Turbomachinery Handbook- Hydrocarbon Processing-1974

TURBOMACHINERY HANDBOOK

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Reprinted from HYDROCARBON PROCESSING

o Gulf Publishing Company o @ 1974 . $t.ZS

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TURBOMACHINERY HANDBOOK

Published by HYDROCARBON PROCESSING This reference manual has been reprinted from the regular monthly issues of HYDROCARBON PROCESSING. Other handbooks and manuals in the series are:

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Table of Contents

TURBOMACHINERY HANDBOOK Page No.

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Centrifugal Compressor Performance Testing Better mechanical testing can improve compressor reliability Test compressor performance-in the shop Test compressor performance-in the field Maintenance Techniques ....... How to improve compressor and maintenance lmprove machinery mainten Why clean turbomachines A closer look at turbomach Hot alignment too complicated Turbines using too much steanl"? ... Better pump grouting .1..... Centrifugal Pumps .. !-. i.. How to improve pump How to control purnp vibration How to prevent pul Turboexpanders ..1 New develop hot gas systems Turboexpanders Turbotmachinery and C GasS histories of : causes Coniprq$sor prol Critical. Related Are couplings Itnik in How tr I

Dlag A new

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5 6 11

14 17

18 24 28 31

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38 41

45 46 49 56 61

62 66 71

72 76

83

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Centrifugal Compressor Performance Testing

Better mechanical testing can improve compressor reliability

Vendor shop testing of critical process compressors can reveal mechanical problems prior to field installation. Here are some points to consider when pertorming compressor mec hanical tests

TABLE l-Objectives of centrifugal compressor mechanicaltests Partial verification of: oThe quality of over-all unit assembly o Fieedom from internal rubs r Bea ring fit, alignment and adequacy of lubrication o Rot or-bearing system dynamic stability and calculated critical speed

oVibration levels o Cor rectness of assembly and tightness of shaft oil seals o Drive coupling fit-up and balance o Lub rication system cleanliness and performance

oTrain component compatibility (optional) o Noise (optional)

Douglos F. Neole, Union Carbide Corp.-Chemicals and Plastics, South Charleston, W. Va. Conrpnrsson MEoHANTcAL tests have proven effective in identifying design and fabrication deficiencies. If these deficiencies are identified before shipment of the compressor corrections can be made and verified under optimum shop conditions. The value of compressor mechanical tests is limited by the absence of significant gasJoad and the low energy involved. The tests are most effective if API test standards and the customer's supplementary test specifications are strictly enforced as a minimum requirement, Revised and new API Standards relating to compressors and oil systems will significantly strengthen test requirements. The standards will also increase the obligations of vendors with regard to test facilities and data acquisition.

Test obiectives. Table 1 is a brief list of specific objectives of compressor mechanical tests. Note that only

partial verification of the considerations listed on the table is claimed. Even the most sophisticated mechanical test falls short of complete and prolonged simulation of actual operation under full gas-load and power. 6

API test slqndqrds. The Second Edition of API Standard. 617, "Centrifugal Compressors for General Refinery Services,"'has provided an accepted basis for mechanical spin tests since 1963. Most manufacturers and purchasers added to the requirements of API 617 as new technology developed. Many of these additional requirements will be reflected in the Third Edition of API 617, scheduled for

publication this

year.2

Table 2 summarizes key differences between the old and new API Standard 617. The new standard provides a significant step toward beter defined, more uniform and more rigorously verified centrifugal compressors, but there are areas where the purchaser must define and exercise his rights 'andf or preferences.

Righf to wilness. The purchaser reseryes the right to observe testing, dismantling, inspection and reassembly of eQuipment "when specified." The vendor must provide "sufficient notice" when shop inspection by the purchaser is required. Sufficient notice is to be defined by mutual consent of the purchaser and vendor. We have found that vendors frequently cannot be counted on to maintain

BETTER MECHANICAT TESTING

TABLE 2-APl Standard 617lor centrifugal compressor mechanical tests Second Edition (1963) Mandatory hydro tests

Third Edition (Planned issue) Essentially same

the right

lnquiry specifications and order

Purchaser reserves

shall specify tests requiring witness lncrease speed in undefined increments from 0-110[ MCS Run at MCS for 4 hours, uninterrupted

observe tests

lnstall oil seals after test and

Use contract shaft seals and bear ings during test

manually turn Demonstrate control systems (E.G.; inlet guide vanes) to the

lncrease speed f rom

0-100[

to

MCS

in 10ft increments Run at MGS for 4 hours

Not required (except lube)

extent practical Make every effort to determine first critical of flexible shafts, short of opening case and unbalancing

The actual first critical shall be

Vibration:

Vibration:

Record throughout operating

determined

Essentially same

speed range

demonstrated "to the extent practical." For centrifugal compressors, we have interpreted this to relate to lube oil systems and inlet guide vane actuators. The Third Edition of API 617 does not contain a comparable requirement. The new API oil system requires that "the completed oil system shall be shop run to test operations," etc., but it is only by rather liberal interpre_t21i6n that this covers inlet guide vane actuators.3 Within the past two years, we have experienced several startup difficulties from miscalibrated or mechanically unsound guide vane linkages. Guide vanes will receive more, not less, attention.

Delerminqtion of criticol speeds. The new edition of API Standard 617 requires that "for flexible-shaft compressors, the actual first critical speed shall be determingd-" whereas the Second Edition required "every sff611-16 determine the actual first critical speed-short of opening the casing to create rotor unbalance." Our specifications have independently paralleled the evolution of the API Standard 617 and have avoided forcing criticals. Present thinking is toward mandatory identification

Not required

Perform and record vibration sweeps

of criticals

Not required

Use purchased vibration probes

Vibrqtion qnd beqring temperqture meqsuremenl qnd qnqlysis. The revised issue of API Standard 617

Not required

Run spare rotors ordered with c0mpress0r

Vendor

to provide certified detailed

logs including vibration and oil temperature data

Essentially same but include additional data on lube and seal oil systems, and rotor dynamics

records of all required inspections or to provide adequate advance notice of inspections. Notification is frequently too late to permit the most knowledgeable machinery specialist to be assigned and made familiar with his assignment. The alternative is for the purchaser to con-

tinually stress during negotiations and order coordination meetings, the requirement for responsible and timely vendor notification, and to follow up with frequent checks by his own inspection/expediting personnel.

Tesl speeds qnd durqtion. The Third Edition of API 617 is strengthened by the inclusion of a definition of the increments of increasing speed during mechanical tests. The four-hour maximum continuous speed run apparently is less rigorous, as an uninterrupted run is no longer required. However, we specify successful completion of an uninterrupted four-hour run, and a repeat uninterrupted run if bearings, seals or balance require modification of the test.

Verificqtion of shoft seqls. A significant improvement in the Third Edition of API 617 is the requirement that contract bearings and shaft end seals be installed during the mechanical test and that the oil leakage from each seal be measured with approximate design differential pressures across the seals. Although not required by the Second Edition, many have imposed this requirement for sometime. More widespread testing with seals in place will occur following the issue of the Third Edition. It will be easier for purchasers to gain confidence in the mechanical adequacy of seals.

Conlrol syslem checks. As indicated in Table 2, the Second Edition of API 617 required control systems to be REPRINTED FROM HYDROCAflBON PROCESSING

as required by the

Third Edition.

reflects the progress made during the past 10 years on vibration measurement and analysis. For example, a sweep of vibration amplitudes at frequencies covering a minimum range of 25 percent of synchronous to twice vane passing frequency is now required. This is interpreted to mean multiples of number of vanes times stationary parts such as discharge nozzles or diffuser vanes. Additionally, purchased vibration probes and detectors shall be used during the mechanical running test. We have preferred this practice for several years. Requirements or recommended practices relating to use of bearing metal temperature sensors during test are not stated in the API Standard. These detectors have become standard for our critical rotating machinery and should be better covered by APL

Spore rotors. Because of the need for reliable process compressors, a mechanical run test of spare rotors ordered with the compressor is now mandatory. The prac-

tice has been valuable by not only ensuring a properly balanced rotor, but by avoiding situations where the spare, although properly balanced, just wouldn't fit in the machine without extensive diaphragm or thrust bearing adjustments.

Vendor tesl dqto. The Third Edition of API 617 requires submittal of more extensive test data to the purchaser. The more stringent data requirements reflect the improved scope of the tests as previously discussed. Additional data includes seal oil temperatures, pressures and leakage rates; rotor balancing and critical speeds, and more extensive vibration data.

Optionol tesls. Optional tests which may now be speiified in the purchaser's inquiry or order include hydraulic performance, complete unit tests, tandem tests, gear tests, helium leak tests, sound level tests and post-test inspections. We consider performance tests on an individual basis as a function of the extent to which the design has been proven by analogous machines, the degree to which

TABLE 3-AP! Standard 614* for lubrication, shaft-sealing, and control oil systemssummary of inspection and test requirements

startup problems are not uncommon. Furthermore, the mandatory use of the system for test will provide a signifi-

cant incentive to thoroughly clean the systems

in

the

vendor shops. o Purchaser has right of inspection oPurchaser's inspector furnished specifications, material certifications, and running test data oPurchaser shall specify whether purchased oil system shall be used during main equipment shop test oComponent hydrostatic tests comparable to previous API Standard-617, Second Edition

o Four-hour shop test under normal system operating conditions o System checked for leaks and proper functioning of controls and alarms .System capable of riding out filter-cooler and oil pump changeovers . System cleanliness criteria defined +

Planned issue

TABLE

4-Vendor test plans

Test installation Shop and contract equipment included in the test train Xnown limitations in driver power and/or speed Measuring instruments Pressure and/or temperature control systems Seal and lube oil console limitations

Test procedure Duration of each speed run Shop data sheets Special procedures to avoid excessive temperatures Post-test inspections

or other limitations

Report outline Scheduled issue date Data to be reported Disbussion subject headings

final operating conditions can be synthesized, the criticality of the machine, the time schedule and the cost. Gears are also subjected to running tests. When duplicate gears are ordered, they are subjected to full speed, full torque load and back-to-back tests. Following the test, a full bearing and tooth-contact inspection is performed. The gear undergoes a standard four-hour shop running test on reassembly. Complete unit, tandem, or shop sound level tests are seldom justified.

Lube, seol ond conlrol oi! systems. A properly de-

signed and functioning oil system is a prerequisite for a reliable compressor train. Thus a discussion of compressor mechanical tests should include consideration of the oil system that will support the train. This consideration is facilitated by the new API Standard 614 entitled "Lubrication, Shaft Sealing, and Control Systems for SpecialPurpose Applications."s Table 3 briefly summarizes key features of oil system tests as specified by API 614. The right of inspection and receipt of finished specifications, material certifications and running test da[p are established. It is the purchaser's responsibility to specify whether the purchased oil system shall be used during the compressor test. Our normal practice has been to require cleaning, flushing and shipment preservation of oil systems. Operation of the contract oil console during mechanical testing of the compressor has been required only infrequently. Ifowever, some purchasers are requiring tests to be run with the contract oil console and this appears to have some advantages. The design and fabrication of oil systems seem to be a sideline with many compressor manufacturers, and 8

Additional points in API Standard 614 bear consideration. Filter and cooler changeovers shall be accomplished without the system's delivery pressure dropping to the automatic start setting of the standby pump. The capability of the control valve shall be demonstrated by starting, running and stopping a second pump (main or standby) without the delivery pressure dropping below 75 percent of the differential between normal and shutdown pressures. The valve shall also be capable of holding oil pressure at minimum oil flow (normal bearing and seal oil plus steady state control oil flows). Problems with new systems could have been avoided by the foregoing types of checks. For instance, we frequently use shaft-driven main oil pumps. After starting a unit on the electric motor-driven auxiliary pump, the main pump is operated briefly in parallel with the auxiliary, and the recycle control valve is wide open. When the auxiliary oil pump is tripped off, the valve response was too slow to prevent a drop in header pressure and subsequent trip of the compressor train. Some vendors have resolved this problem through more responsive and better tuned controls. Other vendors have provided a manually controlled recycle line from the auxiliary pump which can be opened to effectively remove the pump from the system prior to its shutdown. We have made similar last minute modifications in the field. Proper exercising of oil systems during test can avoid these costly and time-consuming nuisances.

Test plans ond preporolions. The latest API Standards provide good general definitions of test requirements. It is only in limited instances, however, that the standards require vendor submittal of a detailed plan for each specific compressor system test. We have found that this is a serious omission. Some vendors stop short of full compliance with API requirements. For several years we have required the vendor to submit detailed test plans for review prior to testing of critical machinery. The contents of an ideal test plan are listed in Table 4. Advance review of the test setup will avoid surprises to inspectors andf or machinery specialists when they report to the vendor's shop. These setups normally have sufficient drive turbine power available to attain specified system speeds. It is common, however, to encounter limitations with electric motor drives. The unavailability of a motor with properly rated speed and the inability to vary speed result in less than adequate tests. The development of excessive casing gas temperatures during tests have also compromised tests. Prior knowledge of these circumstances can permit corrections or alternatives to be implemented in a timely manner. Situations where the purchaser is forced to compromise his test specifications or lose his scheduled time on the test stand can be avoided. The proposed test procedures should define the basis for

determining when the compressor undergoing test

has

stabilized and when the test operator is free to move on to another point. Normally, the temperature of oil leaving

the bearings is the last indicator to stabilize. The purof oil temperatures at each point seldom extends the test significantly. It is desirable, however, for the vendor to know in advance that this requirement will be made. chaser's insistence on stabilization

BETTER MECHANICAL TESTING

spectrum ama)yzerc, X-Y plotters and oscilloscopes to permit any probe or set of probes to be read, analyzed and

The planned shop data sheet (or typical automatic data center printout) provides practical insights into the extent of data to be taken. A quick scan of these sheets can identify unacceptable omissions. For example, it is common to find that seal oil leakages are not to be recorded, or are to be recorded at static conditions or at reduced speeds and oil temperatures. Every effort should be made to simulate design speeds and operating temperatures to get the best possible insight into actual operating per-

displayed. Multichannel tape recorders are becoming increasingly cornmon for simultaneous recording and subsequent display of many data channels. Some vendors have even installed "run-out subtractors" to electronically mask vibration or excessive rotor or probe surface runout. One major rnanufacturer recently commissioned a minicomputer system to simultaneously analyze multiple data inputs and print results within minutes of test completion. Progress has not been so great with regard to bearing metal temperatures. We are strongly committed to using bearing metal temperature sensors on critical machines but have found that even more extensive vendor liaison is required than with proximity probes. Test installations for reading out the sensors are frequently vendor impro'

formance.

The poor legibility of vendor test report forms is a of frequent frustration. Specifications call for legible data sheets, but it seems some will have to be rejected source

before satisfactory quality can be attained. The lack of interpretation of test results is also of concern. Most vendors will provide a formal test report with bare data. Ordinarily, interpreted d.ata are supplied only if required by the purchase order. This is a worthwhile requirement; prior consideration of the characteristics of a compressor can be valuable when diagnosing a real or apparent problem on the midnight shift!

VENDOR TEST INSTAI.I.ATIONS Purchaser emphasis on the reliability of compressors has led to a general upgrading of vendor test installations. It has become the exception when major vendors are unable to comply with the older API test requirements.

Vibrqtion ond temperqlure monitoring. Significant progress has been made in using the purchaser's own proximity probe holders and probes during mechanical testing" In the past, vendors were slow to accept the feasibility of shop installation of the probes----especially two per journal that could be adjusted and replaced while the machine was in operation. The contract probes were frequently not available for the test so temporary shop probes and holders were used. Part of the value of shop mechanical tests as a reference for startup problem diagnosis was lost because surfaces observed by the shop probes differed from surfaces observed by permanently installed probes, It is still a challenging task to develop mutually accept-

able proximity probe installations. Vendor's ability to utilize probes during test has improved, however. Many vendors have installed variable frequency filters, real time

About lhe qulhor DoucLAs F. NsAr,E is

vised and may not be available during the test unless requested

in

advance.

Couplings. Common test stand practice is to use shop couplings except for possibly one hub, during compressor mechanical tests. We have accepted this practice in the past. However, recent experiences indicate that couplings are one of the more troublesome machinery train components. Many coupling manufacturers apparently do not enforce standards for residual unbalance and vibration comparable to those of compressor manufacturers. Additionally, coupling manufacturer's use, during balancing, of coupling spool and sleeve surfaces that were eccentric to

the gear tooth pitch circle has led to "cranking" and excessive vibration.

The use of the contract coupling is being increasingly specified for tests of critical machines for several reasons:

o The contract coupling will permit the closest practical simulation of the final field installation with respect to rotor dynamics.

o The mechanical test can confirm coupling

o The use of the contract coupling reduces the temptation for vendor shop personnel to "touch ,p" the shop coupling to compensate for minor rotor unbalance. The added expense and time required to adapt the contract coupling to test service is well justified. CONDUCT OF TESIS The strengthened API standards supplemented by the preparations discussed in this paper will contribute to

effective mechanical testing mo,nager---ltrocess

macldnerg appl;ication

for

(Jnion Car-

bide Corp.--4hemicals and Plastics,

South Charleston, W.Va. He i,s responsible for the sTtecifi,cation antl process desi,gn of large machinerg systems for capital erpansions and deaelopment of new technology in this area. P,r'i,or erperience with Union Carbide Corp. includes plant process design and a oari,ety

of engineet'ing nxanag enxent as signments. Mt'. Neale holds a B,S. degree i,n mechanical eng,ineering from Rensselaer Polgtechnic Institute and, a M.S. degree i,n meclu,nical engineering from Massachusetts Institute of Technology. He is a registered prof essional engineer in West Virgina and, a member of A.S.M.E.

REPRINTED FROM HYDROCARBON PROCESSING

balance

and provide an indication of machining quality.

of

centrifugal compressors.

The following additional recommendations are

made,

however:

o A competent

machinery engineer experienced in startup and maintenance problems should witness the tests. Onthe-spot interpretations and decisions will materially influence the effectiveness of the test and assure satisfactory performance of the compressor.

a Rotor

residual unbalance data sheets and run-out maps should be reviewed prior to the test. The knowledge gained can alert the test engineer to the possibility of rubs or abnormal noises.

o The vendor's calculations of expected casing temperatures during tests should be reviewed. There are too many

instances where failure to plan a vacuum test has resulted in termination of the run before planned speeds are reached.

o The use of electronic rotor run-out subtractors in conjunction with proximity vibration probes should not be permitted. Vendors should machine probe surfaces well enough to avoid the need for these corrections.

. it

Insist that actual rotor critical speeds be identified. If is necessary to "force" the critical, the removal of. a coupling bolt will normally suffice. If the coupling is too close to a bearing, it may be necessary to open the case and add internal weights.

o

Compressor design and manufacturing deficiencies become apparent with the passage of time. Therefore, dqmand an uninterrupted four-hour maximum continuous speed run regardless of the test stand problems encoun-

tered. If it is a hardship to keep the machine operating for four hours on the test stand, it is likely that the same condition will prevail in the field.

o Be certain that the contract or shop drive coupling removed prior to rotor rebalance.

It

is

can be a costly error to permit rotor or coupling unbalance to be camouflaged by correction of an adjacent removable part.

oA

previously stored or idle rotor should be run at a speed sufficient for self-correction of bow or "set" before being rebalanced. Recently one of our machinery engineers refused to accept the vendor's suggestion to balance and then test. During the initial run, apparent unbalance disappeared. The vendor's proposal to save time by prior

balancing would have resulted in recurring unbalance after the "set"of the rotor was spun out.

Our vendor inspectors and test engineers always check the compressor bearings and shaft end seals following a test. The case is not opened unless unexplained noise or vibrations are encountered.

ment is limited by the inherent shortcomings of a minimal gas-load, low energy operation. It is influenced by the extent to which operating temperatures, pressures and flows are simulated and by the quality and quantity of recorded measurements.

Our experience indicates that mechanical tests are effective in confirming the quality of assembly of compressors primarily with regard to: o Adequacy of

clearances seals

and alignment of shaft and

o The absence of Ieaks in lubrication and seal oil passages o Bearing fit, alignment and lubrication o Rotor balance o Rotor-bearing stability and avoidance of criticals. Several instances of labyrinth seal rubs, bearing misalignment and damaged thrust bearing shoes were identified. In another instance, a wiped journal bearing was discovered. The wipe was explained by the vendor as the result of a "somewhat tight" bearing, and scraping was 10

support plate was improperly aligned relative to the machine's centerline. At mechanical spin test temperatures, casing expansion was insuficient to permit detection. The problem became apparent when the machine was placed in normal operation. A misbored thrust bearing housing had gone undetected until placed in operation. The axial forces placed on the rotor during the mechanical test were insufficient to "seat" the rotor thrust collar and cause the shaft deflection and vibration observed later in full-load operation. An error in the pattern drawing used for the upper half of the diaphragms of one machine went undetected. This defect permitted hot discharge gas to bypass the balance piston and recycle to suction through the balance piston vent line. A performance test would have provided two indicators of this deficiency: (1) reduced capacity and

(2) higher-than-planned

gas temperatures.

A simple

cal-

culation of the predicted mechanical test air discharge temperature might have identified the problem. Mechanical tests provide significant insight into rotorbearing dynamic stability, especially in Iight of the limited gas dampening provided by the unloaded conditions established for mechanical tests. Excessive vibration was noted on one of our machines. Touch-up balancing was performed on the rotor before it and a duplicated coupling were reinstalled and the test rerun. When the vibration persisted, the rotor was stripped for total reinspection, a potentially disastrous flaw was found, and the shaft was junked.

Value of tesls. ft was stated earlier that compressor mechanical tests only partially accomplish the specific objectives listed in Table 1. The degree of accomplish-

wheel labyrinth

proposed. Our engineer insisted that a new bearing be installed and the test rerun. The same situation occurred. It was found that the bearing housing was undersized and caused excessive bearing crush. If scraping of the bearing had been permitted, the plant would have inherited a chronic problem that would arise each time new bearings were installed. Not unexpectedly, there have been instances of undetected deficiencies in assembly. In one instance, a casing

Mechanical tests of compressors with seals in place have also been effective. fn one instance, excessive seal leakage was evident. The seal O-rings appeared to fit but were not as specified. In addition, an out-of-round outer seal bush-

ing (floating ring) and incorrect spacing of parts

at

assembly were found. The vendor replaced the defective

parts and reworked the seals to obtain a satisfactorily performing system. Lube oil consoles have not been tested with the associated compressor, and pump noise and vibration have been encountered in several instances during startup. These faults are usually corrected by field realignment of the pump and piping. Control system deficiencies hav.e been more common and commissioning cleanup is always a headache. Specification of a more extensive workout in the vendor's shop is increasingly justified. Testing of the oil console and compressor together is becoming a necessary requirement. ACKNOWLEDGMENT

l'[dii"l,c-,Yl:48ii:1,ffi r

::is;"ur'"r,"]l:

LITERATURD CITED

API 617-Ccntrifugal Compressors for General Refinery

Service, Seond

Edition, 1963, Amerien Petroleum Institute, Washington, D. C. zAPI 617-Third Edition, planned issue. t API 61,!-Lubrication, Sbaft-Sealing, and Control OiI System for SpecialPupose Applications, planned

issue.

I

Test compressor

performance

-in

the shop

Royce

N. Brown, Dow

Chemical U.S.A., Houstol

RncrNrr,v there has been much written about thc mechanical reliability of compressor trains. There is no question as to the importance of a high degree of rnechauical reliability since it generally results in a "go" or "no go" situation. Good performance attainability is generally not as consequential, because partial load operation may be possible with a deficient compressor. A fact not generally realized is that the subject of performance is a requirement for reliability! For a machine to be truly reliable it rnust not only run well mechanically but must perform at 100 percent of its capacity whenever called upon. It is then very desirable to test prove performance prior to receiving a machine, o.r as an alternative, to field test shortly after installation prior to enterins prodrrction. AS'YIE TEST CODE

The basis for code testing is the ASME Powcr Test Code PTC 10-1965 Compressors and Exhausters.l Several specific points made in the Code were intended as guiding principles, yet are often misunderstood. The following facts must be considered:

o The

Code establishes the rules for a test, including

the definitions of Code or Non-Code.

Compression equipment must pertorm as expected. Failure to do so can result

in many startup and operation difticulties. Here's what can be done to avoid problems betore the machinery is installed

o The Code is not a textbook on testing. o The Code has to assume that the gas properties oi the gases involved are known.

It

recosnizes

that this

is

Fig,_1-Allowable departure from specified design parameters

for Class ll and Class lll tests. REPRINTED FROM HYDROCARBON PROCESSING

11

TEST COMPRESSOR PERFORMANCE-IN THE SHOP

allowable departure from specified design parameters for Class

ITI0OARD ET{D

II

apply to

SECONO

SEC0lrD

STAGE

STAGE

INI.ET

DISCHARGI

Fig. 2-Typical multi-stage centrifugal compressor.

not always the

case,

but must place the burden of knowl-

edge of gas on the test's participating parties.

o The Code establishes a basis on which to agree or disagree. Ultimately the final test procedure and methods must be agreed on by the purchaser and vendor. The Code attempts to categorize testing and thereby establish an inherent degree of accuracy. These categories are based on methods of test and methods of analysis. The Code establishes three classes of tests. Class I includes all tests made on the specified gas (whether treated as perfect or real) at the speed, inlet pressurg inlet tem-

Class III basically differ only in method of analysis of data and computation of results. The Class II tesi may use perfect gas laws in the calculation while Class IIi must use the more complex ,,real gas,, equations. An example of a Class II test might be a suction throttled air comp.ressor. An example of Class III test might be a COz loop test of a hydrocarbon compressor. Fig. I shows code

The Code establishes a basis on which to agree or disagree. The final test procedure and methods must he agreed on by the purchaser and vendor. 12

and

III

tests.

Another facet of the Code is the establishment of instrumentation for the compressor test. The type, number of points and locations on the comp.ressor are defined and

all

classes

of Code tests. The Code is specific

about the number of readings per point and the minimum duration of the run. Calculation methods are also supplied using perfect gas thermodynamic relations for Class II tests and real gas relations for Class III. Class I inherently requires little correction; but when corrections are needed, either perfect or real gas relations can be used depending on the nature of the gas. The Code draws heavily on work presented by Schultz6 using methods directly from his paper. Calculation methods will be further discussed later in the paper. In actual practice very few tests are run as "True Code," or by Code definition, as Code tests. Each vendor uses some form of deviation to either speed up the test or to fit his facilities. Most of these shortcuts are not serious-in fact they contribute to keeping the cost of testing reasonable. The user must understand fully the Code requirements and where these deviations occur. An example of a nonserious deviation would be the waiver of the time limit per point, i.e., the time required for settling out. This may be based on stability of the data only, rather than data stability and a minimum time. Another deviation generally taken, which the user must evaluate, is the gage calibration procedure immediately prior to and after the test. The use of manometers keeps this requirement to a minimum. There are quite a number of the above examples, too numerous to mention, none of which generally is serious. The user should realize, however, in permitting these deviations he may not have a true code test.

Testing lhe unleslqbles. Compressors, which fall outside of Code limits due to the nature of the gas, vendor shop test limits, or machine speed limits, may still be tested and useful information obtained for the user. The first and most obvious question is how and to what degree one evaluates or judges the test. The most obvious reply is "it depends." On this seemingly vague note, let us explore some of the possibilities. On a simple "once through" compressor, the gas properties may dictate a test speed higher than the physical capability of the unit. For a Class II or III test, the Code requires an equivalent test speed higher than rated speed so that the wheels are operated at the design volume ratio. When this is not possible the machine cannot be Code tested. In the case of the "once through" compressor with two or maybe three impellers, a predicted air curve may be developed. The air curve is derived by using the basic wheel characteristics and calculating a wheel-to-wheel rematch of the compressor on air. The shape of the curve will not be truly maintained. However, the design flow area can be reasonably explored to check capacity and head. It is reasonable to assume if the predicted air curve is reproduced, that the wheels did perform as designed and the compressor should rematch on design gas and produce expected flow and head. Power requirements are somewhat less reliable, but a reasonable measure of effi-

ciency may be made.

For more complex compressor configurations with multiple inlets andf or outlets (Fig. 2), even more judgments must be made. The unit may have to be tested one section at a time. A different test speed may be required for each section. Another method of testing requires the removal of some impellers. While tedious and expensive, it might well be justified. In some cases, particularly if more than one section is tested at one time, additional instruments may be required. Instruments may have to be installed internally within the compressor. The internal instruments are required to sepa.rate the individual section performance from the whole. This test normally uses air as the test gas so the vendor must prepare air-expected curves

for

use

in

SECOM)

test evaluation.

When testing requires significant interpretation a judgment must be made as to whether or not the nature of the test will develop good temperature rise data. lt may be necessary to install total temperature probes at the OD of the last impeller prior to the flow leaving the casing. When properly installed this type measurement provides a much more reliable reading of true temperature rise and therefore gives a more accurate basis for compressor efficiency and input power required. Generally the vendors resist the specification of internal instruments. Therefore, good understanding of users' requirements and objectives is essential. If costs are realistically evaluated, air equivalent testing will cost less than loop testing.

Loop fesfing. A closedJoop test may be necessary to performance test within the Code. There is more inherent accuracy in the loop test than in the ai.r test as volume ratio matching can be more closely achieved. The loop test has several limitations that may not be obvious. Loop tests are generally expensive and time consuming. They may become so complicated with complex compressor confieurations that they become impractical to set up. Finally, the number of gases available for sh'op-loop testing is quite limited. Air which is quite available for normal testing is not particularly suitable for loop testing. Fig. 3 shows a typical shop test loop arrangement. In a loop test, air and oil come into contact causing the danger of an explosion. Extreme caution must be used. All combustible, toxic and other gases where safety is involved are disqualified. This narrows the field consider-

THROTTLE VALvE

Fig. 3-Shop test loop arrangement.

ably. The gas cost factor makes the problem even nlore dificult. The problem of known gas properties adds the final cornplication. The limitations just covered, together with the cost factor, help to provide the incentive to get a useful test from an open air test even if the instrumentation is more complex.

fesl correlqtion. Earlier mention was made of gases and their correlation calculation. A point was made that the Code could not assume the role of being the final author-

ity on gases. Do not expect the vendor to be all-knowing in the areas of gas properties. The problem of gas properties is the responsibility of the user. Much data is published on gases which together with high-speed computers make the job of defining gas properties somewhat easier today than it was at the timethe Code was written. This is not to say that all properties of all gases are as well defined as they might be. Gas mixtures in particular are always a problem.

Aboul lhe qulhor Rovcn BRowN is or? Associate Cottstdting Engineer with Dow Chemical, U.S.A,, Engineering and Constru,ction

clud,es u;ork

in

Seroices, Houston. He is responsible f or rotating ma,chinery specifi,cations, bid eaaluation, and technical asgistance for new equipment. Dut:ies also include d:iagnostic assistance for operating equipment and analEtical assistance for field, eaaluations and rebuilds ol rotating machinerg. Prior erperience ,inth,e control, pump, steam turbine and com-

A few suggestions come to mind. Also some useful references are listed at the end of the article. The BWR2,3 equation works well with m,any of the hydrocarbons. The Martin Ho3'a equation works well with refrigerants and chlorine. Once gas properties are established, correlation methods suggested by the Code provide results that are meaningful to the ultimate user of the compressor. ACKNOWLEDGMENT

_ Originally presentcd at the Third Sympmium on Compressor Train ReIrability, Manufrcturing Chemists Association, April 24, tg7c, Chrcago. -- --LITE,RATURE CITED

REPRINTED FBOM HYDROCARBON PROCESSING

13

Test compressor performance field the -in Compressors occasionally tail to meet pertormance exqectations atter instatlation and startup. When this occurs, tield testing is requiredHere's how to get the iob done efticiently and economicallY

due to excessive power consumption or simply that it will not provide the compression needed- Compressor performa.rc" is expressed in terms of brake horsepower required to compress a specified amount of gas flow {rom inlet pressure to discharge pressure and can be calculated from

H. M. Dovis, Delaval Turbine, Inc., Trenton' N.J.

The basic cornpressor characteristic .curve is described in terms of inlet flow, head rise and brake horsepower for each operating speed. A sample of a centrifugal performance curve is shown in Fig. 1. From each set of data, which is comprised of the seven items listed above, one compressor operating point can be calculated and plotted in terms of head, flow and power on the basic compressor characteristic curye. Normally five or six measurements are needed, with a flow adjustment for each, to completely describe the compressor curye for each speed. Usually field testing is not conducted on this large a scale, but rather the normal operating point is measured and the calculated results are compared to a manufacturer's sup-

Cr,Nrnrruc,tl- conlPressors are rugged designs that will perform satisfactorily fot years at a time betrveen scheduled rr-raintenance shutdowns. Flolvever, there are titnes rr'hen the contltressor's perfortnance mal' be

e

5n*

ft.,,

in

question

measurements of the following items: gas composition, inlet pressure, discharge pressure, inlet temperature, dis-

charge temperature, [as flow and speed. The type of instruments and method used to obtain this data have a direct effect upon the accuracy of the results. These requirements will be discussed in detail later.

plied performance curve. These curves are either based on the results o{ a shop test or if a test was not conducted prior to shipment, an estimated curve can be used.

$ lo* E r,000

The scope of the field tests depends on the extent of

the problem. For example, if the compressor performance is suspected to be drastically diflerent than it should be, the manufacturer will need as much information as possible to try to determine from these tests what corrective action is recommended. Many times the reason for the operating difficulty can be determined from the test results u"a if new parts are needed, they can be manufactured ahead of time before the compressor is opened. There are applications where the comP,ressor operating condition cannot be varied because of the process. The user and manufacturer must work together to establish the most meaningful test agenda before the test begins. This helps to eliminate misunderstandings and has proven to be the quickest way to get the job done. Another reason for checking the compressor's performance is to determine if it has changed from the initial it startup. If there is a change,

tr 6_

U

!2 !L

E F o

t0 0

2108

0f,Er cAPACfi. 1,O0

CFI{

Fig. 1-Compressor performance curve 14

bY may indicate that the comPr ed the intern or foreign material is and a rnainterlance shutdorr'll particularly true if nothing significant has changed in the pro..rr, but the comPressor has become the limiting factor in the plant's production.

FIELD DATA ACQUISITION

The acquisition of field test data, if meaningful and accurate results are to be achieved, requi.res planning. These field tests are time consuming inconvenient, and can be quite expensive. The extent of the testing agenda depends on the purpose of the test. If the purpose of the test is to prove or disprove manufacturer's guarantees, then by all means the manufacturer should be consulted on the test procedure and will probably want a rePresentative present during the data acquisition. It is imperative that mutual agreement be reached between the user and the manufacturer concerning the testing accuracy, number of test points, method of recqrding the readings, etc., prior to the actual test in order to have the final results fulfill the intent. When the purpose of the test is more of a routine

nature and is only to provide information for the user's benefit, it is still recommended that the manufacturer be consulted. The manufacturer will usually be willing to provide a test procedure that outlines the minimum requirements, including the type of instrumentation, necessary to achieve a useful test result and a procedure for calculating the basic compressor characteristic curveThe data that will be recorded during the test consists of readings of pressure, temperature, flow, speed, and gas properties. Measurements of. Power consumption are normally not available through direct readings, except when the compressor is driven by an electric motor, but rather are calculated from the other data.

Following

is a table of

instruments (per ASME

PTC-10)1 that may be used to test a compressor: Pressure:

Bourdon tube gages Deadu'eight gages

Liquid rnanometers Barometers

Temperature.' Mercury-in-glass thermometers Thermocouples Resistance thermometers Thermo'"vells

Flou:

Orifice plates

Venturi tubes Flow nozzles S

peed:

\{echanical tachometers Electrical tachometers Digital electricity frequency countcrs Stroboscopes

Gas

properties: Gas sample bottles Psychrometers.

The accuracy of the test instrurrents luust be verified before the test. Some suppliers specify the range of thcir instrurnents; these should be checked against an appropriate standard. Those instruments subject to changes in calibration during use should be checked before and after the test. The ASME PTC-10 also specifies that bourdon tube pressure gages be deadu'eight calibrated at approximately 5 percent intervals over the anticipated working range and that thermocouples and mercurl,-in-glass thermorneters be certified at 20 percent intervals over thc u,orking range. FIow measurements are obtained from either perlDanently or temporarilv installed plant flow meters. lVhen REPRINTED FROM HYDROCARBON PROCESSING

Iilt"T IBPMAIME

4 - tE sPACf,l

$nli€ g,

3TA110N6

tEG.

N[fr SIAIC PnESSUm

4 . TAPS,

SFACED

OO

Fig, 2-Field test instrumentation diagram

these metels are installed properly they provide flow readings that are suffciently accurate for testing purposes. Commercially available flow meters state on the nameplate the accuracy of the instrument and these limits can

be included when determining the over-all accuracy of the compressor test. When testing compressors that handle gases othet than atmospheric air, a gas sample must be taken during the testing to determine the volumetric analyses of the gas mixture. When the gas composition varies during the test, it may be necessary to obtain several gas samples to evaluate the composition of the gas mixture. The analysis of the gas samples is obtained from an independent laboratory after the test has been completed. The recommended location of the pressure and tern-

is shown by the diagrarn in Fig. 2. The cornpressor performance is to be evaluated from inlet to discharge flange and therefore the pressures and temperatures should be measured as close to these connections as possible. This diagram locates four connections on both the suction and discharge of the cornpressor where readings are to be taken. Due to the lirnitation of some compressor installations and the accessibility of the piping, it may not be possible to obtain four readings of pressure and temperature at each measuring plane in the pipe. Under these circumstances, the instrumentation is placed in the pipe in the best arrangement possible and the potential errors in.the readings are considered when evaluating the results. During the testing, the operating conditions of the compressor must be maintained as steady as possible. However, some small fluctuations can be tolerated without aflecting the accuracy of the test. The table in Fig. 3 is taken from the ASME Power Test Code-l0 and gives allowable fluctuations of test readings during a test .run. The ASME PTC-10 has also established allowable deviations for the compressor operating parameters that can perature instrumentation

MEASUBEMENT

INLET PBESSURE INLET TEMPERATUBE DISCHABGE PRESSURE NOZZLE DIFFEBENTIAL PBESSUBE NOZZLE TEI\4PERATUHE SPEED ELECIRIC MOTOR INPUT SPECIFIC GRAVITY TEST GAS LINE VOLTAGE

(l)

UII

I

PSIA

'8 PSIA PSI

'8

FtucTuATroN

{r )

2%

05

q6

2% 2C6

0,5

96

8PM

05

%

KW RATIO VOLTS

10qo 0 259t 2qo

PRESSUFE AND IEI,'IPERATUFE FLUCTUATICI'I Foff THE 0AS ExPft€SSio AS PESCEI{T OF AVERASE AESOLUTE VALI}ES

Fig. 3-Allowable fluctuation of test reaCings during a test

15

TEST COMPRESSOR PERFORMANCE

uiltT

vrnrlBrE

DEPAf,IURE

s8 2 24

PSIA RATIO

RPII

cnt

i

NI (2) (2)

l2l

BTSED ON I}IE SPECIBTD VATIE W]GRE PSETH'RES AIIO ARE ABSO.ITTE.

{I}. IEPIflNNES Afrf

'

TEHPMA]UNB

TIE oorBllED e) - 'I}lAfl

SFECT OF llEt{S

8 PER6II{T OPAf,NNE

Fig.

il

(+

(D) AilO (c} $lr'Lt GAS DBISITY.

t{0T Pf,il)t

G }xn€

ilLEI

4-41|o*"ble departure from specified operating conditions.

esist during a test and still yield valid results- These limits apply to compressors that are operating with conditions that'are different than the original comPressor design conditions. These allowable deviations are shown in Fig' 4. The stated departure allowances from specified operating conditions apply when the object of the test is to establish if the compressor meets the manufacturer's perforrnance guarantees. This table of allowances can also

be useful to the user. When the operating conditions o'f the compressor exceed these limits, it can be expected that the shapi of the compressor characteristic curve, including stable ringe and efficiency may be different than what will be obtained when operating with the specified conditions'

The most important objective of field testing is to obtain accurate measurements in order to calculate the true performance of the complessor. The following is a list of rules of good practice for testing.

o

.

Plan the test ahead of time. Prepare a test agenda that

will

Consult with the compressor manufacturer.

. . . .

Use the best quality of calibrated instruments.

Observe the data for consistency during the test.

recording data.

. Do not rush the test. CO'YIPRESSOR PERFORIYIANC TEST EVATUATION

E

The compressor's performance is calculated from the Ab,out the outhor

is the mnnager of

Cen-

Com,pressor Engineering De-

partment, DeLaoal Ttnbine Dioisiort of

DeLaaal Turbine, Inc,, Trenton, NJ. He is ,responsible for the organizat'ion

direction of all

engineering and

drafting functions related, to the prod-

in Januwg

16

uct. Mq', Dquis recei,ued, a B.S. d,egree ,in meclnnical engineering from Rose Polgtechrri,c Institute, Teme Hau,te, Ind,., in 1956 and a,ssumed lui,s present pos,i,tinn 1970.

Powcr

1971.

LITERATURE CXTED firmpreesoro and .Exhausters,"

miel Enrioerr.

AmerieriPetrolem Institute,

This paper was originally presented at the

o Allow the opefating conditions to stabilize before

and

"ASME

coowirht 1965 z "fdch"niot Dar

Second Turbomachinery Symposium,

Take measurements simultaneously for each test run. Record several sets of data for each test run.

DAVIS

1

accomplish the

o

trifugal

The procedures discussed in this article, although correct, have been simplified and are meant to be an introduction to the subject of compressor field testing. For those who are actually faced with this problem, it is recommended that the refelences be studied.

copyright

required objective.

Hucn M.

measured test data in accordance with well established thermodynamic methods. Section 5, of the ASME PTC-10 for compressors and exhausters, covers comPutation of the test results. The method that is used to calculate the test results depends on the properties of the gas being comPressed. The calculation is simple when the Perfect gas laws applyWhen dealing with a real gas, the deviation from the perfect gas laws must be considered. API Data Book, Second Edition 1970,2 is an excellent reference for the physical and thermodynamic Properties of gases and gas mixtures. Procedures are given for desk calculations with recommendations that some procedures be computerized. In addition to the above reference, the ASME PTC-10 lists a bibliography of 87 different authorities who have published literature on the subject of thermodynamic properties of gases and gas mixtures. The Code also cautions the engineer that considerable variation in thermodynamic properties can be found among the various published papers on certain pure, commonly encountered gases. For this reason, it should be agreed upon prior to the test what thermodynamic data is to be used to evaluate the test data.

Texas A&M University, College Station, Texas, October 1973. The Gas Turbine Laboratories at Texas A&M University have announced their Third Turbomachinery Symposium to be held on the Texas A&M University campus from Oct. 15-17, 1974. The Symposium will consist of lectures, discussion groups and tutorials covering all aspects of design, application, troubleshooting and maintenance of turbomachinery and related components. The object of the Symposium is to provide interested persons with the opportunity to learn the application and principles of various types of turbomachinery, to enable them to keep abreast of the latest developmen s in this field and to provide a forum wherein those who attend can exchange ideas. Enrollment will be limited to 700 participants so early registration is suggested. For more information contact: Dr. M. P. Boyce Gas Turbine Laboratories Mechanical Engineering Department Texas A&M University College Station, Texas 77843

I

Maintenance Techniques

How to improve compressor operation and maintenance Centritugal compressors can be designed and built with operation and ease of maintenance in mind. Here's what to consider when specitying a new compfessor Hugh M. Dovis, Delaval Turbine Division, Trenton, N.J.

Trrn pnocpss TNDUSTRv has experienced trenrendous glowth during the past trvo decades. The grorvth of sinele line and continuous process plants and the increasing use of automation have demonstratccl thc in-rportancc of component reliabilitv. \{achinerv uscrs are norv demanding dependabie performance, simplicitl, of operation and ease of rlaintenance. The sr-rppliers of centlifugal compressors have been lorced to rcvier,r' tlrcir designs and sometimes to design nerv equiplt-rent to satisfv these users' demands. This paper discusses the wa,vs in rvhich a centrifugal compressor can be built and used to satisfy a customer's operating and maintenance reqLrire-

.i

Fig. 1-Multi-stage fabricated case compressor.

ments.

Typicol rnuhistoge compressor design. A tlpical multistage centrifugal compressor, designed to meet a particular custorner's needs, is slrown in Figure 1. This machine consists of 9 inpellers in scries and is designcd to compress 4000 cfm of gas from an inlet pressure of 25 psi to a discharge pressure ol 425 psi. Each inpelier imparts velocity (kinetic) energy to the gas being ccmpressed. This velocity energy is converted into increased pressure in the difluser passage. The cross-ovcr passage and the return guide vanes lead the gas to the next impeller where the compression is continued. The volume of the gas stream is reduced as it is compressed and each stage is designed to accept a successively smaller florv.

Ronge of opplicotion. Fig. 2 is a curve rvhich shows the limits of applications for centrifugal compressors in terms of flow and speed. The speed is limited by thc stresses in the impellers. The small f1611,, high speed compressors have the same working stress levels as the 18

E

d = tr

STEAM TUREIt{E DBIVEE

c E o

a

MOT()fl GEAR ORIVER

H o o

-

StEA'tI TURBINE oAs TlJfEfiE TIOTOR GEAB ORIVES

r,m0 00FRE$$08 sPEm, 1000

Fig. 2-Application chart for centrifugal compressors.

COMPRESSOR OPERATION

AND MAINTENANCE

besides having high rotating speeils, is usually high pres_ sLlre as lvell. Shaft alignrnent is more critical since the

shaft

al pipe forces rnust

of the

ut the

irffi,

5::t."n'T;",1 clear:lnces

re_

;l#: of ihe

internal seals and bearings rnust be ,"vatched more closely ciue to their small ph),sical sizc.

.\ large com diffic The clearances less critical but

the

4, is more

e to and er)t

obtain, although more liberal toleranccs are acceptable, irot easv to nor.,e anci ipeciat lifting facilities ire required. The foundatior. for these lalge rnachines is also of special concern" Unless the because the components are

slrpports are designed, constructed. and maintained properlv the ntacllinery nra), never aclrieve trouble-ft-ee operation.

Fig. 3-Four-stage barrel compressor.

large florr, lorv speed nachines. 'fhe compressor applications in the lorv flou' range arc alnost entirel,v clri,,.en by motors ancl speed increasing The colrpressors in -qe:rrs. the mid-r'angc oi flor.r s are driven by rlotor-gears. steanr

turbines, and sone ges turbines. The large, high florv

all dril'en b1, steam turbines. The size ancl olterating spcccl oI :r centriflrgal compressor har e a dilect effect on the o1;cration arrd the maintaining of rhe conrpressor. A smal1 nrachine such es the onc sho.,r,'n in Fie. 3, compres-cors are practically

Rofor dynomics. I,Ioclern process colrpressors are built accordance rvith the API specification 617., One irnporterlt iten defined by this specification is the natural fi'equencies of the rotor. These natural frequencies must lrot occlu in the variabie speed range of the cornpressor. The dvnamics of a rotor can be studied r,vith the help ot' the computer and the effect on the rotor of operative Lrnbalence ch-re to build-up or ntisalignnrent c:1n be evalLLateci. Tliese rotor r-rnbalanccs r,vill load the bearings. CiompuLr:r analr'sis allorrs the engineer to predict these

in

beering loadings encl to desiqn a dependable m.rintenance free rnacirinc-

Iig. 5 shous the

c.-Llculated

and neasured rotor re-

Fig. 4-Large turbine-driven centrifugal compressor. REPRIilITED FROM HYDROCARBON PROCESSING

19

sponse curves for an eight-stage compressor rotor. The measured values were obtained first during the mechanical test of the compressor. Although the vibration level was less than 0.7 mils and the bearing forces were below the design dynamic load limit, the steepness of the vibration curve near the maximum operating speed was understandably cause for concern. The rotor was modified and retested.

Ioaded journal bearings, such as those used in compressors, can be unstable at high speeds, and a number of solutions to this problem have been used. The tilting pad bearing is widely used in compressors. Each shoe tilts independently to maintain its load carrying hvdrodynamic pressure wedge. Extensive service in many types of com-

the dependability of this bearing.

Off-design operqfion. Most

o'

compressor users take the1,

are concerned with compressor performance. Figure 7 shows a typical compressor performance curve. Uncomplicated and trouble-free operation can be expected in the stable performance region to the right of the surge line. Surging, or unstable operation, can occur in any centrifugal compressor when ,the inlet florv is reduced to approximately 6O/o of the design inlet flou, or lower. Compressors that produce large pressure ratios, ratio of inlet pressure to discharge pressure tend to have more violent surges. When the compressor is operated repeatedly or for prolonged periods of time ir.i surge the pressure forces can damage the internals of the machine. For those applications where frequent surge operation can be expected the compressor internals should be made of steel, instead of the more common cast iron n-raterial. When the comp.ressor is operated in surge continuously absorbs approximate\

a)/o of the rated horsepower,

+

,af

6

-

ANTI-NoDE 0F 2nd

+

MEASURED

CB|T|CAL

= E ,,, U

= ,.0 l6 & 6

o.N

I

0.6

i

o

wrrr ruon-comect H @

PICK.UP

r

1.4

250

= =

a@H

ul

!r

@ 2

ALIOWAELE DYNAMIC'

the compressor and in a matter of seconds,

if

the condiseals

which control the intemal leakage. The compressor performance suflers once the seals are damaged and the machine must be opened and the seals replaced to restore it to the original condition. Excessive temperatures in a compressor having a balance drum labyrinth seal made o[ a soft material with a Iow melting temperature can melt the seal. This will upset the rotor thrust balance and overload the thrust bearing. When the thrust bearing fails the rotor will shift axially and the impellers will rub against the stationary parts causing further damage. To help avoid these operating p.roblems the compressor can be provided with a high temperature balance drum seal made from compressed metal fibers that will withstand several times the normal operating discharge temperatures.

Another safety feature that can be employed 20

is

E o

ioo

E

g

sfEEo. r(tro nPil

Fig. s-Eight-stage rotor /esponse curve. 22r CALCULATED WITH

2'o

+-+

i

[,IEASI]RED WITH

1,0:

a

t.6 r

o

uG

'nl = 1.2.

E

loi'

'r50

d oaL

a

t@

0.6 0.,1

r

2

s 5 F d e

o,l-

0 f 2945

7I910

SPED. .IO(tr

RPM

Fig. 6-Rotor response curve for modified eight-stage rotor.

ll

UNSTAEtf

STABLE

r

110% SPEI

1

e F e u G !

N

fr E

V .

ll{LEI vtLLmE

+

ufilTtil,t 0f

st

EGE

Flg. 7-{ompressor performance curve.

high temperature switch located in the balance drum leakage pipe. It is wrong to locate this switch in the discharge pipe. In this location the switch does not pro-

tect the compressor since there is not sufficient discharge flow to carry the heat to the switch when the compressor is operated in surge. The balance drum leakage pipe is the correct location for this protective device. There is leakage flow in this pipe even when the compressor is being operated completely shut-off.

Normq! mqintenqnce ilems. It is reasonable to expect trouble-free operation

a

150

7C910r'tl2t3

however the flow thru-put is greatly reduced and under some conditions stops completely. The power required to drive the compressor in surge is therefore largel1, converted to heat. This causes excessive tempe.rature build-up inside tions are severe enough, can melt the soft labyrinth

yr

-

0.41.

E g

trouble-free mechanical operation for granted, but

it

I

I

These results are shor,vn in Fig. 6. The shop test shows quite low vibration amplitudes and low bearing loading. The compressor, once in acual service, u,ill become unbalanced due to build-up on the rotor. The diflerence betrveen the calculated and measured response curyes shows that this compressor will be tolerant to considerable rotor deposts before it r'vill have to be cleaned. Lightly

pressors have proven

CALCULATED WITH UNBALANCE AT

-

12t

for

periods

years between internal inspections

up to tw,o or three for heavy duty com-

COMPRESSOR OPERATION

AND MAINTENANCE

mercial machinery. I{owever, some applications where the gas stream being compressed is extremely dirty and internal washing cannot be used it may be necessary to shut down in order to clean the compressor internals to restore full flow capacity. The internal seals that prevent leakage around the impellers are normally of the labyrinth type. They consist of a series of circumferential knife points that are positioned closely to the rotating impeller. In order for the compressor to maintain the design performance these knife points must not be damaged by rubbing, erosion, corrosion, or plugged-up with foreign matter. These seals are normal u.earing parts and spares should be maintained in anticipation that they will need to be replaced after an extended operating period. Labyrinth seals are used in compressors with various design features that will extend their life and make

that approximately /3 of. all compressor failures and loss of production was due to malfunctioning oil film type seals ! Regardless of the cause, whether it was design, operation, or maintenance it points out a particular component of the compressor that commands respect. Oil film seals consist of basically two stationary bushings which surround the rotating shaft with a few thousands of an inch clearance. Seal oil is introduced between the bushings and leaks in both directions along the shaft. "O" rings in the seal housing prevent leakage around the outside of the bushings. The seal oil is maintained at some pressure higher than the gas pressure inside the compressor. The differential pressure across the inner bushing is usually only a few pounds per square inch to limit the amount of inward oil leakage. This leakage is collected in a leakage chamber that is separated from the gas stream by a labyrinth seal and is drained away through a drain trap. If the gas being compressed contaminates the seal leakage, the leakage is discarded" The outer bushing takes the total pressure drop from seal oil pressure to the atmospheric drain. Oil film seals can cause many diflerent types of operating and maintenance problems. The most common is excessive inward leakage due to increased clearances between the bushing and the shaft. When these clearances

become large enough the leakage chamber becomes flooded and oil spills over through the labyrinth and enters the compressor. This malfunction can cause many operating problems ranging from the nuisance of having

Fig. 8-High-speed balancing machine.

them less susceptible to damage. One popular version is to machine the knives in the rotating part and to position a stationary sleeve of a soft material around the points to obtain a seal. The clearances between the rotating and stationary parts of the seal can be reduced because it is intended that the rotating points cut grooves in the adjacent material under normal operation. This type seal allows the compressor to operate with higher efficiency due to reduced leakage and longer life because the thin knife points are made of steel and resist erosion. Also in this design considerable radial rotor motion can be tolerated without altering'the effectiveness of the seal. The stationary part of these seals have been manufacured in babbitt-lined steel, aluminum, non-metallic compounds, and compressed steel fiber materials. Each has its own advantages when considering the particular operating environment of the compressor. In many compressors labyrinths are also used for the shaft seals. Where the small leakage allowed by labyrinth shaft seals can not be tolerated, either the oil fiIm or the mechanical contact shaft seals can be used. Compressor shaft seals

of the oil barrier type present

potential maintenance and operational problems. A recent study of a large cross-section of cornpressor users showed REPRINTED FROM HYDROCARBON PROCESSING

to continuously add oil to the reservoir to contaminating the main process gas. The bushing clearances can increase in time due to corrosion or erosion of the inner surfaces of the bushing, wiping due to radial shaft vibration, or dirt in the seal oil. Another shaft seal, which is not as common in process compressors, is the mechanical contact type. This seal has the advantages of being able to maintain low inner leakage rates with higher oil-to-gas differential pressures and therefore makes the seal oil pressure control system simpler. The seal has a spring loaded carbon face that runs against a face on the seal collar. The complete shal't seal either consists of two mechanical contact seals in a back to back arrangement or a combination of the mechanical contact seal on the gas side and a bushing seal on the atmospheric side. These mechanical contact seals have been used for many years to seal against pressure up to 1200 psi in natural gas service. They presently are being tested for pressures as high as 2500 psi differential across a single sealing face. The mechanical contact seal has an added advantage when applied to high pressure applications in that the radial sealing faces have a minimum effect on the rotor dynamics, whereas oil film bushing seals can lose their free floating featrue and cait upset the stability of the rotor when operating at high speeds.

High speed compression machinery must be properly balanced, especially when it is designed to operate between the Ist and 2nd lateral critical speeds. The correct method is to first dynamically balance each impeller

on an arbor, and to check the rotor balance as each impeller is installed on the shaft. The runout of the shaft must be watched closely as each of the impellers are shrunk on the shaft. Abnormal change in shaft runout indicates that the impeller is not square on the shaft and 21

must be adjusted before checking the balance. \\'hen the balance check indicates that an unbalance exists, the correction is lnade to the last impeller that rvas mounted. This procedure when strictly follolved can ptoduce a rvell balanced rotor even tvhen the balancing is clone in a lou-speed balance machine. Some users routinely check the balance of complete rotors and make corrections on the first and last impellers. This practice is lvrong for flerible rotors that have more than three impellers. The onl1, sure rvay to balance a completely assembled high spced rotor is to nse a high speed balance machine. The rotor unbalance rnust be checked throughout the operating speed range and any corrections that are made to the rotor must be macle at the plane of unbalance. N'Iost balance shops in this country have balancing equipment that operates below 1500 rpm. The photograph in Figure B shorvs a balance n'rachine that will balance a 1000 lb. rotor at speeds up to 7000 rpm. The machine r,vill handle most high speed flexible compressor rotors.

fouilflilO

TOOLT

rloul?aD couPLlro

!

A

Fig. 9-Hydraulically-mounted coupling and tools.

Experience has sholvn that high speed compressor rotors can be successfully balanced in a lorv speed balancing machine when the progressive impeller by in-rpcller method is followed. The compressor is normailr' subjected to a mechanical test in the shop before shipping. During these tests the vibration amplitudes and frequencies are measured. When the compressor vibrates

excessivelv the rotor is rebalanced in the high speed balancing l'iachine, hor'vever this is normally not required. When a spare rotor is purchased or repaired after the compressor has been installed the rotor should be given a high speed balance check in order to prevent delavs later in the users plant. Thc alignment betrveen adjacent shaft ends of all rotatinq equipment in the compressol train is very inrportant to the operation and maintenance of the equipnent. -\11 compressors, gears, motors, turbines, etc. havc some toielance for misalignn-rent. IIow-ever, except for rotor balance, misalignment is the most frequent cause of running problems. Excessive misalignment can force the rotating equipument to vibrate and shorten the life of the bearings and gear tvpe couplings. When misaligned shafts are rotated, the gear teetli in the coupling must slide back and forth on each other. This causes bending lxoments and forces to be imposed upon the shaft ends. The shaft end rvill fail if the conditions are severe enouglr or the coupling hub rvill come loose on the shaft. \,{ost couplings are mountcd on a tapered shaft end. It is con.imon that a key and keyway be provided to transmil tite torqne through the connection. The coupling

hub is prrshed upon the shaft taper by a nut r,vhich also prevents it from coming ofl the taper during operation. This trpc of coupling hub mounting is satisfacton, for lorv speed-1or,v horsepor'ver applications but can be a limitin,r factor for modern high speed machinery. A much better rnethod is to shrink the coupline hub upon the shaft taper- using hvdraulic pressure and jacking tools, as shorr-ir ir.r Fignre 9. This type of mounting allows for at least trrice as much torque to be transmitted through

the sarne size shafts with the same stress levels. This is possible since the key is no longer needed and the key'way

is eliminated. The misalignment forces frorn the coupling teeth rvill not loosen the coust.ress concentration

pling hub on the shaft because it is fitted on the shaft taper rrith considerable interference. To mount a colr22

*ffi* Fig. 10-Compressor case with sliding mounts.

pling hydraulicallv, the shaft end is drillccl to pr.ovidc ar.r oil passage for pun-rping high pressure oil betwccn tlrcr shalt taper and the coupline hub borr:. A liyclrarrlic hanrl pump is connected to the shaft enr.l throrrsh an aclapter'.

is strctcl-ied u,itli hyclraulic pressurr: ancl the jlck scre\vs (Figure 9^\), are used to push thc i:orrpling hrrlr t1p on the taper a specifiecl distance. 'l'he jnr:k scre\\'s hold the hub in the correct position r.vhile thc oil prcssure is released. After the oil has drained the nrounting plate is remor,ed, a nr-rt is rnourted on thc shalt end. '1'1're hLrb

(Figurc 98) to protect thc thrcads. To rcnrove thc coupling hub the procedure is reversed and the hub is "popped-off" against the tools with hyclrarrlic plesstrrc, 'Ihis type sf s6upling n-rounting pro.,,icles Ior lrr casill' mounted hub that will transn-rit as nrur:h torclLrcr ns tlrc sl.raft material u.,iIl allow.

Good shaft alignment mr.rst be mairtairccl uncler dr'namic conditions if trouble-free operation is to be aclrievetl. To accomplish this the thermal srowth of the incliviclrral elements must be taken into account when aligning compressor equipment in the cold conditions. |or instance, the diameter and length of a compressor case will increase clue to the heat of compression. Normallv the casing is supported at the ]rorizontal centerline to allorv the case to grow irr cliameter rvithor.rt

changing the position of the shaft. 'I'he mountins feet on one end of the case are bolted and doweled to the foundation, as shown in Figure 10. The opposite cnd of the case must be allowed to move axially as the casing

COMPRESSOR OPERATION

AND MAINTENANCE

Aboul the qulhor Hucs M. D-q.vts is tlLe manager of Centrifugal Cotnpressor Engineering DeprL,rtnrcnt, DELAVAL Turbine Diaision of DELAVAL Turbine Inc., Trenton, N.J. He is responsible f or th,e organization and direction of all engineering and drafting functions related, to tlte product. Mr. Datsis receiaed a B.S. in ME degree from Rose Polytechnic Institute, Terre trIaute, Ind., in 1956 and, assumed ltis present posi,tion iru January 1970,

provicles centerline sllppolt for thc case, has four mounting pads that are all boltecl and dor,veled to the foundation, and requires no maintenance. The ruggeciness of tiris support n'ill allou, for consideraltle extelnal forces to be cxcrtcd upon thc casing rvithout changing tlie shaft alignment"

lnternql configurqtion. The process market places very clenranding arLd ever changing requilernents upon the selcction and alrangcrlent of the compressor inteJnals end erternirJ casing nozzle configuration. Centrifugal coiuplcssol' selections should be made rvith the ease of o1-relation ancl rlaintenance aspects in mincl as well as tire cornpressing requirerlcnts. Figure 12 shows the most tr irical allangernents of tlie cornpressor internals and ouLcr casirg in thc folrn of simplified diagrams. f ire opelai-ion and mailrtenance aclr,antages are listed

Fig. 1l-Compressor case with centerline support.

I

l

PrnAun-

wsRow

ndlf

fol e:rch as follows: i. Sing-le corrl)ressor- bocly instead oi tlvo or

$Enoil lll0ss

&4SrC

corrEtsoa

ry* ?

I

PARAI.IJT

SEffEs FTO' oNE

fLow

0(n.lrs

sl,cf,fll

PffiI

@

$r

CE*IEi

oo@oo

ll

II

oll fil0

r::,LtNG

)i1t:i Plrltti

1)lYtfai

r

!Y'o-l

lt

H

tenancc problems.

8Eftlsl R.Oil

Wrll

7. Singlc inlct to better suit ertcrnal piping arrange-

OOU8(!

frov{ rIlLEf Ar.r0 sr0l sTREril

ment.

oo

B. Single discharge

Fig. 12-Eight-compressor case and nozzle configurations.

expands. This has been accomplished in the traditional compressor mount with holddown shoulder bolts that allow the casc to slide axially upon lubricated shims but limit the vertical movement of the case. With this arrangement a ve.rtical key and keyway are located between the case and foundation on the vertical centerline to prevent transverse movement of the case. This mounting requires a special founda.tion to support the vertical key and also regular lubrication of the shims. The compressor case mounting shown in Figure 11

allows for thermal growth

more

in a sinrpler s)'stem. 2. Hot dischage at center of case to reduce lubricai.ir:l :Lnc1 oil scal problems. 3. Reduceci po\ver- r'equired to cornprcss gas results in a srnailcr ch'iver. 4. Back to bacli irnpellcrs rcduce natural rotor thrust and nllorrs iol mole internal seal \\,ear befole overloading thrust bearing. Increase tirr-re betrveen overhauls. 5. Cold inlet at center of case to reduce iubrication and oil scal problerns. [i. Srnaller colnprcssor anci higher specd to do the sarDe compression job. Rcduced foundation ancl trainresult,s

of the case in all

REPRINTED FROM HYDBOCARBON PflOCESSING

directions,

to better suit external piping

arrangement.

9. No external balance piston leakage pipe. Compressor can tolerate increased balance seal rvear r'vithout upsctting thrust bal:rncing s1'stern and over'loacling thrust bearing. 10. Hot or colc1 sections oI case arc adjacent to reduces thermal gradients and distortion of the cese. Makes alignment easier to achieve. ACKNO\VLEDGN{ENT

Origiually prcsented at the lst Tcxas A&lV{ University Turbomachincry

Slmposium, October 1972, College Station, Texas. 1

LITERATURE CITED for General Refincry Service, Edition. 1963, Amcrican Petroleum Institute, Washington, D.C.

API 617-Centrilugal

Compressors

Second

23

lmprove machinery mai tenance Reliable operation ol modern turbomachinery requires up-to-date mai ntenance technrgues. Here ate some ideas tor improving your maintenance procedures W. E. Nelson, Amoco Oil Co., Texas City,

Texas

MaNtrrecruRE AND MATNTENANoB of turbomachinery are completely different. The first involves shaping and assembling of various parts to required tolerances while the second involves restoration of these tolerances through series of intelligent compromises. This is the crux of maintenance techniques-keeping the compromises intelligent. The ,process industry has pushed for "bigger and better" turbomachinery until ,operational problems have become tremendous. We are literally "snowed under" by these problems. The failure to provide adequate feedback of reler-ant operational troubles into the design phase is the greatest problem facing the turbomachinery industry today.

a

This lack of ,communication results in much turbomachinery designed with too little regard for the operating and maintenance complexity created. Many of our maintenance techniques are, in a word, inadequate to cope with the troubles we encounter. The mechanics in our various installations are not, in general, technically competent enough on the complexities of our equipment to adequately maintain it under unit operational pressures. The growth of contract maintenance firms has not been sufficient to fill this void. The ability of the original equipment manufacturer (OEM) to provide technical service has also frequently been inadequate. Perhaps some of the pnocedures used at Amoco to combine the best of our capabilities, those of the OEM, and those of specialty service organizations to meet our problem will be of interest to others facing similar situations. 24

The Amoco refinery at Texas City is forrltlr lareest in the nation in crude capacity. Ilt:causc of the sevelitl' o[ some of our reforming plocesscs, generation of 60 nrcglu'atts of power, and large ammonia production it larrks higher in complexity of cqrripment. Much oI our equipment is less than ten years old. It is sophisticatecl, lalse arrd complex. Since modern design trends tolvlrd sin.-le train equipment, we have some of the largest cqrripnrerrt made. The rapid growth of tliis refinery has createrl tlemendous pressures on our nrainLenance pcrsonnel and experience has become extrcrnely scarce. At the l)resent time over two-thirds of the macliinist hourly lrersonrrel har-e under six years experience in the rnachinery ficld. Having problems, we have bccn forced to find solutiotrs. These solutions can be diviclccl into forrr brrsic categorit's:

1. Training of personnel

2. Tools and equiprnerrt 3. Replacement parts

4. Reliability improvement projects TRAINING OF PERSONNET \Ve believe that training must be the central themc to the solutions of our problems. The days of the rucchatric

arned with a ball-peen harnrner, a screwclrir,'cr, ancl a crescent wrench are gone. More and more conrplicated maintenance tools must be placed in the hancls o[ tlre mechanic and he must bc trained to utilize thern. People must be trained, motivatcd, and directed so tlrat thei' gain experience and develop, not into meclranics, but into highly capable technicians. Good training is cxpensivt: but it yields great returns. Machinery has srown more complex, requiring more knowledge in many areas. 'Irhe olcl traditional craft lines must yield to the rrraintenance needs of complicated equipment. A joint eft'olt by craftsmen is necessary to accomplish this.

Bqsic mqchinist trqining. To solve otrr ploblenrs rve have embarked upon ambitious training proglrnrs. Since

IMPROVE MACHINERY MAINTENANCE

niques are rvritten up and distributed to all machinists frequently. The ground rules of this "nervspaper" limit it to two typervritten pages and use of reduced size sketches andf or drawings is encouraged. In addition, the machinists are given copies of all machinist-oriented issues of a similar training newsletter issued by our Engineering and Technical Division. We feel that knowledge breeds more knorvledge.

Mqnufqclurer trqining. Our personnel are sent to

Fig. 1-Models used to teach reverse indicator alignment. 1967, over 100 nerv machinists have completed an intensi-

fied training prosram involving 800 classroom hours of instmction in machinery principles and concepts. Ntlost of this training has been developed and conducted by in-plant personnel. It has been expresslv tailorecl to our equiptrlent and is highly detailed. Training rnust be carefully planned and adrnirristered to fit the requiren'rents of each situation. At present u,e have a full tiurc training Staff of two people in our maintcnance division. The responsibilities of the group range across all crafts. These machinerv related subjects have beren covered recently:

o . o

Rcvelsc indicator :rlignnrent seminars (see Fig. 1) Gas turbine overhaul

Mechanical seal ltaintenante 'a rideo-tape n'as developcd in orrr corporate train;ng grouP.

Prqcticql froining. Aftcr the machinist receives his fundamental training, his on-thc-job expcrience should continue his training and test his skills. For this reason \ve attempt to clo as much repair work u'ith our own people as possible. Each man is rotated among various jobs to accelerate his learning and development and is as familiar with a Iarge cornpressor as he is with a small pump. Over half of our shop personnel can operate a balancing machine expertly. The remainder havc worked on the machine and arc farrriliar with its o1;eration. The spreadins around of the hardest jobs clevclops nrore competent people. If rve restrict a man to onc type of rvork, he r,vill probabl,v become expert in tliat arca but his curiosity, u.hich rve feel is a prime nrotivator, will er.'entually fadc. \Ve cncouragc the participation of the machinist in the solr.rtion of difficult problerns. This policy often causes the machinist to seek out information on his orvn. References to API spccifications are not uncommon on our shop floor. To encourage this type of participation we have a

library adjacent to the shop floor with filed drawings, written histories of thc equipment, catalogs, and other literature pcrtinent to thc machine maintenance field. The "librarian" and "editor" of the weekly summarv of repair histories is an hourly-paid machinist. The job is rotated periodically to develcp interest. FIe also alters equipment drawings and service manuals as revisions and changes are made.

Updote lroining. In order to pass on knowledge and to update the mechanic, unusual experiences or special techREPRINTED FROM HYDROCARBON PROCESSING

manufacturer-conducted training sessions. As an example, a number of our people, including first-line foremetr, have attended the very excellent maintenance school conducted by the manufacturer of the six gas turbines we oPerate. This is a very effective effort on the part of the OEM to provide maintenance instruction on rnachinery to the customer. We would like to see more along these lines.

Servicemen qs trqiners. \Ve look upon nanufacturers' sen,icemen as trainers. Their primary reason for being in

our plant is to train our people. Because of thesc t,iervs, our standards for this type of u'ork arc very high. \Ve attempt to get lasting value frotn their services. LrnfortunateJv, the equipment manufacturer frcquently suffers from the san-re problems that 11's 121'g-2 shortrrge of capable people. For this reason rve train our own lteople as much as possible. We havc one loreman rvho is ottr in-plant serviceman. He supervises no hor-rrly personnel, but acts as a consultant to maintenance jobs. A three-man technical group devoted to turbornachinery rvas establishcd in 1971 to further our internal capabilities.

!nslruclionql books. Manufacturers' instruction

books

are often inadequate. \\Ie have resorted to writing maintenance rnanuals for the mechanic on such subjects as rnechanical seals, vertical pumps. hot-tapping machines, and more recently, gas turbines. The gas turbinc overhaul manual consists of :

o

Step-b)'-step overhaul procedures devcloped largely

from the manufacturer's training school mentioned previously

. Almost a hundred photoeraphs illustrating the stepby-step proccdures using one of our gas turbines

o An arror'v

diagram showing the sequences of the pro-

cedures.

bound into a book form and used as thc basis session to our supervisors and mechanics that combined the best of the experiences of all

This

r'vas

for a four-hour training our people.

Detailed drawings are developed to aid in nrainterrance. Our first large scale effort involved a contact seal assembly which was dcvelopcd after rve held up thc operation of our giant catalytic cracking unit because the "t1'pica1" dimensionless drau,ing supplied bv the OEM rvas not adequate to correctly assemble thc compressor seals. Many other assembly drarvings have been der,eloped since that first effort.

TOOTS AND SHOP

EGIUIP'YIENT

Many maintenance tools should be made available to the mechanic. We attempt to do as much specialized machine work as economicalll,feasible. Fig. 2 shorvs a cooling 25

water pump being machined to replace a suction flange that r.r''as broken. Doing jobs like this keeps us technically compctent enough to tackle other jobs. The catalytic cracking unit air blower turbine nozzle blocks shown in Fig. 3 r^"'ere fabricated in our shops in eight shifts because a ne\v one could not be delivered in under three months.

The "homemade" nozzles have been in service for several months with no loss of efficiencl'. In-plant balancing capability is one of the most valuable tools in a petrochemicals plant. Our 4,000-pound rotor capacity balancing machine has had a very significant effect in development of know-how. Until about six years ago, we relied entirely upon outside balancing facilities and only balanced the more critical equipment. After purchasing balancing equipment, we began to practice on equipment that l,ve once felt did not require dynamic balancing. Vast improvement of pump seal lile as rvell as bearing life on some of the smaller equipment was

\

F

Fig. 2-Versatile tools develop skills.

achier-ed. The static balance of pump impellers is woefully inadequate for long runs and reliable service. In brief, the balancing machine paid for itself in balancing pumps, motors, and small turbines that r,r,,e once thousht did not need dvnamic balancing. The d,vnamic balancing of impellers of ventical pumps is cspecially advantagcous in reducing maintenance. The long'. slender shafts are highly susceptible to any vibration induced by imbalance. Thc man familiar rvith the plant equipment tends to balance closer if the machine's operational history has proven to have a balance sensitive rotor', another advantage. This sparks ihe development of that much sousht-after knou,-

how.

fechniques to check balance gear-t)'pe couplings for the large high-speed con-rpressor and turbine drives as a unit were der-eloped. This led to the solving of many vibration problems. High speed couplings are routinely checkbalanced norv.

SPARE PARTS Spare parts problems are inherent in the machinery maintenance business. Cost of replacement parts, long deliverv times, and quality are problems everyone has faced. Amoco, at Texas City, undsylook to combat this problcnr by forn'ring a spare parts group to consolidate the spare parts activities for its expandins rgfinsly on a companv-rvide basis. We believe in stocking spare parts: our present inventory is over 16,000 items, including 75 complete rotors. The field of spare 1 arts is changing very rapidlv and is much more complex rhan ir the past. \{any pieces of equipment on process units are made up of unitized components from several different vendors. The

traditional attitude has been to look to the packaging vendor as a source of supply. \{any vendors are refusing to handle requests for replacement parts on equipment not directly manufactured by them. More and more specialty companies are entering the equipment parts business. Sorne are supplying parts directly to OEM companies for resale as their "own" brand. Others are supplying part5 directll to the end user. The primary function of our spare parts group is to develop multiple sources of supply for as manv parts as possible. Gaskets, turbine carbon packing, and mechanical seal parts are purchased from loca1 sources almost entirely. Shafts, sleeves, and cast parts such

l:rt;5l;,,*,

are becoming increasingly available from local

We have found some OEN{'s are altering their spare 26

Fig. 3-Shop-made nozzle block in semi{inished condition.

parts system to improve servicc clue to this local compctition, definitely a bright spot in thc lticture.

MACHINERY REIIABIIIIY I'YIPROVEMENT High naintenance costs and lorv operatinlr reliabilitv go harrd in hand. Usually thc lorv rcliability is a (rceter economic factor than the high rnaintenanr.e r:osts. Ahnost one-third of the unschcduled (ancl costly) shutdor,vns irr our refinery are caused by machincrv failures. Machine "revamps" and alterations h:rve been necessary to inrpror-c

reliability. A brief discussion of somr: of this tvpe of rnaintenance r,r'ill give a feel for some of thc benefits of our training, spare parts prosrarn, ancl our major essetpeople.

Governor redesigns. Due to a long history of failures and poor performance u,ith our four process sas turbine governor control systents, rve undertook to redesisn these systems using state-of-the-art electronics and "plug-in" concepts for ease of maintenance. f'he first s1,51s111, dubbcd "Turbotronic," was installed in 1969. Since tliat time two All of these iustallations have been highly successful in that rnaintenauce has been minimal and is usually accomplished on-stream. Also, turbine performance, speed control, and flexibility other gas turbines have been converted.

are greatly improved. The original design has been supplemented to include a self-contained alarm system, a sen-riautomatic sequential start system, and a con'rplete trip and protection system as r,vell as the elcctronic controls. Our

IMPROVE MACHINERY MAINTENANCE

eost oI this svstcrn is considcrably less than the cost of a sinrilar device olTered by thc OEM on nerv ruachines. The controls for the fotrrth and final conversion are due for installation in laic 1973. The original installation r.vill be dismantlecl ancl an trpdated version installed at the same timc. Toial horscpower involvcd is alnrost 80,000. Witlr thesc successful installations, rve turned to stcam turbines. A conrprcssor clrivc on oLrr most vital process unit lracl been plaaucd bv sovernor dri"'e reduction gcar problerns. The tLrrbine speed (approximatcly 9,000 rpm) ivas

drive via a three-shaft gear train. \/ibration induced by the poor design limited the maxirnurn rlrn to about thrcc nionths. The \\roodrvard goverrcduccd

fc.rr gorrernor

nor and the gearing were rcmovcd and our or.vn electronic

sovclnor installcd. This governor n'as built, installed, and test run in foLrr days. Since the installation in May 1971, vibration Jras been minimal and governor performance has be