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The objective of the following work was to determine frequency-dependent rotordynamic force coefficients for a new annular gas damper seal design. Both rotating and nonrotating experimental tests are presented for inlet pressures at 1000psig(69bar), a frequency excitation range of 20-300Hz, and rotor speeds up to 15,200rpm. Two different testing methods were used for determining coefficients: (1) dynamic pressure response method and (2) mechanical impedance method. The dynamic pressure method required the measurement of internal seal cavity pressure modulations in combination with the vibratory motion, whereas the mechanical impedance method used the measurement of external shaker forces, accelerations, and motion of the mechanical system. In addition to the new fully partitioned damper seal (FPDS) tests, the same experiments were conducted for a conventional pocket damper seal (PDS) design. Results of the frequency-dependent force coefficients and the internal seal dynamics for the two different gas damper seals are compared. The conclusions of the tests show that the FPDS design possesses significantly more positive direct damping and direct stiffness compared to the conventional PDS. The experiments also show the measurement of same-sign cross-coupled (cross-axis) stiffness coefficients for both seals, which indicate that the seals do not produce a destabilizing influence on rotor-bearing systems.

The turbomachinery component of interest in this paper, the pocket damper seal, has the dual purpose of limiting leakage and providing an additional source of damping at the seal location. The rotordynamic coefficients of these seals (primarily the direct stiffness and damping) are highly dependent on the leakage rates through the seals and the pressures in the seals’ cavities. This paper presents both numerical predictions and experimentally obtained results for the leakage and the cavity pressures of pocket damper seals operating at high pressures. The seals were tested with air, at pressures up to 1000psi(6.92MPa), as the working fluid. Earlier flow-prediction models were modified and used to obtain theoretical reference values for both mass flow rates and pressures. Leakage and static pressure measurements on straight-through and diverging-clearance configurations of eight-bladed and twelve-bladed seals were used for code validation and for calculation of seal discharge coefficients. Higher than expected leakage rates were measured in the case of the twelve-bladed seal, while the leakage rates for the eight-bladed seals were predicted with reasonable accuracy. Differences in the axial pitch lengths of the cavities and the blade profiles of the seals are used to explain the discrepancy in the case of the twelve-bladed seal. The analysis code used also predicted the static cavity pressures reasonably well. Tests conducted on a six-bladed pocket damper seal to further investigate the effect of blade profile supported the results of the eight-bladed and twelve-bladed seal tests and matched theoretical predictions with satisfactory accuracy.

This paper presents measured frequency dependent stiffness and damping coefficients for 12-bladed and 8-bladed pocket damper seals (PDS) subdivided into four different seal configurations. Rotating experimental tests are presented for inlet pressures at 69 bar (1000 psi), a frequency excitation range of 20–300 Hz, and rotor speeds up to 20,200 rpm. The testing method used to determine direct and cross-coupled force coefficients was the mechanical impedance method, which required the measurement of external shaker forces, system accelerations, and motion in two orthogonal directions. In addition to the impedance measurements, dynamic pressure responses were measured for individual seal cavities of the eight-bladed PDS. Results of the frequency dependent force coefficients for the four PDS designs are compared. The conclusions of the tests show that the eight-bladed PDS possessed significantly more positive direct damping and negative direct stiffness than the 12-bladed seal. The results from the dynamic pressure response tests show that the diverging clearance design strongly influences the dynamic pressure phase and force density of the seal cavities. The tests also revealed the measurement of same-sign cross-coupled (cross-axis) stiffness coefficients for all seals, which indicate that the seals do not produce a destabilizing influence on rotor-bearing systems.

The primary objective was to determine, through experimental testing and rotordynamic computer simulations, the radial stiffness of an angular contact ball bearing with inner race static misalignment and outer race dynamic misalignment under various axial loads. Two different testing methods were used to determine the radial stiffness of the ball bearing: 1) eigenvalue frequency analysis and 2) critical speed transition tests. Each method used experimental measurements in conjunction with computer simulation to determine the radial ball bearing spring rate. Experimental measurements of the forward eigenvalues, backward eigenvalues, backward critical speeds, and forward critical speeds were matched with computer model simulations to determine the stiffness values for each of the cases. The rotordynamic characteristics of the rotating assembly suggested significant nonlinearity of the test bearings and also revealed signs of bifurcation for higher degrees of static misalignment.

It is a common practice for cryogenic turbomachines to utilize stiffly mounted bearings due to the incapability of providing significant damping with conventional methods and designs. With no damping available, the historical motive behind this design practice was to elevate all critical speeds above the maximum running speed. The desire for higher energy density has raised the running speeds of rocket engine turbopumps very near, or even above, the first three critical speeds. A companion paper gives experimental evidence that accurate prediction of these critical speeds with the turbopump rotor on ball bearings with stiff supports is not practically possible. The feasibility of applying a metal mesh damper with soft supports in a cryogenic engine is investigated. To date the focus has been on cryogenic experimental tests, rotordynamic simulations, and possible bearing support design schemes to justify the incorporation of a metal mesh bearing support in a liquid rocket engine fuel turbopump. Success of the design is dependent on the amount of damping that the damper retains at cryogenic temperatures and its ultimate effect on fuel turbopump rotordynamics, especially subsynchronous instability.

Metal mesh is a commercially available material used in many applications including seals, heat shields, filters, gaskets, aircraft engine mounts, and vibration absorbers. This material has been tested by the authors as a bearing damper in a rotordynamic test rig. The test facility was originally used to support the design of a turboprop engine, developing squirrel cages and squeeze film dampers for both the gas generator and power turbine rotors. To design the metal mesh damper, static stiffness and dynamic rap test measurements were first made on metal mesh samples in a specially designed nonrotating test fixture. These property tests were performed on samples of various densities and press fits. One sample was also tested in an Instron machine as an ancillary and redundant way to determine the stiffness. Using the stiffness test results and equations derived by a previous investigator, a spreadsheet program was written and used to size metal mesh donuts that have the radial stiffness value required to replace the squirrel cage in the power turbine. The squirrel cage and squeeze film bearing damper developed for the power turbine rotor was then replaced by a metal mesh donut sized by the computer code. Coast down tests were conducted through the first critical speed of the power turbine. The results of the metal mesh tests are compared with those obtained from previous testing with the squeeze film damper and Show that the metal mesh damper has the same damping as the squeeze film at room temperature but does not lose its damping at elevated temperatures up to 103 °C. Experiments were run under several different conditions, including balanced rotor, unbalanced rotor, heated metal mesh, and wet (with oil) metal mesh. The creep, or sag, of the metal mesh supporting the rotor weight was also measured over a period of several weeks and found to be very small. Based on these tests, metal mesh dampers appear to be a viable and attractive substitute for squeeze film dampers in gas turbine engines. The advantages shown by these tests include less variation of damping with temperature, ability to handle large rotor unbalance, and the ability (if required) to operate effectively in an oil free environment. Additional testing is required to determine the endurance properties, the effect of high impact or maneuver loads, and the ability to sustain blade loss loads (which squeeze films cannot handle).

Pocket damper seals perform a dual function: both sealing the pressurized gas around a rotating shaft and providing large amounts of vibration damping. The annular cavity between the labyrinth seal teeth is subdivided into separate annular cavities around the circumference of the rotor by partitioning walls. Also, the upstream and downstream teeth have different radial clearances to the rotor. These seals have been shown to provide a remarkable amount of direct damping to attenuate vibration in turbomachinery, but they generally leak more than conventional labyrinth seals if both seals have the same minimum clearance. Conversely, brush seals allow less than half the leakage of labyrinth seals, but published test results show no significant amount of damping. They are considered to be a primary choice for the seals in new aircraft engine designs because of their low leakage. This paper will describe a recently invented hybrid brush/pocket damper seal that combines high damping with low leakage. Previous brush seal results were studied and calculations were made to select a brush seal to combine with the pocket damper design. The result is a hybrid seal with high damping and low leakage. A special design feature can also allow active vibration control as a bonus benefit. A computer code written for the original pocket damper seal was modified to include the brush element at the exit blade. Results from the computer code indicate that the hybrid seal can have less leakage than a six bladed (or 6 knives) labyrinth seal along with orders of magnitude more damping. Experimental evaluations of the damping and leakage performance of the hybrid seal are being conducted by the authors. The experimental work reported here tested the damping capability of the new hybrid brush seal by exciting the seal journal through an impedance head. A conventional six-bladed labyrinth seal of the same working dimensions was also tested. The brush hybrid pocket damper seal is found to leak less than the labyrinth seal while producing two to three times more damping than the original pocket damper seal, (orders of magnitude more than the conventional labyrinth).

Selection criteria and design evaluations of several types of bearing dampers with active control for application to aircraft engines were described in a companion paper (Vance, Ying, and Nikolajsen, 1999). A disk type electrorheological (ER) damper was chosen for further study and testing. The results of the tests and the final conclusions of the study are described in this paper. Experimental results including stiffness and damping coefficients are presented for the ER bearing damper with two types of ER fluid, 350 CS and 10 CS (centistokes) viscosity. The vibration attenuation performance of the ER damper was measured on a rotordynamic test rig in the form of free vibration decay, rotor orbits, and runup unbalance responses. The results show that the ER fluid with lower viscosity has the better characteristics for rotordynamic applications. It was found that ER fluids produce both Coulomb and viscous damping. If only the damping is considered, the Coulomb type is less desirable, but with active control it can also achieve control of rotor stiffness as analyzed in Vance and San Andres (1999). A feedback control system was developed and applied to the ER damper with the objective of improving the overall rotordynamic performance of the rotor bearing system, considering both vibration amplitudes and dynamic bearing forces. A “bang-bang” (on and off) simple control logic was found to work better in practice than more sophisticated schemes. The measured runup responses of the rotor-bearing system with this control approximated the desired vibration response curves fairly well. The tests highlighted some of the practical considerations that would be important for aircraft engine applications, such as the ER fluid limitations, the electrical power supply requirements, the electrical insulation requirements, the nonlinear relationship between the voltage and the damping, and the relative benefits of active control. It is concluded that active control of bearing damping is probably not a practical improvement over the passive squeeze film dampers currently used in most aircraft gas turbine engines.

Experimental rotordynamic force coefficients and leakage for a four-blade, two-four pocket gas damper seal are presented and compared to predictions based on a one control volume bulk-flow model. The test rig comprises a vertical shaft and a test seal housing and flexible structure suspended from a rigid centering frame. The experiments were conducted at increasing rotor speeds to 6000 rpm and inlet/exit pressure ratios from 1.0 to 3.0. The seal force coefficients are obtained from impact response measurements of the seal and flexible structure using a frequency domain parameter identification technique. Both measurements and predictions show the seal direct stiffness and damping coefficients are proportional to the inlet/exit pressure ratio and insensitive to rotor speed. The agreement between experimental results and analytical predictions is acceptable. Predicted cross-coupled stiffness coefficients are of small amplitude. However, the test results evidence cross-coupled stiffnesses without journal rotation due to a structural asymmetry induced by the external pressurization into the seal. [S0742-4787(00)04201-6]

A bulk-flow model for calculation of the dynamic force characteristics in a single cavity, multiple-pocket gas damper seal is presented. Flow turbulence is accounted for by using turbulent shear stress parameters and Moody’s friction factors in the circumferential momentum equation. Zeroth order-equations describe the isothermal flow field for a centered seal, and first-order equations govern the perturbed flow for small amplitude rotor lateral motions. Comparisons to limited measurements from a four-pocket gas damper seal show the current model to predict well the mass flow rate and the direct damping coefficient. For a reference two-bladed teeth-on-stator labyrinth seal, the current model predicts similar rotordynamic coefficients when compared to results from a two control-volume bulk-flow model. Force coefficients from a reference single-cavity, four pocket gas damper depend on the rotor speed and pressure drop with magnitudes decreasing as the rotor whirl frequency increases. The multiple-pocket gas damper seal provides substantially more damping than a conventional labyrinth seal of the same dimensions. The damper seal cross-coupled stiffness coefficients are small though sensitive to the inlet circumferential pre-swirl flow.Copyright © 1998 by ASME

Background: There has been an abundant flow of information written about problem‐based curricula in the basic sciences. However, there is little collective information about its use in the clinical setting. This article reviews the literature concerning the implementation of problem‐based learning (PBL) in the clinical setting and draws conclusions regarding its misconceptions and feasibility. Summary: The authors found numerous articles that described PBL programs in undergraduate, graduate, and continuing medical education, but only a handful met the criteria outlined in Barrows’ taxonomy of PBL. Many programs produced interesting innovations in clinical education, but they should not be labeled PBL. Conclusions: When writing about PBL, authors should describe how their program fits or departs from a PBL taxonomy and provide enough detail to permit others to replicate it. In addition, there remains a dire need for more experimental studies to warrant the extensive use of problem‐based learning in the cl...

This is the first (Part I) of two papers describing recent results of the research program directed at developing a vibration damper suitable for high-temperature turbomachinery applications. It is expected that such dampers will replace squeeze-film dampers, which use oil as the working fluid and have limitations at higher temperatures. A novel gas-operated bearing damper has been evaluated analytically and experimentally for its damping characteristics. A theory based on the isentropic assumptions predicts the damper performance characteristics reasonably well. A maximum damping level of 2311 N-m/s (13.2 lb-s/in.) at a frequency of 100 Hz was measured with a single actuator of the gas damper. Since many such actuators could be placed circumferentially around the squirrel cage, considerable damping levels can be realized. The study also shows that significantly higher damping levels can be achieved by modifying the current design. Part I describes the theoretical model that has been developed based on isentropic assumptions. This model is an improved version of the previous theory (Vance et al., 1991) and includes the supply groove effects, dynamic area changes of the inlet feeding holes, and the effects of flow choking on damper behavior. The governing equations are derived and theoretical predictions using these equations have been made for two hardware designs that were experimentally investigated (see Part II for experimental results).

This is the second of two papers describing results of a research project directed at developing a gas-operated vibration damper for high-temperature turbomachinery applications. This part presents the experimental measurements made on a gas damper hardware and compares them with the theoretical predictions given in Part I. It is found that the isentropic theoretical model predicts the damper characteristics quite well. A maximum damping of 2310 N-s/m (13.2 lb-s/in.) was measured at a natural frequency of 118 Hz using the present design and the results suggest that significantly higher damping levels are possible with design modifications.

A gas-operated bearing damper for turbomachinery has been designed, analyzed, and experimentally investigated in the laboratory. The damper utilizes air bled off from the compressor to power an actuator through orifices with area modulated by the vibratory displacement at the bearing support. The design objective for this passive device is to make the actuating dynamic gas pressure phase lead the vibratory displacement by 90 deg. Several variations of the basic concept have been tested. An analysis was performed to guide the experiments. All of the designs tested to date can produce positive damping, and one particular design has produced a damping coefficient of 8756 N-s/m (50 lb-sec/in.) with a power penalty of 5.2 kW (7 hp) at 310 KPa (45 psi). This design was installed on a laboratory rotor with flexibly supported ball bearings, and significant damping of the critical speed response was demonstrated. The experimental results to date suggest that further research can produce significant improvements in performance, and the device appears to be especially adaptable to high-temperature applications for aircraft engines.

The effect of vapor cavitation on the pressure distribution and the force coefficients of a long squeeze film damper executing circular centered orbits is studied in association with the Swift-Stieber cavitation boundary conditions. The existence of vapor cavitation significantly decreases the whole of the pressure distribution. A modified Reynolds equation is solved analytically for a long squeeze film damper to investigate the effect of fluid inertia on cavitated pressure profiles. Increasing fluid inertia lends to extend the region of cavitation to the minimum film gap and to reduce the pressure. The corresponding damping and inertia force coefficients are presented for varying positions of cavitation inception and termination. Experimental comparisons are presented in Part II. Presented at the 45th Annual Meeting in Denver, Colorado May 7–10, 1990

A solution of a cavitated long squeeze film damper executing circular centered orbits, which is obtained from the Swift-Stieber cavitation conditions, is modified considering side leakage of end seals. The cavitated pressure profiles and force coefficients for orbit radii equal to e=0.47 and 0.76 are measured from a squeeze film damper test rig, and compared with the analysis. The Swift-Stieber conditions and leakage factor used can improve the prediction of cavitation effects and end leakage effects on the performance of a squeeze film dampe. Presented at the 45th Annual Meeting in Denver, Colorado May 7–10, 1990

Experimental measurements of pressure distributions and force coefficients obtained from a squeeze film damper lest rig executing a circular centered orbit are presented. The test rig has been designed to study the effect of fluid inertia on the pressure field and dynamic force response on a damper configuration with a relatively large clearance. Past measurements of the squeeze film damper force characteristics have been carried out at squeeze film Reynolds numbers not exceeding a value equal to 10. In the present paper, following contemporary applications, operations at Reynolds numbers up to fifty are tested for CCO's with an orbit radius = 0.8 (Re ≤ 10 at e = 0.5). The results obtained from a fully open ended damper are presented in detail. The effects of fluid inertia, cavitation and the open end geometry on the pressure distributions and force coefficients are discussed. Presented as a Society of Tribologists and Lubrication Engineers paper at the ASME/STLE Tribology Conference in Toronto, Ontario, ...

In Part I, a squeeze film damper (SFD) test rig and measurement procedures were explained, arid the experimental results obtained from an open ended damper were presented. In this paper, the experimental results measured from a partially sealed SFD test rig executing a circular centered orbit are presented and discussed. A serrated piston ring is installed at the damper exit. This device involves a new sealing concept which produces high damping values while allowing for oil flow to cool the damper. In the partially sealed damper, large cavitation regions are observed in the pressure fields at orbit radii e = 0.5 and e = 0.8. The cavitated pressure distributions and the corresponding force coefficients are compared with a cavitated bearing solution. The experimental results show the significance of fluid inertia and vapor cavitation in the operation of squeeze film dampers. Squeeze film Reynolds numbers tested reach up to Re = 50, spanning the range of contemporary applications. Presented as a Society of ...

This paper presents a theoretical model for analytical study of internal friction-induced whirl instability due to shrink fit and interference fit joints in rotating machines. It is the theoretical continuation of an experimental study on the effects of internal friction published in a companion paper. The theoretical study shows that the follower force internal friction model proposed by Gunter in 1965 is a physically incorrect model due to its inconsistency with Newton’s Third Law of Motion. The internal moments model proposed by Lund in 1976 to explain unstable sub-synchronous vibrations due to micro-slip in interference and shrink fit joints is physically consistent with the basic principles of mechanics and can explain rotordynamic instability due to internal friction in rotating machines. Rotordynamic simulations of the experimental single-disk and two-disk rotor based on an internal moments model are carried out using finite element based rotordynamics software. Rotordynamic simulations of the unstable experimental rotors using the linear viscous form of the internal moments model, with selected moment stiffness and damping parameters at the interface, result in correct prediction of the instabilities (negative real part of the damped eigenvalues).

This paper presents an experimental study of internal friction effects on two different test rotors with shrink fit and interference fit joints. One of the setups was a single-disk rotor with an interference fit joint to the main shaft through a mechanically adjustable tapered sleeve; the other setup was a two-disk rotor with a variable cross-section shaft, having the two disks shrunk fit into an aluminum sleeve at the ends. The slippage in these types of joints is known to be a potential source of sub-synchronous rotordynamic instability in rotating machines. The experimental results from the rotor setups show the presence of strong unstable sub-synchronous rotor vibrations initiated at operating speeds above the first critical speeds of the rotor-bearing systems. The experiments with the two-disk rotor demonstrated that different distributions of shrink fit at two locations (loose fit at one end; tight fit at the other end) were necessary to produce rotordynamic instability. The instability in that case caused complete wreckage of the test rig, showing that the sub-synchronous vibrations caused by slippage at interference fit joints are potentially dangerous to the safe operation of rotating machines. Furthermore, a critical radial interference (shrink fit) of 1 mils (25.4 μm) at speed was identified to cause an onset of instability in several of the experiments.Copyright © 2008 by ASME

Rotordynamic instability, commonly observed as subsynchronous vibration, is a serious problem that can cause heavy damage to a turbomachine or make it incapable of operation due to high vibration levels. However, all subsynchronous vibrations are not necessarily unstable. A way to quickly diagnose them would be helpful. In an earlier paper, the authors presented data from experiments that simulated various causes of sub-synchronous vibrations, some causes being genuine rotordynamic instabilities and some others being benign (stable), and identified ways to diagnose and classify the subsynchronous motions. In a continuation of the same study, subsynchronous vibrations due to coupled lateral-torsional effects are experimentally simulated, the objective being to signal-analyze these vibrations to find unique signatures that identify this cause and also be able to recognize if they are a true rotordynamic instability or not. To this end, a test rig was built with parallel shafts coupled by gears, driven by a DC motor at one end and loaded at the other end, to closely simulate a real-world machine. A torsional mathematical model for the test rig is also presented to predict its torsional natural frequencies. Experiments were conducted wherein the first torsional natural frequency was externally excited, with the shaft spinning at a higher speed. The result was a false sub-synchronous “instability” signal in the lateral measurements. A method to distinguish these vibrations from a genuine lateral non-synchronous instability is presented. Also, a new diagnostic method to classify the subsynchronous vibration as benign is elucidated.Copyright © 2008 by ASME

This paper analyzes the effects of air entrainment and cavitation on the synchronous response of squeeze film supported rigid rotors. The fluid film force coefficients are obtained from experimental measurements corresponding to a wide spectrum of operating conditions. These conditions include regimes in which air entrainment effects are dominant. Other conditions where vapor cavitation and fluid inertial effects are dominant are included for comparison. The effects of air entrainment are shown to produce a nonlinear response representative of a softening spring effect not previously known to exist in squeeze film dampers.

Air entrainment in squeeze film damper bearings results in a type of cavitation that is considerably different from the cavitation traditionally assumed for hydrodynamic bearings. The fluid lubricant in this case is highly compressible and results in a drastic influence on the pressure distribution. This describes and experimental procedure for determining the density correlation for and oil-air mixture. This correlation is then utilized in the Reynolds equation to determine the influence lubricant compressibility has on the pressure distribution in a squeeze film damper bearing. Presented at the 35th STLE/ASME Tribology Conference in Fort Lauderdale, Florida October 16–19, 1989

High speed motion pictures have revealed several operating regimes in a squeeze film, damper. Pressure measurements corresponding to these distinct regimes were made to examine their effect on the performance of such dampers. Visual observation also revealed the means by which the pressure in the feed groove showed higher amplitudes than the theory predicts. Comparison between vapor and gaseous cavitation are made based on their characteristic pressure wave, and the effect this has on the resultant force in the squeeze film. Presented at the 44th Annual Meeting in Atlanta, Georgia May 1–4, 1989

Rotordynamic instability can be disastrous for the operation of high speed turbomachines in the industry. Most ‘instabilities’ are due to de-stabilizing cross coupled forces from variable fluid dynamic pressure around a rotor component, acting in the direction of forward whirl and causing subsynchronous orbiting of the rotor. However, all subsynchronous whirling are not unstable and methods to diagnose the potentially unstable kind from the benign are critical to the health of the rotor-bearing system. In this study, methods to demarcate between the two are detailed. Orbit shape, “frequency tracking” and agreement of subsynchronous frequencies with known eigenvalues are used as diagnostic tools. It is shown that a change in synchronous phase angle produced by de-stabilizing cross coupled forces can be used as a definitive indicator of incipient instability. Typical signatures of subharmonic vibrations induced from non-linear stiffness of the rotor- bearing system are examined analytically and through experiments.Copyright © 2007 by ASME

Rotordynamic instability, commonly observed as subsynchronous vibration, is a serious problem that can cause heavy damage to a turbomachine or make it incapable of operation due to high vibration levels. However, all subsynchronous vibrations are not necessarily unstable. If the amplitude of the subsynchronous vibration is large, it can cause damage to seals, bearings, or process wheels. If it is small, the question arises as to whether it has the potential to grow larger (“instability”) or whether it is benign and harmless. A way to know would be helpful. The objective of this study is to signal-analyze subsynchronous vibration in turbomachinery and distinguish benign subsynchronous vibration from true rotordynamic instability. Effort is also made to identify unique signatures to a cause, thereby aiding in faster diagnosis. A computer simulation study is conducted on four rotors, including two gas-reinjection compressors that went unstable, to examine the possibility of using the change in synchronous phase angle as a possible indicator of impending instability. Two other rotors in laboratory test rigs were also studied, both experimentally and with computer simulations of the phase angle response. The computer simulations and experimental results agree very closely on the test rigs. Measured signals from another rotor, with bearings having a dead-band clearance are studied as an example of a benign cause that can result in subsynchronous vibration. The effect is studied with the rotor in both horizontal and vertical positions and clear indicators are discovered that confirm the subsynchronous vibration to be benign in nature. Signatures from another rotor with internal friction are also presented. The experimental data clearly shows indicators that distinguish it as a genuine instability. Dry friction whip is also experimentally produced on a test rig. The measurements show that dry friction whip defies all general rules of thumb for diagnosing a true instability and at the same time is a very violent one as well.Copyright © 2007 by ASME

The effects of two seal design parameters, namely blade (tooth) thickness and blade profile, on labyrinth seal leakage, as well as the effect of operating a seal in an off-center position, were examined through a series of non-rotating tests. Two reconfigurable seal designs were used, which enabled testing of two- four-, and six-bladed see-through labyrinth seals with different geometries using the same sets of seal blades. Leakage and cavity pressure measurements were made on each of twenty-three seal configurations with a four inch (101.6 mm) diameter journal. Tests were carried out with air as the working fluid at supply pressures of up to 100 psi-a (6.89 bar-a). Experimental results showed that doubling the thickness of the labyrinth blades significantly influenced leakage, reducing the flow-rate through the seals by up to 20%. Tests to determine the effect of blade-tip profile produced more equivocal results, with the results of experiments using each of the two test seal designs contradicting each other. Tests on one set of hardware indicated that beveling blades on the downstream side was most effective in limiting leakage whereas tests on newer hardware with tighter clearances indicated that seals with flat-tipped blades were superior. The test results illustrated that both blade profile and blade thickness could be manipulated so as to reduce seal leakage. However, an examination of the effects of both factors together indicated that the influence of one of these parameters can, to some extent, negate the influence of the other (especially in cases with tighter clearances). Lastly, for all configurations tested, results showed that leakage through a seal increases with increased eccentricity and that this phenomenon was considerably more pronounced at lower supply pressures.Copyright © 2007 by ASME

One of the main contributors to rotordynamic instability in modern day turbomachinery is cross-coupled stiffness. The terminology comes from the phenomenon in which the reactive stiffness force is normal to the rotor displacement. Cross-coupled stiffness (CCS) can be generated from fluid pressure forces in internal machine components such as gas/liquid seals, hydrodynamic bearings, axial flow turbomachinery stages, and centrifugal impellers. Some of these machine components exhibit a type of CCS that materializes as a follower-force, which is normal to the whirl orbit radius and therefore tangent to a circular whirl orbit. With the follower force collinear with the whirl velocity, energy is fed into the motion and makes it grow larger with time. This destabilizing force can be represented in a Newtonian coordinate system by opposite-sign CCS, Kxy = −Kyx. However, there are cases of CCS where the forces produced are not always in the direction of rotor whirl, and the net work of the force per cycle is zero. These CCS forces are purely distortive to the orbit and are not de-stabilizing. They are represented by same-sign CCS, Kxy, Kyx, not necessarily equal. This paper investigates the effects of same-sign CCS on rotordynamics. Experiments are presented that show the measurement of same-sign CCS on a pocket damper seal and the required testing scheme to determine the effect on rotordynamic stability.

In a previous ASME paper the second author reported experiments on wire mesh bearing dampers (WMD) incorporated in a power turbine rotor-bearing system in order to enable a direct comparison between WMD and squeeze film dampers (SFD). The results showed that both WMD and SFD perform equally well for reducing the rotordynamic amplitudes of vibration. Moreover the WMD were found to have significant advantages over SFD. The damping provided by the wire mesh is independent of temperature changes and presence of turbine oil. Experiments by another investigator showed that WMD are capable of sustaining more than twice the unbalance as compared to SFD, which promises possible application to withstand blade loss loads. This paper presents empirically developed non-dimensional design equations for WMD, capable of predicting stiffness and damping for a wire mesh ‘donut’ subject to changes in various design, installation, and operational parameters.Copyright © 2005 by ASME

A model of impact of a tennis ball with a flat surface is developed based on a planar, two-mass, linear, four degree of freedom vibration system idealization. The impact is assumed to be incident on a flat surface with friction. The incident parameters of the ball include the centre of mass translational velocity, angle of impact with the surface and the incident angular spin of the ball. The linear, piecemeal vibration model predicts the corresponding rebound parameters of the tennis ball. The model also predicts the duration of contact of the tennis ball with the flat surface, the transition of motion of the tennis ball during contact with the ground from sliding to rolling contact, and the resulting contact forces developed between the tennis ball and the flat surface. The model is computationally efficient because the governing differential equations of motion are linear and their standard solutions can be easily implemented on a personal computer. Predictions of the rebound parameters from the model are compared with experimental findings on tennis balls which are incident on a flat surface with various angles, velocities and angular spins (zero spin, topspin and backspin). For selected parameters of the two-mass model, the comparisons show excellent agreement between the model and the measurements.

Pocket damper seals (PDSs) have been shown to offer an improved rotordynamic performance compared to conventional labyrinth seals, providing large amounts of direct damping and small cross-coupled coefficients. To this date, the rotordynamic characteristics of the PDSs have been determined using linearized models, with validity limited to small displacements of the rotor. The present investigation determines the effect of the nonlinear forces arising from the pocket damper seal on the response and stability of a PDS-rotor system when subjected to large dynamic excitations. It is determined that the nonlinear forces maintain bounded amplitudes of vibration (“limit cycles”) after the rotor passes the threshold speed of instability predicted by the linear theory. It is also found that the critical speed of the PDS-rotor system is well predicted by using both linear and nonlinear models of the seal forces, as both predictions compare well with the experimental measurements.Copyright © 2003 by ASME

Pocket damper seals can produce large amounts of damping in turbomachines where the working fluid is compressible. Even with non-optimum design configurations, they have solved rotordynamic instability problems in high-pressure compressors. They do not rely on viscous effects, so they are ideal for oil-free applications. It is shown here that proper sizing of the seal pockets can maximize the damping. Significant parameters and variables have been identified, both experimentally and theoretically, that reflect the fundamental effects of the pocket damper seal to produce rotordynamic damping. Dimensionless analysis was performed, resulting in significant Pi groups involving damping, stiffness and leakage parameters. Of particular interest is a Helmholtz Pi group that shows 1) the benefit from optimizing the volume of the pockets around the seal, and 2) the frequency dependence of the seal. Dimensionless prediction equations are shown graphically. The graphs were generated by a computer code that uses a single control volume model. The code has been verified experimentally on several different seal test rigs with data taken at different vibration frequencies spaced an octave apart.Copyright © 2002 by ASME

A previous paper described measurements of damping and leakage of a brush hybrid seal in a nonrotating test apparatus. In those experiments most of the damping came from the brush. This paper describes further experiments with this seal in a rotordynamic rig designed to determine if the bristles will still contribute significant damping at large amplitudes after wearing in to a line-line contact. These tests continued to show much more damping and much less leakage than a labyrinth seal of the same dimensions. The damping performance is even better after the brush is worn in. In the rotating experiments it was discovered that the brush elements also produce a surprising amount of friction torque and consequent power loss, especially at high pressure drops. During the experiments it was found that the brush friction is destabilizing to backward whirl, and that support asymmetry suppresses the instability.Copyright © 2001 by ASME

In a previous ASME paper experiments were reported on metal mesh bearing dampers (MMD) that were tested in a power turbine rotor at speeds up to 12,000 rpm. They were made of 0.229 mm stainless steel 304 wire mesh, compressed to 57% density, which is close to the maximum density that was economically available. After balancing, a level of vibration was achieved similar to that previously observed with squeeze film dampers. These experiments showed that the MMD could suppress vibration amplitudes of the 22.7 kg rotor at critical speeds of 4,000 rpm and 9,300 rpm. Much of the testing showed the rotor having little or no response to unbalance on coastdown through the critical speeds. The donut-shaped MMD in those tests were the only bearing supports; no squirrel cages were used. A question was raised about the feasibility of using MMD in parallel with a squirrel cage bearing support so that the stiffness can be controlled independently of the damping. This paper presents experimental results for metal mesh dampers with a squirrel cage as a parallel bearing support. Experiments with copper mesh as seal elements (on another project) had indicated that copper mesh has higher damping than stainless steel, so copper was chosen for these experiments. Both a linear viscous damping model and a hysteretic damping model were investigated. Some hysteretic damping models predict that damping depends on stiffness. A different hysteretic model turned out to be useful and promising as a prediction model for two reasons: a) it fits the measured data, and b) it predicts that the damping is not lost if the MMD is put in parallel with a steel structure such as a squirrel cage bearing support. The measurements reported here support the validity of that prediction.Copyright © 2001 by ASME

This paper describes some of the requirements for bearing dampers to be used in an aircraft engine and briefly discusses the pros and cons of various types of dampers that were considered as candidates for active control in aircraft engines. A disk type of electrorheological (ER) damper was chosen for further study and testing. The paper explains how and why the choice was made. For evaluating potential applications to aircraft engines, an experimental development engine (XTE-45) was used as an example for this study. Like most real aircraft engines, the XTE-45 ran through more than one critical speed in its operating speed range. There are some speeds where damping is desirable and other speeds where it is not. Thus, the concept of a damper with controllable forces appears attractive. The desired equivalent viscous damping at the critical speeds along with the available size envelope were two of the major criteria used for comparing the dampers. Most previous investigators have considered the ER damper to produce a purely Coulomb type of damping force and this was the assumption used by the present authors in this study. It is shown in a companion paper (Vance and San Andres, 1999), however, that a purely Coulomb type of friction cannot restrain the peak vibration amplitudes at rotordynamic critical speeds and that the equivalent viscous damping for rotordynamics is different from the value derived by previous investigators for planar vibration. Control laws for Coulomb damping are derived in Vance and San Andres, (1999) to achieve minimum rotor vibration amplitudes in a test rig while avoiding large bearing forces over a speed range that includes a critical speed. The type of control scheme required and its effectiveness was another criterion used for comparing the dampers in this paper.

Attempts have been made in the past to use Coulomb damping for vibration suppression in rotating machinery. Typically, a dry friction damper is designed to operate on a flexible bearing support. These designs have usually been unsuccessful in practice, partly because the Coulomb coefficient changes with temperature, with ingress of dirt or lubricant, and with the surface wear conditions. It is known that purely Coulomb damping forces cannot restrain the peak rotor whirl amplitudes at a critical speed. The invention of a disk type of electroviscous damper, utilizing a fluid with electrorheological (ER) properties, has recently revived the interest in Coulomb type dampers. Several investigations have suggested that a Coulomb friction model was the best representation for an ER damper with voltage applied. This model was used to study the feasibility of developing actively controlled bearing dampers for aircraft engines. This paper analyzes the imbalance response of two different rotordynamic models with Coulomb friction damping and shows the benefit of adding active control. Control laws are derived to achieve minimum rotor vibration amplitudes while avoiding large bearing forces over a speed range that includes a critical speed. The control laws are derived for purely Coulomb type of damping and assuming a combination of Coulomb and viscous damping effects. It is shown that the most important feature of Coulomb damping for minimal rotordynamic amplitude response is the control of rotor support stiffness, i.e. leading to the relocation of critical speeds, rather than control of a damping coefficient.Copyright © 1999 by ASME

Measurements on a large rotor dynamic test rig in the Turbomachinery Laboratory with a concrete and steel foundation have shown some critical speeds which can only be predicted if the foundation is included in the model. Mobility transfer function measurements on the foundation were made to obtain the required parameters for including the foundation in the rotordynamic computer model. These measurements eliminate the necessity for building a large finite element model of the foundation structure and allow the use of fast transfer matrix codes for modeling rotor bearing systems which have strong foundation participation. A method first described by Nicholas was used to get a set of equivalent bearing coefficients from a combination of the foundation parameters with the true bearing coefficients. Measurements made by the present authors show that the modal damping of the foundation can be obtained easily from the mobility curves, and that the improved critical speed predictions can be obtained in some cases by using only one foundation mode to calculate the equivalent bearing coefficients.Copyright © 1997 by ASME

A swirl damper intended for high temperature turbomachinery applications was experimentally investigated for its damping characteristics. The results were disappointing in that only a small amount of damping was obtained. Instead, an instability was produced at a speed which exhibited backward whirl. The rotor rig that was used exhibited backward whirl over a speed range between two critical speeds. This backward whirl could be prematurely induced by the use of the swirl gas damper. At high damper pressures, the rotor system exhibited an instability. The exact nature of this instability is not known, however, possible causes are presented. The observations discussed in this paper are also important to those who design swirl brakes or anti-swirl guide vanes for labyrinth seals. The principles involved are similar in that the flow into a labyrinth seal is turned to swirl anti-rotationally as it enters the seal, same as the swirl gas damper.

A new type of labyrinth gas seal for damping vibration and whirl, called the TAMSEAL, has been evaluated in both non-rotating and rotating tests at Texas A&M University. Test results of the prototype, along with comparison tests of a conventional labyrinth seal, show up to one hundred times more direct damping than the conventional bladed seal. The new design also has a feature that blocks swirl of the working fluid, which is known to be rotordynamically destabilizing in machines with conventional seals. Coastdown tests of the new seal were conducted at various pressures on a rotordynamic test apparatus with a critical speed at 4000 rpm and compared with identical testing of a conventional labyrinth seal. Rap tests of both seals were also conducted to measure the logarithmic decrement of free vibration, and the leakage of both seals was measured. Test results show large reductions in peak vibration at the critical speed in all cases, with the critical speed being completely eliminated by the TAMSEAL at some pressure drop conditions. The leakage rate of the tested TAMSEAL is higher than the conventional seal at the same clearance, but the large reductions in vibration and whirl amplitudes suggest that the TAMSEAL could be operated with smaller clearances than conventional labyrinth seals.Copyright © 1995 by ASME

Experimental measurements have been made to evaluate the rotordynamic performance of straight-through labyrinth seals under conditions that are realistic for many turbomachines. Both teeth-on-rotor and teeth-on-stator gas seals were tested, each with twelve blades, 173 mm (6.8″) blade diameter, and 102 mm (4″) total length. The nominal blade tip clearance was 0.5 mm (20 mils). The teeth-on-stator seal was tested with the blade tip clearances diverging (in the direction of the flow), uniform, and converging. The teeth-on-rotor seal was tested with uniform clearances. The inlet air pressure to the seals was varied from 1.7 bar to 14.6 bar (25 psi to 200 psig) with the last blade exhausting to the atmosphere. Coastdown tests of all the seals were performed on a rotordynamic test rig to show their effect on synchronous response to imbalance when passing through a 3700 rpm critical speed. For the teeth-on-rotor seal, rap tests at 4500 rpm were also conducted to measure the effective damping coefficient for subsynchronous vibration. The synchronous response to imbalance was generally increased by all the seals at inlet pressures up to about 11.2 bar (150 psig). The worst case was for the teeth-on-rotor seal at about 2.7 bar (35–45 psi) inlet pressure where the rotor whirl amplitude was increased from .1 mm (3.75 mils, peak to peak) to over .13 mm (5 mils). In most cases the rotor whirl amplitude was slightly decreased at inlet pressures above 13 bar (176 psig). The teeth-on-rotor seal provided a small amount of damping to attenuate the 61 Hz subsynchronous vibration with the rotor running at 4500 rpm. A computer model which includes both the rotor and housing dynamics was developed to evaluate the possible range of values of the rotordynamic seal coefficients. Simulations show that the effective subsynchronous damping coefficient of the teeth-on-rotor seal ranges from 175 N-s/m at 5.1 bar inlet pressure (1 lb-s/in at 75 psi) to 876 N-s/m at 10.2 bar (5 lb-s/in at 150 psi). This corresponds to a range of 0.3% to 1.4% of critical damping added by the seal for subsynchronous vibration, even though the seal increased the synchronous response at the critical speed. It is shown that the orbit conditions for the synchronous and subsynchronous tests were radically different, as they likely will be in most turbomachines.

A new type of labyrinth seal for damping vibration and whirl in turbomachinery has been invented and tested at Texas A&M University (TAMU). Test results of the prototype, along with a conventional labyrinth seal, show that the new design has fifteen times more direct damping than the conventional bladed seal. The design also has a feature that blocks swirl of the working fluid, which is known to be destabilizing in machines with conventional seals. This paper shows how the new design is different from conventional bladed seals and honeycomb seals, describes the test apparatus and results, describes a theory which explains the test results, and explains how the seal can be much more effective than a bearing damper for reducing vibration in turbomachines with a compressible working fluid.

This is the second of two papers describing results of a research project directed at developing a gas operated vibration damper for high temperature turbomachinery applications. This part presents the experimental measurements made on three variations of the gas damper hardware and compares them with the theoretical predictions presented in Part I. It is found that the isentropic theoretical model predicts the damper characteristics quite well. A maximum damping of 13.2 Ib-s/in was measured for a single actuator at a natural frequency of 100 hz using the present design and the results suggest that significantly higher damping levels are possible with design modifications.

This is the first (Part I) of two papers describing recent results of a research program directed at developing a vibration damper suitable for high temperature turbomachinery applications. It is expected that such dampers will replace squeeze-film dampers that use oil as the working fluid and have limitations at higher temperatures. A novel gas operated bearing damper has been analytically and experimentally evaluated for its damping characteristics. Theory that is based on the isentropic assumptions predicts the damper performance characteristics reasonably well. A maximum damping level of 13.2 lb-s/in at a frequency of 100 hz was measured with a single actuator of the gas damper and, since many such actuators can be placed circumferentially around the squirrel cage, considerable damping levels can be realized. The study also shows that significantly higher damping levels can be achieved by modifying the current design. Part I describes the theoretical model that has been developed based on isentropic assumptions. This model is an improved version of the previous theory [1,2] and includes the supply groove effects, dynamic area changes of the inlet feeding holes and the effects of flow choking on damper behavior. The governing equations are derived and theoretical predictions using these equations have been made for the three hardwares that were experimentally investigated (see Part II).

Sixteen graduate students in mechanical and aerospace engineering were supported on this grant, producing four M.S. graduates and three Ph.D. graduates during the grant period. Eight students are still pursuing research and studies toward their degree. Two students failed the Ph.D. qualifying examination and left Texas A and M. Six of the program participants are now employed by companies in the aircraft or aerospace propulsion field.