r----r

NASA Contractor Report 178040

_NASA-CR-1780 40) EVA LUATICN 0_ A DOUBLE

GIMBAL IPACS DESIGN _Allied _endix

A£rospace) 28 _ _C A03/_F A01 CSCL I0C

G3/20

N86-247_5

Unclas

42919

Evaluation of a

Double Gimbal IPACS Design

P. R. Burke and P. A. Coronato

Allied Signal Bendix Aerospace

Teterboro, New Jersey 07608

Contract NASI-17658

May 1986

N SA

National Aeronautics and

Space Administration

Langley Research Center

Hampton, Virginia 23665
Table of Contents

Section Page

1.0 Introduction 1

2.0

2.1

2.2

2.3

2.4

Rotor Material Trade Study

Rotor Material Qualifications

Candidate Materials

Parametric Rotor Study

Rotor Material Trade Study Results

2

2

2

3

5

3.0

3.1

3.2

3.3

System Response Analysis

IPACS Spring/Mass Model

Launch Environment

Launch Loads

8

8

12

13

4.0

4.1

4.2

4.3

Belleville Washer Preload Mechanism

Belleville Washer Advantages

Preload Mechanism

Impact on System

15

15

16

21

5.0 Conclusions 22

6.0 References 24
List of Figures

Description

I. Spring-Mass Model of IPACS

2. Frequency Response for IPACS Rotor

3. STS Launch Vibration Environment

4. Load versus Deflection for the IPACS

Belleville Spring

Page

9

I0

12

17

ii
List of Tables

Description Paqe

II.

III.

V,

VI.

VII.

VIII.

IX.

I. Mechanical and Physical Properties of the

Three Candidate IPACS Rotor Materials

Results of a Parametric IPACS Rotor Sizing

Rotor and System Comparisons for the Three

Candidate Rotor Materials

IV. Spring and Mass Values for the System

Response Analysis

System Response Analysis Results

Margins of Safety Against Fatigue Strength

Margins of Safety Against Yield Strength

Belleville Spring Parameters

Physical and Mechanical Properties of the

Belleville Washer Material

X. Belleville Washer Load Cases

XI. Spin Bearing Preloads for the IPACS

II

13

14

14

16

18

19

20

iii
1.0 Introduction

Integrated Power/Attitude Control Systems, (IPACS), were

investigated in the early 1970's to determine if the dual

functions of electrical power storage and spacecraft

attitude control could be efficiently integrated into a

single package. An IPACS inner gimbal assembly (IGA) was

designed and built by Rockwell International and described

in NASA CR-172317 (reference i). A gimbal sizing analysis

was performed on this IPACS by Allied Bendix Aerospace and

documented in NASA CR-172524 (reference 2). This effort

reviewed the IPACS IGA design and produced a preliminary

design for the gimballing assembly.

The scope of the task performed by the Allied Bendix

Aerospace Corporation and described herein progresses from

the results of reference 2. A parametric study analyzing

the suitability of the IPACS rotor material was completed.

This study investigated three materials: 6AI-4V titanium

(current IPACS rotor material), BI20 VCA titanium, and

Custom 455 stainless steel. The preliminary linear

vibration response analysis was updated to include the

stiffnesses and the weights of the gimbals designed in Phase

I (reference 2). Finally, a belleville washer spring

preload mechanism was designed to replace the existing

helical spring and interference fit preload mechanism.
2.0 Rotor Material Trade Study

2.1 Rotor Material Qualifications

A trade study has been performed on three different metals

to investigate their suitability as the IPACS rotor

material. Four main criteria were applied to initially

restrict the material choices. Primarily, the candidate

material had to have a very high strength to density ratio,

excellent toughness, high resistance to stress corrosion,

and possess an extensive history of use for life prediction

security.

These criteria are of critical importance due to the unique

demands placed upon the rotor of an IPACS system. A high

strength to weight density ratio is crucial to the

maximization of the power storage capability of the IPACS.

However, strength without toughness is unacceptable for a

high speed flywheel and even more so in an IPACS

configuration. The rotor must live thru thousands of energy

storage/discharge cycles each of which applies a stress

whose magnitude is a considerable fraction of the yield

strength of the material. Simultaneously, the rotor must

endure the fatigue cycling that occurs when the rotor is

torqued about a gimbal axis. The alternating stress in this

case is rather small, but it is reacted at the rotor spin

rate of up to 450 cycles per second (see reference 2 for the

analytical rational for this maximum speed). Therefore, the

material must have high toughness against both low cycle

crack growth and high cycle fatigue failure. Stress

corrosion is an incipient danger that has been identified in

many high strength alloys being used in the aerospace

industry. Because of the continuously high levels of stress

present in energy storage and control moment gyro (CMG)

rotors, a very high level of stress corrosion resistance

must be maintained. Reference 3 is the standard document

for stress corrosion suitability. Finally, the importance

of an extensive database of mechanical and physical

properties should not be underrated. The hardware

development of any concept must be based upon a solid

foundation of analysis and experience lest the construction

of a prototype degrade into a very expensive form of

destructive testing.

2.2 Candidate Materials

Most high strength metals are quite brittle and have low

fracture and fatigue tolerances. The requirement for high
toughness eliminated a considerable bulk of the ultra high

strength alloys currently available. Lack of availability

and extremely high cost eliminated the cobalt based

superalloys. Maraging steels are unacceptable due to their

tendency towards stress corrosion. Plastic composites,

metal matrix alloys and ceramics were eliminated at this

time due to the lack of substantial property databases which

are necessary to insure the performance of this highly

stressed component.

Throughout this process of elimination, three alloys stood

out and were chosen as candidate materials for the trade

study. These alloys were: 6Ai-4V-Titanium alloy, BI20 VCA

Titanium alloy, and Custom 455 Stainless Steel alloy. All

of these alloys have very good strength to density ratios

while retaining excellent fracture toughness and fatigue

strength. They are all very resistant to stress corrosion

and have extensive backgrounds of testing data and actual

hardware implementation. Bendix has been fabricating high

speed rotors from Custom 455 Stainless Steel since 1979.

Important mechanical and physical properties of the three

alloys are presented in Table I (see references 4 and 5).

2.3 Parametric Rotor Study

The parametric rotor study that was performed on the three

candidate materials attempted to determine which would yield

the highest IPACS power to weight density. The basic

configuration of the original IPACS rotor was retained and

modification of the design was limited to scaling all of the

coordinate axes simultaneously by the same amount. In this

way, the basic rotor shape was preserved and, depending upon

the material, the rotor either shrank or grew but kept the

same proportions. References 1 and 2 contain more indepth

information on the IPACS rotor configuration, and the system

in general. It was necessary to modify the size of the

wheel so that the highest possible rotor speed would develop

the same amount of energy as the original IPACS rotor. By

way of explanation, a Custom 455 stainless steel IPACS rotor

can be made much smaller than a titanium wheel of the same

energy storage capacity because of the steel's higher

density.

The optimum rotor size and speed for each of the materials

was found through an iterative procedure. An estimate was

made as to the required size of the rotor, depending on the

strength of the material and its density. The size of the

rotor and the strength and density of the material

determined the centrifugal stress in the rotor and allowed a
Table I. Mechanical and Physical Properties of the Three Candidate IPACS Rotor Materials

Material

Modulus of Elasticity

Poisson's Ratio

Density

Elongation

Thermal Expansion

Custom 455

Stainless Steel

HT 1000

20.0 x 104 MPa

29.0 x l0 b psi

0.30

7860 kg/m 3 3

0.284 ibs/in

10%

10.6 x 10 -6 cm/cm/_K

5.9 x I0 6 in/in/ F

6A1-4V-Titanium

ii.0 x 104 MPa

16 0 x 106 psi

0.31

4430 kg/m 3 3

0.160 ibs/in

10%

8.8 x i0-6 cm/cm/_K

4.9 x i0 -b in/in/ F

BI20 VCA Titanium

i0.0 x 104 MPa

14.5 x l0 b psi

0.31

4820 kg/m 3

0.174 ibs/in8-10%

9.2 x 10 -6 cm/cm/AK o

5.1 x i0 -b in/in/UF

Ultimate Strength

(293_K)

(68_F)

(344°K)

(160UF)

Yield Strength

(293°K)

(68°F)

(344°K)

(160°F)

Fracture Toughness

1380 MPa

200 ksi

1325 MPa

192 ksi

1310 MPa

190 ksi

1255 MPa

182 ksi

5

II0 MPa*m06 5

I00 ksi*in "

895 MPa

130 ksi

835 MPa

121 ksi

825 MPa

120 ksi

750 MPa

109 ksi

77 MPa*m0l 5

70 ksi*in u'5

860 MPa

125 ksi

785 MPa

114 ksi

825 Mpa

120 ksi

750 MPa

109 ksi

66 MPa*m0655

60 ksi*in "
prediction of the maximum operating speed of the rotor. The

maximum speed prediction and the moment of inertia of the

rotor were used to calculate the energy storage capability

of that particular design. Depending upon the outcome, the

proposed rotor size was altered and the process repeated to

iterate the design to accomplish the task required. In this

way, it was guaranteed that each rotor would achieve the

same energy storage capability at its maximum operational

speed.

2.4 Rotor Material Trade Study Results

A tabular presentation of how the three rotors compare to

each other can be found in Table II. The active constraint

on the rotor designs was that the kinetic energy of each

rotor at its maximum operating speed would equal 1140

watt*hours. This value was chosen because it is the amount

of energy stored in the original IPACS rotor at its maximum

operational speed. The maximum operational speed of a rotor

is defined here as that speed at which the centrifugal

stress in the rotor e_uals the yield strength of the

material at 344° K (160- F).

The 6A1 4V titanium rotor has the greatest energy storage to

weight density, achieving 20.3 watt*hours per kilogram (9.19

watt*hours per pound) of material. While the energy storage

to weight densities of the titanium rotor is superior to the

stainless steel rotor, the stainless steel rotor is

significantly smaller in size. It is this size disparity

that will allow the steel rotor to achieve higher overall

system power to weight and power to volume densities.

The Custom 455 rotor is 16% smaller than the original IPACS

6AI 4V titanium rotor and is 5% heavier. But the reduction

in rotor size brings with it a corresponding reduction in

the size and weight of the inner gimbal, the outer gimbal

and the mounting ring. The gimbal pivots and torquers are

sized for stiffness and will not significantly change due to

this package size reduction. A 15% reduction in required

gimbal diameter will allow approximately a 25% reduction in

gimbal volume while maintaining the same structural

stiffnesses. When this factor is applied to the gimballing

structure as described in Section 3.0 and in references 1

and 2, then the weight reduction for the gimbals is:

5
Table II. Results of a Parametric IPACS Rotor Sizing

Material

Rotor Radius

Rotor Height

Rotor Weight

Rotor Inertia

Maximum Speed

Kinetic Energy @

Maximum Speed

Energy Storage to

Weight Density

Required Output 1

Torque

Momentum @ 1/2

Maximum Speed

Required Gimballing 2

Rate

Custom 455

Stainless Steel

HT i000

19.07 cm

7.51 in

35.56 cm

14.00 in

59.0 kg

130 ibs

2

0.586 N*m*sec 2

0.432 Ft*Ibs*sec

35700 rpm

1140 Watt*hrs

19.3 Watt*hrs/kg

8.77 Watt*hrs/ibs

6A1-4V-Titanium

22.71 cm

8.94 in

42.29 cm

16.65 in

56.2 kg

124 ibs

0.789 N*m*sec 2

0.582 Ft*ibs*sec 2

30800 rpm

1140 Watt*hrs

20.3 Watt*hrs/kg

9.19 Watt*hrs/ibs

BI20 VCA Titanium

22.71 cm

8.94 in

42.29 cm

16.65 in

61.2 kg

135 ibs

0.858 N*m*sec 2

0.633 Ft*lbs*sec 2

29500rpm

1140 Watt*hrs

18.6 Watt*hrs/kg

8.44 Watt*hrs/lbs

27 N*m 27 N*m 27 N*m

20 Ft*ibs 20 Ft*ibs 20 Ft*ibs

1270 N*m*sec

937 Ft*ibs*sec

1095 N*m*sec

808 Ft*ibs*sec

0.025 rad/sec 0.021 rad/sec

1325 N*m*sec

977 Ft*ibs*sec

0.020 rad/sec

I. Original IPACS performance requirement.

2. Gimballing rate needed to produce the required output torque at 1/2 maximum speed.
Inner Gimbal... 8.39 kg * 0.25 = 2.10 kg

Outer Gimbal... 10.4 kg * 0.25 = 2.61 kg

Mounting Ring... 9.07 kg * 0.25 = 2.27 kg

Total Reduction = 6.98 kg (15.4 lbs)

This reduction in gimbal weight would more than make up for

the 2.7 kg (6 ibs) increase in rotor weight. In this way,

an IPACS based upon a Custom 455 steel rotor would have an

overall greater energy storage to weight density than either

of the systems based on the titanium rotors. Additionally,

there would be a reduction in system volume of nearly 40%

below the titanium rotor based IPACS system. Table III

compares the system and rotor weights and volumes.

Table III. Weights and Volumes for the Three Rotor Materials

Rotor Rotor Weight System Weight

Material kg (lbs) kg (lbs)

System Vol_me

m (ft _ )

Custom 455 59.0 (130) 181 (400) 0.57 (20)

6AI-4V-Ti 56.2 (124) 186 (410) 0.91 (32)

S120 VCA Ti 61.2 (135) 191 (421) 0.91 (32)

There are several other advantages to using Custom 455

stainless steel over the titaniums for the IPACS rotor.

Machining a titanium rotor would be a laborious and

expensive process. When machining titanium, high

temperatures develop at the machine tool's edge due to the

low thermal conductivity of titanium. This localized

heating effect causes a plastic smearing of the metallic

surface instead of a cutting action and rapidly wears out

machining tools. The steel selected is considerably easier

to machine than titanium and would therefore significantly

reduce the manufacturing costs of the rotor assembly. The

reduction in gimbal size would also improve producibility

thru the ease of fabrication of a casting or in reduced

forging costs if hog-out fabrication is utilized.

7
3.0 System Response Analysis

The loading on any component of a structural system depends

upon the interaction of all of the components in the system.

Each component actively participates in the attenuation and/

or amplification of the loading environment as it propagates

through the system. For this reason, a system response

analysis was performed on the double gimballed IPACS package

to determine the maximum loading conditions in a launch

environment. This analysis and its results were discussed

in detail in reference 2.

In this Phase II effort, the weights and stiffnesses have

been updated and included in a reanalysis of the system

response. In the following sections, the linear spring-mass

model of the system will be described, the STS launch

vibration environment will be defined, and the launch loads

for each component along each axis will be presented.

3.1 IPACS Spring/Mass Model

A simple seven-degree-of-freedom model was used to describe

the double gimballed IPACS. The model is a single path

model in that there are no nodes that have more than two

elements connected to it (I mass, 1 spring). Figure 1

defines the masses and springs in the model and depicts

their connectivity. Table IV presents the mass values for

the model and the spring stiffneses along the inner gimbal,

outer gimbal and spin axes. The frequency response of the

rotor within the IPACS is given in figure 2.

The modifications made to this model which distinguish it

from the original model developed in reference 2 have to do

with the weights and stiffnesses of the gimballing

structures. Analysis of the inner gimbal (reference 2)

found that the stiffness of the structure was especially

high along the outer gimbal axis (OA). The weight of the

inner gimbal was known exactly since it exists as a piece of

hardware (reference i). The outer gimbal was generally

softer than had been expected, notwithstanding an increase

in weight from 9.07 to 10.4 kg (20 to 23 ibs). Substitution

of an aluminum metal matrix composite would increase the

structure's stiffness without altering any other properties

if it becomes advantageous to stiffen the gimbal.

Currently, the stiffnesses of the outer gimbal are

sufficient for a moderately tight control loop bandwidth.
Mass I = Rotor

pring 1 = Motor/Generator Housings

Mass 2 = Spin Motor Rotors + Rotor Shafts

+ Bearing Inner Races + Lube Systems

pring 2 = Spin Bearings

Mass 3 = I/2 Inner Gimbal + Motor Stators + Misc.

pring 3 = Inner Gimbal

Mass 4 = I/2 Inner Gimbal + I/2 Inner Pivots

Spring 4 = Inner Pivots

IMass 5

._pring •5

= I/2 Outer Gimbal + I/2 Inner Pivots

= Outer Gimbal

Mass 6 = I/2 Outer Gimbal + I/2 Outer Pivots

Spring 6 = Outer Pivots

ISass 7 I = 1/2 Mounting Frame + I/2 Outer Pivots

I

I

Spring 7 = Mounting Frame

//////////2//////

Figure I. IPACS Model for the System Response Analysis

9
0

Figure 2. IPACS Rotor Frequency Response

Normalized Deflection vs Frequency

1,E1-

tO 1.EOQ_

0

©

1.E-1-

1.E-2 -

I.E-3 i J t t

I.E-I 3.E-I I.EO 3.EO I.EI

! i 1 l I I

3.E1 1,E2 3.E2 1.E3 3.E3 1.E4

Frequency in Hz
The mounting ring structure was very stiff primarily because

the bolt down points were positioned quite close to the

pivots. Redesign and analysis of the mounting ring cut the

weight from 11.3 to 9.07 kg (25 to 20 ibs) while reducing

the stiffnesses only slightly. These high stiffnesses do

not pose any problems to the IPACS. In fact, they simplify

the development of a tight gimbal control loop. The penalty

assessed for this gain is usually an increase in weight for

the gimbals and an increase in loading throughout the system

during launch vibration. In this case, the position of the

mounting bolts creates a very stiff structure. Furthermore,

the loads throughout the IPACS during launch vibration are

still comfortably within the capabilities of the double

gimballed IPACS design as will be shown in the System

Response Launch Loads section.

Table IV. Spring and Mass Values for the System

Response Model

Body Mass Stiffness

Along IA Along OA

08 N/m I0_ N/m

_, , kg2.. 106 (XDS sec /xn) 1 ibs/in l0 b lbs/in

1 54.0 57.4 57.4

(0.308) (32.8) (32.8)

2 4.54 1.60 1.60

(0.0259) (0.92) (0.92)

3 14.33 3.24 4.52

(0.0818) (1.85) (2.58)

4 20.0 1.75 5.25

(0.114) (i.0) (3.0)

5 21.0 0.66 0.70

(0.120) (0.38) (0.40)

6 21.0 5.25 1.75

(0.120) (3.0) (I.0)

7 20.3 75.3 43.8

(0.116) (43.0) (25.0)

Along SA

i0_ N/m

l0 b ibs/in

6.46

(3.69)

0.27

(0.15)

2.07

(1.18)

5.25

(3.O)

0.58

(0.33)

5.25

(3.0)

158.0

(90.0)

11
ORIG_AL F_Q_ _$

OF POOR QUALtTY

3.2 Launch Environment

Three launch environments were applied to the IPACS springmass model: I) the original Skylab launch at 4.6 g's rms;

2) the final Skylab launch at 5.2 g's rms; and 3) an STS

launch environment of 9.6 g's rms random vibration. This

final environment was that which caused the highest

component loads throughout the IPACS.

On Earth testing and in space operation impose relatively

benign loading conditions on the outer gimbal, mounting ring

and pivot assemblies. Qualification level vibration testing

on Earth and the actual vehicle launch vibration are the

most severe environments to which these components will be

subjected. The spectral density plot of the STS

qualification vibration environment is presented in

figure 3. The precessional loads generated at 0.02 radians

per second are insignificant relative to the launch or

qualification loads.

N

¢q

.4

C

.u

U

m

0

,-.4

oJ

<

0.50

0.10

0.05C

0.010

!-t ttt, tt*t

_N: I "il-

_r

r_

i ; ,:;;;;:

20 100 200 300

III

ill

!il

ttl

:[

iii

500

I IIi_I-I'_H UIII]I

ltS_IIlt [I!t!ii

:F_ ! :!!_! .......

•f-+_!:iI!M HHIII

{t !:t11I!Ill!lifi

:-i_:t .l.x._ .2'

iiii!ii! .....

f-_ tt.! ! [: iM!!i_i

- __:l::t._-w -t

I I L LiLLil

I

1000 2000

Frequency (Hz)

Figure 3. Spectral Density Plot of the Qualification Level

Random Vibration Environment based on STS Launch

12
3.3 Launch Loads

The system component loads due to qualification random input

as predicted by the analysis are presented in Table V.

These loads are slightly higher than those presented in the

original analysis (reference 2) but are still comfortably

within the capabilities of the current design.

Table V. System Response Analysis Results

Maximum Design Loads (I_)

Body Along IA Along OA Along SA

N (ibs) N (ibs) N (ibs)

1 6180 6225 5735

(1390) (1400) (1290)

2 6670 6760 6625

(1500) (1520) (1400)

3 7780 7825 6445

(1750) (1760) (1450)

4 9245 9290 7290

(2080) (2090) (1640)

5 10450 10890 8670

(2350) (2450) (1950)

6 10980 11335 9510

(2470) (2550) (2140)

IIII0 11425 9910

(2500) (2570) (2230)

On the average, loads along the inner gimbal axis (IA)

decreased by 2%, loads along the outer gimbal axis (OA)

increased by 14% and the loads along the spin axis (SA)

increased by 2%. The magnification of the loads along the

OA axis is largely the result of the high stiffnesses of the

inner gimbal and mounting ring structures. The gimbals are

designed to be stiff in order to enhance the control loop

bandwidth of the gimballing system. These high stiffnesses

also have the effect of increasing the loads throughout the

system. However, these loads are still small compared to

the capability of the structures within the system and no

problem is foreseen with the stresses arising from the

launch vibration environment.

13
Table VI presents the margins of safety that the 16 launch

loads have against the components' endurance strength. The

lowest margin of safety for fatigue is 1.0 which indicates

that the high cycle fatigue loading is about one-half of the

endurance strength of the component.

Table VI. Margins of Safety Against Fatigue Strength

Margin of Safety = (Endurance limit/l_ Launch Load)-l.0

Body Along IA Along OA Along SA

Rotor 185. 185. 13.

Inner Gimbal 2.6 2.0 2.2

Inner Pivots 1.4 1.9 2.6

Outer Gimbal 6.3 8.8 4.7

Outer Pivots 1.4 Io0 1.8

Mounting Ring 66. 30. 93.

Table VII presents the margins of safety that the 3_ launch

loads have against the components' yield strength. The

lowest margin of safety is 0.3 which indicates that the

greatest single load that can occur during vibration is

still 30% below the capability of the component.

Table VII. Margins of Safety Against Yield Strength

Margin of Safety = (Yield Strength/3_ Launch Load)-l.0

Body Along IA Along OA Along SA

Rotor iii. Iii. 7.3

Inner Gimbal 1.7 1.2 1.3

Inner Pivots 0.6 1.2 1.8

Outer Gimbal 4.0 5.7 3.0

Outer Pivots 0.8 0.3 I.I

Mounting Ring 46. 20. 64.

14
4.0 The Belleville Washer Preload System

Currently, the IPACS rotating assembly's bearings are

preloaded with a combination of interference fits at the

bearings and a large helical spring. The interference fits

on the bearings are on both the inner and outer races. At

operating temperatures these interference fits are reduced

to nearly zero, and consequently so is the preload from this

reaction (see reference 2). The large helical spring is

therefore the sole remaining preloading mechanism during

high speed operation. In order to keep the bearing

preloaded during gimballing operation of the IPACS, the

force produced by the helical spring would have to be 245 N

(55 ibs), (see reference 2). This is a fairly high preload

for a high speed energy storage device, resulting in a

considerable power loss. Furthermore, an increased preload

generates higher bearing temperatures and reduces the

fatigue life of the bearing.

An improvement over this type of spin bearing preload system

would be the use of belleville washer constant force

springs. These small conical springs can achieve high

spring rates in extremely small packages, allowing the use

of lower preload forces than the helical spring. In the

following sections the advantages of using belleville

washers to preload spin bearings will be enumerated; a

preload system utilizing belleville washers and designed to

be compatible with the existing IPACS rotating assembly will

be presented; and the resulting improvements upon bearing

reliability, assembly power losses and system torque

bandwidth will be discussed.

4.1 Belleville Washer Advantages

A high performance rotating assembly like an IPACS or a CMG

is a very interactive device in that the alteration of any

component can produce a multitude of reactions throughout

the system. The spin bearing preload mechanism is one of

the most sensitive components in this regard. For this

reason, optimization of the preload mechanism will often

result in significant improvements in the performance of

many of the other components in the assembly.

The use of belleville washers in a preload mechanism allows

the utilization of very low preload forces on the spin

bearings. The impact of this preload reduction throughout

the IPACS system is far reaching. Reducing the spin bearing

15
preload directly produces a reduction in spin bearing drag

torque power losses, and thereby, a reduction in the spin

bearing operating temperature. Decreasing the spin bearing

drag effectively increases the efficiency of the

charge/discharge cycle of the IPACS, the spin bearings

being a significant source of power dissipation. This

reduction of drag torque would then allow a reduction in

motor power to achieve the same run up and run down

performance characteristics. This would in turn cause a

reduction in the operating temperature for the motors and

coupled with the reduced bearing temperatures would result

in a lower flywheel temperature. Because of the inverse

relationship between temperature and the flywheel material's

strength, lowering the flywheel's temperature increases its

strength and would allow a greater top speed for the rotor

and consequently greater power storage.

4.2 Preload Mechanism for the IPACS

4.2.1 Spring Parameters

Parameters of the belleville washer spring are given in

Table VIII. A load versus deflection graph for the spring

is shown in figure 4. Inspection of this graph shows that

the spring bottoms out (runs out of travel) before the

spring can snap thru. Snap thru is the action that occurs

in washer type springs when the differential stiffness of

the spring (dK/dx) changes from positive to negative.

Physically, the top of the washer deflects past the base and

turns inside out.

Table VIII. Belleville Spring Parameters

Outside Diameter:

Inside Diameter:

Free Height:

Thickness:

5.08 cm

3.81 cm

0.508 mm

0.508 mm

(2.00 inch)

(1.50 inch)

(0.020 inch)

(0.020 inch)

Stress at Flat Position: 335 MPa

Spring Load at Flat Position: 52.9 N

(48 ksi)

(11.9 ibs)

4.2.2 Spring Material

Type 302 Stainless Steel was chosen as the spring material.

It would be cold worked to about 40% which greatly increases

the strength of the material. The steel has a high

16
resistance to stress corrosion, can operate at high

temperatures and has excellent fatigue strength. Important

physical and mechanical properties for this stainless steel

are presented in Table IX.

53.4

44.5

35.6

O

26.7

©

17.8

8.9

Thousandths of an Inch

0 2 4 6 8 10 12 14 16 18 20

I--/ ! I

/

i [ i _ !

i

T

i

I

!

!

5 10 5 20 25 30 35 41

Hundredth of a Millimeter

46

12.0

I0.0

8.0

6.0

0

4.0

2.0

0

51

Figure 4. Load versus Deflection for the IPACS Belleville Spring

17
Table IX. Physical and Mechanical Properties of the

Belleville Washer Material

Material: Type 302 Stainless Steel, MIL-5-5059

40% Cold Rolled

Ultimate S_rength:

293VK = 1190 MPa

366°K = 1070 MPa

(60°F = 170 ksi)

(200°F = 153 ksi)

Yield Strength:

293°K = 875 MPa

366°K = 810 MPa

(60°F = 125 ksi)

(200°F = 116 ksi)

Density: 7915 kg/m 3 (0.286 ibs/in 3)

Modulus of Elasticity: 2.0x105 MPa (29xi06 psi)

Poisson's Ratio: 0.30

Fatigue Strength:

293-K = 510 MPa

366°K = 485 MPa

(60°F = 73 ksi)

(200°F 69 ksi)

Stress Corrosion Resistance: High

4.2.3 Belleville Spring Analysis

Analysis of the belleville washer preload system involved

the application of ten different possible loading

conditions. The variables that were considered included

vacuum shrinkage of the inner gimbal, thermal mismatches

between the rotor and inner gimbal, operation with the spin

axis both horizontal and vertical, and an axial pumping

force generated by the rotor at operational speed. These

loading cases are defined in Table X.

The short lengths of the rotor and inner gimbal, and the

fact that there are no huge variations in these components'

temperatures, produced thermal expansions that were

extremely small. These thermal deflections between the

titanium rotor and the aluminum inner gimbal were less than

5% of the other system deflections and were therefore

ignored.

The axial pumping force is generated by the rotation of the

rotor thru the spin bearing runout tolerances. This

oscillation produces a pumping force at the wheel speed

frequency. The natural frequency of the rotor/belleville

18
Table X. Belleville Washer Load Cases

Load Spin Rotor IG 2 Rotor

Case Description Axis Vacuum I Temp Temp Pumping 3

1 Assembly Vertical No RT RT No

2 Assembly Horizontal No RT RT No

3 Evacuation Vertical Yes RT RT No

4 Evacuation Horizontal Yes RT RT No

5 Run Up Vertical Yes 338°K 294°K Yes

150°F 70°F

6 Run Up Horizontal Yes 327°K 294°K Yes

130°F 70°F

7 Max Speed Vertical Yes 330°K 311°K Yes

135°F 100°F

8 Max Speed Horizontal Yes 327°K 311°K Yes

130°F 100°F

9 Run Down Vertical Yes 300°K 311°K Yes

80°F 100°F

i0 Run Down Horizontal Yes 294°K 311°K Yes

70°F 100°F

Notes: i. Vacuum of 2 microns pulled on inner gimbal

assembly = 1 atmosphere external gage pressure.

2. IG stands for Inner Gimbal.

3. Pumping force generated by rotor at speed due

to bearing runout.

4. Extreme temperatures taken for conservatism.

spring system is very low at approximately 8 to I0 hz, while

the IPACS will operate at roughly 300 to 400 hz. Because of

the great distance between the operating frequency and the

natural frequency of the system, there is no significant

dimensional movement of the rotor due to this axial pumping

force anywhere near the operating speed range of the IPACS.

The greatest dimensional changes occurred between the

vacuum/no vacuum conditions and the spin axis

vertical/horizontal conditions. These four loading

conditions primarily controlled the design of the belleville

spring.

Installation of the belleville springs would initially

specify a deflection of 0.305 mm (0.012 inch) for each

spring in a spin axis horizontal configuration. Moving the

IPACS to a vertical spin axis orientation would result in

the lower belleville spring compressing to its maximum

deflection of 0.508 mm (0.020 inch) while the top spring

19
would expand 0.203 mm (0.008 inch) for a resulting preload

deflection of 0.102 mm (0.004 inch).

When a vacuum is pulled on the inner gimbal (assuming there

is 1 atmosphere of pressure outside), the entire structure

shrinks. This shrinkage is approximately 0.203 mm (0.008

inch) and was computed from finite element analyses of the

inner gimbal (see reference 2). If the IPACS was in a spin

axis vertical configuration, this 0.203 mm (0.008 inch)

shrinkage would be applied to the top spring, increasing its

deflection to 0.305 mm (0.012 inch). Upon returning the

unit to a spin axis horizontal orientation, the springs

would equalize their deflections at 0.406 mm (0.016 inch)

apiece.

It should be noted that the horizontal, no vacuum condition

most simulates the situation that will exist in space. The

tabulation of the four loading cases, the spring deflections

for each, and the resulting spin bearing preloads for each

case are presented in Table XI.

Table XI. Spin Bearing Preloads for the IPA_S

Load Case i: Spin Axis Vertical

No Vacuum

Bottom Spring Flat (Snap thru not possible)

Bottom Spring Deflection= 0.508 mm (0.020 in)

Bottom Bearing Load= 578 N (130 ibs)

Top Spring Deflection = 0.102 mm (0.004 in)

Top Bearing Load= 18.2 N (4.1 ibs)

Load Case 2: Spin Axis Horizontal

No Vacuum

Spring Deflections and Bearing Loads Equal

Single Spring Deflection= 0.305 mm (0.012 in)

Single Bearing Load = 40.5 N (9.1 ibs)

Load Case 3: Spin Axis Vertical

IG Evacuated to 2 Microns Pressure

Bottom Spring Deflection= 0.508 mm (0.020 in)

Bottom Bearing Load= 600 N (135 ibs)

Top Spring Deflection = 0.305 mm (0.012 in)

Top Bearing Load = 40.5 N (9.1 ibs)

Load Case 4: Spin Axis Horizontal

IG Evacuated to 2 Microns Pressure

Spring Deflections and Bearing Loads Equal

Single Spring Deflection= 0.406 mm (0.016 in)

Single Bearing Load= 47.1 N (10.6 ibs)

20
4.3 Impact on System

The preload on the spin bearings peaks at 47.1 N (10.6 ibs)

during spin axis horizontal, on Earth testing. This is a

considerable improvement compared to the 245 N (55 ibs) that

would be necessary to preload the system with the original

helical spring preload mechanism. In space, the expansion

of the inner gimbal due to the absence of a pressure

differential across it allows the belleville washer preload

system to expand slightly and drops the spin bearing preload

to 40.5 N (9.1 ibs). The original preloading mechanism has

no such reaction; the change in preload is insignificant

due to the extremely small deflection combined with the low

stiffness of the helical spring.

A comparison between the spin bearing power levels can be

made for an operational load. For comparison, from

reference 2 the loading would be: the titanium rotor IPACS

rotating assembly running at 24000 rpm in space with either

the 245 N (55 ibs) preload from the original helical spring

preload mechanism, or 40.5 N (9.1 ibs) from the belleville

washer preload mechanism.

The drag torque power may be calculated from the component

viscous, thrust, and radial drag torques as was done in

reference 2. The original helical spring system would

develop a drag torque power of 25.5 watts dissipation per

bearing. The belleville washer system develops a drag

torque power of only 8.9 watts dissipation, a reduction of

65%! As a consequence, there would be a considerable

reduction in the motor power required to drive the IPACS

flywheel. Combined with the reduction in heat generated at

the bearings, there would be an overall lowering in the

assembly's operating temperature.

Reliability of the spin bearings would benefit because of

the reduced load and the cooler operating temperatures. The

reduction of preload from 245 N (55 ibs) would increase the

analytically derived survival rate from 99.84% to

above 99.99%.

Associated with the reduced thrust preload on the spin

bearings will be a reduced radial stiffness in the bearings.

However, this softening of the spin bearings will not

adversely affect the system torque output bandwidth. The

electronic control loop that is responsible for determining

the bandwidth would be designed to take into consideration

the different stiffness of the spin bearings. The bearings

would still be very stiff, and this would allow a wide

bandwidth control loop to be established.

21
5.0 Conclusions

A parametric study investigating three rotor materials for

the IPACS revealed the benefits of using Custom 455

stainless steel for the rotor as opposed to the high

strength titaniums. Although this steel cannot match the

titanium's strength to weight density, it is a better choice

for the rotor material because of its high density. Because

the steel is approximately 78% denser than the titanium, a

much smaller and only slightly heavier steel rotor can be

made having the same energy storage capability as the

titanium rotor. The advantage of the steel rotor is that it

is 16% smaller in every dimension than the titanium rotor

and the weight savings due to the smaller gimbals

surrounding the steel rotor more than makes up for the

weight difference between the steel and titanium rotors.

Not only would an IPACS with a steel rotor have a slightly

higher energy storage to weight density than a titanium

rotor IPACS, but the overall system volume would be 40%

smaller.

The system response analysis updated the existing linear

spring/mass model. Modifications to the model included

changing the stiffnesses and weights of the inner gimbal,

outer gimbal and mounting ring. These structures were

analyzed and the results were included in this system

analysis. A mild redesign effort was established for the

outer gimbal and mounting ring in order to lower their

weights without severely altering their stiffnesses. This

was accomplished and an STS launch vibration environment was

applied to the model. The loads on the components in the

system increased slightly due to the gimbal stiffnesses

being higher than were originally estimated. However, the

peak loads are still comfortably within the capabilities of

the system.

A belleville washer spring preload mechanism was designed to

be compatible with the existing IPACS IGA. This mechanism

was designed to replace the helical spring and interference

fit preload mechanism currently in use in the IPACS.

Preload reductions for the belleville washer system are

significant. The existing preload system would require

245 N (55 ibs) for an on Earth, gimballing operation. The

belleville washer spring preload system would require only

40.5 N (9.1 ibs). Drag torque power dissipation at the spin

bearings would also decrease substantially. Operation of

the IPACS in space with the existing preload design would

dissipate 25.5 watts per bearing. The belleville washer

preload system would develop a drag torque power of only 8.9

22
watts per bearing. Additional benefits include the

reduction of the operating temperatures for the spin

bearings and consequently an increase in reliability for

these components. Motor power levels could also be reduced

while maintaining the same run up and run down times.

Because the spin bearings and motors would be dissipating

much less power, the entire rotating assembly would be

operating at a much reduced temperature. The reduction in

operating temperature for the rotor would increase the

allowable strength value for the material and subsequently

increase the margins of safety for the IPACS.

Development of the IPACS concept has reached a point where

hardware implementation is the next logical step. The inner

gimbal assembly (IGA) can achieve 2.8 watt-hours per

kilogram (6.3 watt-hours per pound) energy density. As was

presented in this document significant improvements in the

performance of the IGA may be achieved by using Custom 455

stainless steel for the rotor, and substituting a belleville

washer constant force spring preload system to preload the

spin bearings instead of the helical spring/interference fit

method of preload currently being used.

Gimballing of the IPACS IGA is not difficult. The very

small required output torque does not generate any

significant loads in the system. For this reason, there is

a moderate amount of gimballing capability that is not being

used in this system. A precession rate approximately five

times larger than the 0.02 radians per second currently

required for the IPACS would be within the operating

capability of the system. Although the advantage of the

IPACS system over separate energy storage and attitude

control devices becomes less pronounced as the capabilities

of the system are utilized less and less, there is still a

significant savings in weight and space in this application.

However, with low output torque requirements, as developed

in this IPACS, it may be useful to compare the double

gimballed IPACS concept with a system of high speed energy

storage flywheels configured as single degree of freedom

reaction wheels. Analysis has shown the advantage of

reaction wheel assemblies over control moment gyros in low

output torque configurations and it may also be true for

this application.

23
6.0 References

, Cormack,A. III; and Notti,J.E.: "Design Report for the

Rotating Assembly for an Integrated Power/

Attitude Control System", NASA CR-172317,

September 1974.

• Burke,P.R.; and Coronato,P.A.: "A Gimbal Sizing

Analysis for an IPACS Rotating Assembly", NASA

CR-172524, May 1985.

• MSFC-SPEC-522A, "Design Criteria for Controlling

Stress Corrosion Cracking", George C. Marshall

Space Flight Center, NASA, November 18, 1977.

• MIL-HDBK-5C, "Military Standardization Handbook -

Metallic Materials and Elements for Aerospace

Vehicle Structures", 1978.

• ASM Metals Handbook, American Society for Metals,

8th Edition, 1961.

24
I. Renarz No.

NASA CR_178040

4. Title and S_lxitle

2. Govwnment Acc_ NO.

Evaluation of a Double Gimbal IPACS Design

7. _U_(s)

P.R.Burke and P.A.Coronato

9. _fmm!_ O,_iUt_ _ and A,clclr_

Allied Signal Bendix Aerospace

Route 46

Teterboro, NJ 07608

12. S_moti_ Agency _ and Address

National Aeronautics and Space Administratioz

Washington, D.C. 20546

15. _pplementm.y Norm

3. Recimem's C_,og No.

S. Rel_ort Oata

May 1986

6. Pla'form,n90rganizazicm

8. Plcformir_! Orgln;zat,on Repot_ NO.

tO. Work Umt No.

11. Contract ot Grant NO.

HAS 1-17658

13. Tyl)e of Rel_rt and _iod Covered

Contractor _Dort.

14. S_;_,'_soting Ager_-y

506-46-I 1-03

Technical Monitor: Claude R. Keckler, NASA-Lancley Research Center

16. _lx_a_

A parametric study analyzing the suitability of various

IPACS rotor materials was conducted. This study

investigated three materials: 6Ai-4V-Titanium (the current

IPACS rotor material), B120 VCA Titanium, and Custom 455

Stainless Steel. The preliminary linear vibration analysis

was updated to include the weights and stiffnesses of the

_imbals designed in NASA CR-172524. Finally, a belleville

washer spring Dreload mechanism was designed to replace the

existing helical spring and interference fit preload

mechanism.

17. K_ W_ (Suggesz_ _ Aut.l_(s))

IPACS,CMG, flywheel, energy

storage, attitude control,

gimballing, belleville washer

spring

19 _urltv Oa_f. (of t.is re_t) _ _. _itv Oamf. (of th,s _K_J

Unclassified I Unclassified

1-305

18. Oistributicm Statement

Unclassified- Unlimited

Subject Cate[orv 20

21. _. of Pagez 1 22. _,ce

28 J

F= _le by t_ Nat=o_l Techn,cal Infant,on Serv,ce. Sp;_nEf_eid VJrg,n=a 22161