API 610 notes.pdf

8/14/2019 API 610 notes.pdf

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Updates in Progress for ISO 13709/API 610, Centrifugal pumps forpetroleum, petrochemical and natural gas industries

byRoger Jones

ConsultantSpring, Texas

ABSTRACT:

This tutorial discusses the status of ISO 13709/API 610 and three lesser and three majorissues and/or revisions to the standard. The lesser issues are: NACE Requirements,casing gasket requirements and seal gland plate connections. The major issues are ShaftFlexibility Index, Bearing System Life and Performance Testing.

PUBLICATION STATUS

American Petroleum Institute (API) Standard 610, Centrifugal Pumps for Petroleum,Petrochemical and Natural Gas Industries and International Standards Organization (ISO)Standard 13709 (with the same title) are identical standards. For decades API 610 has

been a de facto international standard for refinery pumps. But, it hasnt officially beenan international standard. This lack of official status has created difficulties forinternational oil companies investing outside of North America. The API has thereforedeveloped a strategy of co-branding certain standards such that the API standards

become ISO, bona fide international standards. API 610 is among those in this program.

The ninth Edition of API 610 was published in January of 2003. At the same time theFirst Edition of ISO 13709 was going through the ballot process. The ISO ballot processinvolves at least two and sometimes three drafts and ballots. The first is the CommitteeDraft or CD. This draft is offered for comment and ballot. If it passes the WorkingGroup resolves all comments, revises the draft and moves it to the next stage. The nextstage is Draft International Specification or DIS. For mature specifications such as API610 the CD stage can be skipped. This is exactly what was done with the 9 th Edition.

When API 610 9 th Edition went through the DIS review and ballot a large number ofmainly editorial comments we submitted. These were duly resolved and the draft wasrevised. There was essentially no significant technical change to the standard. The

document was moved to the Final Draft International Standard ballot level. The FDIS ballot is a yes/no ballot with no technical comments allowed. However non-technical/editorially comments are allowed. API 610 9 th Edition passed FDISunanimously.

However, there was now a problem. In order to adopt ISO 13709 back as API 610 9 th Edition an annex would have to be added identifying every editorial change in thestandard. There were more than 100 of these changes. The only way that ISO 13709


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there wasnt any other standard that could be used for these downstream services. Theauthor is unaware of this ever causing an issue and MR0175 has served our industry well.

In 2002 NACE updated MR0175. This update expanded the scope of MR0175 to alsodeal with Chloride Stress Corrosion Cracking. These changes made MR0175 even less

applicable to downstream applications. NACE then decided to create a new standarddirected specifically at downstream applications. The result was a new standard MR0103 which was published in 2003 shortly after MR0175-2003 was published. The majorimpact of the new standard is that there are numerous, more rigorous requirements forwelding on carbon steel. In fact the bulk of the text is devoted to welding requirements.If you want more detail on the content of MR 0103 and how it compares to MR0175, youwill find An Overview of NACE International Standard MRO103 and comparison withMR0175 by Bush, Brown and Lewis (Corrosion 2004, Paper No. 04649 a good read.

There is a second issue with the application of NACE to pump materials. This issue wasdealt with in the 9 th and 10 th Editions however, questions still arise so the WGTF is

making further revisions in an effort to attain clarity. The issue is that purchasers specify NACE materials. The pump vendor then looks at NACE and sees that it does not requirereduced hardness materials. The pump vendor supplies standard materials which complywith NACE. Did the purchaser get what he wanted?

The author surveyed all user and contractor API members who attended the Fall Refinerymeeting in 2001, the responses we almost evenly divided. Half of the respondents feltthat specification of NACE meant that reduced hardness materials were required. Theother half believed that specification of NACE resulted in the pump supplier looking atthe H2S and water content of the service, determining whether reduced hardnessmaterials were required and then either supplying them or not.

The fact that there is ambiguity in requiring NACE turns this into a commercial issue.For example one manufacturer believes that specification of NACE requires reducedhardness materials. He supplies them and adds the cost into the price of his pump. Asecond manufacturer, who might have a higher threshold of risk, looks at the service andsupplies standard materials, The second manufacturers pump has a lower price and hewins the bid.

The current draft of API 610 contains the following paragraphs to make this issue crystalclear:

6.12.1.x0 The purchaser shall specify the amount of wet H2S that may be present, consideringnormal operation, start-up, shutdown, idle standby, upsets, or unusual operating conditions suchas catalyst regeneration.

Note: In many applications, small amounts of wet H 2S are sufficient to require materialsresistant to sulfide stress-corrosion cracking. If there are trace quantities of wet H 2Sknown to be present or if there is any uncertainty about the amount of wet H 2S that may

be present, the purchaser should consider specifying that reduced hardness materials arerequired.


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6.12.1.x1 The purchaser shall specify if reduced hardness materials are required.

6.12.1.x2 If reduced hardness materials are specified in 6.12.1.x1 they shall be supplied inaccordance with NACE MR0103.

Note: NACE MR0103 applied to oil refineries, LNG plants and chemical plants. NACEMR0103 applies to materials potentially subject to sulphide stress corrosion cracking.

6.12.1.x3 If specified reduced hardness materials shall be supplied in accordance with ISO15156 (NACE MR0175).

Note: ISO 15156 applies to oil and gas production facilities and natural gas sweetening plants. NACE MR0175 is equivalent to ISO 15156. ISO 15156 applies to material potentially subject to sulphide and chloride stress corrosion cracking.

6.12.1.x4 If reduced hardness materials are specified, ferrous material not covered by MR0103 or ISO 15156 (NACE MR0175) shall have a yield strength not exceeding620 N/mm2 (90 000 psi) and a harness not exceeding HRC 22. Components that arefabricated by welding shall be post weld heat-treated, if required, so that both the weldsand heat-affected zones meet the yield strength and hardness requirements.

6.12.1.x5 If reduced hardness materials are specified the following components shallhave reduced hardness:

1) the pressure casing;

2) shafting (including wetted shaft nuts);

3) pressure-retaining mechanical seal components (excluding seal faces);

4) wetted bolting;

5) bowls.

Double-casing pump inner casing parts that are in compression, such as diffusers, are notconsidered pressure casing parts.

6.12.1.x6 Renewable impeller wear rings that must be through-hardened aboveHRC 22 for proper pump operation shall not be used if reduced hardness materials arespecified. Wear rings may be surface-hardened or coated with a suitable coating. Ifapproved by the purchaser, in lieu of furnishing renewable wear rings, wear surfaces may

be surface-hardened or hardened by the application of a suitable coating.

The paragraphs above use xX notation because the paragraphs in the draft have not been reordered and renumber at the time of this writing.

CASING GASKETS

The Issue of casing gaskets was raised in mid-2006 when a user company discovered thatthey had bought a cryogenic pump which used o-rings on pressure casing joints. This


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rings and a minimum 2 mm (0,08 in) chamfered lead-in for dynamic O-rings. Chamfers shall havea maximum angle of 30.

6.8.3 The seal chamber shall conform to the dimensions shown in Figure 25 and Table 6. Forpumps with flange and pressure ratings in excess of the minimum values in 6.3.5, the gland studsize and circle may increase. Larger studs shall be furnished only if required to meet the stressrequirements of 6.3.4 or to sufficiently compress spiral-wound gaskets in accordance withmanufacturers specifications.

Clause 6.12.1.3 is new in the 11 th Edition. This paragraph says that the materialspecification of the seal chamber joint gasket must be selected in accordance with ISO21049 (API 682).

6.12.1.3 The material specification of all gaskets and O-rings exposed to the pumped fluidshall be identified in the proposal. O-rings shall be selected and their application limited asspecified in ISO 21049.

Table H-1 and Note g are unchanged from the 10 th Edition.

g If pumps with axially split casings are furnished, a sheet gasket suitable for the service is acceptable. Spiral-wound gaskets shouldcontain a filler material suitable for the service. Gaskets other than spiral wound, may be proposed and furnished if proven suitable forservice and specifically approved by the purchaser.

Having reviewed the current radial case gasket requirements, it should be noted that theWGTF is virtually unanimous in feeling that spiral wound gaskets should be the standardfor radial split casings. It is also noted that despite requests for data that indicates that o-ring should be disallowed, we have no case (no pun intended) for change. Having nocase for change also means the WGTF does not want to see pump manufacturersredesigning all their pressure casings for o-ring sealed radial joints in lieu of the presentstandard spiral wound gaskets. The 11 th Edition will contain a note stating the preferencefor spiral wound gaskets. The current wording, which continues to be debated is:

6.3.10 Radially split casings shall have metal-to-metal fits, with confined controlled-compressiongaskets, such as an O-ring or a spiral wound type.

Note: The materials table H-1 shows only spiral wound gaskets for casing joints. Spiral woundgaskets are preferred because they typically have had better availability, are more conducive topositive materials identification and historically have had higher temperature limits.

As a final note, vertical suspended pumps almost always have o-rings for gaskets oncolumn joints and bowls.

GLAND CONNECTIONS

Virtually no oil or chemical company today will allow screwed joints in process piping.Yet API 610 and API 682 allow a screwed joint between the piping and the seal gland orend plate. The reasons screwed piping is not allowed in refineries and chemical plantsare pretty obvious. The piping is significantly weakened by the cut threads and the jointsleak, if not visibly in the form of drips then invisibly in the form of detectible HC fugitiveemissions (yes sealants can be used). Allowing either form of leakage is not responsiblein todays environmentally sensitive world. So why do we allow it?


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We allow it because it is perceived that there has is no simple way to replace the taperedthread joint on the seal end plate and still be able to assemble and disassembly typical

pumps. The difficulty varies depending on pump types and is probably most difficult forsmall single stage over hung pumps. How long is the industry going to accept this?

The as new screwed joint may or may not have detectible leakage. However the firsttime the joint is broken and remade, it almost certainly will have detectible leakage. Themechanic or pipe fitter now has a choice. He can let it leak or he can tighten it another 90degrees due to the four holes in the flange (because the rest of the piping is welded). Thesituation is better if tubing is allowed but many user companies do not allow tubing onthese (relatively) low pressure product lines (while at the same time allowing tubing in2500 psig hydrogen systems for instrument sensing lines). A possible solution to the 90degree tightening issue is to specify a lap joint flange on the nipple that connects to thegland. The WGTF has found no data to support this reducing fugitive emissions butintuitively we think it would be an improvement. Lap joint flanges are probably an

improvement but not a solution.API 610 has had a possible solution in it since the 8 th Edition.

2.3.3.3 If specified, cylindrical threads conforming to ISO 228, Part 1 may be used. If cylindricalthreads are used, they shall be sealed with a contained face gasket, and the connection bossshall have a machined face suitable for gasket containment (see Figure 2-1).

Figure 1 Reproduction of Figure 2-1 from API 610 8 th Edition

If a component existed that used this internationally accepted joint and a lap joint, thoseusers that specify on hard piping could eliminate all the weaknesses in the tapered thread

joint. In fact since this is a metal to metal joint it is possible (fabrication issues) that a lap


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joint would not be necessary. For those who specify tubing the last significant fugitiveemissions source is eliminated.

The convenor/chairman of the WGTF has strong feelings that the threaded joint to theseal end plate (gland) should be disallowed. This will be proposed in the DIS2.

MAJOR ISSUES

This tutorial has now covered three smaller issues or changes with API 610. The threemajor issues now follow.

SHAFT FLEXIBILITY INDEX

The quantity L 3/D4 has been used to evaluate single stage overhung pumps since the 50s.It came into common usage in the 70s and 80s and there is no question that it, alongwith requirements in API 610, has moved the pump industry into make stiffer more

robust pump designs. Some companies devised bid penalties based upon values ofL3/D4. The higher the value of L 3/D4 the more flexible the pump shaft and the higher the penalty. In all this time API 610 has remained silent on this quantity. In the 11 th Edition,API 610 will address L 3/D4. To begin with it will be called the Shaft Flexibility Index orSFI. A very simple, standardized method of calculating SFI will be set forth in Annex Kas follows:

To meet the requirements of 9.1.1.3, the shaft flexibility index should be calculated as follows(see Figure K.1):

[SFI] = 25.4 (L 3/ D4) in SI units

[SFI] = (L3/ D4 ) in US Customary unitswhere:[SFI] = shaft flexibility indexL = distance from the centerline of the radial bearing to the centerline of theoverhung impeller, mm (in)D = nominal diameter of the shaft between the radial bearing and the overhungimpeller hub, mm (in)

Before settling on the admittedly very simple method of calculating SFI above the WGTFconsidered requiring actual deflection and critical speed calculations. The WGTF alsoconsidered a more complicated Shaft Deflection Factor, SFD that considers the diameter

of the shaft between the bearings. SFD is calculated as follows:SFD = L13/D14 + L1(L2)2/D24

To judge the desirability of the two methods one (manufacturer) taskforce memberlooked at SFI using both definitions. Calculations were performed for a number of linesof pumps complying with various editions of API 610. The results are shown in Figures2 and 3.


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Figure 2. SFI for various vintages of pumps

Figure 3. SFD for various vintages of pumps


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Why has SFI not already been dealt with in API 610? Primarily because it is not adefinitive tool for evaluation of shaft stiffness and because there has been no consensus inthe industry as to what value is acceptable. For example lets examine deflection in atypical single stage overhung pump shown in Figure 4.

Figure 4, Typical Single Stage Overhung Pump Rotor

La Lb

Da Db

X

SEAL FACEBEARING C

W

L

Shaft deflection is caused by impeller weight and unbalanced hydraulic loads. For a shaftwith two major diameters, shaft deflection can be calculated using:

+=b

b

a

aa

I

L

I

L

E

WLY 3

2

max

where:W = radial loadL b = distance between centerlines of bearingsD b = nominal shaft diameter between bearingsE = elastic modulus of shaftIa = moment of inertia for the shaft diameter between impeller centerline and radial

bearingI b = moment of inertia for the shaft diameter between radial and thrust bearings

Table 1 shows four hypothetical pumps all designed to have an SFI of 76.2. Generallyusers of SFI would find this value acceptable. However two of the pumps do not meetthe seal face deflection criteria of API 610. For this reason some pump manufacturersoppose dealing with SFI and have offered the alternative of testing for deflection. A

pump bearing housing would be rigidly mounted in some fixture and a knownstandard weight would be attached to the shaft end. The deflection would be measureddirectly. This deflection value would be furnished in all proposals and would provide adirect way of comparing shaft stiffness in all pumps offered in a particular case. The


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down side of this method is standardization of the fixture and test/measurement method.With hydraulic loads also varying among pump designs and vendors, this still doesntreally give a direct comparison.

Table 1, Comparison of deflection in four hypothetical pumps with equal SFI

La (in) 12.5 12.5 12.5 12.5Da (in) 2.25 2.25 2.25 2.25L b (in) 7.75 10 7.75 10D b (in) 3 3 2.5 2.5X (in) 8.5 8.5 8.5 8.5W (lbs) 250 250 250 250L3/D4 76.2 76.2 76.2 76.2Deflection at seal face (in) 0.0018 0.0019 0.0021 0.0023Deflection at impeller center (in) 0.0057 0.006 0.0067 0.0073% deflection 100 105 117 128

All of this material has been debated in the WGTF for more than two years. Everyone inthe WGTF recognizes the limitations of an SFI comparison. Further there is no realadvantage to using SFD for a comparison. In spite of the weakness of this sort ofanalysis the WGTF received multiple comments on this issue and there is a faction in ourindustry that strongly desires to have this calculation performed to compare various pumpofferings. It is clear that if one line of pumps has disparately high SFI numbers there is agood chance they cannot meet the deflection and dry bending critical requirements ofAPI 610. One the other hand if SFI numbers are disparately low one might suspect themanufacturer is using overly large and expensive seals or he is exercising his creativity inthe use of numbers.

The WGTF therefore took a look at values of SFI for a number of lines of modern (API7th and 10 th Edition Pumps) with the result shown in Figure 5. It was found that if SFI is

plotted as a function of HQ/N (this number is proportional to shaft torque) on log-logscales, the result is a straight line.

Figure 5. SFI for typical modern pumps


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Overhung Pumps - Rotor SFI

0.01

0.1

1

10

100

1000

1 10 100 1000 10000 100000 1000000 10000000

Kt = QH/N

S F I =

L 3 / D 4

W'ton HN

Pacific SVCN7FLS HPXPump-Turbines

6,100KtE-0.76BWIP SC7

The DIS2 will have a standard simple method of calculating SFI and will refer to a figure based on Figure 5 for guidance. A significant inconsistency with the value in the chartwould be cause for the purchase to perform a more in depth investigation before

purchase.

BEARING SYSTEM LIFE:

The API 610 individual bearing life requirements have b een unchanged since the XthEdition in 19xx. In the current 11 th Edition draft these requirements are as follows:

6.10.1.7 Rolling-element bearing life (basic rating life, L 10h) for each bearing or bearing pair) shall be calculated in accordance with ISO 281 and be equivalentto at least 25 000 hrs with continuous operation at rated conditions, and at least16 000 hrs at maximum radial and axial loads and rated speed.

NOTE 1 ISO 281 defines basic rating life, L 10 , in units of millions of revolutions. Industry practice is to convert this tohours and to refer to it as L 10h . ISO 281 also defines the method required to calculate bearing system life from individual

bearing life.

NOTE 2 For the purpose of this provision, ABMA 9 is equivalent to ISO 281. It is the experience of the authors and many other users that fatigue failure of bearings isnot a significant issue in most plants. Typical bearing failure numbers are 8-10% of all

pump failures for conventional lubrication and fewer than 1% for oilmisted pumps.Bearing failures are almost entirely lubrication related. Either the oil is contaminated orthere isnt enough oil. As a result the WGTF has not paid much attention to paragraph6.10.1.7, feeling that current experience proves it is adequate.


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All API standards contain a paragraph establishing the minimum design life for theequipment covered. In the current draft of 610, the paragraph reads as follows:

6.1.1 The equipment (including auxiliaries) covered by this International Standard shall bedesigned and constructed for a minimum service life of 20 years (excluding normal-wear parts asidentified in Table 19) and at least 3 years of uninterrupted operation. It is recognized that theserequirements are design criteria and that service or duty severity, misoperation or impropermaintenance can result in a machine failing to meet these criteria.

Design is defined in the following paragraph:

3.10designmanufacturers calculated parameter

NOTE Design is a term that may be used by the equipment manufacturer to describe various parameters such as,design power, design pressure, design temperature, or design speed. This term should be used only by the equipment

manufacturer and not in the purchaser's specifications. There is clearly an inconsistency between the bearing life requirement of 25,000 hrs(approximately but just less than 3 years) and the general requirement for the pump to bedesigned for 3 years. This inconsistency comes about as a result of the fact that for the

bearing system to be designed for 3 years, the individual bearings must be designed formore that three years. For two equally loaded bearings the individual bearings wouldhave to be designed for 40,000 hours L 10h . The bearing system life can be calculated bycombining the individual bearing lifes as follows:

L10h (System) = [(1/ L 10h A) 3/2 + (1/ L 10h B) 3/2 + + (1/ L 10h N) 3/2 ] 2/3

where: L 10h A = Basic rating life, L 10h per ISO 281 for bearing A,L 10h B = Basic rating life, L 10h per ISO 281 for bearing B,etc.

So, why not simply require the bearing system to be designed for 25,000 hours or 50,000hours or whatever? In a nutshell longer bearing life requires larger bearings. Over manydecades pump manufacturers have tried to increase bearing life by installing larger

bearings. They have consistently had trouble meeting the bearing temperaturerequirements of 610 when bearings larger than a 7314 are used at 3600 rpm. The authoris only aware of one pump manufacturer that uses a 7315 bearing in his largest OH2

bearing frame. Pump manufacturers are very concerned that requiring a 25,000 hour

system life will force redesign of their bearing housings which costs money and worsemight result in overheating of the oil and ball skidding failures. Further manufacturersand many users view bearing fatigue life as a non issue. Other users strenuously object tothe inconsistency and some other API standard taskforces are receptive to higher bearingsystem life requirements than 610 for other types of equipment.

To understand why bearing life is not an issue in single stage overhung pumps one canlook at the form of the equation for L 10h bearing life. Before looking at the equation lets


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review how the life calculation is performed. First the actual radial and axial forces onthe impeller are determined (HI 1.3 and multiple other texts discuss methods ofdetermining these forces). Then coupling and seal forces are determined. The forces arethen summed and the equivalent radial load, Pr, is calculated for each bearing.

Pr = XF r + Y F a Where:

Fr = Radial LoadFa = Axial Load

andX & Y are factors from a table in ISO 281, these factors vary depending on the

type of bearing and the relative magnitude of the radial and axial loads.

At this point we go to the bearing manufacturers catalog and select the smallest bearing

that will give us an acceptable L 10h life. L10 life is calculated from the followingequation:

L10 = (C r /P r )3

Where:

Cr = Catalog load ratingPr = Equivalent radial load as above

This result is in millions of revolutions so we convent to hours by dividing by the numberof revolutions per hour. Now lets look at the equation. For a system of two equallyloaded bearings to have a 25,000 hr life, the radial bearing and the thrust bearing (40degree angular contact duplex pair must each have a 40,000 hr life. The ratio of 40 to 25is 1.6. Rearranging our life equation we find that as long as C r /P r is 1.17 or greater thelife will be 40,000 hours or greater. So how likely is it that the ratio of C r /P r is 1.17 orgreater?

To begin with we will restrict our discussion to single stage overhung, OH2, pumps.Most manufacturers will have either three or four bearing housing sizes for their OH2

pumps. They will line their sets of hydraulics up against these standard bearing housingsizes based upon the equivalent radial loads the bearings must deal with. Each bearinghousing will have a single set of hydraulics that represents the highest possible loads atmaximum diameter impeller and some arbitrary suction pressure. For the manufacturerfrom which the following example comes, that suction pressure appears to be 250 psig.This seems to be a sensible number to the authors in that a 250 psig suction pressure

probably covers 98% or more of all refinery services. (It is also noted that the pumpmanufacturer has some tricks in this bag for higher suction pressures, such as differentialwear ring sizing and plugging balance holes in the impeller.)


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The equivalent radial load is made up of the weight of the rotor, the radial thrust on theimpeller, the axial thrust on the impeller, seal compression load and the coupling axialthrust. API 610 requires that the coupling axial thrust be input as the thrust at themaximum allowable axial misalignment allowed by the coupling. It should be rare that auser will stretch or compress the coupling to the maximum as opposed to repositioning

the coupling hub somewhere close to the right distance between shaft ends or hubs.Additionally API 610 disallows pump selections that do not allow at least a 5% increasein head. This latter requirement means that except in those cases where revisions to

pump hydraulics encroach on the 5% margin thrust will be lessened by 5% from theworst case plus a decrease in thrust load due to something approaching proper couplinginstallation. Next we have the arbitrary suction pressure (of 250 psig; where individual

bearing life is about 25,000 hrs). The suction pressure for any given pump doesnt haveto be much less than the 250 psig assumed in bearing selection for the ratio of dynamicload rating to equivalent radial load to be greater than 1.17. Figure 6 shows the L 10h lifefor a 4 x 6 x15 OH2 pump at maximum diameter impeller It is seen that L 10h life

becomes a huge number for most suction pressures encountered.

Figure 6. L 10h Life for a typical OH2 Pump

0.00

200000.00

400000.00

600000.00

800000.00

1000000.00

1200000.00

0 50 100 150 200 250 300 350

Suction Pressure (psig)

L

1 0 h L i f e

( h o u

.The last point in this discussion is that this pump is represented as having the heaviestloads of any set of hydraulics used for this bearing housing. All smaller sets of

hydraulics will have even longer L 10h lives. This agrees with most user experience thatfatigue failures of API pump bearings are exceedingly rate. What is puzzling is why ballskidding failures are not more common? The WGTF does not have agreement as to whatwords will be in the DIS2 but it is likely we will simply require a 40,000 hour bearingsystem L 10h life.

PERFORMANCE TESTING:


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The performance test tolerances of API 610 may be the only requirements in the standardthat have remained unchanged since the 1 st Edition in 1954. During this time HydraulicInstitute Standard 1.6, Pump Tests has changed its tolerances and evolved but it has neveragreed with API 610. Additionally ISO has a standard ISO 9906, Rotodynamic pumps Hydraulic performance acceptance testsGrades 1 and 2, 2000 does not agree with

either HI 1.6 or API 610. In spite of this API 610 currently references both HI and ISO9906.

The purpose of the reference is to use the test methods and allowable measurementuncertainties and basically everything except the performance test tolerances. The API610 tolerances are shown in Table 2.

Table 2. API 610 Performance Test Tolerances

As the WGTF has worked through the 11th

Edition drafts we have received multiplecomplaints about the tolerances in Table 2 (14). If one studies the table one will note anumber of interesting things. First if one converts the allowable tolerance in head from a

percent to pressure in PSIG (which is what is measured) the tolerance is shown in Figure7.

Figure 7. Performance Test Head Tolerance Band Width in PSIG

Table 14 Performance tolerances (API)Ratedpoint

Shutoff

% %Rated differential head:

2 + 10

+ 5 10 a 2 + 8

+ 3 8 a 2 + 5

+ 2 5 aRated power + 4 b Rated NPSH 0 NOTE Ef f iciency is not a rating value.

a If a rising head f low curve is specif ied (see 5.1.13), the negativetolerance specified here shall be allow ed only if the test curve still show s arising characteristic.

b Under any combination o f the above (cumulative to lerances are no t acceptable)

Condition

0 m to 150 m (0 ft to 500 ft)

151 m to 300 m (501 ft to 1 000 ft)

> 300 m (1 000 ft)


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Tolerance Band Width

0

5

10

15

20

25

30

35

40

0 500 1000 1500 2000 2500

Head

P S I f o r

T e s

t o n

W a t e r

Band Width

3-.003*Head BW


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In the DIS which was balloted last winter the WGTF proposed equations that plotted atthe purple line. A huge number of people objected to this as too complicated.

Another issue is that Table 2 (14) does not mention flow. The flow tolerance for the 9 th

and 10th

Editions is contained in paragraph 7.3.3.3 b). The tolerance is +/-5%. This is a big number. The tolerances have now defined a rectangle shown in Figure 8.

Figure 8, Typical Performance Test Curve

Figure 9 is a close up look at the Allowable Test Point Region or block.


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The point of Figure 9 is the allowable test region allows plenty of room for discussionand it allows a stack up of tolerances that can result in actually efficiency beingconsiderable lower than quoted and yet the pump would be acceptable. For typicalrefinery pumps this may not be very important but for large pumps and pumps that havemostly frictional system curves such as pipeline pumps, this is a very big deal. It should

be noted that API 610 bases acceptance on power not efficiency. This is becauseefficiency is a derived or calculated value whereas power is measured directly. If onetakes the uncertainties (allowable inaccuracies) in the values measured during the test onecan see that with acceptable uncertainties in other variables the uncertainty in efficiencyif very large. Figure 10 shows the Allowable uncertainties in various test parameters


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8.3.1.2 Performance and (NPSH) tests shall be conducted using the methods anduncertainty requirements of ISO 9906 grade 1, HI 1.6 (for centrifugal pumps) or HI 2.6 (for verticalpumps). Performance tolerances shall be in accordance with Table 15. Evaluation of resultsshall be in accordance with 8.3.3.3 b).

And

8.3.3.3 Unless otherwise specified, the performance test shall be conducted as specified inbelow.

a) The vendor shall take test data, including head, flowrate, power and vibration at aminimum of five points. These points will normally be

1) shutoff (no vibration data required),

2) minimum continuous stable flow (beginning of allowable operating region),

3) between 95% and 99% of rated flow,

4) between rated flow and 105% of rated flow,

5) approximately the best efficiency flow (if rated flow is not within 5% of best efficiencyflowrate)

6) end of allowable operating region.

b) The test data shall be fit to a spline or appropriate polynomial (typically third or fourth order)for head and for power using a least squares method. The rated/guarantee flow shall beinserted into the resulting equation and a value for head and power calculated. Thesevalues shall be corrected for speed, viscosity and density (specific gravity). The correctedvalues of head and power shall be within the tolerance bands allowed in Table 15.

In the case of high-energy pumps (see 6.1.18), integral-gear and multistage pumps, itmay not be feasible to test at shutoff. Some low specific-speed pumps cannot achieve120 % of BEP flowrate for the rated impeller diameter.

c) Unless otherwise agreed, the test speed shall be within 3 % of the rated speed shown on thepump data sheet (see example in Annex N). Test results shall be corrected to rated speed.

d) The vendor shall maintain a complete, detailed log of all final tests and shall prepare therequired number of copies, certified for correctness. Data shall include test curves and asummary of test performance data compared to guarantee points (see 10.2.4, 10.3.2.2 andexample in Annex M).

e) If specified, in addition to formal submittal of final data in accordance with 10.3.2.2, curvesand test data (corrected for speed, specific gravity and viscosity) shall be submitted within24 h after completion of performance testing for purchaser's engineering review andacceptance prior to shipment.

Having said this, we have one more round of review and comments and the performancetesting section will almost certainly draw numerous comments.


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CONCLUSION

The main focus of the 11 th Edition of API 610 is to improve clarity and to deal with alimited number of key issues. The three key issues, Performance Testing, BearingSystem Life and Shaft Flexibility Index have attracted sufficient interest and comment

that the Working Group/Taskforce has decided to submit the document for an additionalround of comment. In the ISO world this is a second Draft International Specification.

The second DIS is expected to be distributed by ISO and API in early 2007 with thecomment period and ballot closing approximately mid year. Comments will be resolvedin the third quarter. Comment resolutions will be presented to the API Subcommittee onMechanical Equipment at the Fall Refining Meeting in early November 2007. The planis for the Final Draft International Specification and API Ballot to take place in early2008. Publication is planned for mid 2008.

References:

This tutorial contains unpublished work by a number of members of the API 610Taskforce/ISO 13709 Working Group. Among these members are:

Mick Cropper, Sulzer PumpsFred Blumentrath, CPCTerry McGuire, FlowserveCharle Heald, ConsultantJim Harrison, Flowserve

Additionally the ideas in this tutorial have been affected/developed through inputs fromthe entire ISO 13709 Working Group/API 610 Taskforce. The authors asknowledge andthank all of them for their contributions.

Hydraulic Institute Standard 1.3


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Appendix 1 BACKGROUND DATA COVERING THE HISTORY OF API 610AND API 682 GASKET REQUIREMENTS

API 610 1 st Edition (tentative) 1954: Totally silent on gaskets or gasketmaterials.

API 610 1st Edition January 1955: Text is silent on gaskets and gasket materials.Datasheet has a block for gasket materials and blocks for Confined or Flat beside theword Gaskets.

API 610 2nd Edition January 1957: Text is silent on gaskets and gasketmaterials. Datasheet references are identical to 1 st edition.

API 610 3rd Edition January 1960: Text is silent on gaskets and gasketmaterials. Datasheet references to gaskets have been removed. Thus 3 rd Edition is totallysilent on gaskets.

API 610 4th Edition July 1965: Text and datasheet are silent on casing gaskets but Seal Gland Plate gaskets are addressed in Section 24, item d.:

API 610 5th Edition March 1971: Addresses radially split casing gaskets in item12. f.:

Seal End Plate gaskets are addressed in item 24. k.:


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Casing Gaskets are addressed in the Materials Section in Table D-1:


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Note that the table does not reference spiral wound gaskets at all and that the only choicesare variations of asbestos gaskets and Teflon. Teflon Casing Gaskets are addressed in theGeneral Notes, Note 10:


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The datasheet has no reference to gaskets.

API 610 6 th Edition January 1981:

Pressure casing gaskets are covered in 2.2.7:

Gland gaskets are covered in 2.7.1.17:

The datasheet has no reference to gaskets for either the casing or the seal gland plate.

The materials table is now E-1 but the requirements are identical to 5 th Edition. The tableis now so large it is impractical to scan and insert into this record. Note 10 of the 5 th Edition is now note 7.

API 610 7th Edition, February 1989:


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In the 7 th Edition there are two references in the text related to pressure casing gaskets.These are 2.2.6 and 2.2.10:

The fact that 2.2.10 gives the requirements for o ring grooves implies that o-rings might be acceptable gaskets. Seal gland gaskets are covered by 2.7.1.16. This latter clausspecifically classifies o-rings as controlled compression gaskets further implying that o-

rings can be used as gaskets on pressure casings.

Datasheet has no blanks or references to gaskets for the casing or seal gland plate.

After the text has created this ambiquity. Annex H completely contradicts theacceptablility of o-rings by only calling out spiral wound gaskets on the pressure casing.This applies to both casing and seal gland gaskets. Further note that this is the first timespiral wound gaskets are mandated. No previous edition calls for them.


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API 610 8 th Edition, August 1995:

The pressure casing gasket is covered in 2.2.7 and the o-ring groove is covered in 2.2.10.For the first time 2.2.2 is crystal clear, o-rings can be used as casing gaskets.

Seal gland gaskets are covered by 2.7.3.23. It is also crystal clear that o-rings are allowed between the pump casing and the seal gland.

Table H-1 has been reduced in size to fit on a single page but the requirements forgaskets are unchanged and only spiral wound gaskets are called out.Note that there isanother paragraph in the mechanical seal section 2.7.37 which says that seal gaskets andhard faces shall be specified from the seal materials tables. This gets us a material for thegland gasket but there is no reference directing materials for other pressure casing gasketsif o-rings are used.


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Note that there is another paragraph in the mechanical seal section 2.7.37 which says thatseal gaskets and hard faces shall be specified from the seal materials tables. This gets usa material for the gland gasket but there is no reference directing materials for other

pressure casing gaskets if o-rings are used.

API 610 9 th Edition, January 2003 and 10 th Edition, October 2004

The pressure casing gaskets and o-ring groove requirements are covered by 5.3.10 and5.3.12. They are unchanged from the 8 th Edition.

Clause 5.8.3 for the first time refers to spiral wound gaskets being used on the sealchamber joint. Note that this paragraph implies that bolting might have to be increased insize to properly crush a spiral wound gasket. This could cause manufacturers to prefer touse o-rings on this joint.

Seal chamber gaskets are covered by 5.8.11 and are unchanged from the 8 th Edition.

Table H-1 continues to only call out spiral wound gaskets for the pressure casing.


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API 682 3rd Edition

API 682 has three paragraphs and two annex sections relevant to o-rings. The first is themost relevant. Paragraph 6.1.6.7.2 mandates o-rings on the joint between the sealchamber and gland plate for services below 350 F.

The other two paragraphs, 6.2.1.2.2 and 6.2.2.2.2, are identical and are also identical tothe requirements in API 610:

API 682 also gives temperature limitations for o-ring materials and a tutorial on theirselection. These two sections follow.


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Standard Paragraphs, Revision 23B, November 3, 2005

The standard paragraphs have only one paragraph relative to pressure casing joints. This paragraph, 6.2.4, discourages the use of o-rings but is really aimed at compressors orsteam turbines.


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Documents

API 610 notes.pdf