PREDICTING CENTRIFUGAL PUMP TYPE AND NPSH IN EARLY

Copyright© 2019 by Turbomachinery Laboratory, Texas A&M Engineering Experiment Station

PREDICTING CENTRIFUGAL PUMP TYPE AND NPSH IN EARLY FACILITY DESIGN

Patrick C. Green, P.E. Principal Machinery Engineer LyondellBasell Houston, TX, USA

Patrick Green is a principal machinery engineer in LyondellBasell’s Global Engineering & Turnarounds organization. He currently works as the rotating equipment lead on one of LyondellBasell's several large greenfield projects. His experience ranges from operation, maintenance, and failure analysis of machinery in oil, gas, and petrochemicals facilities to specification, assurance, and installation of new machinery in the same industries. He is a 2008 graduate of the University of Illinois with a Bachelor of Science degree in Mechanical Engineering, a registered professional engineer in the state of Louisiana, a member of ASME, and a member of the Texas A&M International Pump Users Symposium Advisory Committee.

ABSTRACT For new and expanded facilities, prediction of the size and type of pumps used for applications in those facilities during early plant design can be a difficult and daunting task, especially for new process technologies and for greenfield projects with an abundant number of pump applications. The earlier the size and type of pumps can be predicted in a project, and the earlier net positive suction head (NPSH) requirements can be accurately estimated, the earlier certain aspects of facility hydraulic design can be clarified. Not only can this result in better economical choices regarding facility layout and pump selection, but it can also minimize developmental re-work during the later detailed engineering phases of projects. This tutorial focuses on centrifugal pumps, which are generally the workhorse of industrial oil, gas, and petrochemical facilities, and the tutorial approaches this early assessment concept in several parts. First, it examines some of the theoretical hydraulic selection parameters that can be calculated for pump applications, some of the physical limiting factors pertaining to those hydraulic selection parameters, and the assumptions upon which those calculations might be based. Once the groundwork has been laid for those calculations, the tutorial examines the use of those hydraulic parameters in conjunction with the conditions of service to predict pump type and NPSHR against actual pump selections for two greenfield projects with a combined total of over 150 pump applications and identifies any shortcomings of these predictions. Lastly, this tutorial covers some examples of potential pitfalls due to lack of good design and testing practices not being implemented at the proper stage of project execution that may render the aforementioned predictions less accurate or may result in facility re-design. INTRODUCTION This tutorial focuses on a method for predicting pump type and setting NPSHA for pump applications in early design of new facilities. The ultimate intent is to help the engineer to be more efficient and accurate in selecting pumps and setting NPSHA for projects involving greenfield facilities where there are an abundant number of applications. For projects involving only a handful of pumps, or for projects where a single pump is being replaced in an existing facility, the author recommends using the pump hydraulic selection software or catalogs readily available from pump manufacturers across the industry. The approach taken herein is generally not practical for a single pump application or handful of applications. That said, the reader may benefit from this tutorial in understanding what is possible or not possible for centrifugal pumps for his or her pumping application(s).


HYDRAULIC PARAMETERS Net Positive Suction Head (NPSH) All pumping applications have a hydraulic variable associated with them known as Net Positive Suction Head, commonly referred to as NPSH. It is defined as the amount of specific energy or head of a liquid above the specific energy or head that the liquid has at its vapor pressure at the liquid’s temperature. It is calculated as follows:

(1)

Note that the variable for the liquid head includes the head due to pressure, due to velocity, and due to elevation. There are two terms for NPSH generally discussed in industry, the available NPSH of the liquid (NPSHA) and the NPSH required by the pump (NPSHR) in order to meet a certain criterion. The most common criterion used in industry today and the one that historically has been used to represent NPSHR is a 3% reduction in the differential head of the pump. In many of today’s industry standards, the NPSH at which a pump’s differential head is reduced by 3% at a given flowrate is referred to as NPSH3. For the purposes of this paper, NPSHR and NPSH3 will be used interchangeably to mean the same thing. Note also that the difference between NPSHA and NPSHR is typically referred to as the NPSH margin, with “positive” margin implying that NPSHA > NPSHR. The reduction in differential head is due to vapor bubbles that form in the impeller due to an insufficient amount of liquid head above the liquid’s vapor pressure. This reduces the average density of the fluid going through the pump, which ultimately reduces the differential pressure produced by the pump and is perceived as a reduction in differential head. The reader should be aware that for multistage pumps, this reduction in NPSH is only evaluated for the first stage of the pump since those vapor bubbles will condense back into liquid after the first stage, with subsequent stages of the pump having sufficient NPSH margin to eliminate any cavitation that might noticeably affect the differential head performance of the pump. Specific Speed Perhaps one of the most commonly known hydraulic parameters for centrifugal pumps is specific speed, typically denoted Ns (or Nq when calculated in SI units). Generally speaking, this parameter characterizes an individual impeller’s shape and can be used to identify what type of pump might be suitable for an application. It is calculated as follows:

(2)

Figure 1: Impellers and Corresponding Specific Speeds

(courtesy of the Hydraulic Institute [11], www.pumps.org)

http://www.pumps.org/


Note that the parameters in this equation are specific to the pump itself, rather than the required flow and head for the service. The volumetric flow and differential head values used in the equation are at the best efficiency point of the maximum diameter impeller for the subject pump. Note also that for multistage pumps, this is for an individual impeller and not the pump itself. This means that for multistage pumps TDHBEP from the pump curve must be divided by the number of impellers to calculate the specific speed, assuming each impeller imparts the same differential head to the liquid. Regarding how the specific speed characterizes the impeller shape, many are likely familiar with the chart shown above or one similar to it.

For double-suction pumps, specific speed can be calculated one of two ways: either using the full volumetric flow through the discharge nozzle at QBEP or using half of that flow to indicate the flow through each eye of the double-suction impeller. The former method better reflects the discharge characteristics of the pump, while the latter method reflects the suction characteristics of each impeller eye. Since this tutorial is ultimately focused on assuring sufficient NPSH for pumps during facility design, the convention used for double-suction pumps will be that Ns is calculated using half of the BEP flow for the maximum diameter impeller. Note that relatively low specific speeds indicate that a positive displacement pump may be a better option for the application than a centrifugal pump, whereas relatively high specific speeds are associated with axial flow machinery. It should also be noted that recent developments in centrifugal pump design and hydraulics have allowed for relatively low specific speeds (Ns < 200 USC) to be achieved with the use of a single, fully closed, overhung impeller. Means to achieve this low Ns include large impeller diameters (sometimes greater than 30”), small impeller tip passageways sometimes requiring fabricated impellers rather than cast impellers, and tighter than normal clearances achieved with the use of composite nonmetallic materials as stationary wear rings. These services have historically been satisfied using Barske hydraulics, multi-stage pumps, or positive displacement pumps. Suction Specific Speed Suction specific speed, typically denoted Nss (or sometimes S or SSS), is a parameter that characterizes the net positive suction head required for a pump relative to its rotational speed and flow. It is calculated similar to Ns, as follows:

(3)

Once again, these parameters are calculated at the best efficiency flow for the specified pump with the maximum diameter impeller installed. For double-suction pumps, since this parameter deals with the suction characteristics of the pump, the flow used in the calculation should be half the BEP flow for the max impeller diameter. The work by Jerry Hallam in the early 1980s [1] set the tone for pump selection regarding Nss: namely, that selection of pumps with Nss near or above 11,000 USC should be approached with a higher level of scrutiny than typical. Furthermore, the publication by Lobanoff and Ross [2] indicates a recommended stable operating window for centrifugal pumps based on their Nss, as seen in Figure 2. The theme from this plot is that higher values of Nss result in smaller stable operating windows.

Figure 2: Stable Operating Windows per Lobanoff and Ross (courtesy of Texas A&M Turbomachinery Laboratory [12])


These recommendations are based on occurrence of suction and discharge recirculation phenomena associated with operation of high Nss pumps outside the recommended ranges, which can result in cavitation, vibration, and potential damage to the pump and seal over the long term. Recent work by Cowan et al [3] shows that certain geometric features can be incorporated into the suction side of the pump to alleviate many of the issues associated with high Nss designs of previous decades and recommends a modification of the plot originally put forth by Lobanoff and Ross with larger stable operating ranges than those shown above. It should be noted that the suction characteristics of the pump, which are quantified via the Nss calculation, do not change with impeller trim. The recommended stable operating windows are based on the BEP of the maximum impeller diameter, not the BEP of the furnished impeller diameter. Caution should be exercised when selecting a high Nss pump, even with today’s advancements in hydraulic design. While it may be possible to incorporate certain features into a pump design to alleviate these issues, the recent work by Cowan et al is no guarantee that a manufacturer actually incorporates such features. The purchasing engineer should request supporting information from the pump manufacturer as to how the manufacturer manages to create a wide stable operating range for a relatively high Nss pump (generally greater than 11,000 USC), including not only analytical information but also empirical data corroborating the manufacturer’s claims. The purchasing engineer should also ensure that these modifications to the hydraulic design are clearly communicated as part of the pump purchase. PREDICTING NPSHR IN EARLY FACILITY DESIGN Review of Equation 3 will reveal that the NPSHR of the pump can be calculated based on the Nss, rotational speed, and BEP flow of the pump for the maximum impeller diameter, as follows:

(4)

In the early stages of facility design, these values will not be precisely known for the ultimately selected pump. This presents a challenge as to what minimum NPSHA should be set for the service. Too low of NPSHA may result in a very large, very slow pump to accommodate the service or potentially even installing the pump in a dry sump below grade. Requiring too high a minimum NPSHA may result in unnecessarily elevating large suction vessels, requiring large, complex concrete and steel structures to properly support them. Fortunately, some approximations for the variables in Equations 2 and 4 may be made in an effort to predict the type of pump and the NPSHR for the specified conditions of the pump application, the validity of which will be assessed later in this tutorial. Approximation for Flow Regarding the flow, it can be assumed that the specified rated flow for the pump, for which the designation Q* will be used, is equal to the BEP flow for the max diameter impeller. Industry standards such as API 610 11th Edition [4] require that the rated point is within a range of the BEP for the furnished impeller. Depending on the maximum diameter of the impeller for the selected pump, the total allowable trim range for the impeller, and the hydraulic design of the impeller and stator, the BEP flow for the furnished impeller can shift dramatically to the left of the curve from the BEP flow for the maximum diameter impeller. The effectiveness of the suggested approximation that Q*≈QBEP can change based on the above listed parameters, but especially based on impeller trim. For example, in the instance that the impeller is trimmed to the low end of the allowable trim range, and the rated flow falls significantly to the left of the BEP for the furnished trimmed impeller, then this approximation may not be an accurate one. This will be examined later on. Approximation for Nss Regarding Nss, a relatively blind assumption must be made. Again, assuming too high of Nss may result in predicted NPSHR values that are far lower than necessary, while assuming too low of Nss may result in unnecessarily high NPSHR values, potentially requiring the elevation of suction vessels and process equipment to be unnecessarily high. Nss = 9000 USC will be assumed for multiple reasons. Experience has shown that most pump manufacturers can furnish impellers with Nss near 9000 for most applications. More importantly, Nss = 9000 is a relatively conservative assumption, as higher Nss impellers can be selected to reduce the NPSHR in the case that the assumption of Nss = 9000 does not result in setting the NPSHA for the service sufficiently low during design.


Approximation for Speed The assumed speed is generally limited by the following factors:

• Synchronous power grid frequency (for motor-driven, direct-drive pumps) • Required differential head • A reasonably low NPSHR value

In certain instances, pumps may be run faster or slower than synchronous or fractional synchronous speeds. However, for onshore downstream applications these are usually limited to special applications and are generally not economical for normal applications due to the need for a gearbox or VFD. So assuming the pump is motor-driven, the next limit is the differential head required. The assumption in this instance is that the head required per impeller is the impeller tip head, as follows:

(5)

This formula is derived from Euler’s turbine equation, with impeller tip speed being substituted for the actual fluid tangential velocity. For multistage pumps, TDH*stage is calculated by dividing TDH*pump by the total number of stages. Note also that this approximation can be modified by the known correlation of the ideal head coefficient ψ for an impeller given the specific speed [5]. For simplicity’s sake and for the reason that actual pump specific speed is not known at this point in facility design, this tutorial will forego any variation in ψ in Equation 5 via use of a correlated head coefficient (i.e. assume ψ = 1). Equation 5 allows us to solve for impeller tip speed, which can then be used to check what synchronous or fractional synchronous rotational speed is realistic based on

• Practical upper limits of impeller diameter • Practical limits of Ns (or at this point, Ns*) for centrifugal pumps (see Figure 1)

It should again be noted that both the high limit for impeller diameters and the low limit for Ns have been pushed in recent years. Based on the determined required impeller tip speed, the approximate pump impeller diameter can be determined via the following equation:

(6)

Some reasonable limits on impeller diameter are as follows: • For single-stage, overhung pumps

o 4” minimum o 18” maximum at speeds near 3600 RPM o 30” maximum at speeds near 1800 RPM or slower

• For single-stage, double-suction, between-bearing pumps o 6” minimum o 18” maximum at speeds near 3600 RPM o 32” maximum at speeds near 1800 RPM o 42” maximum at speeds near 900 RPM or slower

• For multi-stage pumps (two or more stages) o 16” maximum for BB1/2 at speeds 3600 RPM o 32” maximum for BB1/2 at speeds near 1800 RPM or slower o 18” maximum for BB3 o 18” maximum for BB5

Where the “BB” classifications are as defined in API 610 11th Edition [4] and in ANSI/HI 14.1-14.2-2019 [6]. Careful examination of the above listed diameters and speeds will reveal that impeller tip speeds for commercially available pumps generally do not exceed 300 ft/second. Some specialized high pressure, high energy pumping applications may exceed this value. In the case that the chosen rotational speed results in an unreasonably high impeller diameter, the assumed speed can be increased, or the assumed pump type can be modified to increase the number of impellers to accommodate the higher head with a lower tip speed per impeller. For unreasonably small impeller diameters, the rotational speed can be reduced to allow for a larger impeller diameter, or a mixed flow impeller or axial flow impeller might also be applied.


Another way of determining a reasonable pump configuration and rotational speed is by simply using the below parameter Ns*/N, which is calculated as follows:

(7)

By substituting the specified flow and differential head for the pumping application (not the impeller) into the equation for specific speed (Eq. 1), i.e. setting QBEP = Q* and TDHBEP = TDH*pump, the resulting parameter Ns*/N can be quickly calculated. The following are general rules of thumb for assuming what type of centrifugal pump can be used for a given range of Ns*/N (in USC units):

• For Ns*/N > 1.2 o Assume either a double-suction single-stage pump at 1800 RPM or less, or a single-stage overhung pump at

1200 RPM or less.

• For 0.2 < Ns*/N < 1.2 o Assume a single-stage overhung pump at 3600 or 1800 RPM, a single-stage double-suction pump at 3600 RPM

or less, or a two-stage pump at 1800 RPM or less.

• For Ns*/N < 0.2 and TDH*pump < 1200 ft o Assume a single-stage overhung pump at 3600 RPM or a multi-stage pump.

• For Ns*/N < 0.2 and TDH*pump > 1200 ft

o Assume a multi-stage pump, high speed single-stage Barske pump, or other type of high head pump.

Note that the above are general rules of thumb and should not be substituted for common sense. There are plenty of exceptions to the rules (e.g. a two-stage pump at 3600 RPM with Ns*/N > 0.2). The above recommendations will be compared to real pump selections in the next section of this tutorial. Once a reasonable impeller diameter, reasonable Ns*, and reasonable rotational speed have been determined, the rotational speed can be checked in Equation 4 to ensure that a reasonable value of NPSHR has been established. The reader might wonder what constitutes a reasonable NPSHR, and therefore a reasonable NPSHA. The answer is that it depends on the application. For pumps taking suction from atmospheric tankage or from open bodies of water, the pressure of the atmosphere will add about 34 feet of suction head to a liquid with SG = 1.0 at sea level. On the other hand, for a service such as a reflux pump which is pumping a liquid near its bubble point, the NPSHA is greatly dependent on the elevation of the suction vessel relative to the elevation of the pump and the specified minimum NPSH margin. In general, discussions should be held among facility design engineers, the purchasing engineer, a senior facility operator, and the end user (which may include the facility design engineer, purchasing engineer, and/or the operator) to determine what a reasonable NPSHA should be. This will almost certainly be an iterative discussion, weighing the pros and cons of elevating vessels or increasing minimum vessel operating levels against purchasing double suction pumps, applying a slower pump with a larger impeller diameter, and/or adding more stages. As a general rule of thumb, the price of a motor-driven overhung pump will increase by about 50% for every doubling of the impeller diameter. For an already expensive pump, this may be a relatively high expense compared with a few more feet of structural steel and pipe to elevate the suction vessel to achieve a suitably high NPSHA for the smaller, higher speed pump.


EMPIRICAL DATA AND VALIDATION OF APPROXIMATIONS The approximations laid forth in the previous section may seem reasonable at first glance, but without real empirical evidence to support these approximations, the reader cannot be assured that they will not lead to more problems than solutions. Two recent greenfield projects with more than 150 real pump applications and the offered pumps for those applications are now reviewed to assess the validity of these approximations regarding prediction of pump type, operating speed, and NPSHR. First, plots of the relative ranges set forth for the values of Ns*/N versus the quoted pump speeds can be seen in Figures 3 and 4, respectively. From the plotted data, it can be seen that the range of values of Ns*/N for each pump type coincides fairly well with the recommended ranges from the previous section. It can also be seen upon close examination that there are exceptions to those recommendations. Recall that the recommendations from the previous section are to be used as a starting point for the assumed pump type, and exceptions to the recommendations are to be expected from time to time.

Figure 3: Ns*/N versus Speed for TDH*pump < 1200 ft

Figure 4: Ns*/N versus Speed for TDH*pump > 1200 ft


Figure 5 below shows how Q* relates to QBEP for both the maximum diameter impeller and for the furnished impeller. Review of the distributions shows that they center around Q* = 100% BEP for the furnished impeller and Q* = 90% BEP for the max diameter impeller. Based on this data, it can be inferred that selection of pumps with Q* considerably far from QBEP may result in the given Ns*/N ranges being skewed.

Figure 5: Count of Pumps by Q* as %BEP

The next and most important assumption to verify is the approximation of using impeller tip head as the pump’s actual differential head. Figure 6 shows how Ns* impacts the accuracy of the impeller diameter prediction for single-stage overhung pumps, not including Barske designs. As can be seen from the distributions for each range of Ns*, the calculation as outlined in the previous section seems to slightly overpredict the impeller diameter on average at very low values of Ns*, while at higher values of Ns*, the calculation seems to slightly underpredict the impeller diameter on average.

Figure 6: Impact of Ns* on Accuracy of Predicted Impeller Diameter, Single-Stage Overhung Fully Closed Impellers


Figure 7 shows the impact of Ns* on predicted impeller diameter for Barske impeller pumps. As can be seen, the approximation overpredicts impeller diameter for Barke designs at a significantly higher rate than standard impeller designs.

Figure 7: Impact of Ns* on Accuracy of Predicted Impeller Diameter, Single-Stage Overhung Barske Impellers

Further evaluation of the approximation of pump head can be seen in Figures 8 and 9, where for a given range of Ns*, the effect of Q* relative to the BEP flow for the furnished impeller is examined. The figures indicate that there is no consistent impact of Q* as a percentage of BEP of the furnished impeller on the prediction of impeller diameter. This is likely due to the fact that most pumps for the referenced greenfield projects were quoted with the impeller diameter somewhere between 80% and 120% of BEP flow for the furnished impeller (see Figure 5). Accepting pump proposals with Q* outside these ranges would likely lead to further departure of the predicted diameter from the actual diameter, with pumps whose rated points are further back on the curve having actual impeller diameters smaller than predicted, and with pumps whose rated points are further out on the curve having actual impeller diameters larger than predicted. This follows from Equations 5 and 6 since at higher flows on a given pump curve, TDH reduces, which results in lower values for U per Equation 5 and hence smaller values for D in Equation 6. Similarly, at lower flows on a given pump curve TDH increases, which results in higher values for U per Equation 5 and hence larger values for D in Equation 6.

Figure 8: Impact of Q* as %BEP of Furnished Impeller on Accuracy of Predicted Impeller Diameter, Ns*=400 per impeller


Figure 9: Impact of Q* as %BEP of Furnished Impeller on Accuracy of Predicted Impeller Diameter, Ns*=700 per impeller

Now that the Ns*/N correlation to pump type has been shown to be suitable, and now that the prediction of impeller diameter has been shown to be sufficiently accurate using the assumption that impeller tip head is equal to the service head, it can be assumed that pump rotational speed can be reasonably predicted. The final question to be answered is whether or not the assumption of Nss = 9000 USC in conjunction with the assumption that Q = Q* predicts NPSHR with reasonable accuracy. Figure 10 plots the ratio of the actual NPSHR (i.e. the manufacturer’s quoted NPSHR) to the predicted NPSHR against the actual pump Nss. Examination of Figure 10 shows that for values of actual pump Nss equal to or greater than 9000, the actual NPSHR is on average a little less than the predicted NPSHR, which means the resulting prediction is a conservative one for Nss > 9000. Furthermore, the dashed line indicates the predicted NPSHR value for any given service over a range of Nss using Equation 8 with Nss1 = 9000, which is derived from Equation 3 using the assumption that the product N√QBEP is constant. Review of the plotted points relative to this line indicates that even for actual values of Nss greater or less than 9000, the actual values of NPSHR are typically less than what is predicted.

(8)

There are some points in Figure 10 where, even when compensated for varying Nss using Equation 8, the predicted NPSHR is still less than the actual NPSHR of the quoted pump, resulting in a value that falls above the dashed line. Further examination of the data reveals that this generally happens when the rated flow is significantly higher than the BEP flow for the furnished impeller or when the furnished impeller is heavily trimmed relative to the maximum impeller diameter. See Figure 11.

Figure 10: Ratio of Actual Pump NPSHR to Predicted NPSHR versus Actual Pump Nss


Figure 11: Impeller Trim and % BEP Flow for Quoted Pumps where Actual NPSHR Exceeds Predicted NPSHR by ≥ 10%

For pumps with Nss < 9000 USC, the predicted NPSHR is on average less than the actual NPSHR. For these pump applications, this may present a real problem due to insufficient NPSH margin. However, review of the NPSH margin relative to the predicted NPSHR for the pump applications as shown in Figure 12 shows that for these services, there seems to be sufficient NPSH margin. This would indicate that for these services, NPSHA is quite high. It is likely, then, that pump manufacturers quoted lower Nss pumps with higher NPSHR for these applications because there was ample NPSHA to allow the use of such a pump.

Figure 12: Margin Between Actual NPSHA and Predicted NPSHR versus Actual Pump Nss


NPSHR, NPSHA, AND NPSH MARGIN – PRACTICAL CONSIDERATIONS There have been many industry papers and standards (such as ANSI/HI 9.6.1-2017 Rotodynamic Pumps – Guideline for NPSH Margin [7]) published about how to determine a suitable NPSH margin, and most major end user companies have requirements pertaining to NPSH margin, suction specific speed, NPSH testing, etc. No reputable industry professional, however, would consider operating a pump with zero or negative NPSH margin an acceptable practice. Whatever the purchaser’s requirements may be for NPSH margin, testing, and the like, there are practical considerations that should be taken into account during facility design, pump purchase, and pump testing to prevent the end user from ultimately having insufficient NPSH margin, regardless of the predictive methods outlined previously in this tutorial. Setting NPSHA Calculation Datums The first and foremost consideration when designing for sufficient NPSH margin should be a conservative and reasonable calculation of NPSHA. When calculating NPSHA, the datums used to calculate the elevation difference between the fluid source and the pump suction must be chosen carefully. For the pump, the reader might think this is a fairly simple task, as most industry specifications require that the datum used for a pump’s NPSHR is the centerline of the impeller (except VS6 and VS7 type pumps, which will be addressed later), so one of the datums used to calculate NPSHA should match that datum. However, it is not possible to know exactly how high above grade the centerline of a pump impeller will be before the pump has been purchased, a dimensional drawing including the mounting plate has been provided by the vendor, and the foundation has been designed. A rudimentary way to predict the impeller centerline elevation is by use of the specified flow for the pump. For larger ranges of flow, the purchaser might assume that the centerline height of the suction impeller is higher than that for smaller pumps. That may be true for a large percentage of applications, and with no additional data with which to work, this may be a suitable assumption. Pump pedestal heights can then be tailored later to match the centerline of the pump to the assumed centerline. The facility design engineer should be diligent in choosing an assumed elevation: too low of an assumption may result in the height from bottom of baseplate to pump centerline being greater than the assumed centerline elevation, while too high of an assumption may render the pump too high to be readily accessible once installed. An alternate assumption would be to use the methods described earlier in this tutorial to calculate impeller diameter in conjunction with some nominal assumptions for foundation height, baseplate thickness, and pedestal height to calculate a more accurate expected elevation of the impeller centerline. Additionally, the facility design engineer should add a nominal value to the assumed foundation height in the instance that the pump will require an external oil circulation system, since increased pump elevation will be required to allow the lube oil to drain back to the oil reservoir. Platforms around the perimeter of the pump pedestal may be required in these instances to allow access to the pump for regular monitoring and operation. The other datum used to calculate NPSHA is the assumed fluid level in the suction vessel. For open tanks or sumps, the lowest allowable operating liquid level should be used. This is often determined by a process engineer or operations specialist. For pressure vessels, the bottom of the vessel is typically used as the datum for horizontal vessels, and the bottom tangent is typically used as the datum for vertical vessels. While these datums may seem unreasonable at first glance, the reader would do well not to increase the height of the datum due to potential unforeseen circumstances during detailed plant design. The previously listed datums should be chosen to allow the user some level of recovery measure in the event that for whatever reason, the NPSHA is reduced as the plant design develops. For example, in the event that the lateral distance of a pump from the suction vessel increases significantly during the development of the plot plan for the facility, thereby increasing friction losses in the suction line, and if increasing the elevation of the vessel is not possible (perhaps, for example, it is an extremely large suction vessel), then the datum can be adjusted inside the suction vessel. In this instance for large tanks at ground level, the lowest operating liquid level can be elevated, and for vessels, the low liquid level might be used rather than the bottom of the vessel or the bottom tangent. Such adjustments should be made with scrutiny but do allow the end user some means to ensure sufficient NPSH margin in light of facility design changes. Assumed Friction Losses in the Suction Piping With suction elevation head established, with a known vapor pressure, and with the suction vessel operating pressure known, the only other major element to consider is friction losses in the suction piping as the liquid flows from the suction vessel to the pump. Obviously it is difficult to calculate this when the final dimensional layout for the facility piping, fittings, and valves is not known.


Fortunately, the design of the facility at this point in the project should be sufficiently detailed to know the difference in elevation between the suction vessel and the pump as well as the approximate lateral distance from the suction vessel to the pump. This allows for a suitably conservative approximation of the suction piping losses. The recommended procedure is as follows: Sum the vertical and horizontal distances from the suction vessel. Note that pipe may not be able to be routed in a straight line from the vessel to the pump, so the pipe lengths should be calculated based on reasonable routing of piping along available support structures. Once this distance has been calculated for each given pipe size, multiply the distance by a nominal factor to account for potential future routing changes as well as the impact of valves and fittings. The author suggests a factor of 1.5. This factor might be reduced if the facility design engineer knows the exact type and number of valves and fittings in the suction piping, but it is not recommended to reduce the factor to 1.0. Next, calculate the losses assuming the highest specified flowrate and highest specified viscosity, and then multiply by an additional factor to account for corrosion or fouling of the pipe if that is expected for the pumping service in question. A commonly used factor in industry to account for corrosion and fouling is 1.2. As always, the NPSHA should be recalculated after the exact pipe routing, size, and number of valves and fittings is known to confirm that the NPSH margin is in fact adequate. Pump Certified Points and NPSHR Considerations Pumps are quite often specified with a Rated and Normal operating point with the Rated operating point flow being some nominal percentage higher than the Normal operating point, typically 10%. The Normal point is the point where the pump is expected to operate during normal operation of the facility, while the Rated point is the point for which the pump is certified, and occasionally may correspond to some infrequent or alternate mode of facility operation. In the preferred operating range of the vast majority of pumps, and typically throughout the entire allowable operating range, the NPSHR for the pump decreases as the flowrate decreases. Therefore, the NPSHR at the Normal flow is typically less than the NPSHR at the rated flow. Additionally, the NPSHA should be slightly higher at the Normal flow relative to the Rated flow as a result of slightly less friction loss in the suction line of the pump at the lower flowrate. Therefore, specifying a Rated flow that is some percentage higher than the Normal flow allows for some additional NPSH margin when the pump is operating at the Normal point. Quoted Pump Speeds versus Actual Pump Speeds The purchaser should be wary when reviewing pump curves and the speeds at which those pump curves are published. Pump curves are sometimes published at speeds that would require a relatively high amount of motor slip. When the final motor torque curves are compared against the pump load curve, the purchaser will often find that the pump will actually run at a speed higher than what the published curve indicates. This is generally advantageous for the pump vendor, as it allows them to more assuredly guarantee the flow-head performance of the pump. However, this ultimately will result in an increased NPSHR, and the purchaser should take note of this before purchasing the pump. Application of Minimum Flow Lines with Restriction Orifices In many instances, facility design criteria will require that some if not all pumps in the facility have recirculation piping incorporated into the design of the pump piping to prevent the pump from operating at a flow below the minimum continuous stable flow (MCSF). These are generally referred to as “minimum flow lines”. The minimum flow is usually controlled by either a restriction orifice or a control valve, and the choice between one or the other is typically an economical one. Minimum flow lines with restriction orifices constantly recycle flow while the pump is in operation, which results in wasted power consumption. If a pump is large enough, it becomes more economical to apply a control valve rather than an orifice, as the control valve will remain closed as long as the pump’s flow exceeds the MCSF. This results in zero wasted energy, but the control valve costs more to purchase and install than a simple orifice. The facility design engineer must weigh the cost of the control valve against the wasted energy from continuous recirculation and design their facility accordingly. There may also be other factors to consider, such as regional, state, or federal regulations pertaining to pumping efficiency and energy consumption. Note that the minimum flow line is usually installed to recirculate the flow back to the suction vessel to allow the recycle flow to mix with a large volume of cooler liquid to help dissipate the heat of pumping as well as to allow disengagement of any fluid that may have flashed across the orifice. See Figure 13 for a typical schematic. Prior to a pump being purchased, the minimum flow is not known, so oftentimes when quoting a pump the supplier is instructed to add the required minimum flow to the purchasing engineer’s specified rated flow at the specified differential head to determine the quoted


rated flow for the pump. However, this higher rated flow for the quoted pump will result in increased friction losses in the suction piping if the recirculated flow is returned to the suction vessel, meaning the NPSHA will decrease. Futhermore, if a maximum Nss is specified, the increase in flow may result in a higher NPSHR than predicted in order to keep the Nss down. This implies that before purchase, the NPSHA should be recalculated. Unfortunately, for a large number of pumps in systems with minimum flow lines and with bids from multiple different suppliers, this can be a considerable task and could significantly delay purchase of the pumps. As a result, recalculation of NPSHA often will not be performed until after pumps are purchased due to rigid work processes or failure to identify this potential issue at the time of purchase.

Figure 13: Typical Minimum Flow Line Arrangement

A simple method to address this issue is to incorporate into the specified rated flow a nominal assumption for what the minimum flow for the purchased pump will be. Since the applications for which restriction orifice minimum flow lines are applied are typically low horsepower due to the aforementioned considerations, the subject pumps will more often than not be single-stage pumps. The author has found that for single-stage overhung pumps at or below 50 BHP, the assumption that 35% of the pump’s total flow at rated conditions recirculates through the minimum flow line will result in sizing of the minimum flow restriction orifice such that the recirculating flow will meet or exceed the MCSF of the purchased pump if the pump discharge were to be completely blocked. Take note that if this method is used, the restriction orifice should be sized for the assumed flow (35% of total flow through the pump) rather than for the pump’s actual minimum flow. If the orifice is ultimately sized for the pump’s actual minimum flow, which is often less than the assumed 35%, this may result in the real operating point moving back on the pump curve, potentially resulting in less reliable operation. Reduced Speed Testing and Scaling of NPSH Research as to how NPSHR scales with the speed of a centrifugal pump has shown that it does not scale according to the affinity laws [8], but rather, NPSHR scales with speed as follows:

(9)

Note that in Equation 9, 1 ≤ α ≤ 2, as opposed to the affinity law used to scale differential head, where the value of the exponent is 2. The conservative and proper method for reduced speed NPSH testing is as follows: The pump is NPSH tested at reduced speed using cold water, after which the NPSHR value or curve from the reduced speed test is scaled to the quoted operating speed using α = 2. The scaled value or curve should then be compared against the contractually guaranteed value, with any positive margin constituting a failure.


This implies that on the test stand, if reduced speed testing is required due to limitations of the test stand flow capacity, pressure rating, driver power, etc., and any value less than 2 is used for α for the purposes of scaling up the NPSHR, the supplier or the test facility (often one and the same) should be made to provide empirical evidence to prove that the value used for α is legitimate. A caution here is that if the value of α is known to be significantly below 2, mandating the use of α = 2 will result in a failure of the test. In this instance, the supplier and the purchaser should agree on what value of α is to be used, and as previously stated the supplier should be made to prove the legitimacy of the scaling factor with empirical evidence. NPSH Considerations for Specific Pump Types For vertically suspended pumps, there are multiple considerations regarding NPSH that the purchasing engineer should keep in mind when specifying a pump. First, the datum used for referencing NPSHR and NPSHA at the pump should be agreed upon by both the supplier and the purchaser. As previously discussed, this is usually the centerline of the first impeller of the pump. However, for vertically suspended “canned” pumps (denoted VS6 and VS7 type pumps), this is a less than ideal location, as the friction losses from the suction nozzle to the centerline of the first impeller are not readily calculable by the purchaser. It is for this reason that API 610 11th Edition [4] specifies that for VS6 and VS7 pumps, the NPSHR datum for these pumps is to be at the top of the foundation. This provides a consistent elevation at which both the purchaser and supplier can clearly communicate the NPSHA and NPSHR without any confusion as to significant friction losses inside the pump. Most other vertically suspended pumps are typically installed in atmospheric sumps or mounted inside pressure vessels or tanks. For those applied in atmospheric sump service (including vertical cooling water pumps, stormwater lift pumps, etc.), atmospheric pressure can contribute a substantial amount of NPSHA. For water at mean sea level, this is almost 34 feet of suction head. In these instances, NPSH margin may be of little interest since performance loss due to vortexing of the fluid surface and ingestion of air as a result of insufficient submergence usually occurs well before NPSH margin becomes insufficient. For vertically suspended pumps mounted in pressurized tanks or vessels, however, there is no contribution by atmospheric pressure to NPSHA. In these instances, the purchasing engineer would do well to confirm which demands a higher operating level on the vessel, NPSH margin or minimum submergence, and then set the low limit for vessel’s liquid operating level accordingly. For further information on proper pump submergence, the author recommends the facility design engineer reads one of the many available texts and papers written on pump intake design and calculation of associated parameters such as Froude number. Axial flow pumps fall outside the calculation methods and statistical analyses from the first and second parts of this tutorial since the mechanism for head generation in these pumps relies almost solely on the change of flow angle rather than the change in tangential velocity. Fortunately, the rule of thumb for predicting NPSH for these pumps is straightforward: Assume that for every foot of TDH the axial flow pump produces, the pump will need one foot of NPSHA to have a reasonable NPSH margin (e.g. ~25% of NPSHR). Reciprocating pumps, which include piston pumps, plunger pumps, and diaphragm pumps, require additional considerations for head losses in the suction piping due to the pulsating nature of the flow. These losses are called acceleration head losses and result from the acceleration of the fluid that occurs with each stroke of the pump. There are several industry standards that provide calculation methods for these losses, and it is recommended that the facility design engineer in conjunction with the purchasing engineer perform these calculations to approximate acceleration head losses prior to purchasing a pump. It should be noted that these calculations depend on the number of heads on the subject pump and whether or not the pump is single- or double-acting. This implies that a pump must be proposed by a manufacturer prior to performing these calculations, or that the purchasing engineer specify the number of heads on the pump and whether it is single- or double-acting. A conservative assumption to enable calculations prior to having specific pump details would be that the pump will have one head and will be single-acting, although for relatively high flows this may be impractical or may result in unmanageable suction head losses. Design considerations to minimize the impact of acceleration head losses include minimizing the length of piping from suction vessel to pump suction, eliminating any significant restrictions in the suction line, increasing the number of heads on the subject pump, and application of pulsation dampeners. Whatever the strategy, ensure that acceleration head losses are included in the ultimate NPSHA calculation. It is also recommended that the purchasing engineer require a pulsation analysis of the piping system be performed for any reciprocating pump whose brake power exceeds 50 HP. API 675 3rd Edition Annex F [9] provides more information on these analyses.


Acceleration Head in Centrifugal Pump Systems Unlike reciprocating pump applications, centrifugal pump applications have zero acceleration head to consider during normal operation. However, during startup of the pump, the fluid in the suction and discharge piping must be accelerated to steady state velocity. Depending on the magnitude of the acceleration of the liquid, the drop in pressure of the liquid in the suction piping may be of sufficient magnitude to cause the suction fluid to cavitate, potentially even to the point where the pump fails to pump the fluid due to ingestion of a large pocket of gas. The head losses due to acceleration can be calculated as follows:

(10)

For pumps with extremely short starting times, or for pumps with a long run of suction piping from the suction vessel, the facility design engineer might consider reviewing acceleration head for the subject pump application. The only real option for addressing this issue after installation without modifying the installation is starting the pump against a blocked discharge and then slowly opening the discharge valve. This can present a challenge for installations where pumps are designed to start automatically. Using Pressurized Gas to Increase NPSH In some instances, the facility design engineer may find that a seemingly simple solution to insufficient NPSHA is to use a source of pressurized gas to artificially increase the pressure of the suction vessel. However, depending on the pressure required and the fluid temperature, the use of pressurized gas may result in the liquid dissolving some of the gas. Liquids with dissolved gases significantly complicate the issue of calculating NPSHA. Like a carbonated canned beverage that is shaken too much prior to opening, the shearing of the liquid by the impeller as the gas-laden liquid enters the pump will cause the liquid to “fizz”, releasing the dissolved gas. This will cause the pump to lose performance similar to insufficient NPSHA. The difference between this phenomenon and typical cavitation is that the gas does not immediately return to liquid form (or in this case, does not immediately dissolve back into the liquid) as it does with cavitation. As a result, entrained gases may travel a significant distance downstream of the pump or may even quickly collect in the impeller eye in significant enough volume to render the pump’s performance useless. With the additional pressure in the suction vessel provided by the pressurized gas, the vessel’s operating pressure becomes higher than the vapor pressure of the liquid. However, the pressure at which the dissolved gas releases from the liquid is at some pressure higher than the vapor pressure of the liquid. This makes difficult the determination of the effective vapor pressure of the liquid to be used in the calculation of NPSHA. The most conservative approach to calculating NPSHA for such services is to assume that the operating pressure of the suction vessel is the vapor pressure of the liquid, but this assumption defeats the purpose of applying the pressurized gas altogether. If the facility design engineer wishes to better understand how to calculate this, the author directs them to the explanation by C.C. Chen [10]. The facility design engineer should take caution when performing these calculations, taking into consideration all reasonably plausible operating scenarios in order to avoid loss of pump operation due to significant amounts of gas fizzing out of solution in the pump. NPSH Reduction Factors NPSHR for a centrifugal pump is typically defined as the NPSHA value at which the pump’s differential head when pumping cold water at a given flow is reduced by a certain amount, typically 3% loss in head. In reality, the head for the pump is not reduced at the outlet of the impeller, but rather the average density of the fluid is reduced due to the bubbles in the liquid that result from vaporization of the liquid at the impeller inlet. Since the density used to calculate the differential head from pressure gauge readings is higher than the average density at the tip of the impeller, the resulting calculated differential head is lower than the real differential head at the impeller outlet. Practically speaking, the differential pressure across the pump is the performance parameter that really matters, so any drop in that performance is obviously of interest to the end user.

(11)


Water is used as the basis for NPSHR since it is relatively easy to use water as a real test fluid. An additional benefit of basing these tests on water is that the difference in density between liquid water and gaseous water at temperatures near ambient is quite substantial. For a great many fluids besides water, the ratio of the densities of the liquid phase to the gas phase at operating temperatures is considerably less than that of water. Therefore, for pumps pumping these other fluids where NPSHA is equal to the NPSHR for water, the perceived loss in differential head will be less. Another way to say this is that for fluids other than water where the ratio of the densities of the liquid phase to the gas phase is less than that of ambient temperature water, the NPSHR value will be lower than the NPSHR for water.

Figure 14: NPSH Reduction Factors

(courtesy of the Hydraulic Institute [13], www.pumps.org) Per Figure 14, the applied NPSHR value for the pump can be reduced for operation on the indicated pure fluids. Caution should be used when applying these reduction factors, as they are for pure fluids only. Any small amount of impurities in the fluid may result in a higher NPSHR than what the chart indicates. For example, a pump pumping a mixture of 90% propane and 10% butane will lose 3% head at an NPSH value higher than if it were pumping 100% propane. Therefore, if the NPSHR reduction factor for 100% propane is used for the application where the pump is pumping the impure propane mixture, then the applied NPSH margin based on 100% propane may not be sufficient to prevent severe cavitation and performance loss when pumping the impure propane mixture. It is the author’s recommendation that the end user only use the above chart to give themselves some sense of reassurance for applications where the NPSH margin based on cold water is less than they would normally be comfortable with. CONCLUSIONS NPSH is an important pump performance parameter that may sometimes not be given sufficient consideration in early design, and it can be complicated by factors such as fluid acceleration, entrained gases, varying liquid compositions, etc. Insufficient NPSHA for pump applications can be detrimental to pump reliability and can result in performance that is insufficient to meet operational needs with few if any options for correction once a facility has been constructed. There are means to accurately and conservatively predict the NPSHR for a given centrifugal pump service prior to ever reviewing actual pumps for the given service so long as the specified flow for the service ends up relatively close to the BEP of the furnished pump impeller and so long as the impeller is not too heavily trimmed. Early identification of potential NPSH issues during facility design using empirically proven predictive methods, applying sufficiently reasonable and conservative facility design practices, and assuring proper NPSH testing in conjunction with proper pump selection will contribute to a functional pump installation with adequate NPSH margin.

http://www.pumps.org/


NOMENCLATURE BEP = Best Efficiency Point C = 720 (in.-seconds / feet-minutes) D = Impeller outer diameter (inches) g = Acceleration due to gravity (32.2 ft/s2) GPM = Gallons per minute hAccel = Head loss due to acceleration (ft) hLiquid = Total head (specific energy) in a liquid, including elevation head, pressure head, and velocity head (ft) hVapor Pressure = Liquid head at a given temperature at which the liquid will begin to vaporize (ft) Δh = Calculated differential head (ft) L = Length of suction piping (ft) N = Rotational speed in units (RPM) NPSH = Net Positive Suction Head (ft) NPSHR = Net Positive Suction Head required for 3% reduction in differential head (ft) NPSHRBEP = Net Positive Suction Head at BEP for max impeller diameter (ft) NPSHR1 = NPSHR at speed N1 (ft) NPSHR2 = NPSHR at speed N2 (ft) NS = Specific Speed (unitless, calculated in USC units) NS* = Specific Speed calculated using parameters for the pump service and not the pump itself (unitless, USC units) NS*/N = Pump selection parameter (“Sheets Parameter”, unitless, calculated in USC units) Nss = Suction Specific Speed (unitless, calculated with USC units) Δp = Measured differential pressure (psi) Q = Volumetric flow (GPM) Q* = Specified volumetric flow for a given pump application (GPM) QBEP = Flow at BEP for the maximum diameter impeller (GPM) SG = Specific gravity of the liquid (unitless, relative to water at standard conditions) t = Pump starting time (seconds) TDH = Total differential head (ft) TDHBEP = Total differential head at BEP for the maximum impeller diameter (ft) TDH* = Specified total differential head for a given pump application (ft) U = Impeller tip speed (ft/s) V2 = Steady state velocity (ft/s) V1 = Starting velocity (ft/s) α = Unitless constant for scaling NPSHR with speed (1 ≤ α ≤ 2) ψ = Centrifugal pump head coefficient (unitless) REFERENCES

1. Hallam, J.L., 1982, Centrifugal Pumps: Which Suction Specific Speeds Are Acceptable, Hydrocarbon Processing Magazine, Volume 61:4.

2. Lobanoff, Val S., and Ross, Robert R., 1992, Centrifugal Pumps: Design & Application (2nd Edition), Gulf Publishing Company.

3. Cowan, D., Liebner, T., and Bradshaw, S., 2013, Influence of Impeller Suction Specific Speed on Vibration Performance, Proceedings of the Twenty-Ninth International Pump Users Symposium.

4. API Standard 610, Eleventh Edition, 2010 / ISO 13709:2009, Centrifugal Pumps for Petroleum, Petrochemical and Natural Gas Industries, American Petroleum Institute, Washington, D. C.

5. Sulzer Pumps Ltd., Winterthur, Switzerland, 1998, Sulzer Centrifugal Pump Handbook (Second Edition), Oxford, UK: Elsevier.

6. ANSI/HI 14.1-14.2-2019, Rotodynamic Centrifugal Pumps for Nomenclature and Definitions, Hydraulic Institute, Parsippany, New Jersey

7. ANSI/HI 9.6.1-2017, Rotodynamic Pumps – Guideline for NPSH Margin, Hydraulic Institute, Parsippany, New Jersey 8. Schiavello, B., and Visser, F., 2009, Pump Cavitation – Various NPSHR Criteria, NPSHA Margins, and Impeller Life

Expectancy, Proceedings of the Twenty-Fifth International Pump Users Symposium. 9. API Standard 675, Third Edition, 2012, Positive Displacement Pumps – Controlled Volume for Petroleum, Chemical, and

Gas Industry Services, American Petroleum Institute, Washington, D.C. 10. Chen, C.C., 1993, Cope With Dissolved Gases in Pump Calculations, Chemical Engineering Magazine, Volume 100:10


FIGURES

11. “Figure 14.1.4.1 – General Impeller Types”, from ANSI/HI 14.1-14.2-2019 12. “Figure 3: Stable window according to Lobanoff & Ross”, from Proceedings of the Twenty-Ninth International Pump Users

Symposium, Influence of Impeller Suction Specific Speed on Vibration Performance by Cowan, D., Liebner, T., and Bradshaw, S.

13. “Figure 14.3.4.10.5b – NPSHR Reduction for Pumps Handling Hydrocarbon Liquids and High-Temperature Water (US Customary Units)”, ANSI/HI 14.3-2019

ACKNOWLEDGEMENTS

The author would like to acknowledge John P. Joseph II and Todd R. Monroe for their unrelenting quest to mentor the next generation of rotating machinery engineers. A great deal of the knowledge in this paper is due solely to their tutelage. The author would also like to acknowledge Ron Adams for his always timely and insightful correspondence pertaining to the highly technical details of centrifugal pumps. Lastly, the author would like to acknowledge Adam Sheets for his insights regarding the parameter Ns*/N.

DISCLAIMER All information (“Information”) contained herein is provided without compensation and is intended to be general in nature. You should not rely on it in making any decision. LyondellBasell/Equistar Chemicals accepts no responsibility for results obtained by the application of this Information, and disclaims liability for all damages, including without limitation, direct, indirect, incidental, consequential, special, exemplary or punitive damages, alleged to have been caused by or in connection with the use of this Information. LyondellBasell/Equistar Chemicals disclaims all warranties, including, but not limited to, the implied warranties of merchantability and fitness for a particular purpose, that might arise in connection with this information.

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