Shop Rotordynamic Testing - Options, Objectives, Benefits ... lectures/2013 TPS...behavior (Design...

Copyright© 2013 by Turbomachinery Laboratory, Texas A&M Engineering Experiment Station

SHOP ROTORDYNAMIC TESTING – OPTIONS, OBJECTIVES, BENEFITS AND PRACTICES

Dr. John A. Kocur, Jr. ExxonMobil Research & Engineering

3225 Gallows Road, Rm 2A-0601 Fairfax, VA 22037-0001

(703) 846-2275 [email protected]

Dr. C. Hunter Cloud BRG Machinery Consulting LLC

703 Highland Avenue Charlottesville, VA 22903

(434) 293-9926 [email protected]

John A. Kocur, Jr. is a Machinery Engineer in the Plant Engineering Division at ExxonMobil Research & Engineering in Fairfax, Virginia. He has worked in the turbomachinery field for 30 years. In his current capacity, he provides support to the downstream, upstream and chemical business lines within ExxonMobil with expertise on

vibrations, rotor/thermo dynamics, failure analysis and health monitoring of rotating equipment. Prior to joining EMRE, he held the position of Manager of Product Engineering and Testing at Siemens Demag Delaval Turbomachinery. There Dr. Kocur directed the development, research, design, engineering and testing of the compressor and steam turbine product lines. He has also held positions with Pratt & Whitney and Amoco Corporation.

Dr. Kocur received his BSME (1978), MSME (1982) and Ph.D. (1991) from the University of Virginia and an MBA (1981) from Tulane University. He has authored papers on rotor instability and bearing dynamics, lectured on hydrostatic bearings, has sat as a committee chairman for NASA Lewis and is a member of ASME. Currently, he holds positions within API as 617 vice-chair, 684 chair and chair of Subcommittee on Mechanical Equipment (SOME).

C. Hunter Cloud is President of BRG Machinery Consulting, LLC, in Charlottesville, Virginia, a company providing a diverse range of rotating machinery technical services. He began his career with Mobil Research and Development Corporation in Princeton, NJ, as a turbomachinery specialist responsible for application engineering, commissioning, and troubleshooting for

production, refining and chemical facilities. During his 11 years at Mobil, he worked on numerous projects, including several offshore gas injection platforms in Nigeria as well as serving as reliability manager at a large US refinery.

Dr. Cloud received his B.S. (Mechanical Engineering, 1991) and Ph.D. (Mechanical and Aerospace Engineering, 2007) from the University of Virginia. He is a member of

ASME, the Vibration Institute, the API 684 rotordynamics task force, and the advisory committee for the Texas A&M Turbomachinery Symposium.

ABSTRACT

Understanding the lateral rotordynamic behavior is critical in determining the reliability/operability of rotating equipment. Whether examining a centrifugal pump or compressor, steam or gas turbine, motor or generator, rotating machinery share the same need to accurately predict and measure dynamic behavior. Industrial specifications determining fit for purpose rely on the accuracy of rotordynamic predictions where direct measurement is impractical or otherwise impossible in an industrial setting. Testing to confirm rotordynamic prediction and behavior provides both the purchaser and vendor the confidence that the design will meet project expectations.

Rotordynamic shop testing has several options available to the project during acceptance tests at the vendor’s shop. These options include mechanical run, string and full load/Type 1 testing as well as verification testing to validate unbalance response and stability predictions. Such testing has numerous advantages; the most important being the avoidance of production disruptions involved with testing at the job site. Each test option has associated costs as well as limitations as to what lateral vibration characteristics are revealed. Understanding these factors is vital to efficiently mitigate the risks associated with the purchased equipment.

Applying best practices and an understanding of the industrial (API) test requirements are needed to derive the maximum benefit of each test option. The best practices not only involve the test procedure but also the associated analytical methods used to post process the measurement information. Whether performing a simple mechanical run test or more complex stability verification during ASME Type I testing, ensuring that a logical, repeatable and proven methodology is followed produces reliable evidence to confirm the rotordynamic model and lateral vibration performance. The rationale behind the API test requirements provides an understanding of why that test is being performed and its correct application to the dynamic behavior.

Proceedings of the Forty-Second Turbomachinery Symposium October 1-3, 2013, Houston, Texas


Test options can be separated into two categories; tests that reveal portions of the dynamic behavior of the equipment to specific operating conditions and those used to verify the analytical predictions of that behavior. API mechanical, string and Type I (or full load) tests reveal the rotordynamic behavior of the equipment to a given set of conditions. These are used specifically to determine acceptability of the design. Unbalance and stability verification testing is used to confirm (or provide confidence in) the rotordynamic model. Confidence in the model permits extrapolation of the design (vendor) and operation (purchaser) beyond the machine’s as-built and specific shop test conditions.

INTRODUCTION

The first task of an annular seal is the restriction of leakage flowrate between a rotating shaft and a stationary housing. As it turns out, annular gas seals can also have a significant impact on dynamic characteristics of compressors and turbines. Demands on turbomachinery continue to push designs beyond experience limits in terms of speed, power, size, pressure development and flow rate all the while demanding higher reliability and operability. To meet the performance objectives of the application, almost absolute knowledge of the behavior of the machine is necessary. To aide in this understanding, advanced analytical methods have been developed in parallel with the increases in computational power. Whether predicting blade flutter due to impingement flow, reaction of a machinery train to alternating torque or rub potential due to sudden imbalance, current tools sets allow designers to study detailed aspects of machinery behavior like never before.

Shop acceptance testing has long been used as the “final” check of the turbomachinery design and is required for all special purpose equipment. With the advancement of computational methods, failure rates during testing have been greatly reduced. However, experience has taught us analytical methods alone are insufficient to guarantee the “right the first time” philosophy that many reliability systems employ. Whether used to provide data to baseline prototype equipment, to benchmark extensions of current experience limits, or for verification of proven practices, testing remains an integral part of all reliability systems.

Understanding the dynamic behavior is critical in determining the reliability/operability of rotating equipment. Whether designing a pump, compressor, turbine, motor or generator, all rotating machinery share the same need to accurately predict and measure dynamic behavior. Literature is swamped with failures that resulted from both inadequate design and testing methodologies. While the potential for failure originates at the design stage, testing represents the final step to identify that potential. The importance of efficiently employing both cannot be understated [1].

Turbomachinery is dominated by two classes of dynamic behavior; rotordynamics and fluid dynamics. As their names imply, each focuses on a specific dynamic behavior; rotordynamics on the rotating system’s vibration and fluid dynamics on the mechanical interaction with the working fluid. As our depth of understanding increases, the more interrelated these behaviors become. Decisions made in the design of one

can impact the other with sometimes disastrous effects. Nowhere is this more evident than centrifugal equipment, especially compression. For this reason, this tutorial will focus on testing the rotordynamic behavior of centrifugal compressors. However, most of the principles and practices are applicable across all types of turbomachinery.

Rotordynamic behavior testing involves both direct measurement and inference. Direct measurement of the vibration is typically made at the journal locations. Internal vibration levels at other critical locations can only be inferred from these measurements using the rotordynamic predictions. Thus, industrial specifications determining fit for purpose rely on the accuracy of rotordynamic predictions where direct measurement is impractical or otherwise impossible in an industrial setting. Testing to confirm rotordynamic prediction and behavior provides both the purchaser and vendor the confidence that the design will meet project expectations.

Rotordynamic shop testing has several options available to the project during acceptance tests at the vendor’s shop. These options include tests to demonstrate operating behavior at a specific condition (mechanical run, string and full load/Type 1 testing) and verification testing to validate unbalance response and stability predictions. Such testing has numerous advantages; the most important being the avoidance of production disruptions at the job site. Each test option has associated costs as well as limitations as to what lateral vibration characteristics are revealed. Understanding these factors is vital to efficiently mitigate the risks associated with the purchased equipment.

Application of best practices and an understanding of the industrial (API) test requirements are needed to derive the maximum benefit of each test option. The best practice not only involves the test procedure but also the associated analytical methods used to post process the measurement information. Whether performing a simple mechanical run test or more complex stability verification during ASME Type I testing, ensuring that a logical, repeatable and proven methodology is followed produces reliable evidence to confirm the rotordynamic model and lateral vibration performance.

Test options can be separated into two categories; tests that reveal portions of the dynamic behavior of the equipment to specific operating conditions (Vibration Demonstration Tests) and those used to verify the analytical predictions of that behavior (Design Verification Tests). API mechanical run, string and Type I (or full load/full pressure) tests reveal the rotordynamic behavior of the equipment for a given set of conditions. These are used specifically to determine acceptability of the design in a pass/fail mode. Unbalance response and stability verification testing is used to confirm (or provide confidence in) the rotordynamic model and analysis. Confidence in the model permits extrapolation of the design (vendor) and operation (purchaser) beyond the machine’s as-built and specific shop test conditions.

The tutorial will cover the following aspects of rotordynamic testing:

Decision to test: Why and on What Basis?

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The user must also consider the safety, health and environmental impact of possible failures that can be attributed to the machinery in question. The impacts can result in the release of gas from component failure due to high vibrations, failure to meet regulatory requirements due to unplanned outage or injury due to parts release during failure. The risks of each can be determined through a failure mode analysis [3] that incorporates the rotordynamic contributions to the identified failure modes.

Experience plays an important role in determining the extent of testing to perform. Experience in this case relates to both the vendor and user. Obviously, prototype machinery or equipment that extends the experience limits of the vendor should be tested thoroughly. What may be overlooked is the experience of the user with that equipment. First application of technology within a user company may benefit from additional testing at the vendor’s shop. The testing can be used to better understand the dynamics and what conditions or operating nuances may affect the rotordynamic behavior. Finally, prototype components within the machinery may require additional component testing to determine their impact on the rotordynamic behavior of the machinery.

Finally, the results of the rotordynamics analysis should be used to weigh the decision to test and which test to select. Machines shown sensitive to destabilizing forces or those predicted to have low stability levels may benefit from stability verification tests. Similarly, a rotor with high amplification factors may influence the decision to perform unbalance response verification testing. Verification testing, as noted earlier, is intended to prove the accuracy of the predictive tools used to model the rotor behavior and thus assess the acceptability of the design. The use of a proven rotordynamic analysis is an effective and efficient mitigation strategy towards reducing risk.

ROTORDYNAMIC TESTING

Options

Fundamentally, there are two types of testing options available for lateral rotordynamic assessment. The first type, referred to as vibration demonstration tests, demonstrates the behavior of the as-built machine and/or train for a specific set of operating conditions. Generally, only the site specific instrumentation is used to measure lateral vibration behavior. The tests do not address the accuracy of the model or tools used to predict the rotordynamic behavior, nor do they attempt to estimate or determine the robustness of the design. Those that fall into this category are:

Mechanical Run – an example is the API required mechanical test [5]

String – API 617 refers to this as a complete unit test [6]

Full load/Full pressure – Referred to as a Type I test by ASME [7]

Many important aspects of the rotordynamic behavior cannot be practically measured (i.e. internal vibration levels, separation margin to modes above operating speed) nor can

every possible operating condition or combination of assembly /machining tolerances be tested. To fully understand the acceptability of the design and the rotordynamic behavior, we must rely on the accuracy of the rotordynamic analysis. Consequently, tests to determine the accuracy of those predictions were developed. These additional tests were developed to provide more than a pass/fail test, which is essentially what the vibration demonstration tests are. They were implemented to determine the robustness of the design. This is particularly useful when operating conditions are widely variable or when design experience limits are exceeded. The second type of testing option is referred to as design verification tests with the two most prevalent being:

Unbalance response – This test would include the more invasive testing required by API [8]

Stability – Not currently specified by API standards.

Verification testing also has the options of where these tests can be performed. In either case, the tests may be performed in a high speed balance bunker, during the mechanical test or during the full load testing. Each option adds additional considerations in what can be measured, dynamic effects included in the test and what portion of the dynamic behavior is analyzed.

Test Preparation

Following the decision to perform a rotordynamic test, the user should decide which specifications to apply to the test. The specification should describe the objectives and requirements of the test. For several of the test options described above, API standards have described a specific test procedure to be followed. The API standard details the procedure, objective and requirements for the test. For the more specialized testing, the user will need to develop their own test specification. This can be done with the assistance of the vendor or by consulting industry specialists.

Test Objectives

The objectives of the tests performed should be discussed at the initial stages of the project. Agreement on the test objectives will assist in the determination of what equipment is needed, measurements to be taken, and conditions to be run. Generally speaking, the objectives of each test are listed below. Other specialty objectives can be added, but the ones listed below would form the basis for each test.

Vibration Demonstration Tests

Mechanical Run Test

The mechanical run test as required by API is primarily a vibration level check. Measured at the probes located at the journal location, vibration levels are checked against the specified limit for both overall and non-synchronous components. General mechanical performance is also examined including bearing temperatures, close clearance rubbing and seal performance typically up to maximum continuous speed (MCS). Supercritical behavior is examined by determination of the amplification factor and separation margin of typically the 1st critical speed. (Obviously, the behavior of modes above MCS remains undetermined.)


String Test

The string test is not much different than the mechanical test mentioned previously. As the name implies, the string test is performed with all or a major portion of the machinery train connected (typically everything but the driver.) The objectives of this test are also similar to the mechanical test of a single body in that vibration levels are checked against limits, mechanical performance is examined, and supercritical behavior is analyzed. However, in this case, these are determined for the coupled train configuration. The string test is run to measure the coupled body dynamic behavior (when rigid couplings are used in the train) or the coupling spacer dynamics (for couplings with unusually long or heavy spacer tubes.)

Full Load / Full Pressure Test

Full load / full pressure tests are rarely performed based on rotordynamic justifications only, mainly due to the costs involved. However, the Type I test does permit vibration level checks at operating conditions, stall determination, impact of internal loading and deflections on dynamic behavior of individual components (seals and bearings) and a binary check for stability (yes or no). Range testing is rarely done during the Type I test. Typically gas properties are held constant, clearances are left at as-built values and alternate configurations are not considered. Thus, while some aspects of rotordynamic behavior are tested, margins and robustness are left unchecked. The machines undergoing these tests leave those factors to analytical studies whose accuracy may remain unchecked. Keep in mind, to predict the stability (stable vs. unstable) of any machine is the first step in developing a good predictive tool. However, it is not the only step. As designs extend the operating or design experience, it becomes necessary to predict the stability threshold, separation margins and overall optimization of the design correctly, thus the need for verification testing.

Design Verification Tests

Unbalance Response

The likely first attempt to publish a verification test in an industry standard was the unbalance response verification test published by API 617. The objective was simply to verify the unbalance response prediction accuracy of the vendor’s rotordynamic analysis with regards to the machine’s unbalance sensitivity within the operating speed range and the location of the critical speeds (below trip speed.) The verification test analyzes both the predictor tool and model employed. A methodology was refined over several editions of API 617 within the limitations of the mechanical shop test. Alternatively, performing this test in the balance bunker has gained acceptance with the increase in at-speed balancing of rotors and the freedom it permits in terms of weight placement and additional measurement points.

Stability

Several methods have been developed analogous to the API unbalance response verification test with the objective to verify the stability predictions of centrifugal compressors. As with the unbalance response, the intent is to measure more than just

“is the compressor stable” but “how stable is it.” The measurements are then compared against the analytical predictions to determine accuracy. Pettinato et al. [9] presented a methodology employing this test during mechanical and performance testing (as required by API 617).

Test Information / Knowledge Gain

To determine the extent of testing to perform, one needs to understand the information or knowledge gained of the rotordynamic behavior of the body or train. While similar information can be obtained from several of the tests, the costs associated with each determine the overall efficiency of obtaining the necessary information to mitigate risks identified in the risk matrix. As before, the tests are separated into vibration demonstration and verification testing. Vibration demonstration testing confirms the machinery can meet project specifications for a given operating and as-built condition. No attempt is made to confirm the accuracy of the analytical prediction beyond confirmation that the specification has been met. Since the accuracy of the analytical prediction remains largely in question, inferred information from the analytical method should also be questioned.

Vibration Demonstration tests

Mechanical Run

Mechanical testing provides information related to the critical speed location and some indication of that modes behavior. The modal information is limited to only those modes located below the maximum test speed achieved (trip speed in most cases.) Typically, this is only the 1st critical speed. The modal behavior information is restricted to the amplification factor which is sketchy at best. The amplification factor can be highly sensitive to the acceleration/deceleration rate and whose magnitude is not restricted by API. The test also validates the balance procedure’s effectiveness in meeting the project vibration limit specification.

For certain low risk applications, this amount of information is sufficient. What isn’t tested however can be significant. For example, subcritical motor applications (incorrectly termed stiff shaft) operate below the 1st critical. This critical speed can have high amplification factors and can be damaging if the separation margin is lost. Performing only a mechanical test will tell the user whether the mode is on or below the operating speed. The amount of separation remains untested and can only be inferred from the unverified analytical predictions. Thus, the impact of support stiffness loss on the location of the mode in the field (due to structural or bearing clearance changes) can have significant risk associated with it.

String Test

String testing will provide the same type of rotordynamic information as obtained with the mechanical testing but at the higher costs of assembling the entire (or high speed portion) of the train. Information again limited to the critical speed location and some indication of the modal behavior. Ignoring other reasons to perform string testing (e.g. fit checks for trains being sent to remote portions of the globe), rotordynamic justification for the test should be limited to the information supplied by the test, namely, dynamic behavior of the coupling


spacer(s) and rigid coupling effects on the dynamic behavior of individual bodies. As noted in API 684 [10], a train lateral analysis should be requested for unusually long or heavy coupling spacers or when rigid couplings are used. In this case, the correct boundary conditions at the shaft ends are obtained when the train is modeled (train in this case refers to the bodies on either side of the coupling(s) in question.)

Testing of this train configuration should mimic the analytical model to verify the behavior in question. With a rigid coupling, the rotordynamic behavior of the bodies rigidly coupled will be affected. For unusually long or heavy coupling spacers, the dynamics of the spacer can only be accurately modeled/tested with the hubs attached to the shafts.

Full Load / Full Pressure (FLFP) Test

Type I tests provide a stable vs. unstable behavior indication of the rotor to specific test conditions. The test also reveals the change in lateral behavior of the measured modes (typically the 1st critical speed as with the mechanical test) due to internal loading of the compressor and application of gas pressures and densities close to the design values at the seals. The latter introduces seal dynamics into the testing that is only achieved during the FLFP test. This is important for both stability and synchronous behavior especially for machines incorporating hole pattern or honeycomb seals. The presence of subsynchronous vibration due to phenomena such as stall, surge or whirl may also be revealed during the FLFP test.

Design Verification Tests

Unbalance Response

The unbalance response verification test (URVT) provides a measurement of how well the analytical predictions match the vibration produced from a known unbalance. This in turn adds confidence to the internal deflection, separation margin and unbalance sensitivity calculations made from that analysis. On the shop floor, the unbalance weight addition is typically limited to the coupling. Some turbines and overhung machines have the ability to add internal weights or weights to the overhung impeller. Optionally, the test can be performed in the vendor’s balance bunker. The bunker permits more freedom in terms of weight placement and measurement of shaft deflection at points other than the job’s proximity probe location (mid-span, for example.) Of course, the analytical model needs to reflect the setup in the bunker; bunker bearings if used and support stiffness of the bunker pedestals. Since the intent is verify the accuracy of the analytical predictions, these differences in configuration should not affect reaching that conclusion.

Stability

The stability verification (SVT) test provides confidence in the analytical predictions regarding the stability level and position relative to the stability threshold. As with the unbalance response verification, options are available regarding the platform or test configuration to perform this test. The results of each platform can be summarized below regarding the accuracy of the analytic method to predict:

Balance bunker – Rotor, bearings, and pedestal support impact on stability (bunker bearings if used) at various speeds.

Mechanical test (Vacuum) – Rotor, job bearings and casing support impact on “baseline” stability (or basic log decrement reflecting the bearings and rotor only) at various speeds.

Performance test – Rotor, job bearings, casing support and reduced aerodynamic and seal behavior impact on stability for a limited speed range.

Type I test – Stability level and margin at nearly the same operating conditions as expected in the field. The range of gas conditions, inlet and discharge pressures and flow rates may be limited as a result of the test setup.

Test Benefits

The benefits derived from the rotordynamic testing can be identified for the two groups of testing. Vibration demonstration testing provides the purchaser the following:

Demonstration that vibration levels and critical speed separation margins (for those under the maximum test speed) specifications have been met – All tests

A stable vs. unstable check is made for a specific test condition – Type I test

Proof of effectiveness of the balance procedure in meeting vibration level specifications – All Tests

Non-synchronous vibration levels examined – Type I test, to some degree all tests

Design verification extends those benefits to:

Determining accuracy of unbalance response calculation with regards to the unbalance sensitivity – Unbalance verification (shop floor)

Determining predictive accuracy for mid-span unbalances and deflections – Unbalance verification (at-speed balance bunker)

Reassurance that mid-span deflections are within operating clearances for close clearance locations – Unbalance verification (higher for at-speed balance bunker)

Verification of stability level prediction for system with no excitations (baseline stability) – Stability verification (at-speed balance bunker and during the mechanical run test)

Reassurance that rotordynamic model of shaft and bearings is accurate – URVT and SVT

Robustness and optimization of machinery design – URVT and SVT


Correctness of the stability model of impellers, annular seals and other destabilizing mechanisms as well as the effectiveness of any components which are intended to reduce the destabilizing effects (shunt bypass, swirl brakes, damper seals) – Stability verification (PTC Type I and Type II testing)

The at-speed balance bunker can extend the benefits of the URVT by permitting weight placement at and measurement of locations that are more sensitive to unbalance and the negative effects of high vibrations. For between bearing machinery, critical close clearance locations that impact performance are located at or near the mid-span where the first critical speed has its peak deflection. In addition, unbalance creation is more likely to occur at the mid-span due to deposits, erosion or corrosion and is more likely to excite the 1st critical speed. Unbalance placement at the coupling (typically the only readily accessible location during the mechanical shop test) does not excite the 1st critical speed significantly and provides minimal information or assessment of the prediction accuracy of this mode.

Typically, URVT on the shop floor is limited to the coupling location. Journal probe vibration levels produced by adding a 40W/N weight to the coupling is less than 0.00025” at MCS. Performing the verification test in the bunker can permit measurement of the mid-span and journal locations to a variety of applied unbalances. Direct measurement of the relative displacements along the rotor allows for closer scrutiny of the predicted shaft and bearing dynamics as they relate to the amount of modal bending and damping at the critical speed. It should be noted that the intent of the test is to determine the vendor’s ability to predict the unbalance response behavior of a model containing a shaft, bearings and support structure. The results may not be indicative of the actual behavior of the job machine and relies on accurate modeling of the bunker pedestal dynamics. However, the assurance gained from verification of the predictive method in the bunker should carry over to the job rotordynamic predictions.

Vibration demonstration tests can reveal the presence of instabilities. However, even if no problematic nonsynchronous vibrations are observed, the machine’s actual stability could be very small (say less than the API design minimum requirement of log decrement of 0.1) and close to the stability threshold. This state of blindness with respect to the machine’s actual stability, as well as the significant uncertainties remaining in current modeling tools for stability prediction [11], can only be mitigated by directly measuring the machine’s log decrement through a SVT.

Unlike the URVT, a SVT requires additional hardware in the form on a temporarily, mounted shaker. While this is an added complication, a SVT is often a cost and technically effective alternative to the very expensive Type I test. Such Type 1 testing is unlikely to receive project approval unless it is specified very early on during FEED. SVT can provide a valuable stability assessment of the machine during the other operations/tests that are often typically specified, namely, at-speed balancing, MRT or PTC 10 Type II performance tests.

When done in conjunction with high speed balance or the

MRT, an SVT can measure the machine’s basic log decrement. Figure 2 shows the results of such testing versus speed for a particular test rig mounted on tilting pad journal bearings [12]. Measurement of the basic log decrement provides evidence as to whether or not the machine’s rotor/bearing/support system has the robustness to counteract the destabilizing mechanisms that it will experience in the field. As shown in Figure 2, such basic log decrement testing also provides the opportunity to assess competing tilting pad journal bearing models, a topic of much debate in the industry [11, 12].

Figure 2) Base stability measurements versus speed

RECOMMENDED PRACTICES OF ROTORDYNAMIC TESTING

This section will focus on the rotordynamic testing of centrifugal compressors due to the complex dynamic behavior of this class of turbomachinery. Concepts developed in this section can be applied to all types of machinery as they all share certain aspects of rotating behavior. In addition, this section will focus on rotordynamics only. The tests listed below, especially the non-verification testing, are also used to prove other aspects of the machine’s acceptability. These aspects will not be discussed.

Vibration Demonstration Tests

Mechanical Run Test

Mechanical run test of the rotating equipment should be viewed as a minimal test to determine the rotordynamic acceptability and should be considered for equipment that is designed-for-purpose in contrast to equipment selected from a catalog.

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Copyright

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Copyright

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Test Procedure

For the FLFP test, the procedure should be developed in concert with the rotordynamic analysis. The FLFP test is intended to study the stability of the centrifugal pressure under similar conditions to the field. By applying load and pressure, destabilizing forces of the seals and impeller/shroud interactions are introduced including seal clearance changes due to internal deflections as a result of reaching full pressure. Operating points during the FLFP test should be determined, in part, by the rotordynamic analysis and reflect operating points of minimum stability. These could represent operation at MCS near surge (highest differential pressure) or, in some cases, partial speed towards stonewall. Since this test as normally carried out is a “pass/fail” (e.g. the rotor is stable or not), test conditions should match as close as possible to the field conditions. The parameters of importance include gas MW, power, suction and discharge pressure and temperature, speed and mass flow. If the exact gas composition cannot be tested, some of these parameters will have to be compromised. The rotordynamic model should be used to determine an appropriate combination of factors to produce the maximum instability drivers or minimum log decrement. Miranda and de Noronha [17] propose modifications to the FLFP ASME Type I test to better assess the stability of centrifugal compression equipment. The modifications were intended to create conditions to submit the rotor to instability mechanisms as near as possible to the design conditions rather than reproduce similarity for performance evaluation. The conditions were developed with the aid of the rotordynamic stability predictions.

As with the Mechanical Test, the FLFP test should consist of a warm up portion where the speed is increased gradually to permit stabilization and examination of the behavior at lower speeds. This is followed by an extended run at MCS to ensure thermal equilibrium of the entire machine is achieved. During this test phase, it is recommended that the operating curve at MCS is explored from the surge control line to the end of curve (stonewall.) This operation may include four to five operating points and may include other speeds as highlighted by the rotordynamic analysis. (Note: Other factors may dictate operation at other points as required, i.e. defining the surge line vs. speed, rated point defined at partial speed.)

Factors such as lube oil conditions and rotor assembly are expected to meet the field configuration and specified operating ranges. When practical, the lube oil operating range should be explored during the FLFP test. Lube oil inlet conditions impact the dynamic behavior of bearings. As a critical factor in determining the rotordynamic behavior, it is important to vary these factors over the allowable ranges during testing.

Test Requirements

Test requirements for the FLFP test are defined by agreement between the vendor and purchaser prior to the test and should be done at the contract stage. Holding the overall vibration limit to the level specified for the mechanical run test is impractical due to the additional rotor forces present during the FLFP test. These include aerodynamic forces of the impellers, stator-rotor interactions, seals forces and power transmission forces. However, raising the limit to the vendor

recommended trip setting does not leave margin for off-design operation in the field nor deterioration of the balance state from erosion or deposits. An agreed level should take into account both factors and fall somewhere in between.

Other requirements for the rotor and case vibration may include:

Components of non-synchronous vibration to be less than 20% of the vibration limit or 0.2 mils p-p, whichever is less

No stall related vibration components

No instability related vibration (associated with re-excitation of the 1st natural frequency)

Maximum housing vibration less than 0.1 in/sec

Limitations regarding thermal instability (Morton Effect) vibrations

Test Deliverables

Deliverables are similar to the mechanical run test and should include data for all purchased components tested (as with the string test.) Increased emphasis is placed on the FFT plots of shaft vibration during the test as this is the best indicator of instability, stall, whirl and other phenomena that produce non-synchronous vibrations. Performance data should be recorded during the test to confirm the input used to predict the seal and impeller dynamic behavior and aide further stability analysis if needed.

Rotordynamic Modeling

When FLFP testing is selected, the rotordynamic model should be expanded as necessary to conform to the Level II stability analysis requirements of API 617 [18]. Given the cost, effort and reasons to perform FLFP testing, a Level I stability analysis is insufficient to predict the behavior accurately. The Level II model will reflect changes in the stability level to MW, gas pressures and temperatures, seal clearances and rotor speed to the best of the vendor’s or purchaser’s analytic capabilities. The Level I model uses an empirical relationship that either estimates these effects or doesn’t take them into account at all when calculating the destabilizing forces present in the machine.

Test Knowledge / Verification

The FLFP test will reveal the presence of instability, stall or whirl for a prescribed set of operating conditions for the specific machine’s as-built conditions. The test is pass/fail as no measurement of the stability level or margin is included in the test as described. Rotor stability at different gas compositions, other clearances within the tolerance range or other operating points is determined by rotordynamic predictions. The ability to operate successfully at these alternate points, which cannot be tested under all combinations, depends on the stability margin (not measured by FLFP test) of the machine.


Design Verification Testing

Such verification testing should be considered for the following types of equipment:

Special purpose (as defined by API)

Loss of production produces an unacceptable financial impact

Services or applications with a history of bad actors (vibration related)

Critical service (as defined by the user)

In addition, verification testing of either the unbalance response or stability should be considered when the rotordynamic analysis demonstrates concerns or higher risks associated with the application. The verification testing can be used to mitigate those risks when applied correctly.

Unbalance Response Verification Test (URVT)

Test Procedure

The URVT test is basically a comparison of measured versus predicted vibration levels for the application of a known unbalance. The test is routinely performed following the four hour mechanical run test. For compressors, the verification weight is applied to the coupling flange. This is the only practical location available. For other machinery (e.g., steam turbines, overhung single stage compressors), other locations may be available. Steam turbines may have field accessible balance planes and overhung compressors an impeller checknut with a balance weight placement provision. The measured response of the machine with the verification weight is compared against the analytical prediction using the same weight and location. While this is not a complete check of the analytic unbalance response accuracy, as it only compares the model’s prediction at the probe locations for one weight placement, it is an important first step in ensuring the accuracy of the model.

It was recognized early on that an important aspect of the URVT test was to compensate for the residual unbalance in the machine. This residual unbalance creates the synchronous vibration witnessed during the mechanical test run and is present before and after the verification weight is applied to the rotor. The residual unbalance left in a rotor after balancing (either low or at-speed balancing) is uncharacterized and, therefore, cannot be modeled. Thus, the analytical model will have only the verification weight as an excitation source for the response.

The initial attempt by API 617 6th Edition to compensate for the residual unbalance was to apply a significant verification weight to the coupling to raise the response to the vibration limit. The implication was that the majority of the response would be due to the verification weight placement. This had two important drawbacks (besides being analytically incorrect): First, the amount of unbalance weight at the coupling needed to raise the response to the vibration limit could reach unsafe levels. Coupling flanges are not designed with the intent of adding unbalance weight. Large rotating forces applied to the coupling had a chance of failing the flange or, more probably,

failing the mechanism used to hold the weight in place. Second, the method relied on larger vibration limits

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are provided ref. [19]. The 8 plots the preed at the couessor for one d

Verification WeStiffness

sponse for thare plotted onsince the verisis) was addede largest respotest run was

hronous Responsssor

000 10000 1200

peed (rpm)

t© 2013 by Tu

shaft bendin

onse magnitudred speed(s) – ured value, thesed on the factoClose clearancbe reviewed fo

equirements a

to illustrate thfirst example i

edicted responsupling locatio

drive end probe

eight for Min and

he baseline ann Figure 9). Afication weigh

d to the couplinonse. The sho

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otice that meaht does not resue to the resrming the vectesponse due sared against th

e 10) 8 Stage Co

gure 10 illustual unbalance dupling unbalancured response fthe rotordynam

rement was enal speed) falls e 8. For this exeeded.

he measured ne and verificd example, tcation weight by 25%, Figutical speed dosis, Figure 11. nse magnitude

N unbalance shmits and sensitliance.

igure 11) Measu

Texas A&M En

asured probe reemble the predsidual unbalantor subtraction solely to the vhe predicted res

ompressor – MeaRespon

trates the needue to the relace. Even at lofalls within thmic response e

nforced. In adwithin the ranxample, no cor

response of cation run is the measured is shown to e

ure 12. The moes occur withi

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esponse with tdicted responsnce response of the two da

verification wsponse on Fig 1

asured vs. Predicnse

ed to removeative insensitivw amplitudes, e maximum leespecially at Mdition, the pea

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a three stageshown on Fig

response to exceed the pre

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onse for 3 Stage

periment Station

the verificatione, Fig. 8. Thi

in the rotorata files isolateeight. This i10.

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he verificatioracy of the roon. Beyond tion demonstrctive versus mht placed on thd bunker canding the mod

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: For some well baot be identifiable oll not be appreciang. Performing thment and greater e

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Verification

n test providotordynamic pthe pass/fail

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rotor sensitivi

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Copyright

easured vs. Predinse

the rotordynamguration of the pport stiffness

cts of the bunkld also reflect te assembly) inuld cover thequired by API r the radial bealy, the analysiscritical speed a

information arequirements.)

f the rotordyn

esent the test ve

n for URVT peest on the shop

arings and peance bunker

ement and mag

obes or other lo

tion if probes a

des a determpredictions in observational

the URVT deonse discrepanormed correctlyditional accuri.e., mid-span ity to locations

damped rotors, thas a clear peak in terification weight nce bunker enablest critical speed for

t© 2013 by Tu

icted Verification

mic model fotest setup. If

s of the modeker bearings anthe ½ weight o

n its correct c.ge minimum t617 7th Edition

aring clearances should specifand verificatioas necessary t

namic modelin

ehicle

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gnitude

ocations as use

are mounted o

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emonstrates thncy to a knowy, the operatinacy assuranceversus journa

s other than th

e 1st critical speedthe vibration data placed at the s mid‐span weightr between bearing

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machinmid‐spaand agr

Laboratory, T

lity Verificatio

Methodology

urrently, there nd-user and OEd be used to acttempt to provie three funda

ynchronous excing ratio estime methodologssion will begiestimation.

Damping Ratio

should be noation” has beurement.” Thits:

. The fundamdamping ra

. Estimation processing resultant daonly estima

. Logarithmifor damping

amping ratio efication. Sincine is operatinl analysis whicm identificationas natural freq

ved much greartance with resrn in every fiel

s mentioned ssing analysis tude of estimaging, each techata. Some arense channel, wce. These tec

munities, such aique designed t

is beyond thets of damping machinery engenges and the ron a turbomac

enges when dea

es. All options regan probes, alternareed during contra

Texas A&M En

on Testing (SV

are no industrEM as to whatccurately perfoide some guida

amental elemecitation design

mation. Since thgy requiremenin with some g

o Estimation

oted as to wheen chosen s choice is m

mental modal atio.

of the dampanalysis of

amping ratio iated.

ic decrement isg ratio.

estimation is ace this identng, such testinch is a subset n. Identificatioquency and mater attention. spect to resonald whereas stab

earlier, the eof the measure

ation techniquhnique basicalle very simple

while others curchniques are bas controls andto utilize time

e scope of thratio estimati

gineers undersrecommended pchine, the estialing with the m

garding the use ofate unbalance weigact development w

ngineering Exp

VT)

ry standards int methodologieorm these SVTance in this areents of the ten, measuremenhis last elemennts for the oguidelines rega

hy the term “instead of “lmeant to emp

parameter of

ping ratio invthe measurems not measure

s its own estim

a subset of mtification occung is often cal

of the more gon of other momode shapes, h

This is likelance problems,bility is not as

estimation invement data. W

ues available aly involves a cand deal with

rve-fit multiplebeing develop

d speech procesor frequency d

is tutorial to ion. However, stand some opractices. Whemation procesmeasurement d

f the balance bunkght locations) showith the vendor.

periment Station

n place to guides and practiceTs. This tutoriaea, specificallyesting processnt process, and

nt dictates muchother two, ouarding damping

“damping ratilog decremenphasize severa

f interest is th

volves a postment data. Thed directly, it i

mation techniqu

modal parameteurs while thlled operationageneral field odal parametershas historicallyly due to thei a predominanmuch so.

volves a postWhile there are

and continuallycurve fitting oh just a single channels’ datped by varioussing, with eachdomain data.

discuss all thit is importan

of the peculiaen conducting ss faces severadata:

ker for URVT (e.g.,uld be specified

n

e s

al y, s: d h

ur g

o nt al

e

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The data response lo

The data significant excitationsmisalignm

The data modes, not

The data characterisstructures.

ach of the abbe used for accon a large n

opriate here beavailable at a f

niques assume e dynamic pniques relying turbomachine

metric cross-cou

erhaps the biggurements fromibutions from

e of interest. Muencies which mation procesacteristic, rotorard whirling amity and share

he easiest wayr mode pairs is the Campb

pressor. Typicae sisters remaine the backwardard mode. The rated with spescopics and sence of gyrosndent on the

mple, unlike thees for an overd due to theent.

is typically ocations.

contains addpresence

s within the ment, aerodynam

likely containt just the mode

is from astics are not lik

ove challengecurate stability

number of senecause responsfew restricted lthe system be

properties are on such an aswhere fluids wupling in the sy

gest estimationm a rotor’s SVT

several modeMany structure

can be diffis. Sharing

r systems posseand one backwe similar modal

y to observe is through a Cbell diagram al for betweenn very close in d mode’s frequsecond mode seed, indicatingsupport anisotscopics and s

overall rotore between-bearhung compres

e overhung im

Copyright

from only a

ditional respoof immeasurmachine such

mics, etc.

ns contributione of interest.

a system wke those for sta

s influences wy estimation. Snsors, meaningse measuremenlocations on a ing tested is a

symmetric ossumption are within the machystem [21,22].

n challenge is tT often consistes, more than es possess vercult to identithis high m

ess pairs of sisward that are l characteristic

the close proxCampbell diagfor a multistn-bearing macfrequency as s

uency is slightsisters becomeg the strongetropy for thessupport anisotr system charing compresso

ssor will be vempeller’s stro

t© 2013 by Tu

few vibratio

onse from thrable, interna

h as unbalance

n from multipl

whose dynamiatic, nonrotatin

what techniqueome techniqueg they are nonts are typicallmachine. Othe

a static structuror self-adjoinnot appropriat

hine create non

the fact that thts of significanjust solely th

ry close naturafy through thmodal densitster modes, onoften in clos

cs.

ximity of thesgram. Figure 1age centrifuga

chines, the firspeed increasetly less than th

e more and morer influence ose sisters. Thtropy is highlracteristics. Foor, the first sisteery sensitive tong gyroscopi

urbomachinery

on

he al e,

le

ic ng

es es ot ly er re

nt. te n-

he nt he al he ty ne se

se 13 al st s,

he re of he ly or er to ic

Tothe asforwarundersby thepeak nonsydirectiwhirlidampiphase and sspecia

Unappliesister the Sbehavmodescirculapeak imodesand, imode.

Laboratory, T

Figure

o date, many psumption that rd mode, is pstand why thise industry. Fir

in the frequynchronous excion with the ing mode. Anding estimationslope or mec

simple formulalist.

nfortunately, eed and only onbackward mod

SDOF dampinvior for a sims’ response canar excitation isis measured, ths are respondinin this case, b.

Texas A&M En

13) Example Ca

previous SVT only the mode

participating ins assumption rst, the measuruency responcitation was apintent of only

d finally, singln techniques, schanical log delas that are

even when fone peak is presde can be excing estimators.

mulated rotor sn be distinguishs being appliehe backward mng, instead of jbroadened, fro

ngineering Exp

ampbell Diagram

investigators e of interest, tyn the responsehas historicallrements often nse functions.pplied in the foy exciting thele degree of fresuch as ampliecrement, invofamiliar to e

orward circulasent in the meaited, reducing t Figure 14 system where hed. Even thou

ed and only a smode is excitedjust one, the p

om that due to

periment Station

m

have relied onypically the firse. It is easy toly been appliedonly show on. Second, thforward circulae first forwareedom (SDOFification factorolve convenienevery vibration

ar excitation iasurements, ththe accuracy oillustrates thithe individua

ugh the forwarsingle responsd. Because twoeak is distorteo just a singl

n

n st o d e e

ar d

F) r, nt n

is e

of is al d e o d e

Fi

Westimconsicomm

Badoptwill iown tto enrotordtestin

gure 14) Modal

While the methmation are stil

dered good munity, that hav

Multiple provide thdamping rabout a characteristechniquesfor stabilibackward measuremefrequency

To resolveMDOF teoutput (Mmultiple lo

Any candvalidated measuremefor shop tesystem whThis validasymmetrystability. Wsuch validthe model

ecause of the tion of estimaindependently trusted methodnd-users’ indedy

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