[American Institute of Aeronautics and Astronautics 44th AIAA/ASME/SAE/ASEE Joint Propulsion Conference & Exhibit - Hartford, CT ()] 44th AIAA/ASME/SAE/ASEE Joint Propulsion Conference

Innovative High-Temperature Compliant Surface Foil Face Seal Development

Hooshang Heshmat* and James F. Walton II† Mohawk Innovative Technology, Inc., Albany, NY 12205, USA

An innovative compliant surface foil face seal being developed for aerospace applications will be described and discussed in this paper. Due to the requirements of gas turbine applications, this compliant surface face seal was designed to accommodate large system thermo-mechanical excursions in all three dimensions. The systematic design methodology and key governing flow equations used to establish the design and predict performance will be presented and discussed. Additionally, the experimental parametric evaluation that was conducted to verify feasibility of the design concept will be presented. Test parameters varied included seal diameter, seal land length to radius ratio, axial preload and corresponding gap heights, surface velocity, and differential pressure. A total of six different seal configurations were tested having outer diameters of 97, 112, and 229 mm. Projected surface area pre-load levels up to 145 kPa, surface velocities from 0 to 425 m/s and differential pressures up to and greater than 140 kPa were used. Seal leakage was measured for each test condition, normalized to flow factor and compared to the developed theoretical/empirical model. Results from the experimental program show that excellent sealing performance and correlation with the theory was achieved. Based on the excellent correlations, flow factor projections to 690 kPa were made. A review of the high temperature seal design and forthcoming testing at temperatures to 800C using Korolon™ coating and a high temperature superalloy substrate will also be presented.

Nomenclature b = breadth C = Clearance D = Diameter g = gravitational constant ho = reference film height under no load condition hfilm = film height hb = bump height hnew = film height under preload condition h = non-dimensional film height L = radial path length (Ro-Ri) m& = mass flow P = pressure Pa = ambient pressure Pu = upstream pressure, absolute p = non-dimensional pressure

Q = volumetric flow r = non-dimensional radius Ri = inner radius Ro = outer radius Re = Reynolds Number U = universal gas constant * President/CEO & Technical Director, 1037 Watervliet-Shaker Rd, Albany, NY 12205.

American Institute of Aeronautics and Astronautics

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† Vice President of Program Development, 1037 Watervliet-Shaker Rd, Albany, NY 12205, and AIAA Member.

44th AIAA/ASME/SAE/ASEE Joint Propulsion Conference & Exhibit21 - 23 July 2008, Hartford, CT

AIAA 2008-4505

Copyright © 2008 by the American Institute of Aeronautics and Astronautics, Inc. All rights reserved.

T = temperature θ = angular extent ρ = density μ = dynamic viscosity ψ = flow factor ω = angular velocity Λ = (6 μ ω/ Pa ) (Ro / hfilm )

I. Introduction uel efficient, durable, affordable and maintenance free gas turbine engines for both future and existing aircraft are needed given the present cost of fuel. While considerable effort continues to be expended to improve pressor, combustor and turbine performance, engine designers must also develop advanced mechanical

components, such as gas path seals. Advanced engine seals have shown promise of reducing losses and maintaining these performance benefits over the service interval of the engine [1]. However, to achieve these benefits as surface speeds and ambient temperatures increase, new technologies that result in extremely lightweight design concepts, with reduced leakage and new materials are needed. For axial or face seals, complex mounting schemes have been implemented with some success but, they remain a troublesome maintenance and overhaul component. A key face seal limitation is its susceptibility to variations in surface flatness and the waviness that results from manufacturing imperfections as well as distortions induced due to thermal gradients, centrifugal forces and pressure gradients [2]. Even the newer designs, such as the non-contacting segmented seal, while showing some success, are subject to coning and misalignment concerns [3]. Damage tolerant, high-performance designs are, therefore, needed to accommodate the high differential pressures, extreme temperatures, high surface speeds and rotor excursions. The seal must also minimize undesirable internal airflow and/or air/oil mist distribution with its attendant negative impact on performance and possible engine failure. Advanced seals must also accommodate large rotor excursions; whether due to axial thermal gradients, hard maneuvers or emergency powered rotor starts. Attempting to meet these stringent requirements with fixed geometries has generally resulted in design compromises with larger than desired steady state operating clearances. Although brush seals have demonstrated significantly lower leakage than conventional designs, contacting seals are subject to wear, which degrades their effectiveness with time [4,5,6]. Good brush seal performance requires a smooth, wear resistant runner interface and wear resistant bristles. Most importantly, brush seals cannot continue to operate under large rotor excursions. Experience has shown that the bristles plastically deform when subjected to large deflections, thereby changing clearance and increasing leakage over time. Additionally concerns related to brush seal durability and damage during construction and installation have been reported [5]. In particular, reverse rotation of the shaft during installation can damage the seal bristles. Thus, being a contacting seal performance of the brush seal is a function of time, wear rate, shaft excursions, environment and the tribo-material system.

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What makes the development of advanced engine seals very challenging is the fact seals must perform well under large differential centrifugal and thermal growth of the rotor while also accommodating large gyroscopic loads. Adding to the challenge is the fact that the requirements for this seal fall outside of the existing experience base when temperature, surface velocity and size are all combined into requirements for a single new seal. Finally, since existing brush or labyrinth seals are limited to maximum deflections of less than 0.75 mm without encountering serious degradation and leakage, the compliant foil seal (CFS) was developed. In this paper a CFS that can accommodate differential pressures to 690 kPa, surface velocities to 335 m/s, large axial motions and temperatures approaching 538°C will be presented. The current compliant foil seal technology is developed based on the advanced technology of a compliant gas lubricated foil bearing [7-10]. The compliant bump foil elements are

F

Shaft

CompliantBumpSprings

05-0246C

HighPressure

LowPresureSide

SmoothFoils

Figure 1. Cross section view of compliant foil shaft seal.


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designed to provide the desired spatial structural stiffness needed to accommodate both the hydrodynamic/hydrostatic pressures as well as rotor motions. Additionally, the bump elements are capable of providing additional system damping for enhanced system stability [9,11-13]. Figure 1 shows the basic configuration of a circumferential or shaft CFS, while the compliant foil face seal is shown in Figure 2 and Figure 3. In each design a thin, high-pressure gas film is formed between the surface of a rotor (journal or runner) and the smooth top foil of the CFS due to hydrodynamic action. The multiple compliant layers shown in Figure 3 are included and designed to permit large axial motions of the rotor while also deforming in a manner that enhances the seal performance. Recent studies have shown that the circumferential shaft CFS performs well in a high temperature section of a small gas turbine engine simulator, even under large rotor excursions [11-13], giving credence to the face seal concept.

Smooth Foils

Bump Foils

Smooth Foils

Bump Foils

Figure 2. Compliant foil face seal solid model.


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The CFS computational analysis was presented previously in which the combined hydrodynamic and structural governing equations were solved simultaneously using a combination of finite difference and relaxation methods. Both laminar and turbulent flow conditions were considered. Turbulence functions, which contained the Reynolds numbers of the flow in circumferential and axial direction, were used to account for turbulence in the governing fluid flow equations that had been modified to include structural compliance [13]. The effect of pressure on the viscosity was also considered; and subsequently the thermal effects were added [14].

Given this background, it is evident that there is a need for a highly efficient, long-life seal that is capable of maintaining consistent performance especially when encountering large excursions. The limitations and inability of existing seals to meet these demands logically leads one to consider novel axial face seal design approaches. Compliant foil bearing (CFB) technology has great application and development potential for use in gas turbine engine face seals [8,15-17]. The CFB technology foundation that has resulted in significantly increased load capacities and damping values has been instrumental in the development of the radial compliant foil seal as reported by Salehi and Heshmat [12-14], in which the successful operation of a 72 mm CFS in a simulated hot section of a small gas turbine engine at speeds to 50,000 rpm and temperatures to 1094°F (590°C) was demonstrated. Subsequent testing of 150 mm and 216 mm diameter seals to 30,000 rpm and 540°C was also completed, including testing of the 216 mm compliant foil seal by NASA, (see Figure 4).

Thin Gas Film Gap ~ .002”

KH

Axial Motion

Smooth Foils

RotatingDisk/Runner

HydrodynamicPressure

Disc OuterDiameter

Disc InnerDiameter

PHigh

Compliant Bump Foil Supports

Smooth Foil Axial Motion

K1K2K4 K3

Separator Foils

Secondary Seals

Thin Gas Film Gap ~ .002”

KH

Axial Motion

Smooth Foils

RotatingDisk/Runner

HydrodynamicPressure

Disc OuterDiameter

Disc InnerDiameter

PHigh



K1K2K4 K3

Separator Foils

Secondary Seals



K1K2K4 K3

Separator Foils

Secondary Seals

Figure 3. Compliant foil face seal operating schematic diagram.

While the above discussion was directed at radial seals, the advances made in foil seals and bearings also benefit axial face seals and address limitations present in current face seals. Unpredictable failures in face seals result in high maintenance costs. One of the challenges for mechanical face seals is the intermittent contact between the stator and the rotor. This intermittent contact results in one of the failure modes exhibited by higher harmonic oscillation [18]. While there are models and analyses of the mechanical face seal in the literature [19-23], there is little evidence of developments directed at large compliant surface seals capable of sustaining large axial motions. As such, a proof of concept compliant foil face seal was tested. The test articles were based on the compliant foil thrust bearing which features:

1. A hydrodynamic pressure profile, which is more uniform over the surface of the seal 2. Spatially variable stiffness to accommodate rotor axial motion, misalignment and distortion 3. High axial load capacity designed to accommodate large axial steady state and transient motions 4. Light weight and low parts count for high frequency response and reliability

II. Test Method and Descriptions


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A. Test Item Description Six different configurations of proof of concept foil

face seals were fabricated in order to assess the impact of flow path radial length, axial preload and surface velocity on leakage. The six test articles are shown in Figure 5 through Figure 10. Two 9-inch OD thrust foil bearings, two 4.37-inch OD and two 3.82-inch OD configurations were fabricated providing different L/Ro ratios, angular gaps between pads and different flow paths. For each test seal, the outer periphery of the compliantly supported foil pads was open to atmosphere, thereby presenting a leakage path along the radius as opposed to the closed ends shown in Figure 3. Figure 11 and Figure 12 schematically show the tested configurations with the open ends and the primary flow paths. The importance of the open ends for these initial tests was to determine the baseline resistance to flow due to the total axial gap (htotal) in the angular segments between pads, the gap beneath the bump foils (hb) and the gap between the top smooth foil and the disc (hfilm), all without the end flanges and secondary seal elements to restrict flow. This would allow for an assessment of critical design parameters for the fundamental seal shape, such as an assessment of the importance of L/Ro ratio. Additionally, tests of the candidate seals with this arrangement and conducted under rotating conditions would, when compared to static/non-rotating tests, verify that the hydrodynamic pressures were generated and reduce total leakage. As shown in Figure 2 and Figure 3, both the angular gap and open ends will be eliminated in the final configuration, thereby only allowing leakage flow to pass through the minimum film height (hfilm). By eliminating the larger gaps associated with the pad angular spacing and the region beneath the compliant bumps in the face seal configuration, the leakage will be substantially reduced from the measured baseline configuration. It should be noted that during testing the total gap height was on the order of 0.031 inch, whereas hfilm was either zero when static tests were conducted or on the order of 0.001 inch when dynamic testing was conducted at speeds from 24,000 to 60,000 rpm. While hfilm initially increases during dynamic testing (see Figure 11), the generated hydrodynamic film pressure resists the radial inflow/outflow of high pressure air. Thus, the air is forced to flow behind the top smooth foil and through the passages formed by the bumps (approximately 0.021 inch high) as well as the gaps between individual pads. It should also be noted that with the high pressure at the OD, the inward directed pressure driven flow will also be restricted by the inherent outward self pumping action of the disc. Finally, while the gap between pads is approximately 0.031 inches high and between 6° and 10°, flow in this gap is turbulent, even for differential pressures as low as 2 psig. Thus, when the end flanges are introduced at the OD and the pads overlap one another, the primary leakage path will be through the very narrow hydrodynamic film region, which, at about 0.001 inch, will result in leakage rates well below any present technology.

B. Test Method

Figure 4. 216 mm compliant foil seal flow factor as measured at NASA

Two rigs were used to assess leakage characteristics of the baseline compliant surface axial face seals, as shown in Figure 13 and Figure 14. Testing was initially conducted under non-rotating conditions to assess the influence of static preload on leakage for each configuration. The static test procedures called for installing the test seal into the rig; applying an axial preload, introducing pressurized air into the seal cavity. Measured parameters included air mass flow, air pressure, axial preload force, and gap between the seal and runner. Following static testing and data analysis, two configurations with different L/Ro ratios were tested under conditions from 24,000 to 60,000 rpm in the test rig Figure 13. The L/Ro ratios and angular gaps tested are listed below:

Outer Diameter L/Ro ratio Angular Gap (Deg) 3.88 0.516 10 3.88 0.327 10 4.37 0.351 6 4.37 0.223 6 9 0.665 10.5 9 0.665 3.5

Figure 5 Nine pad 9-inch seal with offset pad spacing

Figure 6 Eleven pad 9-inch seal with pads inline

2.84”

4.37”6°

2.84”

4.37”6°

Figure 7 4.37" diameter seal with L/R=0.35

3.397”

4.37”6°

3.397”

4.37”6°6°


1.85”

3.88”10°


2.57”

3.88”10°

2.57”

3.88”10°



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Qi, Pi

hTotal

hfilm

Hydrodynamic Pressure

hbump

hbump >> hfilm

r

z

ID

ODRunner

Figure 11 Flow paths in proof of concept face seal

h filmh total

h film << h total 07-0026

Runner/Disk

Figure 12 Sketch of test article gaps as viewed from OD to ID for fully open seal without end flange or overlapping top foils.

Figure 13. Test rig used to assess leakage of 3.88 inch and 4.0-inch diameter test items.

Figure 14. Test rig with 9-inch diameter test item installed.

C. Data Analysis Method The preliminary model used to assess leakage flow is presented in equation 1 below.

3

0

9/44/7

14)(7.1 ⎟⎟

⎠

⎞⎜⎜⎝

⎛⎥⎦

⎤⎢⎣

⎡Δ=

hh

PL

Dq newπ Equation 1

The above equation is based on the Reynolds flow equation but has been simplified so that the influence of the

critical parameters and their impact on design could be readily evaluated. As seen in the above equation, the key parameters are the axial gap, which is, in reality, comprised of both the total gap beneath the seal surface (hbump) and the gas film height (h film) (See Figure 11 and Figure 12). Since flow is affected by the ratio of film height to total gap to the 3rd power, this parameter will be crucial to reducing the leakage flow through the seal once the ends of the hydrodynamic compliant foil pads are designed to close off the radial flow, as shown in Figure 3. It should also be noted that the angular radial gap was not included in the preliminary analysis above. In the final design, the radial angular gap will be eliminated in the hardware by having the pads overlap each other. Additionally, as noted previously, the ends of the pads will be wrapped around the end plat to seal off the gap behind the top smooth foil that is created by the bump foils, leaving only the film height (h film) as the primary leakage path.

In order to be able to compare data between the different seal configurations and account for the size differences data was normalized to flow factor was used. Where Flow Factor is defined as:


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[ ]uPDTm /.

Overall, the importance of and reason for conducting these tests was to determine the impact of radial path length on the leakage since total radial space is often limited.

Comparison of Measurement and Theory4.37" Diameter Seal with Different Length Radial Flow Path

Pressure (psi)-2 0 2 4 6 8 10 12 14 16

Mass Flow

(lb/min)

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3

4

5

Full Length Pad T1R2,3,4Full Length Q_Theory Half-Length Pad T2R2Half Length Theory

Comparison of Measurement and Theory4.37" Diameter Seal with Different Length Radial Flow Path

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Mass Flow

(lb/min)

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5

Full Length Pad T1R2,3,4Full Length Q_Theory Half-Length Pad T2R2Half Length Theory

III. Test Results Based on the data analysis and comparison to the

theoretical model (Equation 1), theory and experiment correlate quite well, especially for the seals with L/R ratios of 0.327 and 0.223 as seen in Figure 15, Figure 16 and Figure 17. When comparing the correlation of the 3.88” full radial length pad data to analysis, the discrepancy becomes large when differential pressure is high. It should be understood that the primary cause for this discrepancy is that the gap between pads is much greater for the 3.88-inch diameter seal than for the 4.37 inch diameter seal. For example, the gap on the 3.88-inch seal is 10° between each pad, whereas for the 4.37-inch diameter seal, the pad gap is only 6° or a 40% reduction.

As seen in Figure 18 and Figure 19, increasing pad radial length reduces leakage flow as expected, since in equation (1) flow is inversely proportional to L. Leakage was also affected by increased axial preload, though less noticeably. The increased axial preload acts to reduce both h total and h bump, hence, the leakage. Thus, the combination of increased path length (L/Ro) and axial preload will contribute to reduced flow. Thus, with this data subsequent parametric analysis, the proper directions for the most efficient design will result. Figure 20 and Figure 21 are composite plots of both measured and predicted flow factors for all static tests of the 4.37 and 3.88 inch OD seals with L/Ro =0.351 and 0.516 for preloads from 80 to 130 lbs, respectively. As seen, the predicted values, which are based upon measured inlet pressure and axial gap, closely match the measured flow factors. When comparing flow factor for cases with lower L/Ro ratios, as shown in Figure 22, flow factor is slightly higher than for cases with larger L/Ro ratios. This is to be expected since flow is inversely proportional to length (L) in equation (1), above. Further confirmation of this relationship is seen in Figure 23 for the two different configurations of the 9-inch diameter seal. Of most interest is the difference between the measured flow factors for the 9-inch seals with inline and offset inner and outer pad rings. Also that the predicted flow factor for the offset pads under predicts flow factor. Upon review, the experimental results are not unexpected since there was a considerably larger angular gap between pads for the 9-pad design than for the 11-pad design. Specifically, the seal with the inner ring having 9-offset pads (Figure 5), had angular gaps of 10° for the inner ring and 11° for the outer ring, whereas the seal with 11-inline

Figure 15 Comparison of measurement and theory for 4.37" OD seal with L/R = 0.351 and 0.223.

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10

12

0 5 10 15 20 25

Mass Flow (lb/min)

Pressure (psi)

T3R1_Flow Qplt-plt_T3R1 Qplt-plt_T4R2 T4R2_Flow Figure 16 Static comparison of measurement and theory for 3.88" OD seal for L/R=0.516 and 0.327.

Comparison of Measurement and Theory4.37" versus 3.88" Diameter Half Pad Seals

Pressure (Psi)-2 0 2 4 6 8 10 12 14

Mass Flow

(lb/min)

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0

1

2

3

4

T2R2 Half-Length 4.37"T2R2 Half Length TheoryT4R1_Half Length 3.88"T4R1 Half Length Theory

Comparison of Measurement and Theory4.37" versus 3.88" Diameter Half Pad Seals

Pressure (Psi)-2 0 2 4 6 8 10 12 14

Mass Flow

(lb/min)

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1

2

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T2R2 Half-Length 4.37"T2R2 Half Length TheoryT4R1_Half Length 3.88"T4R1 Half Length Theory

Figure 17 Seal leakage for 4.37 and 3.88 inch seals with L/R=0.223 and 0.327 respectively


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pads (see Figure 6) had a 3° angular gap for the inner ring and 4° for the outer ring. However, the current model under-predicted the flow factor when using measured data for pressure and geometry. It may be that in systems, such as that tested with larger areas with full height gap, the analysis should be modified to account independently for both the sections relating to the pads and then, also, the sections relating to the gaps.

Pressure (psi)-5 0 5 10 15 20 25

Mass Flow

(lbm/m

in)

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1

2

3

4

T1R2 4.37 OD, Len=0.767" Preload=90 lbT1R3_4.37" OD Len=0.767" Preload=100 lbT1R4 4.37 OD Len=0.767" Preload=130 lbT3R1 3.88 OD Len=0.985" Preload=121 lbsT3R3 3.88 OD Len=0.985" Preload=98 lbs

Length = 0.985 inch

Length = 0.767 inch

Pressure (psi)-5 0 5 10 15 20 25

Mass Flow

(lbm/m

in)

-1

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1

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3

4

T1R2 4.37 OD, Len=0.767" Preload=90 lbT1R3_4.37" OD Len=0.767" Preload=100 lbT1R4 4.37 OD Len=0.767" Preload=130 lbT3R1 3.88 OD Len=0.985" Preload=121 lbsT3R3 3.88 OD Len=0.985" Preload=98 lbs

Length = 0.985 inch

Length = 0.767 inch

Figure 18 Leakage for L/R ratios of 0.516 and 0.351 on 3.88 and 4.37 inch OD seals respectively

Pressure (psi)0 2 4 6 8 10 12 14

Mass Flow

(lbm/m

in)

0

1

2

3

4

T2R2 4.37 OD Len=0.488 Preload=125 lbT4R1 3.88 OD Len=0.625" Preload=68 lbs

Len=0.488"

Len=0.625"

Pressure (psi)0 2 4 6 8 10 12 14

Mass Flow

(lbm/m

in)

0

1

2

3

4

T2R2 4.37 OD Len=0.488 Preload=125 lbT4R1 3.88 OD Len=0.625" Preload=68 lbs

Len=0.488"

Len=0.625"

Figure 19 Leakage for L/R ratios of 0.327 and 0.223 for 3.88 and 4.37 inch OD seals respectively

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0 2 4 6 8 10 12 14 1

Flow Factor

Pressure (psi)

4.37" OD Seal, 0 rpm, L/ R=0.351, Preload=80 to 130 lb

6

T1R2_Ffid T1R2_Ff_80-100lb T1R3_Fflow_100 lb T1R3_FFid T1R4_Fflow_130 lb T1R4_Ffid Figure 20 Flow Factor for OD=4.37” and L/R=0.351 under 80 to 130 lb preload at 0 rpm.

Comparison of Measured and Predicted Flow Factor for L/R=0.516

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Fac

tor

FFid_T3R1 T3R1_Fflow T3R3_P T1R3_Fflow Figure 21 Flow Factor for OD=3.88" and L/R=0.516 for 98 and 121 lb preload at 0 rpm

Having completed static tests that confirmed pressure drop across the seal and excellent flow restriction, the

impact of operating speed/surface velocity on flow was evaluated next to verify that the generation of the hydrodynamic pressure profile within the seal would increase resistance to flow, as hypothesized. Figure 24 through Figure 26 present plots of measured and predicted flow factor versus upstream pressure for tests of the 3.88” OD face seal with an L/R ratio of 0.516. The theoretical Reynolds flow equation was modified based upon in-situ empirical data and used for these correlations. Measured data points are shown as open red circles. In each case, the measured flow factor asymptotically approaches a flow factor of 0.02 with increasing differential pressure. The predicted flow factors, which are slightly higher than measured at approximately 0.025 to 0.027, are shown as the dark blue line and use the measured differential pressure, seal geometry and differential change in film height. Thus, the predictions are very reflective of the actual test conditions and show that the proposed closed form Reynolds based flow analysis is valid for use in preliminary design of the compliant foil face seal. While this Phase I analysis did not account for the fully open channel between the individual face seal pads (a 0.031 inch h total gap versus a 0.021 inch h bump gap), an analysis of the flow in the gaps showed that Reynolds numbers greater than 3000, and often approaching 30,000, indicating very turbulent flow was to be expected, hence, part of the reason for the rapidly falling flow factor with increasing pressure. Following the assessment of leakage flow as a function of speed, an additional test was done at a fixed speed of 40,000 rpm while compressive preload was varied from


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approximately 20 to 150 lb. As seen in Figure 27, once preload approaches and exceeds about 80 pounds, the measured flow factor remains fairly constant at about 0.02 and the predicted flow factor is about 0.027. This indicates that the stiffness of the fluid film and the bump foils has increased sufficiently, such that there were very minor changes in the gap and film height with further change in compressive preload. Thus, from equation (1) with constant film height and differential pressure, we would not expect a change in flow, as is evident in Figure 27. This information also shows that the external pressure was not sufficient to overcome the resistance introduced by the generated hydrodynamic pressure between the runner and the top smooth foil.

Flow Factor Comparison for 3.88" OD and 4.37" OD Seals

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Pressure (psi)

Flow

Fac

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T4R1_Fflow T4R2_Fflow FFid_T2R2 T2R2_Fflow FFp_T4R1 T4R2_FFP

OD=4.37" and L/R=0.223 - Measured

OD=3.88" and L/R=0.327 - Measured

OD=3.88" and L/R=0.327 - Predicted

OD=4.37" and L/R=0.223 - Predicted

Figure 22 Flow factor for 4.37 and 3.88 " OD seals with L/R=0.223 and 0.327 respectively

Comparison of 11-Pad Flow Factors Under 10.5 and 13.2 psi pad loadingvs 9-Pad Inner ring with 9.5 psi Pad Preload for

9-inch OD

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FFid_UniT5R5 UniT5R5_Fflow FFid_UniT5R6 UniT5R6_Fflow FFid_UniT4R1 UniT4R1_Fflow

9-Pad Offset radial leakage paths - Predicted

11-Pad Inline radial leakage paths - Predicted

9-Pad Offset radial leakage paths - Measured

11-Pad Inline radial leakage paths - Measured

Figure 23 Flow Factor for 9 inch OD Seal with multi-rows inline vs. offset L/R=0.665

Flow Factor Full Seal 24 krpm, 100 lb Preload

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FFid_T6R2_24Krpm T6R2_Ff_meas_24Krpm

Figure 24 Flow Factor vs. differential pressure for a 3.88” OD seal with an L/R of 0.516 when operating at 24krpm with a 100 Lb Preload

Flow Factor Full Seal 24 krpm, 100 lb Preload

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Fac

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FFid_T6R2_24Krpm T6R2_Ff_meas_24Krpm

Figure 25 Flow Factor vs. differential pressure for a 3.88” OD Seal with an L/R of 0.516 when operating at 40 krpm with a 100 Lb Preload

Following the dynamic foil seal tests with an L/R ratio of 0.516, a 3.88” OD foil seal with an L/R ratio of 0.327

was tested. In order to maintain a consistent pad loading with the lower projected area, testing was conducted under lower loads of 68 and 80 lbs but at the same speeds of 24,000, 40,000 and 60,000 rpm. These test speeds yielded surface velocities at the periphery of the seal pads of 400, 667 and 1000 ft/sec. Figure 28 through Figure 33 show individual tests conducted under each load speed combination. As seen in these figures, the measured flow factor is slightly above 0.02 at the highest pressures tested. Interestingly, for the tests with the lower L/Ro ratio, correlation between the predicted and measured flow factors appears better than for the higher L/Ro ratio. The improved correlation between the predicted and measured loss factor, in this case, is due most likely to the reduced leakage path length for the seal with the lower L/Ro ratio.


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Full 3.82 Seal 100 lb Preload 60krpm

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FFid_T6R4_60Krpm T6R4_Ff_meas_60Krpm Figure 26 Flow Factor vs. differential pressure for a 3.88” OD Seal with an L/R of 0.516 when operating at 60 krpm with a 100 Lb Preload

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Flow Factor vs Preload at 40krpm

160

Series1 Series2 Series3 Figure 27 Flow Factor vs. preload for a 3.88” OD Seal with an L/R of 0.516 when operating at 40 krpm with a differential pressure of 10-16 psi

Flow Factor Half 3.82" Pad at 24krpm 68 lb preload

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FFid_T7R2_24Krpm T7R2_Ff_meas_24Krpm Figure 28 Predicted and measured Flow Factor vs. differential pressure for a 3.88” OD Seal with an L/R of 0.327 when operating at 24 krpm with a 68 Lb Preload

Flow Factor 80 lb Load, 24krpm, half pad

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Flow Factor vs Pressure for T7R3-40krpm 68 lb load, L/R=0.327

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Flow Factor vs Pressure T7R3 40 krpm 80 lb preload, L/R=0.327

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In order to gain additional insight into the test results, plots of flow factor versus surface velocity, L/R ratio and preload, respectively, were generated (see Figure 34 through Figure 36). In each of these plots it is evident that, for those tests conducted with the shaft spinning, leakage was lower than for static or non-rotating conditions. Further, leakage was lower for larger L/R ratios and was lower for higher compressive preloads. Each of these observations was as expected, confirming the hypothesis that hydrodynamic effects would be at work in the seal design to reduce leakage by providing an additional resistance to flow. For example, a review of the leakage flow model (equation 1) shows that flow is inversely proportional to path length (L) and is reduced by reductions in the ratio of axial gap height to the third power. What is not shown in this closed form Reynolds equation is the influence of the hydrodynamic pressure generated as a function of surface velocity (see Figure 11 and Figure 12). In essence, the hydrodynamic generated pressure will act to reduce the effective gap between the disk and the top smooth foil and will put an additional resistance to leakage flow, thereby, reducing total leakage flow factor. Thus, given the trends observed and the compilation plot of all static tests shown in Figure 37, it is highly probable that large face seals having flow factors approaching and even less than 0.01 will be possible.

Flow Factor vs Pressure T7R4 60krpm 68 lb Preload, L/R = 0.327

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FFid_T7R4_60Krpm T7R4_Ff_meas_60Krpm Figure 32 Predicted and measured Flow Factor vs. pressure for a 3.88” OD Seal with an L/R of 0.327 when operating at 60 krpm with a 68 Lb Preload

Flow Factor vs Pressure T7R4 60 krpm, 80 lb Preload, L/R=0.327

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FFid_T7R4_60Krpm T7R4_Ff_meas_60Krpm Figure 33 Predicted and measured Flow Factor vs. pressure for a 3.88” OD Seal with an L/R of 0.327 when operating at 60 krpm with a 80 Lb Preload


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Flow Factor vs Surface Velocity and L/R Ratio

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Figure 34 Measured Flow Factor vs. surface velocity for L/R Ratios

Flow Factor vs L/R vs Speed

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Static 400 fps 667 fps 1000 fpsLinear (400 fps) Linear (Static) Linear (667 fps) Linear (1000 fps)

Figure 35 Measured Flow Factor vs. L/R ratio and surface velocity

Flow Factor vs Preload and Surface Speed

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Figure 36 Flow Factor vs. Compressive Preload and Surface Velocity

Flow Factor Versus Inlet Pressure

Pressure (psi)

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4.37" Full-T1R3_100 lb Data4.37"Full_T1R4_130 lb Data4.37OD_T1R4_FFid_Theory 4.37OD Half_T2R2_125 lb Data4.37OD_Half_T2R2_Theory3.82OD_Full_T3R1_121 lb Data3.82OD_Full_T3R1_Theory3.82OD_Full_T3R3_100 lb Data3.82OD_Full_T3R3_Theory3.82OD_Half_T4R1_68 lb Data3.82OD_Half_T4R2_80 lb Data9"ODUniT5R5_500 lb Data9"ODUniT5R5_P Theory4.37 OD_T1R3_FFid-Theory3.82_Half_T4R2_Theory

Figure 37 Flow Factor for all static tests with theoretical projections out to 100 psid pressures

IV. Summary and Conclusions Based on the experimental work conducted and the preliminary data analysis, correlation between experiment

and theory was shown to be excellent. Static testing showed that increasing radial path length and axial preload reduces leakage flow. In each of the static test cases, the measured flow factor asymptotically approaches a flow factor of 0.02 with increasing differential pressure. The predicted flow factors, which are based upon the measured differential pressure, seal geometry and differential change in film height, were slightly higher than measured. Thus, the predictions are very reflective of the actual test conditions and show that the proposed closed form Reynolds based flow analysis is valid for use in preliminary design of the compliant foil face seal. While this analysis did not account for the fully open channel between the individual face seal pads, Reynolds numbers greater than 3000, and often approaching 30,000, indicating very turbulent flow was to be expected, hence, part of the reason for the rapidly falling flow factor with increasing pressure.

Dynamic testing was conducted under axial loads of 68 and 80 lbs and speeds of 24,000, 40,000 and 60,000 rpm for two different seal configurations, yielding peripheral surface velocities of 400, 667 and 1000 ft/s respectively. For these dynamic tests, leakage was lower than for static or non-rotating conditions. Further, leakage was lower for larger L/R ratios and was lower for higher compressive preloads. Each of these observations was as expected, confirming the hypothesis that hydrodynamic effects would be at work in the seal design to reduce leakage by providing an additional resistance to flow. Thus, given the trends observed, it is highly probable that large face seals having flow factors approaching and even less than 0.01 will be possible.

Finally, since experience with compliant foil air bearings used in aircraft air cycle machines, where the typical MTBF has been in excess of 130,000 hours, it is expected that life of the compliant foil face seal will be on the order of tens of thousands of hours. The key life issues are the number of start and stop cycles and the maximum operating temperatures. Since there is slight coating wear when speeds are not sufficient to generate the hydrodynamic pressures, which act to separate the foil seal and shaft surfaces, the selected coating needs to sustain the wear at the temperatures encountered during startup and shutdown. Thus for the intended gas turbine engine application the coating needs to have adequate life at both the starting temperature but more importantly at the high temperature shutdown condition. Additionally, the coating system needs to be able to sustain potential high velocity intermittent rubs such as may occur due to the potentially very large total rotor deflections/distortions and loads.

Acknowledgments The authors would like to acknowledge the support of US Air Force Research Laboratory at Wright-Patterson

AFB, OH under contract FA8650-07-M-2783 and Mohawk Innovative Technology, Inc.

References 1Liang, A.D., and Hendricks, R.C., 1993, Seals Flow Code Development-92, NASA CP-10124


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2“1999 NASA Seal/Secondary Air System Workshop", NASA CP-2000-210472, 2000 3“Seals Code Development Workshop", NASA CP10181, Workshop Proceedings, 1995 4Pezzella, G., “Lubricant and Mechanical Component Requirements Assessment for the High Performance Turbine Engine

Initiative,” AFWAL-TR-88-2138, February 1989. 5Hendricks, R., and Steinitz, B. Editors, “Seals/Secondary Flows Workshop 1997, V.1”, AST 028, 1998. 6Flower R., "Brush Seal Development Systems, AIAA Paper 90-2143, 1990. 7Heshmat, H., Walowit, J., and Pinkus, O. "Analysis of Gas-Lubricated Compliant Journal Bearings." Trans. ASME J. Lubr.

Tech. Vol. 105, No. 4, 1983: pp. 647-655. 8Heshmat, H. and Hermel, P., "Compliant Foil Bearing Technology and Their Application to High Speed Turbomachinery."

The 19th Leeds-Lyon Symposium on Thin Film in Tribology - From Micro Meters to Nano Meters, Leeds, U.K., Sep. 1992, D. Dowson et al. (Editors), Elsevier Science Publishers B.V., 1993, pp:559-575.

9Ku, C.-P.R., and Heshmat, H., "Compliant Foil Bearing Structural Stiffness Analysis Part II: Experimental Investigation." Trans. ASME J. of Tribology, Vol. 115, No. 3, 1993: pp.364-369.

10Heshmat, H. "Advancements In The Performance of Aerodynamic Foil Journal Bearings: High Speed and Load Capability." Trans. ASME J. of Tribology, Vol. 116, No. 2, 1994, PP 287-295.

11Salehi, M., Heshmat, H., "Performance of a Complaint Foil Seal in a Small Gas Turbine Engine Simulator Employing a Hybrid Foil/Ball Bearing Support System”, Transaction STLE, Vol. 44, No.3., pp.458-464.

12Salehi, M., Heshmat, H., "High Temperature Performance Evaluation of a Compliant Foil Seal”, AIAA 2000-3376, 36th AIAA/ASME/SAE/ASEE Joint Propulsion Conference & Exhibit, Huntsville, Alabama, 2000.

13Salehi, M., Heshmat, H., Walton, J., and Cruzen S., 1999, "The Application of Foil Seals to a Gas Turbine Engine", AIAA 99-2821, 35th AIAA/ASME/SAE/ASEE Joint Prop. Conference & Exhibit, Los Angeles, CA., 1994

14Salehi, M, and Heshmat, H., "On The Flow And Thermal Analysis Of The Compliant Gas Foil Seals And Foil Bearings", Tran. STLE, Vol. 43, No.2, 2000, pp.318-324.

15Heshmat, H. "Analysis of Compliant Foil Bearings with Spatially Variable Stiffness." AIAA-91-2102, AIAA/SAE/ASME/ASEE 27th Joint Propulsion Conference, Sacramento, CA, 1991.

16Heshmat, C.A. and Heshmat H. "An Analysis of Gas Lubricated, Multi-Leaf Foil Journal Bearings with Backing Springs." ASME 94-Trib-61.

17Heshmat, H. and Ku, C-P R., ”Structural Damping of Self-Acting Compliant Foil Journal Bearings,” Trans. ASME, J. of Trib., Vol. 116, No.1, 1994, pp. 76-82.

18Zou and I. Green. 1997, “Real-Time Condition Monitoring of Mechanical Face Seals.” 24th Leeds-Lyon Symposium on Tribology, Imperial College, London, September 4-6, 1997, 423-430.

19Lee, A. S., and Green, I., "The Coupling of Shaft and Mechanical Face Seal Dynamics Using the Extended Complex Transfer Matrix Method," Proceedings of the Fourth International Symposium on Transport Phenomena and Dynamics of Rotating Machinery, Honolulu, Hawaii (April 5-8, 1992), Vol. A, 305-314.

20Lee, A. S., and Green, I., "Higher Harmonic Oscillation in a Flexibly Mounted Rotor Mechanical Seal Test Rig," Trans. ASME, Journal of Vibration and Acoustics, Vol. 116, No. 2 (April 1994) 161-167.

21Lee, A. S., and Green, I., "Physical Modeling and Data Analysis of the Dynamic Response of a Flexibly Mounted Rotor Mechanical Seal," Trans. ASME, Journal of Tribology, Vol. 177, No. 1 (January 1995) 130-135.

22Lee, A. S., and Green, I., "An Experimental Investigation of the Steady-State Response of a Noncontacting Flexibly Mounted Rotor Mechanical Face Seal," Trans. ASME, Journal of Tribology, Vol. 177, No. 1 (January 1995) 153-159.

23Wileman, J. and Green, I., "The Rotor Dynamic Coefficients of Eccentric Mechanical Face Seals," Trans. ASME, Journal of Tribology, Vol. 118, No. 1 (January 1996) 215-224.

Documents

[American Institute of Aeronautics and Astronautics 44th AIAA/ASME/SAE/ASEE Joint Propulsion Conference & Exhibit - Hartford, CT ()] 44th AIAA/ASME/SAE/ASEE Joint Propulsion Conference