Fred B. Oswald

Glenn Research Center, Cleveland, Ohio

Michael Savage

University of Akron, Akron, Ohio

Erwin V. Zaretsky

Glenn Research Center, Cleveland, Ohio

Space Shuttle Rudder/Speed Brake Actuator—A Case Study. 

Probabilistic Fatigue Life and Reliability Analysis

NASA/TM—2015-218846

July 2015
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Fred B. Oswald

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Michael Savage

University of Akron, Akron, Ohio

Erwin V. Zaretsky

Glenn Research Center, Cleveland, Ohio

Space Shuttle Rudder/Speed Brake Actuator—A Case Study. 

Probabilistic Fatigue Life and Reliability Analysis

NASA/TM—2015-218846

July 2015

National Aeronautics and

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Glenn Research Center

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NASA/TM—2015-218846 1 

Space Shuttle Rudder/Speed Brake Actuator—A Case Study. 

Probabilistic Fatigue Life and Reliability Analysis 

Fred B. Oswald*

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Michael Savage

University of Akron 

Akron, Ohio 44325 

Erwin V. Zaretsky†

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Abstract 

The U.S. Space Shuttle fleet was originally intended to have a life of 100 flights for each vehicle, 

lasting over a 10-year period, with minimal scheduled maintenance or inspection. The first space shuttle 

flight was that of the Space Shuttle Columbia (OV-102), launched April 12, 1981. The disaster that 

destroyed Columbia occurred on its 28th flight, February 1, 2003, nearly 22 years after its first launch. In 

order to minimize risk of losing another Space Shuttle, a probabilistic life and reliability analysis was 

conducted for the Space Shuttle rudder/speed brake actuators to determine the number of flights the 

actuators could sustain. A life and reliability assessment of the actuator gears was performed in two 

stages: a contact stress fatigue model and a gear tooth bending fatigue model. For the contact stress 

analysis, the Lundberg-Palmgren bearing life theory was expanded to include gear-surface pitting for the 

actuator as a system. The mission spectrum of the Space Shuttle rudder/speed brake actuator was 

combined into equivalent effective hinge moment loads including an actuator input preload for the contact 

stress fatigue and tooth bending fatigue models. Gear system reliabilities are reported for both models and 

their combination. Reliability of the actuator bearings was analyzed separately, based on data provided by 

the actuator manufacturer. As a result of the analysis, the reliability of one half of a single actuator was 

calculated to be 98.6 percent for 12 flights. Accordingly, each actuator was subsequently limited to 

12 flights before removal from service in the Space Shuttle. 

Introduction 

The U.S. Space Shuttle fleet was originally intended to have a life of 100 flights for each vehicle, 

lasting over a 10-year period, with minimal scheduled maintenance or inspection. The first space shuttle 

flight was that of the Space Shuttle Columbia (OV-102), launched April 12, 1981. The disaster that 

destroyed Columbia occurred on its 28th flight, February 1, 2003, nearly 22 years after its first launch. 

The Space Shuttle actuators are lubricated with space-qualified National Lubricating Grease Institute 

grade 2 perfluoropolyalkyl ether (PFPAE) grease and were intended to operate for life without periodic 

re-lubrication and maintenance (Refs. 1 and 2). Many of these actuators were operated without 

maintenance in excess of 20 years. 

 

*

Retired. 

†

Distinguished Research Associate. 
NASA/TM—2015-218846 2 

During Space Shuttle actuator inspection and refurbishment after the Columbia Space Shuttle 

disaster, there were external signs of corrosion on one of the Space Shuttle Rudder/Speed Brake (RSB) 

Actuators. Visible inspection of two partially disassembled RSB actuators from the Space Shuttle 

Discovery was made on September 16, 2003. Pitted gears and discolored grease were observed inside. 

The grease-lubricated ball bearings and gears making up the actuators exhibited various degrees of wear. 

In some cases the wear was severe. Both the bearings and gears operated under a dithering (rotation 

reversal) motion when these systems were powered during ground operations (Ref. 2). 

The observations summarized above led to several research efforts to determine the causes of the 

grease degradation, damage and wear to estimate the future life and reliability of the shuttle actuators and 

to investigate design improvements for future heavily-loaded space mechanisms. These works included 

Morales, et al. (Ref. 1), which analyzed the condition of the grease in two actuators after 39 flights, 

Oswald, et al. (Ref. 2), which describes a probabilistic analysis of the life and reliability of Body Flap 

Actuator (BSA) bearings, Krantz, et al. (Ref. 3) and Krantz and Handschuh (Refs. 4), which examined 

wear of spur gears lubricated by PFPAE grease, Proctor, et al. (Ref. 5), which describes experiments to 

characterize gear scuffing damage observed on one gear of the power drive unit that operates the 

Rudder/Speed Brake (RSB) Actuators, and Handschuh, and Krantz (Ref. 6), which investigated a concern 

that a tooth might break off from a planet gear, potentially jamming the mechanism. 

In order to minimize the risk of losing another Space Shuttle, Oswald, et al. (Ref. 2) performed 

experiments on a test rig under simulated conditions to determine the life and failure mechanism of the 

grease lubricated Space Shuttle Body Flap Actuator (BSA) bearings that support the input shaft of the 

space shuttle body flap actuators. The failure mechanism was wear that can cause loss of bearing preload. 

These experimental life data were analyzed using the 2-parameter Weibull-Johnson method (Refs. 7 and 

8) on experimental life test data from bearings. These tests established life and reliability data for both 

shuttle flight and ground operation. Test data were used to estimate the failure rate and reliability as a 

function of the number of shuttle missions flown. The Weibull analysis of the test data for the four 

actuators on one shuttle, each with a 2-bearing shaft assembly, established a reliability level of 

96.9 percent for a life of 12 missions. A probabilistic system analysis for four shuttles, each of which has 

four actuators, predicts a single bearing failure in one actuator of one shuttle after 22 missions (a total of 

88 missions for a 4-shuttle fleet). This prediction is comparable with actual shuttle flight history in which 

a single actuator bearing was found to have failed by wear at 20 missions. 

The work of Oswald, et al. (Ref. 2) was extended to perform a probabilistic life and reliability 

analysis of the Space Shuttle Rudder/Speed Brake (RSB) Actuators to determine the number of flights 

and/or the probability of failure that the actuators could sustain based on rolling-element (contact) fatigue 

of the rolling-element bearings and gears and tooth bending fatigue for a minimum of 12 Space Shuttle 

missions. (The 12 mission life requirement was based on the test data for the Space Shuttle Body Flap 

Actuator (BSA) bearings discussed above.) Accordingly, the objectives of the work reported were to 

(1) determine the life and reliability of each of the gears and bearings in a Rudder/Speed Brake (RSB) 

Actuator and (2) extend the analysis to the four RSB actuators on each shuttle vehicle to estimate the 

system reliability. 

Nomenclature 

a major semiaxis of contact ellipse, m (in.) 

a1 life adjustment factor for reliability 

a2 life adjustment factor for materials and processing 

a3 life adjustment factor for operating conditions including lubrication 

B gear material constant, N/m1.979 (lbf/in.1.979) 

CD basic dynamic capacity of a ball or roller bearing, N (lbf) 

Ct basic dynamic capacity of gear tooth, N (lbf) 
NASA/TM—2015-218846 3 

d diameter of rolling element, m (in.) 

Ft normal tooth load, N (lbf) 

f tooth face width, m (in.) 

fcm bearing geometry and material coefficient 

i number of rows of rolling elements 

J AGMA gear tooth form factor 

K AGMA bending stress adjustment factor 

k gear tooth stress cycles per input shaft revolution 

L life, hr, stress cycles, or revolutions 

Lβ characteristic life or life at which 63.2 percent of population fails, hr, stress cycles, or revolutions

L10 10-percent life or life at which 90 percent of a population survives, hr, stress cycles, or 

revolutions 

l length of stressed track, m (in.) or roller length, m (in.) 

m Weibull slope; exponent or gear tooth module, mm 

N number of gear teeth 

n life, stress cycles 

P diametral pitch, 1.0/in. 

Peq equivalent bearing load, m (in.) 

Pt normal tooth load, N (lbf) 

p load-life exponent 

R reliability for bending fracture, fractional percent

r pitch circle radius of gear, m (in.) 

S probability of survival, fractional percent or material strength MPa (ksi) 

V stressed volume, m3

 where V = alz 

Xn fraction of time spent at load-speed condition n

Z number of rolling elements per row 

z depth beneath the surface of maximum orthogonal or maximum shear stress, m (in.) 

α contact angle, deg 

10t L10 life of single gear tooth, stress cycles 

Ʌ lubricant film parameter 

ρ curvature sum, m1

 (in.1) 

σ gear tooth bending stress 

σ1, σ2 surface roughness of bodies 1 and 2, rms, m (in.) 

τ maximum orthogonal or maximum shear stress, MPa (ksi) 

φ gear pressure angle, deg 

Subscripts: 

1, 2 bodies 1 or 2; load-life condition 1, 2, etc. 

B bearing 

G gear 

g gear reliability 

i individual gear 
NASA/TM—2015-218846 4 

n body n or load-life condition n 

ref reference 

s, sys system 

t tooth 

Rudder/Speed Brake (RSB) Actuator System 

The Space Shuttle Rudder/Speed Brake (RSB) Actuator system is contained in the Space Shuttle 

Orbiter tail section shown in Figure 1. The actuator system comprises a power drive unit and four 

actuators. Each actuator contains two complete gear trains, designated left and right. The system controls 

the RSB panels through the four actuators that are driven by a single power drive unit. 

A schematic of an RSB actuator is shown in Figure 2. The actuators are designed to provide both 

rudder and speed brake functions by a split design in the vertical tail section shown in Figure 1 in which 

two input shafts enter the actuator. By rotating the shafts in the same direction, the panels are moved 

concurrently for the rudder function. When the input shafts rotate in opposite directions, the panels move 

apart for the speed brake function. Both functions can be used simultaneously. 

Figure 1.—Rudder/speed brake actuating components. 

Figure 2.—Schematic of space shuttle rudder/speed brake (RSB) actuator. 
NASA/TM—2015-218846 5 

The 1L (left) rudder/speed brake gearbox shown in Figure 2 (designated Actuator A in Fig. 1) is 

composed of a 19.75:1 offset compound spur reduction followed by a differential planetary with a 

24:1 speed ratio. The overall speed ratio is 474:1. The input shafts are shown in the lower right and left 

corners of the figure. A series of two spur reductions bring the input power to the shaft of a sun gear, 

which drives a nine-planet differential planetary with two ring gears, one of which is fixed. In this 

arrangement, the fixed ring gear is split into two parts that straddle the output ring gear. The sun gear has 

a small face width and drives the rightmost planet set that meshes with the fixed ring gear. It is placed 

near the center of the planet spools. The geometric properties of the gears in both the compound spur 

reduction and the differential planetary reduction are given in Table 1. 

In Table 1, the gear face widths were reduced to estimate the gear tooth contact stresses more closely 

since the analyses only deals with a constant stress across the gear face. In the planetary, the actual 

diametral pitch and pressure angles of the gears differ from the nominal values because the gears operate 

at non-standard center distances. 

The compound spur gear input stage has an offset configuration with a 155° angle between the two 

gear center-lines connecting its three shafts. The input shaft is supported by two symmetrically-placed 

cylindrical roller bearings. The intermediate shaft is supported by a quill of four; one-half inch wide rows 

of 0.125-in. diameter rollers. This quill is modeled as two half-inch long roller bearings at its ends with 

increased capacity. The sun gear shaft is supported by a single ball bearing in the plane of the spur output 

gear. An axial thrust bearing is placed between the sun gear and the nearest of the two planet support rings. 

The two support rings carry the radial loads applied to the nine planets by the fixed and output ring 

gears. Finally, there are two ball bearings placed between the output ring and the housing. These bearings 

carry the radial load associated with the hinge moment of the output arm. They also carry the axial preload applied to the rudder speed brake actuator by a rod through its center. 

The gears comprising the actuators were manufactured from case carburized AISI 9310 steel. Most of 

the rolling-element bearing rings were manufactured from AISI 9310 steel. Two bearing rings were made 

from through-hardened AISI 52100 steel. The rolling elements were made from AISI 52100 steel except 

the balls for the bearings that connect the fixed and movable ring gears were made from AMS 6491 steel. 

TABLE 1.—GEAR GEOMETRIC PROPERTIES IN COMPOUND SPUR REDUCTION 

AND DIFFERENTIAL PLANETARY REDUCTION

Gear No. of 

teeth 

Face width, 

mm (in.) 

Nominal tooth 

size,a

module-mm 

dia. pitch (1.0/in.) 

Nominal 

pressure angle,

deg. 

Working Tooth 

sizea, 

Module-mm 

dia. pitch 

(1.0/in.) 

Working 

Pressure 

angle, 

deg. 

Input (1) 16 10.85 (0.427) 1.814 

(14) 

25 1.814 

(14) 

25.0 

Intermediate (2) 69 10.85 (0.427) 1.814 

(14) 

25 1.814 

(14) 

25.0 

Intermediate (3) 19 23.75 (0935) 2.117 

(12) 

25 2.117 

(12) 

25.0 

Spur Output (4) 87 23.75 (0935) 2.117 

(12) 

25 2.117 

(12) 

25.0 

Sun (5) 54 3.18 (0.125) 2.117 

(12) 

25 2.169 

(11.708) 

27.839 

Planet on Sun (6a) 18 3.18 (0.125) 2.117 

(12) 

25 2.169 

(11.708) 

27.839 

Planet on fixed ring (6b) 18 52.20 (2.055) 2.117 

(12) 

25 2.169 

(11.708) 

27.839 

Fixed ring (7) 90 52.20 (2.055) 2.117 

(12) 

25 2.169 

(11.708) 

27.839 

Output planet (8) 18 37.34 (1.470) 2.54 

(10) 

25 2.479 

(10.245) 

21.797 

Output ring (9) 81 37.34 (1.470) 2.54 

(10) 

25 2.479 

(10.245) 

21.797 

a The tooth size standards in metric and in English units are different. The metric module is the pitch diameter divided by the number of 

teeth. The English Diametral Pitch is the number of teeth divided by the pitch diameter. 
NASA/TM—2015-218846 6 

TABLE 2.—SUMMARY OF ROLLING-ELEMENT BEARING LIVES FOR 

RUDDER SPEED BRAKE ACTUATOR 

Bearing no. 1a 1b 2 3a 3b 4 6a 6b 

Bearing type Roller Roller Roller Roller Roller Ball Ball Ball 

L10 life (hr) 647,000 647,000 15,823 4,875 4,875 54,973 1,509 1,089 

L0.1 life (hr) 9,747 9,747 238 73 73 828 23 16 

Min life (hr) 34,291 34,291 839 258 258 2,914 80 58 

Analytic Procedure 

For the contact stress analysis, the Lundberg-Palmgren bearing life theory (Ref. 9) was expanded to 

include gear-surface pitting for the actuator as a system. The mission spectrum of the Space Shuttle 

rudder/speed brake actuator was combined into equivalent effective hinge moment loads including an 

actuator input preload for the contact stress fatigue and tooth bending fatigue models. 

The reliabilities from the contact stress and tooth bending analyses were calculated for a series of 

estimated loads and load cycles for 100 missions of the shuttle. The loads considered were the output 

hinge moments on the gearbox. The analyses were performed for the applied loads and for a combination 

of the applied loads with a pre-load due to assembly of 50 lb-in. Once the reliabilities of the separate 

analyses were obtained for each loading condition, they were combined to determine an overall reliability 

for the gearbox for the analyzed loading conditions and cycle counts. This analysis was performed for the 

left half of the first actuator, with the assumption that the life and reliability of the right half were 

identical. The life and reliability for the four actuators on each Space Shuttle was assumed equal as well. 

The results were extended to the system of four full actuators on a single shuttle. 

The life and reliability assessment of the gears was performed in two stages: a contact stress fatigue 

model and a gear tooth bending fatigue model. The actuator manufacturer provided L10 life data for the 

actuator bearings, assuming a Weibull slope of 1.11. These bearing life data are shown in Table 2. The 

analyses do not consider the probable reduction in life and reliability caused by boundary lubrication, 

wear and grease degradation. 

The original intent of the Space Shuttle design was for a life of 100 flights with minimal maintenance. 

During the Return to Flight program, post 2003 Columbia disaster, this goal was initially reduced to 

20 flights. The reliability analysis for the gear tooth bending, gear tooth surface fatigue and bearing fatigue 

were calculated separately along with the total system reliability for 1, 12, 20, and 100 Space Shuttle flights. 

The reliability values did not consider other modes of failure, such as wear or lubricant degradation. 

Enabling Equations and Analysis 

TLIFE: Transmission Life and Reliability Modeling 

The method of probabilistic design was applied to the Space Shuttle Rudder Speed Brake (RSB) left 

half actuator gearbox number 1L (designated Actuator A in Fig. 1). A fatigue life and reliability 

assessment of the gearbox was performed in two stages: a contact stress fatigue life and reliability model 

and a gear tooth bending fatigue life and reliability model. The initial effort to determine the fatigue life 

and reliability of the 1L rudder speed brake actuator gearbox was focused on adapting the program TLIFE 

(Ref. 10) to analyze this gearbox. This program performs life and reliability analysis in three stages: 

1. It determines the loads and load cycles on the components of the gearbox based on the overall load 

and speed of the gearbox, 

2. It then determines the lives and reliabilities of the individual components based on these loads and 

load cycles, and 

3. It combines these lives and reliabilities to determine the life and reliability of the gearbox system. 
NASA/TM—2015-218846 7 

This program did not initially include an analysis for the differential planetary described above. It did 

have an analysis for the compound spur gear reduction that precedes the differential planetary. The 

analysis for the differential planetary based on the analysis of the preceding section was added as part of 

this effort. 

The loading for this analysis was obtained from two Shuttle Program Office tables. The first three 

columns of Table 3 are a list of shuttle events, the corresponding rudder speed brake applied moment and 

its duration in minutes. For a flight of 7.604 hr, 7.25 hr are for the ferry from the landing site to the launch 

site. In the original table, specific moments and times are given but no load cycle counts. 

The second table, Table 4 lists load ranges for given cycle counts. The total cycle count in this table is 

35,297,798. However, 91,500 of these cycles are for reversed loading (negative moments) at lower 

moment ranges. These loads are applied to the opposite sides of the teeth and do not add to the positive 

moment fatigue damage. Removing these cycles from the total gives 35,188,298 positive load cycles for 

this analysis. Multiplying this total by the percent times of duration gives the estimates of the cycles for 

each applied moment which are presented in the last column of Table 3. These loads and cycles are the 

basis of this analysis. The loading condition analyzed represents a median load for the conditions 

considered. It should be noted that a 57,827 N (13,000 lb) axial pre-load is included in the loading on the 

output ring support ball bearings. 

The two-parameter Weibull distribution function was used to model the statistical variations in life 

and strength for both models. For the contact stress analysis, the Lundberg-Palmgren bearing life theory 

(Ref. 9) has been expanded to include gear-surface pitting and the gearbox as a system (Ref. 11). The 

mission spectrum of the Space Shuttle rudder speed brake unit 1L gearbox has been combined into a 

single equivalent hinge moment effective load based on the contact stress fatigue model using the Linear 

Damage Rule also referred to as the Palmgren-Langer-Miner Rule (Refs. 12 to 13). This load was also 

used for the bending fatigue analysis.

TABLE 3.—EVENT DURATION

Event Moment, 

(N-m) lb-in. 

Duration,

min 

Cycles 

calculated 

Ferry (1,693) 15,000 420 32,392,514 

Ferry (23,953) 212,000 0.25 19,281 

Ascent (9,039) 80,000 0.2 15,425 

Ascent (5,197) 46,000 1.7 131,113 

Ascent (11,073) 98,000 0.1 7,713 

Descent (15,818) 140,000 33 2,545,126 

Descent (27,116) 240,000 1 77,125 

TABLE 4.—INITIAL APPLIED LOADING INFORMATION 

Load range, N-m (in.-lb) Cycles 

–50,814 to –39,545 (–450,000 to –350,000) 0 

–39,545 to –28,246 (–350,000 to –250,000) 0 

–28,246 to –16,948 (–250,000 to –150,000) 2,939 

–16,948 to –5,649 (–150,000 to –50,000) 88,561 

–5,649 to 5,649 (–50,000 to 50,000) 34,369,892 

5,649 to 11,298 (50,000 to 100,000) 20,768 

11,298 to 16,948 (100,000 to 150,000) 67,044 

16,948 to 22,597 (150,000 to 200,000) 439,135 

22,597 to 28,246 (200,000 to 250,000) 291,459 

28,246 to 33,895 (250,000 to 300,000) 0 

33,895 to 39,545 (300,000 to 350,000) 0 

39,545 to 45,194 (350,000 to 400,000) 0 

 45,194 (0  400,000) 0 

Total cycles 35,297,798 
NASA/TM—2015-218846 8 

Weibull Analysis 

In 1939, Weibull (Refs. 16 and 17) developed a method and equation for statistically evaluating the 

fracture strength of materials. He also applied the method and equation to fatigue data based upon small 

sample (population) sizes, where 

ln where 0 ; 0 1 1 lnln      

















  



 





 

 L S

L L

L L

m

S (1a) 

The location parameter Lμ is the time or life at or below which no failures are expected to occur or the 

will be 100 percent probability of survival. Equation (1a) is referred to as the three parameter Weibull 

equation relating life and probability of survival. If Lμ = 0, the expression is referred to as the twoparameter Weibull equation and is written as follows:

ln where 0 ; 0 1 1 lnln      













  



 







L S

L

L

m

S (1b) 

When plotting the ln ln [1/S] as the ordinate against the ln L as the abscissa, fatigue data are assumed 

to plot as a straight line (Fig. 3). The ordinate ln ln [1/S] is graduated in statistical percent of components 

failed or removed for cause as a function of ln L, the log of the time or cycles to failure. The tangent of 

the line is designated the Weibull slope m, which is indicative of the shape of the cumulative distribution 

or the amount of scatter of the data (Refs. 7 and 15). 

The method of using the Weibull distribution function for data analysis to determine component life 

and reliability was later developed and refined by Johnson (Ref. 17). 

Figure 3.—Weibull plot where (Weibull) Slope of tangent or line is 

m. Probability of survival S of 36.8 percent at which L = L or 

L/L = 1. 
NASA/TM—2015-218846 9 

Rolling-Element Bearing Life Analysis 

Lundberg and Palmgren (Refs. 9 and 18) extended the theoretical work of Weibull (Refs. 16 and17) 

and showed that the probability of survival S could be expressed as a power function of maximum 

orthogonal shear stress τ, life n, depth of maximum orthogonal shear stress z, and stressed volume V: 

V

z

n

S h

c m  ~ 1 ln (2) 

1 ~ 1 ln 



h

c m

z

n al

S (3) 

By substituting the bearing geometry and the Hertzian contact stresses for a given load into Equation (3), 

the bearing basic dynamic load capacity CD can be calculated (Ref. 9). The basic dynamic load capacity

CD is defined as the load that a bearing can carry for a life of one-million inner-race revolutions with a 

90-percent probability of survival (L10 life). Lundberg and Palmgren (Ref. 9) obtained the following 

additional relation: 

p

eq

D

P

C L 















10  (4) 

where Peq is the equivalent bearing load and p is the load-life exponent (Ref. 9). 

Formulas for the basic dynamic load ratings derived by Lundberg and Palmgren (Refs. 9 and 18) and 

incorporated in the ANSI/ABMA and ISO standards (Refs. 19 to 21) are as follows: 

Radial ball bearings with d  25.4 mm: 

 

0.7 2 / 3 1.8 C f i cos Z d D  cm  (5) 

Radial ball bearings with d  25.4 mm: 

 

0.7 2 / 3 1.4 C 3.647 f i cos Z d D  cm  (6) 

Radial roller bearings: 

 

7 / 9 3/ 4 29 / 27 C f i cos Z d D  cm   (7) 

Equation (4) can be modified using life factors based on reliability a1, materials and processing a2, 

and operating conditions such as lubrication a3 (Refs. 22 and 23) where 

L  a1a2a3L10 (8) 

For the boundary lubrication under which the rolling-element bearings in the actuator operate, the 

lubricant film parameter, Ʌ can be used as an indicator of rolling-element bearing performance and life. 

For Ʌ < 1, surface smearing or deformation, accompanied by wear will occur on the rolling surfaces and 

the factor a3 in Equation (8) can vary from 0.2 to 0.5 (Ref. 22). If the effect of boundary lubrication had 

not been considered by the actuator manufacturer, the bearing lives summarized in Table 2 would be as 

much as 80 percent less than those shown. 
NASA/TM—2015-218846 10 

Gear Life Analysis 

Between 1975 and 1981, Coy, Townsend, and Zaretsky (Refs. 24 to 26) published a series of papers 

developing a methodology for calculating the life of spur and helical gears based upon the LundbergPalmgren theory and methodology for rolling-element bearings. Coy, Townsend, and Zaretsky (Ref. 27)

reported that for AISI 9310 spur gears, the Weibull slope mG is 2.5. Based on Equation (2), for all gears 

except a planet gear, the gear life can be written as 

k

N

L t

G

G

m 101/

10

 



 (9) 

For a planet gear, the life is 

 

k

N

L

G G

t

G

t

G

G

m e m m 1/

10 10

1/

10 1 2

        (10) 

The L10 life of a single gear tooth can be written as 

G

t

p

t

t

F

C

a a 















10  2 3 (11) 

where 

0.907 1.165 0.093 C  Bf  l t (12) 

and 



 sin

1 1 1

1 2















  

r r

 (13) 

and t 10 is the L10 life in millions of stress cycles for one particular gear tooth. This number can be 

determined by using Equation (11), where Ct is the basic load capacity of the gear tooth; Ft is the normal 

tooth load; pG is the load-life exponent (usually taken as 4.3 for gears based on experimental data for AISI 

9310 steel); and a2 and a3 are life adjustment factors similar to that for rolling-element bearings. Life 

factors a2 for materials and processing are determined experimentally. The value for Ct can be determined 

by using Equation (12), where B is a material constant that is based on experimental data and is 

approximately equal to 1.39108

 when calculating Ct in SI units (Newtons and meters) and 21 800 in 

English units (pounds and inches) for AISI 9310 steel spur gears; f is the tooth width; and ρ is the 

curvature sum at the start of single-tooth contact. 

The G L10 life of the gear (all teeth) in millions of output shaft load cycles at which 90 percent will 

survive can be determined from Equations (9) or (10) where N is the total number of teeth on the gear; mG

is the Weibull slope for the gear and was taken to be 2.5 (Ref. 28); and k is the number of load (stress) 

cycles on a gear tooth per output shaft load cycle. 
NASA/TM—2015-218846 11 

TABLE 5.—RUDDER SPEED BRAKE ACTUATOR PROBABILITY OF SURVIVAL 

Number 

of flights 

Number of 

actuators 

Gear reliability, percent Actuator 

bearing 

reliability,

percent 

Total system 

reliability, 

percent 

Tooth 

bending 

fatigue 

Tooth 

surface 

fatigue 

Combined 

bending 

and surface 

100 1/2 95.943 99.978 95.922 85.950 82.4 

100 4 71.793 99.822 71.666 29.782 21.3 

20 1/2 99.926 99.999+ 99.926 97.495 97.4 

20 4 99.409 99.997 99.406 81.632 81.1 

12 1/2 99.979 99.999+ 99.979 98.571 98.6 

12 4 99.835 99.999 99.834 89.126 89.0 

For all gears except the planet gears, each tooth will see load on only one side of its face for a given 

direction of input shaft rotation. However, each tooth on a planet gear will see contact on both sides of its 

face for a given direction of input shaft rotation. One side of its face will contact a tooth on the sun gear, 

and the other side of its face will contact a tooth on the ring gear. Equation (10) takes this into account, 

where 1 10t  is the L10 life in millions of stress cycles of a planet tooth meshing with the sun gear, and 

2 10t  is the L10 life in millions of stress cycles of a planet tooth meshing with the ring gear. 

Equations (9) to (13) are for gears in a single mesh only. For the case of collector gears such as 

parallel reduction, planetary gear trains and idler gears, the damage accumulates differently. As the load 

count will be different for these gears, the equations must be modified to account for this variable loading. 

In planets and idler gears each tooth is loaded on both faces in one rotation. Since the surface fatigue 

damage accumulates separately on the faces, the gear faces are treated as separate gears in their own mesh 

in this simulation. The load count factor, lc, is used in other cases. This has units of load cycles per output 

load cycle. 

For gears, this analysis used values for the Weibull slope, m = 2.5, gearing load-life exponent, p = 4 

and dynamic capacity surface strength, Sac, 500 ksi. Based on a mission duration of 7.604 hr, the 

reliability (probability of survival) of the gears based on contact (surface) fatigue is given in Table 5 for 

12, 20, and 100 shuttle flights. 

Gear Bending Life Analysis 

Gear tooth bending stresses are calculated using the AGMA bending stress adaptation of the Lewis 

bending stress calculation (Ref. 29). This adaptation included the Dolan and Broghammer stress 

concentration factor (Ref. 30). The maximum bending stress on each tooth is used for the life and 

reliability calculations. For both the pinion and the gear, this stress, σ, is the bending stress at the root of 

the tooth caused by the full mesh load Ft, applied at the highest point of single tooth contact on that tooth. 

  K  ksi

fJ

F P K MPa

fmJ

Ft t    (14) 

The first part of the equation is the metric form, with the tooth size in Equation (14) defined by the 

gear tooth module, m. The second part of the equation is the English version with the tooth size defined 

by the diametral pitch, P. The symbol J is the AGMA tooth form factor and the symbol K is for the 

AGMA stress adjustment factors. 

This stress is compared to the maximum allowable corresponding fatigue strength, S, in MPa (ksi) for 

AISI 9310 carburized gear steel (Ref. 31). Using a Goodman diagram, one can determine the 

corresponding alternate strength Se, where the ultimate strength, Su is 1889 MPa (274 ksi) where: 

u e S S S

1 0.5 0.5   (15) 
NASA/TM—2015-218846 12 

It is very unusual to have the full alternating strength available in an in-service device and since there 

is the presence of a corrosive environment after a period of time, this alternating strength will be de-rated 

to 80 percent of its full value. The reciprocal of Equation (15) including this de-rating factor gives the 

zero to maximum fatigue strength: 











752

1

1889

1

2

0.8

1 1

2

u e S S

S 1076 MPa (16a) 











109

1

274

1

2

0.8

1 1

2

u e S S

S 150 ksi (16b) 

Using Equations (14) to (16), a bending stress analysis of the gears can be performed to determine their 

life and reliability in terms of the maximum fatigue strength S, load-life factor p and Weibull slope m. 

The bending stress is directly proportional to the load. Accordingly, 

    p p

t L ~ 1/P ~ 1/S (17a) 

The load-life factor can be determined from the slope of the S-N curve in the region between 

one-thousand cycles to failure and one-million cycles to failure. For AISI 9310 carburized gear steel, 

S = 1076 MPa (156 ksi). Due to insufficient statistical tooth bending stress life data, a value of m = 2.5 is 

estimated for the bending stress-life Weibull slope. This is the same Weibull slope as used for the gear 

tooth contact-stress life slope. It is higher than that used for the bearing contact stress life slope. With the 

90-percent reliability fatigue strength equal to 0.9 Su at 103

 cycles and the endurance strength equal to S

at 106

 cycles, the load-life factor p becomes: 

       11.6

ln 0.9 1889 ln 1076

ln10 ln10

ln 0.9 ln

ln10 ln10

ln ln

ln ln 6 3 6 3

1 2

2 1  

      S S S SL L

p

u

 (17b) 

The maximum bending stress on each tooth is used for the life and reliability calculations. For both 

the pinion and the gear, this stress is the bending stress at the root of the tooth caused by the full mesh 

load at the highest point of single tooth contact on that tooth. The load-life relationship for bending 

fatigue is: 

p

g

S L 



 







10,  (18) 

where L10,g is the 90-percent reliability life of the gear for the applied stress, σ. In terms of this life, the 

gear reliability Rg for a given life, L, is given by: 

m

g L g

L

R 



















 



 















10, 0.9

1 ln 1 ln (19) 

 
NASA/TM—2015-218846 13 

With the life and reliability of the gears in a single mesh established, the next step is to combine the 

analyses for all the meshes. For this analysis, the system bending reliability is the product of all the 

individual bending reliabilities: 

Rs  Ri (20) 

The pinion torque for each gear mesh is determined as a ratio to the output hinge moment on the rudder 

speed brake. The number of teeth in engagement in each mesh for a single output ring gear tooth 

engagement is counted. 

The load cycles throughout the gearbox for every output ring motion of one tooth engagement, which 

equals 1/81 of a revolution or 4.444 were counted. Since there are nine planets, the fixed and movable 

ring gears each see nine tooth load cycles. Although each of the nine planets that mesh with the ring gears 

see only one load cycle, collectively they see nine. Each planet-sun mesh sees four load cycles for each 

planet-ring cycle (four teeth mesh at the sun for each tooth at the ring), thus the count for the sun with its 

planets is 36. 

The two meshes in the compound spur reduction are a little different. The spur-sun mesh sees 26 load 

cycles. However, the intermediate gear that meshes with this gear has only 19 teeth. Thus, seven of its 

teeth see two load cycles and 12 teeth see only one load cycle. The input mesh sees the most load cycles 

with 14 of the teeth on the input spur seeing six load cycles for the one tooth rotation of the output. 

The overall reliability for all the meshes taken together is the product of all the calculated reliabilities 

in the analysis. By calculating this analysis with different loads for one-million cycles, the dynamic 

capacity of the gearbox in bending is determined. It is the load that produces a reliability of 90 percent for 

one-million stress cycles. 

System Life Prediction 

The L10 lives of the individual bearings and gears that make up a rotating machine are calculated 

for each condition of their operating profiles. For each component, the resulting lives from each of the 

operating conditions are combined using the linear damage (Palmgren-Langer-Miner) rule (Refs. 12 to 

15) where 

n

n

L

X

L

X

L

X

L   

2

2

1

1 1

 (21) 

The Lundberg-Palmgren bearing life theory was expanded to include gear-surface pitting for the 

gearbox as a system. A second fatigue life and reliability analysis was also conducted for the bending 

fatigue lives of the gear teeth in the actuator. Gear and bearing system reliabilities were calculated and 

combined using strict series reliability where 

S S S S Sn ... 1 2 3     (22) 

At a specified reliability, the cumulative lives of each of the machine components are combined to 

determine the calculated machine system L10 life using the Lundberg-Palmgren formula (Ref. 9): 





















   



















   G

n

B G G

n

B B m

G

m

G

m

G

m

B

m

B

m

B

mLsys L L L L L L

1 1 1 1 1 1 1

1 2 1 2

  (23) 
NASA/TM—2015-218846 14 

Unfortunately, Equation (23) is only an approximation since the system Weibull slope varies with load. In 

a balanced life transmission, the system Weibull slope will be somewhere between the highest and the 

lowest of the components’ Weibull slopes. A form of this equation can be solved numerically for system 

reliability as a function of life and plotted on Weibull coordinates (Ref. 32). The resulting graph can be 

fitted with a straight line to determine the system Weibull slope and the system L10 life. In the event of an 

unbalanced life transmission, the lowest lived component will dominate the transmission failures and thus 

can serve as a good approximation for the system Weibull properties. However, at a given reliability, the 

system life will always be lower than the lowest lived component because other components can also fail. 

Based upon the Lundberg-Palmgren equation, the L10 life for the actuator as a system can be 

calculated where: 

p

T

C L 



 





10  (24) 

C is the theoretical load that produces a life of one million cycles (designated the dynamic capacity), T is 

the equivalent output torque and p is the load-life exponent. Using the Linear Damage Rule, where Xn is 

the time and Tn is the torque at condition n: 

n

p n n p p p

X X X X

X T X T X T X T

T

   

     ...

...

1 2 3

1 1 2 2 3 3 (25) 

A mission spectrum for the actuator as supplied by the NASA Shuttle Program Office was used to 

compute effective hinge moment loads for both gear tooth contact stress and tooth bending stress using 

the Palmgren-Langer-Miner linear damage rule. 

A gear contact stress fatigue model established the output hinge moment dynamic capacity for the 

actuator as 44 860 N-m (397 050 in-lb) with a Weibull slope of 2.5 and load-life exponent of 3.64. A 

similar analysis for tooth bending fatigue produced a dynamic capacity of 25 280 N-m (223 740 in-lb) 

with Weibull slope of 2.5 and load-life exponent of 11.6. 

A loading spectrum for 100 missions consisting of 35 188 292 load cycles was assumed for the 

gearbox. Output hinge moment loads varied from 1700 to 27 100 N-m (15 000 to 240 000 in-lb). An input 

preload of 5.6 Nm (50 in-lb) added 2010 Nm (17 800 in-lb) to the external hinge moment for the analysis. 

Results and Discussion 

Rudder/Speed Brake (RSB) Actuator Loading 

The Rudder/Speed Brake Actuator was designed and manufactured in the 1970’s for an expected life 

of 100 missions over a period of 10 years with no definitive plans for relubrication, maintenance and 

refurbishment. At that time and during the course of the operation of the Space Shuttle Fleet and until the 

time of the Columbia disaster, it was assumed by the Shuttle Program Office that no maintenance would 

be required of the actuators and/or that no failure should be anticipated. In other words, there would be a 

100 percent probability of survival for the designated 100 missions. As a result, design loads were not 

available for the analysis reported by us. 

The Shuttle Program Office did provided two simplified load versus time and load cycle tables shown 

in Tables 3 and 4. These tables give load ranges for an operation in output load cycles. It was assumed 

that the frequency of loading was constant for all missions and that the full cycle count of 35 188 298 

positive load cycles was the total for the 100 missions. On this basis, the cycles were totaled, each time 

was converted to a percent time and this percent was multiplied by the 35 188 298 cycles to obtain the 

number of cycles at each listed load. The average mission duration operating time was given as 7.604 hr. 
NASA/TM—2015-218846 15 

Rolling-Element Bearing Life 

The L10 lives of the bearings, (the rolling-element fatigue life at a 90-percent probability of survival) 

were calculated by the actuator manufacture for the left half of actuator number 2 and were not 

recalculated by us. Based on Equation (1a), for each of the bearings, the location parameter Lμ is the time 

below which no failures would be expected (100 percent probability of survival) where Lμ = 0.053L10

(Ref. 23). These results are shown in Table 2. However, as previously discussed, the analyses reported do 

not consider the negative effects of grease degradation, boundary lubrication and resultant wear on the life 

and reliability of the bearings and gears. 

The lowest lived bearing in the system dictates the time below which no bearing failure should occur. 

From Table 2 this time is 58 hr. Based on an average mission duration operating time of 7.604 hr and 

using the two-parameter Weibull equation (Eq. (1b)), the probability of survival for each bearing was 

calculated for 1, 12, 20, and 100 missions. The product of the reliabilities of the individual bearings 

represents the reliability of the system of bearings in one-half actuator. 

These results are shown in Table 5 for 12, 20, and 100 shuttle flights. Based on the results reported 

for bearing 6b (Table 2) having a life of 58 hr below which no failure should occur, it can be reasonably 

concluded that the bearings as a system will operate with no failure for about 8 missions (58 hr/7.604 hr 

per mission). 

Gear Contact Fatigue Life 

Gear tooth failures can be similar to failures of bearing elements. However, there are a few 

differences due to the complex shape of the gear tooth. Most differences result in sudden tooth breakage 

in a poorly lubricated, overloaded mesh. For our analysis we did not consider this as a failure mode. 

For the gear analysis the values for the Weibull slope, m = 2.5, gearing load-life exponent, p = 4 and 

dynamic capacity surface strength, Sac = 3450 MPa (500 ksi). Based on a mission duration of 7.604 hr, 

the reliability (probability of survival) of the gears based on contact (surface) fatigue is given in Table 5 

for 12, 20, and 100 shuttle flights. 

Gear Bending Fatigue Life 

A bending stress analysis of the gears was performed to determine their life and reliability in terms of 

the maximum fatigue strength S, load-life factor p and Weibull slope m that have been determined. These 

values are: S = 1076 MPa (156 ksi), p = 11.6 and m = 2.5, respectively. 

The maximum bending stress on each tooth was used for the life and reliability calculations. For both 

the pinion and the gear, this stress is the bending stress at the root of the tooth caused by the full mesh 

load at the highest point of single tooth contact on that tooth. 

The result of this analysis is a dynamic capacity of 25 279 N-m (223 741 in-lb). Coupled with the 

load-life factor of 11.6 and the Weibull slope of 2.5, this dynamic capacity enabled the bending stress life 

and reliability to be calculated along with the life and reliability for contact stress. The resulting 

reliabilities based on tooth bending fatigue are summarized in Table 5 for 12, 20, and 100 shuttle flights. 

Rudder/Speed Brake (RSB) System Reliability 

The reliabilities for the gears and bearings are summarized for a single half actuator (the 1L (left) 

rudder/speed brake gearbox shown in Fig. 2) and for four full actuators as a system in Table 5 for 12, 20 

and 100 flights. The bearing reliabilities are summarized under the column “Actual Bearing Reliability” 

in Table 5. The reliability of the gears as a system is summarized under “Gear Reliability for each failure 

mode in Table 5. The column showing “Combined Bending and Surface” reliability is the product of the 

reliabilities for “Tooth Bending Fatigue” and “Tooth Surface Fatigue.” The Total System Reliability 
NASA/TM—2015-218846 16 

shown under the column “Total System Reliability” in Table 5 is the product of the reliabilities shown in 

each line for the “Combined Bending and Surface Reliability” for the gears and the “Actuator Bearing 

Reliability.” 

There are four full actuators on each shuttle. The failure of a single actuator can result in the loss of a 

shuttle. When considering the probability of survival of four actuators together these numbers become 

89.0, 81.1 and 21.3 percent for 12, 20, and 100 flights, respectively. For 20 flights the probability of a 

gear failure in one of four actuators is less than 0.6 percent. However, for 100 flights the probability of a 

gear failure in any one of four actuators increases to 28.3 percent where bending fatigue failure becomes 

the predominant mode of gear failure. This is an unacceptable risk. 

The reliability of the bearings was determined separately from those of the gears. Combining the 

bearing and gear statistical analysis for 8 missions results in a 100 percent probability of survival for all 

four actuators as a system. The reliability of the gears in the actuators for both surface and bending 

fatigue is higher than the reliability of the bearings. This means the reliability of this system is dominated 

by the reliability of its bearings. 

The Weibull plots of Figure 4 illustrate this. They are for the sub-systems and full system of Rudder 

Speed Brake 1L. The gear surfaces sub-system, shown to the far right as a dashed line is by far the 

strongest with the highest reliabilities. The gear teeth bending sub-system is somewhat less reliable and 

weaker and dominates the gear system reliability. Finally, the bearings are the weakest system which 

dominates the full system reliability at high reliabilities or low percent of failures. 

These life and reliability data for a single half actuator were assumed to apply to all four full rudder 

speed brake actuators used in the Space Shuttle. Based on four RSB actuators per shuttle, the probability 

of survival or the actuators as a system is the probability of survival of a single half actuator at a 

designated number of flights to the 8th power or S8

. The Weibull plots of Figure 5 show the comparison of 

the reliabilities of the 1L gearbox and the full system of four gearboxes with their higher statistical 

percent of failures and lower reliabilities. 

The above analysis does not include wear as a failure mode. Wear, as fatigue, is probabilistic and not 

deterministic. We know that wear occurred in the RSB actuators. However, without an experimental data 

base and predefined acceptable wear criterion for critical components, wear as a failure mode could not 

be analytically predicted. 

Figure 4.—Rudder speed brake 1L actuator component and 

system Weibull plots. 

 Statistical percent of failures

95

20

0

12

RSB 1L

actuator

system

Gear

system

Bearing

system

Gear

teeth

bending Gear

surfaces

20 100

Number of flights

100 2000

5

2

1

10 1000

10

50

90
NASA/TM—2015-218846 17 

Figure 5.—System Weibull plot for 1L actuator and four full actuators. 

General Comments 

From Table 5, the 81.1 percent system reliability for 20 flights was deemed an unacceptable risk for 

the Space Shuttle. As a result, the number of future flights per actuator was limited to 12. The reliability 

for 12 flights on a single half actuator was calculated to be 98.6 percent. The reliability of the four full 

actuators as a system in a single Space Shuttle was calculated to be 89.0 percent for 12 flights. This 

means the probability of fatigue failure of the bearings or gears in the system of four full RSB actuators 

on a single shuttle is 11 percent over a life of 12 flights. 

Because of time constraints, a life and reliability analysis was not performed by us on the Space 

Shuttle Body Flap Actuator (BSA) gears and bearings. However, Oswald, et al. (Ref. 2), performed 

experiments on a test rig under simulated conditions to determine the life and failure mechanism of the 

grease lubricated Space Shuttle Body Flap Actuator (BFA) bearings that support the input shaft of the 

Space Shuttle body flap actuators. The Weibull analysis of the test data for the four Space Shuttle Body 

Flap Actuators (BFA) on one shuttle, each with a 2-bearing shaft assembly, established a reliability level 

of 96.9 percent for a life of 12 missions. For the purpose of our analysis, it was assumed by us that this 

reliability is that for the entire BFA assembly. If this reliability is for the entire BFA assembly, then the 

combined reliability with the RSB assembly for 12 missions on a single Space Shuttle was 86.2 percent 

[(0.969 × 0.890) × 100]. This yields a 13.8 percent probability that a bearing and/or gear failure will occur 

on either a BFA or RSB actuator. 

In the 1970’s when the Space Shuttle Rudder/Speed Brake (RSB) Actuator was designed and 

manufactured, a sophisticated gearbox life probabilistic life and reliability analysis of the type presented 

in this article was not available. The Rudder/Speed Brake Actuator was designed for a life of 100 

missions over a period of 10 years with no definitive plans for relubrication, maintenance and 

refurbishment. However, a bearing life and reliability analysis was performed by the manufacturer of the 

actuators. The bearing analysis performed by us and reported herein was an extension of the one 

performed by the manufacturer. In previous related studies (Refs. 1 to 6), are reported the effects of 

boundary lubrication, grease degradation and wear on body flap (BFA) and rudder speed brake (RSB) 

actuators’ life and reliability. For this fatigue study, we did not consider the negative effects of boundary 

lubrication, wear and grease degradation on bearing and gear life and reliability. Statistical percent of failures

95

20

1

12

Four full

actuator

systems RSB 1L

actuator

system

20 100

Number of flights

100 2000

5

2

1

10 1000

10

50

90
NASA/TM—2015-218846 18 

The risk of a rudder speed brake actuator bearing failure (not including the risk of gear failures) that 

could result in the loss of a Space Shuttle at 100 missions was 70.2 percent. For 20 missions that risk was 

reduced to 18.4 percent. However, it appears that the NASA Program Office did not examine or extend 

the bearing analysis to reflect the reliability and/or consider that an actuator bearing failure could be a 

probable cause for loss of a Space Shuttle. This omission was further compounded by a failure to provide 

for a scheduled maintenance program to remove, examine, repair, replace, and/or refurbish these actuators 

based on time and missions flown. 

Summary of Results 

A probabilistic life analysis was applied to the Space Shuttle Rudder/Speed Brake (RSB) Actuator 

1L. A contact stress fatigue model and a gear tooth bending fatigue model were used for a life and 

reliability assessment of the gears. Life and reliability of the bearings in the actuator were analyzed 

separately, based on data provided by the actuator manufacturer using the Lundberg-Palmgren life model. 

The life and reliability results for the gears and bearings in a single half actuator were combined using 

strict series reliability. The life and reliability of each of the four full actuators was assumed equal and the 

results extended to the four actuators on each shuttle as a single system. Although the analyses do not 

consider the probable reduction in life and reliability caused by boundary lubrication, wear and grease 

degradation, the recommendation to limit actuators to 12 flights before refurbishment addresses this 

concern. The following results were obtained. 

1. Based on the analysis, the Space Shuttle Rudder/Speed Brake Actuators were limited to 12 flights 

each in order to maintain a reliability of 98.6 percent for any half actuator and 89.0 percent for the 

system of four full actuators on one shuttle. 

2. The life and reliability of the actuator gears as a system for both surface and bending fatigue is 

higher than the life and reliability of the bearings as a system. Thus, the life and reliability of the 

actuator system is dominated by that of the bearings. 

3. Based on the original design requirement of the Space Shuttle, the rudder/speed brake actuator 

system, comprising four actuators on each shuttle, has a calculated reliability of 81.1 percent for a 

life of 20 flights and a reliability of 21.3 percent for a life of 100 flights. 

Epilog 

The body flap (BFA) actuators on all remaining space shuttles were limited to 12 flights each before 

refurbishment. The remaining two rudder speed brake (RSB) actuators were removed from the Space 

Shuttle Discovery. All four of the removed rudder speed brake (RSB) actuators were replaced with four 

actuators that had been subjected to three flights and had been removed from a sister space shuttle. On 

July 26, 2005 the Discovery returned to flight on a mission to the International Space Station (ISS), the 

first Space Shuttle to fly since the Space Shuttle Columbia disaster on February 1, 2003. This was the 

31st mission and flight of the Discovery. On July 4, 2006, the 32nd flight of the Discovery occurred on a 

mission to the ISS. This was the second consecutive return to flight since the Columbia disaster. The 

Discovery was flown seven more times to the ISS without incident for a total of 39 flights. After the 

39th mission on March 9, 2011 the Space Shuttle Discovery was retired from service and was placed on 

permanent display in the Smithsonian Air and Space Museum, Washington, D.C. The body flap (BFA) 

actuators and the rudder speed brake (RSB) actuators functioned as intended without incident. The rudder 

speed brake (RSB) actuators in the Discovery at the time of its retirement had a total of 12 flights, 

9 flights from those in the Discovery and 3 previous flights before being removed from the Space Shuttle 

Endeavor (Ref. 1). 
NASA/TM—2015-218846 19 

References 

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