Intro4u2u

Centrifugal pumps consist of a set of rotating vanes, enclosed within a housing or casing, used to impart energy to a fluid through centrifugal force. The pump has two main parts: a rotating element which includes an impeller and a shaft, and a stationary element made up of a casing (volute or solid), stuffing box, and bearings. Centrifugal pumps operate using kinetic energy to move fluid utilizing an impeller and a circular pump casing. The impeller produces liquid velocity and the casing forces the liquid to discharge from the pump converting velocity to pressure. This is accomplished by offsetting the impeller in the casing, and by maintaining a close clearance between the impeller and the casing at the cutwater. The fluid enters the pump near the center of the impeller and is moved to its outside diameter by the rotating motion of the impeller. The vanes on the impeller progressively widen from the center of the impeller that reduces speed and increases pressure. This allows centrifugal pumps to produce continuous flows at high pressure. By forcing the fluid through without cupping it, centrifugal pumps can achieve a very high flow rate.

Centrifugal pumps are used in many industries. Some of their most common applications /media transferred include: general purpose fluids, pure water, sludge and sewage, slurry, high viscosity fluids, power generation, the paper industry, the petroleum industry, chemicals and corrosives, gravel and solid materials, high temperature materials, and marine applications.

Centrifugal pumps generate flow by using one of three actions: radial flow, mixed flow, and axial flow. These classifications do not rate the performance quality of the pump, they are merely groupings based upon the pump’s action.

Radial flow pumps are centrifugal pumps in which the pressure is developed wholly by centrifugal force. In mixed flow pumps, the pressure is developed partly by centrifugal force and partly by the lift of the vanes of the impeller on the liquid. Axial flow centrifugal pumps develop pressure by the propelling or lifting action of the vanes of the impeller on the liquid.

The term “centrifugal pump” encompasses a multitude of pump technologies. Centrifugal means “directed or moving away from a center or axis”, therefore a centrifugal pump uses a rotating impeller to move the fluid outward. Fluid enters the pump and is drawn into the eye, or center, of the impeller and then is forced outward through the vanes (blades) via centrifugal force generated by the rotating action of the impeller. The fluid is forced to the outside of the pump casing (or volute) and out the pump’s discharge (see figure 1). The flow of a centrifugal pump depends on the system pressure drop: the higher the system’s pressure drop, the lower the flow (See application note: A Primer on Pressure Drop).

Lytron’s systems use seal-less, magnetically-driven centrifugal pumps, also known as mag-drives. Magnetically driven pumps use two magnets to drive the impeller. One magnet is attached to the motor shaft, generally referred to as the “drive magnet”. The other magnet is attached to the impeller (the “driven” or “impeller” magnet). The drive magnet spins causing the impeller magnet, and therefore impeller, to spin at the same rate. This pump design eliminates pump seals which often wear out from the friction caused by the rotation
of the motor shaft and are a source of leakage. In Lytron’s centrifugal pumps, the drive magnet is integrally molded into the impeller and thermoplastically coated to ensure zero contamination of the pump fluid. Thus, “mag-drive” ensures pump integrity and eliminates any possibility of shaft or seal leakage.

Magnetically-driven centrifugal pumps have many features that make them preferred for chiller/cooling system applications. When operated properly, they do not have any significant wear items therefore the centrifugal pump’s life will significantly exceed that of positive displacement pumps and centrifugal pumps with seals. Also this design does not generate particles that could clog the system’s filters so the pump’s performance will not change over time. All pumps will impart some heat to the fluid but it is important to minimize heat added from the pump to ensure the chiller has tight temperature stability. Since mag-drive pumps have minimal frictional surfaces to generate heat, they transfer far less heat to the fluid than other pump styles.

This long, maintenance-free service life, combined with the other design benefits have made magnetically driven centrifugal pumps a leading choice for Lytron’s cooling systems.

A recirculation diffuser for centrifugal pumps that produces a continuous rising head capacity curve and at the same time reduces levels of impeller vane-frequency sound and other high level acoustic flow tones. The conventional vane diffuser in the pump casing at the periphery of the rotary impeller is modified by separating one of the side shrouds from the diffuser vanes thereby creating a recirculation flow path around the vanes from the outside high pressure to the inside low pressure side.
What is claimed is:

1. A centrifugal pump for liquids having a casing with a vaned impeller rotatably mounted therein and a diffuser section at the periphery thereof comprising:

a pair of facing shroud members forming the diffuser section of the casing;

a plurality of hydrodynamically shaped diffuser vanes attached to at least one of said shrouds; and

a recirculation flow path around said diffuser vanes from the outside high pressure side to the inside low pressure side of said diffuser vanes.

2. The centrifugal pump of claim 1 wherein:

one of said shroud members is separated from the edges of said plurality of diffuser vanes to form a gap between the shroud and the diffuser vane thus providing a recirculation flow path.

3. The centrifugal pump of claim 2 further comprising:

spacers between the edges of said plurality of said plurality of diffuser vanes and said shroud members.

4. The centrifugal pump of claim 3 wherein:

said spacers are thumbtack shaped, the head of which is 20 to 40 mils thick.
Description:
BACKGROUND OF THE INVENTION

The instant invention relates generally to centrifugal pumps and more particularly to pump diffuser means having recirculation around the diffuser vanes.

One of the Navy’s major concerns in the design of its numerous types of ships and submarines is the reduction of radiated noise during operation to avoid detection by the enemy. It is known that centrifugal pump tonal noise contributes significantly to the radiated noise signature of the Navy’s ships and submarines.

While the design of quiet impellers for single-stage volute pumps is well established and used, a significantly more complex area of pump quieting technology is involved in multistage diffuser-type pump design. A multistage diffuser-type pump is characterized by its producing several impeller and diffuser vane-related tones rather than the single, vane frequency typical of the single-volute pumps.

Conventional multistage diffuser-type pumps have a vaned impeller and a plurality of diffuser vanes at the periphery of the impeller for directing the flow of fluid to the next stage. The diffuser vanes fit tightly between the parallel shrouds of the pump casing. This design results in a pump head-capacity curve which may droop close to and at flow shutoff. Another problem with this design is that pure-tone sounds are generated at the impeller vane passage frequency and at other frequencies not multiples of pump rotation.

It is desirable to design a centrifugal pump having a continuously rising head-capacity curve, (i.e. without droop near flow shut off) so as to be capable of stable parallel operation with another pump, and to have minimum vane-related tonal sound levels.

SUMMARY OF THE INVENTION

Accordingly, an object of the instant invention is to provide a new and improved diffuser type centrifugal pump.

Another object of the present invention is to provide a multistage, diffuser type centrifugal pump having a continuously rising head-capacity curve to shutoff.

Still another object of the present invention is to provide a centrifugal diffuser type pump having reduced levels of impeller vane-frequency sound and other high level acoustic flow tones.

A further object of the instant invention is to provide a centrifugal pump having a recirculation diffuser.

Briefly, these and other objects of the instant invention are attained by the use of a recirculation diffuser which permits recirculation flow around the diffuser vanes from the outside high pressure area to the inside low pressure area. The vanes are supported by the casing shroud on one side, and a gap separates the edges of the vanes from the shroud on the other side, thus providing a diffuser recirculation path for the pumped liquid.

Why diffuser recirculation works is not entirely understood, but apparently, providing the recirculation path, changes the hydrodynamic flow regime in the annulus between the impeller and the inner edges of the diffuser vanes enough that the objectionable flow-tone-producing mechanism is upset. It is known from experience that the greater the fluid wake mixing between the impeller exit and the diffuser entrance, the lower the vane noise level. Greater mixing is usually achieved by designing a larger radial clearance between the impeller and diffuser but this results in lower pump efficiency. With a recirculation diffuser, the small amount of flow returned to the mixing space tends to increase mixing and so smooths the impeller vane wakes and thus reduces vane frequency pulsations without excessive loss in efficiency.

Regarding the rising head capacity curve feature attained, it is well known that an external recirculation line on centrifugal pump system can extract from a drooping pump characteristic curve that portion which is continuously rising, but only at the expense of wasting all the energy of the recirculated flow.

The recirculation diffuser of the instant invention achieves the rising head characteristic in the same manner, but with less than half the energy loss of the external recirculation scheme. This is because assuming that the recirculation flow rate is the same in both external and diffuser recirculation, less than half the pump’s total developed pressure is generated by the diffuser. It is only a portion of the pressure developed by the diffuser that is lost in diffuser recirculation flow, and obviously, no pressure is lost that was produced in the impeller, per se.

BRIEF DESCRIPTION OF THE DRAWINGS

A more complete understanding of the invention and many of the attendant advantages thereof will be readily appreciated as the same become better understood by reference to the following detailed description when considered in connection with the accompanying drawing wherein:

FIG. 1 is a partial sectional view in elevation taken through the impeller and diffuser area; and

FIG. 2 is a partial sectional view taken at right angles along the section line 2–2 of FIG. 1.

DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring now to the drawings, wherein like reference numerals designate corresponding views throughout the several views, there is shown generally in FIG. 1 a stage of a centrifugal pump 10, which may be a stage of a multistage pump. The pump has an impeller 12 mounted on a rotatable shaft 14 journaled in a pump casing (not shown). The impeller has a plurality of curved and hydrodynamically shaped vanes, 16 which are conventionally integral with the impeller 12. Vanes 16 are between side flanges 18 forming a closed centrifugal flow path for the fluid from the inlet 20 to the outlet 22 at the periphery, as shown by the arrows.

Radially outward from the impeller outlet 22 is a space or annulus 24, separating the impeller 12 from a diffuser section 26. The diffuser section 26 comprises a plurality of diffuser vanes 28 attached along one edge to one of the casing shrouds 30 and spaced a distance of 20 to 40 mils from an other casing shroud 32. A gap 34 is therefore formed between the vanes 28 and the casing shroud 32, and may be maintained by the head of a thumbtack-shaped spacer 36 which is retained in a bore in the vanes 28. Radially outward from the diffuser section 26 is a collector 38 wherein the liquid is collected and exhausted at an elevated pressure.

In operation the liquid to be pumped is admitted into the pump suction near the impeller shaft 14 at the center of the pump casing. Here the liquid is conducted to the inlet 20 of the centrifugal impeller 12 and thrown outwardly by centrifugal force induced by the rotating impeller, past the impeller vanes 16. The liquid exits the periphery of the impeller at high velocity into the annulus 24 that forms a mixing space between the impeller and the diffuser section 26. The liquid continues by centrifugal force to flow into the diffuser 26 past the diffuser vanes 28 where the flow is straightened and the velocity converted to pressure increase, and thence into the collector 38 at high pressure. In the diffuser section 26 a portion of the flow recirculates back around the diffuser vanes 28 via the gap 34 (see dashed arrows) which creates a flow path from the outside high pressure to the inside low pressure.

While the mechanism is not completely understood, the recirculation flow mixes in the annulus with the main flow from the impeller to upset the even flow which induces the objectional flow tones. Also the recirculation flow, as in external recirculation, creates a rising head capacity curve. These novel features are discussed in detail hereinbefore in regard to the summary of the invention.

Obviously many modifications and variations of the present invention are possible in light of the above teachings. It is therefore to be understood that within the scope of the appended claims, the invention may be practiced otherwise than as specifically described.

The Centrifugal Pump block represents a centrifugal pump of any type as a data-sheet-based model. The pump is parameterized with the polynomial whose coefficients are determined analytically or experimentally, depending on the data available. The relationship between pump characteristics and angular velocity is determined from the affinity laws. The model can be used if the shaft rotates in positive direction only.

The centrifugal pump is simulated with the following equation:
(2-1)

where
p    Pressure differential across the pump
k    Correlation factor. The factor is introduced to account for dimensional fluctuations, blade incongruity, blade volumes, fluid internal friction, and so on. The factor should be set to 1 if the approximating coefficients are determined experimentally.
pE    Euler pressure
pHL    Pressure loss due to hydraulic losses in the pump passages
pD    Pressure loss caused by deviations of the pump delivery from its nominal value

The Euler pressure, pE, is determined with the Euler equation for centrifugal machines [1, 2] based on known pump dimensions. For an existing pump, the Euler pressure can be approximated with the equation

where
ρref    Fluid density
c0,c1    Approximating coefficients. They can be determined either analytically from the Euler equation [1, 2] or experimentally.
q    Pump volumetric delivery

The pressure loss due to hydraulic losses in the pump passages, pHL, is approximated with the equation

where
ρref    Fluid density
c2    Approximating coefficient
q    Pump volumetric delivery

The blade profile is determined for a specific fluid velocity, and deviation from this velocity results in pressure loss due to inconsistency between the fluid velocity and blade profile velocity. This pressure loss, pD, is estimated with the equation

where
ρref    Fluid density
c3    Approximating coefficient
q    Pump volumetric delivery
qD    Pump design delivery (nominal delivery)

The pump characteristics, approximated with four coefficients c0, c1, c2, and c3, are determined for a specific fluid and a specific angular velocity of the pump’s driving shaft. These two parameters correspond, respectively, to the Reference density and Reference angular velocity parameters in the block dialog box. To apply the characteristics for a different velocity, the affinity laws are used. First, the new reference delivery is computed with the expression
(2-2)

where q and ω are the instantaneous values of the pump delivery and angular velocity. Then the pressure differential across the pump at a different angular velocity and density is determined with the formula

where pref is the pressure differential computed with Equation 2-1 at pump delivery determined according to Equation 2-2.

The pump efficiency is assumed to be the same as it is at the reference parameters. It is computed with the following equations:

where
η    Pump efficiency
Nref.hyd    Power of the flow at the pump’s outlet
pref    Pressure differential across the pump at delivery q = qref
qref    Pump reference delivery
pEref    Euler pressure at reference parameters
Nref.br    Mechanical brake power at the pump’s driving shaft
Nmech.loss    Power of mechanical losses in the pump drive train

Assuming that the efficiency remains the same at similar regimes, the torque at the driving shaft is determined from the following equation:

The hydraulic power at the pump outlet is computed with the equation

where p and q are the current values of the pump pressure differential and delivery, respectively.

The block positive direction is from port T to port P. This means that the pump transfers fluid from T to P as its driving shaft S rotates in the globally assigned positive direction.
Basic Assumptions and Limitations

The model is based on the following assumptions:

*

Fluid compressibility is neglected.
*

The pump rotates in positive direction only.
*

No reverse flow through the pump is allowed.
*

The pump efficiency remains the same at similar regimes.

Dialog Box and Parameters

First approximating coefficient

Approximating coefficient c0 in the block description preceding. The default value is 362 Pa/(kg/m^3).
Second approximating coefficient

Approximating coefficient c1 in the block description preceding. The default value is 1.65e4 Pa*s/kg.
Third approximating coefficient

Approximating coefficient c2 in the block description preceding. This coefficient accounts for hydraulic losses in the pump. The default value is 1.69e7 Pa*s^2/(kg*m^3).
Fourth approximating coefficient

Approximating coefficient c3 in the block description preceding. This coefficient accounts for additional hydraulic losses caused by deviation from the nominal delivery. The default value is 2.34e6 Pa*s^2/(kg*m^3).
Correction factor

The factor, denoted as k in the block description preceding, accounts for dimensional fluctuations, blade incongruity, blade volumes, fluid internal friction, and other factors that decrease Euler theoretical pressure. The default value is 0.85.
Reference angular velocity

Angular velocity of the driving shaft, at which the pump characteristics are determined. The default value is 1.77e3 rpm.
Pump design delivery

The pump nominal delivery. The blades profile, pump inlet, and pump outlet are shaped for this particular delivery. Deviation from this delivery causes an increase in hydraulic losses. The default value is 130 lpm.
Reference density

Fluid density at which the pump characteristics are determined. The default value is 920 kg/m^3.
Mechanical loss power

Power of mechanical loss in the pump drive train at reference parameters. The default value is 350 W.

Global Parameters

Fluid density

The parameter is determined by the type of working fluid selected for the system under design. Use the Hydraulic Fluid block or the Custom Hydraulic Fluid block to specify the fluid properties.

Ports

The block has the following ports:

T

Hydraulic conserving port associated with the pump suction, or inlet.
P

Hydraulic conserving port associated with the pump outlet.
S

Mechanical rotational conserving port associated with the pump driving shaft.

Centrifugal Pump Fundamentals

A-9 System Curves
For a specified impeller diameter and speed, a centrifugal pump has a fixed and predictable performance curve. The point where the pump operates on its curve is dependent upon the characteristics of the system In which it is operating, commonly called the System Head Curve. ..or, the relationship between flow and hydraulic losses* in a system. This representation is in a graphic form and, since friction losses vary as a square of the flow rate, the system curve is parabolic in shape.

By plotting the system head curve and pump curve together, it can be determined:

1. Where the pump will operate on its curve.
2. What changes will occur if the system head curve or the pump performance curve changes.

NO STATIC HEAD - ALL FRICTION
As the levels in the suction and discharge are the same (Fig. 1), there is no static head and, therefore, the system curve starts at zero flow and zero head and its shape is determined solely from pipeline losses. The point of operation is at the intersection of the system head curve and the pump curve. The flow rate may be reduced by throttling valve.

Fig.1 No Static Head All Friction

POSITIVE STATIC HEAD
The parabolic shape of the system curve is again determined by the friction losses through the system including all bends and valves. But in this case there is a positive static head involved. This static head does not affect the shape of the system curve or its “steepness”, but it does dictate the head of the system curve at zero flow rate.

The operating point is at the intersection of the system curve and pump curve. Again, the flow rate can be reduced by throttling the discharge valve.

Fig. 2 Positive Suction Head

NEGATIVE (GRAVITY) HEAD
In the illustration below, a certain flow rate will occur by gravity head alone. But to obtain higher flows, a pump Is required to overcome the pipe friction losses in excess of “H” - the head of the suction above the level of the discharge. In other words, the system curve is plotted exactly as for any other case involving a static head and friction head, except the static head is now negative. The system curve begins at a negative value and shows the limited flow rate obtained by gravity alone. More capacity requires extra work.

Fig. 3 Negative (Gravity) Head

MOSTLY LIFT- LITTLE FRICTION HEAD
The system head curve in the illustration below starts at the static head “H” and zero flow. Since the friction losses are relatively small (possibly due to the large diameter pipe), the system curve is “flat”. In this case. the pump is required to overcome the comparatively large static head before it will deliver any flow at all.

Fig. 4 Mostly Lift - Little Fricition Head

*Hydraulic losses in piping systems are composed of pipe friction losses, valves, elbows and other fittings, entrance and exit losse (these to the entrance and exit to and from the pipeline normally at the beginning and end not the pump) and losses from changes in pipe size by enlargement or reduction in diameter.

Centrifugal Pump Fundamentals

A-8 Affinity Laws
The affinity laws express the mathematical relationship between the several variables involved in pump performance. They apply to all types of centrifugal and axial flow pumps. They are as follows:

1. With impeller diameter D held constant:
Where:
Q = Capacity, GPM
H = Total Head, Feet
BHP = Brake Horsepower
N = Pump Speed, RPM
2. With speed N held constant:

When the performance (Q1, H1, & BHP1) is known at some particular speed (N1) or diameter (D1), the formulas can be used to estimate the performance (Q2, H2, & BHP2) at some other speed (N2) or diameter (D2). The efficiency remains nearly constant for speed changes and for small changes in impeller diameter.

Example:
To illustrate the use of these laws, refer to Fig. 8 below. It shows the performance of a particular pump at 1750 RPM with various impeller diameters. This performance data has been determined by actual tests by the manufacturer. Now assume that you have a 13″ maximum diameter impeller, but you want to belt drive the pump at 2000 RPM.

Fig. 8 Composite Performance Curve

The affinity laws listed under 1 above will be used to determine the new performance, with N1 1750 RPM and N2 = 2000 RPM. The first step is to read the capacity, head, and horsepower at several points on the 13″ dia. curve in Fig. 9 below. For example, one point may be near the best efficiency point where the capacity is 300 GPM, the head is 160 ft, and the BHP is approx. 20 hp.

This will then be the best efficiency point on the new 2000 RPM curve. By performing the same calculations for several other points on the 1750 RPM curve, a new curve can be drawn which will approximate the pump’s performance at 2000 RPM, Fig. 9.

Trial and error would be required to solve this problem in reverse. In other words, assume you want to determine the speed required to make a rating of 343 GPM at a head of 209 ft. You would begin by selecting a trial speed and applying the affinity laws to convert the desired rating to the corresponding rating at 1750 RPM. When you arrive at the correct speed, 2000 RPM in this case, the corresponding 1750 RPM rating will fall on the 13″ diameter curve.

Centrifugal Pump Fundamentals

A-7 Pump Characteristic Curves
The performance of a centrifugal pump can be shown graphically on a characteristic curve. A typical characteristic curve shows the total dynamic head, brake horsepower, efficiency, and net positive Suction head all plotted over the capacity range of the pump.

Figures 5, 6, & 7 are non-dimensional curves which indicate the general shape of the characteristic curves for the various types of pumps. They show the head, brake horsepower, and efficiency plotted as a percent of their values at the design or best efficiency point of the pump.

Fig. 5 below shows that the head curve for a radial flow pump is relatively flat and that the head decreases gradually as the flow increases. Note that the brake horsepower increases gradually over the flow range with the maximum normally at the point of maximum flow.

Fig. 5 Radial Flow Pump

Mixed flow centrifugal pumps and axial flow or propeller pumps have considerably different characteristics as shown in Figs. 6 and 7 below. The head curve for a mixed flow pump is steeper than for a radial flow pump. The shut-off head is usually 150% to 200% of the design head, The brake horsepower remains fairly constant over the flow range. For a typical axial flow pump, the head and brake horsepower both increase drastically near shutoff as shown in Fig. 7.

Fig. 6 Mixed Flow Pump

Fig. 7 Axial Flow Pump

The distinction between the above three classes is not absolute, and there are many pumps with characteristics falling somewhere between the three. For instance, the Francis vane impeller would have a characteristic between the radial and mixed flow classes. Most turbine pumps are also in this same range depending upon their specific speeds.

Fig. 8 below shows a typical pump curve as furnished by a manufacturer. It is a composite curve which tells at a glance what the pump will do at a given speed with various impeller diameters from maximum to minimum. Constant horsepower, efficiency, and NPSHR lines are superimposed over the various head curves. It is made up from individual test curves at various diameters.

NPSH and Suction Specific Speed
In designing a pumping system, it is essential to provide adequate NPSH available for proper pump operation. Insufficient NPSH available may seriously restrict pump selection, or even force an expensive system redesign. On the other hand, providing excessive NPSH available may needlessly increase system cost.

Suction specific speed may provide help in this situation.

Suction specific speed (S) is defined as:

Where
N = Pump speed RPM
GPM = Pump flow at best efficiency point at impeller inlet (for double suction impellers divide total pump flow by two).
NPSHR = Pump NPSH required at best efficiency point.

For a given pump, the suction specific speed is generally a constant - it does not change when the pump speed is changed. Experience has shown that 9000 is a reasonable value of suction specific speed. Pumps with a minimum suction specific speed of 9000 are readily available, and are not normally subject to severe operating restrictions, unless the pump speed pushes the pump into high or very high suction energy.

An example:
Flow 2,000 GPM; head 600 ft. What NPSHA will be required?

Assume: at 600 ft., 3500 RPM operation will be required.

A related problem is in selecting a new pump, especially at higher flow, for an existing system. Suction specific speed will highlight applications where NPSHA may restrict pump selection. An example:

Existing system: Flow 2000 GPM; head 600 ft.; NPSHA 30 ft.; Specific Gravity 1.0; Suction Nozzle 6 in. - What is the maximum speed at which a pump can be run without exceeding NPSH available? (NPSHMargin Ratio = 1.5 from above @ S.E. = 173 x 106)

Running a pump at this speed would require a gear and at this speed, the pump might not develop the required head. At a mini-mum, existing NPSH A is constraining pump selection.

Same system as 1. Is a double suction pump practical?
For a double suction pump De = .75 x 6″ = 4.5
S.E. = 4.5 x 3550 x 9000 x 1.0
S.E. = 136 x 106 (High S.E.)

For a double suction pump, flow is divided by two.

Using a double suction pump is one way of meeting system NPSH and obtaining a higher head.

The amount of energy in a pumped fluid, that flashes into vapor and then collapses back to a liquid in the higher pressure area of the impeller inlet, determines the extent of the noise and/or damage from cavitation. Suction Energy is defined as:

Suction Energy = De x N x S x Sg

Where D e = Impeller eye diameter (inches)
Sg = Specific gravity of liquid (Sg - 1.0 for cold water)

High Suction Energy starts at 160 x 10 6 for end suctabtion pumps and 120 x 10 6 for horizontal split case pumps. Very high suction energy starts at 1.5 times the High Suction Energy values. For estimating purposes you can normally assume that the impeller eye diameter is approximately 90% of the suction nozzle size, for an end suction pump, and 75% of the suction size for a double suction split case pump.

According to the Hydraulic Institute, ans NPSH margin is required above the NPSHR of the pump to supress incipient cavitation. The amount of margin is a function of Suction Energy and the critical nature of the application as follows:
Suction Energy     NPSHMargin Ratio (NPSHA/NPSHR)
Low     1.1 - 1.3
High     1.2 - 1.7
Very High     1.7 - 2.5

Suction specific speed 9,000, pump speed 3550 RPM, suction nozzle size 6 inch, specific gravity 1.0, and the pump type is end suction.

De ~ .9 x 6″ = 5.4″
Suction Energy = De x N x S x Sg
= 5.4 x 3550 x 9,000 x 1.0
= 173 x 106

Since 173 x 106 > 160 x 106 , this is a High Suction Energy pump.

Net Positive Suction Head (NPSH) NPSH Available is a function of the systetm in which the pump operates. It is the excess pressure of the liquid in feet absolute over its vapor pressure as it arrives at the pump suction. Fig. 4 shows four typical suction systems with the NPSH Available formulas applicable to each. It is important to correct for the specific gravity of the liquid and to convert all terms to units of “feet absolute” in using the formulas.

PB = Barometric pressure in feet absolute.
VP = Vapor pressure of the liquid at maximum pumping temperature, in feet absolute.
P = Pressure on surface of liquid in closed suction tank, in feet absolute.
Ls = Maximum static suction lift in feet.
LH = Minimum static suction head in feet.
hf = Friction loss in feet in suction pipe at required capacity

Fig. 4 Calculation of system Net Positive Suction Head Available for typical suction conditions.

In an existing system, the NPSH Available can be determined by a gauge on the pump suction. The following formula applies:

Where
Gr = Gauge reading at the pump suction expressed in feet (plus if above atmospheric, minus if below atmospheric) corrected to the pump centerline.
hv = Velocity head in the suction pipe at the gauge connection, expressed in feet.

Cavitation is a term used to describe the phenomenon, which occurs in a pump when there is insufficient NPSH Available. When the pressure of the liquid is reduced to a value equal to or below its vapor pressure the liquid begins to boil and small vapor bubbles or pockets begin to form. As these vapor bubbles move along the impeller vanes to a higher pressure area above the vapor pressure, they rapidly collapse.

The collapse, or “implosion” is so rapid that it may be heard as a rumbling noise, as if you were pumping gravel. In high suction energy pumps, the collapses are generally high enough to cause minute pockets of fatigue failure on the impeller vane surfaces. This action may be progressive, and under severe (very high suction energy) conditions can cause serious pitting damage to the impeller.

The accompanying noise is the easiest way to recognize cavitation. Besides possible impeller damage, excessive cavitation results in reduced capacity due to the vapor present in the pump. Also, the head may be reduced and/or be unstable and the power consumption may be erratic. Vibration and mechanical damage such as bearing failure can also occur as a result of operating in excessive cavitation, with high and very high suction energy pumps.

The way to prevent the undesirable effects of cavitation in standard low suction energy pumps is to insure that the NPSH Available in the system is greater than the NPSH Required by the pump. High suction energy pumps require an additional NPSH margin, above the NPSH Required. Hydraulic Institute Standard (ANSI/HI 9.6.1) suggests NPSH margin ratios of from 1.2 to 2.5 times the NPSH Required, for high and very high suction energy pumps, when operating in the allowable operating range.

Net Positive Suction Head (NPSH) and Cavitation
The Hydraulic Institute defines NPSH as the total suction head in feet absolute, determined at the suction nozzle and corrected to datum, less the vapor pressure of the liquid in feet absolute. Simply stated, it is an analysis of energy conditions on the suction side of a pump to determine if the liquid will vaporize at the lowest pressure point in the pump.

The pressure which a liquid exerts on its surroundings is dependent upon its temperature. This pressure, called vapor pressure, is a unique characteristic of every fluid and increased with increasing temperature. When the vapor pressure within the fluid reaches the pressure of the surrounding medium, the fluid begins to vaporize or boil. The temperature at which this vaporization occurs will decrease as the pressure of the surrounding medium decreases.

A liquid increases greatly in volume when it vaporizes. One cubic foot of water at room temperature becomes 1700 cu. ft. of vapor at the same temperature.

It is obvious from the above that if we are to pump a fluid effectively, we must keep it in liquid form. NPSH is simply a measure of the amount of suction head present to prevent this vaporization at the lowest pressure point in the pump.

NPSH Required is a function of the pump design. As the liquid passes from the pump suction to the eye of the impeller, the velocity increases and the pressure decreases. There are also pressure losses due to shock and turbulence as the liquid strikes the impeller. The centrifugal force of the impeller vanes further increases the velocity and decreases the pressure of the liquid. The NPSH Required is the positive head in feet absolute required at the pump suction to overcome these pressure drops in the pump and maintain the majority of the liquid above its vapor pressure. The NPSH Required varies with speed and capacity within any particular pump. Pump manufacturer’s curves normally provide this information.