TURBOMACHINERY HANDBOOK
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Reprinted from HYDROCARBON PROCESSING
o Gulf Publishing Company o @ 1974 . $t.ZS
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TURBOMACHINERY HANDBOOK
Published by HYDROCARBON PROCESSING This reference manual has been reprinted from the regular monthly issues of HYDROCARBON PROCESSING. Other handbooks and manuals in the series are:
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Table of Contents
TURBOMACHINERY HANDBOOK Page No.
i I
Centrifugal Compressor Performance Testing Better mechanical testing can improve compressor reliability Test compressor performance-in the shop Test compressor performance-in the field Maintenance Techniques ....... How to improve compressor and maintenance lmprove machinery mainten Why clean turbomachines A closer look at turbomach Hot alignment too complicated Turbines using too much steanl"? ... Better pump grouting .1..... Centrifugal Pumps .. !-. i.. How to improve pump How to control purnp vibration How to prevent pul Turboexpanders ..1 New develop hot gas systems Turboexpanders Turbotmachinery and C GasS histories of : causes Coniprq$sor prol Critical. Related Are couplings Itnik in How tr I
Dlag A new
c+ t,
ffi,t p6il
{ --- t' 4
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5 6 11
14 17
18 24 28 31
35
38 41
45 46 49 56 61
62 66 71
72 76
83
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d*e:
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<:n cxcellent.
To date, eiqht more nrulti-stage steam turbines driving pro( ess conlprcssors with a total horseporver of 58,000 havr: been converted. These electronic govcrnor installations agair-r are less cxpensive and rnore effectivc (in or-rr opinion) tlran any conrmcrcially available product. Beoring conversion. Rcaring failurcs rank high
oI rotaling equipment failrrres but bearing
as callscs
design has
Iaggccl belrincl as rotational speecls hare risen. Parameters
fol bearing design are u'ell defined by Herbage' and Abrarnovitzr. It has becn dernonstrated that the reliabiliti, ol rotatine ccluipn'rcnt c.an bc increased through the use oI bcaring dcsien irnprover)rent. Dcsign modifications
fa1I
irrl,r tirret' nre.j,,r' 11 p65, 1 . Tlier clir ngins of taPcrcd Iancl thrust bearings to tiltirre pacl thrust bcarings rvith leveling links: Experiencc orrrs ancl othtrrs (:t, 'tl-in dicates that for equal an'as. irnprovcments in loecl-carrying capabilitv can be ac
hievccl. Srrclclen lo:rd surses and liquid slugs
rvill
occur.
Thnrst lrcarinss r.r.liich do not have the ability to absorb such lo:rcls r.r,ill Iail; tht:rclore, rvc fcel rve should have thc nrarinrrrtrr possiblc thrrrst c.apacit1,. 2. Changine radial bearinss fr:orl cvlindrical and/or prcssrrre prcl babbittcd slccve t,vpe to tiltina pad L,earings. 1-his irrrProves rotor stabilitr', elirninatcs chances of oi1 uhill rnd irrproves srrrlivll oI jor-rrnal bearings under lrtor: inrb:rlance and dtrring <:r'itical speed drvclls. Present state-of-the-art indicatcs thrrt the tilting pad bearing is scnerallv tlrc lrest in bcaring clesign because of self-stabil-
bearings. These type bearings are generally supplied with steel-backed babbitted pads. Changing to copper alIo,vbacked, thin-babbitted pads conducts heat away from the babbitted surface at a faster rate than the steel'backed shoes, improving load-carrving capacit,v. Experience indicates that a 50 to 100!t load-carrl,ing improvetrtent can be obtained.
Over 150,000 horseporvcr has been ccnverted since late 1969. Two dozen more machines have been recommendcd for conversion. We do not believe we should \vait for a failure before deciding to change because of the process penalties resulting from a bearing failure. The conversions have been done using the services of some of the best bearing specialists in the country. Amoco participates in the fundamental design by supplying basic dimensions and some machine peculiarities. One equiltment manufacturer has begun to offer bearing upgrading "kits" for the machines alreadv in serr,ice. Again a bright spot in the over-all picture. Rotor redesigns. \\/e havc found that many steam turbines of stacked rotor design are built by the manufacturer r.r,ith substantial imbalance couples. It has proved advantageolrs in manv instances to unstack a rotor, indil idually
balance each lvheel. and then proeressivell, balance the rotor as it is reassembled, the rvay most exPerts agree it should be done. Manr machines, once highly sensitive to vibration, have been t'trade lttore reliablc by tliis technique. In addition, the shrink of the rvheels on tlie shaft pror,ided by thc manufacturer olten proves to be inadecluatc for the steam operating temllcratllres. This allows the r'r"heerls to move on the shaft. Shafts have been redesigned to pror-ide locliinq rings for each rvheel. Six rnachincs of one make havc been rerlanufactuled to date. Continual impt'ovcrnent and ulldating of our mac.hiner,v is necessary to maintain long runs. We can no longer "fix like r'r,e did that last time" but must colrtinually strive for newer. bctter r'va1's to make repaiLs. T'he rnachinist is often called upon more frequently to rl-rake revisions than he is to "fix" a machine. 'fhere are no simple solr-rtions to the maintenance difficulties we face. The principal ingredients of strccessful turbomachincry rnninlgnance are training, eqtripment, parts, and people (both hoLrrli' and technical) . T'hese factors must be couplecl ',vith a determined managenent if rve are to move ahead and find solutions to thcsc ltrobIems. Our experience has been that each of the foLLt' areas contributes to our solutions. A11 are nccessary.
izins charac tr:r'istics. 3. A matcrial chause of tilting pad journals and thrust
Today's process jndustrr- cannot be dependcnt ul)on anv one source for solutions to its troubles. Original equipment manufacturcts, contract service shops, contract trtaintenance companies, and specialtv companies can only supplcment the basic skills that tuust be developed rr-ithin orlr o\\ n organizations.
Aboui lhe oulhor W. E. Npr,soN is Manager of Maintenance Seraices at the Amoco Oil Com-
Orieinallv prcscnted et the 2nd Annual Texas ,A.&tr{ Turllornaclrinery
pnny refinery
in
Teuas CitE, Teras. He Amoco f or nineteen gears
has been uith i,n oarious enginecring, purchasing, a,nd tnai,ntena,ncc assignments. " E d" rece'ia ed o, B.S" degree in Mechanical Engineet'-
ACKNOWLEDGMENT Sympo.irm, October 23-25. 1973. Col)ege Stati.n, Texa..
BIBLIOGRAPHY
ing from Teras A&M Uniuersitg in
1951 and, is a Registered Professi,onal Engi,neer in Teuas.
REPRINTED FROM HYDROCABBON PROCESSING
27
Why clean turbornachines onstream?
Cleaning turbomachinery while in service is tricky but the benefits are signiticant. Here's how to clean your machinery sately and eftectively Briqn Turner, Imperial Oil Enterprises, Ltd., Sarnia, Ont.
"ON-srneau CleeNrNc" is defined as the periodic of accumulated deposits while the equipment continues in service. This definition excludes from the discussiorr the various techniques of preventing the deremoval
posit accurnulation. It also excludes those cleaning methods rvhich do not require disrnar.rtling, but do require that the machine is either idling or stopped. The paper cliscussed the rcasons for cleaning, fouling indicators and abrasive and solvent cleaning techniques. Details of turbine washin_q are used to illustrate the applicetion r,I solverrt cleaning.
Why sleon? There are at least three reasons for "onstreanr" cleaning. The first is to restore the system capability. If the unit is a driver, it's naximum horsepower wiil probably drop as it becomes dirty. Cleaning will restore this limit. If the machine is a dynamic corlipressor, the fouling may have reduced its head, and therefore, the maximum gas flcw rate. Cleaning r,v'iil restore the capacity limit. The second ieason is to increase the machine's efficiencv" fn most, but not all cases, fouling rviil increase
the frrel or power required to do a certain task. The flow contours, R.emoval of the deposits u'iil restore the originai nrcfiles and thr efficiencv. The tl:ird leason for cleaning is to prevent failures due to abnormai operating modes. Fouling on the rotor blacies of stearl turbines can cause ttrrust bearing failures. Deposil-s on steam turbine governor valves and trip and throttle valves are suspected of causing overspeed failures. Foulins in balance piston labyrinths and in balance lines has caused thrust bearine failures in centrifueal rnachines. mecha:-rism is tharl the deposit changes the
28
Any rotor deposit can cause vibration due to unbalance if it is not laid down uniformly or if it sluffs off nonuniformly. There could be other similar effects which will cause failure of the unit.
Fouling indicqtors. A prerequisite of a cleaning program is some kind of fouling detection system. Naturally, thrs system must cover the prime reason for cleaning, If the machine is a gas turbine, then the prime reason mav be horsepo'rver capability or it may be efficiency. On the other hand, on a back-pressure steam turbine, the prime reason may be either horsepower capability or
thrust bearing protection. On a centrifugal compressor. the prime reason for cleaning may be to restore capacitl-, to improve efficiency or reduce thrust loading. The selection of a fouling detection system will be strongly influenced by the safety and complexity of cleaning procedure. For example, the procedure may be to throw 10 pounds of rice into the suction of a gas turbine. Or, it mav involve injecting a quart of water into a ""ingle-stage, mechanical-drive turbine, with a 30oF superheated inlet. In either case, ihe risk of damage and the lnanpower required is so low that no monitoring can be justified. The cieaning should be frequent and routine. On the other hand, the cleaning may invoive renoving 3000 of superheat from 200,000 pounds/hour oi sieam entering an eight-stage turbine. This is a much more compiex case. If the turbine really is dirty, and is not washed, you may end up u,ith a nreck. Alternativelr.. to wash, especiaily the firit time, ma1, result in misa.iignment due to piping stress, water slugging, ioss of clearance due to differentiai contraction, vibration due to nonuniform cieposit removal, or thrust trearing failure. This case obviolslv calls .trr a reiiable indicator. Irouling indicators ir"hich have been used
inch.rcie:
o
Gas turbine exhaust temperature
rD
Steam iurbine stearn cirest Dressitre (singie governor vatrve)
.
Steam turbine P1 pressure (multi-valve)
. b and c above modified to correct for steam
flow
o The exponent n-1/n on a compressor or gas turbine where k is either known or is relativell constant o The
exponent n-1/n in one section of a machine relative to another section handling the same gas
o The pressure ratio in
one section of a machine rel-
ative to another
. Thrust loading or thrust .
bearing metal temperature
Balance line to suction differential pressure.
Cleoning techniques. There are two basic approaches to cleaning. These are abrasion and solvent cleaning. Abrasion is the simplest of the trvo methods, but is usually the least eflective. The more common abrasives are nut shells, sized about /16 inch or rice. The abrasive must have sufficient mass to achieve the momentum required to dislodge the dirt. However, high mass particles do not follow the gas stream. Also, they are hit by the ieading edge of the moving wheels and blades. Consequently, the traiting edges are not abraded. The closer the dirt is to the point of injection, the less significant the asymmetrical distribution. The abrasive must also be sufficiently tough to resist breakage on impact. This is a problem r'vith rice, since it shatters readily. Again, the closer the injection to the deposit, the less significant the toughness. Another problem with abrasives is what happens to them after they have done the cleaning. In a simple cycle gas turbine, they will probably be burnt. However, on a regenerative unit they can deposit in the regenerator. Some regenerator burnouts have been attributed to this. In steam svstem, they would probablv plug up traps throughout the system. Duling discussions about abrasive cleaning, the possibilit;r of causing labyrinth damage is alrva;-s laised. fn fa-ct, these apprehensions have proven groundless. We do not i,now why this is so. It couid be that the particlcs are too big to enter the clearance space. On a centrii,-lgai compressor, a iypical radial clearance on the inl,erstage shaft iabyrinth is 0.008 inches, as compareci "with a particle size oi 0.060 inches. The e1'e labvrinth iLas a ;ruch iarger clearance, but a particle rvould have to rnake an unguided 1B0o turn to reach it' it is unlikely tilac it would do so. ilor''r are the abrasives rntrociuced into ihe machine? iTith air compressors, the abrasive can be thrown into rhe open suction. If the suction or point of injection is
must then be removed from the system before the solute is re-deposited. Each solvent cleaning application presents different problems.
Wqfer wqshing steqm furbines. The cleaning of rvatersoluble deposits from steam turbines is a classic example of solvent cleaning and the complex auxiliary considerations. Since the solvent is water, the intent must be to bring water into contact with the deposits. In this case, the working fluid is steam and water will only remain in equilibrium rvith it if the temperature is at saturation for the pressure at that point. Water must be added not only to provide the solvent, but to cool the fluid to saturation. We would assume sufficient water must be added to make the inlet steam l/o wet. Therefore, the rvater injection rate could be as high as 25/o of the initial steam mass flow. It is the injection of such larse quantities of liquid that creates the potential problems.
The potential problems of water washing steam turbines are:
o Misalignment due to piping stress as the tenperature is reduced. c Water slugging. .
Loss
of clearance due to diflerential contrzrction be-
tween rotor and stator.
. Vibration due to non-uniform
deposit rernoval.
. Thrust failure due to almost conplete plugging of one stage.
.
Damage
to blading if it hits water retained in
the
exhaust casing.
On most machines, the misalisnment due to pipe
stress
will not be significant. After all, we are only disposing of the superheat, whereas during run up) the machine is exposed to a temperature change at least tu-o times as Iarge. However, if piping strains are a problem on startup, one must make sure all sliding supports are free before atiempting to wash. We have had no problems other than an increase (doubling) of axial vibration due
pressurizcd, the abrasives can be introduceci using a blow pot. An eductor should be used to put the abrasive ieaving the blor'v pot into a fluidized state before introclucing it to the main sas stream. A good starting point for ilre injection rate is 0. 1 weight per cent of ,,as flor,v.
to misaiisnrnent. trVe trlr to avoid rvater slugging by tr,vo rleasures. "rVe alwavs use a venturi nozzie lor desuperheating. -\lso. we insist that the piping fall continuouslv betr'r,eeu ri're desuperheater and Lhe rnachine ialet. Er.en if the rvater is :iot broken up into cirooiets in the ciesuperheater, it rviil pass inoccuously through the turbine as a constarlt stream. To prevent loss of clearance. .,ve alu,ays limi'c the rate of temperature chanse to 180oF per hour. The greatest ':tazard wouid be failure of the injection pumps rvhen at maximum injection rate. Such a iailure ".vouki produce ;!. very high rate of change of temperature ald vrould raost likely resuit ur an axiai rub. To guard :his. rue iiry tc use boiler feeciwater, since these p'"rmns are the ncsi
Sclvellt cieaning is a :nuch :nore delicaie lecitnique tha:r ihe blgte force of abrasic:e" Iit :'e;:iiti,-, there ,vill :-iinost z;i'.';ays be solae ahrasive a":ti':.: i:rvc1'zed. Tle idea is ic rjLissoLve ;.he dep*sit j.n r- s*iteai. The soirrticl
retiable in the plant. tr'Ye a-ttempt te reiuce rhe cha-nces of r,li'l-lr:lfcrrn di:posit reinoval i;i, ha'.tirg riie ilcrease oI i'rjet:iicl rate u'h.e;r:.i.:i depcsi: is actr-rel1y bsilg r-::l:.o.,rec. T':ris cot-
fi EFH ! liT
ES
FR0fl"4
ll';'il
RCIeA
REOI{ Pi10C
h$*q;
i',1
G
WHY CLEAN TURBOMACHINES ONSTREAM?
accelerated louling rate durins
dition is detected by neasuring the conductivity of the exhaust condellsate. It is alleged that thrust bearing failures have occurred bccause oI stage pluggin-q n'hen an upstream wheel has sluffed its deposit before a dorynstream one. If this is so. the halting of injection increases tvhen material is beir.iq remor,'ed rvould prevent it. One of the criterion \re use to check a turbine design before purchase is "Can all the condensate be removed from tlre exhaust?" Some turbine designs are such that the l:laclir.rg is 'nvithin abour 1 inch of the bottom of the casir,g. Others don't have a casing drain at the lor,r.est point. Others have a /2 inch or z/a inch casing drain. All thcse designs are suspect unless the condensate can
drain frcely out of the exhaust.
\\Ie consider that a stage is washed adequately when the conclensate conductir.itv falls to half its peak level. \\'lien this point is reached, the rvater injection rate is agair.i increased. The rvash is considered completed when the inlet stearn is saturated and the exhaust conductivity is dou,n to 200 micromhos. After the rvash is completed, the inlet temperature is raised to norrnal at a maxinum rate of change of 180" per hour. Initiation cf the normal steam flow path, bypassing the desuperheater'. is the last hazard. We have found that rrater builds up in the line upstream of the valve, even rvhen the br'pass is left open. We norv al'"vays leave the main r.alve cracked open to pre\/ent the water buildup. Over tlie past 13 years, \\'e have successfully completed about 30 turbine \\'ashings. These involved six diflerent machines, Iocated irr four plants. Based on this, u,e concludc- that on-load \\,ashing is safe provided reasonable cale is exerciscd. \\'e have, howe\/er, observed that deposit solubilities vary considerablv between subscqLrent washes on the sanre machine. This same variabiliry has been observed on tr,vo machines supplied by steanr fronr the same source for the same time period. Some machines can be successfully cleaned rvithout makine the inlet saturated but most have required a r,vet inlet. Durins our earliest washes, rve believed that condensate n,as essential as the superheating medium. We reasoned that any other water u'ould leave salts behind during the remperature increasing phase. Five of the machines have now been washed using boiler leed water, ',vithout ani, sb...r^tle problems or deterioration of the cleaning. AIso during our earliest u,ashcs we noted an apparent Aboul the quthor BnrnN TTnNr,r" is an engineering associ,ate in Imperial Oi,l'a Dttginaering Diaision at Sarruia, OnL Hie functione
include prepat ation of maclvinerg speci-
nd
consu
testing,
ce
and,
has been in the Engineet ing Diui,sion
for 76 cess en mouing
a rat"i,ety of gland.. He usas e ical^engi.neering in Adwrio. i,n-
30
util:ities, prohinety. Before ttso,e emmloued,
ou)et, lectri prof e
rL_
n_
er
a wash. We nomenon.
It
ha,,-e
tlie first
fe,,v dz.ys after
no reasonable explanation of this
phe_
levels out quicklv and does not appear to
affect eithel the maximum mass flow or the efficiencr,. Currently, r,r.e merely rvarn the operators to disregard it
if it is obserr.ed. In spite oI the abor.e hazards, steam turbine rvashinq is an idcal application of solvent cleanins. If the inlei steam is unclei 1400 psig, is initially superh"eated, and the turbine is under load, the exhaust n'ill al,,r,ays be r,r,etter than the inlet. Therefore, u.herer.er the water is injected. the rvashing uill occur first on the dou'nstream blades. Further, as the u-ashing proceeds, the solution passing through parts of the turbine alreadl' cleaned is alr,var.s more dilute than r.vhere the deposit is going into solutior-r. There is no chence rhat dissolred nrrierial uill be redcposited. The water soluble deposits in these machir.res are usually highly soluble and rvashing is rapid. Al1 the deposit removed remains in solution in the condensate and can be removed from the discharge end by dr-ainir-rg.
Other solvent cleoning opplicotions. Other nrachines are not as ideally suited to solvent rvashinq as turbines. fn compressors, for example, the tendency is for the so1-
vent to e\/aporate as it passes through tlre machine. In addition, the solvency of n-rost solvents is much belo,,r that of rvater for sodium chloride. The problem is to find a solvent that rvill attack the deposit, and ."vill transport it out of the machine. In addition, the solution must either be r.vithdrau,able from the discharge s)/stem or must be acceptable in the product. The solvent ma,v be a t\\.ocomponent svstem if the deposits are at tlie cold end-a light one to clo the actual dissolving and a heavy one to pre\-ent dcposition as tlie forrner er.aporates. \\/hen the deposits are at the dischara.e, a light sol'u'ent can be used. prorided it is inj.cterl rs close as possible to the doposits. in large droplets, and a high rate. These three criteria help to prevent premature solvent evaporation. \A'c hacl an example of premature evaporation recentlv.
Dilute caustic soda \\,as inadvertently admitted into a centrifugal cornpressor. Thc rvater evaporated leaving a solid deposit betrveen the sccond and third wheels. \\'e did not \vant to use water as the soivent for tu'o reasons. The machine rvas equipped rvitli alurninum labyrinths and rre u'ere afraid ihe wet caustic rvould dcstroy this metal. Further, since the discharge temperature u,as 300o.
it seenred unlikely that enough water could be added to carn' the solution out of the discharge. We found that a liquid hydrocarbon with a final boiling point of about 4500 would move the deposit. Using this as a rvash, the deposit was moved to betrveen the fifth and sixth rvheels. Hou,ever, even r,vith injection into the stage drain before the fifth wheel, it r,vas not possible to move it out of the clischarge. Finall,v, \ve gave in, shutdor.vn,
and rvashed "off line" w.ith mildly acidic lvater. As rvith turbine washing, there must be a u,ay of remo\-ing the solution from the discharge casing. If the discharge flange points down, this will not present a problem. Ilowever, if it points up, and espccially if the discharge velocitv is lou,, an adequate sized casing drain is essential. ACKNOWLEDGMENT
Originally presented at the 2nd Anoual Texas A&M Turbomachinerv Symposium, Oct. 23-25, 1973, College Station, Tex.
A closer look at turbomachinery alignment to physically remove and inspect the shims at every machine support just prior to final alignment. Obvious problems with shims include rust, improperly cut shims, folds and wrinkles, burrs, hammer marks and dirt. I once found a turbine installed on shimpacks shimpacks is
Successful shatt alignment requires consideration of all factors that influence movement of the machinery. Here's what to look tor before and during the actual alignment
wrapped in masking tape. It is good practice to use as few shims as possible and replace many thin shims with ferver shims of greater thickness. Stainless steel shims will pay for themselves many times over by minimizing alignment problems associated with shim deterioration. It is equally important to see that surfaces of equipment supports and soleplate/baseplate are clean and in good condition.
Jock N. Essinger, Shell Chemical Co., Houston Gooo runsoMACHrNERy ALTcNMENT consists
of
three
interrelated parts. Each part is an equally important step in realizing the goal of an equipment train capable of operating safely for long periods without the many problems resulting from severe misalignment.
l. The sysfem survey. Many steps vital to good turbomachinery alignment are taken well ahead of the actual "cold alignment." IJnless the person responsible for alignment has first-hand knowledge that pre-alignment preparations have been properly carried out and completed, verification is essential. Some of the items which warrant specific attention are: 1. Piping. A visual piping inspection by those responfor alignment is vital. This inspection will assure that the piping is installed in apparent agreement with design criteria, is complete and in its functional state. Obvious items to look for are proper placement and adjustment of guides, anchors and supports; proper adjustment of tie-bolts on expansion joints; correct positioning of spring hangers; complete make-up of flanges with gaskets in place and bolts tightened; absence of slip-blinds which may have been installed for hydrotesting of pipe lines, and proper orientation of check valves. In short, it must be verified that the system is in order so post-alignment piping modifications will not nullify the alignment
sible
eflort.
2. Grouting should be checked to be sure plete and apparently well done.
it is com-
3. Foundation bolts should be checked for tightness.
4. Check all shimpacks. Shims are the vital link between the machine and the foundation and are essential to maintaining alignment over long periods. It is my experience that the only reasonable assurance of proper REPRINTED FROM HYDROCARBON PROCESSING
5. Check for misalignment of machine supports relative to the soleplate. A relatively simple test for problems in this area can be made when shimpacks are checked. Mount a dial indicator on the machine supPort with the indicator stem resting on the soleplate. Watch the indi-
cator as the hold-down bolts are loosened' If movement of the indicator is more than 0.001 to 0.002 inches, it is an indication of a problem that must be defined and eliminated. Remove the shimpack and check with feeler gages to be certain the machine suPPort is parallel with the soleplate. If not, re-grout, re-machine the support or prepare tapered shims. Be suspicious if one support moves less than others. If three of the supports move 0.001 to 0.002 inches, and
the fourth shows no movementJ
it may indicate that
the
support is carrying more than its share of the load.
6. Check the casing for distortion. Despite the massive of turbomachinery casings, they are flexible and easily distorted. This relatively simple test for gross appearance
distortion can be made when shimpacks are checked:
a. With three supports tightened d6rt'n, remove the shimpack from the fourth support. b. Determine the total thickness record the dimension.
of the shimpack
and
c. Let this corner of the machine down (the machine is now supported only at three points) . Using feeler gages,
determine the distance from the soleplate to the machine support. Record the dimension. d. Subtract the feeler gage dimension from the shimpack thickness. This is the total deflection of the machine casing with no support at the corner being checked. e. Repeat the procedure at each of the four suPPorts and compare the deflection of each. Gross differences in deflections at arly of the four supports is an indication of probable casing distortion. 31
4. Decide on tolerances for theoretical cold alignment settings. Allowable deviations which are unrealistically strict cost time and money, and may not be attainable. Tolerances which are too loose are an invitation to trouble.
TURBOMACHINERY ATIGNMENT
COUPLING HUB
IML
5. Make a list of tools and instruments required for the alignment and make certain they are available and in good working condition. These include such items as alignment brackets, dial indicators, tools required for hot alignment checks, pre-cut shims of various thicknesses, tools for lifting the machines during shim changes, Fag. l-Typical arrangement for "Face-OD', readings termine cold alignment of shafts.
to
de-
7. Check for piping strain. Piping strain is seldom detectable by visual observation. Ilowever, gross problems can be detected by a rather simple test. Following the check for casing distortion, place dial indicators on the machine to monitor both vertical and horizontal movement of the casing or shaft. Now loosen all the holddown bolts. If the machine moves more than the average observed when checking individual supports, it is obviously the result of an external force-probably the piping.'
in the mathat they are lubricated and that the bearing
B. See that bearings are properly installed chines,
covers are properly tightened. These are rudimentary pre-
cautions made
on the basis of expelience through at-
tempted alignment of machines r+'ith the bearings removed.
9. For turbomachinery trains with high-speed gearing, extraordinary precautions should be taken to set and align the gear itself. High-speed gearing is normally the most precisely made equipment in the train; it is also the most sensitive and most vulnerable to catastrophic failure as a result of poor installation. IJnfortunately, installation instructions for rnost gears are grossly inadequate. One fundamental precaution is offered: if a gear is involved, be very, very careful. Planning. Since turbomachinery alignment normally involves the participation of many people in various functions, it is imperative that an over-all plan be developed to assure that the goals of the program are known and understood and that the method of attack is consistent with the desired results. Moreover, such planning is valuable from the standpoint of minimizing the time involved (hence, the cost) to achieve satisfactory alignment. Plans should include:
1. Determine the desired placement of the shafts (cold settings) considering anticipated thermal growth of the various components. Define movement of shafts within
bearing clearances (rising-pinion gears, for example),
hydraulic loading and any other factors expected to produce relative movement of shaft centerlines when the machines are operated.
wrenches of the proper size for hold-down bolts, coupling tools and a device for rotating the shafts.
6. Provisions should be made for permanent recording of alignment data. Many good sets of alignment readings, written upon the side of a compressor, have been covered by the painter's gun.
2. Cold olignment. The term "cold alignment" refers to the position of a turbomachine's shaft centerline relative to the shaft centerline of a connected machine, with both machines in a non-operating or "cold" condition. Offset and angularity are both implied by the term. Cold alignment is important because it is normally the only check made to directly determine the relative position of
the two shafts. Results of the check form the basis for determining shaft alignment during operation. The only cold alignment techniques discussed here are those using dial indicators. Procedures ale old and well established, but do have problems and pitfalls. Fig. 1 illustrates the most widely used of the traditional alignment methods, commonly referred to as the ('p46s-
OD" method. A bracket is attached to one shaft and extends near the coupling hub on the adjacent shaft. Dial indicators are attached to the.bracket as shown, with the stem of one indicator resting on the face of the coupling hub. The stem of the other indicator is resting on ttri OD of the same hub. Oflset of the shafts is determined by the OD readings. Angularity is determined by ,,face',
readings.
The method has and continues to serve well in the vast majority of industrial alignment problems. Indeed, it is the specific method outlined in virtually every maintenance manual for industrial rotating equipment. For high-speed, high-hp turbomachinery, where requirements for precise alignment are more stringent, shortcomings of the method become increasin$ly important. Consider the following points:
o The indicator readings from this method reflect not only misalignment of the shafts of the two machines, but inaccuracies in geometry of the coupling hub upon which readings are made. Typicalty, the measurements should have an accuracy approaching 0.001 inches. It is not uncommon to find runout of the face or the OD of a
in excess of this amount, even on god quality, high-speed couplings. This error can be eiimicoupling hub
2. The sequence of alignment must be determined for multi-unit trains. For two-component trains, determine which of the two machines is to be moved. 3. Select a specific method for determining relative shaft positions. 32
o When making face
measurements
it is absolutely
that the axial floats in both shafts be accounted for. This is of little concern with small equipment where shafts are fixed axially by ball bearings. It becomes a problem, however, on machines equipped with hydrodynamic thrust bearings (or no thrust bearings at all) . The normal procedure is to separate the shafts axially each time a reading is taken. The axial position of the shafts as they come to rest depends on how hard the craftsman pushes. Thus, consistent readings are difficult to obtain. The practice also leads to the use of makeshift tools to necessary
pry the shafts apart, and a coupling seldom comes through the alignment procedure unscathed.
o The face diameter upon which readings are taken to determine angularity of the two machines is relatively small, especially in high-speed equipment-sometimes no larger than 3 inches. For example, a reading of 0.002 inches (normally termed "acceptable") on the f.ace of. a 4-inch diameter coupling hub represents an angularity of 0.006 inches per foot or an error of 0.030 inches at the opposite end of a machine 5 feet long. Add to this a 0.001 inch error in the face of the coupling and the outboard end of the machine is 0.045 inches from the desired po. sition. In most instances such an error is not catastrophic, but it is an error that should not exist. o The tool normally used to hold the dial indicators is either a "universal" bracket, or makeshift device contrived on the spur of the moment. Brackets specifically designed for the machines being aligned are not normally considered necessary and are, therefore, not used. The assumption is usually made that the bracket being used is stiff; i.e., it does not deflect under its own weight. Normally there is no attempt to verify the assumption, which is a very poor one, especially for couplings with relatively long spacers. I have found many brackets which deflect 0.005 to 0.010 inches over a l2-inch span. In an extreme case (a machine in trouble) , it was found that the indicator bracket which had been used deflected about 0.040 inches.
o
The method requires that the coupling spacer be removed to make readings. For routine alignment checks this means added ef{ort and the ever-present risk of coupling damage during disassembly and re-assembly. For new installations, it normally means that the coupling is left open until after cold alignment is complete. The interval is usually several weeks from the time equipment is first set upon the foundation
until the alignment is completed,
during which time the coupling is subjected to corrosion,
dirt and physical
abuse. Some couplings do
not survive
this initial treatment.
Alignment hints
o Assign a responsible, technically astute individual It is too important
follow details of the alignment.
to to
receive second-class attention.
o
Spend adequate time in preparation.
It
is money well
sPent.
o Make sure the craftsmen understand both the method and the intent of the alignment procedures.
o
Check the tools.
A sticky indicator
has no place
among the alignment tools.
o Assure yourself that the indicator brackets are sturdy. The best test is a healthy shake; if tJre indicator readings do not return to the original position, the bracket is inadequate. The use of magnetic indicator brackets is never permitted for shaft alignment.
o Find a good way to turn the shafts. If you are lucky, the person who specified the machine had the manufacturer put wrench flats on the shaft extension at the outboard end of the machine. If not, find another means, such as a strap wrench or a clamp-on fixture made specifically for the purpose. o
it
When placing the indicator, be sure
it
starts out at
rests squarely on the shaft. An indicator which sets askew gives bad readings. Likewise,
mid-range, and be sure
be sure that the range of the indicator is not exceeded when readings are taken.
o Before making any moves (and before final acceptance), get at least two sets of identical readings. If the readings cannot be repeated, they are not acceptable.
o Check the data before making a shim change or horizontal move of the machine. It takes very Iittle time to check the data-it takes a long time to move a large machine.
o \4ake one move at a time, then check
the work. Don't
worry about horizontal movements until the elevation is proPer.
o Be alert to peculiarities of the specific machines being aligned. Tilting-pad bearings, for example, sometimes give trouble by permitting the shaft to rock back and forth on the lower pad, making repeatable measurements difficult. It is sometimes necessary to put temporary shims in such bearings to assu,re that the shaft is stabilized for alignment purposes.
o Stop the shaft at precise 90-degree increments for measurements. Turn the shafts in one direction only. If you pass the 9O-degree mark, go around another complete turn. Extremely handy for tJre purpose of determining the quarter turn is a tool with four spirit
Reverse indicqtor method. fn contrast 1o d1s "p2ssOD" method, indicator readings can be taken by the reverse indicator method on the OD of the coupling hubs (or the shaft) only. Two brackets are used simultaneously, which is normally preferred. Equally acceptable but less convenient is a method that uses a single bracket switched back and fourth for each set of readings.t Analysis of information gathered by this method yields suflicient data for determination of relative shaft positions. Advantages of this method over .,Face_OD,, are:
most cases (shafts must be sufliciently in alignment to prevent binding of the coupling, of course) . This feature provides three distinct advantages:
1. By proper design of the alignment brackets, the required removal of the coupling spacer is eliminated in
a. Wear and tear on the coupling is reduced. b. Since both shafts turn as a unit with the spacer in-
REPRINTED FROM HYDROCARBON PROCESSING
levels mounted
at
9O-degree increments.
o Be wary of planned changes that do not turn out right. Alignment changes are calculable and predictable; if they consistently go wrong, look for something amiss in the system.
33
TURBOMACHINERY ATIGNMENT ALIGNUENT
stalled, errors caused by coupling hub runout are entirely eliminated. c. By spanning the entire coupling, angular misalign-
ment is greatly magnified and more precisely diagnosed. For example, a span of 16 inches gives angular misalignment readings eight times as great as if measured on the face of a 4-inch-diameter hub. 2. Since face readings have been eliminated, there is no concern about axial float of shafts.
3. This method is not successful if makeshift tools are used-special tools must be constructed. Alignment results are simply better when proper tools are used. I wonder why alignment tools for high-speed couplings are not available in standard designs from the coupling manufacturer
?
('Face-OD" method that The single problem with the is not solved by the "reverse indicator" method is deflection in the alignment bracket. This is a matter of real concern, and one which must not be ignored if proper cold alignment is to be achieved. The problem is readily handled, however, by determining the deflection in the alignment fixture and making appropriate corrections in alignment data. Fig. 2 indicates one method for determining the error. As shown, barstock of appropriate size is chucked in a lathe, and the end is turned to the proper diameter to accept the alignment bracket. Without removing the barstock from the lathe, the alignment bracket is attached. With the bracket on toP of the barstock, the indicator is set to zero. The entire assembly is then rotated 180 degrees, so the indicator now reads directly on the bottom of the barstock. The indicator reading thus obtained is reflective of the deflection in the alignment bracket. Once determined, deflection should be stamped prominently upon the bracket, as should the designation of the equipment for which the bracket was specifically made. The alignment bracket becomes a special tool for that machine, and should be cared for as such.
3.
The hot olignmenl check.
It
is now generally ac-
cepted that the traditional "hot check" for turbomachinery alignment is'of little value. Not only is it costly and time consuming to bring a machine up to temperature, stop it, break couplings and attempt to determine alignment before it cools off, but results are highly questionable. In some instances a hot check is dangerous because it creates an unwarranted sense of security. It is not possible to make the check quickly enough to accurately determine thermal growth of the equipment.
For machines operatine at temperatures well away from
About the quthor ESSINGER is a staff engineer with Sh,ell Chemical Co., Houston. His
JAcK N.
duties inclucle specifi,cation, eaaluati,on, insta,llation, startup, troubleshooting ond maintenance of rotating mo,clnnery in chemical plants. Mr. Essinger holds a
B. S. degree in mecho,nical eng'ineering from Arizona State Uni,uersitg. He is a member
of the
Amet"ican Societg of Mechanical Engineers anil the American Soci,etg
34
Fig. 2-Setup for determining deflection in alignment brackets.
for Metals.
Fig.
3-Basic elements of mechanical device for checking
hot
alignment.
ambient, it is not uncommon to see a thermal change of 0.001 inch per minute when the machine is first shut down. At this rate, the effectiveness of the hot check is certainly lost before the alignment data can be obtained. Misalignment resulting from hydraulic forces and torque reactions, which can be significant, are never revealed by the traditional methods because the forces disappear when the machine is stopped. It is generally accepted today that a superior alternative is to use the cold position of the shafts as a benchmark and deduce the hot alignment by monitoring the movement of the machine casings, or shafts, from the cold position to the hot position. Several methods have been used for this monitoring, with a variety of techniques applied to the actual determination of casing or shaft movements. The most widely known methods use optical or electronic techniques.
Details of these techniques have been well publicized.2 Flowever, a purely mechanical method for measuring hot alignment is also being used successfully. This method uses permanently mounted tooling balls for measurement location references" A spring-loaded mechanical gaging device measures relative movement between the reference tooling balls in increments of 0.001 inches. Fig. 3 shows the basic elements of this method as applied to a typical turbomachine. Again, the cold position is used as a benchmark for alignment corrections. ACKNOWLEDGMENT Abstracted from the paper "Alignment of Turbomachinery," oiginally pre sented to the Scond Simposium on Compressor Train Reliability, Manufacturing Chemists Association, April 4, 1972.
LITERATURE CITED 1fackrcn. C..
"Hw to align barrel-t?e centrifugal compresom," .flydroiarbon Proeessins, 50 No. 9 p. 189 September 1971. 2 Jacksoo, C., ':Succesful ihatt hot aligment," Hydrocarbot Processitg. 48 No. t p. 100 January 1969. t
Hot alignment too complicated? Existing measurement technigues tor hot alignment of rotating machinery are complex and dilticult to use, Here is a simplilied way to obtain the necessary data Jock N. Essinger, Shell
Nolr- refer to the cold ancl l'rot measln'ements previousll' n'rade (A, A', R, ancl B'), arrd dctcrrrrine the rrrovernent of the bearing housing along vec.tors A ancl 13 by taking
the differences between c:old ancl hot rneasulemerrts (1A and -tB) for each location. Lrry out these nrovemcnts along vectors A and B using any convenient scale. sa1'
tf
Chen-rica1 Co.,
inch equals 0.001 inch, to cstablish points n and b. Nou'
Ilouston
Ixroart.lrroN FoR TrrE hot alignment check is nolmallr' obtained b1' measuring vertical arrd horizontal movements of shafts or bearing ho-.rsings of turbomachinerl components relative to fixed references. When optical alignment techniques are emp1o1'ed, the references are vertical and horizortal planes of sight established by the optical instrument,i. Por electronic or mechanical techniques, the fixed reference is norn-rally the machine foundation. For the typical installation, horvever, the foundation is not in the proxir.nitv of the bearing housing, and it is necessarl' to install pedestals extendins lrom the foundation to a point near the bcarins housing, these pedestals bccomins the refet'euce fronr uhich measurelrents are taken. To asstrre the ecculacv of tl're alignnrent data, the pedestals must be substantial, and thev mr-rst be maintaincd at constant terrlper:rtlu'e to minimize ineasurement inaccuracies resulting fron thelnral grorrth of the pedestals themseh'es. The perlestals arc nornally rvater cooled. The alternate method presented here also uses the nrachine forrndation as a reference, but climinates the requirement ior alignment pedestals b,v using permanent refercncc points rvhich are affixecl directly to the foundation ancl to the bearing housings (or the machine case) , as shoun in Fig. 1. All four refercnce points lie in a planc irerpendicLrlal to the ccnterline of the rnachinc shaft.
Fig. l-Typical placement of benchmarks on foundation bearing housing.
Similar relcrence points are mor-rnted at each bear:ing
hoLrsins in the tr:rin.
Follorving cold alignment of the compressor train, refefence dimensions A and B, and angles O and { are detern'rinecl et cach bealing housing, and are recorcled. \\'hen the rllrchine is brotrght on line, din-rensions A'and B'are
at each position in the hot runnins condition. Thc clata thus obtained is adequate for determining the measulecl
vertjcal and horizontal movement of each bearing housing in the nrachine train relative to the foundation. See Fig. 2. Using c.ommon grid paper (4 x 4 grid is usuallv a convenient size), lay out reference vectors A and B at angles O ancl 95, having these vectors closs at one of thc grid intersections. The intersection of these vectors represents the centerline of the machine shaft in the cold position.
Fig. 2-Graphical determination of shaft in hot position ielative
to cold position.
HOT ALIGNMENT TOO COMPLICATED?
q
FiE.3-Stainless
steel ba
wirh protective cover.
Fig.5-Calibration of the alignment gage with the lnvar
Fig.4-Precision tool set for hot alignment checks.
stand ard.
drarr ljrrcs Llrrorrgh a ancl b Pcr'Pcndicular to \-ectors A arrrl l:i. -f.lrcsc lines rcprcsent arcs of racliuscs A'and B'
elin-rinating i.he concern Iol tirermal srorvth oi the alignment gaLlge. Thc inclinornetel is used lor cietermining angles at rrhich the measLrrLrilents arc made.
dr:rwrr ironr tlrc foundal.ior-r Jrr:ncJrmalks. The intersectiort tlrr:sr: linc-q de[incs t]re Position oi tire machine shaft
oI
irL thr' hot position relativc to the cold positiorr. dr:lr:rrrrinc tlrr: nrovurnent in vertical and holizontal dirc'ctiors, it is necr:ssary c-,nli, to scalc off tile dirur:ttsious reicllecl to as AII:Lnd AV, using the same scaie zrs used in plotlirru A.A. and Ati. A sirnilar piot for the ciata secured at crr:h bcarirr.g ]roLrsing afforcls srrfficiert infonnation for plottin.g tlri' hot a'lisnmcnt of the entire trrrbomachinerv
.\
t:orto'lirt'
f'o
tr:rir
r.
Tlrc Lcrrchnrarks for this techniquc are /2 inch diam-
c[cr' prccision ba]ls. llecr.rse t]resc benchmarks becon-re an intceral Part ol the installation, arrd because the accuracv oI alignnrent rci:ords, over thc long tern, are depcndent uyrol thesc Lel{:renr;cs, it is reconrrrrended t}rat they be rrraclc of stainlcss sti:cl to prc\cr)t corrosion end that thel' br: nrounted strbstantially to avoicl inadvertant movement. Fig. 3 shorvs a bcnchmark and protcctive cover specificall-r' clesigncd for ttrrbomach ineri. rvork. Fig. .l shows :r set of precision tools specifically n-rade for the l)rrrposc of turbornachincrl' alignment by this tech-
rricluc. 'I'he alisnmcnt gage cmploys a long range dial indicatol set into :r si;ring-loaded telescoping column u'ith |/r inch dianreter spherical seats at either end. B), t sing six extensions, thc tooi covers a measurement range of incrcments of 0.001 inch. 'Ihis range is adequate for nearly all turbomachinery in use today. The star.rdard pror,idcd for calihration of tlie basic tool is con25-60 inches,
structed 36
in
of Invar', as are all of the
extensions, thereby
steprrise procedut'e
for alisntuent bv this :letl-rod
is
as loliorr s:
1.
Aflir
bcnchmarks to thc rnachinc anC lo ihe [ounda-
tion.
2. Aiign the tur:bomachinelv train in thc nolrnal mAnneflecording coupling alignment dat:r.r,r
3. Using thc Inr.ar standard, calil:rate thr alisnrrent gauge 3s rorr'n in Fig. 5. 'l 4" Usine the appropriate gauqe extension, clcttt'tnine ani record colcl measlrremcnts at cach locatior-r. a: shorrn in Fig. 6. 5. Deternrine the anq-les at rvhich measLrrcmelrr; arc being made bt' usine the inclinorneter as shou't'L in Fig. 7. Recorci the data.
6. \\Ihen the train is brought on line. obtain end record hot aiignment data by repeating Steps 3 and I at cach reference station.
7. Plot the data as indicated in Fig. 2 to detenline vertical and horizontal movement of each bearing housing. B. Using combined data from the individua[ locationsplot the hot alignment of the entire equiprnellt train.
9. Make appropriate alignment changes as dictated by the above data.
Fig. 7-lnclinometer for determining angle at which rneasurements are taken.
Fig. 6-Reference measurement between bearing housing and
;-
machine foundation.
Fig. B shous t1.pical placcment of ber.ichmarks at the various bearing housing locations in a turbine-contplcssor tra1n.
This alignment technique has been used successfi-rlly on severai turbomachinery trains over the ltast three vears. Some of the apparent advantages of the methoci are:
o Thc
tcchnique is simple, and no speciiic technical skills
are required.
. \rcrtical and horizontal movernents of the rnachinc are obtaincd r,vith equal ease. o Once the bencl'rmarks arc established, setup tirnc for hot checks, or le-checks. is minirnal. o The benchmarks are permancnt, and are readiiv maintained for long-term monitoring of aiignment. e The method is highly resistant to accidental loss of reference; i.e., to being "bun.ipcd" during use. o It
is readilv adaptable to crarnDed quarters.
Fig. 8-Typical benchmark placement on a turbocompressor
train,
About lhe quthor
JAcK EssINCER is a stalj engineer taith Shell Clrcmical Co., Houston. His duties irtclude speci,fi,cct tion, euoluation, installation, startuTt, troubieshooting and maintenance of rotating m,achiner1l
in clrcmi-
cal plants. Mr. Essinger holds a
B.S.
degree in mechctnicctl engineet"ing from Aq'izona Skttc Utzirersity. He is a member of the American Society of Mechanical Engineers and the American Society f or Metals.
REPRINTED FROM HYDROCARBON PROCESSING
o Datataking is a one-n'ran operation. o Initial investment in harclr,r,arc is modest. Rc-use is high. o The equipment is highll'portable. o The hardrr.are is simple, reliable, and rerquires rnirimal caliblation. LITERATURE CITEI)
rJaeLsorr C..."llo_w r-o- ali^en barrel-t1pp cenrrifuqal ,unrlircs\r,,\. ftr tlrurrrLon Procc.sinS. 50. No. a. p. lBr), Septcmber Ir;1. :- Es\iree:. J..^tt.{. Ioo-k ar turhnmaclrinery rlignment." H,drmrrbnn Procc..irrq. 5:, N,'."1o.". 9, p. 1BJ, Seprcmber 1973.'
37
Turbines using too much steam? Steam tlow in single-stage turbines is relatively easy to calculate. Multistage turbines Present tlow calculation uncertainties to which must be added losses in couPlings and compressor llow uncertainties H. Steen-Johnsen, Consultant, East Hampton, N.Y. nBrrNBn was concerned that therr single stage turbines were using too much steam. And so they were. Some
A
of their turbines were rated at 100 hp. The manufacturer wanted to make sure the Power was there, so he nozzled
t#,o
[c 56
3t
-c& o
loo .l€ it6 44 42
rfr control valve. So, of course, they used too much steam at 100 hp.
Couta the operators have determined how safe they
turbine is still on governor control. When the ring pressure is 75 percent of throttle pressure, close the hand valve.
Nozzle flow meter. The first-stage nozzle is an excellent flow meter.. For a single valve turbine with critical drop over the nozzles, calcuiate the steanr florv rvitlr the fornrrrla Q:pK lb.lhr.lin.2, where p is the ring Pressure in psia and K is from Fig. 1. Example. Assume a 600 psia steam throttle pressure at
7500 F. The ring pressure is 550 psia. The enthalpy at the
throttle is 1381.-From Fig.
Q: 38
SSO
(43.1)
=
23,900
l, K = 43.4 and the flow is: lb.lhr.linj' Then, if the tur-
rr00
IIILET E{T}IALPY
'r-ar,,,"al ,,n. flow per sq. in. of nozzle area for steam' (Flow = Abs. press, x Factor K, lb.,/hr. in.') bine has a nozzle area of 0.23 inJz, the turbine is passing, Q= 23,900 (0.23) = 5,500Ib'/hr. which is within 2 or 3 percent of being corect. When the first-stage pressure is greater than critical, as it frequently is for a multistage turbine, the critical flow must be corrected, using a correction curve obtained from the manufacturer. Multistage turbines. For larger multistage turbines with multiple inlet valves, the problem is more complicated. Total flow is the sum of the flow through the several valves that are open, and requires a multiplicity of gages. A turbine of this type should be equipped with a flow nozzle in the throttle line to obtain a reasonable measurement. And here we turn to the PTC 6 Report,l Fig. 4.2' 16-18. pages -
tt
"
first item noticed is that "liquid" measurement is
bine flow as recorlmended in PTC 6, page 21, is from PTC 19.5;4, page 57 Equation 6:
TURBINES USING TOO MUCH STEAA,T?
un:359CFdzF,Y,Vh.iA zo
Fig. 2-Location of the metering nozzle here produces a flow measurement uncertainty of + 18 percent.
is rate of flow,
lb./hr.
C is coefficient of discharge F is the velocity of approach factor : I IVF-- P' When the flow section is calibrated, the product of. C F is established most accurately. d is the nozzle throat diameter. Fo is the thermal expansion factor for ttre nozzle. Yo is the adiabatic expansion factor to allow for the change in ur. It is a function of the pressure ratio over the nozzle prf pr, which for 600-pound steam pressure might be (607) l$l+) = 0.987. The curve Fig.43 A, gives Fo as 0.984. Therefore, an error here is a minor uncertainty, and, it k: 1.3 or 1.4 it makes little diflerence. lr. is the differential pressure, inches water at 680 F
(not Hg).
5lfltrln*nrt
Fig. 3-Location of the metering nozzle here produces a flow measurement uncertainty
of +
1.75 percent.
Fig. 4-y;n;rum tlow nozzle for compressors.
arrangement
rated better than superheated steam measurement. This "liquid" procedure, of course, requires a unit boilerturbine-condenser system, which is rare in a process plant. The liquid measurement allows accurate determination of the condensate and the feedwater. The next item noted is the requirement for inspection. A flow nozzle in a steam line is subject to considerable abuse when a plant is started. Welding beads, welding rod and other debris go through on initial blowdown. The resultant turbulence increases A,p, and a steam nozzle reading 5 percent higher than the feedwater is not unusual.
Location of the metering nozzle is important, as noted by the diflerence between Figs. 2 and 3. Flow measurement uncertainty. If the arrangement in Fig. 2 is used without inspection before perrnanent installation, the PTC 6 Report rates an uncertainty as t 18 percent. If the arrangement in Fig. 3 is used with satisfactory results of an inspection immediately before test, the report rates the uncertainty as -+-1.75 percent. For the primary element, PTC 6 recommends tfre use of the throat tap nozzle with a coeflicient of discharge, C, given in Fig. 4.4, page 23 of PTC 6.1 As background for comparing uncertainty of turbine and compressor flow measurement, the expression for turREPRINTED FROM HYDROOARBON PROCESSING
ur is the specific volume, cu. ft./lb., at the inlet of the primary element. The diflerential pressure, /r-, should be read on a manometer for best results. The use of a recorder for direct flow reading results in an uncertainty. The flow formula is programed in the recorder, and the linkage must be adjusted accordingly. The proper installation of a manometer is discussed in the code.1
Power to coupling. With the flow established with an acceptable degree of uncertainty, the next question is how much power is delivered at the coupling. In other words, how efficiently is that flow used? In a back-pressure turbine with more than 25o F superheat in the exhaust, the heat drop in the turbine can be established, and thus the internal power may be computed. Using ,a reasonable allowance for mechanical losses, the power at the coupling may then be calculated. For a condensing turbine, the expansion line may be available, and there may be some interstage taps. Pressure readings at these points are of value only to establish gradual fouling. The interstage readings should only be used as a source for internal efficiency determination, when steam is extracted for feed-water heating, and reliable readings without turbulence can be obtained from the piping to the heater. And look out for gland leakoff returns that may upset the reading.
After the flow is established, there are two means for determining the coupling power. The manufacturer's efficiency data in conjunction with first stage pressure, or a dynamometer with reasonable uncertainty. For a mechanical drive, the dynamometer may be either an absorption type or a transmission type. For a field test with the driven machine in place, the coupling power is best determined by a transmission dynamometer. This device measures the torsional twist in an extended type coupling, and there are several good ones on the market. Properly installed and calibrated, it will quickly resolve whether or not the prob-
lem is in the turbine or the driven comPressor or PumP.
lnstrumenlotion qnd gos composiiion. For those situations where the turbine is 98 percent efficient, and the compressor 60 percent, the instrumentation should be checked first, and of equal importance, the gas composition must be carefully evaluated and matched to the compressor design.
39
Gompressor flow meqsurement. While the turbine flow is being established by a Fig. 3 flow measuring section, take a look at the means for compressor flow measurement. A flow nozzle in the suction to a process compressor should, as a minimum, conform to Fig' 4, which
About the quthor H. Strox-Jonwsow is a consulting engineer, East Ham,pton, N.Y, He recently q'etired as cluief staff engineer wi,th Elliott Co. After graduating from Colunbia Unwe,r'sity with a B.S. ond, M.S.
corresponds to PTC 10 requirements.l No specific guidance is available from PTC
l0 for the deviations or auangement this with uncertainty asiciated 3, Fig' than shorter is section meas.rring it. This from -l if the PTC 6 Report for uncertainty is applied to this installation, we arrive at-+4-?5 Percent, and without flow straighteneis +8 percent (disregarding deviations in gas
degree in mecha,nical engineering, Mr. en-J ohnsen joi,ned Westinghouse Corp,, where he did prod,uct d,eoelopment in turbi,ne control and design. Wi,th Dlliott, he also was ch:ief turbine eng!neeq' and, chief product engineer superS te
uising turbine and com,pressor
deti,gn.
conditions).
For a further look at the uncertainties of compressor testing, take a look at the expressions fo-r flow and shaft ho.r"il*", in PTC 10. The expression for flow is:- W = p,) lb./min. 31.501 Kd' Yo Vl, (p' expression for turbines, lb./hr. the This is the same as (P, substituting by seen be as can - P) =h*127'7^ and : multiplying by 60, where 27'T inches of water at 680 F I psi. K = C F is the flow coefficient, and represents an item of potential controversy, particularly for an orifice' C is the coefficient of discharge, and F = llt/ 1-p' T t!" velocity of approach faci-or. For a nozzle with high Rp and a iow B iatio, K is close to 0.99 with a small uncertainty that can become large, if the straight pipe and flow straiihtener are not specified. For an orifice with Ra = lff, K isiqual to 0'6205 for F = 0'5' an{ gsual to 0'699 for 12.7 percent, and the B -- O.i. That is an increase in K of 'pfC of +'3'5 Percent to uncertainty an O Report assigns K when it is increased from 0.5 to 0.7. compressors are frequently equipped with Venturi tubes in th-e suction. The p essure loss is small, and K is well established. Just remember that the humidity must be included in the computation of p. F, is the thermal expansion factor of the metering element. It is of minor eflect. Y is the expansion factor that corrects for the volume increase, I" fbr a nozz'le, Yo fot an orifice, as the compressible fluid expands from pt -to pe in the measuring ^element. The expansion factor for a nozzle is given in
Air
PTC 10 as:
," :
l(f),u
where: r :
"
"
(L-
f-)!"
")l'''
l;.i,
^
^7'''
fu and nr is the isentropic exponent' The expansion factor for an orifice is given in PTC 10 pzf
as:
Yo= |
-
(0.41
+ 0.35
9) lQ,- P)/@, n")l
and originates in the Flow Measurement Report by the Fluid Meters Committee. For p ratios in the 0.5-0'6 range, and (fu- Pd in the range of 0.03 pr or less, this factor is in'ihe trigtr O.Ss, and the uncertainty may be on the order of one percent. p' is the density in the lb./ft.s at pr For some gasses it is a reliable number, for some Process streams it is subject to conjecture. The conjecture stems from the deviation from a perfect gas as expressed bY
p40
04+ P) IQRT)
:
(t4+ PM) l(ts+s zT)
In this expression the compressibility f.actor Z is not always known within -f several percentage points for certain gasses. And if p is not what the compressor is designed for, the mass flow (zo) delivered will be an unpleasant surprise. In a centrifugal compressor, a heavy gas picks up more pressure in a wheel than a light one, and with more pressure goes less volume, and less volume means a different area schedule for the wheel from hub
to tip. Thus uncertainty in p does not only 'affect w directly, but it also upsets the flow geometry of the wheel. The expression for shaft horsepower is:
P"n:lu (haNote:
zu, the mass
ha)
* Q,+ Q** Q'i142.408
flow, is subject to some uncertainty,
as discussed above.
haand hr the enthalpy out and in, depends on accurate pressure and temperature readings and. a good Mollier diagram or similar gas information. And of course, where (ha- he) is small, it is doubly important that the instrumentation is the best and that the readings are precise for minimum uncertainty of the difference between them. Q, is the casing heat loss by radiation, and depending on air circulation, this quantity rr:oy vary quite a bit. It must be evaluated as a percentage of the whole.
Q* is the heat equivalent of the mechanical losses
which should be quite well established. Q,"l is the heat equivalent of the seal losses. This quantity may be quite debatable for a process gas, and the seal conditions may have deteriorated. As a percentage of the total, it may be a sigrrificant portion and affect the total uncertainty of the power required. With flow and shaft horsepower computed, the performance of the compressor is established, except for an
evaluation
of the uncertainty. The
head vs. flow per-
formance is established by the data for (ha
-
hr)
.
From this discussion of compressor test, the
several
origins of uncertainty can be listed, and the parties to the test can arrive at some -l- percentage numbers, that rnay be combined and applied to the firial computation. Considerable help with this evaluation can be obtained from the PTC 6 Guidance Report, which also outlines a procedure for combining the individual percentages into a
total uncertaintY.
,
,ITERATURE GIITED
.
Better pump grouting Deficient grouting is an otten unrecognized contributor to high vibration and poor alignment retention. The result can be premature tailure of bearings and seals. Follow these guidelines to good grouting Mqlcolm G. Murray, Jr., Murray & Garig Tool Works, Brownsville, Texas Evr,n rexB a tapping tour?
Try it-you may be sur-
prised at what you find. Take a sma1l hammer and walk arrrong the pumps in your plant, tapping the tops of their baseplates at random. You will probably hear more hollorv "thuds" than crisp "clinks." The thuds indicate voids or cavities, rather than the even support supposedly provided by the grout.
Fig.
1-An example of failed Portland cement-sand-water grout.
Groufing. Let's back up a bit. What is grouting, anyway? Grouting, as concerned with here, is a procedure r'vhere a semi-fluid mixture called grout fills the space between machine foundation and baseplate, including cavities and irregularities. After curing, the handened grout provides a strong support, evenly distributing the weight and forces from the machine through the baseplate and into the foundation.
Why is grouting important? The trend for the last
15
years has been to use fabricated steel baseplates and centrifugal pumps with higher rotational speeds. These baseplates are less rigid than the older cast iron designs, and more likely to vibrate excessively at the higher speeds, especially if grouting is deficient. Excessive vibration can cause early bearing and seal failures. Poor grouting can also result in poor alignment retention, with further premature wear. Good grouting generally costs more than poor grouting, but in improved reliability and reduced maintenance cost, oflers an attractive return on investment.
Pump $rouling-usuql. Much has been written
on
grouting of c'ompressors, turbogenerators, and other heavy machinery. By contrast, very little has been published on
pump grouting. Pump grouting may lack the appanent importance or difficulty anticipated with the larger machines, but this assumption can be deceptive. Pump baseplates can be trickier to grout than compressors or their soleplates. Also, a pump often requires several times the REPRINTED FROM HYDROCARBON PROCESSING
Fig. 2-Estimating graph-pump horsepower and rpm vs. cubic to $60 per cubic foot for epoxy grout and $25 per cubic foot for a good proprietary inorganic grout. For grouting labor, flgure 1.5 manhours per cubic foot, with a minimum of 6 manhours.
fe'et of grout required. For material cost, figure $40
41
BETTER PUMP GROUTING
16 MINUTE ANCHORING
(21
FOFMS REMOVED
AFTER GROUT HAS
LEVELING SCRETIIS REiAOVED AFTEB
GnouT HAS HAnDENEq
HARDENED.
SEAI-ANT.
\
Fig. 3-Grouting sequence illustrations. This procedure provides an easy means of leveling, eliminates shims and wedges and allows pressurizing the main cavity without overflowing
Fig. 4-Grouting equipment tool cart used at a major overseas oil /efinery.
the form,
I
Fig. S-Pump baseplate leveled with screws resting on small bearing plates,
Fig. 6-The rapid-cure mixtureo is quickly troweled in place to close the gap around the baseplate lower perimeter. This mixture will be hard to the touch in 15 minutes.
Fig. 7-Using. a pneumatic.diaphragm pump to place the grout (inorganic only-not epoxy).
Fig. 8-The baseplate cavity 1s nearly filled. Weighted covers are used to prevent overflow while grout is force-pumped into remaining voids.
42
grout volume of a colnPressor hlvinq equal horsepower. Finally, the typical Process plant has about 10 times as many pumps as all other rtrachines Ptrt togeiher, so they are statistically important eveIl if less noticeable individually.
in grouting a comwith engineers, erectors, and top field forces making every effort to use the best materials and procedures to get a good job. A typical'Pump grout job,'on the other hand, is usually left to the discretion of the plant or contractor labor crew. Generally a rather soupy Portland cement-sand-water mix is used, perhaps with some iron powder to offset shrinkage. The result is often poorFig. 1 shows such a grout job altet a few years. Great care normally is exercised
pressor,
Pump grouting-improved. Good pump grouting is not difficult, trut it does require careful preparations and planning, and attention to numerous details, as follows:
(
o
Baseplate and foundation. Certain baseplate design features, such as leveling screws) grout filling and vent holes and corrosion protection, are recommended.s These
provisions cost
l-
little and make good grouting much
easier. The foundation is also important. It should extend at least 3 inches outside the leveling screws. Its surface should be chipped back /z to 1 inch to remove the
laitance, or weak top layer, to provide a rough surface
Fig. 9-Using a simple topping pump to fill the small space beneath the baseplate top surface. This type of pump can handle any grout, including epoxies and coarse aggregates.
for bonding.
o Leveling. With vertical scrervs, the baseplate can be Ieveled easily without using wedges or shims (Fig. 5). Such scrervs also permit "single-pour" grouting, thus saving time and labor. After the grout cures, the screws can be removed to let the baseplate be supported evenly on the grout. . Materials selection. No single grout material is best for all jobs. The traditional Portland cement-sand-water mix should not be used, hor'vever. When soupy enough to flow easily, it gives a rveak grout subject to shrinkage and cracking. Although much stronger if dry-packed,5 it is difficult to place in the typical pump baseplate cavity' Iron-additive grouts also are not recommended. Cases on record show such grouts have continued expanding for many yeam after installation, causing repeated machine misalignment.
The choice of grouts should be made from among the "good" proprietary mixtures. Such grouts fall into two categories. These are the "organics," or plastics, generally epoxies or polyesters, and the non-metallic "inorganics"basically Portland cement with controlled silica aggregate and non-metallic anti-shrink and cure-acceleration addi-
Fig. 10-For the inorganic grouts a paddle mixer speeds the iob. ln this design, the mixing carriage detaches and is wheeled to the grouting site. The mixture can be pumped directly from carriage to baseplate.
tives.
The plastic grouts are the most exPensivg-two to three times the cost of the inorganics. On a single large pump, this cost diflerence can easily run $1,000. AIso, the plastics are subject to creep above 200o F and should not be used where baseplate temperatures are expected to exceed 175o F over major surface areas. Ilowever, they are easy to use, cure adequately in a'bout 36 hours and have high compressive strength. Plastics are also resistant to many chemicals. As a general rule, use plastic grout on small pumps, up to about 6 cubic feet maximum grout volume. Beyond this, the ease of grouting is more than offset by the higher material cost. The proprietary inorganic grouts require mixing with REPRINTED FROM HYDROCARBON PROCESSING
,."a to retain is compte,";ilt. Ilrr-o.ortins "r" keeps them moisture for a three-day cure. A watering manifold wet automatically with minimum water waste.
43
BETTER PUMP GROUTING
o Grouting equipment. Grouting can be done with
simple and inexpensive equipment. If more than a few pumps are to be grouted with inorganics, it pays to obtain additional items such as small vibrator, a diaphragm pump and a paddle-type mixer (Fig. 10). Fig. 4 shows a portable grouting equipment tool cart and its contents. Such a kit is handy for keeping these tools together and having them on hand for each grout job.
o Grouting procedures are most effective if thought out in advance, step by step, and written down. Copies then can be distributed to all concerned, and a short meeting held several days prior to the job. At this meeting, procedural details are cleared up, and resp,onsibilities ,are agreed upon for getting necessary personnel, equipment and materials to the job site. With modern, rapidcuring grouts, the middle of the job is not the time to discover that something essential is missing! Literature citation 2 includes detailed procedures for several grout systems.
Fig. 12-Sealing the leveling screw hol'es with epoxy.
In broad terms, the gncuting procedure should include the following information: A. Preparatory steps such as foundation cleanup and oil removal.l B. Forming instructions. Generally the form top should
be z/ainch above the baseplate bottom edge.
C. Equipment and materials required.
D. Grout mixing
and placement instructions, including
pump, mixer and vibrator usage where appropriate. Sometimes timing is critical and must be emphasized. One inorganic grout requires a minimum of 2/z minutes mixing,
but becomes too stiff to pump if mixed longer than Z/z minutes! See Figs. 3,6,7, B, 9 and 10.
E. Curing instructions.
See
Fig.
11.
F. Leveling screw removal and hole sealing instructions. See Fig.
12.
Fig. l3-Pumping epoxy through grease fittinos to fill voids baseptate to satvage a poor-grout job and 99lp_an_a_pymp avoro the need for complete /egrouting. avoided.
water (sometimes icewater), and are somewhat trickier to use than the plastic gr-outs. 'fhey will, hou,e,",er, rvithstand temperatures up to 5000 F (1.0000 F for one mix-
Solvoging q poor grout iob. Sometimes a poor grout job can be salvaged, saving the trouble of completi regncuting. This is done by injection of epoxy or polyester liquid, without aggregate, to fill the cracks, voidi and cavities in the old grout. Fig. 13 shows this procedure. LITERATURE CITED
44
Centrifugal Pumps
a ,D
!!3 -...,. _
F
"iI+ -
)
Howto improve pump performance lnducers have been used tor some time to incre;ase centritugal pump IVPSH, A new design teatures an inducer that is integral with the impeller. Results are better suction pertormance and efticiency. Here are the details of tfris new development Sh. Yedidich, Worthington Standard Pump Corp., East Orange, N.J.
Trrn usnrur, wonr performed by a centrifugal pump of its speed. For example, when the operating speed of a pump is doubled, its flow-rate is doubled and the head increases four times (i.e. the power output increases eight times). It would, therefore, be most economical to operate a varies as the cube
centrifugal pump at the highest speed possible. However, the use of centrifugal pumps at high operating speeds has been restricted, for many years, by Net Positive Suction Head (NPSH) limitations. Only in recent years has there been a dramatic breakthrough with the commercial application of inducers.l Fig. 1 shows a centrifugal pump equipped with such an inducer. Several pump manufacturers offer such inducers as an option, and they have become widely accepted by users. Fig 2 shows how such an inducer is capable of improving the suction capability of a pump and, at the same time, of preserving its head-capacity relations and efficiencies. As flow rates increase and pumps become larger, the
Fig. l-Centrifugal pump with inducer.
-
IUPELLEF ONLY IMPETLEB PLUS INDUCEB
normal two piece impeller-inducer combination experiences some problems in both hydraulic and mechanical design. These are ove.rcome by the natural joining of the two pieces into a one piece "inducerpeller". Figs. 3, 4, 5 and 6 show different variations of inducerpellers. As can be seen from these illustrations, the main features of such an inducerpeller are as follows:
o The blades of the inducerpeller extend far into the eye of the impeller where they form a geometrical shape which resembles, 46
to a certain degree, either
a
Flg. 2-Performance of a stendard 6 X 4 X without an inducer.
I
pump wlth and
HOW TO IMPROVE PUMP PERFORMANCE
Fig. 3-Compressor wheel with integral inducer blades.
Fig.
4-prrp
impeller with axial blade extensions.
Fig. S-Experimental impellers with integral inducer blades. screw-conveyor
o
or an axial flow impeller.
Downstream of the axial inlet, the orientation of the blades starts to change gradually as does their shape.
. Near the outlet,
the shape of the blades assumes the
form of regular impeller
than the irrpeller. Such a con'rbination is impossible when the inducer blade and the impeller vane form one continuous surface. In this case, both the inducer-part and the impeller-part of the blades have to be designed for exactly the same flow rates.
vanes.
The basic idea of such a blade configuration is not new. has been used for years in the compressor industry (Fig. 3), and is used frequently in pumps (Fig. a). However, in compressors, the sole eflect of such a blade configuration is the increase of the pressure-ratio. In pumps, the geometry represented by Fig. 4 is often used for the same purpose as in the modern inducerpeller: to increase the suction capability. However, the improvements achieved with such a geometry were usually very small. Even the most successful designs were much inferior to the results
It
represented in Fig. 2. It rvas evident, therefore, that
to develop an inducerof matching the perforbe capable peller which would mance of a pump provided with an extra inducer, an extensive research program lvas needed.
The problems to be solved were not easy ones. For example, one of the "secrets" of the success of an impeller p.orii"j with an extra inducer (Figs. 1 and 2) is the fact that the inducer is designed for a different flow rate REPRINTED FROM HYDROCARBON PROCESSING
Fig. G-lmpeller used for test shown in Fig.
9.
47
3.500 FPM TAReET
rFFrcrENcy/
f i!?
Fig. _ 7-lnducerpelter
suction performance,
with good efficiency but poor
Fig, 8-lnducerpeller with good suction performance but poor efficiency.
sf
In
cgrlain way, the awareness of this fact happened to .be useful. It showed us the need for a new appro4ch, with geometries completely different from that shown in Fig. 4. a-
80r
7ol
E Ei
oof
20,000 @
s 5,0[ t00[
'ml
25.00
t5.000
8ol
*l ^ol E .ro
zo
,ol
mf
0-0
However, each of the variations tested furnished useful
information, which was used for improving the results of the subsequent tests.
After extensive testing and analytical work, we arrived at a design which was capable of fulfilling all requirements. It has shown performance and efliciencies which were formerly believed to be attainable only with con_ ventional impellers, and suction performance equal to a Aboul the quthor Srr. YSDIpHH is o
eeialist
pump yth Wwthingbn Corp., Eost Orange, iea aie to det;elop and prouide specifi.c solutions to problems relating to centrifugal
[iS, Slneformance of inducerpeller versus impeller with dividual inducer.
in-
pump provided with an extra inducer. Fig. 6 shows a photogiaph of the final design, and its pJrfo.m.rc" is given in Fig. 9. For comparison, test data fiom a standard pump with an individual inducer are shown as the dashed line on Fig. 9. The successful results of the ,,inducerpeller,, design are clearly seen from these graphs. The successful completion of the research program has opened up new horizons for the pump industry. Larger p"-ry operating at higher speeds can be built thus keiping the size, space requirernents and total investment costs
pum,ps, to create neut designs and a,ct a,s cornpanA hgdrauli,c design consul-
tant. Before ioininS Yedidiah serued as t
and, consulting eng ooerseas marutfacturers. He has pub_ Lished many technical papers and articles ond, holds seieral patents. He is a registered professional engineer in an, ouer_ seas countrA and a member of ASME. 48
efliciency, in turn, can bring about signfficant savings in power consumption. LITERATURE CITED 'Prrr.rl", J. H., "Inducer Blades Cut Into Required NPSH,,, pouer, Oct-,
Trra or;rcr of the study was to determine the effect on vibration levels of those factors inherent in the process pump type and the severity of service. Therefore, before selecting the pump sample ,to be studied, it was essential
to eliminate those factors affecting vibration levels that may be attributed to faulty manufacture, installation, and operation.
How to control
pump vibration
The sample of about 100 pumps chosen, therefore, had been mechanical balanced, well aligned, free from cavitation, and operating stably near peak efficiency. The pumps were in an age range of 2 to 25 years. Vibration levels were then charted weekly for one year for each pump. Readings were taken on the two bearing housings in three directions (horizontal, vertical, axial) using a vibration velocity meter indicating "unfiltered peak" velocity values. At the end of the year-long study, an average value of vibration velocity was computed from the 312 (2x3x52) readings taken for each pumP.
Furthermore, in order to ascertain the correlation between pump reliability and vibration level, the main-
END SUCTION
Many factors are involved in causing a pump to vibrate. Smooth operation begins with proper selection and design tor a specilic service. lf you want reliable pertormance, here is what to consider when matching pump to seryice
ci
u
I
j
MDIAL SPLIT
CASING
1A2STG
E
5
B tr G
=
W. P. Honcock, Shell
Curacao, Netherlands Antilles
IEMFERAIURE,
Fig. l-Vibration leVel according
perature of service.
REPRINTED FROM HYDROCARBON PROCESSING
t
(T)
to
pump type and tem-
49
CONTROL VIBRATION
ON
PROCESS PUMPS
TABIE
f-gummqry ol
Max.
VIBRATION DATA
eervlce
temperature
Norm.
acceptable
Max.
DanEler
mm/eec. lns/eec.
mm/oec. lns/sec.
mm/sec. lns/sec
epeclfed
Pump type
258
3.0
520
End(Top)......
Suction
425
Overhung...,.... Radial Split.
1Stage.....
800
Fig. t
800
to Fis,.2
level
20.o
8.0
150 o.32
o.32
on
4.8 0.19
tenance histories were studied
for the last three years. Regarding the inherent factors affecting vibration levels, the following emerged in descending order of rmPortance:
o Pump design a Service temperature o Size effect (u) by increasing size at constant rpm, or (b) by increasing rpm of a constant size o Head per stage and number of stages o Comparison of packed gland pumps with mechanical seal conversions. FACTORS AFFECTING VIBRATION LEVETS Pump type. Pump type is of fundamental importance to vibration behavior. Consequently, samples of each of the following pump types were studied. (a) End suction-overhung impeller type (b) Single and 2 stage - radial split case-volute (c) Multi-stage radial split-difluser (d) Vertical shaft-in line branches-direct coupled
type
The vibration norrns according to pump design and other factors are shown in Fig. 1 and Table
1.
fempercture effect. Undoubtedly the most significant factor affecting vibration levels is the temperature effect which produces pump and driver.
a
change
in
alignment between
If a pump is designed so that its movements due to expansion are controlled and equalized, then a change of alignment with temperature is not possible. The pump may then be termed "temperature compensated." Most present-day standard hot pumps are either "uncompensated" or "pa.rt compensated" and the eflect of this is clearly illustrated on the vibration vs temperature norrns (Fig. 1). Vibration level increases rapidly with 50
More flexible design. Uncompensated designs suitable only for cold
0.60
060
Dynamic performance see above. "Back to back" impeller combination glves thrust problems. Front to front is much improved but gives higher
15.0 0.36
(932oF),
servlce,
150
8.0
to
850
capable of withstanding periods. cglnpqn-
_t-e-r!p. . ins .F). to 500oc
0.79
Refer
453
excellent temperature compensated design. Vibratiou is independent of temperature.
o.32
Fie,2
800
Comments
An
8,0
8.0
Refer
425
Multi-stage. . Diffuser Barrel casing.
o72 Refer to
425
Case-volute
process pump designs ond vibrqiion crilerio
0.60
Existing compensated designs satisfactory to 340.C (644'F) need further re6nement to reach specified 425.C (800"F).
stumng box pressure.
A robust design rvell compensated for temperatul efiect suitable to 450.C. The design susceptible to internal deformatior rrmation causing sticking of hot pumps q'hen priming. Im ing. Improvements required on method of pump suDports suDport for application above 250'C (4820F).
temperature. The standard end-suction designs perform extremely well. They are compensated vertically by suspending the pump about the shaft center-line (in common
with almost all process pump types) but are not compensated in the horizontal plane. The design also has the advantage of a robust shaft between a short bearing span and consequently runs below the first critical speed at 3600 rpm. The standard radial split volute designs having much longer bodies are more susceptible to horizontal movements and the uncompensated types vibrate excessively above l00o C (2120 F). At 3600 rpm they generally operate above the first critical speed, See Fig. 3(b). Designs which are part-compensated by using longitudinal and/or transverse casing keys perform much better to a maximum temperature of 3350 C (6350 F) . For this type of design and the multistage barrel designs, compensation is essential above 1000 C and is desirable even for cold service.
The temperature compensated pump. To maintain alignment under all service temperatures, the pump must expand freely and equally on its pedestal supports. Most existing designs are not permitted freedom to move since they are solidly bolted to the pedestal supports, thereby subjecting the casing to considerable forces in the hot state. This is the major reason why many pumps become stuck when primed with the hot product. In practice, the only solution to this problem is the use of excessive running clearances (up to 0.050 in. at B00o F) which incur significant efficiency loss. The pump feet must be free to move and yet prevented from lifting by special hold down arrangements (Fig. a) which are fairly standard in steam turbines and high-speed boiler feedwater pumps.
TETPEiAII'RE
AH
=
t (f)
GROWTH OF PEOESTAL WITH TEMPERATURE
g (t08 E
-'
t'
III&IATII'E 8[ITE
DOUil
rIffA
0.@
E g
{
ii
E E 0.0'l
l0 PEDESTAL IE|GHT, H,
5:n;,5""'O
fiH
alignment correction curves for pumps on hot
(8) lacfirlGD
Rm h0fit LFflls
w
SHER
t,I tu.oynE mExHI urEmtty ]
Fig. 4-Suggest'ed improved radial split case design for hot service.
the supporting pedestals are maintained at ambient temperature. In practice the pedestals conduct appreciable heat from the pump body (Fig. 2) and the pedestals can grow up to 0.014 in. (0.036 cm) for a large pump at 750" F (3980 C). The pedestal growth can be determined using Fig. 3 and a cold alignment correction made. Vertical displacement of the driver also takes place,
but generally for motor drives, the temperature rise is so small that the displacement is negligible. For turbine drives, the vertical displacement must be taken into account with the pump vertical displacement in order to ascertain the relative vertical displacement. The turbine displacement is usually specified on the manufacturers drawings or can be estimated using Fig. 3 by substituting steam temperature for product temPerature.
For turbines it is also very desirable to use center-line
suPPort.
Horizontal compensation. Pump movements in the horizontal plane can be controlled by using longitudinal and transverse keys. On the bigger designs, some manufacturers use two longitudinal keys, while others use
REPRINTED FROM HYDROCARBON PROCESSING
mtT
lxurlv
Ct (ti[i.)
both transverse and longitudinal keys. However,
lrs{to
XErl@S PftEvBfi
the
biggest fault with existing designs is the solid boltingdown arangements dealt with previously. Of the existing key arrangements in use, two longitudinal keys gave the best results for the radial split casing and barrel types. Of the end suction types, very few were fitted with keys, but the use of one transverse
key at one pump foot and a center peg beneath the pump body proved to be very effective. Although the uncompensated end suction pumps gave good results up
to 3500 C (660" F), the few compensated pumps vibrated ,much less and should be suitable up to 5000 C (9320 F). The multistage barrel casing types are well compensated and suitable for temperatures up to 4000 c
(7520
It
F).
can be seen that the biggest need for improvement
lies with the radial split voluted designs. A suggested improvement is illustrated in Fig. 4. The revised holding down arrangements would be necessary on the longer bodied pumps operating at temperatures in excess of 2500 c (4820 F). Even better compensation can be obtained by designing
the pump and its pipework in order to achieve perfect thermal symmetry. Several vears ago a compensated high speed feedpump 51
CONTROL VIBRATION
ON
PROCESS PUMPS
Size efiect. The effects of size can be produced in two ways.
FIIN
1. Increasing size (eeomctricalll' sintilar designs) at
constant speed. Most process purnps fall rr,ithin a fairll, narrow range of scale factor for ear:h particular design and consequently size factor had littlc eflect orr vibratiorr level. However for general service pumps such as the single stage horizontal split casing types, it is possible to have a wide range of sizes from 2 to l2O in. diameter discharge. For these pumps, size effect can be predicted to have considerable eflect on vibration level. This is because the ratio of unbalanced dynamic forces to pump stiffness iircreases disporportionately with size, thereby increasing the amplitude of vibration and hence vibration velocity. In practice of course, the eflect is always offset
by progressive speed reduction. It is important however to realize that for a given design (retaining the same speed) there is a limit to which the size can be increased if vibrations are to be maintained at a reasonable level.
Fig.
s-An
tr
V
ideally "temperature compensated" design.
MULTI STG.
2 STG. RiqDlAL SPLIT
nPx
Fig. 6-Effect of rpm on vibration level.
was designed and built with twin suction and twin discharge branches arranged symmetrically together with the pipework systems. This thermally compensated designl greatly alleviated problems of distortion and alignment.
IJnfortunately, the duplicated pipework systems are more costly and this may be the reason why the design appears to have lost popularity. However, there are few advanced refinery duties which would justify such a design.
Ideally compensated designs. The pump designs that are ideally temperature compensated are those types mounted integrally with their'driver (usually a motor). The type which is in common use in recently constructed refineries is the direct coupled in-line branches pump (Fig. s). A limited sample of the in-line types was tested and gave excellent results to 2000 C (3920 F). 52
2. Increasing speed of a given pump. For process duties this is by far the most common manner of inducing size effects which greatly influence vibration levels. Considering the pump as a simple undampened spring subjected to imbalance forces, it is possible to predict the amplitude of vibrations with rpm.
Imbalance forces are proportional to (rpm)2, therefore, amplitude is also proportional to (rpm)r. Furthermore, frequency is directly proportional to rpm, hence, vibration velocity is proportional to (rp-)r. In practice, the influence of dampening effects and the pump critical speeds modify the theoretical relationship considerably. A number of pump types were tested in order to obtain an empirical relationship between vibration velocity and rpm. The results are shown in Fig. 6. All the end suction pumps operated as rigid rotor types below 4,000 rpm. Moreover, vibration velocity is approximately proportional to (.p-)'. For the flexible rotor types having the first critical speed below 4,000 rpm, it is not possible to establish a speed correction factor. It is, however, enlightening to notice the close proximity of the critical speeds to the two pole motor speeds of 3,000 rpm (50 cps) and 3,600 rpm (60 cps). This factor certainly accounts partly for the relatively poor dynamic behavior of the flexible rotor pumps compared with the stiff rotor end suction types.
API Standard 610s is very specific with regard to recommendations on critical speeds and it is hoped that these may be fully implemented in future process pump designs. Most existing standard process designs are rated for a maximum of 3,600 to 4,000 rpm and most process duties require a speed of 3,000 or 3,600 rpm for single stage pumps.
Efrecl of heod per s?oge ond number of sloges. It is reasonable to expect that head pe.r stage affects vibration level since all pump casing types give some unbalanced hydraulic reactions, the magnitude of which depends on the casing design utilized and the head/stage generated.
For
process pumps, most manufacturers use single-
volute casings up to 100 M. of head/stage and doublevolute casings for 100-400 M. of head/stage. For multi-stage designs, the multi-channel diffuser inside a barrel casing is in common use. The diffuser gives the best radial hydraulic balance over a wide flow range. Single-and two-stage pumps use
single-and double-volute casings because they are simple and inexpensive. These types provide good hydraulic balance at best efficiency flow but relatively poor balance below 25 percent of best efficiency flow. These designs show signifi'cant increase of vibration level below 25 percent of best efficiency flow (Fig. 7). Consequently, they should not be run continuously below 25 percent of best efficiency flow even though they may be suitable for low flow operation on the basis of permissible temperature rise across the pump. Diffuser designs do not sufler to the same extent because of better hydraulic balance at low percentage flows. The study revealed that
E
E
5
J
Fig. 7-Effect of part load operation on vibration level for a
double volute pump.
for pumps operating near best efficiency point, head/stage
had negligible effect on vibration level.
It
must
.
be
PUMPS WITH PACKED GLANO E PUMPS CONVERTEO TO SEAL
stressed, however, that the impeller-volute generating circle gap should be a certain minimum distance.2 In the past, certain pumps were found to be vibrating excessively and the cause was that this gap had been
reduced to only l/o of impeller diameter by re-welding new casing tongues. When the tongues were modified to give a l0/o gap, the vibration disappeared.
Number of stages.
ft
lElPInAIURe,
rigid rotor single stage pumps, a decrease in operating speed lowers vibration levels and incJeases reliability.
For multistage flexible rotor pumps, it is a diflerent matter, since vibration levels above the first critical speed remain fairly constant with increasing speed.
Furthermore, the advantages of less stages and a shorter, less complicated pump are obvious. High pressure feed pumps are now built to run at speeds exceeding 8,000 rpm and generating over 100 kg/sq cm (1422 psi) per stage.
Effect
r(T'
was shown previousln that for
of pocked glond in dompening vibrolions.
to popular belief, the study revealed little to substantiate the theory that packed gland pumps (with the advantage of additional support) viContrary
evidence
brated less than pur4ps converted to mechanical seals. A sample of 24 identical 4x6 in. end suction pumps was examined, eight of which had been converted to mechanical seals. The results (Fig. 8) clearly indicate that the change had no effect on vibration level. It is conceded, however, that on the very long flexible shafts of the multistage pump, the packed gland may slightly dampen vibration. New process pumps designed especially to incorporate seals permit a much shorter rotating element thereby reducing the maximum static deflection of the shaft to one-half to one-third of the packed gland pumps. This pump type is expected to show a clear advantage over the Ionger packed gland type.
VIBRATION II'WTS The vibration norrns will enable the maintenance engineer to judge the running condition of each of his process pumps against the average vibration level expected for the particular pump design and service. The other vital factor that now has to be considered is the REPRINTED FROM HYDROCARBON PROCESSING
Fig. 8-Effect
of packed gland/mechanical seal
on end suction type pump.
conversion
vibration limits that each design can withstand. In practice it is necessary to define two vibration limits:
tlqximum qcceptqble vibrqtion level. This should be the criterion to be used for all vital process pump (i.e., those that heavily penalize plant production when they breakdown) also those pumps on highly hazardous service. It is the vibration level above which production losses and maintenance costs begin to soar. It can be shown that for a high penalty pump, a pump availability of 0.995 has to be achieved, (availability : hrs. in service
hrs. in repair/hrs. in service) . In order to determine the vibration level consistent with this order of availability, maintenance histories were carefully studied for the last three years. Availabilities were then plotted against average vibration level during the current year for each pump. IJsing this curve (Fig. 9), it is possible to define approximately the maximum acceptable vibration level for any required availability. For vital pumps the vibration limit is therefore 8 mm/sec. -
Donger level. Many
process pumps
category described previously, inctrrring
fall
outside the
little or no pro-
duction losses when out of service. A much lower availability can therefore be tolerated, and hence it is permissible to allow such pumps to operate above 8 mm/sec' However, a much coarser vibration limit has still to be used and this may be termed "danger level." This can be defined as the vibration level above which there is a high probability of imminent pump failure. This is the level which justifies immediate shutdown since there is
a risk of
consequential damages. During the last three 53
CONTROL VIBRATION
ON
PROCESS PUMPS
time to effect a planned repair.
Mrl. AOCFIAETEAVAI.^SUTY
oErt{E f,A(
ACGPIABI.E VIORAN$I Lh,ET
peratures by hand. Any major change
will
certainly be noted and
to call on expert
assistance
in pump behavior
will be suffilieni
incentive
to make a detailed examinastill be a number of complete r,, the financial consequences ere is no financial justification
Fig. 9-Relationship of pump availabitity with vibration levels. years, several trials have been conducted on pumps subjected to very high vibration levels. One end suction pump was run at 50 mm/sec for three weeks without breakdown, but this was exceptional. Another run at 30 mm/sec lasted only 48 hours. A further twelve pumps running between 15-25 mm lasted from two months to
one year.
On the basis of the practical findings, the following tentative danger limits are proposed: (a) Rigid Rotor pumps (e.g., end suction) 20 mm/sec. (0.8 inlsec.).
(b) Flexible Rotor pumps (i..., 1st critical speed (below 3,500, e.g., two-stage volute, multistage) l5
mm/sec. (0.6 in/sec.).
will serve only as a tentative guide. In practice, it will be found that for certain designs on severe service, risk of failure will be imminent at only l0 mm/sec. These
These special cases will be known and will account for only a small percentage of the pumps, the bulk of which will behave in a predictable manner. PRACTICAL CONSIDER.ATIONS
Generc!. Vibration behavior is a sound basis for use with preventive maintenance or controlled breakdown maintenance procedure. In the former case, 8 mm/sec. would be the criterion used for scheduling a pump repair. Pumps in this category would receive vibration logging on a routine basis together with other checks such as noise, bearing temperatures, heads, flows, and power readings. With a skilled running equipment inspector, early diagnosis of future problems is possible in at least 95/o of all cases. The number of unforeseen failures is small and the production and maintenance savings easily justify the costs of regular inspection. Vibration trend lines should be charted for each pump. A sudden rise of 2-to 6 mm/sec. vibration level may be more indicative of early bearing failure than a pump running at a constant 15 mm/sec. for the last six months. In such a case the inspector would check all possible operational causes such as Iow flow operation or cavitation. The frequency of vibration would be checked since a high frequency vibration is indicative of wear of antifriction bearings and in its more advanced stages produces the high pitched sound which signifies that the bearings life is quickly drawing to a close. From early wear in-
dications to advanced wear nornally takes .anything frgm one week to several months and usually gives sufficient 54
Hol olignment. For noncompensated pumps, hot align-
ment is very useful, but must be carried out after the plant has been in operation for sometime and then executed as quickly as possible. It is not sufficient just to prime the pump with hot product because this gives asymmetric heating to the pipework, since the discharge pipework stays cold.
Ifowever, hot alignment can be both dangerous and time consuming and is the partial cure rather than the prevention of the problem. Moreover, hot pipework is constantly moving the uncompensated pump dependent on ambient temperature and wind conditions. Robusiness of the pump design ond its pedestols. The pump desigrr needs to be sufficiently robust to withstand forces and moments from the hot pipework. Most older designs are extremely solid and withstand quite large forces without change of alignment. IJnfortunately, economic pressures are forcing some manufacturers to produce much lighter designs which perform well on cold service but vibrate exessively in hot service. In this respect, it is felt that the section of API 610 dealing with vibration criteria could be improved, since it applies only
to vibration levels of the pump tested under cold service in the manufacturers works. Pedestals should also be fairly rigid. On one end suction boiler circulating pump with high pressure, high temperature pipework, the pedestals were so flexible that the coupling alignment changed 0.150 in. when the pump was heated. Considerable reinforcement was necessary to obtain a satisfactory result. GONCLUSIONS
o Vibration
measurement provides a sound basis for establishing the running condition of process pumps. The
vibration data evolved by the study will provide an excellent foundation on which preventive/corrective maintenance programs can be built.
. For the standard process pump, the principal
inherent factors affecting vibration level are pump design, temperature of service, and rpm.
. Vibration behavior under ideal cold conditions will not be sirnulated under actual hot working conditions unless the pump is temperature compensated. Many existing process designs are not temperature compensated and perform poorly at elevated temperatures. The in-line pump and the barrel casing designs are
well compensated and vibration is virtually independent temperature. Most end-suction and single/2 stage
of
Fig. 10-Process pump selection chart showing limits of ilows, heads and temperatures.
radial split designs are nontemperature compensated and the temperature effect on the larger bodied radial split designs is quite considerable.
o All standard process designs have solid bolting down arrangements which do not permit free (yet controlled) movement of the pump with temperature. Consequently, the casing distortion duing priming with hot product often makes the pump rotor stick. The only practical remedy is the use of excessive running clearances which considerably reduce pump efficiency. The most suscep-
Aboul fhe quthor Paur, HaNcocr
is head of rotating
equipment engineering for Shell Curacao, Netherlands Antilles. He is responsible f or application, troubleshooting, failure aho,lysis and field deaelopment
studies
of refinerg rotating
equipment
ulhich cotttprises ouer 2,500 pllnlps, conxpressors, blowers and turbines. Prior etperience includes seraiec as a pump design engineer u;ith, Matlter and Platt Ltd. and project engineering with the Central Electricity Genero,ting Bcard. Paul uas born in England and receiaed, a B.Sc.Teclt. degree in mecltaruical eng'ineering from the Unittersity of Manchester. He is a chartered engineer of the Institute of Mechanical Engineers (London) and a member of ASME.
REPRINTED FROM HYDROCARBON PROCESSING
tible of all pump types to this phenomenon are the long multistage designs. For designs of this nature, considerable improvement could be made by using holding down arrangements incorporating either the "paw foot" design or the shouldered bolt which allow free lateral and lon-
gitudinal movements but prevent vertical movement. These more sophisticated designs would be justified for pumps operating above 2500 C (482' F).
. End-suction and in-line designs give excellent mechanical performance. Furthermore, they have low initial cost, are easy to repair, and have a wide degree of interchangeability. Consequently, they should be chosen as widely as possible for process duties. In certain cases, two-stage and even multistage designs are used for duties that could be satisfied by end-suction designs. o A complete guide to the limitations of head flow and temperature are shown in Fig. 10. The vibration norms according to design and severity of service are given in Fig. I and Table 1 which summarizes the main findings of the study. ACKNOWLDDGMENT The author is indebted to the practical assistance rccciycd from Jr B"i1: in the substantial field worked involved with the study. Thir papcr origin4h Eneineering Technial conlcrcncc, cin' 3i.",ill.l"3n1l,,$.13_Yf ,De5ign
ITERATURE.IrED
""
jjii,'ffi 'I 55
Fig. 1a-By viewing the impeller through the suction portion of the pump case, the initial stages of cavitation 'can be
Fig. 1b-Little noise or reduction
in
pump p.erformance is noted in lhis stage of cavitation. lmpeller vanes may shdw some pitting after extended oferation under these conditions, especially if there is any tendency toward corroSion
seen with vapor beginning to form on the entrance edges of the vanes.
present.
Fig... 1c-As a result of increasing the suction on the pump, which decr6ases the pressure in the entrance of the impeller vanes, the vapor extends over a large portion of the entrance. Noise reduction of efficiency should be evident
by this stage of cavitation.
How to prevent pump cavitation Cavitation can destroy process pump perlormance and, in time,
will destroy the pump itself. Caretul attention to fluid conditions at the pump suction will minimize the possibility
at the impeller
vanes
by further increasing
of cavitation
marginal at this point,
0""?3,?li'l;,."L1.'"",:ZE
Fig. l-Photog{aphs_.91 cavitation in action courtesy of peerless Pump Division, FMC Corp.
Chorles Jockson, Monsanto Polymers and Petrochemical Co., Texas City, Texas Cevrrerrorv with a ma
specific thing clran and result at
is a breach of contract!
IJsers c,rntract that will perform
su'rf ac
Cqvilqfion occurs when the pump suction pressure drops causing some liquid to vaporize and bubbles to form. As these bubbles enter the pump vanes, the pressure increases until the bubbles implode or burst inward making
e'
c avi tati
on'$":ii"
T;,I",:,':Jfi
".i,f ,r:!:n ;ilf
ching the contract
i." the basic
against the vane surface. A pump impeller will have a "shot piening,, on the
exposed to cavitation
th parties. Some-
factors that lead to cavitation and how"." to avoid them.
56
a "ping"
:,"
"
":ft
iT"?ilI
.:ffi ,f,'t'iji:;
The threshold pressure required for a vaporized liquid
to implode is called the "suppression preisure.,, pump manufacturers perform a suppression test used for NpSH data.
Fig. 1 illustrates the cavitation process as seen on impeller vanes viewed through a transparent suction line.
HOW TO PREVENT PUMP CAVITATION
Fig. 2 shows a damaged impeller taken from a 30-inch cooling tower pump from our Chocolate Bayou, Texas, plant. Note the picture through the mirror showing cavitation damage progressing through the vane.
Whqt is NPSH? It is the Net (remaining)
Positive
(greater than zero) Suction Head (absolute pressure at the pump suction flange expressed in feet of liquid pumped above the liquid's vapor pressure at the pump-
ing temperature). To the system designer, NPSH is the remaining pressure (net) at the suction flange of the intended pump af.ter all negative forces that restrict liquid from getting into the pump are subtracted from all the positive forces that assist liquid in getting ino the pump. NPSH varies strongly with change of flow. Most pump curves are plotted with NPSH versus capacity, Q, in gpm. But NPSH characteristics allow flows not frequently plotted by pump manufacturers. It is not uncommon for the NPSHR curve to increase or turn upward at low flows (see dashed lines on NPSH curve, Fig.3). Low flow ef-
Fig. 2-Cavitation damaged impeller. Note reflection from the mirror of damage to back side of impeller.
fects can often be amplified by selecting too large a pump for the service. This forces recirculation within the im-
pellerl (sirnilar to "surging" in a centrifugal compressor) . An instability point on the H-Q curve is now indicated by some pump manufacturers (see dashed lines, Fig. 3). NPSHR. Purchased pumps should have received a care' suppression test following the original pump design. This test should accurately predict the pump's suction performance on some standard fluids. Water is usually used. The NPSHR (Net Positive Suction lleads"n,1."6)
ful
by the pump manufacturer is confirmedlby this test. To the pump manufacturer NPSHR is the head of Iiquid required, at each proposed flow rate, for the liquid to travel from the suction flange into the pumping vane area without vaporizing. Vaporization can be thought of as "cold boiling" from the pressure drop in the liquid as it flows from Point A to B (see Fig. 4) . If NPSHp colrditions are met and entrained gas is low ( ( 1 percent by volume),' the pump manufacturer states that the pump should perform according to the H-Q curve provided for that size, impeller and speed. (See Fig. 5 for effects
of entrained gas on H-Q
curves.)
NPSHA. The purchaser or user must calculate the NPSH1 (Net Positive Suction lleade,,rrurr") at the suction flange
of the pump for the range of
system Head-Capacity
(H-Q) cu.rves. All equations for determining NPSHa are simply
equa-
tions in fluid flow. Push fluid to
.
o
Pressure over the fluid. Static head of fluid above center
of suction flange.
NPSH^=1
Note: Sometimes the reference point is to the pump centerline or an agreed reference (datum)
Fig. S-Typical H-Q curve. (Courtesy Bingham Pump Co.)
Suppression lesl. The pump manufacturer provides a suppression test (also called cavitation test). The test procedure is outlined in ASME PT 8.2 for centrifugal pumps. A test set-up is shown schematically in Fig. 6. Measured flow is sometimes provided by tested flow nozzles changed out consecutively. An accttate, often calibrated flow measuring station is also used. Example. A 200-gpm nozzle is installed and the pump is throttled on the suction until the purnp cauitates. The NPSH value is plotted versus 200 gpm. Actually, audible cavitation is seldom obtained but the beginning point of cavitation is considered to be the condition at which a three-percent deviation in total head occurs. Next, the 300-gpm flow nozzle is installed and the procedure repeated progressing to maximum pump flow capabilities. Push fluid back from pump
o Friction
loss in piping system (including entrance or exit loss from vessels).
o Vapor pressure
of liquid.
elevation.
o Atmospheric pressure. o Velocity head. REPRINTED FROM HYDROCARBON PROCESSING
57
&
i.r,
?+:
lr' ENTRANCE LOS 1:'
I
l:.
lr,
ft ;$ al, ar>
l^,
e. q. C'
,.,
INCREASING PRESSURE DUE
E at
URBULENCE, FRICTION,T
ET
ENTRANCE LOSS AT VANE TIPS
G (J IE
\
TO IMPELLER
POINT OF
PRESSURE VAPOFXZATION SIARTS
-WHERE POI'IITS ALONG TIOUID PATr
Fig. 4-Cross-section of a typical pum.p shows the pressure drop area (A to B) for vaporization (courtesy Bingham pump Co.). Relative pressures in the entrance section of a pump are showrr on the ihart airight, Note pressdre drop'betw;en poinis'A an'o B (courtesy The Duriron Co., lnc.).
o
Plugged suction or suction screen,
.
Entrained
ts
gas.
More often, conditions at the suction flange to occur because:
6 a
causing perforrnance failures
P
o Pumping temperature
. 2p@
increases,
Suction vessel pressure-over-the-liquid drops, blanket
e.g..
gas,
o
Suction level is extremely low,
o
Change
in iiquid composition increases vapor
o Flow rate is significantly o
change
pressrre,
increased purposely,
System head pressure decreases causing an unexpected
flow rate increase,
Performqnce. The performance curve (Fig. 3) indicates that at 180 gpm and a head of 215 feet, the NPSHp is 10 feet. Calculated NPSHA might be 72 feet for this example. Note: This pump will cavitate at 216 gpm if NPSHA is 12 feet.
o
Partial suction line restriction,
.
Gas is drawn into the suction exceeding trvo percent by
volume of gas, blocking flow2 but often causing an inlet p.ressure drop, forcing cavitation,
Provided the pump manufacturer made the suppression
test, the pump could dition because of:
fail to perform at operating
o Increased circulation from wear rings or
con-
Deteriorated impeller, e.g., inlets,
imperfect vane,
plugged passageway, corrosion/erosion holes, vane exits,
58
Poor suction pipe layout,3
o
Gas-liquid multicomponent mixtures.a
clea.rance
settings upsetting flow into the impeller,
.
o
Suction specific speed, Iike specific speed, is a design for pumps and is defined by the equation, S = (n VQ) (h,'/o) parameter
HOW TO PREVENT PUMP CAVITATION
Fig. 7 is shown as a guide for lower threshold limits of NPSH based on suction specific speed limits. Note that capacity and suction speed are determined at B.E.P. (Best Efficiency Point) for the pump in question;
NPSH versus speed. NPSH will vary as the square of the speed. This relationship was illustrated well by Karassiks using the suction specific speed relation: S
Since
: (r, V b
I (h","/n) = (n, l/6 (h",1 h",)s/4 : (nz V@) I @,
Q2/Qt: nz/?l,
Then (h"2/h"r.)s/' And hu2f hs=
-
I (hu,"/o)
l/ e,)
(nr3/2) f (nrs/z)
(nrln,),
Where subscripts: 1 = initial condition 2 = final condition
Fig. 6-Manufacturer's test loop to make a suppression test (courtesy Peerless Pump Division, FMC Corp.).
The speed squargd versus NPSH relationship is useful. had occasion to evaluate a 1200-rpm, 28-inch-diameter, Canadian built rubber-lined pump to the manufacturer's U.S. maximum 750-speed ratings. The extrapolated NPSHR indicated approximately 32 feet. The manufacturer agreed but our specification requested no infofirration on NPSHR. IJnfortunately, NPSHa was considerably less than 32 f.eet. The rubber lining was eliminated and the vortex breaker modified to correct the problem.
f
DESIGN AND SPECIFICATION SUGGESTIONS
Fig. 7-Approximate minimum limits for required NOTES:
1. Calculate the NPSHn carefully considering all conditions, i.e., startup, original fill of lines, winter operation, all control valve pressure drops, summer operation, piping condition, exit and entrance losses.
NPSH,
l.
Caution should be used in applying pumps with NPSH require- ment. Iess than shown above-- 2. NPSH limits apply only to the besr efficiency point. 3. R^equired NP-SH limits are based on a suction-specific speed of 13,000. ITpellers with lower suction specific speeds williequire increased NPSH.
4. For double suction impellers, use
f oI pump
capacity.
Where
S: suction specific speed n - rotauve speed, rpm Q: flow, gpm /2"
= NPSHg (required) for satisfactory
operation, feet. For double-suction pumps, use r/2 total flow, i.e., ,/,
total Q.
About fhe qulhor CrrARlES Jrcrsow is an
problems at the
Monsa,nto
Polymers ond, Petroleurn Co., Tetas
Citg, Tenas. He has been with the compang for 22 Aeurs ,worlring in Dioi,si,on Engineering, Production, Maintenance and, Plant Design, before setting up the
present depo,rtment in 7960 later being appointed eng'ineering specinhst, then senior engineering specialist. Mr. Jackson holds a B.S. degree 'i.n mechanical engineering from Tenas A*M Uni,aersitg, plus
an AAS
degree
in electronics technologg from College of
Mainland, Teras C,itg, Teuas. He is eng'i,neer,
o,
3. Provide NPSHA
at
2-feet
minimum over worse
conceiaed NPSHR curve by the manufacturer. (For carbonate solutions, consider more.6) If entrained or evolved gas is possible, consider self-priming volute impeller designs or at least open impeller construction. Note: The effect of entrained gas is not cavitation hut it can con-
tribute to cavitation. 4. Request a suppression test if NPSH margin of "available" over "required" is less than 1 foot. 5. Concern yourself with tle conditions at the pump suction, fluid composition, vapor pressure, etc. 6. Consider the suction specific speed, ^S, of a pump in service application.
engineering fellout in a Mechanical Technologg Section utorking on rlechanical and methoil-
oriented,
2. Be skeptic'al of saturated liquid-gas compositions, e.g., flash tanks at suction lines.
the
7. Do not use hydrocarbon corrections in actual practice. If you get a plus from.it, accept it with a smile, and go on about your business. LITERATURE CITED I a a
ccqtrifueal and _rcciprocating pmpr,tl D. l0ll Feb. 2a, tE70,Centrilugal P.mps, F. W. Dodge Corp.,
registered prof essi,onal
a menlber of ASME, API and S,ESA.
REPRINTED FROM HYDROCARBON PROCESSING
59
Notes
60
Turboexpanders
!4 r'.-
:r
"E:i Y
,
New developments in hot gas expander compression systems Better machining techniques and proven analytical methods provide a new approach to compression systems. Here is a technology update. Hons D. Linhcrdl, Airco Cryogenics Division of Airco Inc., Irvine, Calif. SrNor,B and two-shaft hot gas expander compression systems will soon replace conventional multi-train, lowspeed equipment due to significant economic and process advantages. The single-train concept is a direct result of the recent advances in high pressure ratio compressor, high performance radial-inflow turbine and high performance steam turbine technology. The application and custom engineering of single train compression systems is discussed, with emphasis on performance and reliability. Conventional compression systems, such as required for nitric acid plants, consist of low speed multi-train systems which have been manufactured for the past three decades. While this approach has been constantly improved, t}le
underlying technology is unchanged. Such a system generally consists of two trains of multi-stage low pressure ratio per stage compressors, one multi-stage axial hot gas expander train and one steam turbine train all coupled together and installed on a large machinery platform or mezzanine.L Over the years the capital cost of such a system has increased from 10 percent to about 30 percent of ,the capital plant cost. While these conservative and heavy designs work reliably, they do not allow much process flexibility and do contribute to decreasing plant contractor's profit. Considering current stringent environmental plant speccifications, new process operating conditions must be imposed on the turbomachinery systems. These conditions are increased hot gas temperature and increased over-all compressor pressure ratio. Since most conventional equipment has reached its upper limit in flexibility and performance output, a new look is required at modern innovative technology. Indeed, there is a wealth of significant advancements to be considered: radial compressors with pressure ratios as high as 12:I,'z high temperature radial inflow turbines3 and high performance, low specific speed axial turbines suitable for high speed steam turbine requirements.a,s In addition, proven analytical methods provide confidence in developing stable hydrodynamic 62
bearings and rotating assemblies. New construction materials help solve problems with highly stressed temperature subjected turbine and compressor wheels. The terminology 'tustom engineered" is not a new sales slogan but is an established program of optimized technology as demonstrated in the natural gas processing industry.G These systems are operating reliably and comply with all the process specifications. In light of these significant achievements, it is time to incorporate the latest technology into custom engineered compression systems such as for nitric acid plants, large SNG tail gas expanders and methanol tail gas power recovery systems. In comparison to the significant achievements of the gas turbine industry these technical advance-
nents are conservative and modest. Ilowever,
they
provide a significant growth potential and process flexibility which cannot be realized from existing conservative, conventional designs.
IN TURBOTIACHINER,Y TECHNOLOGY New developments in methods of analysis and construcADVANCETIENTS
tion allow custom engineering of major turbomachinery
TIIfEBL M'ASAMETF
,.ta+!q0rt0-t mEm r .TIEUFE
f . rt&+ tqrl
Fig. l-Miller-Larson Parameter.
tfa
HOT GAS EXPANDER COMPRESSION SYSTEMS
auftl klT maatco X-X-c5A
cnEEP
Fig. 4-Titanium impeller forgings re-
sulted in the design ol a 4.7 to 1 pres-
sure ratio air compressor driven at 20,000 rpm.
8rt{liurrr!r!laro.l Ff lA.(@nrPl
this tool available and understanding the design of reliable, stable, high-speed hydrodynamic bearings,8 the custom engineering of single-train multiple impeller designs becomes
a reality, Besides rotor dynamic considerations, of complex three-dimensional turbine and
stress analysis
Fag.2-A-286 stress data as a functlon of Mlller-Larson Parameter.
r /o.* r' .Till-i.ip" r!=si{aFl
sPEeo (RPM) Qr:EiHAUST VOLI,ME FLOTY (CU fT/SEC, Ats:IOEAL EilTHALPY OIFFEREI{TIAL (SIU/LB)
compressor wheels is required. The effort related to the development of efficient gas turbines has provided refine-
ment of this method. For example, a stress analysis method is available for turbines and compressors that includes the effect of thermogradients.e With development of high strength materials, it is possible to design reliable turbine and compressor wheels with tip speeds of 2,000 feet per second.1o The understanding of material, stress, rupture and creep is another factor which allows reliable and safe prediction of long duration operation of highly stressed turbine and compressor wheels. One of the methods which allows correlation of stress
and long duration is the "Miller-Larson" para-
meter,11 presented in Fig. 1. Pig. 2 presents the allowable stress as a function of the "Miller-Larson" parameter for
ipEcrFtc
aPE€D,
tr!
Fig. 3-Efficiency of various expansion turbines vs. specific speed,
components for gas processing and petrochemical plants. For example, new methods of analysis allow accurate determination of complex rotor dynarnics including the effect of hydrodynamic bearings. The latter includes the stiffness of the hydrodynamic film and the dampening
effect of the film and the mounting structure. Accurate rotor dynamic computer programs allow detailed analysis of multiple impeller shafts including the effect of gyroscopic moments due to overhung wheels.T The rotor dynamic programs developed within the last 10 years allow prediction of critical speeds due to bending or hydrodynamic effects within an accuracy of 2 to 3 percent. Having REPRINTED FROM HYDROCARBON PROCESSING
A-286, one of the most interesting materials of consEuction for hot gas expanders. With stress, rotor dynamics and bearing design all well in hand, areodynamic advancements encourage the reliable design of single-train units. The specific speed concept provides a reliable guide for selecting the optimum parameter for the turbine and compressor designs. 6'12 For example, Fig. 3 demonstrates the correlation of ma>
radial compressors.2,'2 Considering the present state-of-the-art of petrochemical process compression systems, it is obvious that modern radial turbine and high pressure ratio compressor technology has not been applied. This is because axial turbine and compressor technology was developed four decades ago while radial turbine and high pressure ratio compressor technology has been developed in the last decade. It is surprising that the radial turbine has not found its 63
TABLE l-lmpeller processlng steps Sequrnce Decrlptlon Melhod ol taak 1. Selectlon ol basic Gomputer program (COMPO) 2. 3. 4.
geometry Design llow passages Wheel stress analysis Shaft dynamlcs analysis
if 2:1 press ratlo (COMHP) Gomputer program (WHEL) Computer program (STRES) Computer program (CRTSP) Computer program (THRS) Computer program (KITPNCH) or
8.
Thrust analysis Prepare tapes for contour and flow passage Cut master tool Turn and bore impeller
9.
forging Machine flow passages 3D duplicatlng mlll
5. 6. 7.
E
i4OIOF SP€ED
a o
N
.
T785 FPM
Pinhr
aO do
.
t3-oD PSIA
Tinhr.gOoF
E
:
ICV'r
60o
IHP G/BOX. l0O
G
o
NC 3D mill
NC contour lathe
(blades)
Blend end polish bladlng 11. Finish turn and bore 12. Cut spline 13. Blade and disc vibration test 10.
14.
Batance
15.
Spln test
Assemble, rebalance and lnstall 17, Test-operatlon and performance 18. Data feductlonperformance 16.
Bench
Its
NC contour lalhe
-t
6
Broach
Test lab. Electronio dynamlc balancer Spin pit Assembly room, etc.
Test facility Gomputer program (OFFD)
la
,6 L0s€c
Fig. S-Performance ol a 4.7:1 pressure ratio air compressor.
place in the field of hot gas expanders while the earliest turbines for hydraulic plants were the radial or Francis type of turbine. The detailed aerodynamic design of high pressure ratio, radial inflow turbines and compressors can be accomplished with considerable confidence due to better understanding of three-dimensional flow and the wealth of experimental d,ata supporting detailed efficiency predictions and correlations. Having the aerodynamic design in hand, it is necessary to develop new machining techniques which allow custom
machining
of
various wheel designs optimized
for
one
plant or process. Table 1 for example, summarizes the processing steps of advanced impeller designs. The significant achievement in manufacturing custom engineered turbomachinery is the application of nrmerically controlled machines for the manufacturing of turbine and compressor wheels. Applying advanced but practical engineering computer programs, it is unnecessary to develop a drawing of the three-dimensional blade shape. A computer tape is processed which can be applied directly to the three-dimensional milling machine, which uses the tape input to directly generate the required blade shape. This method makes it feasible to custom engineer optimum aerodynamic blade shapes for each application. The tooling can be programed to produce blades with resonant frequencies far above operating frequency. With the above methods of analysis and the available produc-
tion tools, significant advancements have been demonstrated in the aircraft industry and are now being applied to the gas processing industry. The most significant achievement in the aircraft industry has been the develop54
ment
of the high-pressure ratio radial compressor. This for the design of a high per-
development was necessary
formance lightweight gas turbine. Continuous research and refinement of analysis has resulted in the demonstration of a 12:7 pressure ratio, radial compressor which has achieved an efficiency of 74 percent., The design tip speed is 2,000 feet per second for 70oF ambient conditions. A detailed stress analysis shows that by using titanium forgings, reliable 100,000-hour operation can be achieved. Applying these advanced techniques to commercial applications resulted in the design of a 4.7 to I pressure ratio air compressor driven at 20,000 rpm by a 2,000-hp electric motor through a step-up gear box (Fig. a) . The performance of the 4.7 ratio compressor is shown in Fig. 5. Eficiency is 82 percent and margin between choke and surge is larger than 15 percent. Similarly Fig. 6 shows a high pressure ratio compressor for starting abcraf.t and Fig. 7 shows a custom engineered 2: 1 pressure ratio methane compressor. Having demonstrated the feasibility of designing reliable high pressure ratio compressors, the next step should be to eliminate ,the multi-stage axial compressors by increasing the speed and reducing the number of stages. High pressure ratio radial compressor technology is one of the key elements in development of reliable single-train compression systems. The other important
component is the high performance radial inflow turbine.
The radial inflow turbine has attained significant stature in the field of cryogenics.G, 13 Machines up to 10,000 hp have been custom engineered. In the field of hot-gas turbines a 1,600-hp radial inflow gas turbine has recently been developed.la Numerous smaller gas turbines as well
HOT GAS EXPANDER COMPRESSION SYSTEMS
Fig. 6-700-hp high pressure ratio
compressor,
alr
Fig. 7-2:1 pressure ratio natural gas compressor with diffuser assembly.
as turbo super-chargers have been used for some time. ffowever, it is surprising that the radial inflow turbine has not been applied to large power output hot-gas expanders. The radial inflow turbine has a large advantage over the axial turbine since it can operate with a larger pressure ratio and therefore the average blade temperature is much lower, assuring higher reliability and safety, The most significant work in high temperature radial inflow turbines demonstrated the feasibility of a 2,000oF, high tip speed, radial inflow turbine wheel.s Applying these results to 1,300oF inlet temperature radial inflow turbine requirements, a reliable solution results by using A-286 forgings.
The third component of interest in high performance high performance steam turbine. The high performance steam turbine is a unit basically developed from APU (auxiliary power unit) technology. In addition, significant work has been done in the field of liquid metal turbines.l6, 18 High pressure ratio turbines have been developed for exhaust moisture contents as high as 20 percent.lo Contemporary design of
Fig. 8-Steam turbine and methane compressor wheel.
The gas processing industry has developed methods for testing compression systems prior to shipment under conditions similar to anticipated operating conditions.6 Single train compression systems can be tested economically by using external, natural gas fired combustion chambers in conjunction with laboratory steam power sources. The complete system test is a requirement in order to assure minimum problems during plant startup. This acceptance test assures the utmost in reliability and safety. In addition the machine can be tuned to the highest performance posible. ACKNOWLEDGMENT
systems is the high pressure ratio,
LITERATURE CITED
reliable supersonic nozzles and selection of corrosion resistant materials allows reliable and erosion-free operation. When high pressure ratios are expanded in turbines of relative small geometry, a high expansion velocity results which in turn assures sub-micron condensation droplets. These droplets follow the stream lines with minimum slip and therefore erosion is eliminated. Fig. B shows a high performance steam turbine operating with a 1,700 feet per second tip speed and expanding steam from 350-psi steam to ambient pressure. Steam turbines of this type are
machined by the electro-discharge method from a solid forging. This method of tunbine design can obsolete the conventional multi-stage low-speed steam turbine. Besides reliability in steam turbines, advances in reliability are required for the accessories. These high performance turbines have no problem conforming to the latest API requirements. Having developed the capability of custom engineering of advanced turbines, compressors and steam turbines within one train, the next question is: how can this custom engineered unit be tested prior to shipment? REPRINTED FROM HYDROCARBON PROCESSING
65
Turboexpanders recover energy When large gas flows must be reduced trom high to low pressure or when high-temperature process streams (waste heat) are available, expansion turbines should be considered. Here's how to calculate the work recovered
ali power-recovery drives. This type of turbine normally operates at relatively low flows and high rotational speeds. APPlICATIONS
V. H. Abqdie, Brown & Root, Inc., flouston AN nxpaxsroN
TURBTNE converts
the energy of a
gas
stream into mechanical work as it expands through the turbine. The expansion process occurc rapidly, and heat transferred to or from the gas is usually very small. Consequentln tJre internal energy of the gas decreases as work is done, and the resultant temperature
or vapor
of the
gas may be quite low giving the expartder the as a refrigerator as well as a work-producing
ability to act device.
furboexponder lypes. Expansion turbines may be into two broad categories, axial-flow and radial-
Power recovery. A potential application for expansion turbines exists whenever a large flow of gas is reduced from a high pressure to some lower pressure or when high-temperature process strearns (waste heat) are available at moderate pressures. When such conditions exist, an expansion turbine can be used to drive a pump, compressor or electric generator, thus recovering a large portion of otherwise wasted energy. In applications of this type, careful consideration should be given to the temperature drop which will occur in the expander. It may sometimes be necessary to heat or to dry the inlet gas to avoid low exhaust temperatures or the formalion of liquids.
Expansion turbines can be designed to cover a wide range of operating conditions. One manufacturer lists current limix on inlet conditions as 1,300oF at 300 psig and 950oF at 1800 psig with no limit on minimum temperature or on minimum or maximum horsepower.
Refrigerotion. Turboexpanders may be used in closed refrigeration cycles with a pure gas such as nitrogen, which is alternately compressed and expanded to provide
classed
the required refrigeration through a heat
flow.
Various types of open cycles may also be devised so that
Axial-flow turbines are those in which the gas flow is essentially parallel to the axis or shaft of the turbine. Turbines of this type resemble a conventional steam turbine and may be single-stage or multistage with impulse or reaction blading or some combination of these designs. Turbines of this type are not normally used for producing low temperatures but are essentially power-recovery devices and find application where flow rates, inlet temperatures of total energy drops are relatively high. Radial-flow turbines are those in which the gas flow is essentially at right angles to the turbine shaft. Radialflow turbines are normally single-stage and have combination impulse-reaction blades and a rotor that resembles a centrifugal-pump impeller. The gas is jetted tangen-
tially into the outer periphery of the rotor and flows radially inward to the "eye" from which the gas is jetted backward by the angle of the blades so that it flows axially away from the rotor. Radial-flow turbines have been developed primarily for the production of low temperatures but may also be used 66
exchanger.
the process stream to be cooled passes through the expander thus the need for low-temperature heat exchanger. Liquid products can be produced directly from the turboexpander, provided it is specifically designed for this type of service. Similar cycles can be used to recover condensable products from process streams. The use of open cycles requires the removal or control of contaminants such as water vapor and carbon dioxide to prevent foul-
ing of the equipment. This problem is avoided in closed cycles by using a pure, dry gas as the working fluid. To make an economic analysis, it is necessary to understand the expansion process and to be able to predict the turbine performance, specifically, the work output and exit temperature produced by the turbine. EXPANSION
PROCESS
All real gas processes, either ideal or non-ideal, are irreversible to some extent. Ilowever, it is necessary to assume reversibility to illustrate an expansion process on a thermodynamic diagram. A typical expansion is illustrated on the pressule-enthalpy diagram, Fig. 1. In
TURBOEXPANDERS RECOVER ENERGY
of the expander and the gas inlet conditions and outlet Pressure.
fsentropic efficiency is defined by the following equation:
\": W/IA,
(1)
and the work done by the gas is W = rlsWs As shown above: W": Ho- H6 and therefore
W: \, (H"-
H1
Ho) =
Ho- He
(see
Fig. 1).
(2) (3)
The actual final temperature,7", can be read from the Mollier chart at the intersection of H" and Pe. If the expanding gas is a complex mixture for which a Mollier chart does not exist, another approach must be used. One possibility is to use a computer program which will calculate the enthalpy and entropy for the mixture to determine the isentropic enthalpy difference. Some of the turbine vendors use special programs which calculate actual turbine performance. When computer calculations are not available, sufficiently accurate results can be obtained by either polytropic or isentropic analysis.
H. EflTHALPY
-
Fig. l-Pressure-enthalpy diagram showing typical expansion through a turbine,
this diagram, the initial state of the gas is assumed to be at point a having pressure Pr, specific volume Vy temperature To, enthalpy Ho, and entropy So. Point b represents the terminal point of an isentropic expansion, and
the path from a to b represents the sliccession of states through which the gas would pass when expanding, without heat transfer in a reversible machine. An actual irreversible expansion from point a would terminate at some point such as e having the same pressure as that at b, but with higher temperature, enthalpy and entropy. If the expansion from point a were extended to some pressure lower than P2, terminating at point f, it would have entered the two-phase region under the saturation curve, resulting in partial condensation of the gas. If the expanding gas were a mixture rather than a pure component, there would also be a composition change as a result of the condensation. The relationship of heat, work and energy for an expander may be expressed by the first law of thermo-
Polytropic onolysis. Equations developed by Schultz' can be modified to suit expanders. To use this approach, imagine a reversible expansion path between points a and a in Fig. 1. Point a represents the actual condition of the gas at the expander inlet, and point a the condition at the outlet. This reversible expansion path is defined by the equation pVn = constant (4) The work done by the gas expanding along this path is the polytropic work.
w,:
*, : (;:-)
- s,1rs,-',,*f "#?lr The terminal temperature at point e can be calculated
AI{: Q_W A well-insulated expander with normal mass-flow velocity will be very nearly adiabatig and Q may be assumed to equal zero, giving the fundamental expression for work done by the gas:
W=-A,H This equation shows that the work done by any
gas
expanding adiabatically is equal to the actual change in enthalpy. If the expansion is also isentropic, the work
W": Ho- Ht
Turbine performonce. If a Mollier chart for the expanding gas is available, enthalpies may be obtained directly from the chart; and the work output, W, canbe calculated directly knowing only the isentropic efficiency REPRINTED FROM HYDROCARBON PROCESSING
(5)
The actual work is Vl/ = noW, (6) The work W can be calculated by Equation (6) when I4le is known. Wo is obtained by integration of Equation (5) using Equation (4) and n = constant. The results of this integration, when combined with the equation of state, PV : ZRT, yield the well-known expressions for polytropic work, one of which is the following:
dynamics as follows:
done is defined as the isentropic work
* fi' ,0,
(7)
as
T"
:
To
(Pr/P)*
(8)
Equations (7) and (B) can be solved for the required values of work and exit temperature using the exponents n and m obtained from the following equations:
n:lffYooo-m(l+X",r)) Zroo R , m:#(rto+X*o) I I Ooo6oo
(e) (10)
The compressibility functions X and I were calculated by Schultz and plotted as functions of reduced pressure 67
6 o a l J 9 ilt 3 =
02 04 06 08 t0 t2 t.4 t6 r8 z0 22 24 26
o
2
f,EOUCEO PFESSURE, Pn
REOUCEO
and reduced temperature. (See Figs. 2 and 3.) These functions may be difficult to determine accurately from the curves in the regions of low Pn and Ta. Ilowever Schultz has indicated that X and Y can be closely approximated by the following expressions in the regions
<
0.9, 7E
<
1.5
andZ
)
X = 0.18116 (8.361,t, Y :0.074 (6.651,t, +
A7"
To
0.6: 0.509
X, Y, Z, and cp are avetages expansion path. It may be necessary to break the path into several increments, using the appropriate average values for each increment. The calculation must be done on a trial basis using assumed or estimated tem,peratures to obtain X, Y, Z and cp, a,nd adjusting them until an adequate check is achieved. lsentropic onolysis. Using isentropic efficie-ncn it is necessary to calculate the isentropic enthalpy drop Hr- Ha. Equations for this calculation are similar to those for polytropic work2 and are obtained by integrating Equation (5) along an isentropic path defined by
=
I (___!u_\z,RT"l- ,,,o,,r ,r, ,, _ (12) t ns-Dln. " - \2" _ l/ z./l L, \t 2/ 7) ) Using Equation (12) , W" may be calculated with trial
values of. n", and the results checked by Equation (13) until a satisfactory value of ft has been obtained.
:
fcrooo
(7" - 7'r)l / 0 * Xoo)
W't
:
(l + X)
(14)
cPo,o
- To :ZAT"'
To: m,:
(15)
(13)
f(Zo,oR)
T"(Pz/P)^"
(r6)
/ (778cr",,)l(l * Xoo)
(17)
The isentropic work is:
W,:2W,i:Ho-Ho The actual work, using Equation (3).
W: Ha-He
( 1B)
can then be obtained
The isentropic work may also be obtained by the use F,quation (18) and the following approximate ex-
qf
pressron:
W,i
constant
This integration gives
W"
!0
T6 maf also be calculated from the following equation using m" as defined by Schultz:
1.539
Note that the values of
"
u ?4 26 ?8
The isentropic temperature drop is then
for the
PV*
20
PRESSUFE, Pn
Fig. 3-Generalized compressibility function, y.
Fig. 2-Generalized compressibility function, X.
where Ps
08 t0 t2 tt t6 t0
02 04 06
-
l(RT ",0 Z,,s) /
77
8l ln(P, /
Pz) t
(t e)
This expression should only be used in small incremental steps. The temperature and compressibility used in Equation (19) are average values foi the increment. Equation (14) should be used to check the assumed temperature drop for the increment.
An equation for estimating the actual terminal
tem-
perature may be developed by examination of the geometry of Fig. 1. From this figure, it can be seen that
T,
-
T6
+
lW,
(l - n)l/t,*,
(20)
As with the polytropic calculation, the path may be broken up into small increments if it is thought that the fluid properties rnay vaty significantly between inlet
T" may, therefore, be estimated by Equation (20) and the previously calculated value of Tt. In Equation (20), cpaas reptesents the average specific heat between ft and 7".
(PrlPr) r giving incremental work Wui.
Exomple. Calculate the work obtainable and the exit temperature for the expansion of propane from 400 psia, 30O:F to 100 psia using a turbine with an isentropic
and exit conditions, and the pressure ratio (P2lP) in Equation (12) becomes the ratio for the .i.ncrement,
of n", Z, co, and X will not change much in a small increment, Equation (13) can be Since the values
written: 68
efficiency of 0.75.
EOUIPfrIENI SETECTION
TURBOEXPANDERS RECOVER ENERGY
For propane i P": 617 .4 psia, T" :
Manufacturers should be given adequate data for the design and selection of a turboexpander and to assure the purchaser of comparable bids. Data should include the following items:
666.3oR and R :
1545lM.L:35.0 Assume
(7"-To):
Toono:760
:
75oR. Then:
75/2 :722.5"R,p,00:400
-
-
l. Ambient
300/2
250 psia
2.
Tno,o: 722.5/666.3: 1.083, Pno,o : 250/617.4: 0.405 From Figs. 2 and 3i Xo,o:0.+7, Yors: l.l+ For propane at 722.5oR,250 psia: cpo,s: 0.552 ko,o = 1.175 and h"oro
=
kooo/Yoro:
at initial
conditions:
e. Molecr.r-lar weight Ratio of specific heats co/cu g. Inlet weight and volume flow (include the range of expected variations as well as the design value) . h. A description of the amount and nature of any gas contaminants such as liquid droplets or dust
f.
The isentropic work obtained is calculated using Equation 12:
: (--roi-) :
tr,H*]
(, -
*
#,
particles.
3.
38.9 Btu/Ib.
_ 0.552(7" - Tb) | + o'47
from which (7"- Tt) = 103.5oF. Using this new temperature difference, the procedure is repeated. Only the results of the calculation will be shown here.
\il"=39.5Btu/,b.
(7"-
=
To)
:
10B.6oF
from which To=
760.0
651.40R
-
needed.
The actual work obtainable is then,
:
:29.6 Btu/lb. The actual temperature drop (I, - 7,) is calculated rl,W,
using Equation (20)
=
tions.
In addition to
furnishing vendors adequate design data, certain basic mechanical features should be specified for reliable operation. Such features should include the following:
l. A heavy-duty thrust bearing (preferably
108.6
This temperature difference is close to the assumed temperature difference and further refinement is not
W
Operating requirements
a. Driven equipment (type, speed, horsepower and direction of rotation). b. Speed control and governing requirements. c. Shaft sealing requirements including leakage limita-
Substituting into Equation 13:
aao
Gas conditions a. Inlet pressure
b. Inlet temperature c. Outlet pressure d. Inlet composition
1.03
From a compressibility chart, Zo:0.84
,,,
conditions-barometric pressure, tempera-
ture,
0.75 (39.5)
2. Bearing temperature sensors. 3. Vibration sensors. 4. Low lube-oil pressure trip. 5. A highly reliable overspeed trip device. NOMENCLATURE = specific heat at constant pressure, Btu/lb"F c. = specific heat at constant volume, Btu,/lb'F 6e
:
T":65t.4 +
[39.5 (1 -,75)]/0.502 671.1 oR = 211.1 oF
I{ - enthalpy, Btu,/lb
=
,?,
f = absolute pressure, psia = critical pressure, psia P,'= mixture pseudo critical pressure, Pr = reduced pressure p/P" ot P/PJ P"
T" = critical temperature, "R 7"'= mixture pseudo critical temperature, oR Tn = reduced temperature, T/7" or T/7.'
7 = specific volume, cf/lb
= actual work done by gas, BtMb W, = polytropic work, Btu/lb l1z" = iseritropic work, Btq/lb X = compressibility function, Tn (62/8Tn) Y = compressibility function,l -l(pR/Z) (62/6pn)Tn) Z = compressibility factor, PV/RT ?, = isentropic efficiency ?p = polytropic efficiency . = increment
V. H. An.tnrs is on eng'ineeq'with Broun & Root, Inc., Houston. His work inuolues project nxo,nager/Lerlt. Mr. Abadie
W
holds a B,S. d,egree ,i,n mechanical engineering from The Uninersity of Te*as
at Austin and an M.S. degree in mechanical engi,neering from The Unioet'sitg of Houston. He is a member of the American Soci,etg of Mechoni,cal
REPRINTED FROM HYDROCARBON PROCESSINU
psi4
O = heat transferred to or from the gas R = individual gas constant = !t{g/mol-wt, ft-lb,4b'R ,S = entropy, Btu,/Ib 7= absolute temperature, oR
About the qulhor
soci,ation,
= polytropic temperature exponent
- polytroplc volume exponent ,n" = rsentroprc temperature exponent z, =isentropic volume exponent = k/Y P: absolute pressure, psfa 2
From a Mollier chart for propane, the corresponding figures are: \il, = 38.5 Btu/ib, 7u: 19BoF, T"= 2l6op. Breaking the expansion path into three increments each having equal pressure ratios will give the following results u,hich are in slightly better agreement with the Mollier values: W":39.35 Btu/Ib., Tb: 197.24 oF, T": 215.12 "F.
Engineet's and the American Gas As-
of the Kings-
bury type).
I
LITERATURE CITED
M., "Th" Pg-lylropic -{nalysi -of ASME, Jan., Apr., 1962, pp 69 and 222.
Schulq,_-J_._
Trans.
Centrifugal Cmpresms,"
Abadie, V. H,, "Piedicting Expmsion Tubine pa;rcr number 69-WA/PID-19, Nw. 1969.
3
Pufmmce."
ASME
I
69
Notes
70
Turbomachinery Problems
and Case Histories
.-t ti'#
E..
iI i;:i "
-
Case histories
of specialized turbomachinery problems
Wolier von Nimitz, t. C. Wochel and F. R. Szenqsi, Southwest Research Institute, San Antonio, Texas
INvnsucerroN oF sERrous vibration or failure problems in rotating equipment has many aspects of the classic
murder mystery: "Who did it and why?" The engineer uses similar investigating techniques to find the "culpr.it." Since many problems cannot be readily diagnosed rvith typical monitoring equipment, he must use specialized equipment such as real time analyzers, high frequencv accelerometers, pressure probes, strain gages, etc., to gain additional clues. Once the cause has been defined, the question of "u-hy" must be ans."vered. The normal causes are usuallv related to design, installation or operation. Some failures are due to design deficiencies which indicate use of inadequate mathematical models which cannot take into account a11 system variables. Three case histories illustrate
the instrumentation and
analysis techniques that can be used to detern-rine the causes of r,,ibration and failure problems. The first case concerns a steam turbine that had repeated blade failures. When blades failures occur, the1, are normally near the root and associated with fatigue, most probably at the blade natural frequency. The manufacturer usually increases the cross-section near the root to reduce the stress concentration facto.rs, lvhich is usuallladequate as past successes verif,v. If this method does not work, obtaining meaningful data becomes increasinglv
There is no such thing as ordinary ot routine machinery problems but some are
more unusual than others. Here are some case histories that illustrate what to do if you are laced with an uncommon maltunction
difficult because measurements other than blade strain gaging only infer ',l'hat the blade is doing. To properlr, measure the blade response requires a tremendous effort, both in time and money. especiallv considering dou.ntime costs in some process applications. The ideal technique
would be to make external measurernents u.hich could be related to blade response. In this example, blade response was monitored by measuring the bearing housing vibrations with accelerometers. Using a real time spectrum analyzer and other automatic analyzing equipment to obtain continous spectral display versus speed produced a version of the Campbell diagram with the amplitudes of
vibration superimposed. The second case discusses the field balancing of a unit which had a bearing housing support resonance. The unit had several available balance planes and sufficient data was taken to develop influence coefficients for four of the
of the discussion is centered around the inconsistencies that can be obtained planes. The primary emphasis
tz
SPECIALIZED TURBOMACHINERY PROBLEMS
using vibration amplitude and phase data where nonlinearities appear in the support structure. The third case describes the excitation of a compressor
shaft at its'fourth critical r,vhich r,r,as 15 percent above running speed. The shaft was excited by an acoustical resonance jn the suction piping caused by variation of the inlet guide vane angles. CASE
l: BIADE FAIIURES tN
STEAM TURBTNE
A thret stLrse. l.lJ00-hp ste1lr tur.binc clri'"'i1.g a (.cntr-ifugal air cornpr.essor in a catall,tic cracliing unit had a historl of viltratiorr 1;roblenrs. incltrd.ing cxcessir.c shaft vibrrrtirrns- g()\rernol gcar r.ibr-ation problcrns ancl [ailur.es oI tht- llrst.rnd third stagr bladcs ancl shroucl barrrls. A measured horizontal response
at the inboard end of
the
gential modes of the first and third stage blades in the normal running speed range (3,850-4,800 rpm). Vibration signature data rvas obtained on the turbine
FREOUENCY, HZ
Fig. 2-Outboard bearing housing
v ib r at io
showing nozzle excitation frequencies.
n
versus speed
excitation of rnany blade frequencies rvhen passing through the critical speed at 4,600 rpm. When a unit runs on a critical speed, it can excite blades at their own natural frequencies.l Blade natural frequency components were also excited at speeds well below the critical speed. Although it can be argued that the response at these frequencies on the bearing housing may not indicate that the
\ ata presentation technique reinforces the I it is blade natural frequencies since the sa is excited by other multiples of the blade p cy. Another interesting phenomenon occurs near some of the calculated blade natural frequencies; an increase in the.natu,ral frequency is observed as a function of speed, until the speed reaches a certain level and then the frequencv remains the sarne. This increase in blade natural blades are
and tltird stage blacles. \\rhcrr rhe c.rnrPonert at 50 )(. N
Irasses throrrqh thc ll.1(10-+.300 TI;, rrtnse. l:rrse ,,.ibretion componcnts are ileasured on the bearing housir.rg. \\rhen the 100 X N and 150 X N conrponents pass thr.ough this frequency range, they also excite these modes. This data presentation method is significant because apparently neg_
ligible responses in the vibration signatures ,"-, "ui 1r"i_ tinent inforruation. Fig. 3 is a si:ectral analysis (0-1,000 IJz) shorvinq the
frequencv could indicate the effect
of centrifugal
force
on changing blade end fixity. Unequal pressure distribution around the nozzle periphery caused partial adn'rission effects, as shown in Fig. 4,
resulting in larger amplitudes at the lor"e. harmorrics. This proved to be a major source of energy for exciting
the blade vibrations. TANGENTTAL
M0DE NUMBER
s=
F 3E
(T)
AXTAL (A)
i/^ .tr- -sno lsr STAGE (1)
srAGE (3) oR
f; E* s"PE =BFE
E
E
5000
o=
e
6;4000
T! U
fi
gooo
(9
z.
7zw
.E =
1000
FRE()UENCY.Hz
FBEOUENCY. Hz
Fig. l-Campbell diagram. REPRINTED FROM HYDROCARBON PROCESSING
Fig. 3-Excitation
function of speed.
of lower blade
natural f/equencies as a
CASE
2: MUTTIPLANE
BAIANCING
LOWER TMCEUNEOUAT PRESSTJRE
OF STEAM TURBINE lJpon startup of a steam turbine in a chemical plant, difficulties were encountered in the field balancing due to a resonance associated with the inboard bearing housing support structure. The foll'owing data were monitored to determine the vibration characteristics: proximity
0tsTBlBUTlOl,l0N
Fig. 4-Comparison of vibration for partial and full admission steam inlet.
of a balancing digital comthe authors, 'r''"'as used to by developed puter program determine the required balance correction *'eights. This computer program uses the least squares mathematical technique similar to those of Goodman' and Lund.3 The influence coefficient data is obtained by locating an LrIlmerous vector plots and use
A.
balance on each plane individually and obtaining the
detailed analysis of all balance data, including nu-
l-lqlqn6s
Af,OUND NOTZLE BLOGG
BLOCI(S
FREOUEI,ICY, HZ
resonance.
TABTE
NOzzU
INLET
plobes-vertical and horizontal, bearing housing-vertical and horizontal, and shaft absolute-veritical and horizontal. This preponderance of data quickly leads the engineers to ask which measurements are best to use in the balancing. Simple balancing theory indicates that any two measurements in a plane can be used to balance a rotor, since they will define the high spot. If the system is linear and the rotor and its supporting structure have no resonances at or near the selected balance speeds, theory indicates that any combination of probes should give the same balance solution. In actual practice inconsistencies in data commonly occur. Arbitrary selection of balance planes may not lead to a satisfactory balance condition unless extensive trial and error techniques are pursued. Since a resonance of the bearing housing support structure was verified by a shaker test, all of the listed data points were carefully measured to obtain an accurate and complete record. The bearing housing suPPort structure was braced in an attempt to eliminate the effect of the
A
UPPER TRACE-
AFTEB EOUALIZINC STEAM INLET FLOW
for combinqtions of
corrections
plones
Single plane belauce solutiors Balance Plane 3
Balance Plane 4
1
4 5 6 7 8 0 10
Max.
No. oI
Run no.
Aogle
speeds
Proximity vertical Proximity horizontal Proxim. yert. & bor. Shaft absolute, Yert. Shaft absolute, hor. Shaft abs. vert. & hor. Prcxim. vert. & ]ror.. Shaft abs. vert. & hor. . . Proxim. vert. & hor. . Sbaft abs. yert. & hor. , .
co 83
55 110 109
67
7t 67 112
resid. 21
5 358 8 4 345 353 359 346 4 354
54 56 89 85
4t 62
90
Lo wes t
I Angle I
resid.
18 353 3 344 340 341 I I I 327 I 4 I 343 I
1.17 0.93 1.18 1.81 1.69 1.84 0.87 1.36 1.19 1.80
predicted
Mr*. Angle
f31 .{i,r, 81
1.16
I
115
13 16 48
120 208 130 55 268 65
0-86
28 12
r7{
32
67
123 281
31 20 118 m I 17
1.22 2.16 2.42
0.90 1.06 2.03 1.66 2.95 0.91
90 51 80 {6 85
178
residual
BPl BPl BP1
BPI BP1 BP1
BPl BP] BPl BPl
B. Two plane balance solutions BalaocePlanes3&2
BalancePlanes4&2 Bal. Plane Run no.
w-,I
Type of data
1I
Proxioity vertical
77
12 13
Proximity horizontal
98
r4
Shalt absolute, vert. Shalt absolute, hor., Shaft abs- vert. & hor-
15
l6
Proxim vert &hor..
90 119 112 135
4
l,gl" w. I A,El" 316 357 358 356 357
352
-B"t. Pl"""
Bal. Plane 2
48 18 34 41 42 39
Max resid
248
Wt. Angle \Yt, ti8 88 79
255
t01 t36 t16
259 318
3 -B"l-PI""" ,
330 350 318 341 346
Aogle
resid.
Lowe st predicted residuel
Max.
24t
0.85
BP.l &
31 17 51 61
261
BPl&3
266
0 81 101 136
269
1,15
18
288
159
252
BP]&2
BPT&2
BPl&2
C. Three plane balance solutions Bal. Plane Run no.
t7 18 19
20
2t 22
74
4
I
Bal. Plene 2
Bal. Plane 3
w" I
Type oI data
Proximity vertical PruxiutlJ lorrzontal Proxim. vert. & hor. Shaft absolute, vert. Shaft abmlute, hor. Shaft abs. vert. & hor
380
17
359
264
12
96 152 98
355 339
300 150 187
201 178 198
t7 84 46
113 27
A"ct"
42 14
153 107
29 38 58 38
179 269
-
NIax. resid.
038 031 016 0.67
070 296
3
BPI&3
1.07
tions can be traced back to nonlinear or resonant effects
SPECIALIZED TURBOMACHINERY PROBLEMS
in the vibration data (amplitude and phase angle) near each bearing. All possible balancing combinations are calculated based on selected test points and speeds. Optimizing criteria are applied to determine minimum expected vibration, minimum added weight, maximum reduction per added weight, and minimum sum of squares vibration at all test points. Since this leads to nume.rous solutions, the engineer must still use his judgment to determine which is best. Theoretically, the best balance will be obtained by adding weights to all available balancing planes. Ifowever, in field balancing, keeping the balance planes to a minimum is desirable since considerable effort and downtime are required to change weights in all balance planes. A "perfect" balance cannot be obtained nor is it necessary, because only a satisfactory balance within acceptable vibration criteria is needed. Considerable time can be saved if a satisfactory balance can be obtained using only one or two balance shots and a minimum of balance planes. Table 1 consists of 40 single-plane balance solutions, 18 two-plane solutions, and 6 three-plane solutions. Using the computer makes it possible to compare all of these calculations.
A study of the data in Table 1 shows that:
o The balance solutions for the different probes are not consistent. For example, for the single plane balance solution, the ratio of maximum to minimum weight required was l48l17, 4819, 90124, and 712155 for balance planes (BP) 1-4 respectively. The deviation in angle was 57,2L3,51 and 23 degrees. o If the comparison is made for only the cases where four speeds and four probes were used (Runs 3 and 6), the difference between solutions based on the proximity probes and the absolute shaft vectors become closer. Statisticalln this means that more consistent data will be obtained if more probes and speeds are used.
BPl BP2 BP3 Max./min. weight (gms) Angle
difference
data.
o For the three-plane balance cases, considerable difin the predicted solutions. If only the proximity probe data is considered and the influence ferences are noted
for balance planes 2, 3 and 4, then the three plane solution would be as given in Run 19. Imagine the anguish and indecision that an engineer would have if the computer told him to put approximately one-half pound on BP4. When large weight solutions are predicted, the next balance plane will coefficients determined from the data
normally have a large weight diametrically opposite, i.e., Run 19 has 264 gms at 12 degrees for BP4 and 187 gms at 198 degrees for BP3, or 186 degrees away.
o The detailed analysis of the balance data indicated that the best single plane balance could be obtained at BP4. All 10 solutions are plotted in Fig. 5. Since only a 23-degree deviation was calculated, the location was straightforward, and a weight was selected which was the average of the predicted amplitudes. This first attempt showed considerable improvement; and the second trim balance further reduced residual unbalance. The trim balance reduced the vibrational amplitudes through the speed range from 5,700 to 7,200 rpm to less than 1 mil peak-to-peak. A satisfactory balance of the turbine was attained with one balance plane (plane 4) . o A further examination of Table I shows that a better balance could be attained by using BP2 with BP4. Fig. 5 indicates the balance weight region for the second balance plane which is centered at approximately 270 degrees, and a weight magnitude of approximately 18-48 grams required. This was not attempted in the field due to limited time, as startup was imminent and the single plane balance was adequate. BP2 would have been the logical choice for the next balance shot. Ifowever. the vibrations
BP4
33/31 23113 85137 109/55
44o 1510 21"
. For the single plane balance,
15o
BP4 had the lowest
expected residual for all cases (except Runs 3 and 9) and the smallest angle of deviation. Based on these criteria, BP4 appears to be the optimum plane for a single plane balance.
.
For a minimum added weight criterion BP2 is best, based on the data in Run 3 indicating 13 gms at 2O8 degrees attains the same results as 55 gms in BP4. However, due to the inconsistent phase angle and the 5:1 ratio betr,r,een maximum to minimum weight predictions, the solution cannot be trusted.
o Solutions such as giverl in Runs 11, 12 and 13 for BP3 and BP4 where exceedingly large weights are predicted is one problem that is common when using least squares computer analyses. Usually, large weight soluREPRINTED FROM HYDROCARBON PROCESSING
Fig, S-Balance weight corrections using different probes
75
=
iCUfllH
CIq
this, high inlet guide vane settings were necessary to control pressure rise. The critical speed map for the compressor is given in Fig. 6 and the mode shapes for the first four criticals are given in Fig. 7. The unit runs between the third and fourth criticals. It can be seen from the mode shape that the anti-nodes for the fourth critical speed are near the two impellers.
q-
-slJJl'
G J o
t
i
I
E E O 4 E
I
I
I
snrF_b,
I
NOT|CE THAI 252 Hz 252 Hz_4It CRIT;CAL 5p569 C0tap0NENT C0MES tN
I l0l
1U
B[ABIiiIO SIIFFNESS. LBJIII
nunrurruc
It
tl^,iETrrNG
rrs-
60'
-u' FREOUENCY, Hz
Fig. 6-Critical speed map
Fig. 8-Compressor shaft vibrations versus guide vane setting
were already within specified limits. CASE
3:
ON
ACOUSTICAT EXCITATION SHAFT VIBRATIONS
A gas turbine driven compressor unit had several reoccurring problems, including (1) slipping of compressor coupling on shaft, (2) bearing failures in turbine, (3) rotation of diaphragm, (4) failure of inlet guide vanes and (5) excessive noise levels. The unit ran from 13,000 to 14,200 rpm. The compressors did not operate on their original design point due to changes in plant operating characteristics after the units were pu.rchased. Because of
N0TICE THAT 252 Hz
_h
COMPONENTCOMES IN G V SETTING IS
-AS
FREOIJENCY, HZ
Fig. 9-Compressor suction pulsation versus guide vane
A field study rvas conducted to measure the vibration, pulsations and noise of these units. Pigs. 8 and 9 sr-rnmarize the significant results. At certain inlet guide r-ane settings, the shaft vibrations drastically increased and the
noise in the compressor and piping also increased. Shaft vibrations and pulsations were recorded simultaneouslr.. Fig. 8 is the spectral anal1.sis of the shaft vibrations versus the guide vane settings. The unit speed was held constant
and the guide vane settings were gradually changed. Each vertical step represents about 5 seconds. At the lower guide vane settings, the predominant frequency was at the running speed. When the guide vane setting was increased above 45 degrees, a 1 mil vibration component at 252 Hz (15120 cpm) suddenly appeared matching the calcu-
Fig. 7-Vibration mode shape
76
lated fourth critical. The amplitude is as large as the running speed component. Suction pulsation recorded during this same period (Fie. 9) revealed the cause of those nonsynchronous vibrations. Very little pulsation rvas present at the compressor running speed (220 Hz) or at
SPECIALIZED TURBOMACHINERY PROBI.EMS
252 Hz. Ilowever, when the guide vane setting was increased above 45 degrees, an acoustical resonance of 110 psi peak-to-peak was excited at 252 Hz. This acoustical resonance caused the shaft to vibrate at its fourth critical
which was above running speed. The iinpeller was at an anti-node which helped the energy to couple into the lateral shaft vibrations.
About the qulhors
W. voN NrMrrz is assistant directo,t' of the Department of Applied PhEsics of Soutlttoest Reseac"ch Institute
in
San
Antorio, Tera,s. He is a 7950 graduate of the Technical Uniaersitg of Munich utith an M.S. degree in electrical engi-
neering. Since joining Southwest Research Institute in 1957, Mr. Nimitz has
superaised the deuelopment o,nd operation of the SGA Compresso,t' Installation Design Labora,tory and the deaelopment of the SGA onalog in particular. His major field of erperience is natural gas, petrochem;ical and chemim'l plant design and etsaluo,ti,on engineering ttith, emphasis on pulsation and oibration control methods and techriques, design of acoustic filters, compressor efficiencg and noise control studies, flou; nxed,surenxent problems and design problems associated utith unsteady fluid, flow in general. He has superuised the design studies of oaer 2,000 compressor and, pu.m,p installo,tions for more tho;n 3oo conlpanies th,t'oughout the world and has conducted seueral hundred fi.eld eualuation studies and consulti,ng engineering sem.tices for the industry. Mr. Ninritz is the author of numerous published papers o,nd articles, is a member of the Americaru Society of Mechanical Engineers, and
is listed in "'American Men of
Science."
J. C. WACHET, holds an M.S.M.E. degree from tlte Uniaersity ol Tenoe. He lws been
uith
Southutest Research Institute
since 19ti1. His oiliaitics luoe centered, in the fields of vib,ratians, pulaotions, dynamic simulations and acoustics, with
particular emplnsis on uibration and ailure problems in turbomachinerg. His
f
ertensiue fwld experience has led to the deuelnpmmt of impro'ueil experimental tech,niquea for comelating oibration signatures of rotating equipment to potentio,l problems. He also has been responsible for the deoelopment of computer prog,r"ams uthich- are used to solt;e lateral and, tot,sional critical speed problems, as utell as other luilure problems in turboma,cfuinerg. He is a member of Tau Beta Pi ond Pi Tau Sigma.
F. R. SzeN,tsr joined Southutest Research Institute in 7965 as a member of the Department of Applied
Physi.cs uhere his utork has included both theo-
I
analysis of the fi,elds of
s and, acoustics. His usork experience in mechanics
includes the analgsis of lateral and tot-
sional uib,t'ation response of mechonical sEstems, prediction of uibrational displa,cement, stress and, methods of failure deteetion. He graduated from The (Jni-
Since the excessive vibrations were found to relate to the high negative guide vane settings, avoiding these operating conditions eliminated the problems. The compressor wheels in these units were changed to more properly match plant pressure and flow requirements.
coNclustoNs The three case histories of vibration problems illustrate the problems that can occur when available analytical solutions are not sufficient to accurately take into account the system variables in the design stage. Based on data obtained and the analysis of these problems, certain conclusions can be made:
(1) By measuring vibrations on the bearing housing of a turbine, vibrating blades can be verified. Real time spectrum analyzers and automatic data monitoring instrumentation used to display vibration signatures as a function of speed like a Campbell diagram make blade vibrations mo.re obvious.
(2) Bearing housing support or rotor resonance near the running speed increases the difficulty of rotor balancing, because the vibration amplitude and phase data generally have inconsistencies which hinder the solution for the balance weights. Nonlinear eflects in a system also complicate the solution for the balance weights. (3) When complete vibration data is taken, including relative shaft vibration, bearing housing vibrations, absolute shaft vibrations, plus shaft orbits, many probe conibinations theoretically can be used for the balancing procedure. The arbitrary choice of only a few probe positions, as usually done, can lead to uruealistic solutions and can result in costly trial and error balance procedures. (4) A digital used
computer program was developed and
for multiplane balancing for any combination
of
balance planes, test speeds, and any number of vibration measurement points. This program calculates all possible balancing combinations, which the engineer must analyze and compare to determine the best balancing procedure. Several optimization schemes have been programed to help eliminate the inconsistencies caused by nonlinearities and by resonance in the rotor and support structu.re.
(5) The evaluation of all possible sets of vibration data leads to a zor,e of correction weights and angles for particular balance planes. The engineer then must decide on the amount of weight and angle. (6) Nonsynchronous shaft vibrations can be excited by acoustical resonances in the piping system. This illustrates the desirability of performing detailed spectral analysis of vibration and pulsation data so that the exact
cause
can be established. ACKNOWLEDGMENT
originally Teru A&M Univcnity, gralclully rknow-lcdge I This paper
Elrmcntatron lor
omachinery Symposiu,
ber,
,1973.
ifhe
authom
who helped develop ttre
gen
LITERATURE CITED
uersity of Colorad,o with an M.S,M.E. degree.
REPRINTED FROM HYDROCARBON PROCESSING
77
Compressor problems: Gauses and cures
Most compressor problems can be prevented with a good purchaser
in a lathe, The If a coin could job was accept
,,n
,f; ical
specification based on lield experience. Here are some experiences and solutions John J. Dwyer, Air Products & Chemicals, fnc., Allentown, Pa. Successrur, coMpRESSoR o,peration not only
depends
on an experienced and capable maintenance team but to a greater exte specifications. Too often the written by the contractor or e ApI. Extramaintenance.
Dynomic boloncing of rotors is now a routine procedure for almost all manufacturers but there are still occasional problems. A fer,r, years ago we had a small rotor dynamically balanced, shipped to the field, assembled and started up. We immediately lost a set of bearings and damaged the rotor. The rotor vu'as repaired and rebalanced but the Some compressor problems are unique to a particular location or set of conditions. An example is the time an operator ran headfirst, with his safety helmet, into a lube oil pressure switch causing the compressor to trip and shut down the entire plant. But most problems are caused by design features that can be corrected in future installations.
By far the most important type compressor installed in the process industry today is the centrifugal machine. It is the product of more than 40 years of industrial bxperience and development. Most prdblems in earlier years resulted
from poor balancing and high vibration. It was not unusual for a repair shop to statically balance the rotor 78
same failure happened. We then discovered ancing machine was not properly calibrated.
that the bal-
As a result of these and similar problems we have improved our compressor specifications. Today we require all rotors to be dynamically balanced with not more than two impellers added to the rotor between each balancing step, and require the 'balancing machine to be calibrated at the conclusion of the operation. Vibration specs. We specify the amount of residual unbalance'that we will accept. We also require proximity probes at each radial bearing and require the surface under the probe to be finished to the same standard as the bearing journal. Max,imum shaft vibration limits at operating speed are specified and filtered vibration ampli-
COMPRESSOR PROBLEMS: CAUSES AND CURES
tude at frequencies other than operating speed are limited to a small percentage of over-all amplitude. In critical services, two probes at 90 degrees are required at each bearing so an orbit analysis may be made on an oscilloscope. On critical high-speed units we insist on tilting-pad bearings to eliminate the possibility of oil whip.
Beoring fqilures. Although many problems associated with bearings can be analyzed and prevented, it is almost impossible to eliminate human error. Several years ago at one of our plants a large centrifugal compressor was turning backwards at trow speed, without lubrication, due to a check valve leak. The bearings were removed, inspected and were s plvoted shoe thrls steel retainer rode
n reassembly one backward so the ar. The machine
was started and ran satisfactorily for about a year but subsequently the failure did occur, at night, on a weekend.
away from operating speed. On new designs or critical installations, at times, we have employed consultants to analyze the lateral and torsional characteristics of the complete train including the instantaneous torque iharacteristics
of the motor.
High-speed couplings have been the subject of much discussion over the past several years. Problems associated
with flexible coupling lock-up, where they transmit axial forces, have resulted in many purchasers requiring the compressor and driver vendor to design thrust bearings to accommodate this possibiiity. In addition, minute dust particles in the lubricant have blocked oil passages by separating-out through centrifugal force within the coupling. One widely used method of minimizing dust problems, especially on large compressors, is to install filters (one micron or less) in the coupling oil rfeed line. One major vendor completely eliminates the flexible coupling problem by using solid 'couplings. flowever, extra precautions must be exercised when using this approach to assure near perfect hot alignment and sufficient axial clearances on all labyrinth seals, especially in long compressor trains. Another promising solution to the coupling lock-up and dirt problem is the fleible diaphragm coupling. For the present, we are specifying that compressor thrust bearings be designed to accept some axially transmitted force and require sludge drain holes in gear-type couplings.
duct. By this method we have not only eliminated many fi.eld problems but have also obtained in-terchangeability of filier cartridges, control valves and switches which reduces our sPare parts inventory.
lqlerql qnd torsionol problems. Most
centrifugal
machine, and the first lateral critical is normally checked on the manufacturer's test stand. Torsional problems are prevalent with synchronous motor drivers since the instaniur,"ort torque outPut varies considerably during motor acceleration with an exciting frequency starting at 720
hertz and diminishing to zero at synchronous speed. We have tried to eliminate lateral and torsional problems by specifying the acceptable range of lateral critical speeds ai a percent of operating speed, and requesting that torsionai nodes of resonan'ce be a given Percentage margin
Specify lateral critical speeds as a percent of operating speed. Require torsional nodes of resonance he away from operating speed REPRINTED FROM HYDROCARBON PROCESSING
Lobyrinth seqls. The design and fitting of interstage and casing seals is an irn:portant part of over-all design and an area that causes its share of problems. Interlocking labyrinths where both the rotating and stationary members have labyrinth teeth are superior to a single labyrinth in reducing leakage. However, axial clearances and float are critical. In one design it was possible to install the stationary part backwards which resul'ted in almost lineto-line contact of the labyrinth teeth. Since this type of assembly error was possible, it naturally occurred, damaging the balance drum on startuP. We have since modified our specifications to require the stationary parts of interlocking labyrinths to be designed with an off-center key. It is easy to improve compressor efficiency by minimizing interstage labyrinth clearances. The new compressor will operate with minimum leakage, but when the rotor becomes slightly unbalanced or vibrations occur from an-
other source, the seals will wear rapidly, opening the clearance and allowing greater leakage with a commensurate decrease in efficiency. To prevent this situation, we specify that the minimum clearance shall be equal to the bearing clearance plus rotor run-out at the seal due to its weight plus a given amount of mass unbalance.
Mqteriqts of conslruction. The selection of materials of construction is left almost solely to the purchaser when corrosive gases are to be compressed because he frequently
has more experience with the gas than the machinery manufacturer. For most gases, the casing and diaphragm
79
stainless impellers and aluminum labyrinths because of
On moist gas compressors,
When discussing materials we should not overlook the inter and aftercooler. Several yeani ago we ran into a series of failures on cooler tubes in air serwice. After a
install a large low-velocity suction drum with a moisture separating pad ahead of the compressor inlet
corrosive fumes
in the atmosphere.
flloisture s€porotors. No
discussion
o[ centrifugal com-
until a large section of deposit breaks oflthe wheel, causes un'balance, a sud,den rise'in vibration and an unscheduled shutdown. We have been insisting on high-efficiency moisture separators; and ,on moist gas process machines we install a large low-velocity suction drum with a moistureseparating pad immediately before ihe compressor inlet.
Reciprocoting compressors. Most reciprocating compressor failures can he related to high vibration, pulsation, lubrication and wear. Analyzing each type of failure and placing it into a general category assists in developing a cule.
Vibration switches. Several years ago we experienced a number of different types of problems all resulting in a major failure. A seized crosshead .caused by insufficient
resonator in the line. stalled with a tee in
is to know the puls
resonator frequencv quency.
t
in_
rick the fre_
To minimize pulsation problems an in the field u'e specify a maxim and discharge pulsation level. fn a installation of pipe taps in the inlet a all pulsation snubbers so \^r'e may easily check the actual pulsation levels in the field with a pr.rr,rr. pick-up and checks
oscilloscope.
clearance
breakage broken Although A
:f": rod. and
several other problems were all different, they had one in common; if the compressor was shut down immediately, damage would be minor. We equipped all of our large balanced-opposed compressors with aLieleration type vibration switches. At first we had some dificulty with spurious shutdowns, but the problem was eliminated after. rve learned proper switch positioning and setting. I4Ie had trvo problems shortlv after installing vibratioln snitches on a 2,500-hp five-stage machine. Oncc lve lost
factor
80
Lubricqnts. Air compressor cylinder Iubrication and com-
COMPRESSOR PROBLEMS: CAUSES AND CURES
one problem can be the cause of another. We found that the synthetic lubricant is an excelient paint remover.
I-ubricant that traveled along the rod entered the crank case and the crank case paint began to soften and peel. As a temporary rneasure we installed an extra set of rod scraper rings to reduce the lubricant carryover into the crank case. and monitored the percentage of synthetic Iubricant in the crank case oil. When the peroentage of sl,nthetic got too high we would simply drain a barrel of the mixture and add a barrel of fresh lubricant during operation. We non, specify ail reciprocating compressors lvhere we might want to use synthetic cylinder lubricants, to have special lubricator seals and a coating of epoxy base paint inside the crank case. Lubricators. Cylinder lubrication is another potential problem area. Some cylinders had onlv one lubrication point in the bore. A cylinder failure would be caused by a failure of the pump feeding that point or a line ieak. We partially corrected this problem on some units by installing a feed point in the cylinder inlet pipe. Although this does not result in an even lubricant distribution it can prevent a major failure. The lubricator itself mar' also give rise to a problem. If the operator neglects to filI the lubricator, failure r,r,ill result when the resen'oir runs dn'. This type of failure could be protected bv a reservoir level switch, but the level switch is of no help if there is a failure in the lubricator drive mechanism.
Another method of protection is to install pressure in each lubricator feed line, but on large units rvith 20 or rlore lubricator points this rneans 20 pressure switches, alarm Iights and high installation costs. We use a short suction pump and pressure switch mounted in the end irosition of the lubricator reservoir opposite the driver. If the lubricant level runs low or the drive mechanism fails, this one pumping unit will activate an alann. srvitcires
Pocking. Breaking in a new set of high Pressure packing used to be an experience capable of causing the operator and mechanics no end of grief. One method u'as to increase the pressure slorvly over many hours. If the packing started to smoke, the pressure was reduced. What no one r.vanted to admit \vas that iI the packing started to smoke it rvas probabiy too late. On one particularly troublesome installation, after instailing the packing rve removed one of the packing case tie rods and installed a therniocouple adjacent to the first seal ring. '{'he compressor was gradually brought up to pressure by monitoring the packing About lhe qulhor JonN J. Dwtun is mano,ger of ma-
ch;i,nerg engi,neering, Air Products & Chemicals, Inc. Allentown, Pa. He 'is ,rasponsi,ble for cost estimntes, prepq,r-
ing put'chase speci,fi,cations and eoaluations of all machinerll procured and furnished bg the compana for process appl;icati,on, Mr" Dwger holds a B.S. degree in mechanical eng'ineet"ing f rom Newark College of Engi,neet"ing and an M.B.A. degree lrom Lehigh Uni,aersitg. He is a member of ASME the Acoustical Society of America and, i,s a regi,ster.ed professional engineer in Pennsgluaruia.
REPRINTED FROM HYDROCARBON PROCESSING
In a later installation, we converted to a lubricated-filled Teflon breaker and seal-ring design and built up pressure at a much greater rate. Our specifications for high pressure packing now require cooled packing cases and Teflon rings. case temperature.
Nonlube compressors. Some operators have reported fantastic life from what they term nonlube service, but investigation shows that they do lubricate the packing or have lubricant carryover from the crankcase. In other nonlube compressors the compressed gas was wet. Even atmospheric air has some lubricant value from the moisture in the air. We have a number of reciprocating compressors in nonlube service handling dry nitrogen. The compressors are fitted with extended-distance pieces with flingers mounted on the rod to prevent oil carryover. At first, we found a great deal of variance in s,ervice life between different units. We have been able to correlate the service life with what we term a PV factor; that is, the mean cylinder pressure multiplied by the average piston velocity. Although this correlation is not precise, we are able to predict with some accuracy the possibility of satisfactory operation. To improve nonlube operation we specify maximum rider-band loading, cylinder-bore finish, and equip all horizontal cylinders with rod-drop detectors to monitor ring wear in operation. In some designs, we have found excessive rider-band loading caused when gas pressure gets under the band and causes it to act like a seal ring. We have corrected this in the field by either drilling holes in the rider band to relieve the pressure
or by
segmenting
the ring.
Mqintenqnce. Not all problems associated with compr,essors can be attributed strictly to design factors. Preventive and major maintenance also play an important role in achieving high onstream factors. We have established a centralized control system for all compressor maintenance and a computerized preventive maintenance system. Each month a deck of computer cards indicating preventive maintenance jobs to be performed ar,e mailed to each plant. The majority can be accomplished with the compressor in operation, such as sampling lube oil for analysis, inspecting filters and testing alarm devices.
Major maintenance on centrifugal compressors is only scheduled alter a thorough analysis of operating parameters, We have developed computer programs to check actual compressor performance data against design values. We also analyze shaft vibration data and look for trends of increasing vibration or indications of bearing instability. If either performance or vibration analysis indicates deterioration, major maintenance is planned. Since most of our facilities are of a size that do not warrarrt a complete maintenance staff, major maintenance is handl,ed by specialists from our Central Maintenance sta.ff. By this method we can limit our local maintenance personnel to a minimum, yet provide the expertise for major overhauls.
The final and perhaps most important aspect of major compressor maintenance is that we do not open a centrifugal compressor on some hypothetical periodic schedule. It is opened only when analysis of performance or vibration indicates a need. Following this approach, we have consistently achieved an onstream factor above 98 percent. 81
Notes
82
Critical Related EquiPment
.-'*"
l't
'. tF-;.
Are couplings the weak link in rotating machinery systems? The answer is yes it the coupling is improperly applied or designed. Here are the facts about selectrng this critical component Ken Krqemer, Allis-Chalmers Compressors Division rnilwaukee, Wis.
Trro appr-rcerroN effort involving the signers. The user, by
can either help or hi experience level of his operations and maintenance people when selecting the basic coupling style.
n'ould be an unbalance force requiring a shaft u,eight of 17,000 pounds to insure that the unbalance force ivould be only 10 percent of the rotor weight. The dr bine rvith
pling rvas weight n' eccentricity was great. The redesign in this case was basically the incorporation of a pilot fit to insure concentricity of the coupling hubs and spacer. The user's problem was to undeistand the
caused trouble, the errors could be corrected and the cou-
ever, greater than this expense implies. This feedback, that
tempted to permit using the gear coupling configuration instead. The coupling was successfully applied, and is still in use today.
-
Had the eqqipment been modified to suit the dynamic
forces possible with the original coupling design, the driven shaft overhang would have had to be increased in diameter
to handle a potential force of 850 pounds, in addition to the weight of the coupling itself. More important, this 84
With both gear and disc couplings, the procedure of taking a predesigned catalog item and applying it without modifications to suit the adjacent componenls is being eliminated.
The coupling manufacturers' engineers are becoming
more willing and able to merge with the rotating equipment engineers in adjusting the couplin.q design after the
ARE COUPTINGS THE WEAK LINK?
compressor and driver have been designed for optimum performance. With this approach, consistent design success is being achieved. This leaves a burden on the user, because couplings are
and have lower tolerance to damage when handled by personnel unfamiliar with the specific design. Due to this feedback, efforts have been made to seek out a coupling which requires less care and experience in the field. This factor is more often considered today and has supported the development of the industrial flex-disc coupling. An exercise in this relatively new application technique will point this out and illustrate quickly a few of the reasons for the many recent applications of disc couplings. Iess standard
APPLICATION EXERCISE Driven rofor-3-stage centrifugal compressor with a rotor weight of 350 lbs., rated at 12,000 rpm, driven by an elecrric motor through a gear increaser.
Geor increoser-Double helical gear transrnitting 2,500 hp with a ratio of 6.675 to 1, with a pinion static weight of 100 lbs., and a dynamic weight (at rated load) of 2,500 lbs.
Molor-l,800 rpm
synchronous motor drive
with a 6,000-
lb. rotor. Acceleration time to speed is 12 seconds at noload conditions. It is immediately apparent that 12 seconds is not a very long time. If any rotating component has an unexpected radial residual force condition, in excess of the strength of the materials, the initial start may be the last one. Every design eflort must be made to eliminate the possibility of high radial forces. Since we are talking about couplings, let's stay with that component.
Starting with the most extreme conditions, the
to be coupled is a
compressor shaft-end
system
with a
static
weight of 175lbs. and a pinion static weight of 50 lbs. Because even at light loads the pinion bearings will see some dynamic load, a weight, or resistance, of 100 lbs. will be used for the pinion shaft end. Upon reaching speed the acceptable radial unbalance
Fig.
l-Gear
coupling-male teeth integral with hub.
Fig. 2-Gear coupling-male teeth integral with spool. REPRINTED FROM HYDROCARBON PROCESSING
r\
Fig. S-Schematic of gear used in coupling applications.
force of the coupling would tle 15 percent of driven and driver shaft resisance, or 26 and 15 lbs., respectively. This indicates the driven shaft-end half of the coupling should have been less than 0.1 in.-oz. and the pinion shaft-end half 0.06 in.-oz. residual imbalance, respectively. Quite a target for a gear coupling designer. If we assume for the moment that the driven shaft and the pinion have no inherent unbalance (a poor assumption), the exercise consists of selecting a coupling, suitable for the horsepower, which can be manufactured, and which can be selectively assembled and maintained to operate within these tolerances. Two types of gear couplings have been applied to this type of service. One has male teeth integral with the hub, Fig. 1, and the other has male teeth integral with the spool, Fig. 2. Both have a pilot incorporated into the male tooth form to support the loose member of the coupling in a
concentric manner at speed, Fig. 3.
Couplings carrying catalog ratings
of
4,500
hp
and
16,000 rpm, based on a standard shaft gap, were selected. One coupling catalog wouldn't give the speed rating until
the shaft gap was determined, a very good position. The standard sleeve and spacer weight of the first coupling is 19 lbs. The spool weight for the second type is 4 lbs. One-half of these weights is carried by the coupling hubs mounted on each shaft end. Assuming the coupling hubs are mounted without producing any additional eccentric unbalance (another presumption) an eccentricity of 0.0001 inch on one end of the first spacer would produce 0.015 in.-oz. of unbalance. The second coupling spool piece u,ould produce 0.0035 in.-oz. per 0.0001 inch eccentricity. These pilots, acting at speed can permit only 0.0004 inch eccentricity and 0.0015 inches respectively, if we presume the coupling components are perfectly balanced. By their catalog rating both couplings pass. The next step is to study the operating characteristics of the design to determine if, for instance, the coupling will have an active pilot at operating conditions. In the first coupling, the heat generated at the teeth flows diflerently into the shaft than it does through the sleeve to the surrounding air. The sleeve will heat up and expand more than the hub. This plus centrifugal force acting on the sleeve will cause it to grow radially as much as 0.003 to 0.004 inches more than the hub in proportion 85
to the horsepower being transmitted. Such a differential would permit an eccentricity of 0.002 inches, or 0.30 in.-oz. at the rated speed or three times the permissible amount. The second coupling also develops the same amount of heat, but the hollow bored spool will accept heat in a manner similar to the sleeves to the extent that no differential growth occurs. Therefore, only centrifugal force should be acting on this pilot, amounting to 0.001 radial growth permitting only 0.0005 inches eccentricity, or 0.02 in.-oz. of force. By these numbers the first coupling could not do the job, and just as clearly the second could. Note that I said could do the job. If this coupling can be expected to normally produce 0.02 in.-oz. and 0.06 is the limit we want, there isn't much left for error in the manufacture or assembly. The user of this coupling must understand this condition and arrange to component balance the half coupling on the shaft it is to be mounted on, so,as to eliminate any accumulated unbalance as a result of balance mandrel or shaft eccentricities, or errors in factory balance of the coupling itself. A balancing plane of tapped holes is needed for that purpose. In the field, where replacement is necessary without dismantling the units, field balance checks are almost mandatory.
There is another area of evaluation-sliding friction coefficient. This produces a resistance to the axial movement necessary as rotors heat and expand. For the same catalog size rating, both coupling types have about the same pitch line diameter, so the unit load on the teeth
Fig. 4-Schematic of a typical disc coupling.
eccentricities. Almost all transient factors can be accounted
for by the designer in this selection of the components used to fabricate the coupline, but in this case, he must know shaft characteristics before designing the discs. The same manufacturing errors that can cause eccentricity in gear couplings can occur in disc couplings, but these can be detected before operating the unit and the chances of them occurring are greatly minimized. The same assembly eccentricities can occur in the field. but the absence of dl,namic radial residuals in the assemblv
would be the same. This load and a selected friction coefficient result in a force which must be handled by the thrust bearings and, in this case, one helix of the double helix gear before the teeth will slide and the normal op-
IIEOLIGIBLE AXIAL
EXPAIHOI{
erating position is achieved. A reaction will occur during every thermal change the system goes through. This is a subject of more interest for the user than it may be for the designer, since the user must contend with tooth surface finishes other than new or as manufactured. This is the reason the words "a selected friction coefficient" was used.
This characteristic of the gear coupling therefore requires a serious study of the capabilities of the driven shaft thrust bearing and the loading service factors used in the gear design "in order to suit the coupling." Intensive design efforts and complete design changes have been made in recent years to gear couplings to diminish the sliding friction coefficient, with some success, but again the burden is on the user to understand and maintain ideal conditions to insure continuous production. Closely associated with this latter problem is the matter of lubrication required by gear couplings. A failure in this area results
in the infinite sliding friction factor-the
locked coupling.
This exercise could continue, for there is much more, but this is sufficient to illustrate the many characteristics users must understand to utilize gear couplings. Taking the same conditions, let's evaluate a disc coupling (Fig.4) . The same maximum permissible unbalance values e4ist, but in this coupling there are only the static pilot fits to consider. There are no changes due to operating temperatuie or centrifugal forces to cause operating 86
I
I
sennumrrc rrnusr eemrrc
Aarse sppmr Fig. S-Thermal movements of shaft and casing.
ARE COUPLINGS THE WEAK LINK?
permits more use to be made of the concentricity tolerances allowed. The one area the user must be aware of is the tolerance permitted for axial growth or axial growth transients shown in Fig. 5. This can be a more serious problem with disc couplings than with gear couplings on certain types
of units. A rather specific control is placed on the disc deflection range, and the equipment has to be adjusted axially to suit with more accuracy than with gear couplings. Sorne units are very difficult to move once they have been set-a gas expander, for instance. Where this is a problem, horvever, arrangements can be made to permit
repositioning the coupling by remachining the components for a one time per coupling fit-up, or by adding a spacer plate. The application of a disc coupling therefore almost completely eliminates the every generation re-education of coupling users in the intricacies of the design itself-once the coupling is designed, because in order to design it, the coupling characteristics must be determined and adjusted to suit the connecting rotors. Furthermore, there are no changes or wear taking place in the coupling characteristics
throughout the life of the unit. In the design stages, the effects of concentricity, alignment, overhung bending moment during mis-alignment, transient thermal grorvth and net growths are incorPorated. The resonant pattern of the design is then matched to the pattern of the connected rotors. In the manufacturing stages it is relatively easy to balance the assembly to lorv residuals as dictated by the shaft design, and the mechanical fits are simple turns, bores and faces. As with gear couplings, balancing is done on a mandrel, so potentially some eccentricity can occur when the cou-
pling is mounted on the shafts in the field. Since the coupling flex members are rigid radially, field balancing is not complicated by dynamic changes to the pilot fits. One of the main features of the disc coupling is its known axial deflection load value. When compared to the varying unknown sliding friction factor in the gear coupling, this feature eliminates the greatest concern of the user and to an equal extent the designers of the rotating equipment. In the case of the previous example a specific load can be used in sizing the gear and thrust bearings. About the ouihor KlN Kn-q.ulrpn 'j oined
Alli.s-Chalmers'
Tool and Die Department in 1941, During the usar he was sta,tioned in lron in the combat militarg police. When the
ended, he was ivr, the Of f icer's Canclidate Scltool i,n Fort Benning, Ga, After getting his d:ischarge, he attended, the Unioersity of Wisconsin in Mihtsaukee and, reioined Allis-Chalmers in 1947 as a fi,eld seruicematn. In 7952 he trans-
uar
e,rt'ed to Allis-Chalmers' Com.pressor Engineering Department, tohere he utas 'instramental in the deuelopment of high pressure oil seals, laigh' speed rotor designs, precision coupl'ings, pioot shoe beat"ings, aibration analEsi,s, anil lube ond seal systems. In 1960, he uas transf
This eliminated a big unknor'vn factor in system
designs.
In this exercise, both the disc, and the gear coupling can be applied. The diflcrence between them can be noticed, but even though one has advantages over the other, both will succeed as couplings. The one overriding difference to users is the obvious promise that the disc coupling will be less susceptible to usage than the gear coupling, and that it lvill require less attention.
It should also be very apparent that
turbomachinen'
couplings, whether gear or disc type, should not be simpll' picked lrom a catalog. There will also be trvo phases of coupling applications. One will be new applications on new rotating systems. The second rvill be the replacement of existing couplings with one of a new t)'pe. This latter phase poses the most problems in application. As mentioned earlier, an intimate knorvledge of the rotating equipment is necessarv to proper coupling design. If an existing coupling is to be replaced rvith a nerv type,
there is good justification to review the nature of the rotating system to be coupled. Some installations are ver\' old and some have been revised in other rvays in the field.
Such engineering revie'"vs are not easy to arrange rvith busr equipment suppliers. The tendency is therefore to match the obvious characteristics of the existing coupling and see what happens.
With many older
designs having relatively heavy and
larger diameter shafts the retrofits have been, as far as rve
can tell, ver)/ successful and trouble free. Part of this success is due to the consideration siven to the retrofit br cooperating engineers of the coupling manufacturer and the rotating equipment manufacturer. A large part is due to the dedication of the first cornpanies offering the disc coupling r.r,herein extra efforts to insure success have been made.
In addition to new unit applications, an increasing number of users of our compressors have been replacing the original gear coupling with the disc type. All of the new unit installations have been successful and, with one exception, all of the retrofit units have been successful. The one case where it r'vas not, the gear coupling characteristics ivere simply copied. \{hen the driver and driven shaft characteristics were suitably considered, and the coupling modified, success was achieved. The purpose for changing the coupling varies unit by unit in accordance lvith the problem stated earlier. In verl' few cases was it necessary because the gear coupling could not be made to work. In most cases, it was because of the training necessary to make the gear coupling work. In one application the retrofit disc coupling has never been installed, because after careful fitting, the gear coupling is still operating after nearly t'wo years of almost continuous duty. ACKNOWLEDGMENT Orieinallv presented at the 2nd Annual Texas A&M Turbomachinery
poiu6,
Sym-
Coilege Station, Texas, October 1973.
ferred to the Compressor Department Marketing Group to
deaelop the Compressor Customet' Seruice Group, which he superaises tod,ag.
REPRINTED FROM HYDROCARBON PROCESSING
87
How reliable is your high pressure seal oil system? Current designs tor high pressure seal oil systems have inherent problems that limit reliability. Here is an explanation of these problems and what to do about them Briqn furner, Imperial Oil Enterprises Ltd.,
Sarnia,
Ont.
Loss on sEAL orr- on a centrifugal compressor creates a very hazardous condition. Without seal oil, most compressors will leak gas from the seal areas. The higher the compressor suction pressure, the greater the leakage and the higher the risk of a fire or oI a fatality. We became aware that the reliability of our system was inadequate when we lost both seal oil pumps in succession twice within a two-year period. Detailed examination of our system suggested that the source of the problem was extrapolation from low pressure system designs. Some modifications had been incorporated to permit the system to operate, but a thorough design review had not been made. To understand the situation, consider the system proposed for API 614, Fig.321 (Fig. 1). This systern works fine for low pressure systems. There are hundreds of them in operation. The tank temperature can rise to the average seal oil return temperature. IJsually, however, it r.r'ill stabilize slightly below this temperature because of heat loss by radiation from the tank, and Irom the pressure controi bypass. Provided the pump can accept oil of the corresponding viscosity, there is no problem.
As ihe discirarge eressure increases-. the energ.v aclded bv ihe pumps increases. ;\t some jlressure, the reservoir temperat'rre wiii rise above the seai return temperature. Irultl:er, as tl:e pump presslrre is raised, the minimum accrpiaicie cil viscosity increases. 'iherefore, at some Dressure, tlie oil wilt become too hot for tir.e pumps. Expediencv ciictated ihat as pressures rose, the i:ypass r::ust l:e cooied. The new' connection is srtrown in Fig. 2, Tiris syste*i is cperable oniy il the :'eliei vaive does not actuate.
It tirerelole
almost doubles. Consequently, the spare pump is destro,ved as ruthlessly as the first one.
The disclosure of such a major error begged the question "IIow many other built-in problems are there?" The analysis revealed other unacceptable deficiencies. This article describes our analysis of the potential problems of current designs and proposes an alternate system. We submit that the alternate will eliminate the selfdestruct loop and will increase the over-all reliability. RELIABITITY CRITERIA
fn
reviewing the current design of seal oil systems,
rue
decided on three objectives. The primary criterion had to be that the seal oil supply must be reliable. Secondly, the
probability of main pump failure must be reduced. The third requirement was that failure of one pump must not cause failure of a second. For the purpose of the analysis,
the system shown in Fig. 2 was assumed to be a good illustration of current practice.
Seql oil sysfem reliobility. We defined seal oil reliability as the continuous supply of oil to the compressor seals at a pressure above the gas pressure, Anv of the follcrving conditions could result in a seal oil supply failure: o leadequate power on ihe pump e' I'ump wear o Faulty reiief valve e tr,ow oil ievel in the reservoir e Dirq' filter s Controi ioop iaiiure.
ceir:taies a self-destruci looo, a possi-
'lie operai.i-:rg mode r,-hich wiil result in :apici destruction of rhe pru:::s. Tl-.is loop ,:f pump luear, pir,rgged filters, debris anri hi;;ii tr,:mperature is -qcl.r'-perpr:irating (Fig. 3), Suppose tir;,.t Ior soille reascn ;he purnp starts to wear" The.wear pa:rtrcies wili siart tc plr-rg4 tire filter" The c.iscirarEe oi:esslire rrii! rise. 'itre hea-t input ."viil ,ncrease. Afler a '"vkik:, the reiie.f vatve on the pi.,a;: rviil o;:etr. This 'nl'iii rei$ce thi: cii flcr,r'iri tlh-e cooirr a:rC ihe s,1,st€n1 heat rejection. It wiii alir: diirnp det'ris hi;ck lrrts tire 'iank a-r-rd i,:-lto the ,:rumn :uc;ion. E,r,er.tr:::ii;,r, :hs f lie,: r1eEE
posits will reduce the flow to the seals. The reduced oil flow to the seals automaticaily starts the spare pumP. IIowever, the usual tlpe of pump for this service requires a higher viscosity oil when starting under pressure than when operating. At this point, the viscosity is too low for an operating pump, and the oil is loaded with debris. When the spare pump starts, almost all its capacity goes through the relief vaive, and the heat into the reservoir
Fig. ?*Typicai i:lgh preEsure ceiirjlreBsai sea! oii sysiEi:,
HIGH
PRESSURE
SEAI OIT
SYSTEMS
High pressure systems are another story. Failure of the lubrication film in the pump causes the rapid generation of fine metal particles. In addition, the high pressure filter size is usually restricted because of cost. Use of metallic screens instead of depth filters also tends to increase filter plugging rates. For a given flow rate and pressure drop, the screen requires a smaller con-
Fig. 2-Typical system modified for cooled bypass.
tainer than a depth filter. At high pressures, the container cost is significant and the screen filter costs less than the depth type. However, the debris-holding caPacity of the screen is less than 10 percent of that of a pleated paper unit. Examination of Fig. 1 reveals another interesting fact. In neither the low pressure case nor the high pressure modification does starting the spare pump significantly increase the flow to the seals if the filter is plugged. In both cases, the only increase in differential pressure is caused by a slight increase in safety relief valve accumulation.
A further problem with the high pressure modification If the spare pump starts, the pressure drop will increase by a factor of four. To prevent collapsed filters or blpassing, either the elements must be changed at one-fourth of their collapsing pressrue, or the collapsing pressure should be greater than the differential between normal pressure downstream and the PumP relief valve setting plus accumulation. A solution to all the problems with filters on these pumps would be to provide two separate trains of pump: cooler and filter. The filter differential pressure could be used to initiate starting the standby train. Starting the auxiliary pump would not increase the flow through a filter. is the filter flow.
s,r$?sra ?,i-i:ws - :i.'! Fig.
3-Self destruct loop for high
pressure seal oil pumps.
Inadequate power to pump-pump wear-faulty relief valve. Each of these items u'ill cause a fall in head tank level and actuate the automatic startup of the auxiliary pump. There is no problem if there is sufficient storage in the overhead tank between the auxiliary pump start level and the trip level. The 10 minutes called for in API 614 is adequate for either a motor or a turbine start. However, high pressure vessels are verrar costly. Vendors balk at supplying five minutes even on low pressure sJ/stems. Our high pressure s)rstems had six seconds storase!
Low oil level in reservoirs. Loss of pump suction due to iow reservoir level will cause loss of seal oil. The trow oil Xevei couid be caused by a cooler ieak, too high a level in the head tank or a faiiure of an inner seal bushing. With systerns designeci according to API 614, Fig. 32, the loss of oil on a coojer ieak can be very hieh. At 2,000 psi, a /2-inch hole wiil pass aooroximateir' 14 U.S. gpm! C'arrently, the onlv protection against loss of oil inventory is a iow level aiarm. Futting on the sDare cooler '.o,ouid be a ieason:1o1e action in the event nf a io,,r, level ziavv:t. I{owever, a 'beiter way to l:randle cocier teairage u,ould be to have the cooler ail cressure onlv a little ;:irove i,.,ater ilressure. 'lhis rvould give the operalor iime ic takc c ,.:i'ec'.ive rc(it.r.
ilirti"
fiIters, Iiirers on icrv pressure seai si'51srns do not foul at r:lacceptable raies. iVhen the high diffc|r:;itial Xlr:essure :+larrr cn e filt-n:r socrri{]ls, t};e oi,ei'ator can usually v,,a1k over and swilrg ;:vei" thc irai:{*r v,,;lve. ;ri:r:l:.11.v
REPHIniTEE FR0tui t'r"\jDRGCAfrECIl,i PFIOCESSiiliG
Contr,ol system failure. Failure of the level control loop on the head tank can cause a loss of seal oil. Opening the bypass valve will cause the loss directly. Closing the valve will dump seal oil into the machine, and the pumps will lose suction. In spite of the history of other level control loops, ou.r experience in seal oil service has been good. However, there are bound to be some failures in the life of a compressor. The API 614 Standard recognizes this and requires a separate Ievel switch to shut down the machine. However, it ignores the hazard of
blowing the seals. It would seem reasonable to provide a redundant system for levei control. One way would be to install an additional transmitter. A pressure switch set to alarm at a fixed error between the two outputs would warn of failure of either transmitter. Conserve nol repoir. Iligh pressure oil pumps are more lrrlnerable than most other mechanical equipment. To varying degrees, dependent on type, they can be damaged by the following:
o Low oil o Dirt
viscosity (related to temperature)
. Low speed c Loss of suction o Excessive differential pressure o Entrained gases o Loaded starts and stops. Eg
Viscosity control. If dirty seal oil is returned to the rEservoir, the viscosity can be affected by contamination. As stated earlier, the use of a cooled bypass was an expediency step to make the oil in the ,"r".rro-i, acceptable to the pump. It was a foolish step for two reasons. No provision was made to ensure that the necessary cooling oil was circulated. This would have required an orificed flow of say 50 percent of pump capacity. Secondln it is a most inefficient method of cooling. We need as much of a safety factor in viscosity as reasonable. There are two ways of meeting this objective. One is to use an oil with as high a viscosity as practical. The second is to cool the suction to as low a temperature as practical. With a given water temperature, a suction cooler will give a lower temperature than the high pressure arrangement illustrated in Fig. 2. Filtration. Almost all vendors recommend that the pump suction be filtered. This creates a dilemma. Plugged suction filters would cause a seal oil failure and may destroy the pump. Putting automatic bypasses on these filters would render them useless at the most critical time and probably result in a pump failure. The use of pressure filters on the suction of the high pressure pumps would not eliminate the problems. Ilowever, the greater acceptable filter Ap with this approach would give the operator more time for corrective actions. It would therefore reduce the chance of a pump suction failure due to a dirty filter. Speed. The lubrication film on screw pumps will fail if, at a given viscosity and differential pressure, the speed is
reduced below a certain limit. The speed limit increases as the viscosity decreases. It also increases as the differ-
ential pressure increases.
One way of avoiding the low speed problem is to use a motor-driven pump, fn separate systems, with a reasonably sized head tank, this is possible since interruption of the oil supply during turbine start is not a problem. Many people do not believe a turbine is a satisfactory driver for the standby oil pump in lubrication service. Therefore, in combined systems where the main oil pump is turbine-driven, speed monitoring is essential. The
higher the diflerential pressure in a given pump, the more critical this becomes. Loss of suction. As stated earlier, loss of suction will quickly destroy a high pressure pump. Tripping the pumps on loss of reservoir level offers maximum protection to the pumps, However, it may unnecessarily cause a seal oil failure. A reasonable balance would be to risk damage to one pump while safeguarding the other. One way of achieving this would be to lock out the spare pump on low oil level alarm. Another would be to switch to the spare at low oil level alarm. Excessive differential pressure. The use of safety relief valves as currently specified should ofler adequate protection against excessive diflerential pressure. Ifowever, the arrangement shown in Fig. 2 is most unsatisfactory. If this system must be retained, the pump should be shut down by flow or temperature switch when the relief valve actuates. Operation of the relief valve will destroy the pump by temperature rise or by debris. A relief flow shutdown would almost always result in a seal oil failure because the spare pump would also shut down on overpressure.
90
Entrained gases will cause pump failure by cavitation and by overheating. The following features are essential to prevent failure. All high pressure returns to the reservoir should enter belor,v liquid level and the ends of the pipes should have target plates. Lolv velocity lines, such as returns from the machine, should enter at or below liquid level. Dirty seal oil returns must be thoroughly degassed external to the reservoir.
Loaded starts and stops. Starting and stopping some types of high pressure pumps under pressure can cause
damage. This tendency is aggravated if there are tu'o pumps, a centrifugal and the booster, in series on the same driver. It is difficult to see lvhy this should be, although it may be because of high oil r.iscosity durir-rg starting. One proposal to deal lvith this problem is to use pilotoperated safety relief valves on the high pressure pump discharge. The safety relief valves would be held open during startup of the pump, using solenoid controlled pilots and closed after startup.
One wreckscl-en6 lo go. If we
assume
that
one
pump is failing, how do we protect against failure of the other? The second pump has two minimum requirements. The fluid must have a satisfactory viscosity and must contain no debris. Clearly, the high pressure sysiem sho*'n in Fig. 2 will not meet these criteria. If the pump is failing. the debris will either be passing through the relief valve
or the filter
element
will be
bypassing. Also, the heat
rejected by the cooler will have decreased. Therefore. the oil will probably be too hot. Also, when the second pump starts, the heat load to
the reservoir doubles. This is because all the additional capacity must pass through the relief valves into the tank. Even if the second pump survived the startup, it rr-ould not live for long. This additional heat load could be avoided by shutting down the damaged pump. This requires instrumentation to ensure that the spare is operat-
ing before the other is shutdown. The best solution to this situation is to specifl. that the oil to high pressure pumps be cooled and filtered. Such a requirement rvould prescribe trvo pumps in series. A low pressure pump to raise the oil pressuie to abor-e cooling water pressure, and the high pressure booster,
In
between the pumps would be the cooler and the suction filter. Tu,o such trains would ensure the spare pump would not fail on starting because of high tempera-
ture or debris. Pig. 4 is a section of Fig. 3+,35 and 36 in API 61-l. It illustrates another possible way of destroying the second pump. The filters follorving the coolers are considered to be the pump suction fiiters. No other filters are required. Hor,vever, suppose the main pump started to destroy itself. After a while the filters u,ould plug, reducing the florv to the system. The level switch on the overhead tank rvould start the spare pump. The oil florv from this r,r,ould overpolver the main pump. If debris should lodge in the check ,,,alve holding it open, the good pump would circulate through the damaged one. All the debris contained in the main pump and that being generated would be dumped into the suc-
HIGH
PRESSURE SEAL
OIt
SYSTEMS
Fig:s. 3a, 35 and 36 of API 6t4, that is what was intended. No one is going to change from an established system
unless they are convinced it is unsatisfactory. What then, is a good system? We would propose the system shown in Fig. 5 for all seal oil systems over 1,000 psig.
Fig. 4-Typical dual pump train high pressure seal oil system with cooled and filtered suction and filtered discharge.
There should be two complete trains of pumping equipment as shown, eaeh train coarsisting of: o Suction gate valve o Low pressure pump o Cooler o Filter, 10 microns, 3 psi pressure drop with oil at 100 SSU, not more than 20 psi with oil at 650 SSU, and a collapsing pressure of at least 50 psi. o High pressure pump o Filter, 10 microns, 3 psi pressure drop with oil at 100 SSU, not more thar. 20 psi with oil at 650 SSU, and a collapsing pressure of at least 50 psi. o Check valve
o
Discharge gate valve.
A pilot-operated relief valve shall be provided between the discharge of each high pressure pump and its high pressure filter, discharging to between the low pressure pump and the cooler on the same train. The low pressure filter on the main pump train shall be a duplex unit. fn each train, the centrifugal and the high pressure pumps shall have a common driver. The low pressure
Fig. S-Proposed system to eliminate the pump self destruct loop.
tion of the spare. This wouldn't last long. The above is fact-not fiction. This reinforces the proposal that there be two separate trains of equipnrent with an absolute minimum of interconnection. Two down. One mechanism for destroying both pumps is low reservoir level. If the main pump can not get sufficient oil to satisfy the seals, the oil pressure drops and the spare is started. With high pressure pumps, it is unlikely that either will survive. This possibility has existed for years in the low pressure systems. Why raise it now? There are two reasons.
High pressure pumps are much more susceptible to cavitation damage than low pressure pumps. Also, the rate of flow through high pressure cooler leaks is much greater. In addition, operators were probably more readily available. The obvious answer to this is low reservoir level pump trips. Since tripping both pumps would cause a seal oil failure, the level alarm should be duplicated, one for each pump. An alternate which would save one pump, would be to start the spare pump on low seal oil level. If the main pump were arranged to shut down, as proposed earlier, when the spare is running, only one pump could be damaged.
Whot should we do? If the preceding has seemed a condemnation of Figs. 32 and 33, and to a smaller extent, REPRINTED FROM HYDROCARBON PROCESSING
pump shall be coupled to the driver to facilitate flushing. The high pressure pump may be coupled to the driver. On separate seal systems with a head tank, the main pump shall be motor-driven. On all other systems, our preference is that the main pump shall be turbine-driven. The turbine-driven main pump shall be equipped to alarm if its speed is less than a specified value. The alarm shall cancel if the pump is off-line. In addition to alarming, the reservoir low level alarm shall lock out the spare pump if it is turbine-driven, and start spare pump if the spare has an electric motor-driver. The high pressure pump shall be started unloaded. The pilot-operated safety valve shall be opened by solenoid valve approximately 10 seconds a"fter the pump is tripped and shall close 10 seconds after the pump restarts. The solenoid shall be actuated to cause the safety valve
to
byryass.
A differential pressure of one-half the collapsing pressure of either main pump filter shall cause the spare pump to start.
If the auxiliary pump starts up for any reason, the main pump shall shut down. The shutdown circuit shall verify, by pressure switch at the high pressure level, that the spare is operating before shutting down the main. The high pressure pumps shall be screw type in a steel casing. The capacity shall be based on the seal flow required to give a clean seal temperature of 160oF. The pump shall be sized for this flow plus 10 U.S. gpm or 200 percent of this flow, whichever is the greater. There shall be a high temperature alarm on the cornmon discharge of the pumps. A second pilot-operated safety valve shall be provided on the spare train to vent the system of air. A handoperated vent shall be provided on the main train. The head tank level control shall actuate the two control valves on split range. The bypass valve will be fully closed with a 3 psig air signal, and fully open with 91
a 9 psig air signal. The throttling valve will be fully open at 9 psig and fully closed at 15 psig. The level control shall be proportional only with no reset. The head tank shall have two level transmitters. One shall operate the pressure switches, the level controllers, plus the low level shutdown. The second shall monitor the first and shall alarm if a pressure difference of more than -+ 1 psi exists between the output from the two transmitters. JUSTIFICATION
Moin pump driver. With separate systems, there is normally plenty of time to start a spare turbine-driven pump. Two advantages of running the motor pump are:
.
Lower power cost
o The emergency pump is maintained in first-class condition (as new). When the steam driven pump is normally operated the system is very vuinerable to a power failure when this pump is being repaired. On combined
systems, the rate
of fall of the lube
sys-
o Prevent cross-contamination in event of pump failure o Limit filter flow to capacity of one pump
tem pressure creates a problem. Our experience is that we can normally get a motor-driven spare in service without tripout. However, this is not possible with a turbine spare without an accumulator. Our experience with accumulators has been unsatisfactory.
. . .
Low reservoir profection. The intent is to ensure that one pump survives the emergency. At the same time, the system is protected at the possible expense of a pump.
Duql troins. The purpose of the dual trains is:
Ensure that the spare pump starts with clean filters Cost saving on transfer valves Ensure that spare filter cannot be plugged by oper-
ator error by starting a damaged pump.
Single filters. The only purpose of the pump discharge filters is to retain the debris in the case of a pump failure. Therefore, if they need cleaning, the pump needs repairing. There is no justification for dual filters.
Main pump duo! filters. The main pump is in service all the time. The suction filter will gradually become dirty. If only a single filter were provided, only one train would be available when it was being cleaned. To limit
Stortup unlooding has been found necessary on manv it seems expedient to pla,v it safe. Since there is the possibility of a control malfunction causing a seal oil loss, the solenoid on the safety valve must be activated to unload. applications. However,
Moin pump shuldown. If the spare pump starts up. the most probable cause is a failure on the main train. Keeping the latter in service is asking {or trouble. The best solution is to get the suspect equipment ofl-line as
this downtime, dual filters should be provided.
quickly as possibly.
Single coolers. There are various means of cleaning coolers on stream-shock chlorination, acid cleaning, abrasive washing, etc. Therefore, there should be no
Piston pumps. Screw pumps were selected for this service because they are the only suitable pumps ar.ailable with a steel case.
justification to have a spare cooler on the main train.
High temperqture qlqrm. Screw pumps, like piston pumps, have a minimum viscosity of about 70 SUS. Since
Pilot-operoted sofety vqlves. Standard safety valves have a high accumulation. Pilot-operated safety valves are much more reliable and have a much lower accumulation. Further, the pilot feature can be used for startup unloading.
Single driver for pumps. With the single train system, where the low pressure pump is essential for the high pressure unit, it is a prerequisite that both pumps should have one driver. The alternate is a much more complicated control system.
Aboul the quthor
Bnr,tN TnnNER is an engineering associate in Imperial Oi.l's Engineering Diuision at Sarnin, Ont. His functions incluCe preparation of machinery speci,fications and, consultation on machinerg sclection, tesling, installat:ion, startup, ntuintenance and trouble-shooti,ng. He Ir.as been in th,e Engineering Diuision
16 gears u.torhing oyr. utilities, ptrocess engineering and, macluinery. Bef ore mooing to Canada, Brian utas em,pl,oyed in a uarietg of positions in thermal pouter stations i,n England. He was educated in England in electri,cal and mechanical engi,neering and, is registered a,s a p,rofessional engineer in Ontario.
for
this is easily reached, the operator must be the oil temperature increases.
if
Overheqd tqnk level monitoring. The mean time between failure of level control loops is much less than the life of the machine. The approach taken is to moniior the main transmitter, using a second identical unit. and a pressure switch. The switch would activate an alar-la. With only one redundant unit, it would not be practical to take any automatic action. Two redundant unils would be needed before corrective action could be taken. A number of failures have indicated that the c rrrent design of seal oil systems for high pressure centlifr,rgal compressors is inadequate. One major deficiencl, is the presence of a self-destruct loop that can destroy the high pressure pump. Other significant defects and rvcaknesses are present. These imperfections can be reducecl at little cost. However, the improvements will onlv con're 'f designs conflicting with past practice on lor,r, pressurc svstems are accepted. ACKNOIVLEDGT,IENT Pressure Seal Oil Systems," originally presented at the ASMD Petroleum Mechanical Engineering CoofereLrce. Los Angcles, Sept. 16-20, 1973. Abstracted
lrom "Design of High
j API 614-I,ubrication,
LITERATURE CITED Shaft-Sealing, and Control
Purpose Applications, First Edition, 1973.
92
r,r.arnectr
Oil
Systems
for Spcial
What to specify for gear reliability John C. Finney, Shell Oil Co., Ilouston WrrE,N spEcrFyrNc high speed gears
for critical
services,
a few reliability items should be considered over and above those included in the two accepted standards, AGMA 421 and API 613. The API Standard is written as a user's modification to the manufacturer's AGMA Standards and, if followed, should result in a satisfactory gear installation. The following items are suggested for additional gear reliability.
Geor hordness. If you trace the history of AGMA Standards, you will find that through the years the allowable hardness of gear teeth has been increasing with a subsequent increase in the gear rating. In other words, the average gear which was built, say in 1960, could probably be used today for about twice the horsepower simply by increasing its tooth hardness. Considering the problems that have occurred over this same period, it strikes me that upgrading gears by means of hardness alone is a questionable practice. I would recommend that gear sizing be based on a moderate hardness, say 350 BHN pinion, 300 BHN gear. After the gear has been sized, I would have no objection to increasing the tooth h'ardness during manufacture but not to increase its rating.
Service fqclors. My experience has indicated that AGMA recommended service factors are inadequate. Most users I have talked with share my view on this, although rnost manufacturers still attempt to sell gears based on AGMA factors. I notice that the most recent AGMA Standard 421 (Sixth Edition) has finally relented and increased its recommendations somewhat. This is certainly a step in the right direction, but I still feel they have not gone far enough. As a rule of thumb, my recommendation would be to add about 0.25 to each of the AGMA recommended service factors.
Geqr mounting. Generally speaking, the details of the gear mounting are provided by the supplier of the driver
or the driven equipment or possibly by the contractor.
I believe the average gear manufacturer has definite ideas about mounting of gears, he seems reluctant to make too many engineering recommendations to these suppliers or contractors, who, of course, are his customers. To avoid this pitfall, I recommend that the ultimate user specify that the gear ideally be set directly on a sole plate using no shims, thus requiring mating equipment to be Although
REPRINTED FROM HYDROCARBON PROCESSING
shimmed to the gear. If shims are necessary, specify that they be of equal thickness under each mounting foot and made of r/s-inch minimum ground plate. All gears should be doweled after the hot alignment has been verified.
Spore geqrs. Purchase of a spare set of gears (pinion and bull gear) is highly recommended. Having these long delivery items on hand has been a real salvation on several occasions that I can recall. A fringe benefit which can be derived from a spare gear for a nominal extra is having this set at a diflerent ratio (usually higher) than the main set. This allows some protection for inadequate comPressor performance or for slight uprating (be sure you buy the driver large enough to cover this higher horsepower re-
quirement)
.
Geqr conslruclion. I recommend that you rule out gears made of cast components-this is my feeling based on a bad experience we had with a cast gear. Our failure occurred after about six weeks of operation when a spoke of the gear wheel failed in fatigue. The fatigue fracture emanated from a small crack which, because of casting complexity, had not been detected by magnetic particle inspection. A catastrophic failure resulted.
Oil choking is a conditi'on caused by flooding of the gear housing with oil, resulting in overheating and subsequent fire. Causes are high pitch line velocity, low ratio, inadequate oil drainage, down mesh, and too small a gear housing (gears should occupy only about one-third of the volume of the box) . We had two fires in the same box, one with the main gear and then another with the spare set of gears. Although rboth gears were scored, it was necessary that they be used another time (in desperation). Both sets failed after several weeks additional operation; but, fortunately, we had been able to obtain a new gear by that time. Our report on this failure concluded: After an investigation, it was determined that a condition known as choking had caused the fires. This condition is sensitive to high pitch line velocity, down mesh, low ratio, small box relative to size of gears and inadequate oil drainage' This gear had all of the above conditions and yet had run successfully for several years before the steam turbine driver speed was increased about 100 rpm (still below the design speed) . The slightly high pitch line velocity appears to have set off the choking condition. Apparently, therbull 93
[ilngh-speed geep unnHs
About the quthor
JonN C. FTNNEv is a staff eng,ineer with the M anuf acturing E ngi,ne ering D epartment, Shell Oil Co., Houston. He is i,n charge of rotating mechanical equipment for all of Shell's U.S.A. refineri.es, M,t. Finneg holds a B.S. degree in mechuni-
gear begins to entrain oil from the bottom of the case and pumps it to the top of the box" This tends to over-lubricate
the down mesh creating excess heat in addition to oil foaming and vortexing which blocks the oil drain. The oil level begins to rise aggravating the condition. Oil choking occurs very rapidly and has been described as a tornado inside the gear box. In this case, hot vapors and oil escaping from the vent ignited upon contact with the overheitel gear case. The problem was overcome by the addition of adequate drainage (four additional drains were added to the box). With a new set of gears, the unit has since run successfully at speed several hundred rpm higher than previously.
Quolity control. A most important step in obtaining a succesful gear installation is follow-up of quality control in the factory. In addition to the usual quality control considerations of inspection and testing, I recommend considerable attention be given to quality control of engineering. This consists of an early review of the gear design in the manufacturer's engineering office to assure compliance with the specifications and to ascertain that critical
cal engineering from Tulane (Jnioetntg, New Orleans, La, He is a member of the API Subcommittee on Mechanical Equipment and is clruirtnan of the Task Force on API Standaril 617,"Centri.fugal Compresso,t's for General Refinerg Seroices."
items such as stress levels, tooth ioading, bearing design. etc., cornply rvith good engineering practice. In conclusion, let me stress a conservative approach to gear design. On ani' proposed project, I would recom-
mend a thorough review of the various driver requirements concentrating on elimination of gears rvltenever possible. \'Vhen speed increasers must be applied, specif-v a conservative design. ACKNOWLEDGMENT Based . on intro_d_ucrory remarks made by Mr. .f inney to a panel se:sion on qear units. ASME Petroleum Mechanicai Engineering Con[erence. Lo. An-geles, Sept. 18, 1973. f
High-speed gear fai luresuser experrences R. M. Dubner, Standard Oil Company of California, San Francisco
"1
?;n ars
CA
The phra yu: had
ar \\ras basically two steel diaThe outer forged tooth ring of the diaphragm,s. The gear about tlree weeks nhen increased noise and high vibration were observed, and the unit was shut dorvn. Several teeth were found to be broken at the apex of the helix on both the pinion and gear. The lveld where the diaphragm joins the outer rin,g was cracked approximately two-thirds of the way u.ounJ the circumference.
up r Ioad, back-totightening
factors than
r
service
ble,
full-
In
some
special cases, a consultant is used during the design and engineering stages.
Cqse histories. The types of gear problems experienced are'best illustrated by a few actual case histories. The first was a gear unit on a motor-driven centrifugal gas compt"rro.. The gear was rated at 7,700 hp m"a*irirr-, increasing the speed from 1,800 rpm to 7,900 rpm. pitch line velocity was approximately 16,600 feet per minute.
a
A metallurgical analysis r.vas made by an outside laboratory to determine the cause of this failure. Their analysis indicated a fatigue failure at the weld. It rvas found that tle rveld at this joint was not full penetration. There was a crevice on the inner side of the diaphragn weld. Fatigue cracks emanated from that crevice. A SOCAL metallurgist looked at the same failure and felt that fatigue cracking was only part of the story. There n,as evidence that hairline cracks had been formed by hydrogen embrittlement which could have resulted from a faulty welding procedure or from moisture in the welding rod coating. It was determined that the gear did not receive the required post-weld heat treatment and that the welding rod coating did contain some moisture. These tr'vo factors combined to produce hvdrogen embrittle-
ffiWWgwumffi What to specify for gear reliability John C. Finney, Shell Oil Co., Houston WnBw spEcrFyrNc high speed gears for critical services, items should be considered over and above those included in the two accepted standards, AGMA 421 and API 613. The API Standard is written as a user's modification to the manufacturer's AGMA Standards and, if followed, should result in a satisfactory gear installation. The following items are suggested for additional gear reliability.
a few reliability
Geor hordness. If you trace the history of AGMA Standards, you will find that through the years the allowable hardness of gear teeth has been increasing with a subsequent increase in the gear rating. In other words, the average gear which was built, say in 1960, could probably be used today for about twice the horsepower simply by increasing its tooth hardness. Considering the problems that have occurred over this same period, it strikes me that upgrading gears by means of hardness alone is a questionable practice. I would recommend that gear sizing be based on a moderate hardness, say 350 BHN pinion, 300 BHN gear. After the gear has been sized, I would have no objection to increasing the tooth hardness during manufacture but not to increase its rating.
Service fqclors. My experience has indicated that AGMA recommended service factors are inadequate. Most users I have talked with share my view on this, although most manufacturers still attempt to sell gears based on AGMA factors. I notice that the most recent AGMA Standard 421 (Sixth Edition) has finally relented and increased its recommendations somewhat. This is certainly a step in the right direction, but I still feel they have not gone far enough. As a rule of thumb, my recommendation would be to add about 0.25 to each of the AGMA recommended service factors.
Geqr mounting. Generally speaking, the details of the gear mounting are provided by the supplier of the driver or the driven equipment or possibly by the contractor. Although I believe the average gear manufacturer has definite ideas about mounting of gears, he seems reluctant to make too many engineering recommendations to these suppliers or contractors, who, of course, are his customers. To avoid this pitfall, I recommend that the ultimate user specify that the gear ideally be set directly on a sole plate using no shims, thus requiring mating equipment to be REPRINTED FROM HYDROCARBON PROCESSING
shimmed to the gear. If shims are necessary, specify that they be of equal thickness under each mounting foot and made of /g-inch minimum ground plate. All gears should be doweled after the hot alignment has been verified.
Spore geqrs. Purchase of a spare set of gears (pinion and bull gear) is highly recommended. Having these long delivery items on hand has been a real salvation on several occasions that I can recall. A fringe benefit which can be derived from a spare gear for a nominal extra is having this set at a different ratio (usually higher) than the main set. This allows some protection for inadequate compressor performance or for slight uprating (be sure you buy the driver large enough to cover this higher horsepower re-
quirement)
.
Geqr consfruclion. I recommend that you rule out gears made of cast components-this is my feeling based on a bad experience we had with a cast gear. Our failure occurred after about six weeks of operation when a spoke of the gear wheel failed in fatigue. The fatigue fracture emanated from a small crack which, because of casting complexity, had not been detected by magnetic particle inspection. A catastrophic failure resulted. Oi! choking is a conditi,on caused by flooding of the gear housing with oil, resulting in overheating and subsequent fire. Causes are high pitch line velocity, low ratio, inadequate oil drainage, down mesh, and too small a gear housing (gears should occupy only about one-third of the volume of the box) . We had tw,o fires in the same box, one with the main gear and then another with the spare set of gears. Although rboth gears were scored, it was necessary that they be used another time (in desperation). Both sets failed after several weeks additional operation; but, fortunately, we had been able to obtain a new gear by that time. Our report on this failure concluded: After an investigation, it was determined that a condition known as choking had caused the fires. This condition is sensitive to high pitch line velocity, down mesh, low ratio, small box relative to size of gears and inadequate oil drainage. This gear had all of the above conditions and yet had run successfully for several years before the steam turbine driver speed was increased about 100 rpm (still below the design speed) . The slightly high pitch line velocity appears to have set off the choking condition. Apparently, therbull 93
filngh-speed g]eap unnHs
About the quthor
JonN C. FrNuoy is a staff engineer uitlt the M anuf acturi,ng E ngineering D epartment, Shell Oil Co., Houston- He is in charge of rotating mechanical equipment for all of Shell's U.S.A. refineri.es. Mr. Finneg hold,s a B.S, ilegree in mechuruical engineering from Tulane Uni,uet sity, New Orleans, La. He i,s a membet of the
gear begins to entrain oil from the bottom of the case and pumps it to *re top of the box. This tends to over-lubricate
the down mesh creating excess heat in addition to oil foaming and vortexing which blocks the oil drain. The oil level begins to rise aggravating the condition. Oil choking occurs very rapidly and has been described as a tornado inside the gear box. In this case, hot vapors and oil escaping
from the vent ignited upon contact with the overheated gear case. The problem was overcome by the addition of adequate drainage (four additional drains were added to the box). With a new set of gears, the unit has since nrn successfully at speed several hundred rpm higher than previously.
Gluolity conirol. A most important step in obtaining a of quality control in the factory. In addition to the usual quality control considerations of inspection and testing, f recommend considerable attention be given to quality control of engineering. This consists of an early review of the gear design in the manufacturer's engineering office to assure compliance with the specifications and to ascertain that critical successful gear installation is follow-up
API Subcommittee
on Mechani,cal Equi,p-
ment and is chuirman of the Task Force on
API
Sta,ndard,
pressors
6
1
7, "
C
mtri.fugal
C
om-
for General Refinerg Seruices."
items such as stress levels, tooth loading, bearing design. etc., comply with good engineering practice. In conclusion, let me stress a conservative approach to gear desien. On an1, proposed project, I rvould recom-
mend a thorough review of the various driver requirements concentrating on elimination of gears ujtenever possible. \,Vhen speed increasers must be applied, specif1,. a conservative design. ACKNOWLEDGMENT Ba.ed.on intro_ductory remarks made by Mr, Finney to a panel.eision on sear units. ASME Petroleum Mechanicai Engineerini Conference. Lo. Angeles, Sept. 18, 1973. -
High-speed gear tdluresuser expenences R. M. Dubner, Standard Oil Company of California, San Francisco o 1
c
tightening up on gear requirements, using higher service factors than required by AGMA. Whenever possible, fullload, back-to-back factory test runs are specified. In some special cases, a consultant is used during the design and engineering stages.
Cqse hislories. The types of gear problems experienced are best illustrated by a few actual case histories. The first was a gear unit on a motor-driven centrifugal gas compressor. The gear was rated at 7,700 hp maximum, increasing the speed from 1,800 rpm to 7,900 rpm. pitch line velocity was approximately 16,600 feet per minute. 94.
The phra was had
ar
was
The o of the
about
steel dia_
ooth ring
The gear u,hen in-
creased noise and high vibration were obsen,ed. and the
unit was shut down. Several teeth were found to
be
broken at the apex of the helix on both the pinion and gear. The weld where the diaphragm joins the outer ring was cracked approximately two-thirds of the way around the circumference. A metallurgical analysis was made by an outside laboratory to determine the cause of this failure. Their
analysis indicated a fatigue failure at the weld. It u,as found that tfre weld at thi,s joint was not full penetration. There was a crevice on the inner side of the diaphr.agm
weld. Fatigue cracks emanated from that crevice, A SOCAL metallurgist looked at the same failure and felt that fatigue cracking was only part of the story. There rvas evidence that hairline cracks had been formed by h,udrogen embrittlement which could have resulted from a faulty welding procedure or from moisture in the welding rod coating. It was determined that the gear did not receive the required post-weld heat treatment and that the welding rod coating did contain some moisture. These tn'o factors combined to produce hydrogen embrittle-
ment, probably the principal cause of this fabricated gear failure. Other contributing factors uncovered were that the input shaft of the gear was misaligned 0.035 inch and the compressor had apparently been surging during operation. The controversy which developed was rvhether or not the gear would have failed in service even if it had been fabricated properly. The company said, no, the vendor said. yes. And that question has never been resolved. The gear was entirel1, rebuilt. The welding procedure was changed and provisions made for post-weld heat treating. Welding rods were kept rvarm to prevent moisture absorption. The repair on this gear was successful. It has been in continuous operation for over four vears
now rvithout giving any additional trouble. Because of this problem, Socal has discontinued the use of fabricatedtype gear constrr-rction for high-speed gear units. A pitch line velocity of approximately 15.000 fpm has been set as a reasonable maximum for fabricated gears.
Above that, solid-forged gears or a solid-forged gear pressed on to a forged shaft are required. The next case concerns a long, continuing problem rvhich really isn't resoived yet. This gear connects a gas turbine to a centrifugal air compressor in an amnonia plant. The horseporver is about 16,000, increasing the speed from 4,860 to 5.815 rpm. This low-ratio gear is of double-helical, forged construction. The unit was first started in May, 1967 and was immediately noisy and rough. The teeth rvere scored. In early June of that vear, there rr.as more pinion scoring and an extreme-pressure additir,e rvas blended into the lube oil. Later in Jr-rne, the pinion rvas inspected and the scoring seemed to be healing slightl;,. Ifowever, a new gear box was ordered in anticipation of a failure. After an additional examination in September, 1967, the gear box continued to run successfully for a period of five vears. Finally, in November, 7972, the output shaft of the gear broke between the bearing and the coupling hub. The new sear bor ordered in 1967 r.vas installed. A four-hour test was run and the tooth contact pattern examined. The pattern \vas jroor. In January, 1973, the gear teeth were found scored and pitted at the pitch line. Later the vibration increased. Loacl was reduced by about 10 percent. In Februarv, 79i3, the vibration continued to increase. Inspection revealed some chippine of the teeth. A gear of a different make from another refinery r'vas bolrorved. It is still rr-rnning satisfactorill,. SOCAL and the consultant hired to investigate the problem believe that the gear was of marginal design and unsatisfactorv for the service. When the plant was new, it was operating at reducecl capacity. The gear experienced some initiai pitting, went through a healing process and then ran satisfactorily for apprcximately five years. The ne\'v g-ear box started up under full load conditions and so it did not have a partial load break-in period like the first gear. The solution will be replacement of this gear with one of stronser design. Silr.er plate on the nelv pinion will be required to improve break-in. This plating technique has worked quite successfully in past installations. For replacement of a critical gear like this, a thorough design analysis is needed, r,vith some outside help. Bending and compressive loads must be calculated and construction rnaterials reviewed with particular considerations for long life. REPRINTED FROM HYDROCARBON PROCESSING
About the qufhor RTcHARD M. Dunxpn
i,s
staff engineer-
mechanical equipment, Standard Oi'I Co. of California, San Franci,sco. Mq', Dub-
ner superaises the Mechanical Equiy ment Group. After recei,uing a B,S, de'
gree in mechanica,l engineering from the Uni,aersitJ of Cali,fornia at BerkeleE ha was emploged by Tidewater Oil Co. and
then Philldps Petroleum Co. at their Aron refinery. He joined Stomdard Oi,l Co. of Califoq"ni.a in 1968, ulorking for the Corporate Engi,neering Department. He represents SOCAL on the API Subcommi,ttee on Mechanical Equipm,ent.
Oil inlet orifice. There have been other gear probierrrs not caused by fundamentai design errors but are nevertheiess of interest. The first occurred on a gear connectinq a gas turbine driver to a tandem pair of centrifugal compressors in a hydrogen plant. The gear unit was rated at 24,000 hp, with approximatel'1 a 2 to 1 ratio. This unit had all the benefits of an independent design analysis and is running quite well. After an initiatr run-in period at reduced speed, the unit was brought up to full speed. Shortly thereafter, the gas turbine tripped out on high exhaust temperature, The gear box was quite hot. Oil was squirting out every opening.
A restriction orifice in the oil inlet line to the gear had been omitted. The manufacturer's procedure was to deter-
mine the oil flow requirements on the test stand and then size the orifice. The orifice ivas correctly sized but sonehow, it was never installed. The box was overfilling with oil. The power consumed by the gear in this overfilled condition was many times design which caused the turbine to trip out on high exhaust temperature. Fortr-rnately, no damage was sustained by the gear.
The orifice was not shown on the piping drawings. Those involved with installation and startup did not knor'v that the orifice was required before starting up the unit in the field. It's better to buy a gear which doesn't require an orifice on the inlet of the oil supply line. But if it must have an orifice, make sure it's clearly shown on drawings and make a field check to be sure it's installed. Breather vents. Another example happened to a gear connecting a 12.500 horsepower gas turbine to a centrif-
ugal compressor in a reinjection plant. At startup, this unit had high vibration and the gas turbine tripped out on overload. The breather vents in the top of the gear box had not been installed. These breathers were shown on the ..,endor's dralr''ings. The problem was cured by simply installing the breathers. The gears have now operated satisfactorily for l/z vears. These last two case histories indicate that failures like these are always possible and are not necessarily caused by poor design. Mistakes will occur during installation and startup but in many instances these things are easily identified and corrected. Fundamental design problems are much more difficult to solve and require cooperative effort between the gear manufacturer and the user. These fundamental problems should be avoided during the engineering stage by a conservative design philosophi, and
careful review of critical details. ACKNOWLEDGMENT
Broed on introductory remarks made by Mr. Dubner to a panel session on gear units, ASME Petroleu Mechanical Engineering Conference, Los Angeles, Sept.
lB, 1973.
I
95
ffingh-speed gean unnhs
Relationship of vibration to gear quality DBsrcNens can engineer every gear. They can define the design criteria, design loads in bending, etc., on gear teeth. However, when the engineering design is turned over to the shops to build a gear and somebody makes a mistake, problems begin. They have a lead error or a helix mismatch or the shop misses on the tooth-to-tooth spacing or they have some sort of apex wander. AII of these mistakes have the effect .of ,shifting the pinion with respect to the bull gear. Calculations on a 7,000-hp gear indicated
a 0.004-inch wander, or axial vibration, and the
gear
failed.
Axiol vibrqtion. There is a new AGMA Standard 426.01, "Specification for Measurement of Lateral Vibration on Herringbone Gear Teeth." The measurement of lateral vibration is described in great detail and one of the first things stated is that the lateral vibration shall be taken perpendicular to the shaft's axis of rotation. Some
operating company engineers contend that one of the most significant measurements is the axial vibration measurement and that the axial vibration measurement on a_highspeed, high-horsepower gear should be no more than t mil peak-to-peak in amplitude at any given frequency. It has also been the experience ,of operating company
engineers that as gear errors were corrected, to achieve this low amplitude vibration, noise problem diminished greatly. In other words, the pinion doesn't hammer axiall,v against the bull gear. How much Ioading can the pinion take due to the separation of horsepower between the pinion and bull gear? This power can be calculated and is rather significant. One engineer reported seeing two mils in line rr.ith the riding point of the pinion on the pinion bearings, r,r,here it is supposed to ride beside the buil gear. The recommendation of the AGMA says, "satisfactorv amplitude is that plane is two mils at 5,000 rpm." If lou rvould check the forces, some believe you would finc1 these forces beyond the realistic capabilities of the gear or the bearings.
It has been suggested that all the vibration spectra should be recorded on tape during tests. It is quite significant to see two mils vibration on the pinion at bullgear frequency indicating a pitch line runout of the bull gear. Sometimes this vibration is disregarded by the manufacturer. They want to look only at the level of pinion vibration at pinion speed and completely neglect toothpassing frequency.
I
AGMA views on shop testing and vendor required data F. A. fhomo, Delaval Turbine, Inc., Trenton, N.J.
Trm AtvronrcaN Gnan MeNura,crunp,ns AssocrATroN (AGMA) has or is taking action on new standards or
information sheets on shop testing and vendor required data. Such action will be the base for this discussion.
in i* entirety does not readily lend itself to standardization. The range in the size oi gear units Shop testing
96
in shop facilities is the primary reason. Certain general tests are normally performed, however, rvhich include such things as full speed and overspeed operation, demonstration of satisfactory oil flow and drainage, gear tooth contact patterns, backlash, clearances, etc. The load and
at which these tests are performed ranges from zero to full power. The relative value of a spin test vs. a partial load or full load test, depends to a large extent on the type of unit, intended service, and the character-istics of
the system into which the gear will be placed. Obviously, the more elaborate the test, the more costly.
The two primary aspects of shop testing that have rapidly been gaining interest are: vibration and noise. I will describe AGMA's recent action in these areas. Vibrotion. For years, test engineers have been measuring the vibration of equipment operating under part, full, and no-load conditions; in the lab or in the field; and isolated or in the middle of a system. This led to the necessity of establishing what might be considered good, bad or tolerable vibration levels. The absence of a generally accepted standard has led many user companies to write their own specifications,
'#t
with the result that a number of differing requirements can be specified to the gear builder, adding unnecessary cost and confusion. Many of these specifications have a stair-step limit on allowable (broad band) amplitude in terms of shaft rpm, reflecting the over-simplified concept that a significant vibration occurs only at a frequency corresponding to shaft rpm. Frequently, the specifications also use words to the eflect that "balance shall be such that the vibration shall not exceed . . .," thus ignor-
ing the many other
sources
of vibration. Rarely
does unit has
an existing specification recognize that a gear at least two rotational speeds. The gear wheel speed is often as low as 900, 1,200 or 1,800 rpm, while in contrast, the pinion speed is 4 to 8 times as much. Obviously, the permissible vibration is not the same even though the parts share a common housing.
fn our efforts to
develop an AGMA standard, the
committee literally pooled its collective background data on fi.eld performance, factory tests, and technical skills to produce a realistic and workable approach. Notwithstanding the importance of definition, instru-
ment sensitivity and calibration, a key portion of the standard is Fig. 1 which specifies acceptable vibration limits. The committee's object was to develop a "damage avoidance" curve with a built-in margin. For several obvious reasons, it is desirable to express such a curve in terrns of displacement, velocity, acceleration, or at least some mathematical expression. As it turned out, no single parameter or expression would be satisfactory. The "damage avoidance" curve could be best approximated by a combination of : (a) constant displacement at low frequencies; (b) constant velocity at intermediate freq.uencies, and (c) constant acceleration at the higher frequencies.
However, in view of the fact that the American Petroleum Institute had just recently revised Standard 612 "Special Purpose Steam Turbines for Refinery Service" and had adopted the expression (12,000/tpm)L/2 as a peak-to-peak amplitude limit, the AGMA committee felt it desirable to use this expression [(200/l)1/2 where I is vibration frequency in Hz] instead of constant velocity for intermediate frequencies so as not to be in confict with the new API limit. The fnrits of this committee's efforts are: AGMA Standard +26.01 "Specifi.cation for Measurement of Lateral Vibration on High Speed Helical and Herringbone Gear Ijnits."
Noise. As those of you who have worked with noise appreciate, it's not a simple subject. For example, in specifications dealing with most other subjects, the euors asREPRINTED FROM HYDROCARBON PROCESSING
vmATI{rI ffiEouErffi $r4
Fig. l-Acceptable gear unit vibration limits.
with the environmental conditions or instrumentation acc:':racy are usually small enough to be disregarded in all but the most precise laboratory problems. In acoustical measurements, however, many things combine to alter this condition and severely complicate the measurement problem. The (dB) scale of measurement is logarithmic which expands the apparent influence of lowet sound pressure levels. This is so because the sociated
ear hears, or perceives, differences logarithmically and to facilitate objective/
instruments have been designed subj ective considerations.
Since all airborne acoustical measurements are without contact between the instrument and the source of sound, the environmental influence should be considered. fn acoustical measurements the environmental influence becomes critical as the compression waves generated by the sound source interact with any and all surfaces and obstruction. The resulting energ'y loss or reflection can change the measured sound pressure level many orders of magnitude. It is altogether possible to get changes of 3,000 percent by varying the environment alone. Such a variation expressed in acoustical notation would be over 14 dB. To assist in defining the acoustical environment location ("field"), many terms have been developed. The measurement position may be indicated as being in the fieldr" field" or t'reverberant field." fn each
"direct "near of these generalized fields diflerent conditions exist, which
must be considered to obtain reasonable measurement accutacy.
In addition, sources do not radiate sound equally in all directions. This results in large variations caused by such things as source geometry, radiation efficiency, maximum energy areas and others. Variations due to source directionality are usually in the order of -r- 5 to 20 dB. The instrumentation used is poor at best, when one considers typical laboratory accuracies in other disciplines. A Class II sound level meter, which is the recognized basis for compliance with U.S. federal noise level standards has, at best, an accuracy of -+ 3 dB. This means 97
ffingh:speed gean unnHs it reads the airborne pressure changes to a tolerance of plus tr00 percent, minus 50 percent. Class I, or precision sound level, meters are good to approximately that
-f i.5 dB.
To complicate the problem further, certain other conditions arise in, for example, reverberant spaces, The creation of standing waves can change the Ievel at aty given spot in the area 15 to 20 dB as compared to other iocations. This variation is frequency related which means that the level for one narrow frequency band may be plus 10 dB over the adjacent frequencies at a given location while the levei for a different naryow band may be i0 dB below the adjacent frequencies. Despite these complications, the AGMA Acoustical Technologv Committee has recently prepared a proposed standard called "Sound for Enclosed Helical, Herringbone and Spiral Bevel Gear Drives." Its purpose is to present the instrumentation and procedure to be used for sound measurements. It also presents typical maximum sound levels (A-rveighted sound pressure levels). This particular standard is limited to gears operating
at a maximum pitch line velocity of 5,000 fpm and a rnaximum speed of 3,600 rpm. The standard is presently being balloted by AGMA member companies and hence is not available at this time. The future plans of the AGMA Acoustical Technology
Committee include:
some
consideration arrd, hopefully, to establish better communication betlveen "system engineers" and gear manufacturers.
The term "overload" will be used to describe a condition that overstresses component parts, and may lead to failure. Overload can be of momentary duration, periodic. quasi-steady state, or vibratory in nature. Depending on its magnitude and the number of stress cycles accumr-rlated at overload, it can be a fatigue or a yield stress consideration. Overload on a gear unit can result fronr internal or external causes. Internal causes of ove;load
as poor acciracy, faulty assembly, bumps or usually found by routine ir.rspections before the unit is put into service. Externai sources of overload result from the operational characteristics of the svsten-r into which the gear is placed, and
-such bruises on gear teeth-are
are more complex and difficult to identify. Some commor.. sources are:
l. Moximum conlinuous power. If a gear unit has
for the nominal design point load, overload inevitable. The pump or compressor designer can predict the potr.er requirements at the design point u'itl.r fairly good accuracy. Arriving at the maximum continuous pou,er is another story. It has to be estin'rated and is a combination of been sized
:
A.
(a) Fundan-rentals of measurement (b) Sources of gear noise (c) X,{ethods Ior quieting
Changes
in
specific gravity or density
of the
n'redia
beir.rg pumped,
B. Carry out,
B. tr,tethods for predicting gear noise from design and manufacturing data. gear
drives.
D. Handbook on gear noise. DATA REQUIRED BY VENDORS To discuss data requirements presupposes effective communication. Unfortunately, the industry,s eflorts at communication haven't been that good. AGMA's High Speed Gear Committee is currently engaged in the preparation of an information sheet which, hopefully, will help improve the industry's ability to communicate, particularly in those areas that affect the capacity and reliability of gear units. The following discussion has been borrowed from the committee's working draft.
The increasing demands for "mechanical reliability" 98
The follorving points are intended to highlight
of the subjects for
is
1. Similar standards for other products as the need arises. 2. Consideration of parallel activities such as: A. Educational program on gear noise for AGMA members andf or gear users.
C. Recommendations on acoustical enclosures for
can best be satisfied by a coordinated technical exchange between system designer, user, equipment builder and e.recting engineers. This coordinated effort is properli, called "system engineering," and is normally performed by the design agent or his technical representative. Obviously, the investigation of potential problems during the design stage is preferable to a search for solutions after the equipment is in operation.
C.
Overspeed,
D. Variations in pressure ratio across due to abnormal operating conditions. Specific gravity or density. Changes
in
a conrpressor
specific gravirr.
of the fluid medium handled by a oump, or
chanqes
in the densitv of the gas handled by a compressor. affec, the horsepo\ver transmitted in direct proportion. Cir boiler feed pumps, for example, this occurrence can be encountered during startup, upon malfunction of preheatine equipment, or during boiler cooldorvn follorvius tube failure. The overload can be appreciable-in tire order of 12-15 percent under average conditions. ar.id as high as 35-40 percent in some instances. Design horsepou,er for air handling centrifugal compressors is usually based on the normal maximum anrbient, frequently about 1000 F. On a rvinter dav, u.hen the temperature drops to 30o F, the density of the air varies with the absolute temperature, i.e.,560/,190 = 1.1.1. In still colder weather, say -20o F, the density increases to 1.27 times that of design conditions. Compressors handling other gases are usually encountereci in process systems under greater control, where temperature vari-
ations are less. Holvever, other variables may become
In refinery practice, for example, the cornposition of the gas car, vary widely, and in other process work the inlet pressure may not be a fixed value. Carry Out is an expression used by the pump and compressor industries to indicate performance on a head serious.
I
r- -r--r-
,,nrol,,*rr..o
curve beyond the so-called design point. Fig. 2 illustrates a typical compressor percentage performance curve. Note that at 100 percent speed, as the head drops off and the flow is increased, the horsepower increases to
.rto
^n
ffi
*"d1_
levels as high as 115 percent load. Carry out is an everyday reality. It comes about through such things as improper estimation of system performance during design stages, altered system requirements of*existing processes, gradual deterioration of processes, systems using half size of multiple units where shutdown or failures of one increases the requirements on the remaining units, or through leaks or failures. Overspeed is just what the names implies, and is obviously limited to applications with variable speed prime movers. Because the power absorption of the driven machine varies approximately with the third power of
Fig.
2-A
typical centrifugal compressor performance curve'
the speed, overspeed is a large contributor to overload. Referring again to Fig. 2, the performance curve indicates that at 110 percent speed and 100 percent flow, the horsepower is increased to 125 percent. Carry out at this speed can increase the horsepower still further, to levels approaching 140 percent of design horsepower.
Normal practice for a turbine-driven centrifugal compressor is to set the overspeed trips at 115 percent design speed. Governor settings are generally established to permit continuous operation between 105 percent and 110 percent of design speed. Keep in mind the fact that operators can and do reset governors to avail themselves of maximum output of the system, regardless of original settings. 2. Nqturql torsionql vibrqtion. Practically every mechanical drive system in operation is oscillating in one or more of its natural modes of torsional vibration. The amplitudes and corresponding torques vary widely-being a function of the proximity of forcing frequencies, amount of excitation, anC damping. If the frequency of an exciting force is close to, or coincident with, a natural frequency-resonance exists and the results may be disastrous. If the frequencies of exciting forces are sufficiently far from natural frequencies, violent vibration is avoided the system will still vibrate in its natural modes at -but low amplitudes, excited by random torque pulsations.
3.
Forced torsional vibrotion. Forced torsional pulsations result from blade passage, eccentricity of impeller scrolls, surge (both steady state and random), pipe strain and foundation movement as they affect housing and rotor displacement with the resulting loss of pressure balance, coupling misalignment, gear runout, apex wobble, motor air gap eccentricity, engine firing rate, etc. Recent testing on pumps has indicated torque pulsations between 10 percent and 40 percent mean torque. Compressors with vaned diffusers generally run about 3-10 percent. Where vaneless, or volute diffusers with large tongue gaps are used, the torque impulses disappear
for all practical
purposes.
The torsional vibration spectmm of most any mechanical drive system will show torque pulsations from most all of the above-mentioned sources simultaneously. Torque REPRINTED FROM HYDROCARBON PROCESSING
0 10 & m O S0 B0 7A
o0 .90,.100
FnEilE{CY (}tr)
Fig.
3-A
"real time plot" of a telem'etered strain gage slgnal
measuring torque.
in Fig. 3. The figure is a "teal time plo?' of a telemetered strain gage signal measuring torque. Of the seven more prominent spikes in this spectrum, three are natural frequencies and four are forced vibration. The total vibratory torque approaches the summation of the major discreet frequencies. In the absence of specific knowledge of a vibratory system, average vibratory torque summations for typical equipment and applications can be estimated from the following: pulsations are illustrated
o For motor or turbine driven
geared
sptems-5-l0 per-
cent mean torque
o For well treated
engine geared systems (soft couplings and dampers)-15-30 percent mean torque
o For stiff
4.
engine geared sptems-50 percent or higher.
Gommunicotion.
If
system operational problems are established between the "systems engineer" and the machinery
to be avoided, effective communication must be
builders (includes not only gear and coupling suppliers, but prime mover and driven machine as well). The various system ahalyses, in at least preliminary form, should precede detailed equipment purchase specifications. 99
frJtgh-speed gean unnHs
About the quthor
This sequence will permit the design to be based on a more nearly correct load and operating conditions. In the past, system engineers and industry in general, have relied heavily on the so-called service factors as recommended by gear builders to provide the necessary margin to accommodate overload. Gear builders had been maneuvered into the position of furnishing this kind of information through self-protection (gears Le frequently the safety valve, or weak link in the chain) . When a gear unit failed, a bigger one was provided until it worked. The difference in size became the service factor.
The more complex, higher power and higher
speed,
mechanical drive systems being designed today make this procedure and the reliance on "seryice factor" obsolete. Gear manufacture.rs simply do not have the technical depth nor the detailed information to adequately analyze system overload. This function must be performed by
SUMMATION OF VIBRATORY TOBOUES
Fnoomrc A. THoMA is manager, gear engineeri.ng, Delaoal Turbine, Inc., Trenton, N.J. He holils a B.M.E. degree from
Georgia Institute of Technologg. He hns a transmtission designer (machine tools) with both Gi.dding a Lewis and Warner & Swaseg. Mr. Thoma hns been with Delaaal ooer 20 Aeo,rs o,s a designer of marine and high-speed industrial gears. He has held positions as utorked, as
gear
eer
and chief engineer before He is curt'entlg chairtnon of e,r.s Association's High Speed Gear Committee and the Marine Gear Comrrti,ttee. He is also chuirmon of the API Marutfac-
turers Subcommi,ttee on High Speed Gears.
Neither of these misalignments can be readily calcu-
Iated, but they warrant discussion so that the "svstems engineer" can take precautions to minimize their effect. "Flexible" couplings, whether of the dental tooth. spring element, flexing disc, or elastomeric type, produce forces and moments on their supporting shafts when operating misaligned. The analytical determination of the ma_enitude of these forces and moments is not fullv understood. It can be generalized that:
A. The sense and direction are such to try and bring the supporting shafts in line. B. Bending moments have been measured in supporting of the torsional
shafts that have exceeded 25 percent moments transmitted. ESTII/ATEO MAX. CONTINUOUS TOBOUE
C. That the magnitude of the forces and moments inwith larger angularity across the coupling.
creases
D. Notrvithstanding catalog claims for angular capacitv, flexible couplings should not be looked upon as unir.ersal joints; they should be given the best possible alignment. There are several general rules which will help the designer obtain better alignment of the coupled shafting Fig.
l-1rr;"al
mechanical drive system torque components.
specialists under the responsibility of the systems engineer.
There is no set format for communicating these data. The required information is the magnitude of overload and a description of the operating conditions under which occurs, such as when, how long, and natu.re (i.e., steady state, vibratory, etc.). Fig. 4 is a graphic illustration of the combined eflects
it
of maximum torque plus vibratory torque as compared to the nominal design point torque for a centrifugal compressor operating at overspeed and carry out.
5. Misolig,nment while operating produces overloadsnot in the sense that increased torque does, rather in the sense
of increased
stress
on one or more component parts
for a given torque. There are two general (and interrelated) kinds of misalignment in geared systems. The first concerns shaft end misalignment across the couplings or both, high and low speed. The second con-either cerns misalignment of pinion and gear axes-brought about primarily by foundation or bedplate twisting. 100
system:
o Make
as comprehensive an assessment as possible of
operating alignment. This is a system study and must include all elements of the system including bedplates and/ or foundations. An accurate evaluation of thermai grol.th for all components from a valid and common reference line is required. Journal displacement within bearinqs. though generally smaller in magnitude, should be considered, particularly as it affects cold or static alignment checks. The forces and moments imposed in pumps, compressors and turbines by their inlet and discharge piping are major factors in deflecting this equipment and causing
operating mal-alignment. All efforts should be made to minimize piping effects. After determining the probable magnitude of alignment change from static and cold, to dynamic and hot (including any periodic cyclic changes that may occur), select a coupling arrangement that provides enough length or span between flexible elements to keep the angularity low-in the region of 6 to B minutes or lower.
o
Reinforced concrete foundations with grouted-in sole-
gene.ra plutes in ter-ms of iion of adequate piling, is the best plates are
fabricated steel bed-
A concrete foundasoil or on sufficient d unequal settling or s.
twisting from other sources.
Fabricated steel bedplates make convenient shipping
and handling frames, but are generally designed for strength, not rigidity. They are frequently designed so that the various piping andf ot oil sumps cause unsymmetrical thermal expansion. Out-of-door installations on steel bedplates are particularly subject to cyclic bowing caused by the daily rise and fall of the sun. When steel bedplates are used, the designer should endeavor to achieve two things: First, he should arrange oil
weather protection to minimize unsymmetrical thermal expansion. Second, he should
sumps, piping and
thoroughly investigate elastic deformation of the bedplate
caused by piping forces and moments-then design the bedplate to eliminate twisting at the gear suPPorts'
o Last, but certainly not least, the "system designer"
should issue complete and comprehensive installation instructions covering, as a minimum, such things as:
1. Soleplate, bedplate, machinery position and leveling detail,
2. Foundation bolting and grouting detail, 3. Cold alignment data-including method of measuring, relative position and sequence of alignmen! 4. Keying, pinning, and torqueing detail as required, 5. Pipe support and flange makeup details. 6. All other relevant details that would otherwise left to the judgement of the job site mechanic.
be
r
Testing can reduce gear failures Shipley, Mechanical Technology, Inc., Latham, N.Y. E. E.
Trrs cnen vENDoR knows from past practice horv his gear units should perform in the field from an internal heat and deflection point of view. BLrt relying wholly on past performance is not justified since many costly gear failures do occur. It appears that there is room for shop testing-the type of testing that trulv brings out the weaknesses of a given design and its associated hardware. The more test effort applied at the factory, and the closer the factory test is to actual field operations, the less trouble should be experienced in the field. Full load torque testing at design speed (back-to-back) in the factory comes the nearest to duplicating field practice.
It high reliabilitv is the first order of preference, then of the full load, full speed veriety may
nrore test work be the answer.
Cquses of geor fqilures. There are four general causes oI failure that are often identified without question as the prime reason for failure. The approximate percentages of thcse failures, from actual field investigations covering a five-year period, are: 50 percent heat and lubrication ofterr in thc form of scoring 15 Perccnt contact str-ess su|face
pitting
15 perccnt wear-metal removal rorn the sutface more f
or less ur"riformly 10 percent bending stress-gear tooth breakage 10 perccnt metallurgical.
The "metallurgical" category has becn included to indir:ate that sometimes the prime cause of failure can be REPRINTED FROM HYDROCARBON PROCESSING
traced back to poor metallurgy such as dirty steel, irn. proper heat treatment, poor microstructure, segregation,
in the final analysis is not directly a result of design but rather poor quality control of the material procurement and heat treating Process.
etc., which
Heot ond lubricqtion. Since the category "heat and lubrication" accounts for 50 percent of gearbox failures, it rr,ay be interesting to consider this problem in a little more detail. It is also a problem which could be solved by shop testing. The heat loss of a gear mesh can be calculated from: (1) geometry of the gear teeth, and (2) the steady-state load being transmitted. Mesh loss is a prime source of heat and must be taken into consideration when the lubrication requirements are specified. In high-speed gears, there are other important sources of heat generation which may lead to gear mesh difliculties and must be considered carefully. Examples are: 1. Windage and churiring losses within the gear casing, 2. Spray patterns and oil distribution, within the casing, 3. Improper drainage of oil from the gear casing, 4. Improper use of baffies and gear shrouds, 5. Loss of backlash caused by heating of gear rotors, 6. Excessive oil flow into'the gear unit, and 7. Poor surface finish on contacting surfaces'
It may be interesting to see how heat affects a pair of meshing gears. Fig. 1 shows what happens to high-speed pinions when they increase in temperature operating under load. The lubricant carries away some of the heat but the remainder is carried away by conductionl and the net result is that the center of the pinion may operate at a higher temperature than the ends of the pinion. The 101
As the pinion or gear, for that matter, increases in
ffitgh-speed geanuntHs
temperature, the tooth also changes shape. Fig. 2 shows
bottom pinion shows the exaggerated growth experienced when the pinion is cooled or heated nonuniformly. This
condition is called "barrelling" and is a rather common occurrence.
Although Fig. I shows a double helical gear as an example, this conditiou also applies to a single helical gear as
well.
It is quite easy to understand how barrelling can concentrate load near the apex or middle of a double helical gear. Load concentration tends to increase the temperature even more. It is not uncommon at all to find the apex of a double helical gear scoring locally.
a tooth cross-section before and after heating higher temperature level.
in to
up to
a
:Ti: X,:",:fL"'5:J"H:11 temperature.
2. The top dotted curyes show an expanded pinion tooth profile after an increase in operating temperature. 3. The pinion grows more radially at the outside diameter than at the base diameterl this growth has about the same eflect as changing the base pitch of the involute gear teeth.
in tooth from engin
ce
roaded gears
in
action.
ear
It
or
for
5. An expanded pinion will cause heavy root bearing with its mating gear profile. It wiil also have a tendency to mesh deeper into the root section of the gear.
6. The expanded pinion may also help to initiate an early scoring type failure because of the tendency for high load concentrations in the area where sliding velocitv is the highest. Because of the tendency of the smaller gear member (pinion) to run hotter than larger gear members (gear). it is customary to make some initial adjustment during the design stage to help counteract this phenomena.
Flg. l.-Pinion, barreling is seen as exaggerated growth when the pinlon ls heated or cooled nonuniformly.
highest temperature of the meshing pair. If, on the other hand, the installation requires a gear reduction, that is, the pinion driving the gear in a step down mode, the pinion should be driving the gear with the apex of the pinion leading. Most gear units for the refining and petrochemical industries are purchased to API standard 613. This standard states that all gear units shall be given a mechanical
'-trx6x10-'raTl
of this nature.
(,o
l.+r-
lt
-
PITCH
= ows oe nno us
mlI$
EASE RAotuS
-t
Fig. 2-Tooth cross-section before and
higher temperature.
102
Torque tesl-slow roll. All gear units should be given a no-load spin test as a minimum. However, there are other types of gear tests that reveal more about the potential field performance of a given gear unit. Additional tests must be evaluated in terms of the reliability required
l I
I I I I
after heating
by a given customer for a given application. Some thought
to
a
must also be given to delivery schedules; that is, if you desire more testing, then more time is required. A slow
roll torque test will produce additional performance rnformation. One of the important considerations of any gear unit is to assure that the gear structure is strong enough to support the load. In parallel axis gearing, such as is used on double helical gearing, the main objective is to keep the shafts parallel. If the gear casing deflects, it must deflect uniformly. The torque test at slow rol1, as shown
in Fig. 3, permits you to study the contract pattern of the gear teeth at any desired load condition. Very often, just ohserving the contact at full load is not sumcient since you rnust know rvhat the contact rvas like at 50 percent torque to adequately evaluate the 100 percent
r{oi-of,irN
*^:'^.':a/
CLAMP6
torqLle conditions.
The slorv roll test as shown in Fig. 3, consists of a production test gear unit. The input and output shafts have bcen connected to the rotor shafts of h1'draulic r.otary actuators. Each actuator will make onl1' a perccntage of a full revolution. The stator portion of the actuaior is connected in each case to a fixed base pedestal.
1'he gear is loaded in the follorving manner: highoil is fed into the actuator on the output shaft rvhilc at thc sarnc time, high-pressurc oil is fed into the actuator on the input shaIt. In effect, the trvo shafts are rotatcd against each other under the desired torque conclitions bv carefully adjusting oil pressures. At this tinle, the gcar rrheels are not moving. The tooth surfaces, on one member, are painted with prussian blue and the other member is painted with red lead (this is a conmon method of checking contact statically). Now tliat the input ancl output shaft lias beerr torqued agair-rst each other, at 50 percent torque for example, the contact can be rolled out simply by either decreasing the output torque s1ightl1' by reducing oil pressure on the output actuator or bf increasing the input torque slightly by incleasing oil pressure on the input actuator' After thc gear mesh has progressed through several sets of teeth, the actllators are stopped and a 50 Percent tooth contact has been rviped out on the teeth. The contact pattern can be lifted by tape and placed in a book for record purposes. The contact pattern for any desired Ioad can be determined in this manner. By close obser\rance of these contact Patterns, you can determine if the gcar casing is moving under pressurc or just how the load is deflccting the teeth into a load carrying
pressure
position.
Fig. 4-Back-to-back torque test using two production gear speed increasers.
testing is used. This is basically a locked torque test set up simulating full power operation. It is by far the most involved test procedure and it also produces the best over-all test results, since it simulates field conditions almost 100 percent. The duration of the test must be limited but it is usually run a considerable length of time.
Fig. 4 shows a locked torque test set up of two production gear speed increasers, using any type of prime mover as a driver. The object of such a test is to verify that production gear units will meet the design specifica,
tions, at full speed, with loads comparable to those they will be subjected to in field operation or even higher.
The test is conducted by mounting two identical drive gear units adjacent to each other, back-to-back and connecting the nligh-speed shafts together. The torque in the gear mesh is achieved by reverse loading the low-speed
coupling with large lever-type torque wrenches (see Fig. 5). By attaching a force measuring device to the lever arrr,, a reading can be obtained indicating the exact torque in the system. With the desired torque achieved,
chanical distortion due to centrifugal force acts in a sirnilar manner as thermal distortion and must bc considered uhen applopriate.
the high-speed couplings are bolted in position thereby locking the torque in the gear meshes similar in nature to a wound-up spring. Since the prime mover oftentimes does not have sufficient torque to start the loaded system from zero speed, a special starting-motor-gearbox arrangement is necessary. This consists of a low-horsepower motor driving through a high-reduction speed reducer clutched to the test unit. When the test stand is running at slow speed, the driver is started and the starting-motor-gearbox clutch is disengaged. The torqued units are then accelerated by the prime mover until the desired test speed is obtained (the gear units are now running at an equivalent horsepower rating).
Torque lesting bqck-to'bqck. To full load test highhorsepower gear units, the technique of back-to-back
for workmanship to insure that correct assembly pro-
A test of this nature is very beneficial in determining rvhat can be expected from the gear unit at full torque provided, of course, that the usual full speed no-load test has been run. The major '"veak point in tests at slow roll is the fact that the gear units does not evah-rate the high s1;eed distortion of the gear wheels, if an1 , and the temperature characteristics of the assembled gear unit.
As stated previously,.the heat generated by thc rotating gears call cause considerable distortion to the rncsh and cause unever-r loading across the teeth' Nonuniform me-
REPRINTED FROM HYDROCARBON PROCESSING
Pre-test. Prior to testing the units, all parts are inspected
103
ffifrgm-W@@d
and 414 torque. Suficient running time should be permitted so that the running contact at these load levels can be observed and evaluated. Once full torque operation has been approved, the test operation should continue until the slowest speed element in the gear unit has accumulated 10 x 106 fatigue cycles.
gean wnnfis
Acceptobility limiis. If you torque test a gear unit or simply spin test it, here are some recommended limits oj acceptability. Some of these limits are overall appraisal limits pertaining to design and arrangement while others pertain to acceptance after testing. In addition to meeting the general specification vibration and noise limits, a close examination of the gear tooth contact must be made. 1. There should be no evidence of scoring. 2. If the gear units are subjected to high-speed runs (light torque-load) the acceptable contact pattern after test should be 80 percent contact across each helix rvith no obvious heavy end bearings at either the apex or the base. The only exception would be on a high L/D ratio design where the helix angle corrections were machined into the teeth to compensate for twist and deflection under
Fig. 5-. Applying. torque to the low-speed coupling with large rever-rype wrencnes.
cedures have been followed. This final check is performed
in addition to the normal quality control inspection provided throughout manufacturing. After manufacturing, the gearing is located and rotated in the housing manually to note proper gear-tooth patterns.
Each housing is flushed with hot oil before final
as-
sembly, and then checked for cleanliness.
Test stand set up. Gear units to be tested are mounted on a special foundation so that they may be adjusted to the proper height of the prime mover output shaft. Lubrication is supplied from lube consoles having sufficient capacity to supply oil to the units at design specifications (Note: both the proper oil flow at the proper inlet temperature.
)
Thermocouples are installed at bearing positions and temperature gages are placed to read the inlet and sump oil temperatures. The oil flow rate is measured at the in-
Iet with an oil flow gage. Load test operation. The entire system is started using the high-torque starting motor reducer. As the gear train is broken loose and begins to turn, the friction drops ofi, the prime mover is started and the clutch disengaged. The test gear units are then accelerated by the driver to full test speed. Speed is attained over a period of about two hours in discreet steps. During this acceleration period, the lube system pressures, temperature and flow, as well as bearing temperatures, are carefully monitored and adjustments made to assure proper operation of the system.
When a full-load torque test has been specified, the duration of the test must be detailed along with many of the other variables. As a minimum, the load level on the gear units should be run and inspected at l14, 2l+,314 104
full
load.
3. If the gear unit is subjected to a high-speed, fu1Itorque-load run, the acceptable contact pattern after test would be 95 percent contact across each helix with no obvious heavy end bearings at either the apex or the base.
4. The total heat rise across the gear unit based on oil flow and temperature rise must be within the vendor's estimated total heat loss. It also should be evaluated against other gear units of this nature as a double check on the
gear vendor's calculation. 5. Backlash-Gear rotors are designed with backlash to compensate for: center-distance tolerances, deflection under load, thermal expansion of rotors and casing, as
well
as variations in tooth thickness due to manufacturing. The backlash is usually checked by mouirting the gears at the proper center-distance, fixing one gear, and rocking the other back and forth against a dial indicator arranged tangentially at the pitch circle. For helical or double helical gearing, the backlash is measured normal to the teeth rvith either an indicator or feeler gages. The backlash can also be measured by holding the gear stationary and measuring the total axial movement of the
pinion.
If meshing teeth interlere with one another, the temperature inside the casing will increase rapidly creating more tooth interference which will quickly cause excessive damage. A gear must have enough backlash to operate satisfactorily under transient and abnormal operating conditions. Some conditions that may require more than the minimum backlash recommended by the gear vendor are as follows:
. If more oil is permitted to enter the gearbox at the job site than on the test floor, o If the horsepower load is built up rapidly on the gear teeth before the gear casing has an opportunity to heat uP,
o
If
the oil tends to foam as the bulk temperature of the
oil is
increased,
o If a partial restriction of the drain causes a disturbance to the internal oil distribution, o If the gear elements tend to barrel in the middle because of excessive heat buildup in the center, o If a higher viscosity oil is substituted in the field for th: normal test gearbox lubricating oil, o If a back pressure builds up in the drain system. The drain system at the job site may be much longer and more complicated than on the test floor at the factory,
a,
Massachusetts.
Method of manufacture Pinisn s1 gttt
Recommended design limits
Pitch line velocity
.
The following recommendations may be helpful:
,Fabrication OK (forged rims welded in webs) Forgings with integral shaft OK
o Specify a false bottom on the gear unit so that the oil falls from the turbulent area into a quiet sub-reservoir and then drains away.
Shrunk on rim not Permissible
o Specify an extension to the bottom of the gear casing so that the oil level is at least two feet from the bottom of
Disk forging pressed on shaft OK
is very Iow and vendor calculates combined unless stress
ft.,/min. ft./min..
Between 15,000
and 30,000
is a member of AGMA' registereil professional engineer in
assumi.ng his present position. He
ASLD, ASME and,'is
6. Gears and pinions'
Under 15,000 f.t./min.,
About the quthor E. E. SHrPr,Eu is manager-mecho'nical transmissions, Me-
Forgings with integral shaft OK .Disk forging pressed on shaft OK Fabrication OK if gear vendor has had previous experience and calculates deflection and weld stresses Shrunk on rims not permissible Casting generally not permissible
unless considerable
successful demon-
experience can be strated by the r.endor.
7. Gear casine design-The gear casin.o; shoulcl bc stifl enough to support the gears and bearing so that their shafts remain parallel. The casing also must act as a con-
tainer so that oil is not flung around the area. Many of the difficulties rvith internal oil flow in gear units results from the fact that not enough space has been left to adequately drain the oil away. High pitchline velocity gears are particularly subjected to this difficulty. Below 10,000 ft./min. (Windage is generally not strong enough to prevent adequate drainage, even in a tight fitted gear unit).
the large diameter gear to escaPe windage effects. A shield should'be designed over the drain so that windage cannot directly hit it.
o Check the design for ample room at the top part of the gear casing so that oil has a chance to hit the end walls
and flow down relatively unrestricted'
. In gear units that appear to be critical from a heating point of view, specify an upward meshing gear- This arrangement has a tendency to form windage paths that help to scour the walls of the casing in the direction of the oil flow down the walls and into the sump.
. The gear unit must have a vent and the drain line {rom the gear unit must also be vented. Often it is desirable to specify that two drain connections be built into the box. Another precaution is to specify that the gear unit drain line to the main reservoir be independent of all other drains. 9. Oil nozzle-The location of the oil nozzle has been a controversial item over a long period of time. Ilowever, the following table can be used as a guide: PitchJine velocity Berow I 0,000 ft. / mh... . . . . It
10,000-15,000 ft./min. (Windage effects can be felt, hor'vever, drainage can be accomplished with little diffi-
rfjTi,l,tj9,"utf:'?ffi",#T;:
as a sufficient
iifi'""ft[.1;
culty).
g side is more
ft./min. (Windage is often r, problem, particularly with gear units that were designed for lorv speed and have been uprated with little, if any, consider-
effective'
15,000-20,000
10,000_15,000
rt.,/rnin.
20.000-30,000 ft./min. (A gear unit must be designecl rvith the rvindage cffect in mind so that proper pror-isions are made for adequate removal of all cooling and lubricating oil). B.
Drain
considerations
for high pitch-line
15,000-30,000
ft./min. .....Oil
oil quantity should
be B5/o for cooling. In addition, oil nozzles on the ingoing side directed slightly off in-
going mesh, the oil quantity
velocity
ft./min.). The drain must be protectcd from the turbulence of the windage from the mesh or arranged so that the turbulence helps to lorce the oil in the proper direction for drainage.
nozzles on the outloing side directed into the outgoing
mesh, the
should be lSVo 6or lubrication.
gears (20,000-30,000
REPRINTED FROM HYDROCARBON PROCESSING
"i#f,$ufT:TuJiu,l:".,]; squarely into the mesh.
ation for internal lubrication flow).
Abstracted from
ACKNOWLEDGMENT speed geu testing and evalution," origi-mhy
"High
at the Thirt Compressor Train Reliability Symposium, i".turins Chemists Asseiation, Chicago, April 24, 1973.
n.esented
Manrr-
I
105
[iln9ffi-se[!)@@d g]@aLe
Shipping and installation specs fllqrk W. Becrd, Fluor Engineers & Constructors, Inc., Los Angeles Srrrpprrve rNSTRUcrroNs should be considered as two separate problems. First is "Preparration for Shipment."
Should
it
be shipped assembled or knocked down? What sort of shock loading is likely to be encountered? How can the equipment be protected from the elements? Second, how should it be crated or boxed for the common carrier's requirements? W&rile the "manufacturer standard" is often adequate, it is foolhardy to rely on his best guess of your field storage and installation environmental problems.
Covered storage is rarely possible and coastal environments are extremely hostile during long storage periods which often exceed 12 months. Personalln I favor preservatives which are compatible with the lube oil to be used and do not require disassembly of the gear before placing the unit in service. I also insist on shipping the unit fully assembled to minimize field installation problems.
Exporl boxing ond shipping poses special problems, depending upon the type of vessel. We have used sea-land containers, converted LSTs and ocean going barges
as
well as conventional freighters. Equipment is usually stowed in cargo holds but frequently it is lashed on decks and is subjected to ocean spray. Companies specializing in export boxing are available in m,ost major ports and
cerning maintenance, trouble shooting guides and overhaul. They are well illustrated and include pictures of damaged bearings and teeth to guide the neoplyte. Startup checks. Installation manuals discuss the routine checkouts required immediately prior to startup and are very useful to the startup engineer. These incluC.e as a
minimum: 1. Check the oil level and type in reservoir. 2. Check the direction of rotation since many
3. 4. 5.
6.
ing initial loading. During startup
Check for unusual noises and record vibrations fron'r minimum speed to maximum operating speed. ) Monitor pressure drop across filters. Ner,y filters should be available during startup for changeout. 1.
J.
Oil
+.
After about 30 minutes, the gear should be shut dou,n and the case opened to inspect gear tooth contact. A final shutdorvn for gear tooth inspection should be made after full load operation and stabilized oil tem-
are recommended.
After the gear has arrived in the field, a routine inspection procedure should be established to protect the equipment from atmospheric breathing, dust, and mechanical damage. All temporary covers and screens should be carefully tagged to be sure that they are removed before startup. Our only gear failure occurred when temporary screens in the oil supply lines adjacent to the bearings were inadvertently left in place when the unit was placed in service.
-gears are
uni-directional. All shafts must turn freely. The engineer should verify that cold shaft alignnent. have provided for thermal growth of all equipment. Foundation bolts must be tight and casing pr-operh, shimmed to prevent casing distortions. Three spots around the gear should be cleaned arrd coated r",'ith Prussian Blue to check tooth contact dur-
5.
temperatr-rres should be watched closely.
peratures have been established.
Vibrqtion monitoring. For critical installations rr-here downtime is extremely expensive, a good vibratioli n-ronitoring system is cheap protection. There is no question that failures can be minimized or avoided b1, uiing vibration monitoring equipment now available. I still far,or
lnstqllotion. There are several excellent gear installation manuals available from the manufacturers. They cover detailed instructions on :
o Handling o Mounting of couplings
.
Foundation and alignment
o Cold and hot check alignment o Axial positioning o Doweling o Tooth contact pattern checks In addition, the manuals contain 106
computer. Based on introductory
excellent details con-
gear units, ASME petio seles, Sept. 18,1973.
to a paDel session on Cortjre."", -* fo.- '-l e"-
Turbomachinery Problem Diagnosis
r
'ilr i
1-l .I ,l
f-+:
a-q:,S
A new approach to turbomachine ry analysis Large process plants containing many rotating machinery systems present a unique data collection and analysis problem. How do you physically get to all the machinery with the necessary instrumentation? Here is how one company solyed the problem Dole Lorio, Shell Chemical Co., Deer Park, Texas
440-volt extension cord rnounted on a reel in the rear of the truck (Fig. 3). Electrical po\yer is obtained from welding receptacles located throughout the process units. The analysis r.an is primarily designed to be used in conjunction u'ith noncontacting eddy current vibration probes but is b1' no means limited to this type of transducer. Velocitv transducers, accelerometers, miclophones. and pressure transducers have all been used rrith equal success. Normalll', the anall'sis van has to be parked from 50 to 100 feet from the machine being anah'zed. lloiveve.r,
ExcessrvB Ir.{TNTENANCE costs resulting from high vibration levels on rotating machinery can be significantly reduced if problems are quickll' detected, analyzed and corrected. Shell Chemical Company's llouston plant accomplishes this by using a mobile analysis unit consisting of an oscilloscope, vibration monitors, tape recorder, spectrur)t analyzer and an X-Y plotter. The use of this unit has .resulted in improved efficiency, better operational performance and increased life of the rotating machinery by decreasing the frequency and severity of failures. For vears 'r,ibration and noise levels have been useful indicators of machinery condition. But vibration and noise
levels, or amplitudes, are primarily of value only as indicatols of over-all machinery condition. In order to diagnose the source of unrest within a given machine, more specific information is required. Vibration frequency, noise frequeno', phase angles, shaft orbit patterns, vibration amplitude as a function of speed, and slow roll shaft motion levels are only a few of the parameters necessary to form a conclusive diagnosis. To obtain all the required data and do it quickly and efficiently, Shell Chemical Company realized that more sophisticated equipment, used in conjunction with existing vibration equipment,
the high frequency detector-drivers for noncontact
vibration prbbes urLrst be mounted near the actual ltrobe
"*.
t*--.
{
1
0
l:
LIEOIIIORY
ffiTltilICAL
I Fig. 1-Mobile machinery analysis unit.
EErrtrrI
was mandaton'. A self-contained mobile
unit that could be driven to and located near the machinery to be analyzed was selected for this service. Fig. 1 shows the van truck that is currently in use as a mobile machinery analysis unit. The van is equipped with a one-ton, roof-mounted air conditioner and a dehurnidifier to maintain constant temperature and humidity levels within. The interior of the van was modified from its off-the-shelf status to accept the instrumentation, rvork desk, lights and circulation fans (Fig. 2). Pou'el- to the instrumentation, air conditioner, fans and lights is supplied through a 440lll0-volt step-down transfolmer mounted in the rear of the truck. The power system includes a
Power 108
circuit breaker and distribution facilities.
is brought into the truck through 200 feet of
Fig. 2-lntbrior work area of mobile analysis unit.
.
TURBOMACHINERY ANALYSIS
Two vibration monitors which include low-pass filters for removing high frequency noise
oA
dual-channel X-Y oscilloscope, either directly or indirectly through the lorv-pass filters
Thus, it is possible to simultaneously monitor two points on the vibration meters, two points on the oscilloscope and a fifth point on the spectrum analyzerfX-Y plotter. Or, all points can be fed directly into the tape recorder and selectively analyzed at a later date. The oscilloscope was initially incorporated into the van to be used as both a general purpose test instrument and as an aid to interpreting and/or analyzing vibration signals. The dual-channel feature of the oscilloscope permits orbital analyses of the shaft on any of the journal bearings equipped r'r.ith both horizontal and vertical vibration probes. The oscilloscope is also used in conjunction rvith an engine/compressor analyzer not included in
the van instrumentation.
The spectrum analyzer has the capability of operating in three distinct modes:
.
Spectrum analysis
o Rpm
.
tracking
Signature ratio
Primarily the unit is operated in the spectrum analysis fn this mode, the analyzer receives the vibrationtime signal from a machine operating at a fixed speed and converts it into a vibration-frequency sigrral. This signal is fed into the X-Y recorder which traces vibration amplitude versus frequency (Fig. 5). The analyzer automatically sweeps through a preselected frequency range. In our case, this is generally from 6 through 1,666 cycles per second for the high-speed compressors and turbines. The spectrum analyzer accomplishes the same thing as the tunable filter on a conventional vibration atalyzer. Ilowever, it has the advantages of doing it automatically, quickly, provides a permanent record, and has greater mode.
Fig. 3-Electrical power distribution system.
sensitivity.
Fig. 4-Field mounted housings and interconnecting signal cables for proximity probe detector-driver devices.
Iocation (10 feet ir.r otrr case) . To accommodate the situation, the analr.sis van is equipped rvith a 100-foot input cable and field boxes. The field boxes house the high frequencv detector-drir.er devices for the noncontact probes and are coml;lete rrith the 1O-foot interconnecting cable (Fig. 't) . This allorr.s the field box to be connected directli, to the probe lead. The input cable connects to a receptacle in the side of the analysis van. From here the sisnal is fed into a series of palallel selector srvitches. The signal input cable supplies -18 volt DC po\'ver to the detector-driver der.ices as \\'ell as routing the signals back to the van. 'I'he selcctor srvitchcs cnable the operator to route the incoming signals to:
. -\ series of signal conditioning amplifiers o ,\ ser.en-channel F\I tape recordcr for data storage o Spectmm analyzer for anall'sis and recording on X-Y recorder REPRINTED FROM HYDROCARBON PROCESSING
The rpm tracking mode of operation is also useful. In this mode, the analyzer will track changes in speed and plot vibration amplitude versus machine speed (Fig. 6). When used during startup, it enables the operator to
pick out critical speeds and provides him with an excellent indication of whether the vibration is a result of imbalance. T'he mode is also very useful in establishing overspeed trip points. Signature ratio analysis is used primarily on nrachines whose operating speed fluctuates. This includes gas engines, diesel engines, jet engines and machines u,ith poor speed control capability. Fluctuations in machine speed causes the recorded frequency of a vibrating member not to be in its true ratio with respect to the main shaft speed, i.e., high speed pinion with respect to lor,v speed bull gear on an engine, speed increaser, pump train. This makes interpretation of the results difficult and time consuming. For this reason, the signature ratio mode records data directly in vibration amplitude versus ratio of this component to fundamental speed. The relationships are automatically maintained at the same value regardless of the speed setting or fluctuations in speed. The seven-channel tape recorder provides the capability
an
of storing input data in raw form. These data can be played back at a later date and transformed into X-Y 109
plots. The recorder allows rapid accumulation of data on startups and other instances when operator time is at
SPECTRUM ANALYSIS
a premium.
.t+
I
r,2m R.P.t
Uses. With the equipment contained in the van, possible to detect malfunctions such as:
o Imbalance o Oil whirl a Resonance o Critical speeds of component
it
._
_Lt+
is
I8,MO RPJ
MACHINE-9100 R.P.M. COMPRESSOR 1. 4,200 R,P.M. VIBMTION WAS CAUSED BY OIL WHIRL IN THE SEAL BUSHINGS. 2. 18,200 R.P.M. VIBMTION RESULTEO FBOM MISALIGNMENT.
R .t.
t
paJts as well as that of
the main shaft
o Misalignment o Mechanical looseness o Gear problems o Hydraulic effects
Fig.
5-X-Y plot of vibration amplitude versus
frequency.
Analysis techniques are the same as those with conventional equipment. The van is employed on startups of all major equipment after overhaul. Such utilization
a "fingerprint" for future comparative analysis. fn addition, all maj,or equipment is analyzed periodically (quarterly at present) to establish trends and pinpoint any potential problem areas. Examples of problems deprovides
7f-l-14.P.1u. -HAtrilG-If,8tilE
tected and diagnosed using the analysis equipment are
-
as follows:
o
Tn 0(l1s
tmE-
1. VIBRATION
INCREASES AS A SOUABE I )F THE SPEED INDICATING IMBALANCE. THIS PARTICI ,I.AB PBOBLEM WAS CAUSED 8Y A THROWI\ BUCKET.
lllllrl
Imbalance-established on a high-speed steam turbine using the rpm tracking mode on shutdown. High vibration levels were indicated and the plotted curve was a function of the square of the speed. Disassembly
verified the diagnosis; the turbine had lost a blade assembly closing block.
o Oil whirl-establjshed when a
significant peak was found at approximately 43 percent of rotating speed on a barrel-type centrifugal compressor. It was found that oil whirl was occuring in the seal bushings rather than in the bearing. The vibration amplitude was reduced to an acceptable level by machining pressure dams in the seal bushings and by increasing the shaft to seal bushing clearance.
o
Mechanical looseness-established on the same steam turbine mentioned in the imbalance example. Vibration amplitude detected a two times shaft speed was approximately double that at shaft speed. Disassembly
About the quthor Dlr,p Lonro is a, seni,or maclrinety
ca,tion stuclies, deaelopment and itmrylementati,on ot' alignment teclmi,ques and machinerg malfu,nction data collect:ion,
110
e,
T,,-inid,q,d,
and bearing bracket. Temporary addition of shim stock provided the proper bearing crush and elimi-
nated the conditiorr.
o
Reciprocating compressor pulsation-verified the need for pulsation dampeners by using pressure transducers
in conjunction rt'ith velocity vibration transducers. Pressure pulses were detected at 2 percent of peak pressure. In some cases, this level of pulsation might nant frequency of some attached piping. Under these conditions, the pressure pulsations aCt"d as the driving !or9e for high-level resonant vibration in the piping-. Serious damage could have resulted had the pioit"* not been accurately diagnosed and quickly corrected.
machinerpy, Supporting o,r eas i,n cluile sound mrcasut'ement ond, analysis, lubri-
Colleg
indicated excessive clearance betr,yeen the bearing shell
compressor, the pulsation frequency matched the reso-
neer with Shell Chemical Co., Deet Park, Teras. He is responsible f or application, selection ond troubleshooting of totating
State Junior ASME.
macntne spee0 cnanges.
be considered acceptable. However, on this particular
engi-
,reduction and, analysis. Pt"i,ot erperience 'includes maintenance engi,neering and project engineering u)ith Shell Chemical. B.S. degree in mecholvical engin Unitereitg, an A.A.S. degree ,in
Fig..6-RPM tracking mode showing vibration amplitude versus
The mobile analysis unit is an extremely Mr. Lorio holds a
valuable
machinery malfunction diagnostic tool and r,r,ill become even more valuable as newer techniques and instruments are developed. This approach to machinery analysis has proven to be a formidable method for greatly reducing maintenance costs, improving stream factors and reliability and preventing catastrophic failures. I
Optimize your vibration analysis procedures Many existing vibration analysis procedures are inadequate. Either too little or too much data are available. Here's how to match analysis instruments to machinery behavior knowledge and reduce maintenance costs in the process Chorles Jockson, \Ionsanto Polymers and Petrocht:micals Co., Texas Citr'. Texas
A rotrrar, \'IBR.\TIoN analysis system can save millions in rotating nrachinery failures. The type, speed, and critical nature of the equipment should dictate the vibration program needed. There are many spin-ofls from such a system, e.g. balance equipment and procedures, alignn-rent equipment, bealing gauging arrd maintenance procedures. The program outlined is for a plant employing seventeen hundred people rvith 80,000 plus horsepou'er in compression equipnrent. 1200 pumps and other special equipn-rent. 1'1,000 horsepou'er is the largest single component size, rvith 60,000 horseporver in turbomachinery at speeds up to 52,000 RPfI. This paper covers the back-
ground, present s1'stem, practices, and some recent experiences of a svstem in existence for 10 years. Four engineers and one engineering aide (average engineer experience is trventv vears) make up olrr analvsis groltp. Prior to the use of formal vibration analysis techniques, a vibrating lieht vibi-ometer) a nickcl, or sensitir-e hancls \A.ere our main analvzers. The first instrument purchased was a hand helcl vibrograph. It gave us a pressure sensitive tape of the vibration waveforrn for records and frequency/displacement anal-
Group under Engineering Services, we intended to simplv define problems by proper measurements' A program of 90 or 180 da1, vibration surveys r'vas established for critical rotating n'rachinery. The scheduling was complicated by an equal amount of emergenci, t1'pe inspections. Historical data files were established. These files contained pure ancl manually tuned vibration spectra and unfiltered velocitv level trend charts. These trends were recorded at three positions in the shaft bearing area, horizontal and vertical in the radial direction, and in the axial direction (paral-
lel to the shaft) . The trend charts are helpful because they illustrate horv far graclual excursions are drifting o\/er a given period of time. The chart shou's the diminishing health of a machine. The pure (unfiltered) velocity data is the most reasonable plot to make. If one n'ill recall simple harrronic motion, peak velocity is the only product tel'm at the first power of frequency and peak amplitr.rcle. Displacement
Velocity
:X:A
(sinwt)
: * : Aw
Acceleration =i:Aw'
(cos
-A
wt) : Arv
Peak
Value Peak
\/alue
(sinwt) =-rv'? X
Peak
Value
T'his t,vpe of analyzer evolved into the current lightr,veight solid statc analyzers with frequency meter. displace-
br--
,vsis. It rvas totallv
mechanical, small, lightweight and could be used on housings, structures, or directly on shafts rvith a "fish tailed" shaft stick. The tape can easily be copied
for reports. This instrument is still in continued use. It is most useful in confirmation of non-periodic rnalfunctiorrs sLrch as oil rvhirl. Our first electronic vibration arrd balancing analyzer used a seismic, or velocity, transducer and cost about $2800. This unit put us into the preventive maintenance business on a large scale. I might add that that was not the original intention. As a Nlechanical Development REPRINTED FROM HYDROCARBON PROCESSING
G
Fig. l-Labyrinth seal rub on a centrifugal compressor. Machine shutdown 14 seconds after startup.
111
the legs and you do not have a fire. Removing one component of MACE renders it incomplete. We attempted to expand our measuring hardrvare at a rale equal to our understanding. As a result, r\.e are adding capital to our department at approximately $10.000 per vear. In 1966, we had about $10,000 in portable nreasuring equipment, about six plant installed seismic tvpe
OPTIMIZE YOUR VIBRATION ANALYSIS PROCEDURES
protection monitors, and a small lab building. In 1973, lve have approximately $85,000 in portable eqr-rrpn.renr as listed belorv:
o
Real time spectrum analyzer with
dual channel orbit capabilities
.
l4-channel FN,{ tape recorder and
preamplifiers
o Digital
vector tracking filter r.vith dual channel tuned orbit capabilities
. . .
o
Two and four channel AM tape recorders Five manual tunable vibration analyzers Vibrometer, vibrographs One
7/s,
\/to octave band vibration,/
sound analyzer rvith recorder Fig. 2-Rotor shaft crack at coupling kbyway on 9,000 horsepower steam turbine.
ment, velocity, acceleration and sound measure, strobe iight tLrnable filter for 10-1,000 Hz use.
Proximity meqsure. As long as the equipment
was accessible and the force transmission lrom rotor to bearing housing r,vas high, this program rvorked well. However, a ne\v generation of machines tvere accontpanl,ing large single train plants.
c Normal speeds were now 10,000-52,000 RPM. o Ratios of total machine rveight to rotor rveight were running 20:1 up to 125:1.
o Coupling and shafting rvere total enclosed tinuouslv lubricated coupling drives.
for
con-
first
critical speed and often above the second critical.
o
Horsepower per train '"r,as ty'picalli, 10,000 to 18,000 horsepou'er. Other companies are up to 40,000 horsepou'er irer train. Protection monitorins systems using eddy current (in-
ductive type) sensor probes became necessary. We have approximately 150 of these probes installed on bearing housings observing shaft motion relative to the bearing.
for early
detection of machine malfunctions and are perfect for r.ibration analysis. Protection monitoring is all togcther a different philosophy lrom vibration analysis though nany view it the same. How'ever, if one will provide a rvell designed monitoring svstem. the data can be easily obtained for vibration analr.,sis.
Present qnqlysis syslem. About 1966, rve adopted a progran called MACE (Measurement, Analysis, Correction, Enqineering). This rvas'viewed similar to the fire triangle of fuel. oxygcn, and ignition, i.e. remor.e one of 112
9,500
$20,000
$ 5,000
$
1,000
$10,200
$
1,000
$ 3,100 $ 9,000
S 8.000 S 5.000 S 2.000 S 3.000
$ +.0i-rtl $ 1.0110 S 2.100 S 3,500 SST.OOO
One might question the amount of investment. but oLrr rotating equipment maintenance cost is $2 MM,'r'ear. The compressor repair cost are only $% MM. The turbo compressor maintenance cost with drivers is approxirr-rateh' $3.00/horsepor,ver/year. The 30,000 horsepoiver ir.i a trvo
ycar old 1,000 ton/day' methanol plant operates less rhan
. Equipment was operatine generally above the
These "electronic eyeballs" give good sensitivity
o Alignment-24-channels proximity o Alignment-3 optical instruments and accessorres o Oscilloscopes and accessories o Strairl and fatigue gage facilities . Test instruments and cables o $sns615-piezoelectric, geophone,inductii,e . Tu,o digital tachometers o X-YY' recorder (amplitude vs phase vs RPI.I plotting) o LonS rvheel base van rvith rvork table
$
$2.00/horsep ower f year.
Prqctices. The API Standards recently published are good standards for rotating equipment. Monsanto co:-rtributes to these standards and uses these standards :s do the refineries. Some of the practices for protection :,onl-
toring are:
o
Trvo 90o circumferential oriented probes are attached
at each bearing of critical equipment. . A phase marking, one event per revolution. probe rs provided on each driver. o Dual channel "high read" alarm and trip readourr are provided.
c Should all measurins points not be monitored. tire information is routed to the control room for dia-qnostic measure.
o Explosion proof, NEMA to the rrrachinery location.
7 type enclosures are limited
o ^{n attempt is made to provide installations that allorv probe replacements while operating. o Leads are protected in rigid and flexible conduit. o Thrust displaccment probes generally read the thrust
collar as opposed to the shaft end near the thrust collar. . Thrust displacement probes are always on automatic shutdown. Calibration of 100 mv/mil with a minimum linear range of 80 n-rils specified.
. Five mils (peak/peak) radial vibration meter range is normally specified. . Two level alarns are always specified. The first warns andf or allows call-out for troubleshooting. The second level commits either automatic or manual shutdown. . Electrically deenergized systems which are energized to trip are preferred over contrnuously energized systems. o Coaxial connectors are kept to a minimum and these
are wrapped in Teflon tape and encapsulated in RTV silicone rubber. o Sensor svstems are standardized, generally lr,'ith 15' probe lead lengths.
o The monitoring hardrvare is located on the same side of drivers and driven equipment. o Vibration
sensors are arranged so that one sensor is predominatelv horizontal and the other is predon.rinately vertical.
o The horizontal sensor must sense the rotating assembly first in rotation with respect to the vertical sensor. . Great care is given to safety by using grounded system with care that only one ground exists, i.e. no ground loop noise.
There is a philosophy on vibration analysis which must
It needs to grow in normal steps. One can always measure more than can be explained. The training and supporting education comes slowly. Many times we be realized.
investigate a pin pointed area not knowing what we will find, but having at least pin pointed the area. At some
point, science finally or.ertakes art, but there is a lot of art left. Intuition becomes a strong tool. Often, \\'e can anticipate what rvili happen without fully understanding why. To be eflective, one nrust be thoroughly familiar rvith machine internal construction. EXPERIENCE
One might presume that many "saves" must occur to justify the continued purchase of expensive portable and fixed vibration equipment. This is true. Keeping this paper limited and current, some selected results will be listed more to show the t,vpe of coverage rather than the amount.
Cqse l. During commissioning, a 9,000 horsepower, 11,000 RPM steam turbine developed a 14/o non-synchronous oil whirl instability that eventually reached approximately
5 mils peak/peak vibration, journal to bearing. Several tests and
orbit studies indicated the instability was initiated
by steam valve sequence loading. The shaft first orbits elliptically from 11 to 5 o'clock. Then after the next quadrant of valves opens, the orbit shifts to 9 to 3 o'clock. The pressure dam type bearing locking pin was relocated to damp the 9 to 3 o'clock motion and the machine was operated here for two weeks until a 4 lobe bearins design by the turbine builder was manufactured. This bearing was installed in one shift with no loss in production. Cose ll. A compressor rotor on a 14,000 horsepower train was slowly fouled with a clay like deposit over six months, suddenly unbalancing the rotor during slough off. REPRINTED FROM HYDROCARBON PROCESSING
Fig. 3-Heavy salt deposits on steam turbine rotor blading which caused thrust bearing lailure,
The synchronous response jumped from 0.6 mils peakto-peak to 2.7 mils peak-to-peak. The equipment was closely monitored for two more months for a scheduled shutdown and the spare ro'tor installed. During the next six months, the process separation design equipment was revised and a second rotor switch with separation revision put us back in control. IJnbalance amounts confirmed rotor res'ponse tests during the mechanical run test at the factory before shipment to job site. Cose lll. A 5,000 horsepower motor-gear-compressor train was shutdown fifteen seconds after commissioning start.
A severe rub was detected by the vibration probes, limiting the damage to the compressor sleeve and labyrinth seals (Fig. 1) . The unit was repaired in forty-eight hours. The reason for the rub was a mismatched top seal half.
!V. Severe misalignment of an 8,000 horsepower turbine-compressor train caused a coupling-end unbalance which rose to 4 mils peak-to-peak. Shutdown and inspection at the coupling hub-shaJt keyway section indicated a fatigue break 60/o through the shaft (Fig. 2). Gose
Cose
V. Startup of a refrigeration
compressor accom-
panied by vibration and hot alignment measures indicated that the compressor discharge pipe to the condenser, moved the compressor over 100 mils horizontally out of alignment with the drive turbine. The refrigeration design manufacturer rechecked his expansion joint calculations and found an effor in number of convolutions required. Vibration signals were typical for misalignment, i.e. high axial vibratibn and twice running speed radial vibration. Gqse Vl. A turbine oR a five case train developed 3 mils peak-to-peak vibration at 10,300 RPM. The coupling shaft keyway was short a full key, thereby causing an unbalance condition. Cose Vll. A 14,000 horsepower, 11,000 RPM steam turbine was field balanced with two external balance rings to 113
Cqse lX. At anothcr \.{onsanto loaction, a 15,C00 holsepower stcam turbine \^,as shllt dou,n manuallv on both tlrrust :rlarnrs pltrs the radinl vibration alarms. 'Ihe turbine had a heavy deposit ol salt on trto stagcs (Fig. 3). f-he thrrrst bcarins had lailed and the blading shror,ds contactcd tlre cliapl-rraems cutting into thern altc',ttt ,/g" .Figrtre 4 shou's rlzrtrrge to the thrrrst r:ollal fronr overireatinq... 1'lrouglr the trailing edges on trr,o blaclinq- ro\\'s \\ere cllair'n and lcqrrirecl gr:incling. the rotor rvas savercl and put back
OPTIMIZE YOUR VIBRATION ANALYSIS PROCEDURES
in scrvice one rrecli later. Nlonsanto l-ras paid fronr $f,/; I\,IfI to $1 lI\I 11,1te oI u reck in the i:ast. Cose
thrust collar damage resulting from salt deposits
Fig.
-Severe turbine blading. Note the heat tears on the thrust on steam collar face.
X. At another
Nl[onsanto locrtion,
oLr rirrs
the coLiplirr
housing filled rvith oil clur: to an iurpropcrly sizecl trsket rvhich clrokcd the drain. f'his carLsccl an overfilling rriLrclr took serrerll hoLrrs fronr startr-rp. A or-rc-ha1f nrnning frecluenc)' vibration u,as the inciic:rtion. Tlre nraclrinr: rr'as shrrt c1orr,n to inspecl the bearilg. A couplc of entr(rti,: nraintenance men startecl liftinq thc top half oI tire coLrpline guarcl onlv to find the flooclecl condition. ir ir':r. a lrLckv acciclcnt.'fhe bcarine \\'as sllslrect btrt tirc c:,r.rpling, Ior thc first tirne in its lifc. sarv the opprrrtrrr,,r,. ', bccorne a bealing, tlrough a iloor one. Tlris brings rrp an intcrcsting thoLrglrt rrhoLrt i'ibr'-.ii, i, annlrsis. Sever'.rl rnallLrnctior-rs ma1' scncl thc
serr' t.,r ' .i
ileclLrcncl'signal, i.c. one miglrt think oI tlris.ri LirL.'-r' rnorc reclio st:rtions transnritting on somcone clsr s iie-
rcdLrce srnr:lironous vilrratiorr leverls lronr 2-6 rrrils Pcak-to-
to
c]l tel
tc\,.
1.25 nrils peali-to-pcali in one shi[t rvithotrt clis-fhis .rsscnrblr'. cxtcrrdcd this rnaclrine's trvo-\'r:rr rLrn irrto 1973 ancl possibly lonqer. Scisrrric anirllzcr witlr strobr:
A loose thrtrsl collar, loosc cotrpling hrrb, clorLblr' ;rrr,ke.i shaft. oL nrisaligrrrrertt rrright n,cll scrrcl a trvit c t 'Lri,ir.:l
trr o 1;luue cotrltle colrcction rr,:rs .rraclt: orre urontli l:rtcr in less than onc hoLrr to ellcct a 1.1 rrril ltcak-to-pelLk I'ihla-
llcceusc 60(b o[ our ribration problcrns resLrlri nrisaliqrrit'nI o\ cr oLrr first fir'c vears of vibratiorr vr"c bcqan :r prt'cisc nlcusLrrenrcnt oI slraft and c:rsr r-ncrt rrsiug the edclv crrrrclrt probes. Wr.l liarc 21 oI this ecluii-lrrcnt, plus ll optical instnrrncnts tir rvork. lioth tcclrnirlLres have been pLrblishecl.l':
lreak
reference w'cre coupk:d u ith rrrbital phast: rr:rrking to placc u't'ieht ancl corrct:t irr:r rrininrLrrn oI trvo rulrs. 1\
tion lcvel.
Cose
Vlll.
I
)ue to an inrProPcrlr' lrositiorrccl clisclrerse-clcl
bcaring Jrorrsinq, tlre thlLrst orr I balrcl corrrprcssor lr:rcl u'oln [he ba]ancc piston labvlinth r'lcar:rnce to triltle orieirr:rl clcer;rrce.'['hcr thnrst dcflcction graclLutllr, progrcssccl Lrntil :rt 17 trrils frortr oriqinrl load point, IJLe coltpr.cssor rlas shrrtcIow n rr'ilh :r snrr.urecl lrrrt Parti:rLlv ftii1t'cl he aring bLrt undenr:rgccl c:olrprrssor r'otor essr':nbl\'. Static baLrncc
charnber l)ressLlre colfirntccl loec[ing esit oil terrlleraturcs clicl rrot.
but tltrrrst
bcar.ing
CrtlRtos J.LcxsoN is
o.n a.nglimeering
f
el-
in a X[eclrutt,ical Technology Section utorking on mechanical u,ncl methndlout
oriented problenrs ctt the l\Ionsanto Polyttters ancl Petrcc|teyrical Com.pany, Teras lm.s ltccn witlt. tlte cotn-
City, Texas. He
pa,ny for 2") yea,rs tLorking in Di,uision EnJtineering, Prodtt.ctiott, Maintenance o,n.rl Plctnt Dcsign, bef ore setting up the
in 1f)60 later being s p e ciali s t, then
nppointcd engineering
senior cngineering speciulist. Mr. Jctckson ltolds a B.S. degree in nreclnnical engineering lront Teras AcQM Uni.uersity, plus an AAS degree in elcctronics teclmology from College of the ll[uinland, Teras City, Texas. He is a registcrecl prof essional engineer, a tnember of ASME, API, SESA, Pi Tctu Sigma,
Beta Pi, Vibration Institttte, and the Texo,s A&X[ Turbonruchinery Symposium Aduisory Committee. Tau.
114
.
Severiiy. 'l'hc nrost diflir:rrlt problern irr vibration .rr-r1.,.i. is cstablislrr rent of sevclitl, lirnits. M:it lrirres. rr LL, jr iii.e pcolrlr:. lrar e cli[Tcrent sersitir']Lir:s end rcsist:inces ro r rjrr'-r-
tion lirrc es.
In
gc'rcr-al.
\rc lra\c founcl tlrat 0.2 irrclrcs':ec,:,i,d
(peak) r'clocitv to retlLrilc:rlarrrr action on beerirq- hoLrsir: scisnric tr'pc reacling. L}r'eater vlrlrrcs. e.g.0.il to 0.5 iir. ,':ii
ronrrll1, Iorcc
About the quthor
ytresent rle1nrtm,ent
spcccl lreclLLer-rcy signal.
slrLLtclorr Lr.
The shaft-to-bcllir,g
r',
i.
r.i
'
rreasrrre lirrits :Lre clilTrlent. \\'itJr cle:rn sh:rlts. rr(' -.irerallr,' sct lliq.lr spted rrechines u,itli 2.2 liils pcrl.r-t, -ir,':rl. rr arnirq arrd '1.2 rnils peak-to-pcak slrutdo\\,n lirrrits, Rotor tlirLr,st displrcernunt rrronitors.irc s( t rlitlr li 1,,i.
orcr normal loacl clcflection as an alarn-L:rncl ,'i1]11'1'l('r r,: 25r mils 1;rst nolnral lord dr'flcction:ls alr atrtorrr.rtii ..,.riclorr'n position. Ilascrl on planI cxpericnce, it is felL th^rr vou crlnnot risli rnarLr:Ll slruti]orvr-t irt thrust. The rorrr.r i failLrrc tirne is
thirtr',st'conds!
I
AC]KNOWI,EI)GN{ENT '[ he autlror rvishes to acknorrledqe the contr ibulions of linginccLs E- .f()lrcrslenr, ParrI S. (iu1rlorr. Cart A. Drrhon. Jlorrerrl Ill;rckhrrrn. ancl Jrrrrcr
[[ Inqrerrr, NIrtrrials & \IcclranirrL 1cc]rnologr, )\forrsa:rto,'lrras Clit.. Tc\,rs to tlr: \l\(lE I)rogranr irs rcportcd. C)opyriglrtcd lrv rrrrrl origir.,lLr' Jr r'\crtr(l at tlrc Tnstitutc ol [,lcctricrLl ancl l]lrctroric [irrqirrcc s, ['etLolcrrm Dirisiorr ('rrrlrrcrrcr'. Selrtcrrrlrcr 1!73, IIotLstorr,'l'txrs, BIBI,IOGRAPHY r flrcl.on. C., "Horv to aliqn lrnrrcl-type ccntrilrrqal cornprcssors," Hldro,'arlrort ['rocessing, ir0, No.9, p.1{19, Scptcrrrlrt:r 1!71. 2,fecksrrn, Ci., 'SrrcccssIul slraIt hot alJgnnrcnt,'' Ilydror:arlron J)rocessing, l{1, \o. l. p. t00, .[anrLary I969.
Index
Alignment cold, 32,33 hints on,33
hot,34-37 problems in, 1 13 tooling tor,32 Balancing, dynamic compressor, 7 8, 1 13-1 1 4 shop, 21 steam turbine,74,75 Cleaning onstream solvent washing, 30 water washing,28,29 Com pressors, centrif ugal
applications for, 18, 19,23 construction materials for, 79, 80 mai ntenanc e ot, 20-22, 24-27 operation of, 20 rotor dynamics of , 19, 78 seals for, 80 seal-oil systems for, 88-92 testing of , 5-16 mechanical, 6-10 performance, field, 1 4-1 6
performance, shop, 11-13, 39, 40
Couplings, high-speed applications tor, 84-87 specifications for 9, 79 Data acquisition
compressor field performance, 12, 13 vibration, 108-113
REPRINTED FROM HYDROCARBON PROCESSING
Gear units, high-speed
failures of,94,95 shipping and installation of, 106 shop testing of , 101-105 specif ications for, 93-94, 96-100
purchase,94,95 test, 96-1 00 vibration of , 96-106 Pumps, centrifugal cavitation, 56-59 grouting, 41-44 performance, 46-48, 58, 59 specifications for, 59
vibration,49-55 Training mai ntenance personnel, 24, 25 Turbines, steam balancing, multiplane, 74-74 blade failures, 73
design,63 fouling, 28,29,113 water washing, 29 Turboexpanders applications for, 66 construction materials for, 63 fabrication of, 64 performance of , 67-69 power recovery f rom, 66-69 types of, 66 Vibration analysis, 108-1 13 compressor,T6, 109, 1 13, 1 14 gear units, high-speed, 96, 106 instrumentation, 108, 1 13 pump,49-55 severity limits, 54, 55, 1 14 specifications, 78
115
Notes
116
