Standard specifies requirements for single and multistage centrifugal pumps, including pumps running in reverse as hydraulic power recovery turbines, for use in petroleum, petrochemical and gas industry process services. This Standard is applicable to overhung pumps, between-bearings pumps and vertically-suspended pumps . This Standard is not applicable to seal-less pumps. API 685 Seal-less Centrifugal Pumps This standard covers the minimum requirements for seal-less centrifugal pumps for use in petroleum, heavy duty chemical and gas industry services. Single stage pumps of two classifications, magnetic drive pumps (MDP) and canned motor pumps (CMP), are covered by this standard. API 674 Positive Displacement Pumps - Reciprocating This standard covers the minimum requirements for reciprocating positive displacement pumps for use in service in the petroleum, chemical, and gas industries. Both direct-acting and power-frame types are included. API 675 Positive Displacement Pumps - Controlled Volume This standard covers the minimum requirements for controlled volume positive displacement pumps for use in service in the petroleum, chemical, and gas industries. Both packed-plunger and diaphragm types are included. Diaphragm pumps that use direct mechanical actuation are excluded. API 676 Positive Displacement Pumps - Rotary This standard covers the minimum requirements for rotary positive displacement process pumps and pump units for use in the petroleum, petrochemical, and gas industry services. ISO 13709 Centrifugal Pumps for Petroleum, Petrochemical and Natural Gas Industries This International Standard specifies requirements for centrifugal pumps, including pumps running in reverse as hydraulic power recovery turbines, for use in petroleum, petrochemical and gas industry process services. This International Standard is applicable to overhung pumps, between-bearings pumps and vertically suspended pumps. It is not applicable to seal/less pumps. PROCESS ISO 9905:1994 Technical specifications for centrifugal pumps -- Class I The technical requirements refer only to the pump unit. Includes design features concerned with installation, maintenance and safety of such pumps, including baseplate, coupling and auxiliary piping. The selection of the class to be used is made in accordance with the technical requirements for the application for which the pump is intended. The class chosen is to be agreed between purchaser and manufacturer/supplier. ISO 5199:2002 Technical Specifications for Centrifugal Pumps - Class II This International Standard specifies the requirements for Class II centrifugal pumps of single-stage, multistage, horizontal or vertical construction, with any drive and any installation for general application. Pumps used in the chemical process industries (e.g. those conforming to ISO 2858) are typical of those covered by this International Standard. This International Standard includes design features concerned with installation, maintenance and safety for these pumps including baseplate, couplings and auxiliary piping, but it does not specify any requirements for the driver other than those related to its rated power output. ISO 9908:1993 Technical specifications for centrifugal pumps -- Class III Covers class III requirements for centrifugal pumps of single stage, multistage, horizontal or vertical construction (coupled or close-coupled) with any drive and any installation for general application. Includes design features concerned with installation, maintenance and safety of such pumps including baseplate, coupling and auxiliary piping but excluding the driver, if it is not an integral part of the pump. ISO 14847:1999 Rotary positive displacement pumps -- Technical requirements This standard specifies the technical requirements, other than safety and testing, for rotary positive displacement pumps and rotary positive displacement pump units. This standard does not apply to rotary positive displacement pumps for fluid power applications.


ISO 15783:2002 Seal-less centrifugal pumps - Class II – Specification This International Standard specifies the requirements for seal-less centrifugal pumps that are driven with permanent magnet coupling (magnet drive pumps) or with canned motor, and which are mainly used in chemical processes, water treatment and petrochemical industries. Pumps will normally conform to recognised standard specifications (e.g. ISO 5199, explosion protection, electromagnetic compatibility), except where special requirements are specified herein. This International Standard includes design features concerned with installation, maintenance and operational safety of the pumps, and defines those items to be agreed upon between the purchaser and manufacturer/supplier. ISO 16330:2003 Reciprocating positive displacement pumps and pump units -- Technical requirements This International Standard specifies the technical requirements, other than safety and testing, for reciprocating positive displacement pumps and pump units. It applies to pumps which utilise reciprocating motion derived from crankshafts and camshafts and also direct-acting fluid driven pumps. It does not apply to reciprocating positive displacement pumps, nor pumping water, where the whole pump is lubricated with the liquid being pumped. ANSI/ASME B73.1 Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process This standard covers centrifugal pumps of horizontal, end suction single stage, centre-line discharge design. This Standard includes dimensional interchangeability requirements and certain design features to facilitate installation and maintenance. It is the intent of this Standard that pumps of the same standard dimension designation from all sources of supply shall be interchangeable with respect to mounting dimensions, size and location of suction and discharge nozzles, input shafts, baseplates, and foundation bolt holes. ANSI/ASME B73.2 Specifications for Vertical In-Line Centrifugal Pumps for Chemical Process This Standard covers motor-driven centrifugal pumps of vertical shaft, single stage design with suction and discharge nozzles in line. It includes dimensional interchangeability requirements and certain design features to facilitate installation and maintenance. It is the intent of this Standard that pumps of the same standard dimension designation, from all sources of supply, shall be interchangeable with respect to mounting dimensions and size and location of suction and discharge.

ASME B73.3 Specification for Seal-less Horizontal End Suction Metallic Centrifugal Pumps for Chemical Process This Standard covers seal-less centrifugal pumps of horizontal end suction single stage and centre-line discharge design. This Standard includes dimensional interchangeability requirements and certain design features to facilitate installation and maintenance. It is the intent of this Standard that pumps of the same standard dimensional designation from all sources of supply shall be interchangeable with respect to mounting dimensions, size, and location of suction and discharge nozzles, input shafts, baseplates, and foundation bolt holes

ASME B73.5M Specification for Thermoplastic and Thermoset Polymer Material Horizontal End Suction Centrifugal Pumps for Chemical Process (not often used in Australia) This Standard covers centrifugal pumps of horizontal, end suction single stage, centreline discharge design, which components are made of thermoplastic and thermo-set polymer materials either reinforced or non-reinforced. It includes dimensional interchangeability requirements and certain design features to facilitate installation and maintenance. It is the intent of this Standard that pumps of the same standard dimension designation from all sources of supply shall be interchangeable with respect to mounting dimensions, size, and location of suction and discharge nozzles, input shafts, baseplates, and foundation bolt holes. This Standard does not include lined or non polymer components. ANSI/HI 3.1-3.5 Rotary Pumps (not often used in Australia) This standard applies to industrial rotary positive displacement pumps. It includes: types and nomenclature; definitions; design and application, operation and maintenance. The updated standard also includes:

• Capability tables in both metric and US customary units providing comparisons of rate of flow, pressure, viscosity, solids and abrasive handling, reversible rotation, and power range for 11 different types of rotary pumps;

• A consolidated range chart providing a comparison of pressure and rate of flow in both metric and US customary units for 11 different rotary pump types;

• Detailed explanation of each rotary pump type including basic operation, design features, typical applications, and driver requirements;


• New and updated cross-sectional illustrations of popular rotary pump designs; • Revised and expanded list of definitions; • Listing of viscosities of common fluids; • Plot of efficiency vs. rate of flow for different pump types; • A completely new section on multiphase pumps for oil and gas applications.

ANSI/HI 4.1-4.6 Seal-less Rotary (not often used in Australia) This standard applies to industrial rotary positive displacement pumps. It includes: types and nomenclature, definitions; design and application; installation, operation and maintenance; and test. It does not include standards on magnetic drives for seal-less pumps. ANSI/HI 5.1-5.6 Seal-less Centrifugal Pump Standards (not often used in Australia) This standard is for seal/less centrifugal pumps that are driven by canned motors or magnetic couplings. It includes types and nomenclature; definitions; design and application; installation, operation/maintenance and testing. The testing section includes:

• Hermetic Integrity Test • Mechanical Integrity Test • Winding Integrity Test

ANSI/HI 6.1-6.5 Reciprocating Power Pump Standard (not often used in Australia) This Standard applies to industrial/commercial reciprocating power pumps. It includes:

• Types and Nomenclature • Definitions • Design and Application

- Basic Speeds - Pump Torque Characteristics - Calculating Volumetric Efficiency

• Installation - Protection of Pump Against Seepage or Flood - Drive Alignment after Piping Installation

• Operation and Maintenance

ANSI/HI 7.1-7.5 Controlled Volume Metering Pumps (not often used in Australia) This Standard applies to Controlled-Volume Metering Pumps, which are reciprocating power pumps used to accurately displace a predetermined volume of liquid within a specified time. It contains sections on:

• General Description • Types and Nomenclature • Definitions • Application and sizing

- Typical performance curves - Materials of construction - Control methods

• Installation - Storage recommendations

• Operation and trouble solving

ANSI/HI 10.1-10.5 Air Operated Pumps (to our knowledge the only standard available for this type of pump and to date not seen in Australia) This standard is for air-operated pumps and includes those positive displacement reciprocating pumps used for general fluid transfer, which are driven by means of a compressed gas (usually air) from an outside source. The pump may be designed with a single diaphragm or double diaphragms connected to a reciprocating shaft in which one side of the diaphragm is in contact with the liquid being pumped and the other side is in contact with the compressed gas. The standard includes the following sections:

• Types, configurations, and nomenclature • Definitions • Design and Application • Installation, Operation, and Maintenance

ANSI/HI 12.1-12.6 Centrifugal Slurry Pumps (not often used in Australia) This standard for centrifugal slurry pumps for nomenclature, definitions, applications and operation addresses a long-standing challenge in the wastewater pump industry for development and acceptance of a test standard written specifically for slurry pumps. The scope of this standard includes:

• Slurry pump types • Types of slurries


• Slurry limitations • Slurry solids effect on pumps • Pumping of froth, pump wear, and application

The standard provides requirements for wet-end and bearing housing shaft seals, establishes allowable nozzle loads, and provides a data sheet that can be used by purchasers and vendors to exchange information. GENERAL PURPOSE ISO 2858:1975 End-suction centrifugal pumps (rating 16 bar) -- Designation, Nominal duty point and Dimensions The standard covers the ISO requirements for general purpose pumps primarily for use with water. The pump designation consists of three numbers detailing suction and discharge connection sizes and nominal impeller diameter in mm e.g. 125 x 100-250. The table covers flange sizes from 50 mm up to and including 200 mm. This is primarily a dimensional standard. DIN 24255 End-suction centrifugal pumps (normal rating 10 bar although some may claim 16 bar) This specification details single stage, end suction overhung impeller centrifugal pumps for general purpose applications design to be used primarily with water up to a maximum of 120 C. This is primarily a dimensional standard. The majority of DIN 24255 pumps supplied in Australia are manufactured in China.

Pump Clinic 43 Impeller Trimming Page 1 of 3

Impeller Trimming

Performance curves for centrifugal pumps normally detail the performance for a number of impeller diameters between a maximum and minimum allowable diameter. Impeller trimming means the reduction of the impeller diameter from maximum usually to adjust the pump performance to a required duty point.

Several things can happen when the impeller vane diameter is reduced. The diagram below will be used throughout this paper.

• Gap "A" describes the clearance between the impeller shrouds to the volute or casing and

• Gap "B" describes the clearance between the impeller vanes and the casing or volute.

• "D" describes the diameters of the vanes and shrouds

1. CHANGE IN PUMP PERFORMANCE

The change in pump performance with changes in impeller diameter can be predicted similarly to that with speed change utilising the Affinity laws

a) Pump flow rate (Q) varies directly with the diameter (D) i.e. Q1/Q2 = D1/D2 b) Pump head (H) varies with the square of the diameter (D) i.e. H1/H2 = (D1/D2)² c) Power absorbed varies with the cube of the diameter (D) i.e. P1/P2 = (D1/D2)³

These relationships (not laws) allow adjustment of the H-Q curve but there is a detrimental impact on efficiency especially for impeller reductions greater than 10% of maximum.

There are several reasons why this is true:

• The affinity laws assume the impeller shrouds are parallel. This is true only in low specific speed pumps.

• There is increased turbulence at the vane tips as the impeller is trimmed because the shroud to casing clearance (Gap "A") is increasing. This is sometimes referred to as "slip".

• The liquid exit angle is changed as the impeller is cut back, so the head/capacity curve becomes steeper.

Note that mixed flow impellers are more affected than low specific speed, radial vane impellers (high head/low capacity). Simply put, the greater the impeller reduction and the higher the specific speed of the impeller, the more the pump efficiency will decrease with impeller trimming. More accurate information can be obtained from a complete performance chart with different impeller diameters detailed.

PumpClinic…….Issue 43 July 2010


2. NPSH IMPACTS

In general, impeller diameter reductions greater than 5% to 10% of the maximum will increase the NPSHR (net positive suction head required) dependent on impeller specific speed.

The interaction of the physical geometry of the pump inlet, inclusive of the casing, impeller, and all associated wetted parts within the inlet field of flow determines the NPSHR characteristic of a pump. The value of NPSHR for any centrifugal pump is determined through performance testing. From NPSH test data, Suction Specific Speed (S) is calculated using the following equation, where Q represents flow at the best efficiency point of the pump.

S = rpm x √Q NPSHR¾

In can be seen in the above equation that NPSHR should not change with changes in impeller diameter as long as flow and RPM remain constant. There is no factor in the S equation that relates to impeller diameter. Suction specific speed (S) remains constant, for any defined inlet geometry, as long as the field of flow into the impeller eye is not disrupted by events taking place downstream of the impeller inlet. When trimming impellers on pumps that are of a low specific speed (Ns < 30 SI, 1500 US), tests have shown that there is little effect on NPSHR within the allowable impeller cut range. Beyond the allowable impeller cut range, recirculation between impeller discharge and the impeller inlet start to disrupt the inlet field of flow, increasing the NPSHR. For low Ns applications, full diameter NPSH values may be used for estimating NPSHR for cut impeller performance. For applications with Ns values above 30 (1500 US), a NPSH test is recommended to determine the NPSHR for any impeller trim.

3. MECHANICAL IMPACTS

i) Excessive shroud to casing clearance (Gap "A") and the resultant recirculation to the low pressure side of the pump will produce "eddy flows" around the impeller causing low frequency axial vibrations that can translate to mechanical seal problems. This can be a real concern in large pumps with powers over 200 kW or pumps pumping heads in excess 200 metres.

ii) For many years pump people have been machining the vane tips to reduce the vane passing frequency vibrations (Gap "B") while carefully maintaining Gap "A". The pulsating forces acting on the impeller can be reduced by 80% to 85% by increasing gap "B" from 1% to 6%.

iii) For impeller diameters up to 355 mm, gap "B" should be at least 4% of the impeller diameter to prevent "Vane passing syndrome cavitation" problems. For impeller diameters above 355 mm, Gap "B" should be at least 6% of the impeller diameter to prevent this type of cavitation. This type of cavitation damage is caused when the outside diameter of the impeller passes too close to the pump cutwater. The velocity of the liquid increases as it flows through this small passage, lowering the fluid pressure and causing local vaporisation. The bubbles then collapse at the higher pressure just beyond the cutwater. This is where volute damage occurs. Unless damage


has penetrated to the outside of the volute, a flashlight and mirror will most likely be required to inspect the damage. The damage is limited to the centre of the impeller vane. If it's a closed impeller, the damage will not extend into the shrouds.

iv) Although both the vanes and shrouds are often cut in end-suction, volute-type centrifugal pumps, it is not a good idea to do this in double suction designs. With these types of pumps you can reduce the vane diameters, but the shrouds should remain untouched.

v) Structural strength is a consideration when deciding how much to reduce the vane diameter in double ended pumps because you could leave too much unsupported shroud. Some manufacturers recommend an oblique cut that will improve the vane exit flow and add some strength to the shrouds.

vi) Machining a radius where the trimmed vane meets the shroud is another good idea to add strength to the assembly. Square corners are never a good idea.

4. NOISE

When writing a pump specification, many practicing engineers limit the impeller diameter to 85% of its maximum diameter. Such a limitation is actually a misunderstanding of a design concept known as "quiet pump operation." This misunderstanding may force the selection of a larger pump for the application. The idea here is not that the impeller diameter should be 85% of the maximum published diameter, but 85% of cutwater diameter (0.85 cutwater ratio). To fully understand the quiet pump operation design concept, refer to figure below.

In designing a pump casing, a design engineer first determines the volute scroll (A) necessary to handle the desired volume of water. This volute scroll terminates at the volute cut water (B) at the base of the discharge nozzle (C). The volute scroll is drawn around a base circle (D), which is sufficiently large enough to allow insertion of the impeller. The distance from the shaft centerline to the volute cut water is called the cutwater radius and twice this distance is the cutwater diameter

Hydraulic noise becomes a factor when the periphery of the impeller passes too close to the cutwater. In designing a pump, the distance between the impeller and the cutwater is a compromise between the pump efficiency and pump noise. Typically, cutwater ratios (D/F) of 0.9 and above produce higher noise and cutwater ratios of 0.8 and below produce significantly lower pump noise. Cutwater ratio of 0.85 is commonly specified by practicing engineers, thereby realising a minimum reduction in efficiency with a mean reduction in noise level.

From the above, it may be understood that a specification should more properly read "impeller diameter not to exceed 85% of the volute cutwater diameter” rather than "impeller diameter shall not exceed 85% of the maximum impeller diameter capable of being installed in the pump casing."

Specifying the later statement is safer since the impeller diameter would be even smaller than the desired maximum. In some cases this may force selection a larger pump than necessary.