Bruce M. Steinetz

Glenn Research Center, Cleveland, Ohio

Scott B. Lattime

The Timken Company, Canton, Ohio

Shawn Taylor

University of Toledo, Toledo, Ohio

Jonathan A. DeCastro

QSS Group, Inc., Cleveland, Ohio

Jay Oswald

J&J Technical Solutions, Cleveland, Ohio

Kevin J. Melcher

Glenn Research Center, Cleveland, Ohio

Evaluation of an Active Clearance Control

System Concept

NASA/TM—2005-213856

November 2005

AIAA–2005–3989
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Bruce M. Steinetz

Glenn Research Center, Cleveland, Ohio

Scott B. Lattime

The Timken Company, Canton, Ohio

Shawn Taylor

University of Toledo, Toledo, Ohio

Jonathan A. DeCastro

QSS Group, Inc., Cleveland, Ohio

Jay Oswald

J&J Technical Solutions, Cleveland, Ohio

Kevin J. Melcher

Glenn Research Center, Cleveland, Ohio

Evaluation of an Active Clearance Control

System Concept

NASA/TM—2005-213856

November 2005

National Aeronautics and

Space Administration

Glenn Research Center

Prepared for the

41st Joint Propulsion Conference and Exhibit

cosponsored by the AIAA, ASME, SAE, and ASEE

Tucson, Arizona, July 10–13, 2005

AIAA–2005–3989
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: Malcolm Robbie, Art Erker, and

Tobie Mintz for their design and fabrication support; Richard Tashjian, Kevin McCormick, Tom Lawrence, Mike

McGhee, and Joe Wisniewski for their technical assistance during the test rig installation and troubleshooting. And

finally, the authors acknowledge Keith Blackwell for his efforts in characterizing the stepper motor load versus speed

characteristics, and Robert Hendricks and Chris Daniels for their 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—2005-213856 1

Evaluation of an Active Clearance Control System Concept 

Bruce M. Steinetz

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Scott B. Lattime 

The Timken Company 

Canton, Ohio 44706 

Shawn Taylor

University of Toledo 

Toledo, Ohio 43606

Jonathan A. DeCastro 

QSS Group, Inc. 

Cleveland, Ohio 44135 

Jay Oswald 

J&J Technical Solutions 

Cleveland, Ohio 44130

Kevin J. Melcher 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Abstract

Reducing blade tip clearances through active tip 

clearance control in the high pressure turbine can lead to 

significant reductions in emissions and specific fuel 

consumption as well as dramatic improvements in operating 

efficiency and increased service life. Current engines 

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 forcing cold 

build clearances to set the minimum clearances at extreme 

operating conditions (e.g., takeoff, reburst) and not allowing 

cruise clearances to be minimized due to the possibility of 

throttle transients (e.g., step change in altitude). In an effort 

to improve upon current thermal methods, a first generation 

mechanically-actuated active clearance control (ACC) 

system has been designed and fabricated. The system 

utilizes independent actuators, a segmented shroud structure, 

and clearance measurement feedback to provide fast and 

precise active clearance control throughout engine 

operation. 

Ambient temperature performance tests of this first 

generation ACC system assessed individual seal component 

leakage rates and both static and dynamic overall system 

leakage rates. The ability of the nine electric stepper motors 

to control the position of the seal carriers in both open- and 

closed-loop control modes for single and multiple cycles 

was investigated. The ability of the system to follow 

simulated engine clearance transients in closed-loop mode 

showed the system was able to track clearances to within a 

tight tolerance (≤ 0.001 in. error). 

I. Introduction 

Gas path sealing continues to be a fundamental concern 

in aircraft and ground-based turbine engines. Technical 

drivers include the need to operate the engines under 

increasingly more demanding temperature and pressure 

conditions at high efficiency and to minimize maintenance 

requirements to keep operating costs at their minimum. 

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 600 psi, high surface speeds 

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

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

for a controllable seal design. Clearances change during 

engine operation as a result of both mechanical (pressure 

and rotational) loads and thermal loads. Discrete structural 

mounts coupled with thermal and flight loads result in nonuniform distortions and hence non-uniform clearances. 
NASA/TM—2005-213856 2

Figure 1.—HPT blade tip seal location in a modern gas turbine engine.1

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), compressor stall margin and engine efficiency as 

well as increased payload and mission range capabilities. 

Implementation of active clearance control systems, 

especially in the HPT, can dramatically improve engine 

service life or time-on-wing. 

A. 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 over the operating points of the engine (e.g., ground 

idle, takeoff, cruise, decel, etc.) as well as over the cycle life 

of the engine. 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), aero-dynamic (external pressure), and 

gyroscopic loads. Engine loads can produce both 

axisymmetric and asymmetric clearance changes (see fig. 2). 

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 and allowed to 

windmill for a while followed by restart 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 2500 °F and higher rotor inlet gas 

temperatures. The cooling air is also used to purge the 

leading and trailing edges of the shroud segments, providing 

a positive back-flow margin from the hot rotor inlet flow. 

This cooling flow is shown in figure 3 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 
NASA/TM—2005-213856 3

Figure 2.—(a) Axisymmetric clearance change, 

(b) asymmetric clearance change. 

Figure 3.—Axial pressure distribution across 

HPT blade tip seal cross section.2

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 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-psi across the 

shroud. 

Wear mechanisms of blade tip seals can be generally 

separated into two categories, namely, rubbing (blade 

incursion), and erosion. Blade rubs and erosion are expected 

to occur throughout the service life of an engine. Engine 

build clearances are in fact sometimes chosen to operate 

line-to-line at extreme operating conditions (e.g., takeoff, reaccel). In this case, manufacturers set cold-build clearances 

such that during the engine green-state (“run-in”) the blades 

will lightly rub the seal shrouds thereby achieving tight 

clearances at one operating point and mitigating the effects 

of manufacturing tolerance stack-up. Rubs can also 

contribute to accelerating the effects of both erosion and 

thermal fatigue by wearing protective coatings (e.g., thermal 

barrier coatings) or distorting cooling passages of the blade 

tips during an incursion event. 

Lattime and Steinetz2

 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. 

Kawecki3 presented trade studies of a variety of approaches 

for active clearance control identifying a variety of fastacting mechanical and novel thermal control systems. 

Melcher and Kypuros4

 outlined NASA Glenn’s general 

approach for developing a fast-response ACC system and 

DeCastro and Melcher5

 examined the control systems 

requirements for fast acting ACC systems. 

B. 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,7 and Wiseman and Guo8

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 reduction in of 0.010-in tip clearance 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 improves 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 temperature can result in increased life of 

hot section components and engine time-on-wing (up to 

1000 cycles). Additional benefits include increased payload 

and mission range capabilities. 

C. 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 
NASA/TM—2005-213856 4

upon the case-cooling methods employed today. The 

objective of the current work is to present the preliminary 

evaluation of a first generation mechanical active clearance 

control (ACC) system that utilizes independent actuators to 

position a series of seal carriers or hangars that would 

support seal segments or shrouds in a future engine system. 

Specific test objectives of this study include the 

evaluation of: 

• Individual component seal leakages (e.g., actuator 

rod piston rings, face seals, flexure/spline seals) 

under engine simulated pressure conditions at 

ambient temperature. 

• Overall system leakage both statically and during 

motion. 

• Candidate actuators’ ability to position the seal 

carriers at the required rate, accuracy, and 

repeatability under engine simulated pressure 

conditions. 

• Candidate clearance sensors as part of the ACC 

closed-loop feedback control system. 

II. Test Apparatus and Procedures

A. Apparatus 

An active clearance control system test rig has been 

developed and installed (fig. 4). The key features of the test 

rig examined during this study are summarized below. For a 

comprehensive review of the test rig design approach, the 

reader is directed to Lattime and Steinetz.6 

1. 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 

Figure 4.—ACC test rig with housing lid and 

chamber cover plate removed for clarity. 

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 requirements. Table 1 compares the 

main characteristics of both the ACC test rig and a typical 

modern high bypass ratio engine. 

The general design of the test rig is shown in figures 4 

to 7. The rig main housing consists of two concentric 

cylinders, which form an annular cavity. An annular radiant 

heater made of upper and lower halves surrounds the seal 

pressure 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 the 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 

 

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 
NASA/TM—2005-213856 5

Figure 5.—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 

 

Figure 7.—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. 

Figure 6.—Actuator rod foot, two adjacent seal carriers, 

flexure seals, materials, and nominal dimensions 

in inches. 
NASA/TM—2005-213856 6

Figure 8.—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. 

are secured to one another through a pinned and slotted link 

as shown in figure 6. This link (or “foot”) positions the 

carrier segments radially while allowing relative 

circumferential movement or dilation of the seal carrier 

segments. Figure 5 shows that the carrier segment is 

supported by two independent actuator rods. A photograph 

of test rig is shown in figure 8. Here the precision alignment 

fixture used to establish concentricity of all of the seal 

carriers 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 at the 

beginning of each test to set the nominal or “home” position 

for subsequent tests. 

2. Chamber and Seal Detail 

A pressurized 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. 5 and 7). The joints between adjoining carrier 

segments are sealed with thin flexure seals otherwise known 

as spline seals (figs. 6 and 7). 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. 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 fixtures, and the actuator rod guides. These are sealed 

in their respective bosses via a shrink fit qualified during 

hydro-testing. 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 innerand outer-flange C-rings. Only the outer C-rings carry 

pressure during operation. The pipes supplying high 

pressure to the chamber (fig. 5) are sealed using two sets of 

concentric piston ring seals. 

3. 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 found in table 2. The stepper motors are 

rated for 500 lbf of actuation force which 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 lowcost 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 servo-hydraulic 

actuators will come equipped with linear variable 

differential transducers (LVDTs) to accurately measure 

actuator piston location. 

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 (lb) 5 

Manufacturer Haydon Switch and 

Instrument Inc. 

Model no. 87H43–12–002 

The stepper motor shafts were connected to the actuator 

rods through a coupling, as shown in figure 9. The coupling 

was manufactured with two diameters to serve two 

additional purposes: (1) as a mechanical stop to prevent 

excessive axial motion; and (2) as a convenient means to 

measure absolute positions of the actuator shaft relative to 

the actuator mount. The actuators were mounted to the 

mounting flanges through a series of stand-offs. 

4. 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 
NASA/TM—2005-213856 7

Figure 9.—Stepper motor mounting arrangement showing 

means for measuring actuator shaft/coupling position 

relative to the actuator mount. 

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 meters 

Low Range 

 Manufacturer Teledyne-Hastings 

 Model no. HFM-200 

 Accuracy (SCFM; lbm/s) 0.2; 0.00027 

 Range (SCFM; SLPM) 173, 6.11 

High Range

 Manufacturer Teledyne-Hastings 

 Model no. HFM 200 + 3000 

SLPM LFE 

 Accuracy (SCFM; lbm/s) 1.06; 0.0014 

 Range (SCFM; SLPM) 106; 3000 

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 (lb) 0.04 

Manufacturer Capacitec 

Model no. HPC 150 

clearance probes were used to measure the radial 

displacement of the seal carriers (fig. 7) at three different 

circumferential locations. 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. 

5. 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. 

6. 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 

the 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. 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. 

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 

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.

B. Procedures 

1. Seal Component Leakage Rates 

A goal of the current work is to identify how much each 

seal component contributes to the overall leakage. This will 

guide future seal development efforts to reduce component 

leakages and overall ACC system leakage to their practical 
NASA/TM—2005-213856 8

minimums. Low ACC parasitic leakage is required in order 

to minimize P3 cooling air loss and to net the engine 

efficiency gains from the tighter tip clearance control. 

To quantify the individual leakage levels of each sealing 

component utilized in the ACC test rig, five different static 

flow test setups were used. These flow tests were conducted 

at room temperature using pressure differentials ranging 

from 0 to 120 psig in 20 psig increments. Each test setup 

was strategically designed to allow the leakage of each seal 

component to be isolated from the total leakage of the 

composite test system as a whole. For the low flow 

component tests (all except the flexure seals), air was 

supplied to the system through a 3/8 in. port on the upper 

chamber cover. Chamber pressure was monitored through a 

second 3/8 in. cover plate port, and flow rate and an 

additional pressure measurement were collected from the 

system air supply line. Each seal utilized in the ACC rig is 

listed below with a description of the configuration used to 

quantify its contribution to total system leakage. 

C-Seal Leakage Evaluation: To evaluate the 

performance of the C-seals, system leakage was limited to 

the outer C-seal on the lower chamber cover. This was 

achieved by plugging the actuator rod and air inlet ports, 

placing elastomeric O-rings in the upper cover plate instead 

of the upper C-seals, and installing a solid aluminum shroud 

in place of the seal carriers. This shroud was machined to 

have the same dimensions as the ring of seal carriers 

normally present in the test chamber, and utilized O-ring 

seals on both top and bottom to limit chamber pressurization 

to the outer half of the test chamber. Because the lower 

C-seals were difficult to remove, they were left in place for 

all subsequent leakage tests. The leakage rates for those tests 

were corrected to compensate for the additional 

leakage path. 

Air Inlet Piston Ring Seals: The effectiveness of the air 

inlet piston ring seals was evaluated using a test setup nearly 

identical to that used for the C-seal tests; however the plugs 

were removed from the air inlet ports and replaced with 

blanked air supply pipes and corresponding piston ring 

seals. All three air inlet piston ring station leakages were 

measured simultaneously. 

Actuator Rod Piston Ring Seals: To determine the 

leakage of the actuator rod piston rings, the actuator rods 

and piston rings were installed in their respective locations 

and the actuator rod feet were installed. Aluminum spacer 

blocks were positioned between the outer chamber wall and 

the back side of each foot to hold the rods in place during 

pressurization. All three air inlet ports were plugged and the 

solid shroud and cover plate with O-rings were positioned 

into their respective locations. Air was supplied through the 

3/8 in. port as described earlier. 

Face Seals: During face seal testing, both the air inlet 

and actuator rod ports were plugged, and O-rings were used 

in the upper cover plate. During these tests, however, the 

O-rings were removed from the solid shroud and the upperand lower-face seals sealed against the sides of the solid 

shroud. The E-seals used to preload the face seals were also 

installed. 

Flexure Seals: The flexure seal leakage was considerably 

greater than the previous leakage rates. To avoid choking 

through the 3/8 in. inlet port, air was supplied using one of 

the 3 air inlet ports. The other two remained capped with 

their respective O-ring sealed plugs. Due to the higher 

leakage flow rates, a higher capacity (106 SCFM, 

0.143 lbm/s) flow meter (Teledyne Hastings) was used. All 

other seals were installed for these tests (C-seals, face seals, 

actuator rod piston rings). The leakage of the flexure seals 

was extracted from the total leakage data by subtracting the 

known leakage contributions of the other seals in the system 

from the composite flow rate. The measured flow is an 

aggregate for all nine flexure seals. 

2. Effect of Carrier Position on Leakage 

Static Leakage: To gauge the ability of the ACC rig seals 

to block flow at multiple radial positions, static flow tests 

were conducted at pressures including 10 and 30 psig. The 

upper test pressure was limited to 30 psig due to the load 

capabilities of the stepper motors used for actuation. Open 

loop actuator control was used. The test chamber was 

pressurized with the seal carriers in their nominal or “home” 

position. The nominal position was defined as the diameter 

established by the alignment fixture shown engaging the 

seal carrier pins in figure 8. Once pressurized, the seal 

carriers were actuated radially outward to a position 

0.040 in. from the “home” position. Next, the carriers were 

moved radially inward at a rate of 0.005 in./sec., pausing at 

three locations (0.035 in. radially outward from the “home,” 

“home,” and 0.035 in. radially inward from “home”) to 

collect static leakage values. (For reference purposes, 

positive sign convention represents movement inward 

radially from the “home” position). Inward motion was 

stopped when the carriers reached a location 0.040 in. 

radially inward from the “home” position. At this point, 

motion was reversed and leakage values were collected at 

the same three locations during the outward stroke. This 

method was applied to determine whether static seal leakage 

was dependent on the direction of seal carrier travel prior to 

stopping. 

Dynamic Leakage: Dynamic leakage values were 

gathered using the same open-loop displacement control 

scheme employed for the static leakage tests described 

previously. However, instead of stopping to record static 

leakage values, system leakage was continuously logged 

(2 Hz) while the seal carriers were actuated through the 

displacement range. Seal carrier motion was limited to 

0.001 in./second, as dictated by the “settling time” of the 

flow meter used to monitor leakage rates. 

3. Open Loop Positional Accuracy Test 

The ability of the ACC rig to move as directed was 

evaluated through a series of open loop positional accuracy 

tests conducted at ambient temperature and pressures of 5, 
NASA/TM—2005-213856 9

10, 20, 30, and 35 psig. To start each test, the seal carriers 

were sent to their nominal position (“home”) and the test 

chamber was pressurized to the desired test pressure. Once 

pressurized, positional reference values were collected by 

measuring the clearance between the actuator rod hard stop 

couplings and the actuator mounts using gage pins. These 

reference values were recorded for each of the nine actuator 

axes. From “home”, the seal carriers were actuated radially 

inward 0.035 in., back to home, radially outward 0.035 in., 

and then back again to home (one complete cycle). At each 

location, positional reference values were recorded using 

gage pin measurements. The positional reference values 

were then used to determine whether the actuator rods 

moved the directed distance. Positional reference 

measurement accuracy was limited by a 0.001 in. gage pin 

resolution. 

4. Displacement Repeatability Study 

The ability of the ACC rig to track a series of 

displacements and then return to the “home” position was 

evaluated by imposing a commanded sine wave 

displacement (0.035 in. amplitude) on the system while it 

was pressurized to 20 psig. Initial reference locations were 

recorded using gauge pin measurements once the test 

chamber was pressurized. The seal carriers were actuated 

radially inward and outward for 20 cycles with the seal 

carrier speed limited to a maximum of 0.010 in./sec. Post 

cycling reference measurements were recorded for 

comparison with the original seal carrier locations. 

5. Control System and Simulated Engine Clearance Versus 

Time Study 

Figure 10 summarizes the overall control logic strategy 

used in the ACC test rig system. The control system is 

composed of an inner loop where the position of each of the 

nine stepper motors is monitored via encoders. The outerloop or executive level controls the overall motion of the 

seal carriers receiving input from three capacitance 

clearance probes (see table 4 for probe details). The control 

system constantly checks to see which probe is at the 

minimum clearance and uses that probe for control for a 

conservative approach. To investigate the control system’s 

behavior during transient events such as take-off, desired 

carrier position (e.g., set point) signals are fed to the control 

system at the outer-loop level. 

During a take-off transient, the rotor expands due to 

centrifugal loads and blades expand quickly due to 

combined thermal and centrifugal loads. The case does not 

expand nearly as fast, and consequently, the ACC system 

must move the seal carrier segments radially outward 

against pressure loads to maintain blade-tip clearance. 

Clearance control must be demonstrated with small tracking 

errors sufficiently small to avoid blade rubs (0.005-in. or 

less). A 200-second take-off transient profile representative 

of a large commercial engine that exhibits a clearance 

change of 45 mils in 10 seconds was used to demonstrate 

closed-loop control. Clearance control tests were performed 

using chamber pressures of 10, 20, and 30 psig. 

 

Figure 10.—Closed-loop active clearance control implementation. 
NASA/TM—2005-213856 10

III. Results and Discussion 

A. Baseline Seal Component Leakage Results 

Leakage rates for each of the ACC test rig seals were 

measured using the procedures described above at ambient 

temperature. Figure 11 shows air leakage rates and for the 

most part the leakage rates vary linearly with pressure drop. 

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. 

Table 3 provides flow measurement accuracies for both the 

low and high flow ranges examined. 

B. 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, we make an initial 

comparison of the measured losses in terms of effective 

leakage flow area per inch of circumference to an engine 

industry reference level. 

Using the total measured leakage rates of all seal 

locations, we back-calculated an effective leakage flow area 

for the entire active clearance control system, 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.9

 The leakage flow was considered 

choked since the pressure ratio of 8.3 

(i.e., Psupply/Pexhaust = 122.7 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/lbf-sec

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 

Figure 11.—Seal component leakage rates 

versus pressure at ambient temperature. 

Figure 12.—Comparison of ACC test rig effective flow area 

per inch circumference (back-calculated from leakage 

flow data) to engine reference flow area per inch 

circumference for forward and aft seal locations. 

(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 12 compares the effective flow area per unit 

circumference found for the rig 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. 
NASA/TM—2005-213856 11

Figure 13.—Flow versus displacement at three carrier 

radial positions: nominal and displaced ±0.035-in. 

from nominal. Positive is radially inward. Figure also 

shows effect of shroud movement-direction (prior to 

stopping) on leakage rate. Conditions: pressure 

30 psig, static condition. 

Figure 14.—Flow versus displacement for dynamic test. 

Positive is radially inward. Figure also shows effect of 

shroud movement-direction on leakage rate. Conditions: 

pressure 30 psig, seal shroud moving at 0.001-in./sec. 

C. Effect of Seal Carrier Position on ACC Leakage: 

Static and Dynamic Tests 

Figure 13 shows the result of the static leakage versus 

seal carrier position tests at a chamber pressure of 30 psig. 

The ACC flow rates show a hysteresis effect and flows are 

greatest when the seal carriers are closer together and are 

moving radially outward. The largest difference in the flow 

is 0.006 lbm/s and occurs at the inward position. This 

represents about a ±15 percent variation in flow relative to 

the average flow. The static flow rate at 10 psig is similar to 

the 30 psig results—just at a lower flow level, as one would 

expect. 

Figure 14 shows the result of the dynamic leakage versus 

seal carrier position tests at a chamber pressure of 30 psig. 

Figure 15.—Diagram of flexure seals in seal carriers in 

two states. (a) Seal carriers positioned radially-in but 

moving apart; (b) Seal carriers positioned radially-out 

but moving toward one another. Note when gap is 

open the normal loads on the flexure seals are 

greater. The flexure seal is shown in a deformed 

condition (enlarged for clarity). 

The dynamic leakage again shows a hysteresis effect and is 

also greatest when the seal carriers are closer together and 

moving outward radially. Similar to above there is about a 

±15 percent variation in flow relative to the average flow. 

The correlation between static and dynamic leakages 

indicate no strong dependence on actuation rate, at least at 

the 0.001 in./sec tested herein. However comparing the 

dynamic flows from figure 14 (~0.024 lbm/s) to the total 

static leakage flow in figure 11 (0.03 lbm/s), we note that 

the dynamic flows are slightly less. A likely explanation is 

that motion and pressure are seating the seals lowering the 

leakage flows. The dynamic flow behavior at 10 psig was 

similar to the 30 psig results—albeit at lower leakages, as 

would be expected. 

Figures 13 and 14 indicate that the ACC flow rates are 

greater when the seal carriers are closer together than when 

they are further apart. Figure 15 depicts the a flexure seal 

supported by two adjacent seal carriers in two states: 

(a) Seal carriers positioned radially-in but moving apart; 

(b) Seal carriers positioned radially-out but moving toward 

one another. A possible explanation for the noted difference 

in flow appears to be related to the unsupported pressure 

loads in the two states and the direction of motion. Note 

when the seal carriers are fully apart, the unsupported area 

and hence the resultant force is greatest. Under these 

conditions (e.g., radially outward), the higher flexure seal 

contact loads would tend to reduce the leakage flows. The 

direction of motion also seems to play a role in the leakage 
NASA/TM—2005-213856 12

flow. One possible explanation is that as the seal carriers are 

moving toward one another friction loads tend to further 

increase the contact loads on the slightly deformed flexure 

(see deformed condition in diagram) in the interface 

between the flexure seal and the carriers. Another possible 

explanation is possible leakage-directional dependence of 

the face seals. Further investigations are required to examine 

these theories. 

D. Open Loop Positional Accuracy Test 

Figure 16 shows the results for the open-loop positional 

accuracy test at a pressure of 20 psig. Results for this 

pressure are shown as being representative of other 

pressures tested. In the 20 psig tests the nine stepper motors 

were issued a command to move the equivalent number of 

motor steps to move the actuator rods ±0.035 in. from the 

nominal or “home” position. In this test, the stepper motor 

encoders were not providing feedback to the controller. 

Figure 16(a) shows the actuator position as measured with 

the pin gages. Figure 16(b) shows the error. The maximum 

error observed for these tests was about 0.003-in. These 

results indicate evidence of either “slipped counts” or 

“positional hysteresis” which is typical for open-loop 

stepper motor operation near their load limit. In the actual 

application feedback control would be required. 

The positional accuracy tests performed herein were 

done with pressure loading to simulate the “loaded” 

condition in the application. Without pressure loading openloop positional accuracy was worse due to small clearances 

between load train components (or backlash). Without 

pressure loading backlash up to 0.013 in. was measured. 

E. Repeatability Study 

The ACC rig was cycled in open-loop position feedback 

mode to evaluate the system repeatability and ability to 

return to its nominal or “home” position. In this test, the 

system was pressurized to 20 psig and cycled 20 times. 

Carrier positions were measured using three capacitance 

probes: two diametrally opposite one-another and a third at 

90°. Actuator positions were checked using precision pin 

gages as described in the Test Procedures Section. 

Table 5 provides the results of these tests. After 

20 cycles all nine actuators returned to their original or 

“home” position to within ±0.001-in. showing excellent 

repeatability. For reference purposes, pin gages used were 

calibrated in 0.001-in. increments so the measured deviation 

was within the limits of the measuring technique. The 

convention used herein is positive deviation is radially 

inward from home and negative deviation is radially 

outward from the home position. Note that the there is some 

variation in pre-/post-test position relative to the actuator 

mount measurements for the nine axes due to slight 

differences in actuator rod lengths and shims used in the 

mounting of the stepper motors. This in no way affected the 

position repeatability results. 

Figure 16.—ACC open-loop position control evaluation for 

three positions: nominal and ±0.035-in. for a complete 

cycle. (a) Actuator position measured using manual pin 

gages, (b) Positional error: commanded versus actual. 

Conditions: pressure 20 psig, ambient temperature. Sign 

convention: positive error indicates inward radial relative 

to the commanded position. 

F. Simulated Engine Clearance Versus Time Study 

Figure 17 shows the result of the simulated engine 

clearance versus time study where the set-point position 

defines ideal carrier position movement versus time for a 

simulated take-off condition. These tests were performed at 

a pressure of 20 psig. Figure 17(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 
NASA/TM—2005-213856 13

TABLE 5.—REPEATABILITY TEST RESULTS: 

CONDITIONS: 20 CYCLES STARTING AT NOMINAL 

POSITION, MOVING ±0.035-in. FROM NOMINAL 

POSITION AND RETURNING HOME. 20 psig 

PRESSURE, AMBIENT TEMPERATURE. 

ERROR IS DIFFERENCE IN POSITION 

BEFORE AND AFTER TEST 

Individual Actuator 

Axis Position 

Axis 

number 

Pre-test position 

relative to 

actuator 

mount 

(in.) 

Post-test position 

relative to 

actuator 

mount 

(in.) 

Error 

(in.) 

1 0.356 0.356 0.000 

2 0.416 0.415 0.001 

3 0.423 0.424 –0.001 

4 0.345 0.345 0.000 

5 0.393 0.393 0.000 

6 0.404 0.404 0.000 

7 0.390 0.389 0.001 

8 0.428 0.428 0.000 

9 0.403 0.404 –0.001 

Individual Capacitance 

 Probe Clearance 

Probe 

number 

Pre-test 

clearance 

(in.) 

Post-test 

clearance 

(in.) 

Error 

(in.) 

1 0.051 0.051 0.000 

2 0.054 0.054 0.000 

3 0.053 0.053 0.000 

Notes: 

• Positive deviation is radially inward from home and negative 

deviation is radially outward from the home position. 

• Probe 1 and 3 are diametrically opposed. Probe 2 is clocked 90° 

between probes 1 and 3. 

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 17(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 this 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 (e.g., 

30 psig) pressure load and corresponding frictional effects 

Figure 17.—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. 

were just above what the stepper motor could reliably 

actuate without slowing down. Future hydraulic actuators 

are designed to overcome this shortcoming. 

IV. 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, 

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 investigated the abilities 

of the ACC system concept to control leakage to acceptable 

levels and to position seal carriers at the appropriate 
NASA/TM—2005-213856 14

actuation rates and accuracies. 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. 

Based on the tests performed herein, the following 

observations are made: 

• The kinematic system permits movement of the 

seal carriers over a range of 0.080-in. This range 

easily accommodates the anticipated range required 

for future turbine engines. 

• The system leakage was tested at pressures up to 

120 psig—comparable to pressure differentials 

anticipated in future engines. Leakage unit flow 

area was comparable to an engine industry 

reference level. 

• The system repeatability was examined at the limit 

of the stepper motor loads (20 psig pressure loads) 

showing that after 20 cycles (±0.035 in. 

inward/outward radial movement) that the 

positional error was within the 0.001-in. resolution 

of the gage pin measurement technique. 

• Tests showed that the closed-loop position control 

followed the set-point to less than 0.001-in. for a 

simulated engine take-off clearance change with 

capacitance clearance probes in the loop. This was 

an encouraging result as it showed the kinematic 

system could track simulated clearance changes at 

the correct rate and range under load with an 

acceptable error. 

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, Scott L., Steinetz, Bruce 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) MayJune 2005. 7

General Electric Aircraft Engines, “HPT Clearance Control 

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

Contract NAS3–01135, April 2004. 8

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. 9

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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AIAA–2005–3989

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20

Evaluation of an Active Clearance Control System Concept

Bruce M. Steinetz, Scott B. Lattime, Shawn Taylor, Jonathan A. DeCastro,

Jay Oswald, and Kevin J. Melcher

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

Turbine; High pressure turbine

Unclassified -Unlimited

Subject Category: 37

Prepared for the 41st Joint Propulsion Conference and Exhibit cosponsored by the AIAA, ASME, SAE, and

ASEE, Tucson, Arizona, July 10–13, 2005. Bruce M. Steinetz and Kevin J. Melcher, NASA Glenn Research Center;

Scott B. Lattime, The Timken Company, 1835 Dueber Ave. SW, Canton, Ohio 44706; Shawn Taylor, University of

Toledo, 2801 W. Bancroft Street, Toledo, Ohio 43606; Jonathan A. DeCastro, QSS Group, Inc., 21000 Brookpark

Road, Cleveland, Ohio 44135; and Jay Oswald, J&J Technical Solutions, 14880 Timber Lane, Cleveland, Ohio 44130.

Responsible person, Bruce M. Steinetz, organization code RSM, 216–433–3302.

Reducing blade tip clearances through active tip clearance control in the high pressure turbine can lead to significant

reductions in emissions and specific fuel consumption as well as dramatic improvements in operating efficiency and

increased service life. Current engines 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 forcing cold build clearances to set the minimum clearances at extreme operating conditions (e.g., takeoff,

reburst) and not allowing cruise clearances to be minimized due to the possibility of throttle transients (e.g., step change

in altitude). In an effort to improve upon current thermal methods, a first generation mechanically-actuated active

clearance control (ACC) system has been designed and fabricated. The system utilizes independent actuators, a segmented shroud structure, and clearance measurement feedback to provide fast and precise active clearance control

throughout engine operation. Ambient temperature performance tests of this first generation ACC system assessed

individual seal component leakage rates and both static and dynamic overall system leakage rates. The ability of the nine

electric stepper motors to control the position of the seal carriers in both open- and closed-loop control modes for single

and multiple cycles was investigated. The ability of the system to follow simulated engine clearance transients in closedloop mode showed the system was able to track clearances to within a tight tolerance (≤ 0.001 in. error).