Bruce M. Steinetz

Glenn Research Center, Cleveland, Ohio

Shawn Taylor

University of Toledo, Toledo, Ohio

Jay Oswald

J&J Technical Solutions, Cleveland, Ohio

Jonathan A. DeCastro

QSS Group, Inc., Cleveland, Ohio

Seal Investigations of an

Active Clearance Control

System Concept

NASA/TM—2006-214114

February 2006

120–ISROMAC–11
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Bruce M. Steinetz

Glenn Research Center, Cleveland, Ohio

Shawn Taylor

University of Toledo, Toledo, Ohio

Jay Oswald

J&J Technical Solutions, Cleveland, Ohio

Jonathan A. DeCastro

QSS Group, Inc., Cleveland, Ohio

Seal Investigations of an

Active Clearance Control

System Concept

NASA/TM—2006-214114

February 2006

National Aeronautics and

Space Administration

Glenn Research Center

Prepared for the

11th International Symposium on Transport Phenomena

and Dynamics of Rotating Machinery

sponsored by the Pacific Center of Thermal Fluids Engineering

Honolulu, Hawaii, February 26–March 2, 2006

120–ISROMAC–11
Acknowledgments

The authors would like to acknowledge the support of NASA Glenn’s Ultra-Efficient Engine Technology (UEET) projects:

Intelligent Propulsion Systems and Foundation Technologies (IPSFT) and Propulsion 21 for their financial support. The authors

also acknowledge the contributions of the following people: Scott Lattime, Malcolm Robbie, Art Erker, and Tobie Mintz for their

design and fabrication support, and Richard Tashjian, Kevin McCormick, Tom Lawrence, Mike McGhee, and Joe Wisniewski for

their technical assistance during the test rig installation and troubleshooting. Finally, the authors acknowledge Robert Hendricks

for his thoughtful review of the manuscript.

Available from

NASA Center for Aerospace Information

7121 Standard Drive

Hanover, MD 21076

National Technical Information Service

5285 Port Royal Road

Springfield, VA 22100

Trade names or manufacturers’ names are used in this report for

identification only. This usage does not constitute an official

endorsement, either expressed or implied, by the National

Aeronautics and Space Administration.

Available electronically at http://gltrs.grc.nasa.gov
NASA/TM—2006-214114 1 

 

Seal Investigations of an 

Active Clearance Control System Concept 

Bruce M. Steinetz 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Shawn Taylor 

University of Toledo 

Toledo, Ohio 43606 

Jay Oswald 

J&J Technical Solutions 

Cleveland, Ohio 44130 

Jonathan A. DeCastro 

QSS Group, Inc. 

Cleveland, Ohio 44135 

Abstract 

In an effort to improve upon current thermal active 

clearance control methods, a first generation, fast-acting 

mechanically actuated, active clearance control system has 

been designed and installed into a non-rotating test rig. In order 

to harvest the benefit of tighter blade tip clearances, lowleakage seals are required for the actuated carrier segments of 

the seal shroud to prevent excessive leakage of compressor 

discharge (P3) cooling air. The test rig was designed and 

fabricated to facilitate the evaluation of these types of seals, 

identify seal leakage sources, and test other active clearance 

control system concepts. 

The objective of this paper is to present both experimental 

and analytical investigations into the nature of the face-seal to 

seal-carrier interface. Finite element analyses were used to 

examine face seal contact pressures and edge-loading under 

multiple loading conditions, varied E-seal positions and two 

new face seal heights. The analyses indicated that moving the 

E-seal inward radially and reducing face seal height would lead 

to more uniform contact conditions between the face seal and 

the carriers. Lab testing confirmed that moving the balance 

diameter inward radially caused a decrease in overall system 

leakage. 

Introduction 

Blade tip sealing has remained a challenging problem since 

the development of the gas turbine engine. Environmental 

conditions at the tip seal location include gas temperatures up 

to 2500 °F, pressures up to 500 psi, high surface speeds 

(1900 fps), as well as unburned jet fuel and contaminants (dirt, 

sand, etc.) that make for very challenging surroundings. 

Clearances change during engine operation as a result of both 

mechanical (pressure and rotational) and thermal loads. 

Discrete structural mounts coupled with thermal and flight 

loads result in non-uniform distortions and asymmetric 

clearances. Designers of clearance control schemes must 

properly account for all of these effects to implement a 

successful active clearance control (ACC) system. 

Improved blade-tip sealing in both the high pressure 

compressor (HPC) and high pressure turbine (HPT) can provide 

dramatic reductions in specific fuel consumption (SFC), lower 

emissions, increased compressor stall margin, and engine 

efficiency, as well as increased payload and mission range 

capabilities. Current clearance control systems employ 

scheduled cooling of the outer case flanges to reduce high 

pressure turbine tip clearances during cruise conditions. These 

systems have relatively slow response and do not use clearance 

measurement thereby limiting possible efficiency gains. 

This paper describes work done to-date to evaluate a fastacting, mechanically-actuated, active-clearance control concept 

with clearance measurement to help overcome some of the 

noted limitations. 

Nomenclature 

Flow Area flow area (in2) where flow is choked

Circumference π×Diameter (in.) 

m& measured flow rate (lbm/sec)

R gas constant for air (53.3 lbf-ft / lbm-°R)

T temperature (°R) 

gc gravitational constant (32.2 lbm-ft/lbf-sec

2

) 

Psupply supply pressure (psia) 

h face seal modified height 

δ radial movement of E-seal for figure 9b 

analysis 

α radial movement of E-seal for figure 9c 

analysis 
NASA/TM—2006-214114 2 

Background 

Figure 1 shows the high pressure HPT blade tip seal 

location in a modern gas turbine engine [1]. The figure shows a 

cross section of the combustor and two-stage HPT. The turbine 

disk, blade, and tip seal of the first-stage turbine are labeled. 

Blade tip or outer air seals line the inside of the stationary case 

forming a shroud around the rotating blades, limiting the gas 

that spills over the tips. Blade tip clearance varies both with 

engine operating points (e.g., ground idle, takeoff, cruise, decel, 

etc.) and cycle life. These clearance variations are due to a 

number of loads on both static and rotating parts and wear of 

these parts. 

Load mechanisms can be separated into two categories, 

namely engine (power-induced) loads and flight loads. Engine 

loads include centrifugal, thermal, internal engine pressure, and 

thrust loads. Flight loads include inertial (gravitational), 

aerodynamic (external pressure), and gyroscopic loads. Engine 

loads can produce both axisymmetric and asymmetric clearance 

changes. Flight loads generally produce asymmetric clearance 

changes. Load mechanisms generally act to temporarily alter 

blade tip clearance. Wear mechanisms permanently change tip 

clearance. The ACC system must be designed to handle even 

the worst case transient conditions such as a stop-cock event 

where the engine is shutdown in-flight, allowed to windmill for 

a while, then restarted to full power. 

The backside of the HPT shroud (blade outer-air-seal) is 

generally cooled with compressor discharge air (T3 air: 1200 to 

1300 °F). This cooling is necessary for the shroud segments to 

survive the rotor inlet gas temperatures in excess of 2500 °F. 

The cooling air is also used to purge the leading and trailing 

edges of the shroud segments, providing a positive back-flow 

margin. This cooling flow is shown in figure 2 for the first 

stage shroud of a two-stage HPT. The pressures surrounding 

the shroud segment can be expressed as a function of the 

compressor discharge pressure (CDP; P3). Flow path pressure 

adjacent to the HPT shroud varies axially due to the work 

extracted by the turbine blades. For large commercial engines, 

the pressure of the cooling air behind the shroud is about 60 to 

80 percent of P3 pressure [2]. Pressure in the tip clearance 

region varies axially from the leading to the trailing edge of the 

shroud about 70 to 30 percent of the P3, respectively. To 

maintain a positive backflow margin from the rotor inlet air, the 

cooling pressure on the backside of the shroud must always be 

higher than the rotor inlet side. The radial pressure difference 

across the shroud creates a load inward toward the shaft 

centerline. A resultant moment also exists on the shroud that is 

created by the non-uniform axial pressure distribution. P3 

Figure 1.—HPT blade tip seal location in a modern gas turbine engine [1]. 
NASA/TM—2006-214114 3 

Figure 2.—Axial pressure distribution across HPT blade tip 

seal cross section [2]. 

pressure is highest during maximum thrust events such as 

takeoff and re-accel. For large commercial engines this 

translates to a maximum cooling air pressure differential of up 

to 150-psid across the shroud. 

Lattime and Steinetz [2] provided a comprehensive review 

of the mechanisms of tip clearance variation, their effects in gas 

turbine engines, methods of controlling tip clearance and the 

benefits associated with reducing tip clearance. Kawecki [3]

presented trade studies of a variety of approaches for active 

clearance control identifying a variety of fast-acting mechanical 

and novel thermal control systems. Melcher and Kypuros [4] 

outlined NASA Glenn’s general approach for developing a fastresponse ACC system and DeCastro and Melcher [5] examined 

the control systems requirements for fast acting ACC systems. 

Lattime et al. [6] designed a first generation, fast acting 

mechanical active clearance control system integrated into a 

test rig to evaluate its performance. Steinetz et al. [7] assessed 

leakage, positional accuracy, repeatability, and control of the 

fast-acting, mechanical active clearance control system. The 

current work builds on that initial investigation. 

Benefits of Active Clearance Control 

Blade tip clearance directly influences gas turbine 

performance, efficiency, and life. Reducing air leakage over the 

blade tips increases turbine efficiency and permits the engine to 

meet performance and thrust goals with less fuel burn and 

lower rotor inlet temperatures. Running the turbine at lower 

temperatures increases the cycle life of hot section components, 

which in turn, increases engine service life by increasing the 

time between overhauls. 

Lattime and Steinetz [6], GE [8], and Wiseman and Guo 

[9] provide overviews of the many benefits of advanced active 

clearance control systems. Some of the more noteworthy 

benefits of implementing fast mechanical ACC systems in the 

HPT of a modern high bypass engine are provided herein for 

completeness. In terms of fuel savings, a tip clearance reduction 

of 0.010-in. results in ~0.8 to 1 percent decrease in specific fuel 

consumption. By reducing fuel burn significant reductions in 

NOx, CO, and CO2 emissions are also possible. Reducing tip 

clearances by 0.010-in. decreases exhaust gas temperature 

(EGT) ~10 °C. Deterioration of EGT margin is the primary

reason for aircraft engine removal from service. Running the 

engine at lower operating temperatures can result in increased 

life of hot section components and extend engine time-on-wing 

(up to 1000 cycles). Additional benefits include increased 

payload and mission range capabilities. 

Study Objectives 

This work is part of a larger research effort to develop 

approaches for clearance control systems for use in the HPT 

section of large commercial aircraft engines to improve upon 

the case-cooling methods employed today. The objective of the 

current work is to present both experimental and analytical 

investigations into the nature of the face-seal to seal-carrier 

interface evaluated in the active clearance control test rig. 

Specific objectives of this study include: 

• Experimentally examine whether the face-seal to sealcarrier interface can be improved to lower parasitic 

losses. 

• Analytically determine the nature of the contact 

pressures between the face seal and the seal carriers for 

a variety of pressure differentials, E-seal positions, and 

seal carrier motion conditions. 

• Analytically determine other candidate face seal 

designs to improve upon the baseline seal design. 

Test Apparatus and Procedures 

Apparatus 

An active clearance control system test rig has been 

developed and installed (fig. 3). The key features of the test rig 

are summarized below. For a comprehensive review of the test 

rig design approach, the reader is directed to Lattime and 

Steinetz [6]. 

Figure 3.—ACC test rig with housing lid and chamber cover 

plate removed for clarity. 
NASA/TM—2006-214114 4 

TABLE 1.—COMPARISON OF ACC RIG DESIGN AND A 

TYPICAL MODERN HIGH BYPASS RATIO ENGINE 

OPERATING CHARACTERISTICS 

Parameter ACC Rig 

Design 

Reference 

Engine 

Shroud backside pressure (psia) 135 500 

Pressure differential (psid) 120 150 

Shroud Backside temperature 

 Current (°F) 1000 1250 

 Future (°F) 1250 to 1300 1250 to 1300

Diameter (in.) 20 30 

Shroud face width (in.) 2 2 

Number of shrouds/seal carriers 9 16 

Pressure induced load on 

actuator (at pressure 

differential) (lbf) 

1650 1750 

Clearance change (e.g., stroke) 

(in.) 

0.190 0.05 nom. 

Clearance change rate (in./sec) 0.01 0.01 

Clearance Measurement 

Technique 

 

 Current capacitance not used 

 Future (under development) microwave capacitance/ 

microwave 

General Overview 

The ACC test rig simulates the environment surrounding 

the backsides of the turbine shroud segments. The purpose of 

the test rig is to evaluate actuation systems in a “static” 

environment without blade rotation. The rig design 

concentrated on simulating the temperature and pressure 

conditions that exist on the backsides of the seal segments, 

without the need for a rotating turbine, which greatly simplified 

the rig design. Rig specifications were chosen to closely 

simulate engine conditions. Table 1 compares the main 

characteristics of both the ACC test rig and a typical modern 

high bypass ratio engine. 

The general features of the test rig are highlighted in 

figures 3 to 5. The rig main housing consists of two concentric 

cylinders, which form an annular cavity. A two-piece annular 

radiant heater (upper and lower halves) surrounds the seal test 

chamber to simulate the HPT tip seal backside temperature (T3) 

and pressure (P3) environment. At the heart of the rig is a 

segmented shroud structure (seal carrier) that would structurally 

support the tip seal shroud segments in an engine. Radial 

movement of the seal carriers controls the effective 

position/diameter of the seal shroud segments, thereby 

controlling blade tip clearance. The carrier segments are 

connected to one another through a pinned and slotted link as 

shown in figure 4. This link (or “foot”) positions the carrier 

segments radially while allowing relative circumferential 

Figure 4.—ACC cut-away showing detail for one of nine actuator rods and attachment 

foot, actuator mount, seal carriers, proximity probe clearance sensor, inlet air supply 

pipe, air flow directions, and radiant heater [6]. 
NASA/TM—2006-214114 5 

Figure 5.—Chamber detail showing seal carrier and 

adjacent face seals with E-seal preloader, actuator rod 

piston ring seals (2 places), chamber cover C-rings (4 

places), the materials used in construction, and chamber 

high and low pressures. Red arrow indicates location of 

trial seal, as discussed in the Procedure Section. 

Figure 6.—ACC test rig with carrier alignment fixture 

installed. Note gage pins insert into precision centering 

holes in carrier pins. Alignment fixture used to set 

nominal or “home” position for all carrier segments. 

movement or dilation of the seal carrier segments. A 

photograph of the test rig is shown in figure 6. Here the 

precision alignment fixture used to establish seal carrier 

concentricity relative to the rig centerline is installed. Tapered 

gage pins mounted in the alignment fixture engage precision 

alignment holes machined in the ends of the carrier pins. This 

alignment fixture or “gage plate” was used to define the 

nominal or “home” position for subsequent tests. 

Chamber and Seal Detail 

 The pressurized test chamber encloses the carrier segments 

inside the annular heater through which pressurized air (heated 

for future tests) is supplied to simulate the P3 cooling/purge air 

pressure on the seal carrier backsides. The pressurized air is 

sealed along the sides of the seal carrier segments by contacting 

face seals that are energized via metal “E-seals” imbedded in 

the upper and lower chamber plates (figs. 4 and 5). The space 

between adjoining carrier segments is sealed with thin flexure 

seals otherwise known as spline seals (figs. 4 and 5). The face 

seal width was selected to seal the sliding interface between the 

edges of the flexure seals as they move radially inward/outward 

during actuation at moderately slow speeds (e.g., 0.01 in./sec). 

To minimize leakage, the face seal should always be in sealing 

contact with the edges of the flexure seals. The chamber 

penetrations for the nine actuator rods are sealed using two sets 

of concentric piston ring seals. A series of radial tubes 

projecting outward from the chamber’s inner and outer side 

walls serve as supports, air supply and exhaust ports, probe 

mounts, and the actuator rod guides. These are sealed in their 

respective bosses via a shrink fit qualified through hydrotesting. The chamber functions to support and align the carrier 

segments and actuator rods, as well as to house instrumentation 

and to seal the pressurized air from the radiant heater which is 

not designed to carry any pressure loading. The chamber cover 

plates are sealed using inner- and outer-flange C-rings. Only 

the outer C-rings carry pressure during operation. The pipes 

supplying high pressure to the chamber (fig. 4) are also sealed 

using two sets of concentric piston ring seals. 

Component Seal Leakage Evaluation.—The relative 

contribution of each seal component to the total system leakage 

was determined by Steinetz et al. [7]. The flexure seals 

contribute the majority of the leakage at 85 percent of the total. 

The face seals contribute the next highest percentage at 

7 percent of total flow. The piston rings for the nine actuator 

rods contribute 6 percent. The rig supporting seals (e.g., air 

inlet piston rings and the four C-ring seals) contribute 

approximately 2 percent of the total. 

Actuators and Position Measurement 

Electric stepper motors were used to position the seal 

carriers for the current set of tests. Each stepper motor is 

equipped with a quadrature encoder to measure both shaft 

rotational position and direction of motion. A complete list of 

specifications is presented in Table 2. The stepper motors are 

rated for 500 lbf of actuation force. This limited the maximum 

chamber pressure that could be tested. Although calculations 

indicated that chamber pressures of 35 psig would result in 

500 lbf actuator loads, actual tests indicated that the practical 

upper pressure was in the 20 to 30 psig range. It is expected 

that this noted difference is due to component friction. Stepper 

motors were selected as low-cost first generation actuators 

because of their high reliability and conduciveness to accurate 

position control. These actuators are therefore useful in 

evaluating the kinematics, seals, sensors, and control 

electronics under moderate loading conditions. Researchers are 

procuring a series of servo-hydraulic actuators designed to be 

able to handle the full 120 psig test pressure. These servohydraulic actuators will come equipped with linear
NASA/TM—2006-214114 6 

TABLE 2.—STEPPER MOTOR SPECIFICATIONS 

Maximum load capability 

(lbf) 

500 

Stroke (in.) 0.5 

Displacement resolution: 

without micro-stepping (in) 

0.0005 

Displacement resolution: 

with micro-stepping (in) 

10 micro steps, 0.00005 

Encoder resolution (in.) 0.000125 

Voltage (V) 24 

Current draw-peak (amps) 1.4 max 

Weight (lbs) 5 

Manufacturer Haydon Switch and 

Instrument, Inc. 

Model no. 87H43–12–002 

TABLE 3.—INSTRUMENTATION SPECIFICATIONS 

Thermocouples Type K

 Manufacturer Omega 

 Accuracy (°F) ± 4 

 Range (°F) –328 to 2282 

Pressure Transducers 

 Manufacturer Druck 

 Model no. PMP4010 

 Accuracy (psi) 0.12 

 Range (psi) 0 to 300 

Flow meter 

 Manufacturer Teledyne-Hastings 

 Model no. HFM-306 

 Accuracy (lbm/s) 0.0022 (1.2% full scale) 

 Range (lbm/s; SLPM) 0.19; 4000 

TABLE 4.—CAPACITANCE CLEARANCE 

PROBE SPECIFICATIONS 

Calibrated operating maximum 

temperature 

1500 °F 

Measurement range (in.) 0 to 0.125 

Accuracy (in.) 0.0002 

Resolution (in.) 0.00005 

Excitation voltage (V) 15 

Probe diameter (in.) 0.375 

Stand-off distance for current tests (in.) 0.025 

Weight (lbs) 0.04 

Manufacturer Capacitec 

Model no. HPC 150 

variable differential transducers (LVDTs) to accurately 

measure actuator piston location. 

Instrumentation 

 A variety of thermocouples, pressure transducers and flow 

meters were used to collect the necessary data. Table 3 provides 

a list of the transducer specifications and their respective 

accuracies. High temperature capacitance clearance probes 

were used to measure the radial displacement of the seal 

carriers (fig. 4) at three different circumferential locations for 

the mission simulated clearance tests performed in 

Steinetz et al. [7] and discussed in the Control Implementation 

section below. Table 4 provides a list of the probe 

specifications. Two capacitance clearance probes were mounted 

diametrally opposite one another and the third was clocked at 

90° relative to the first two. 

Data Acquisition System 

A National Instruments (NI) LabVIEW-based program 

running on a PC host computer was used to implement data 

acquisition and stepper motor control. Temperature, pressure, 

and flow are measured at a rate of two samples per second, and 

position data (motor encoders and clearance probes) are 

sampled at 20 samples per second. The data is digitized with a 

12 bit analog/digital (A/D) card. The stepper motors are 

controlled with two 8 axes motion control cards located in a 

remote chassis. Communication delays between the PC and the 

chassis limits the control loop time to approximately 50 ms. 

Control Implementation 

The active clearance control system was implemented in 

NI LabVIEW with an NI motion control blockset that allows 

the nine stepper motors to be controlled simultaneously via 

motion control cards. The advantage of using motion control is 

that stepper motor “positional moves” commanded by 

LabVIEW are generated by the motion control cards and hence 

operate independently of the software (limited to 50 ms 

updates). When LabVIEW commands a “positional move,” a 

desired position and velocity is fed to the motion control board 

and the move begins. 

The actuators were controlled in either open-loop or 

closed-loop modes. For the leakage vs. carrier position tests 

performed herein, the actuators were controlled in open-loop 

mode. In open-loop mode, a single position move was 

commanded to all nine actuators. Motion stops after the 

commanded number of motor steps have been executed. 

As described in Steinetz et al. [7], closed loop control was 

used for evaluating the ability of the ACC system to maintain a 

tight clearance set-point during simulated engine transients. An 

on-off set point tracking controller was implemented in 

LabVIEW to track simulated clearance transients with the 

stepper motor actuators. An error signal was computed every 

50 ms by subtracting the actual clearance (as measured by the 

proximity probes) from the desired clearance set-point. A 

position move equal to the magnitude and direction of the error 

signal was commanded to the stepper motors. If the move 

completed before 50 ms elapses, then the motor would stop 

until the next cycle began. If the move did not complete, the 

motor would remain in motion, but, with updated commands. 

In order to minimize tracking delays due to error propagation, 

the motor velocity was set to a value of 0.010 in./sec. 

Simulated Engine Clearance versus Time Study.— 

Steinetz et al. [7] examined the closed-loop positional accuracy 

for the ACC system while simulating an engine clearance 

versus time study. Figure 7 shows the results of those tests here 

for completeness. The set-point position defines ideal carrier 

position movement versus time for a simulated take-off 
NASA/TM—2006-214114 7 

Figure 7.—ACC test rig simulation of engine clearance 

transient during take-off conditions. (a) Commanded 

(e.g., set point) and measured carrier position (using 

capacitance probes). Negative position indicates 

movement radially outward from start position or 

simulated “flight idle” condition; (b) Positional error 

versus time. Conditions: pressure 20 psig, ambient 

temperature, control based on minimum clearance of 

3 capacitance probes [Steinetz et al. [7]]. 

condition (clearance change of 0.045 in.). These tests were 

performed at a pressure of 20 psig. Figure 7(a) shows the 

set-point and the measured carrier position at three clock 

positions: probe 1(0°), probe 2 (90°), probe 3 (180°). 

Throughout these tests the controller examined all three probes 

and controlled to the minimum clearance which in this case was 

at probe 1. Though it may difficult to see, the motion of the 

carrier adjacent probe 1 tracked the set-point very well. The 

carriers adjacent the other two probes were slightly further 

away from the probes from assembly set-up and those carriers 

remained further outward radially throughout the test. 

Figure 7(b) shows the error defined as “set-point minus 

actual position” on a highly refined vertical scale. (Note that for 

this chart, negative error indicates the actuator did not move 

outward radially as far as the set-point). For the entire 

simulated take-off transient, the error between the set-point 

(green line) and probe data (black line) was less than 0.001-in. 

This was a very encouraging result as it showed that the 

kinematic system could track simulated engine clearance 

changes at the correct rate and range under load with an 

acceptable error. Note: Tests were also performed at chamber 

pressures of 10 and 30 psig. The results were generally the 

same except that the error tracked with pressure. This is the 

expected result as the higher pressure load (e.g., 30 psig) and 

corresponding frictional effects were just above what the 

stepper motor could reliably actuate without slowing down. 

Future hydraulic actuators are designed to eliminate this 

shortcoming. 

Procedures: Experimental 

Face Seal Contact 

During the testing reported in Steinetz et al. [7], it was 

noticed that minor scuffing indications were present on the seal 

carriers near the outer diameter of the face seal. Contact at that 

location could indicate a flow path that was not fully 

considered in that study. If the contact between the face seal 

and the seal carriers was away from the edges of the flexure 

seals (e.g., face seal contact not coincident with the edges of the 

flexure seals), then leakage could travel inward radially past the 

face seals and exit over the top of the flexure seals through a 

wedge-shaped gap formed between the face seal and the seal 

carriers. Tests were performed herein to examine the nature of 

the face seal contact using pressure sensitive film. This pressure 

sensitive film is placed between two mating surfaces and 

exhibits a color change with color density proportional to the 

unit contact pressure. The pressure sensitive film used in the 

current study was obtained from Sensor Products Inc., East 

Hannover, NJ. Two contact pressure ranges were used in this 

study including the micro range (2 to 20 psi, black in color) and 

ultra low (range 28 to 85 psi, pink in color). 

Face seal in contact with granite base.—A test was 

conducted to determine the uniformity of the contact pressure 

of the face seal in contact with a precision flat granite base. In 

this test, the face seal was loaded against the granite base with 

the ultra-low range pressure film positioned in 90° increments. 

The film width was approximately 1 in. The seal was then 

loaded with a bar placed across the center using gauge pins so 

as to load the center of the face seal. The bar was then loaded 

with approximately 200 lbf. For reference purposes, the face 

seal flatness specification was 0.0003 in. and roughness was 

8 μin. 

Face seal installed against seal carriers in test rig.—

Pressure sensitive film was installed between the seal carriers 

and the face seal at six locations around the circumference of 

the face seal. Both the ultra low range film (~0.005 in. thick), 

and the micro range carbon paper based film (~0.012 in. thick) 

were used. The film types were alternated on a 60° angular 

pattern, leaving 120° between similar film types. The rig was 

built up with the E-seals in place behind the face seal. Air 

pressure was not applied. The test goal was to determine the 

face seal contact loading pattern when assembled. The bolts 

were tightened to 70 in.-lb, as typical for a room temperature 

build-up. For reference purposes, the seal carriers were made to 

the following specifications: flat to within 0.0003 in., opposite 

carrier faces parallel to 0.0005 in., and surface roughness of 

10 μin. 

 Effect of Carrier Position on Leakage 

Low ACC parasitic leakage is required in order to 

minimize P3 cooling air loss and to net the engine efficiency 
NASA/TM—2006-214114 8 

gains from the tighter tip clearance. To quantify the ACC 

system leakage and possible improvements, two different static 

flow test setups were used. 

Baseline Configuration.—In the baseline configuration, 

E-seals were assembled in their normal locations behind the 

face seals. Flow measurements were made with the following 

test conditions. Air or “engine” pressure was ramped from 

10 psig up to 109 psig. For each of these pressures, the seal 

carriers were positioned at a variety of carrier positions ranging 

from 0.02 in. radially inward through home (e.g., 0.00 in.), to 

0.08 in. radially outward, as indicated by the capacitance 

probes. During carrier movement, the air pressure was 

removed. 

Trial Seal Configuration.—To evaluate the theory that 

increasing the engine pressure behind the face seal (e.g., 

changing the balance diameter) could better seat the face seal 

against the seal carriers and flexure seal, the upper E-seal was 

removed from the rig, and a circle of monofilament line 

(0.015 in. diam.) was installed at the mid-point of the land 

between the cover plate and the E-seal as indicated with the red 

arrow in figure 5 (radial dimension from face-seal centerline of 

10.4 in.). The monofilament would act as both a seal and as a 

preloader. Vacuum grease was used to hold the monofilament 

line in place and a lap joint was used at the ends of the line to 

form a continuous loop around the back side of the face seal. 

Flow measurements were made using the same test conditions 

of air pressure and carrier position as described above. 

Procedures: Analytical 

Finite Element Model 

To further understand the source of edge loading observed 

between the outer diameter of the face-seal and the sealcarriers, a series of finite element analyses were performed 

using ANSYS V9.0. In these analyses, the cover plate, the face 

seal, and the inner/outer chamber walls were modeled using 

axisymmetric elements. The elements used were 2-D, six-node, 

triangular structural solid elements (ANSYS: PLANE2). 

The contact surface of the face seal was modeled with 2-D,

three-node, surface-to-surface contact elements (ANSYS: 

CONTA172). The carrier segments were modeled as rigid 

contact targets (ANSYS: TARGE169). Symmetry was invoked 

and only the upper face-seal to seal-carrier surface was 

analyzed. The cover plate was modeled to determine whether 

an interaction between the cover plate and the face seal could 

be affecting the face seal contact. Contact was not observed 

between the cover plate and the face seal during any of the 

finite element analyses. The seal carriers were assumed to be 

rigid. This assumption was followed as the seal carriers have air 

pressure on both their outer and inner surfaces minimizing the 

potential for bending of the seal carrier walls adjacent to the 

face seal. The face seal was modeled to evaluate what role the 

E-seal location (e.g., balance diameter), friction forces, and 

direction of motion might be playing in the observed face seal 

edge loading. The finite element analysis was also used to 

investigate alternate face seal designs. 

Dimensions, Loads, and Boundary Conditions.—The 

relevant face seal dimensions used for the analyses are shown 

in figure 8. A summary of the different analysis cases examined 

are shown in figure 9. Face seal geometries modeled included: 

(a) baseline or as-built; (b) E-seal moved inward radially either 

Figure 8.—Relevant face seal dimensions (in.) for 

analysis. E-seal baseline radius is 10.513 in. Note all 

dimensions emanating from left are radii. Phantom 

lines denote cover plate groove for reference. 

Figure 9.—Face seal geometries and air pressure loads (P) modeled. (a) Baseline; (b) E-seal 

moved inward radially (δ) either 0.1 or 0.2 in. relative to baseline, and (c) overall face seal 

height (h) reduced to either 0.25 and 0.1 in., with associated E-seal movement (α). 
NASA/TM—2006-214114 9 

Figure 10.—Summary of load cases modeled for 

each geometry. Cases examined: Carrier segment 

motion: no motion, inward and outward motion; 

pressure loads (P), 30, 60, 120 psig; E-seal preload 

(F) 12 lbf/ circumferential inch; coefficient of 

friction = 0.3, contact load determined analytically 

using contact elements. 

0.1 or 0.2 in. relative to baseline, (c) overall face seal height 

reduced to either 0.25 or 0.1 in., with associated E-seal 

movement (α) relative to baseline. Figure 10 summarizes the 

load boundary conditions and load cases modeled. Three carrier 

segment motions were considered: no carrier motion, inward 

motion, and outward motion. Applied air pressures considered 

included 30, 60, and 120 psig. A linear pressure distribution 

was used across the face of the face seal (0 to maximum air 

pressure). Air pressure was applied vertically and horizontally 

on other wetted surfaces up to the point of the E-seal to face 

seal contact. The E-seal preload (F) was held constant at 

12 lbf/circumferential inch. The coefficient of friction between 

the face seal and the carrier was assumed to be 0.3. Coulomb 

friction was used therefore friction forces always resisted 

motion. 

Grid refinement study.—A grid refinement study was 

performed to assess the effect of grid size on contact pressure 

solutions. In this study the element size was reduced in size by 

more than half in the interior of the face seal and reduced by a 

factor of 4 along the frictional interface. 

Results and Discussion 

Experimental Results 

Face Seal Contact 

 Figures 11 and 12 show the results of the tests performed 

using pressure sensitive film to investigate the face seal contact 

pressure profile. Tests performed with the face seal against the 

precision flat granite base (fig. 11), showed uniform pressure 

across the seal face (inner to outer diameter), suggesting that 

the face seal was not deformed. Tests performed with the 

Figure 11.—Close-up of one of the pressure sensitive 

films used to characterize face- seal to granite-base 

contact pattern showing near uniform contact 

pressure across the face. 

Figure 12.—Photo of test rig with chamber cover removed 

showing pressure sensitive film documenting face-seal to 

seal-carrier contact pressure for two neighboring carrier 

segments. Observations: the left hand seal carrier is less in 

contact with the face seal than the right hand carrier. Both 

carriers show higher contact pressures near the face seal outer 

diameter and lower contact pressures near face seal inner 

diameter. 

pressure sensitive film between the face seal and the seal 

carriers showed two interesting findings, as shown in fig. 12. In 

this figure the film is laid across two adjacent seal carriers. The 

left carrier did not develop the film as much as the right hand 

carrier when preloaded using the E-seal. However in both cases 

the face seal exhibited edge-loading near the outer diameter of 

the face seal where it contacts the seal carriers. The line of load 

concentration fell on the centerline of the scuff marks 

Outer 

Diameter 

Inner 

Diameter 

Edge 

loading of 

face seal 

outer 

diameter 

against 

carrier 

Pressure 

sensitive 

tape 

Seal 

carrier 

scuff 

indication
NASA/TM—2006-214114 10 

on the seal carriers, which was expected. This proved that the 

scuffing was likely an artifact of the uneven load pattern 

between the face seal and the seal carriers. A similar test was 

performed with the pressure-sensitive film as above but in 

addition to the preload from the E-seal, air pressure was 

applied. This test did not provide any conclusive results. 

Effect of Seal Carrier Position on ACC Leakage 

Figure 13 shows the result of the static leakage versus seal 

carrier position tests at a chamber pressure of 109 psig for the 

two build configurations: E-seal baseline and monofilament 

trial seal installed in place of the E-seal in the top cover plate 

location. Flow data is shown only for the high pressure 

differential where the greatest leakage vs. carrier position trend 

was observed. 

E-Seal Baseline.—Relative to the home position, the 

system leakage was 0.052lbm/s lower when the carriers were 

moved 0.080 in. radially outward. This corresponds to a 

leakage reduction of about 50 percent, indicating that a 

considerable amount of flow is leaking through the face-seal to 

seal-carrier interface. It is believed that this leakage is greatest 

at the three-way intersection between the face seal, seal carriers 

and flexure seals. Leakage reductions were expected because as 

the seal carriers were moved outward, the face-seal to sealcarrier contact zone (e.g., scuffing area) better aligned with the 

diameter of the flexure seals. 

Monofilament Trial Seal.—Relative to the home position, 

the system leakage was 0.027 lbm/s lower when the carriers 

were moved 0.080 in. radially outward. This corresponds to a 

leakage reduction of about 29 percent. 

Comparison of Configurations.—Comparing leakage 

results, a couple of observations are made. First the overall 

leakage of the monofilament seal build is about 13 percent 

(0.014 lbm/s) lower than the E-seal build’s leakage at the 

“home” or nominal position (e.g., carrier position = 0.00 in.). 

Furthermore the E-seal build leakage vs. carrier position curve 

is steeper than that of the monofilament in the 0.08 in. outwardto-home range. Comparing slopes, the E-seal build leakage vs. 

 

Figure 13.—Leakage versus actual seal carrier position for 

109 psig pressure differential and two face seal backside 

configurations: Metal E-seal and monofilament trial seal. 

Figure 14.—Photo of two adjacent seal carriers and flexure 

seal showing face seal contact zone (minor scuffing 

marks) shifted near the inner diameter of the seal carriers. 

E-seal replaced with trial seal: monofilament behind the 

face seal. 

carrier position slope is almost twice (1.85:1) that of the 

monofilament build. This is an interesting finding in that one 

recalls that only the upper E-seal was replaced with the 

monofilament trial seal. One might speculate that the slope of 

the monofilament trial seal build might be flatter should both 

upper and lower E-seals be replaced with monofilament trial 

seals. From an optimal design view point, a relatively flat 

leakage versus carrier position is desired. This condition would 

indicate seal leakage is insensitive to seal carrier position. In 

summary, the trial seal positioned such that engine pressure 

was exerted over a greater face-seal back-side surface area 

reduced the seal leakage and reduced the sensitivity of leakage 

to seal carrier location. 

Figure 14 is a photograph of the hardware taken after the 

monofilament trial seal was tested in place of the E-seal. The 

photo shows two adjacent seal carriers and flexure seal showing 

the face seal contact zone (minor scuffing marks) shifted near 

the inner diameter of the seal carriers. It appears that changing 

the balance on the face seal is an effective way to improve 

sealing performance. 

Comparison of Effective Seal Flow Area to Industry 

Reference Level 

As stated earlier, low ACC parasitic losses are required to 

harvest the engine efficiency gains from the tighter tip 

clearance control. In the current section, a comparison of the 

measured losses in terms of effective leakage flow area per inch 

of circumference is made to an engine industry reference level. 

Using the total measured leakage rates of all seal locations, 

an effective leakage flow area for the entire active clearance 

control system was back-calculated to compare to an engine 

industry reference level. The method used to back-calculate the 

effective leakage flow area was taken from an isentropic 

flow condition with compressibility at the choked-flow 

condition [10]. The leakage flow was considered choked since 
NASA/TM—2006-214114 11 

the pressure ratio of 8.4 (i.e., Psupply/Pexhaust = 

123.7 psia/14.7 psia) was above the critical pressure ratio for 

air. The equation used was: 

g P Circumference

m R T

Circumference

Flow Area

× c × supply ×

× × = 0.6847

&

where: 

Flow Area flow area (in2

) where flow is choked

Circumference π×Diameter (in.) 

m& measured flow rate (lbm/sec)

R gas constant for air (53.3 lbf-ft / lbm-°R)

T temperature (°R) 

gc gravitational constant (32.2 lbm-ft/lbfsec

2

) 

Psupply supply pressure (psia) 

Industry Reference Effective Flow Area.—If one were to 

idealize the ACC system as an elastic structure (e.g., rubber 

band) that could move radially inward/outward, seals would 

only be required between the sides of the seal carriers and the 

static structure. Engine designers have acknowledged that flows 

in these areas less than ~0.1 percent core-flow (W25) would be 

an acceptable loss considering the potential for the significant 

gains possible through the tighter HPT tip clearances. 

Converting this level into an effective flow area per unit 

circumference results in approximately 0.00048 in2

/in. of 

circumference. Since there are two seal locations the total 

effective flow area per unit of circumference would be twice 

the above or 0.00096 in2

/in. Figure 15 compares the effective 

flow area per unit circumference found experimentally for the 

two build conditions (E-seal and monofilament trial seal at the 

carrier “home” position) to this industry reference level. 

Though this current data is only for ambient conditions, the 

effective flow area for the ACC rig falls within the industry 

reference level. 

0.0000

0.0001

0.0002

0.0003

0.0004

0.0005

0.0006

0.0007

0.0008

0.0009

0.0010

Engine Reference

(Forward and Aft

Seal Locations

Combined)

ACC Test Rig:Eseal InstalledACC Test Rig: Trial

seal installed

Flow Area per inch circumference

 (in.^2/in. circumference)

Figure 15.—ACC test rig effective flow area per inch 

circumference (back-calculated from leakage flow 

data) compared to engine reference flow area per inch 

circumference (for forward and aft seal locations). Seal 

carriers at “home” position for two builds: E-seal and 

monofilament trial seal.

Analytical Results 

Baseline Face Seal Design 

Figure 16 shows the results of the finite element analyses 

performed for the baseline face seal design with the E-seal 

positioned at the as-built location. Calculated contact pressures 

(in psi) are shown for three conditions: (a) no carrier motion, 

(b) carrier moving inward, (c) carrier moving outward. These 

studies were performed to assess whether either pressure 

loading or friction were contributing to the edge loading 

between the outer diameter of the face seal and the seal carriers. 

The results do indicate that pressure does play a role in edgeloading the face seal, as indicated for the 120 psig air pressure 

for all cases, and for pressures > 30 psig air pressure when the 

carrier is moving inward. These edge loading results were 

corroborated by the experimental observations presented above. 

In the case where the carrier is moving inward, the contact 

friction tends to rotate the seal and causes edge loading, even at 

the lowest pressures (e.g., 30 psig). 

Color plot scaling.—The reader is alerted that though all 

contact pressure results are plotted on the same scale, ANSYS 

scales the image to fill the ordinate space. Hence, note that on 

select plots the contact pressure is low (e.g., <28 psi) but it is 

still scaled to fill the ordinate space. 

E-Seal Position Study 

Figure 17 shows the calculated results of the face seal 

contact pressures for the baseline face seal configuration (e.g., 

original height) but with the E-seal analytically moved to two 

positions: 0.1 and 0.2 in. radially inward from baseline, for 

three conditions all at 120 psig engine pressure: (a) no carrier 

motion, (b) carrier moving inward, and (c) carrier moving 

outward. Of the two cases examined, positioning the face seal 

contact in 0.1 in. resulted in a fairly uniform contact stress for 

each of the three motion cases considered. Positioning the 

E-seal inward 0.2 in. caused relatively high edge loading 

(contact pressures up to 200 psi) on the inner diameter for the 

no-motion and carrier moving outward cases. Positioning of the 

E-seal requires careful balance to minimize face seal leakage 

while not causing excessive edge-loading for wear 

considerations. 

Face Seal Height Study 

Thinner face seals were considered in this study for two 

reasons. Thinner face seals present smaller areas over which the 

air pressures can act, which can mitigate the combined 

pressure/friction rotational moments that can occur on the seals. 

These moments can amplify edge-loading effects. Thinner face 

seals also are more flexible which may allow the face seals to 

more easily conform and seal to adjacent seal-carrier to sealcarrier out-of-plane movements. 

Figure 18 shows the calculated contact pressures (in psi) 

for two reduced height face seals: 0.25 and 0.1 in. thick. In 

these cases, the E-seal was positioned inward 0.05 and 0.1 in. 

relative to the baseline, respectively. These E-seal positions 

were chosen for better contact pressure distribution. The results 

shown are for 120 psig engine pressure for three conditions: (a) 

no carrier motion, (b) carrier moving inward, and (c) carrier 

moving outward. The 0.25 in. face seal exhibited the lower 

overall contact stress and did not show edge loading. 

E-seal 
NASA/TM—2006-214114 12 

Figure 16.—Calculated contact pressures (in psi) for baseline face seal configuration with E-seal 

positioned as-built for three conditions: (a) no carrier motion, (b) carrier moving inward, (c) carrier 

moving outward.
NASA/TM—2006-214114 13 

Figure 17.— Calculated contact pressures (in psi) for baseline face seal configuration but with E-seal 

analytically positioned to two different positions: 0.1 in. in radially inward and 0.2 in. from baseline, 

for three conditions all at 120 psig engine pressure: (a) no carrier motion, (b) carrier moving inward, 

(c) carrier moving outward.
NASA/TM—2006-214114 14 

Figure 18.—Calculated contact pressures (in psi) for two reduced height face seals: 0.25 and 0.1 in. 

thick. E-seal positioned inward radially as indicated for best pressure distribution, all at 120 psig 

engine pressure for three conditions: (a) no carrier motion, (b) carrier moving inward, (c) carrier 

moving outward. 
NASA/TM—2006-214114 15 

Figure 19.—Results of grid refinement study. (a) Baseline grid used to run analysis cases. (b) Refined mesh with 4 times the 

elements along the frictional contact surface. Note peak contact pressure (in psi) changed less than 4 percent. 

Furthermore it was the least sensitive to the motion cases 

considered. The 0.10 in. face seal design did not show any edge 

loading. However it did exhibit higher contact stresses in the 

center, presumably because the thinner seal was more flexible. 

The features of these designs make them attractive for further 

experimental study. 

Grid Refinement Study 

The first case, (default geometry, no carrier motion, 

120 psig pressure), was modeled with different face seal 

element sizes to determine whether or not the mesh was 

sufficiently refined. Figure 19 shows that the contact pressure 

changed less than 4 percent with the higher element density 

mesh. Therefore it was determined that the default element 

mesh was acceptable. 

Summary and Conclusions 

An active clearance control system concept and associated 

test rig has been fabricated and installed. The system is being 

used to evaluate different kinematic, seal, actuator, controller, 

and clearance-sensor approaches to achieve a fast-acting, 

mechanical active clearance control system to allow tighter 

turbine tip clearances in future turbine engines. The current 

study experimentally investigated the contact pressure 

distribution between the face seal and the carrier segments and 

the effects of carrier position on ACC system leakage rates. The 

current study investigated the system performance at 

appropriate pressure differentials but at ambient temperature. 

Future studies will examine performance under engine 

simulated pressures and temperatures. 

The current study also analytically investigated, via finite 

element contact analyses, the effects of face-seal to seal-carrier 

contact pressure and edge-loading under several loading 

conditions, several positions or balance diameters of the E-seal, 

and two proposed face seal heights. 

Based on the investigations performed herein, the 

following observations are made: 

• Positioning the seal carriers such that the face seal 

contact zone was nearly coincident with the flexure seal 

diameter resulted in considerably lower leakage. 

System leakage was cut in-half with the E-seal build 

and cut by about 29 percent with the monofilament trial 

seal build. 

• Using the trial seal positioned such that air pressure was 

exerted over greater back-side face-seal surface area 

decreased seal leakage sensitivity to carrier position—a 

desired result. Photographic evidence showed that the 

face seal contact shifted toward its inner diameter. 

• Finite element analyses showed that at high (120 psig) 

air pressures the face seal contact pressure for the 

baseline design was greatest at its outer diameter—

consistent with both the pressure sensitive film 

observations and original scuffing marks observed on 

the seal carriers. 

• Analyses showed the benefit of moving the face seal 

balance diameter inward to increase face-seal back-side 

pressure loading and to more uniformly distribute load 

across the face seal under both inward and outward 

carrier motion, minimizing face seal edge loading. This 

result was corroborated by the leakage flow 

measurements. 

• Analyses also showed the benefit of a thinner face seal 

and inward positioned E-seal. A 0.25 in. thick face seal 

exhibited more uniform contact pressure load across the 

face seal under both inward and outward carrier motion, 

minimizing face seal edge loading. 
NASA/TM—2006-214114 16 

References 

[1] Halila, E.E., Lenahan, D.T., Thomas, T.T., “Energy 

Efficient Engine, High Pressure Turbine Test Hardware 

Detailed Design Report,” NASA CR–167955, 1982. 

[2] Lattime, S.L., Steinetz, B.M., “Turbine Engine Clearance 

Control Systems: Current Practices and Future Directions,” 

Journal of Propulsion and Power, vol. 20, no. 2, NASA/TM—

2002-211794, also AIAA–2002–3790, presented at the 

AIAA/ASME/SAE/ASEE conference, July, 2002, Indianapolis, 

IN. 

[3] Kawecki, E.J., “Thermal Response Turbine Shroud Study,” 

Air Force Aero Propulsion Laboratory Technical Report 

AFAPL–TR–79–2087, 1979. 

[4] Melcher, K.J., and Kypuros, J.A., “Toward a Fast-Response 

Active Turbine Tip Clearance Control,” Proceedings of the 

XVI International Symposium on Air Breathing Engines, 

Cleveland, OH, August 31–September 5, 2003. 

[5] DeCastro, J.A., Melcher, K.M., “A Study on the 

Requirements for Fast Active Turbine Tip Clearance Control 

Systems,” Proceedings of the 40th Joint Propulsion Conference 

and Exhibit, AIAA–2004–4176, 2004. 

[6] Lattime, S.L., Steinetz, Bruce M., Robbie, M., “Test Rig for 

Evaluating Active Turbine Blade Tip Clearance Control 

Concepts,” NASA/TM—2003-212533, also AIAA–2003–4700, 

presented at the AIAA/ASME/SAE/ASEE conference, July, 

2003, Huntsville, AL. Journal of Propulsion and Power, 

vol. 21 (3) May-June 2005. 

[7] Steinetz, B.M., Lattime, S.B., Taylor, S., DeCastro, J.A., 

Oswald, J., Melcher, K.A., “Evaluation of an Active Clearance 

Control System Concept,” NASA/TM—2005-213856, AIAA–

2005–3989. Presented at the 2005 AIAA/ASME/SAE/ASEE 

Joint Propulsion Conference, Tucson, AZ. 

[8] General Electric Aircraft Engines, “HPT Clearance Control 

(Intelligent Engine Systems)—Phase I—Final Report” NASA 

Contract NAS3–01135, April 2004. 

[9] Wiseman, M.W., Guo, T., “An Investigation of Life 

Extending Control Techniques for Gas Turbine Engines,” 

Proceedings of the American Control Conference, IEEE 

Service Center, Piscataway, NJ, IEEE Catalog No. 

01CH37148, vol. 5, pp. 3706–3707, 2001. 

[10] Shapiro, Ascher H., The Dynamics and Thermodynamics 

of Compressible Flow, The Ronald Press Co., New York, 1953.
This publication is available from the NASA Center for AeroSpace Information, 301–621–0390.

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NASA TM—2006-214114

E–15372–1

WBS–22–714–92–56

22

Seal Investigations of an Active Clearance Control System Concept

Bruce M. Steinetz, Shawn Taylor, Jay Oswald, and Jonathan A. DeCastro

Seal; Active clearance control; Design; Leakage; Flow; Actuator; Clearance sensor;

Turbine; High pressure turbine; Turbine engine; Experiments; Analyses; Finite element

Unclassified -Unlimited

Subject Category: 37

Prepared for the 11th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery sponsored

by the Pacific Center of Thermal Fluids Engineering, Honolulu, Hawaii, February 26–March 2, 2006. Bruce M. Steinetz,

NASA Glenn Research Center; Shawn Taylor, University of Toledo, 2801 W. Bancroft Street, Toledo, Ohio 43606;

Jay Oswald, J&J Technical Solutions, 14880 Timberlane, Cleveland, Ohio 44130; and Jonathan A. DeCastro, QSS Group,

Inc., 21000 Brookpark Road, Cleveland, Ohio 44135. Responsible person, Bruce M. Steinetz, organization code RXM,

216–433–3302.

In an effort to improve upon current thermal active clearance control methods, a first generation, fast-acting mechanically

actuated, active clearance control system has been designed and installed into a non-rotating test rig. In order to harvest

the benefit of tighter blade tip clearances, low-leakage seals are required for the actuated carrier segments of the seal

shroud to prevent excessive leakage of compressor discharge (P3) cooling air. The test rig was designed and fabricated to

facilitate the evaluation of these types of seals, identify seal leakage sources, and test other active clearance control

system concepts. The objective of this paper is to present both experimental and analytical investigations into the nature

of the face-seal to seal-carrier interface. Finite element analyses were used to examine face seal contact pressures and

edge-loading under multiple loading conditions, varied E-seal positions and two new face seal heights. The analyses

indicated that moving the E-seal inward radially and reducing face seal height would lead to more uniform contact

conditions between the face seal and the carriers. Lab testing confirmed that moving the balance diameter inward radially

caused a decrease in overall system leakage.