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USAAVLABS ltr 14 Jul 1979
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\1
USAAVLABS TECHNICAL REPORT 67-46
DESIGN STUDY OF
HEAVY LIFT HELICOPTER
EXTERNAL LOAD HANDLING SYSTEM
ir
Lester R. Burroughs
Harold E. Ralston
D D C
Ifr- ■ . . T;
mi 3 mi
ag&gilj'x~
November 1967
U. S. ARMY AVIATION MATERIEL LABORATORIES
FORT EUSTIS, VIRGINIA
CONTRACT DA 44-177-AMC-467(T)
# SIKORSKY AIRCRAFT
DIVISION OF UNITED AIRCRAFT CORPORATION
STRATFORD, CONNECTICUT
*
This document is subject to special
export controls and each transmittal
til foreign governments or foreign
nationals may be made only u ith
prior approval of i'S Army Aviation
Materiel Laboratories, Fort F.ustis,
I irginia 23604.
Disclaimer s
The findings in this report are not to be construed as an official Depart¬
ment of the Army position unless so designated by other authorized
documents.
When Government drawings, specifications, or other data are used for
any purpose other than in connection with a definitely related Government
procurement operation, the United States Government thereby incurs no
responsibility nor any obligation whatsoever; and the fact that the Govern¬
ment may have formulated, furnished, or in any way supplied the said
drawings, specifications, or other data is not to be regarded by implica¬
tion or otherwise as in any manner licensing the holder or any other per¬
son or corporation, or conveying any rights or permission, to manufac¬
ture, use, or sell any patented invention that may in any way be related
thereto.
Trade names cited in this report do not constitute an official endorsement
or approval of the use of such commercial hardware or software.
Disposition Instructions
Destroy this report when no longer needed. Do not return it to the
originator.
DEPARTMENT OF THE ARMY
U 5 ARMY AVIATION MATERIEL LABORATORIES
FORT EUSTIS. VIRGINIA 23604
The basis for this design study was obtained from previous investi¬
gations of problems associated with the mechanics of cargo handling
by aerial-crane-type aircraft (USAAVIABS Technical Report 66-63).
This report is one of three contract studies of the same problem
with varying technical approaches. The conclusions drawn by this
contractor are based on sound analytical techniques. They are partic¬
ularly appropriate to a single-rotor, aerial-crane-type configuration.
In this context, this command concurs in general with these findings.
The preliminary designs developed by the contractor are complete,
accurate, and in sufficient detail to provide a basis for component
development programs.
Future work anticipated by this activity relative to this area
includes an analysis of the three preliminary contract designs so
as to define an optimum system based on the best features of each.
This may be followed by conponent development and test of critical
items as appropriate and the detail design, fabrication, and test of
an experimental system.
Task 1X130901D332
Contract DA 44-177-AMC-467(T)
USAAVLABS Technical Report 67-46
November 1967
DESIGN STUDY OF
HEAVY LIFT HELICOPTER
EXTERNAL LOAD HANDLING SYSTEM
Sikorsky Engineering Report 30441
By
Lester R. Burroughs
and
Harold E. Ralaten
Prepared by
Sikorsky Aircraft
Division of United Aircraft Corporation
Stratford, Connecticut
for
U. S. ARMY AVIATION MATERIEL LABORATORIES
FORT EUSTIS, VIRGINIA
*v
■tsji
This document Is subject to special export controls and
each transmittal to foreign governments or foreign
nationals may be made only with prior approval of US Army
Aviation Materiel Laboratories, Fort Eustls, Virginia 23604.
This report presents the results of a two-phase feasibility and prelimi¬
nary design study of load suspension configurations capable of meeting the
external cargo handling system requirements of a 40,000-pound-payload
heavy lift helicopter.
In Phase I, Design Analysis, both separate function configurations (those
that incorporate individual single- and multi-point hoists) and combined
function configurations (multi-point hoists used to perform both single-
and multi-point missions) have been investigated for two external load
handling system arrangements: single- plus two-point suspension and
single- plus four-point load suspension*
This phase was primarily concerned with the investigation of hoxst types;
the methods of power transmission (to the hoists); and the selection of
mechanical, hydraulic, and electrical components. A comparative evalua¬
tion of 13 system arrangements was made on the basis of weight, power, re¬
liability, in-flight safety, versatility, and productivity.
The single- plus two-point and single- plus four-point systems deter¬
mined to meet the heavy lift requirement best were presented to USAAVLABS,
and the latter system was recommended for the Phase II, Preliminary Design.
Upon receipt of approval, the preliminary design of a load suspension sy¬
stem incorporating a mechanically driven single-point hoist and four hy¬
draulically driven multi-point hoists was initiated. This phase of the
study included the preparation of preliminary (layouts) drawings, load and
stress analysis of all major components, and a maintainability and relia¬
bility analysis, as well as the preparation of a component development
plan.
The single- plus four-point system described herein weighs 4974 pounds for
a capacity of 40,000 pounds. The system has been designed such that the
hoists of both systems are readily removable when missions requiring mini¬
mum empty weight are to be undertaken. For single-point operation only
(four-point hoists removed), the system weighs 2738 pounds; for four-point
missions (single-point hoist removed) the system weighs 2704 pounds*
Both single- plus four-point systems have potential for growth to 50,000
pounds with a minimum of modification. This increased capacity can be
realized for a total weight increase of 270 pounds for both systems.
A typical sequence of operations for both single- and four-point systems
is outlined in Appendix IV,
FOREWORD
This report covers a two-phase evaluation of external cargo handling sy¬
stems for a 40,OOO-pound-payload heavy lift helicopter. This project was
conducted during the 10-month period from July 1, 1966, through April 28,
1967, for the U.S. Anqy Aviation Materiel Laboratories (USAAVLABS) vnder
Contract DA 44-177-AMC-467(T). Pertinent data upon which portions of thin
study were based were provided by the following: Bergen Wire Rope Company;
The lyconing Division of AVCO; Eastern Rotorcraft Corporation; Vickers
Incorporated, Division of Sperry Rand Corporation; Taylor Devices, Incor¬
porated; end Holex, Incorporated.
USAAVLABS technical direction was provided by Mr. J. Vichness, Chief, Air
Cargo Systems Branch.
The principal investigators for Sikorsky Aircraft were L. R. Burroughs,
Assistant Supervisor, Mechanical Design and Development Section, and H. E.
Ralston, Supervisor, Mechanical Accessories Group. Also making signifi¬
cant contributions to this effort were A. Korzun, Design Engineer, J. Kish,
Senior Design Analyst of the Mechanical Design Section, and R. Fidler,
Design Analyst of the Hydraulics Section.
TABLE 07 CONTENTS
l
I
f
r
Page
SUMMAHT
ill
FOREWORD
V
LIST OF ILLUSTRATIONS
X
LIST OF TABLES
xiv
LIST OF SIMBOLS
xvl
INTRODUCTION
1
PHASE I - DESIGN ANALYSIS
2
DISCUSSION
2
BASIC DATA
3
DESIGN REQUIREMENTS
3
MISSION RBQUIREMBIT8
3
AIRCRAFT DESCRIPTION
4
INVESTIGATION OF VEHICLES
9
HOIST SYSTEM AND COMPONENTS DESIGN
12
HOIST LOCATION AND TYPE
12
HOIST DESIGN
22
POWBt SOURCES
39
CLUTCH-REVHtSER UNIT
42
CONTROL SYSTEMS
42
ISOLATORS
49
HOIST CABLES
49
CARGO HOOKS
50
MISCELLANEOUS COMPONENTS
54
HOIST SYSTEM CONFIGURATIONS
60
LOAD ACQUISITION AND RELEASE
91
SINGLE-POINT MODE
91
TWO-POINT MODE
91
FOUR-POINT MODE
94
AIRCRAFT-LOAD INTHtACTION
96
STABILITY OF SLUNG LOADS
96
CENTBl-OF-GRAVITI SHIFT
99
TOWING CAPABILITY
99
AIRCRAFT CONTROLLABILITY
101
VffiTICAL OSCILLATION
107
POD JETTISON
107
▼li
I
wuTMti
IN-FLIGHT ADJUSTMENT OF MULTI-POINT HOISTS
111
PROBLEM AREAS AND PROPOSED SOLUTIONS
MECHANICAL LOAD RELEASE FROM COCKPIT
WEIGHT
SYNCHRONIZATION OF KULfI-POINT HOISTS
113
113
114
115
COMPARATIVE RELIABILITY AND MAINTAINABILITY ANALYSIS 116
EVALUATION PROCEDURES
INTRODUCTION
DISCUSSION
DESIGN OBJECTIVES
QUALITATIVE EVALUATION
SUMMARY
118
118
118
120
127
131
PHASE II - PRELIMINARY DESIGN 134
DISCUSSION 134
SINGLE-POINT HOIST SYSTEM 136
HOIST 136
CLUTCH-REVERSES UNIT 137
CABLE 139
MISCELLANEOUS COMPONENTS 140
CONTROLS AND INDICATOR SYSTEM 141
FOUR-POINT HOIST SYSTEM 143
HOIST 143
HYDRAULIC SYSTEM 144
ERROR ANALYSIS 148
niRT.it 149
MISCELLANEOUS COMPONENTS 151
CONTROLS AND INDICATOR SYSTEM 153
LOAD AND STRESS ANALYSIS 157
INTRODUCTION 157
DESIGN CRITERIA 157
SINGLE-POINT HOIST 159
SINGLE-POINT HOIST DRIVE SYSTEM 191
FOUR-POINT HOIST 197
WEIGHT ANALYSIS
MAINTAINABILITY AID RELIABILITY
INTRODUCTION
P1CT.TAR TT.TTT AND MAINTAINABILITY CHARACTERISTICS
FAILURE MODE AND EFFECT ANALYSIS 218
SAFETY CONSIDERATIONS 218
INSTALLATION AND REMOVAL 220
GROWTH POTENTIAL 225
215
218
218
*
vlll
SINGIE-POINT HOIST SYSTEM 225
FOUR-POINT HOIST SYSTEM 226
COMPONENT AND SYSTEM DEVELOPMENT PLAN 228
DISCUSSION 228
TEST PROGRAM 228
ESTIMATED DEVELOPMENT COSTS 230
SCHHjULE 231
CONCLUSIONS 232
BIBLIOGRAPHY 234
APPENDIXES
I SURVEY OF MILITARY VEHICLES 236
II MECHANICAL VARIABLE SPED) DRIVE 245
III CARGO HANDLING SYSTEM DRAWINGS 251
IV TYPICAL SEQUENCE OF OPERATIONS 267
DISTRIBUTION
271
ILLUSTRATIONS
Fluor* Pjgg
1
Single-Rotor H.L.H.
5
2
Tandem-Rotor H.L.H.
7
3
Four-Paint Hoist Load Attachment
Point Spaco Envelope
13
4
Conventional Design - Single-Point Hoist
15
5
Two-Part, Double-Reeved Hoist
19
6
Zero-Moment Hoist
23
7
General Gear Arrangement - One Spur and
Three Planetary
33
8
General Gear Arrangement - Two Spur and
Two Planetary
34
9
General Gear Arrangement - Compound Planetary
35
10
General Arrangement - Conventional and
Compound Planetary
36
11
Hydraulic Schematic - Single- Plus Two-Point
Suepension
44
12
Multi-Point Hoist - Equalized Load System,
Cargo Attitude vs C.G. Location
46
13
Hydraulic Schematic - Single- Plus Four-Point
Suspension
48
14
CH-54A Book - Swivel Assembly
51
15
Traction Sheave
55
16
Conductor Reel
57
17
-1 Configuration
61
18
-2 Configuration
63
19
-3 Configuration
65
20
-4 Configuration
67
x
t:
I
f
i
l
\
i
t
t
0
I
i
i
Figure
Page
21
-5 Configuration
69
22
-6 Configuration
71
23
-7 Configuration
73
24
-11 Configuration
75
25
-13 Configuration
77
26
-14 Configuration
79
27
-15 Configuration
81
28
-17 Configuration
83
29
-18 Configuration
85
30
Special Tow Gear - H.L.H.
92
31
Restoring Moment for Two- and Four-Point
Suspension with 25,000-Pound Load
97
32
Typical Yaw Divergence of 25,000-Pound
Helicopter Fuselage
98
33
Low-Speed Towing Characteristics
(Gross Weight 38.000 Pounds, C.G. at Sta. 550,
Zero Skew Angle)
100
34
Trim Characteristics vs Gross Weight -
Single Rotor
102
35
Vertical Response to Release of
Self-Propelled Mortar in Hover
103
36
Vertical Response to Release of
Self-Propelled Mortar at 60 Knots
104
37
Pitch Response to Release of 5-Ton
Wrecker in Hover
105
38
Pitch Response to Release of 5-Ton
Wrecker at 60 Knots
106
39
Vertical Bounce Mode Frequencies vs
Cable Length Without Decoupler
108
40
Decoupler Spring Rate vs Load
109
k
7
Xl
F£gurs
Page
41
Vertical Bounce Mode Frequencies tb
Cable Length With Decoupler
no
42
H.L.H. Fuselage Mathematical Model
112
43
Clutch-Reverser Unit Operation
138
44
Four-Point Hoist Attitude at Cable
Extremes
143
45
Hydroelectrical Feedback System,
Four-Point Hoist
146
46
Difference in Cable Length vs Cable
Travel, Four-Point Hoist System
150
47
Controls and Indicators, Single- and
Four-Point Hoists
155
48
Gearing Schematic, Single-Point Hoist
160
49
Center-of-Gravity Shift Due to Hoist Load at
Maximum Limits of Cable
161
50
Drum and Support Structures,
Single-Point Hoist
162
51
Load Isolator and Support Structure,
Single-Point Hoist
163
52
Input Housing, Single-Point Hoist
165
53
Drum and Support Housing,
Single-Point Hoist
166
54
Mounting Arrangement, Single-Point Hoist
168
55
Critical Section, Level Wind Mechanism,
Single-Point Hoist
170
56
Bellaouth, Scrub Roller, and Ball Screw
Assembly; Single-Point Hoist
174
57
Belleville Washers, Free Reeling Clutch,
Single-Point Hoist
179
58
Free Falling of Load, Single-Point Hoist
181
59
Load vs Free Fall Velocity,
Single-Point Hoist
183
xli
Page
60
Major Structural Members,
Four-Point Hoist
198
61
Side Plate, Four-Point Hoist
199
62
Induced Axial Loads in Level Wind
Ball Screw, Four-Point Hoist
200
63
Load vs Free Fall Velocity
Four-Point Hoist
206
64
Reliability Block Diagraa of
Cable Cutter Systen
219
65
Typical Military Vehicle
238
66
Variable Speed Drive
247
67
Output Parer and Torque vs Output
Speed, Variable Speed Drive
249
68
Single-Point Hoist Installation,
Single-Rotor H.L.H.
251
69
Four-Point Hoist Installation,
Single-Rotor H.L.H.
253
70
Single- Plus Four-Point Hoist
Installation, TandesnRotor H.L.H.
255
71
Single-Point Hoist
257
72
Clutch-Re verser Unit and Angle Gearboxes
259
73
Four-Point Hoist
261
74
Cargo Hook, 12,000-Pound Capacity
263
75
Isolator, Four-Point Hoist
265
xlll
TABLES
T»bl« Page
I
Basic Design Data
3
II
U.S. kray Vehicles - Length and Width
Dimensions, 15,000- to 40,000-pound Class
9
III
U.S. Army Vehicles with Heights of 130 Inches
or More, 15,000- to 40,000-Pound Class
10
IV
topical U.S. knay Vehicles
10
V
Hoist Location - Single-Rotor Aircraft
14
VI
Hoist Location - Tandem-Rotor Aircraft
18
VII
Drum Material Mechanical Properties
28
VIII
Drum Thickness Summary, Single Cable
Layer Designs
30
IX
Drum Thickness Sumnaxy, Multiple Cable
Layer Design
32
X
Hoist Gear Ratios
32
XI
Sumnaxy of Single-Point Cargo Hoists
37
XII
Sumnaxy of Multi-Point Hoists
38
XIII
Hydraulic Pump Summary
39
XIV
Hydraulic Motor Summary
40
XV
Helicopter Cargo Hooks
52
XVI
Basic Data - Cargo Hooks
53
XVII
Single-Point Plus Four-Point Load
Suspension Configurations
87
XVIII
Single-Point Plus Two-Point Load
Suspension Configuratioxis
89
XIX
Estimated Parasite Drag
Id
xx
Reliability/Maintainability Conparison,
H.L.H. External Cargo Handling Systems
117
xiv
V
\
?
i'
Table
\
i
r>
‘ji
j
Page
in
Summary, Productivity Analysis
129 i
mi
Heavy Lift Helicopter Qualitative
Evaluation Matrix
330
«
mn
Major Support Structure Bearing Reactions at
P " 150,000 Pounds, Single-Point Hoist
164
in?
Gear Sumnary - Single-Point Hoist
186
XXV
Sunniary of Bearing Lives and Loads -
Single-Point Hoist
188
XXVI
Stannary of Bearing Lives and Loads -
Single-Point Hoist
189
XXVII
Critical Section Shaft Stresses -
Single-Point Hoist
190
XXVIII
Sumnary of Gear Tooth Bending and
Compressive Stresses - Clutch-Re vereer Unit
193
xnx
Summary of Bearing Loads and Lives -
Clutch-Revereer Unit
194
xxx
Bearing Lives and Loads -
Upper Angle Gearbox
195
xra
Bearing Lives and Loads -
Lower Angle Gearbox
196
►
nm
Gear Summary - Four-Point Hoist
208
XXXIII
Critical Section Shaft Stresses -
Four-Point Hoist
212
anv
Summary of Bearing Lives and Loads -
Four-Point Hoist
233
*
XXXV
Weight Sumnary - 40,000-Pound
External Cargo Handling System
216
...
XXXVI
Reliability and Maintainability
Characteristics
220
1
XXXVII
Failure Mode and Effect Analysis
222
|
XXXVIII
List of Military Vehicles
238
1
XXXIX
Design Data, Mechanical Variable Speed Drive
246
XV
'A
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4
5TKB0LS
*
A
B.L.
C.G.
C 0
CONVKN
d c
d g
<i
djn
d o
dp
D
E
^d
®L
%
h
f C
*r max
ft
ft aax
P
Acceleration
Area, Strength Factor
Buttock Line
Center of Gravity
Basic Static Capacity of Bearing
Conventional
Cable Diameter
Gear Diameter
Inside Diameter
Mean Diameter
Outside Diameter
Pinion Diameter
Diameter
Modulus of Elasticity
Drum Modulus of Elasticity
Cable Longitudinal Modulus of Elasticity
Cable Transverse Modulus of Elasticity
Bending Stress
Compressive Stress
Maximum Radial Stress
Torsional Shear Stress
Maximum Tangential Stress
Gear Face Width
Axial Clutch Load
xvi
g
H.L.H.
HP
I
J
K
Ka
H
It L
^drum
^SCrSW
m
“b
*bY
**bH
HP
MMH
M.S.
Allowable Bending Stress
Allowable Compressive Strese
Coopresolve Yield Point
Flight Hours
Normal Force
Fuselage Station, Factor of Safety
Ultimate Tensile Strength
Gravitational Constant
Gallons per Minute
Heavy Lift Helicopter
Horsepower
Moment of Inertia
Polar Moment of Inertia
Drum Pressure Constant
Knots
Stress Concentration Factor
Length
Lead of Drum
Lead of Screw
Mass
Bending Moment
Vertical Bending Moment
Horizontal Bending Moment
Hydraulic Motor
Maintenance Man-Hours
Margin of Safety
“t
MTBUMA
n
”1
n.a.
Ny,
OGB
Pc
psi
P all«
r c r
P e
PF
P limit
p ult
rp»
B
RES
S
t
T
W.L.
Torsional Moment
I
Mean Time Between Unscheduled Maintenance Action
i
Number of Friction Surfaces in Brake ojr Clutch
Number of Layers of Cable
it
Nautical Miles I' 1
Change in Directional Restoring Stability
Out of Ground Effect !
I
Cable Pitch
Pounds per Square Inch
Axial Load, Pressure on Clutch or Brake Plates
Allowable Load
Cable Load
i
Critical Buckling Load
External Pressure on Drum
Hydraulic Puap
Cable Limit Load
Cable Ultimate Load
Revolutions per Minute
Bearing Reaction
Reservoir
Reduction Ratio
Distance
Thickness, Time
Torque, Tension
Weight
Water Line
4
*
xvili
X
z
Zero Mom
a
y
8 A
2
V
e
t
p
b
V
or
Tangential Tooth Load
Tooth Form Factor
Section Modulus
Zero Moment
Lead Angle, Angular Acceleration
Cable Pressure Angle
Deflection
Summation
Efficiency
Fore and Aft Cable Angle
Side Cable Angle
Density
Coefficient of friction
Poisson's Ratio
Angular Velocity
six
\
■A-
*“"*•«* ■ *****
INTRODUCTION
The primary purpose of the heavy lift helicopter ie that of an aerial
hoist and transporter for heavy loads including combat vehicles and ether
large, bulky items which cannot be lifted by ether helicopters. There¬
fore, an initial consideration in the design of a 40,000-pound-payload
heavy lift helicopter Mist necessarily be concerned with its external load
handling winch and hoist system.
To provide good reliability and maintainability, adequate in-flight safety,
and simple and accurate controls at a mini mm weight, the cargo handling
system must be designed concurrent with the helicopter airframe, particu¬
larly in those areas of interface. The basic airframe configurations used
for this evaluation are those of the previous heavy lift transmission and
rotor system studies conducted by Sikorsky Aircraft for USAA7LABS
(References 3 and 4). The missions used for productivity analyses are
also those of the previous studies.
■0
1
PHASE I
DESIGN ANAUSIS
DISCUSSION
In this initial phase, both separate function configurations and combined
function configurations will be investigated for two basic external load
handling system arrangements: single-point plus two-point load suspension
and single-point plus four-point load suspension.
Separate function configurations are those systems that incorporate indi¬
vidual single-point and multi-point hoists. Combined functions are those
where the multipoint hoists are used to perform both the single and
multi-point missions. Die hoist systems to be e \luated will Include
those aircraft related components necessary for control, load attachment,
suspension, hoisting, and isolation of the load from the aircraft.
The systems will be evaluated on the basis of power requirements, system
efficiency, weight, reliability, safety, maintainability, cost, and tech¬
nical confidence. At the completion of Phase I, Design Analysis, the
single- plus two-point and single- plus four-point systems that best meet
the cargo handling requirements of a 40,000-pound heavy lift helicopter
will be selected and one of these will be recommended for preliminary
design in Phase II.
«
2
BASIC DATA
DESIGN REQUIREMENTS
The heavy lift helicopter external load handling system evaluated herein
has been designed to meet the requirements shown in Table I.
TABLE I
BASIC DESIGN DATA
Single-
Point
Two-Point
(Per Hoist)
Four-Point
(Per Hoist)
Load (lb)
40,000
23,100
11,550
Ultimate Load Factor
3.75
4.2
4.2
Usable Cable Length (ft)
150/80
50
50
Cable Angle - Static
±30°
±30°
±30°
- Dynamic
±15°
±15°
±15°
Minimum Cable Speed (fptn)
60
30
30
Minimum Service Interval (cycles)
1200
Minimum Retirement Interval (cycles)
3600
System Weight Goal (lb)
4000
MISSION RB3UIREMEWT3
As an aid in evaluating the various external cargo handling system con¬
figurations covered herein, the following mission spectra were assumed
for the 40, OOO-pound-payload heavy lift helicopter. It has also been
assumed that there is an equal frequency of occurence of each mission.
Transport Mission
t
■9
2 Min. Hover
Sea Level
did Day
3
***
Transport Mission
Payload
Radius
Vcruise
Vcruise
Hovering Time
Reserve Fuel
Hovering Capability
Mission Altitude
Fuel Allowance for Start, Warm-up, and Takeoff
MIL-C-5011A
Heavy Lift Mission
10 Min.
Hover
Sea Level
Std D«y
20 n.m.
Payload
Radius
Vcruise
Vcruise
Hovering Time
Reserve Fuel
Hover Capability
Fuel Allowance for
MIL-C-5011A
AIRCRAFT DESCRIPTION
Figures 1 and 2, pages 5 and 7 describe the single- and tandem-rotor heavy
lift helicopters, respectively, used as the aerial vehicles for the cargo
handling configuration studies covered herein.
20 tons (outbound)
20 n.m,
95 knots (20-ton payload)
130 knots (no payload)
5 min. at takeoff
10 min. at destination with payload
10£ of initial fuel
Sea level, standard atmosphere
Start, Warm-up, and Takeoff
12 tons (outbound)
100 n.m.
110 knots (12-ton payload)
130 knots (no payload)
3 min. at takeoff (with 12-ton payload)
2 min. at midpoint *
1 Of, of initial fuel
6,0C0 ft, 95°F (OGE), takeoff gross
Sea level, standard atmosphere
4
672
1243(l03'-7") FUSELAGE LENGTH
1381 (115-1") OVERALL LENGTH
B.
ROTOR DATA
DIA.
BLADE CHORD
NO. OF BLADES
BLADE AREA
OVERLAP
ASPECT RATIO
70.58
3.76’
3 ,
795.00
23.30’
9.40
Figure 2. Tandem-Rotor H.L.H.
STA
300
4 I- 10*
100
198
1402(117
7
INVESTIGATION OF VEHICLES
To aid in the evaluation of various external load handling systems, it is
necessary to survey the types of equipment to be carried. This informa¬
tion is basic, since it defines the aircraft-load interaction, the place¬
ment of hoists on the aircraft, and the means of load acquisition. Since
the type of equipment to be carried was not specifically described in the
contract, it was necessary to conduct a survey of the vehicles presently
in use by the U.S. Array. A summary of the vehicles reviewed is contained
in Appendix I.
Of the 94 vehicles listed, only 37 (items 49 thru 86 of Appendix I) were
in the 15,000- to 40,000-pound weight class. Since vehicles below 15,000
pounds in weight are well within capabilities of other aircraft in the
U.S. Amy inventory (i.e., the CH-47A and CH-54A),they were not considered
to have a major influence on the cargo handling system design. Table II
shows average, maximum, and minimum dimensions of vehicles in the 15,000-
to 40,000-pound weight range. These envelope dimensions were used as an
aid in the placement of the multi-point hoists on the aircraft.
TABLE II
U.S. ARMY VEHICLES - LENGTH AND WIDTH DIMENSIONS,
_15.000- to 40.000-PQUND CLASS_
Minimum
Average
Maid mum
Length, Inches
169
265.5
600
Width, Inches
92
106.5
145
Vehicle Designation
M-114
None
CH-47
Item Number*
49
-
56
Net Weight, Pounds
15,000
-
16,000
♦Item number in .'^pendix I
A list of vehicles greater than 130 inches in height is shown in Table III.
Three vehicles (the CH-47, the M-129 STIii Van, and the M-292 Truck Van)
allow less that what is considered to be a practical clearance for any
form of ground type pickup. The unloaded ground clearance of both single-
and tandem-rotor heavy lift helicopters is 156 inches (Reference Figures
1 and 2, pages 5 and 7)* Therefore, a clearance of lees than 17 inches is
considered inadequate due to the landing gear oleo compression during the
hoisting of vehicles. This dictates hovering pickup for vehicles of this
size.
9
at--*—.Ayal .fii) - 1 »VtTH' -.Mir*
TABLE ZIX
U.S. ARM! VEHICLES WITH HEIGHTS OF 130 INCHES OR MORE*
CLASS
M 313 STLR Van Equip.
M 220 Truck Van
M 109 Truck Van
CH-47 Helicopter
M 129 STLR Van
M 292 Truck Van Equip.
♦Item number in Appendix I
It was impossible, within the scope and tine frame of this study, to com¬
plete investigation of the aerodynamic properties of all 37 vehicles in
the single- , two- , and four-point hoist suspension modes. Therefore, a
representative nunber of vehicles, as shown in Table IV, were selected as
tjp uO and formed the basis for the hoist system configuration
and aircraft-load interaction study.
TABLE IV
TTPICAL U.S. ARM* VEHICLES
Length Width Height Weight
oMfbEiBHfCiEjCMw
Iu2J
155 m Howitser 190
Persomel Carrier 169
5-Ton Wrecker Med. 310
Self-Propelled Mortar 221
12,700
15,276
33,320
43,200
•Item number in Appendix I
Part of the efficiency of the four-point system is derived from its
ability to be hooked directly to fittings on many types of loads (see
page 94 ). For this reason, more data on the size, location, and struc¬
tural adequacy of pickup points on all of the 37 vehicles is needed to
finalize the hoist locations.
10
While these data are not required if the two-point system is used (see
page 91 ), it would be valuable if the size, location, and structural
adequacy of the pickup points were known. The use of slings that can be
attached directly to fittings on the vehicle will be sinpler and more
efficient than the development of methods for attachment of a standard
nylon web sling to the underside of all the 37 vehicles.
L
11
HOIST SYSTEM AND COMPONENTS DESIGN
HOIST LOCATION AND TYPE
Introduction
The external cargo handling systems to be evaluated herein will incorporate
provisions for lifting and cariying 40,000 pounds in both single- and
multi-point modes. Systems that incorporate individual single- and multi¬
point hoists (separate function) and those that employ one system to per-
fora both modes of operation (combined function) will be studied for both
single- and tandem-rotor aircraft.
Hoist Location - Single-Rotor Aircraft
The single-point hoist for the single-rotor aircraft of Figure 1
is located directly under the main rotor at F.S. 550 to minimize the
effect of load oscillations on aircraft stability. It is . seated in a
well in the fuselage and will not extend below the airfraiut when the hook
is in the full-up position. This will permit a can > or personnel pod to
be carried by the multi-point hoists without removal of the single-point
hoists. The main gearbox support structure can be utilized to provide the
required hoist mounting with a minimum increase in weight.
The two-point hoists are located on B.L. 0 at F.S. 406 and F.S. 694. The
horizontal spacing of 288 inches was based on the survey of military ve¬
hicles in the 15,000- to 40,000-pound weight clasB. It allows most of
tho vehicles in this category to be lifted off the ground in the ground
pickup mode. The hoists are located in wells in the fuselage so that they
do not extend below the airframe when the hooks are in the full-up posi¬
tion, This permits the pod to be pulled up and locked to the fuselage.
This relatively high hoist location also reduces the effect of lateral
load oscillations on aircraft controllability, since the cable reaction
point is quite close to the location of the center of gravity of the air¬
craft. Hoi'it well size requires the addition of approximately 40 pounds
to the airframe structure. The selection of these locations for the two-
point hoists may conflict with the airframe designers desirable location
for fuel cells.
The four-point hoists are located on B.L. 70 and at F.S. 406 and F.S. 694.
Horizontal and lateral spacing was selected to achieve compatibility dur¬
ing ground pickup with the widest variety of loads. The hoists are uni¬
versally mounted on a davit type structure with suitable aerodynamic fair¬
ing, and they do not extend below the fuselage. This permits a pod to be
pulled up and locked to the fuselage. As in the two-point system, this
relatively high hoist location also reduces the effect of lateral load
oscillations on aircraft controllability, since the cable reaction point
is quite close to the location of the center of gravity of the aircraft.
The universal-type mounting permits the hoists to be pivoted in order to
reach attachment points on outsized vehicles without inducing heavy side
loads on the hoist. Figure 3, page 13, shows the load attachment point
space envelope and its variation with distance between hoist and the
12
t
Figure 3. Four-Point Hoiet Load Attachment Point
Space Envelope,
Note: Area enclosed by circles gives physical di¬
mensions of pickup points on loads that can be
lifted without exceeding the peradssible 60°
cone angle for the four-point hoists at cable
lengths specified,
i
b
&
23
ground with ths hoist locations and cable angles selected. Page 22 con¬
tains an explanation of the cable angle requirement.
TABLE V
HOIST LOCATION* - SINGLE-ROTOR AIRCRAFT
Single-Point
Hoist
Two-Point
Hoist
Four-Point
Hoist
Fuselage Station
550
406 & 694
406 & 694
Buttock Line
11.25 right to
11.25 left**
0
70
Waterline:
Longitudinal Cable Swing
210
200
200
Lateral Cable Swing
175
200
200
♦Locations described refer to actual cable reaction point
♦•With 150 feet of cable
The actual cable reaction points given in Tables V and VI
indicate a variation in the waterline location for lateral and longitu¬
dinal cable swing for the single-point hoist. This variation, as illus¬
trated in the sketch below, is due to the basic design of the hoist.
FRONT VIM SIDE VJM
H
Figure 4. Conventional Design - Single-Point Holst.
A .
15
BELLMOUTH
SIDE SUPPORT
INCHES
c
Hoist Location - Tandem-Rotor Aircr>ift
In the tandem-rotor aircraft (Figure 2, page 7) the single-point hoist is
located approximately midway between the rotors to minimize the effect of
load oscillations on aircraft stability. It is installed in a well in the
fuselage and does not extend below the fuselage when the hook is in the
full-up position. This allows pods to be carried by the multi-point
hoists without removal of the single-point hoist*
The two-point hoists are located on B.L. 0 at F.S. 413 and F.S. 701. The
horizontal spacing of 288 inches was based on the survey of military ve¬
hicles in the 15*000- to 40*000-pound weight class* It allows most of the
vehicles in this category to be lifted off the ground in the ground pick¬
up mode.
The hoists are located in wells in the fuselage so that they do not extend
below the airframe when the hook is in the full-up position. This permits
pods to be lifted up and locked to the fuselage* Since the hoists are
located relatively high in the fuselage* the effect of lateral load oscil¬
lations on aircraft controllability is quite small. Hoist well size re¬
quires the addition of approximately 40 pounds to the airframe structure.
In addition, this could conflict with the most desirable fuel cell loca¬
tion. No attempt has been made to assess the importance of this conflict
in this study.
The four-point hoists are located on B.L. 70 at F.S. 413 and F.S. 701*
since this ensures compatibility during ground pickup with most of the ve¬
hicles in the 15,000- to 40,000-pound weight category. The hoists are
universally mounted on a davit type structure with suitable aerodynamic
fairings. Full-op position of the hooks permits pods to be pulled up and
locked to the fuselage. As in the two-point system* this relatively high
location reduces the effect of load oscillations on aircraft controlla¬
bility. The universal mounting permits the hoists to be pivoted to reach
attachment points on out sized vehicles without Inducing heavy side loads
on the hoists.
Single-Point Holst
The conventional (one part, single reeved type) hoist offers the most ad¬
vantages for application in the single-point location. Because it need not
be mounted on a universal-type joint, it can be driven mechanically. It
can also fit into a well which has limited vertical space. The drum axis
is mounted at right angles to the longitudinal axis of the aircraft* and
the level wind assembly is allowed to pivot about the drum axis. This per¬
mits large cable angles during tow operations without inducing high loads
on the bellmouth. The simplest and most reliable version is the one that
requires only one layer of cable wrapped on the drum. Multiple layering
is possible but it requires the use of a more complicated, less efficient*
and less reliable type of feed screw to ensure even winding of the cable.
Figure 4* page 15* shows the conventional* single-layer hoist similar to
that used in the CH-54A.
17
TABLE VI
HOIST LOCATION*
- TANDBC-R0T0R AIRCRAFT
Single-Point
Hoist
Two-Point
Hoist
Fuselage Station
557**
423 & 701
413 ft 701
Buttock Line
0
0
0
*
Waterline:
Longitudinal Cable Swing
165
200
200
Lateral Cable Swing
190
200
200
♦Locations described refer to actual cable reaction point
♦"With one-half usable cable length extended
For the single-rotor aircraft of this study* a maxi nun cable length of 100
feet can be carried on a single-layer hoist without exceeding desirable
lateral cyclic stick movement. The limitation on cyclic stick movement is
based on that presently attained in the CH-54A.
For the tandem-rotor aircraft of this study, the single-point hoist drum
can be mounted with its axis located parallel to the aircraft longitudinal
centerline. With this arrangement, the only limitations in cable length
are those iiqposed by permissible C.G. range and/or hoist well siae. This
installation, however, restricts the towing capability of the tandem-rotor
aircraft from the single-point hoist.
Another type of hoist investigated in this study was the two-part, double-
reeved type shown in Figure 5, page 19. This type is used extensively in
coamerclal practice. The two-part, double-reeved hoist has two primary
disadvantages when compared to the conventional (one-part, single-reeved)
hoist. A cable backlash suppressor will be required to keep the cable
from jumping off tno drums, both pulleys on the traveling block, and the
upper pulley when a load is air dropped. While a suppressor has already
been developed for the conventional hoist, the suppressor required for the
double-reeved type will be considerably more cooler. Cable cutters for
the double-reeved hoist have to be mounted in four places to ensure that a
sheared cable will not jam up in one of the three pulleys. This require¬
ment for four cable cutters reduces the inherent safety features of the
single-point suspension.
18
The use of a conductor reel to provide a means of transmit ting electrical
power to the hook adds approximately 50 pounds to the weight of the single-
point hoist system* In addition, the use of the reel also reduces the
reliability of the system because of the added complexity and the rela¬
tively unprotected location of the conductor cable.
Two-Point Holsts
Both the conventional and the eero-moment hoists are applicable to the two-
point requirement. A zero-moment hoist is one which maintains the same
line of action for the load for all lengths of cable extended and hence
always has the same reaction point on the supporting structure.
Since there is no requirement to have the two-point hoists pivot laterally
in order to pick up outsised loads, the use of a zero-moment hoist is not
mandatory. Its greater vertical dimension, required to mount a load
isolator most efficiently, requires the use of a deeper well in the fuselage.
The conventional hoist, as described in the Single-Point Hoist section,
fits in a shallower well in the fuselage and is equally adaptable to both
mechanical and hydraulic power sources. Therefore, since there is only a
limited amount of vertical space available in the fuselage of either the
single- or tandem-rotor aircraft, a conventional hoist is the most advan¬
tageous.
Due to variations in the size of loads to be carried in the multi-point
mode, it is necessary to increase the design load rating of the two-point
hoists to account for cable angle. A cable angle (with the vertical) of
30° has been selected to permit variations in longitudinal dimension of
load. The required hoist rating is then
Rating - - 23,100 pounds (l)
Four-Point Hoists
The use of the zero-moment hoist is mandatory in the four-point system.
Its ability to be pivoted to any position allows attachment to a wide
variety of cargo sizes and shapes. A mechanical drive system for the aero*,
moment hoist will be extremely complicated, while hydraulic power offers
automatic load equalizing, and, by use of a feedback system (see Figure 13,
page 48), permits synchronised operation.
Sikorsky Aircraft experience has shown that a capstan type hoist offers
weight advantages only when the long cable lengths art required, as in
rescue winch applications. For hoists requiring only 50 feet of cable at
low speed and relatively high load capability (such aa the H.L.H. multi¬
point application), the capstan principle offers no advantages; in fact,
there may be some penalties in weight and cable life.
21
A single drum design, universally mounted, offers lighter weight, greater
cable life, and somewhat more reliability. In configurations that elim¬
inate the single-point hoi"t. it is necessary to use three layers of cable
in oruer to retain the -ero-tA&ment capability. As discussed in the Single-
Point. Hoist section, multiple layering requires the use of a more compli¬
cated and somewhat less reliable type of feed screw than is needed in a
single-layer design. Figure 6 , page 23, shows a typical single-layer, zero-
moment hoist*
It should also be noted that it Is necessary to increase the design load
rating above the theoretically possible rating of 10,000 pounds, since it
is not possible to lift the wide variety of loads required with this rating.
If cable angles (with the vertical) of 30° are utilized to permit the
necessary variation in load length and width, a hoist rating of 10,000/Cos
30° " 11,550 pounds is required. Figure 3, page 13, shows the allowable
load attachment point space envelope with the selected hoist locations anl
the 30° cable angle requirement.
HOIST DESIGN
Discussion
The initial consideration in the design of a hoist is the selection of a
cable that meets the load and functional requirements (i.e., nonrotating
construction with electrical conductors in the core) for this study. Once
the cable diameter and the wire size to be used in the cable have been
selected, the drum diameter can be calculated. Using standard conmorcial
practice which requires cable drum diameter to be a minimum of 400 times
the individual wire diameter, it is possible to determine minimum drum
diameter requirements. Other considerations may dictate larger drum diam¬
eters, however. In the case of the single-point hoist, the determining
factor Is the requirement to carry as much cable as possible without exceed¬
ing desirable aircraft control limitations (see Single-Point Hoist section,
page 17)* After the drum diameter has been established, the drum thickness
Is determined by analysis. Once this is completed, the gearing system can
be established and the load brake can be integreted into the primary gear
train. The auxiliary gear train is then designed into the level wind arms
to provide power for scrub rollers and level wind screw drives. For a
single layer of cable, a simple ball screw and nut can be used, whereas in
multiple layer hoists, a more complex double helix screw is required.
The accessory drive gearing can then be designed to provide drives for slip
rings, cable length Indicator potentiometers, and limit switch actuators.
Cable cutters, cable backlash clamps (or covers), guide rollers, and limit
switches are next Integrated into the final design. The support structure
is designed to conform to established mounting structures, and load cell?
and Isolator units are designed into the support structure.
22
nr—— LOAD BRAKE
3
—CABLE TENSION —
ROLLERS
LEVEL WINO
FEED SCREW
Three cable sizes are needed to meet the requirements for the single-point,
two-point, and four-point hoists. All cables are designed to carry seven
electrical conductors in a centred core and are of nonrotating construc¬
tion. Five conductors are used to operate the cargo hook; two are spare
wires. All are of an extra flexible 18 x 19 construction. A 1.39 diameter
is required to meet the 150,000-pound minimum breaking strength require¬
ment of the single-point hoist. A 1.06 diameter cable meets the 97,000-
pound requirement for the two-point hoist, and a 0.79 diameter cable neets
the 48,500-pound requirement for the four-point hoist.
Since the aircraft's angular accelerations in a maneuver are combined with
the linear acceleration effects, the loads at all multi-point hoist attach¬
ment points are increased. This load magnification is due to the location
of the hoist attachment points some distance from the aircraft's center
of gravity. Therefore, an ultimate load factor of 4.2 is required for all
multi-point hoists instead of the 3.75 value used for the single-point
hoist. Further details on hoist cables are found in the Hoist Cables
section, page 49*
The basis for hoist drum analysis for both single and multiple layers of
cable is given on the following pages.
Single Layer of Cable:
25
Stresses in the hoisting drum shell are the result of loads imposed by:
External pressure (P # ) from coiling of ropes under tension
Bending maaent (M^) from rope tension
Torsional moment (M^) due to pcwer transmission from the gear
train to the rope
These loads and stresses can be calculated from the following equations:
Pc
where
A is equal to
P„ L
\ - -v
where
Z is equal to
\
2Z
4
1 - ( 1 - -&-)
d
o
( 2 )
( 3 )
( 4 )
( 5 )
( 6 )
( 7 )
( 8 )
( 9 )
Por the hoist drums designed in this study, the length-^to-diameter ratio
is approximately equal to 1. Therefore, for the purposes of finding the
approximate magnitude of the compressive, bending, and torsional stresses,
the following assumptions can be made:
L " d o " " D
26
Combining equations (2), (3)> and (4),we obtain
*c
Pc*
( 10 )
Combining equations (5)> (6), and (7), we obtain
Combining equations (7), (8), and (9X we obtain
By using the previously mentioned assumptions, it can be seen that the
bending and torsional stresses are of the same order of magnitude.
In ord#r to compare the magnitudes of the stresses further, the following
assumptions can be made:
t
D
.038
D - 25 t - .95
1
These assumptions are within average values used in this study.
Substituting these values into equations (10) and (ll),we obtain
f r
\S\
-*c
( 13 )
c
.914
*b
rj
'V/
f. A/
t rJ
P c
Twil
( 14 )
Therefore, the bending and torsional stresses in the hoisting dnm shells
used in this study are in the order of magnitude of 70 times less than the
compressive stresses and may be neglected when designing the drums.
Material Trade-off Investigation
The following materials have been tentatively selected for use in the
drum:
27
TABLE VII
PRIM MATERIAL MECHANICAL PROPERTIES
Material
P tu
Ultimate
Tensile
Strength
Held
Strength
Density
(lb/in.3)
7079-T6 Aluminum
72,000
65,000
.099
AZ8QA-T5 Magnesium
42,000
25,000
.0652
4340 Steel
180,000
179,000
.283
The drum weight per inch le given by
W - t d c (-f
-1 )p *
(15)
Since the design is based on compressive stress only, equations (13) and
(15) mj be selred simultaneously te give
V -
P ▼ d c
( 16 )
For a given hoist drum, d o> Pc* and will be constant. Hence, the
lightest druB will be the drum with the lowest value of p? c /t c »
The following is a design relationship for cable load:
P c (ult) - 1.304 P c (yisld) ( 17 )
Ths weight of the drum must be investigated under ultimate and yield con¬
ditions. To accomplish this, the following constants can be evaluated:
ult - Ci -
yield - C 2
P* c
Jdt_ . 1.30 U P T c yield
"tu
P P c yield
F_1
cy
r tu
(16)
(19)
28
The values of P , F tu , and F^ for different materials may new be sub¬
stituted Into the preceding equations. The lightest drum based ea ultimate
load conditions will be tbs drum with the lowest value of C^. The lightest
drum based on yield load conditions will be the drum with the lowest value
of C 2 . The design will then be based on the load conditions which produce
the highest value of or Cg for a particular material. Substitution of
the values of p and ?_ Into equations (IS) and (19) for various
materials is given below:
1. 7079-T6 Aluminum forging
C, - Ic ylg3,d . 1#?9 x 10 -6 p ^ eld
1 F tu C
C 2 , P P c ■ 1.52 x 10" 6 P c yield
cy
2. AZ8QA-T5 Magnesium forging
C - 3 ,« 3°4 P P c d . 2.02 x lCf 6 P„ yield
1 F tu C
C 2 - - - 2.60 x 10” 6 P c yield
cy
3. 4340 Steel forging
C. - - 2.05 x 10 -6 P c yield
"tu
C 2 - ..P I sl ^? )A — . 1.58 x 10" 6 P c yield
cy
Therefore, the 7079-T6 aluminum forging, based on ultimate load conditions,
will result in the lightest hoist drum design. The aluminum drum will be
31 pet lighter than a magnesium dnm (designed for yield conditions) and
12 pet lighter than a steel drum (designed for ultimate conditions).
29
'«***»*>* wwaamtiWEB
Notes The use of a maraging steel of P^ u - 250,000 psl will re¬
sult In a significant weight savings. This design will, how¬
ever, have a very low thickness to diameter ratio and may pre¬
sent machining problems because of distortion of the thin-walled
drum. Maraging steels have also been known to present
stress corrosion problems. In addition, the cost is approxi¬
mately six times that of alumintn. Therefore, high strength
steel was not used for the hoisting drum. However, it is
worthy of further study as a possible means of weight saving
along with the titanium alloys and glass. An evaluation of
these materials will be conducted as part of Phase II if time
permits.
Drum Thlckn e e Analysis
The equation .. r cos*)receive stress may now be solved directly for thick¬
ness to diameter ratio using P^ u for 7079-T6 aluminum r
1 -
d o
Table VIH
two-point.
1 _ l6qaO£Lr.. P A ujt
2 / 144,000
( 20 )
summarizes the drat thickness calculations for the single-point,
and four-point hoists with a single layer of cable.
TABLE VIII
DRUM THICKNESS SUMMARY,
SINGLE CABLE L'YER DESIGNS'
»
Two-
Four-
Point
Point
Point
P c - Cable load (ult), lb
150,000
97,000
48,500
p - Cable pitch, in.
1.5
1.1875
.875
dg - Mean drum dla., in.
38.99
29.69
22.19
P e - External pressure (ult), pel
5290
5500
5190
t/d 0 - Calculated
.0382
.0398
.0376
d 0 - Effective drum CD, in.
37.806
28.905
21.568
t - Drum thickness, in.
1.44
1.15
.81
N - Weight of drum, lb per in.
16.40
10.03
5.22
•Drum Material: 7079-T6 Aluminum Alloy
30
Multiple Layers of Cable
Since the external pressure caused by coiling of the cable under constant
tension is the major factor in the design of drums considered in this
study, it is important to determine this pressure accurately. When there
is more than one layer of cable on the drum, the pressure increases, but
not as a direct ratio of the number of layers. The outer cables tend to
compress the inner cables and drum, thereby relieving some of the pressure
originally caused by the inner cables. The general equation for the
pressure on the drum ie given by
P
e
2 K P c
P c *m
( 21 )
where
K is a factor less than
(number of cable layers)
For n-|_ - 1, K m l. K can be calculated from
K
J -1
2 A El
p fd 0 +d c siny(2j-2)]
c
1
E d
K -
j(K 2 -l) + (nr-K 2 )
A »c(*l -J)
'(“l-J-l) K
\
"1 - 1
Bt(nr2)
(n x -l) d 0
(1-1) 1
<*o + d c
sin y (2j - 4) J
( 22 )
where the last tern is equal to 0 for n.< 3 (term with in
denominator)
The trial and error solutions to this equation for single and multi-point
multiple layer hoist drums are presented in Table IX.
31
TABLE IX
DRUM THICKNESS SUMMARI,
MULTIPLE CABLE LAYEB DESIGN
Single-
Point
2 Layers
Single-
Point
3 Layers
Two-
Point
3 Layers
Four-
Point
3 Layers
Pc
- Cable load (ult), lb
150,000
150,000
97,000
48,500
Pc
- Cable pitch, in.
1.5
1.5
1.1875
.875
- Drun naan dim, in.
38.99
38.99
29.69
22.19
d o
- Drua effective OD, in.
37.806
37.806
28.905
21.568
K
- Drun pressure constant
1.495
1.939
1.961
1.930
p .
- External pressure, psi
7,670
9,950
10,790
9,740
t/do
- Calculated
.0565
.0747
.0816
.0730
t
- Drun thickness, in.
2.14
2.82
2.36
1.57
W
- Drun waight, lb per in.
23.978
31.026
19.661
9.887
Hoist Gearing Configurations and Load Brake
Three different gear arrangeaents were considered. Table X swuarizes the
overall gear ratios required for the single- , two- , and four-point
hoists using a high- and low* speed input drive.
TABLE X
HOIST GEAR RATIOS
Single-Point
Two-Point
Pour-Point
High-speed input, rpa
6000
6500
7200
Law-speed input, rpa
3000
2500
2750
Output rpa
6.1
4.02
5.34
Overall ratio (high)
983.6
1616.9
1348.3
Overall ratio (low)
491.8
621.9
515.0
Min. stages of gearing (high HR)
5
5
5
Min. stages of gearing (low RR)
4
4
4
32
I
As can be seen in Table X, four gear stages are required for the low input
drive speeds,while five gear stages are required for the high input drive
speeds. These numbers refer to conventional gear stages such as a spur
gear and pinion or a conventional planetary arrangement (sun gear driving,
ring gear fixed, cage driven). If a compound planetary is used, the num¬
ber of stages can be reduced by one or two because of the higher ratios
obtained.
In general, a spur mesh will be lighter for low torque and a planetary
will be li gh ter for high torque applications because of the load splitting
capabilities. All of the hoist designs use a Weston brake for controlling
the load. This type of brake holds the load when the power source is shut
off, locks as a unit when raising the load, and slips at the same speed
as the driver when lowering the load. The Weston brake should be located
in the gear train a minimum of one gear stage from the input drive. This
is to assure locking in the event of a mechanical failure at the drive
source,which is usually the weak link. The small drag provided by the
first stage will prevent load runaway by providing the torque necessary to
lock the brake plates when lowering the load. Since the Weston brake is a
purely mechanical device, the load can be maintained in the event of a
hydraulic failure or a mechanical failure of the drive train which occurs
before the brake input.
The following discussion refers to a hoist utilizing low-speed inputs.
For high-speed inputs, another planetary stage can be added. Figure 7
shows a hoist arrangement utilizing three conventional planetary stages
and a spur gear input stage.
|
I
I
f
1
DRIVE “7 i— WESTON BRAKE
Figure 7. General Gear Arrangement - One
Spur and Three Planetary.
33
... ■ ysjjjg
In the arrangement shown in Figure 7, it is required that the motor be
located off the centerline of the hoist. This may cause interference
p rob leas with the dim and motor mount. Figure 8 shows another gear con¬
figuration using two spur gear meshes and two conventional planetary stages.
In this arrangement, the drive is located on the hoist centerline, and the
two side plates can be rigidly connected.
Figure 8. General Gear Arrangement - Two
Spur and Two Planetary.
Figure 9 is a compound planetary driven hoist arrangement. It can be de¬
signed with reduction ratios of 3*1 to 150*1. A design of this type can
replace three conventional planetary stages at a AO pet weight savings.
Its disadvantage is that its efficiency is somewhat lower than that of
three conventional plane tar ies. A 131:1 compound planetary was designed
with an efficiency of approximately 92 pet, while the efficiency of an
equivalent system using three conventional planetar ies was approximately
98 pc*.
TUB) RING
GEAR
CAGE
(NO LOAD)
OUTPUT
RDD
GEAR
INPUT SUN
GEAR
L PLANETARY
PINIONS
HOIST
Figure 9. General Gear Arrangement
Compound Planetary.
35
I
Figure 10 shows a general gear arrangement for a compound planetary with
a conventional planetary for the first stage. This arrangement Is the
lightest considered* but approximately 6 pet In efficiency Is sacrificed.
Figure 10. General Gear Arrangement -
Conventional and Compound Planetary.
36
syawor
A complete eunaary of all the cargo hoist types evaluated in this study is
given in Tables XI and Cl, pages 37 and 3d. The weight given includes
the weight of hooks and cables but omits the weight of the power source,
since both mechanical and hydraulic pcsrer sources were considered.
The single-point hoist designated "E" was included since it offers a modi¬
fied zero-moment capability and a significant weight savings. Its dis¬
advantages, as discussed on page Id, outweigh these advantages; hence,
further study of this type is not considered necessary. All hoists have
been designed to permit removal from the aircraft without removal of the
power source. Gear drives, hydraulic motors, and lines will remain with
the aircraft.
POWER SOURCES
^raullc
Hydraulic motor drive for either the two- or four-point hoists is possible
with existing hardware. However, the pump and motor required for the
single-point hoist will require some development. This development and
modification is required to adapt to normal temperature environmental
operation, since these unite are presently designed for the extreme tam-
perature environments on the XB-70. The motor will be a modification of a
pump used in this aircraft. The development effort could be reduced if
two smaller motors, geared to a coiroon input shaft, were used in place of
a large single motor drive. All hydraulic motor combinations offer vari¬
able speed control for the single-point hoist. The pump required for con¬
figurations that use hydraulic power for the multi-point hoists only, while
not a production unit, is of conventional design and hence should not re¬
quire any development effort.
TABLE XIII
HYDRAULIC PUMP SUMKART
ADDlication
HP
OutDUt
HPM
Diap.
(cu in./rev)
^Flow
Weight
(lb)
Single and
Multi-point
150
3500
5.873
68
55
Multi-point
only
100
4000
2.80
45
35
‘rt w*‘ vxmv
■ n**r f.-
Both high- ani low-speed hoist motors were considered in the preliminary
system evaluation in order to determine which type would best meet the re¬
quirements. Preliminary calculations indicated that the weight advantage
offered by the high-speed motors was offset by the weight of the added
transmission system drive gearing. Since weights are nearly equivalent,
the low-speed motors were selected because they offered a greater relia¬
bility and longer life as compared to the high-speed designs. The motor
selected for both traction sheaves and the conductor reel is a standard,
off-the-shelf component.
TABLE XIV
HYDRAULIC MOTOR SUMMARY
Hoist
Designation
Output
RPM
Disp.
(cu in./rev)
Wow
(oau)
Weight
(ibT
A, AA, B, C, D,
104
3000
5.25
68
44
X
104
6000
2.70
68
24
H, J, N
70
4050
2.50
45
22
A, B, C, D, X
52*
4400
1.80
34
19
P, G,
35
2500
2.35
22.5
22
L, M,
35
6500
0.95
22.5
11
K, P,
17.5
2750
0.95
11.1
10
Conductor Reel
17.5
7200
0.367
11.4
5
Traction Sheaves
1.0
1700
0.095
0.70
2.6
•Two motors required per hoist
Mechanical
All mechanical drive systems for the single-rotor helicopter are driven by
an auxiliary power plant and rotor powered accessory gearbox. This permits ,
ground operation without the rotors turning. In flight, the accessory
gearbox is shaft driven by the main gearbox. An identical system is used
to power auxiliary drives on the CH-53A. Another version, in which the
auxiliary power plant drives the accessory gear train in the main gearbox
during operation, is used on the CH-54A. These concepts facilitate •
ground check-outs of all systems, since a pilot is not required to "run up*
the aircraft. In a tandem-rotor aircraft, similar design principles can be
utilised to permit this type of operation.
A clutch-reverser unit mounted on the gearbox is used to provide the oppo¬
site shaft rotation required for raising and lowering. By actuating both
clutches, the mechanical drive system can be disengaged from the accessory
40
gearbox when hoist operation is not required. An alternate type of power
takeoff unit is described in Appendix II. It was not used in this phase
of the study because it is not a fully developed unit. The concepts* how¬
ever* are now being used in similar units for constant speed drives In
several operational aircraft. The mechanical variable speed drive offers
variable speed drive for the mechanical system and is considered worthy
of further study.
The angle gearboxes and drive shafts utilized follow standard Sikorsky
Aircraft design practice similar to that utilized in tail rotor drive
system gearboxes and tail drive shafts. No development problems are antic¬
ipated for these units.
The individual hoist clutches required in the mechanical drive versions
for the multi-point hoists follow standard automotive and marine practice.
They are multiple disc types in which the actuation force is supplied by
oil at 250 psi. This oil* provided by an accessory gearbox mounted pump*
is also used to cool the clutch plates.
Auxiliary Power Plant (APP)
All the external cargo handling systems considered in this study will be
powered either hydraulically or mechanically from the aircraft accessory
drive gearbox. This unit is driven from the primary rotor drive train
when the rotor system is operating and from an auxiliary power plant (APP)
on the ground to permit ground check-out and acquisition of loads when the
rotor system is locked.
A separate gas turbine as the sole source of power for the hoist systems
was considered and was rejected because it was heavier and less reliable.
For an APP driven cargo system* an APP with a hot day output power ranging
from 100 to 170 horsepower (reference Tables XVII and XVIII* pages 07 and
89) depending on the system configuration* is required. In addition* this
system requires either continuous operation of the APP in flight or re¬
start for raising or lowering cargo at the acquisition or release site.
Electrical
There are no electric motors of aircraft quality available in the 100-
horsepower class required for the 8 ingle-point hoists nor in the 3 5-horse¬
power class required for the two-point hoists. A ant or is available in
the 17.5-horsepower class required for the four-point hoists. However*
its weight of 17*5 pounds* compared to 5 pounds for a similar high-speed
hydraulic motor* plus the requirement for large electrical lines and con¬
siderably larger alternators* would require a considerable weight increase.
Therefore* an electrical motor drive was not considered feasible for the
external cargo handling system power source.
U
CLUTCH-REVERSES UNIT
A clutch-reveraer unit is mounted on the aircraft accessory drive gearbox
and tranmite power in either direction of rotation to the hoists. It
consists of twin reversing clutches. By declutching both of the reversing
clutches, the mechanical drive system can be disengaged from the power
source when hoist operation is not required. Oil required to actuate the
piston which clamps up the plates in these clutches is supplied at 250 psi
by a puap mounted on the accessory gearbox. By controlling the pressure
rise to 250 pel, smooth, shock-free operation is obtained. Ibis same oil
supply (transmission oil) is also used as a coolant for the clutches when
they are dieengaged. The use of a clutching arrangement, which can be un¬
coupled from the power source, is possible because each hoist incorporates
an automatic load holding brake. By providing smooth, shock-free accelera¬
tion of loads, the need for a variable speed drive can be eliminated.
In the event that it becomes mandatory to provide variable speed operation
of the single-point hoist, the clutch-reveraer unit can be replaced by a
toroid drive unit of the type described in Appendix II. While units of
the else required for this application have not been built, the design
concept has been proven by use of smaller units for constant speed drives
presently installed lu the Navy A4E.
CO NTROL SYSTEMS
Dlscwslon
The control system employed for both single and multi-point hoists is de¬
pendent upon the type of power source utilised for each system. For each
of the two power sources considered in this study (mechanical and hydrau¬
lic), a separate control system approach must be devised. Basically, the
mechanical drive concepts require mechanical clutching operations, while
the hydraulic drives are flow controlled. In the multi-point systems, the
mechanical drive concepts require electromechanical feedback, while the
hydraulic drives employ both electromechanical and fcydromechanical feed¬
back to provide load equalisation and/or synchronisation. All systems
utilise electrical control of mechanical and/or hydraulic components in
order to enable control functions to be accomplished by the pilot, the
copilot, and a dismounted craw member.
Single-Point Hoist Control System - Hydraulic Power Source
A single pump mounted on the accessory gearbox supplies the hoist motor.
This pump is also ussd to supply the four-point hoists in ths -2 config¬
uration. The hydraulic system is a pressure demand type in which the rate
of flow is varied by controls external to the pump. This control is
achieved by a torque motor controlled servo which varies pump displace¬
ment. The pressure developed it only that required by system loading plus
line losses. This system wee chosen to eliminate heat generation inherent
in systeam using a pressure compensated pump at anything lasa than full
load.
42
Single-Point Hoist Control System - Mechanical Paw
A power takeoff shaft is used to drive the hoist. A clutch-reverser unit
mounted on the accessory gearbox enables the hoist to raise or lower the
load. It also permits disengagement of the power source at which time the
load brake in the hoist maintains the cable position.
« The clutch-reverser unit is similar to the types widely used in marine
applications in small power boats. It utilizes oil-actuated, multiple*
disc clutches. The oil is supplied at 250 pai by an accessory gearbox
mounted pump. Appropriate solenoid operated valves direct the flow of oil
to the proper clutch.
41
Two-Point Holst Control System - Hydraulic Power Source
A single pump mounted on the accessory gearbox supplies the hoist motors.
The hydraulic system is a pressure demand type in which the rate of flow
is established by controls external to the pump and the pressure is deter¬
mined by the line losses and the load. It is a closed system with a pump
supercharge of 50-100 psi. Fluid from the pump is delivered to the appro¬
priate subsystem and is then returned directly to the pump inlet. A sepa¬
rate replenishment pump system supplies fluid to replace that lost from
the closed hoist system due to pump and motor leakage and bypass cooling
flow. It also provides pressure for pump control.
A feedback control system is used to provide synchronized lifting. This
system is described schematically in Figure 11, page 44. The division of
flow between the forward and aft hoists is established by servo controlled
flow dividers. The signal to the flow dividers is derived from a compari¬
son of the signals from rotary potentiometers mounted on each hoiat.
Operation of a single hoist is accomplished by supplying a bias signal to
the flow dividers to block flow to one of the hoists. Clutches are pro¬
vided to disengage the potentiometers from the hoists during beeping and
are reengaged for collective operation. This enables the cables to be ad¬
justed for any extremes in cargo shape and then permits the established
cable lengths to be maintained during collective operation. This system
eliminates errors due to differences in motor efficiency and initial set¬
tings of the divider valves. The maximum error is estimated to be in the
order of 3-7/8 inches in 50 feet.
, It will be possible to deenergize the feedback system during the initial
stages of hoisting, thus permitting the basic system, which ia inherently
load equalizing, to equalize loads on the cables automatically. This
method of hoisting is feasible only if the lifting points are located sym¬
metrically about the center of gravity (C.G.) of the load. If the C.G. ia
« not symmetrically located, the vehicle could assume an unsafe attitude.
Consider, for example, a vehicle with its C.G. located at a point 40 pet
of distance between the pickup points to be lifted with 30 feet of cable
extended. The vehicle would then assume a nose-down attitude of 43° when
the cable lengths were adjusted to maintain equal cable loads. Figure 12,
page 46, shows resultant load attitude, with varying C.G. locations, re¬
sulting from the use of an automatic cable load equalizing feature. Thera-
43
fore, in the lifting operation of the above example it is desirable to keep
the feedback system energized to maintain a level attitude of the load.
Two-Point Holst Control System - Mechanical Power Source
Power takeoff shafts are used to drive the hoists. A clutch-reverser unit
mounted on the accessory gearbox enables the hoist to raise or lower the
load. A separate clutch is used to disengage the single-point hoist, and
clutches are used to disengage either the forward or aft hoists as re¬
quired.
Synchronized operation is attained by engaging both forward and aft
clutches and then engaging the clutch-reverser unit. This operation is
automatically sequenced so that only one control motion will be required.
Equal cable loading is attained by disengaging the clutch driving the
heavily loaded hoist and allowing its load brake to maintain the load
while operating the lightly loaded hoist until equal cable loads are
attained. Load indicators will make it possible to determine when equal
loading is obtained. A more elaborate feedback system, utilizing the out¬
put of the load cells, could be devised to accomplish automatic load
equalizing. Development of such a system is feasible but probably not
warranted, since, by operating the hoists individually and using the load
indicators to equalize cable loads, the same result can be obtained with
considerably less complication.
Four-Point Holst Control System - Hydraulic Power Source
In one of the separate function systems investigated, incorporating single-
and four-point hoists, a single pump mounted on an accessory gearbox
supplies the hydraulic power to both. The systems will be isolated from
each other and from the pump by electrically operated shutoff valves.
For the hydraulically powered combined function systems, a 68 gpm, 3500 psi
pump is the source of power. Two of the separate function systems that
utilize mechanical drives to the single-point hoist require a 45 gpm, 3500
psi pump as a power source for the hydraulically driven multi-point system.
The hydraulic system is a pressure demand type in which the rate of flow
is established by controls external to the pump and the pressure is deter¬
mined by the line losses and the load. It is a closed system with a pump
supercharge of 50-100 psi. Fluid from the pump is delivered to the appro¬
priate subsystem and then returned directly to the punp inlet. A ~*parate
replenishment pump supplies fluid to replace that lost from the closed
system due to pump and motor leakage and bypass cooling flow and also pro¬
vides pressure for pump control. A feedback control system is used to pro¬
vide synchronized lifting, This system is described schematically in
Figure 13. The division of flow between forward and aft hoists and between
port and starboard hoists is established by servo controlled flow dividers.
The signal from the flow dividers is derived from a comparison of the
signals from the rotary potentiometers mounted on each hoist. Two poten¬
tiometers are mounted on each of the starboard hoists, and only one will be
mounted on each of the port hoists. The first potentiometer on the star-
45
(% DISTANCE BETWEEN SUPPORTS)
Figure 12, Multi-Point Hoist - Equalised
Load Systao, Cargo Attitude vs
C.G, Location.
46
board hoists controls the division of flew between the two forward and the
two aft hoists. The second potentiometer on the port and starboard hoists
controls the distribution of flow between the two port and the two star¬
board hoists. Operation of the hoists individually (beeping) is accom¬
plished by supplying a bias signal to the two relevant flow dividers,
blocking flow to three of the hoists.
• Clutches disengage the feedback potentiometers from the hoists during
beeping operation and reengage for collective operation. This enables the
cables to be adjusted for any extremes in cargo shape and then permits the
established cable lengths to be maintained during collective operation.
This system eliminates errors due to differences in motor efficiency and
initial settings of the divider valves. The maximvm error is estimated
to be in the order of 7-3A inches in 50 feet (see Error Analysis, page
148).
By deenergizing the feedback system during the initial stages of hoisting,
the cable loads are automatically equalized, since the basic system is in¬
herently load equalizing. Figure 12, page 46, shows the resultant load
attitude with varying C.G. locations if an automatic load equalizing
system is used.
Four-Point Holst Control System - Mechanical Power Source
A clutch-reverser unit mounted on the accessory gearbox provides power to
both the single- and four-point hoists. Clutches mounted on angle gear¬
boxes adjacent to each of the hoists permit individual operation as re¬
quired. A separate clutch is used to disengage the single-point hoist
during operation in the four-point mode.
Synchronized operation is attained by engaging the four-point hoist
clutches and then engaging the clutch-reverser unit. This operation is
automatically sequenced so that only one control motion will be required.
Equal cable loading is attained by disengaging the clutch driving the
heavily loaded hoist and allowing its load brake to maintain the load
while operating the lightly loaded hoists until equal cable loading is
attained. Load indicators will make it possible to determine when equal
loading is obtallied.
. Since this system is awkward to use effectively, it would be desirable to
incorporate a feedback system to give load equalization automatically.
Such a system utilizee the output of the load cells to supply a control
signal to the proper clutches. This signal permits clutch slippage until
equalized loading is attained. A temperature sensor in the separate
• clutch units would provide protection against overheating by locking up the
clutch to prevent excessive heat buildup.
47
|
l
I
I
i
i
i
Vi
•ft
K
r'
•I.
r.
-*-n v n>>~
FOUR-
POINT
HOISTS
FLOW
DIV
FLOW
DIV
r \
f \
f \
r \
FLOW
MF
MF
MF
(mf
DIV
L _^ —
111 1
—0—i—O— f!_
REPLENISHMENT
PUMP
MAIN
HOIST
HOIST PUMP
Figure 13* Hydraulic Schematic - Single -
Flue Four-Point Suspension.
48
ISOLATORS
>
I
tf
i
Isolators are required on all hoists to eliminate the vertical bounce
phenomena (see page 107 for a discussion of vertical bounce).
The isolators are of the hydraulic cylinder type similar to that used on
the CH-54A main cargo hoist. This type of isolator incorporates an
isolator, a load cell, shock struts, and a charging cylinder in one unit.
The shock struts serve to retard the return stroke of the isolator when
loads are air dropped. The charging cylinder which is pressurized by the
aircraft utility hydraulic system compensates for temperature induced
pressure changes in the isolator and makes up for any leakage that may
occur.
When applied to cargo hoists of conventional design, the isolator reacts
the cable load through a linkage. On zero-moment hoists, it is mounted
directly in line with the cable so that no linkages are required.
HOIST CABLES
Background
The main cargo hoist used on the CH-54A has one of the largest capacities
in existence. It is capable of raising and lowering a 15,000-pound load
at 45 feet per minute and has a static lift rating of 20,000 pounds. A
7/8 diameter (.923 actual) nonrotating cable of 18 x 7 construction with
individual wires .058 inch in diameter is used to support the load. The
core of the cable contains seven electrical conductors wrapped in a re¬
silient Teflon jacket. The electrical conductors are used to operate the
hook indicator lights and to power a solenoid which opens the hook. The
cable has a minimum breaking strength of 58,000 pounds. The cable is
wrapped on a drum whose basic pitch diameter is 24.5 inches, which gives
a drum to cable wire diameter ratio of 422 to 1. The drum is magnesium
to *hich a 0.25-inch-thick polyurethane rubber jacket is molded. The
nibber Jacket serves to reduce chafing, thus prolonging both drum and
cable life. Experience with this hoist in the Southeast Asian theatre of
operations for 11 months has not resulted in a single cable fatigue
failure.
Single-Point Hoist Cable Design
To meet the 40,000-pound single-point hoist capability of the H.L.H., it
will be necessary to use a cable diameter of 1.39 inches. An 18 x 19 con¬
struction is used to obtain a flexibility greater than that of the 18 x 7
■* construction used on the CH-54A.
Normal nonrotating cable construction requires the use of 18 strands of
wire with either 7, 19, or 37 wires in each strand. Therefore, cable
flexibility is increased by using a larger number of smaller diameter
wires in each strand rather than by increasing the number of strands. The
18 x 19 construction cable uses a wire diameter of 0.060 inch compared to a
I
49
1
I
I
wire diameter of 0.058 inch in the 18 x 7 cable. Using commercial prac¬
tice (for cables without conductors in the core) which requires a cable
drum diameter 400 times the individual wire diameter, it would be possible
to use a cable drum diameter of 24 inches. Use of 18 x 7 construction
cable with 7 wires of .065 diameter per strand would, by the 400 to 1 rule,
require a minimum cable drum diameter of 34 inches.
While the drum diameter selected is greater than the 24 inch minimum allow¬
able, the extra flexible construction has been retained, since it increases
the fatigue life of the cable. The determining factor in selection of the
drum diameter is the requirement to carry as much cable as possible with¬
out exceeding desirable aircraft control limitations (see Single-Point
Hoist section, page 17) rather than the minimum ratio of cable drum to
wire diameter (the 400 to 1 rule).
Multi-Point Hoist Cable Design
For the 23, 100 -pound-capacity two-point hoists, a 1.06 diameter cable is
required, and a 0.79 diameter cable is required for the 11,550-pound-capac¬
ity four-point hoists. All cables are stainless steel and are of 18 x 19
nonrotating construction. Wire size for 1.06 diameter cable is .047 and
is *Q35 for the .79 diameter cable. The use of nonrotating construction
in all of the multi-point hoists is desirable, since it will permit them to
carry separate cargo on individual hoists. The four hoists, for example,
could be rigged to cany separate fuel bags, or sling loads of amnunition,
to individual sites. All cables will contain seven electrical conductors
suitably protected by a resilient Teflon jacket in the central core. This
construction provides the maximal protection for the conductors from both
the elements and from damage due to rough handling.
No development problems are expected in the fabrication of ary of the
three cables described above.
The one major development problem to be solved is that of providing a cock¬
pit controlled mechanical release of the load that can be integrated with
the load suspension cables. Several proposed solutions to this problem
and an alternate method of providing the needed redundancy in release
methods are described in the PROBLEM AREAS AND PROPOSED SOLUTIONS section,
page 113.
CARGO HOOKS
Background
Cargo hooks used by helicopters for support of external loads have under¬
gone extensive development since the original manually released hooks were
first introduced. Capacities have increased from 2000 pounds to the
20,000 pound capacity hook presently used in both the CH-53A and the CH-54A.
Electrical release modes have been added and the automatic touchdown re¬
lease has been developed. Indicator lights for load beam attitude also
have been added.
Table XV shows the weight trend of cargo hooks presently in use. It should
be noted that the hooks rated at 20,000 vary in weight by 27.2 pounds, with
the hoist mounted hook being heavier. This increase in weight is due
solely to the requirement for a swivel-slip ring assembly and not for
other reasons such as cable straightening requirements.
Heavy Lift Helicopter Cargo Hook Design
To meet the cargo hooks requirements for hoists of this study, the basic
data shown in Table XVI, page 53, were generated and design proposals were
solicited from several manufacturers of airborne cargo hooks.
TABLE XV
HELICOPTER CARGO HOOKS
Aircraft
Type of
Suspension
Normal
Operating
Load (lb)
Ultimate
Load (lb)
Weight
(lb)
Conroe nts
H-34
Sling
5,000
25,000
10.7
H-37
Sling
10,000
50,000
27
Requires manual
relatch
CH-3C
Sling
10,000
50,000
24
CH-53A
Pipe
20,000
90,000
40
CH-54A
Hoist
20,000
90,000
67.2
Requires swivel-
slip ring assy
Proposals received from hook manufacturers indicate that all these require¬
ments can be met with no major advances being required in the state of the
art. The weight of the 40,000-pound-capacity hook swivel assembly will be
192 pounds, that of the 23j 100-pound unit will be 87 pounds, and the 11,550-
pound unit will be 59 pounds. All units will be similar in design to the
assembly presently in use on the CH-54A, shown in Figure 14, page 51. It
is the opinion of one hook manufacturer that the weight of the 40,000-*
pound-capacity unit can be significantly reduced by a change in the rela¬
tionship of the structural parts. Such a change will be actively pursued
in Phase II of this study.
Field experience with the CH-54A has borne out the fact that the swivel-
slip ring assembly is the most sensitive part of the total hook assenfcly to
environmental conditions. Such assemblies must be designed to provide the
utmost protection from the elements. In addition, they must be nigged
enough to withstand repeated abuse. Environmental testing is mandatory to
ensure that the required protection has been provided. A refinement,
currently being studied under a separate contract, is the incorporation of
an AN 4064 type dehydrator for both the hook and swivel-slip ring assemblies.
52
TABLE XVI
BASIC DATA - CARGO HOOKS
GENERAL DESIGN REQUIREMENTS
1. Open throat design of hook
2. Automatic relatch of load beam
3. No safety lock
4. Both manual and electrical release modes required
5. Swivel-slip ring assembly required to allow hook rotation*
6. Slip ring to have 7 circuits
7® Design life: 5000 full load releases •
6. Indicator signals for hook open and hook closed
9. Environment: -65°F to 130°F, sand and duet, fungus,
water immersion
10. Hook detachable from swivel; swivel detachable from
cable
11, Hook supported by a single cable with hollar core for
electrical and/or mechanical conductors.
Capacity
A
B
C
Cable Size
1.39
1.06
0.79
Max. Operating
Load
40,000
23,100
11,550
Limit Load
100,000
64,700
32,350
Ultimate Load
150,000
97,000
48,500
*Swivel assembly, combined with nonrotating cable construction,
permits individual loads to be carried on multi-point hoist
systems (see page 50).
53
This will provide field-level personnel with a method to check for mois¬
ture contamination without disassembly.
MTSTKI.T-AMBnus COMPONENTS
PiRtch Degign
All hoist distribution clutches are of the multiple-disc, wet-plate type.
They will be either hydraulically or electrically actuated. They are
spring loaded to the disengaged position. The clutches are mounted on
the angle gearboxes and are easily removable.
Traction Sheaves
All traction sheaves will be hydraulically powered and universally mounted.
The power required to drive the sheaves was established by assuming a mini¬
mal cable sag of 1 inch in 19 inches of span and a requirement that the
sheave operate at a speed 5 pet above that of the hoist during the lowering
mode of operation. Driving friction of the cable in the sheave is assured
by the use of adjustable pressure rollers. A slip clutch is used in series
with the motor. The clutch will be imnersed in oil to permit proper cool¬
ing during operation.
The sheave will be removable} both the cable cutters and the bellmouth will
be slotted to permit installation and removal of the cable. Tandam-dual
cable cutters are mounted on the bellmouth. The use of pressure rollers
and a cover for the sheave provides an effective backlash suppressor in
the event that the cables must be sheared. A typical design is shown in
figure 15, page 55.
Cable Cutters
All cable cutters are of the tandem-dual type presently installed on the
four-point hoists being developed for uso on the CH-54A.
Two cable cutters of the electrically ignited, explosively propelled knife
type are mounted in tandem on both hoist and traction sheave bellmouths.
Each cutter's explosive charge can be fired by either of the two bridge-
wire circuits. Separate routing of the wiring for each of the circuits
provides additional redundancy. All firing circuits are actuated simul¬
taneously. Attachment of the cable cutters to the bellmouth will permit
easy tear-out in the event that the lower cutter should trap the cable in
the housing instead of shearing it.
All hoists incorporate the tandem-dual cable cutter concept to meet emer¬
gency release requirements. Since the circuitry required to fire the
cable cutters bypasses the hoist and hook slip ring assemblies as well as
the load suspension cable, this concept meets all requirements for a re¬
dundant release system. All wiring to the cutters will be Installed with
adequate slack to allow free hoist movement and will be armored to pro¬
tect from accidental, damage.
54
HYDRAULIC
MOTOR INPUT
SLIP CLUTCH
SLIP RING
ASSEMBLY
The "all-fire" current will not exceed 16 amperes per hoist. Thus the
emergency release system can be operated on battery power only in the
event of the failure of both aircraft generators.
Conductor Reels
A conductor reel is required in several of the combined function config¬
urations investigated to provide a method of supplying electrical power
to the 40,000-pound-capacity cargo hook. In addition, the conductor reel
is used to stow the hook out of the way when it is not in use.
The basic design will consist of a hydraulically powered hoist capable of
storing 150 feet of electrical conduit. The conduit will have 7 electri¬
cal conductors in the core and will be suitably protected by a braided
wire jacket. A slip ring assembly will transmit electrical power from the
aircraft system to the electrical conduit. The use of a flat coil spring
to provide power for reeling in was not considered feasible because of the
150 foot length of cable required. A 1.0-horsepower hydraulic motor driv¬
ing through a slip clutch will permit the hook to be lowered without slack.
The motor, which will be synchronized to operate with the hoists, will also
operate at a speed such that no slack will be permitted in the cable during
lifting. The slip clutch will prevent the electrical conduit from lifting
more than the hook weight alone and will permit the hook to be lifted into
a storage well when not in use.
The conduit will be attached by means of its steel braided outer jacket to
the swivel assembly of the hook with a suitable electrical connector to
provide electrical power.
While such a unit is not comnercially available at the present time, de¬
sign and fabrication will present no major technical problems. Its weight
should not exceed 50 pounds. A typical design is shown in Figure 16,
page 57.
59
HOIST SYSTEM CONFIGURATIONS
Thirteen basic hoist system configurations have been investigated to meet
the heavy lift helicopter external cargo handling requirements. Seven con¬
figurations are designed to meet the single- plus four-point requirements
and six are designed to meet the single- plus two-point requirements.
Three variants of each type are combined function arrangements with the
single-point function being replaced by combined operation of the multi¬
point hoists. Tables XVII and XVIII, pages 87 and 8$ include a brief
description and summary of pertinent data on each system. Schematic
drawings of all the systems are presented in Figures 17 through 29, pages
61 through 85 . The weights given are based on 60 feet of cable for the
single-point hoist.
60
-1 CONFIGURATION
General Description
Single-point hoist mechanically driven. Hydraulically powered
zero-moment hoists for the four-point system.
System Components and Weights
Single-Point Hoist (Type A) I960 Pounds
Clutch-Reverser Unit 124
Angle Gearboxes (2 required) 43
Drive Shafts 6
Hoist Pump 35
Four-Point Hoist (Type K, 4 required) 2000
Hydraulic Motor (4 required) 40
Plumbing and oil 188
Total System Weight 4396 Pounds
Single-Point Mission Weight
(Remove four-point hoists) 2396 Pounds
Multi-Point Mission Weight
(Remove single-point hoist) 2436 Pounds
Power Required
Single-Point Mission 94.8 HP
Multi-Point Mission 84.8 HP
Figure 17
-1 Configuration
CLUTCH-REVERSER UNIT-
A.
-2 CONFIGURATION
General Description
Single-point hoist hydraulically powered by one or two motors.
Hydraulically powered zero-moment hoists for the four-point system.
System Compor.ents and Weights
Single-Point Hoist (Type A)
I960 Pounds
Four-Point Hoist (Type K)
2000
Hydraulic Pump
55
Hydraulic Motors (5 required)
84
Plumbing and Oil
188
Total System Weight
4287 Pounds
Single-Point Mission Weight
(Remove four-point hoists)
2287 Pounds
Multi-Point Mission Weight
(Remove single-point hoist
2327 Pounds
Power Required
Single-Point Mission 132.5 HP
Multi-Point Mission 84.8 HP
>
*
Figure 18, -2 Configuration.
63
-3 CONFIGURATION
General Description
Single- and four-point hoists mechanically driven.
System Components and Weights
Single-Point Hoist (Type A) I960 Pounds
Clutch-Reverser Unit 124
Angle Gearboxes (4 required) 80
Clutches (5 required) • 51
Shafts and Couplings 45
Four^-Point Hoists (Type P, 4 required) 1952
Total System Weight 4213 Pounds
Single-Point Mission Weight
(Remove four-point hoists) 2261 Pounds
Multi-Point Mission Weight
(Remove single-point hoist) 2253 Pounds
Powered Require^
Single-Point Mission 94,2 HP
Multi-Point Mission 58,0 HP
*
Figure 19. -3 Configuration,
65
-4 CONFIGURATION
General Description
Single-point hoist mechanically driven. Two mechanically driven
dual drum hoists, with cables reeved over hydraulically powered
traction sheaves for the four-point system.
System Components and Weights
Single-Point Hoist (Type A) I960 Pounds
Clutch-Reverser Unit 124
Angle Gearboxes (2 required) 42
Shafts and Couplings 29
Dual Drum Hoist (Type N, 2 required) 1980
Traction Sheaves 240
Clutches (5 required) 55
Total System Weight 4430 Pounds
Single-Point Mission Weight
(Remove dual drum hoists) 2450 Pounds
Multi-Point Mission Weight
(Remove single-point hoist) 2470 Pounds
Power Required
Single-Point Mission 94.1 HP
Multi-Point Mission 64.9 HP
Figure 20. -4 Configuration.
67
iiW'
TRACTION SHEAVES
STA
694
-5 CONFIGURATION
General Description
No single-point hoist. Hydraulically powered zero-moment hoists
for the four-point system with a frame, lockable to the aircraft
to provide single-point capability.
System Components and Weights
Four-Point Hoists (Type M, 4 required) 2200 Pounds
Hydraulic Pump 35
Frame (with slings & 40,000-lb hook) 1172
Plumbing and Oil 188
Hydraulic Motors (4 required) 88
Conductor Reel 50
Total System Weight 3733 Pounds
Single-iPoint Mission Weight 3733 Pounds
Multi-Point Mission Weight
(Remove frame with slings & hook) 2561 Pounds
Power Require d
Single-Point Mission
Multi-Point Mission 171.8 HP
BL —
70
l
\
Bl
70
WL * —
260
WL-
230
WL
50
Figure 21
-5 Configuration
-6 CONFIGURATION
General Description
No single-point hoist. Hydraulically powered zero-moment hoists
with cables joined to a common hook for a single-point capability
and reeved over hydraulically powered traction sheaves for the g L
four-point system. 7 (
System Components and Weights
Four-Point Hoist (Type L, 4 required)
2200 Pounds
Hydraulic Pump
35
Traction Sheaves
240
Hydraulic Motors (4 required)
88
Conductor Reel
50
Hook and Swivel ( 40 , 000 -lb Capacity)
190
Plumbing and Oil
188
Total System Weight
2991 Pounds
Single-Point Mission Weight
2991 Pounds
Multi-Point Mission Weight
(Remove hook, swivel,&
conductor reel)
2751 Pounds
Power Required
Single-Point Mission
Multi-Point Mission 179.1 HP
WL —
260
I
WL
50
Figure 22, -6 Configuration,
A.
71
.tv--.'
-7 CONFIGURATION
General Description
No single-point hoist. Hydraulically powered zero-moment hoists
for the four-point system. Cables joined by a master hook
carried by one of the four-point i o Bts to provide single-point
capability.
System Components and '//eights
Four-Point Hoists (Type L, 4 required) 2200 Pounds
Hydraulic Pump 35
Hydraulic Motor 88
Plumbing and Oil 188
Hook and Swivel (40,000-lb Capacity) 192
Conductor Reel 50
Total System Weight
Single-Point Mission Weight
Multi-Point Mission Weight
2753 PoundB
2753 PouniB
2753 Pounds
Power Required
Single-P.iint Mission
Multi-Point Mission
171.8 HP
Figure 23. -7 Configuration,
73
ACCESSORY GEARBOX
4
-11 CONFIGURATION
General Description
Single-point mechanically driven. Hydraulically powered zero-
moment hoists for the two^point system.
System Components and Weights
Single-Point Hoist (Type A)
I960 Pounds
Clutch-Reverser Unit
124
Two-Point Hoist (Type G, 2 required)
2004
Hydraulic Motors (2 required)
44
Angle Gearboxes (2 required)
43
Hydraulic Pump
35
Shafts and Couplings
6
Plumbing and Oil
174
Structure
40
Total System Weight
4430 Pounds
Single-Point Mission Weight
(Remove two-point hoists)
2426 Pound8
Multi-Point Mission Weight
(Remove single-point hoists)
2470 Pounds
Power Required
Single-Point Mission
94.0 HP
Multi-Point Mission
86.3 HP
Figure 24. -11 Configuration.
75
A*
-13 CONFIGURATION
General Description
Single- and two-point hoists mechanically driven.
System Components and Weights
Single-Point Hoist (Type A) I960 Pounds
Clutch-Reverser Unit 124
Two-Point Hoists (T^pe F, 2 required) 2044
Angle Gearboxes (4 required) 83
Shaft and Couplings 45
Clutches (3 required) 48
Structure 40
Total System Weight
4344 Pounds
Single-Point Mission Weight
(Remove two-point hoists) 2300 Pounds
Multi-Point Mission Weight
(Remove single-point hoist) 2384 Pounds
Power Required
Single-Point Mission 94.0 HP
Multi-Point Mission 61.1 HP
Figure 25. -13 Configuration
-13 CONFIGURATION
General Description
Single- and two-point hoists mechanically driven.
System Components and Weights
Single-Point Hoist (Type A) I960 Pounds
Clutch-Reverser Unit 124
Two-Point Hoists (Tjrpe F, 2 required) 2044
Angle Gearboxes (4 required) $3
Shaft and Couplings 45
Clutches (3 required) 48
Structure 40
Total System Weight
4344 Pounds
Single-Point Mission Weignt
(Remove two-point hoists) 2300 Pounds
Multi-Point Mission Weight
(Remove single-point hoist) 2384 Pounds
Power Required
Single-Point Mission 94.0 HP
Multi-Point Mission 61.1 HP
WL-
260
WL —-
230
WL
50
Figure 25. -13 Configuration.
A.
77
-14 CONFIGURATION
General Description
Single-point hoist is a dual drum type mechanically driven.
Cables are reeved over hydraulically powered traction sheaves
for the two-point system.
System Components and Weights
Dual Drum Single-Point Hoist ('^’ype D) 2138 Pounds
Clutch-Reverser Unit 124
Hook and Swivel (40,OOO-lb Capacity) 192
Traction Sheaves (2 required) 120
Conductor Reel 50
Clutches (2 required) 32
Angle Gearboxes (2 required) 36
Shafts and Couplings 6
Plumbing and Oil 24
Structure 40
Total System Weight
2762 Pounds
Single-Point Mission Weight 2762 Pounds
Multi-Point Mission Weight
(Remove 40,000-lb
capacity hook) 2570 Pounds
Power Required
Single-Point Mission
Multi-Point Mission
WL-
230
90.6 HP
63.8 HP
V
WL
50
Figure 26. -14 Configuration.
-15 CONFIGURATION
General Description
No single-point hoist. Hydraulically powered zero-moment hoists
for the two-point system with a beam, lockable to the aircraft,
tc provide single-point capability.
System Components and Weights
Two-Point Hooks (Type H, 2 required)
2144 Pounds
Suspension Beam with Hook & Slings)
1106
Conductor Keel
50
Hydraulic Pump
55
Hydraulic Motors (2 required)
44
Plumbing and Oil
174
Structure
40
Total System Weight
3613 Pounds
Single-Point Mission Weight
3613 Pounds
Multi-Point Mission Weight
(Remove beam with hook & slings)
2507 Pounds
Renuired
Single-Point Mission
Kulli-Point Mission
174.8 HP
WL-
260
WL—-
230
WL
50
Figure 27. -15 Configuration.
81
ACCESSORY
CONDUCTOR
REEL
GEARBOX —'
HYDRAULIC PUMP
sta
406
S'
5:
\
zf
'kztt f
£
-17 CONFIGURATION
General Description
No single-point hoist. Mechanically powered conventional
hoists with cable joined to a coraaon hook to provide single -
point capability.
System Components and Weigh* s
Two-Point Hoists (Type F, 2 required)
2144 Pounds
Clutch-Reverser Unit
124
Angle Gearboxes (4 required)
60
Shafts ar.i Couplings
38
Clutches (2 required)
32
Conductor Reel
50
Hook and Swivel (40,000-lb Capacity)
192
Structure
40
Total System Weight
2680 Pounds
Single-Point Mission Weight
2680 Pounds
Multi-Point Mission Weight
(Remove 40,000-lb capacity hook)
2488 Pounds
Power Required
Single-Point Mission
Multi-Point Mission 123.9 HP
Figure 28. -17 Configuration
-18 CONFIGURATION
General Description
No single-point hoist. Mechanically powered conventional hoists
with cables reeved over hydraulically powered traction sheaves
and joined to master hook for single-point capability.
System Components and Weights
Two-Point Hoists (Type F, 2 required)
2144 Pounds
Clutch-Reverser Unit
124
Traction Sheaves (2 required)
120
Hook and Swivel (40,000-lb Capacity)
192
Angle Gearboxes (4 required)
50
Shafts and Couplings
40
Clutches (2 required)
16
Plumbing and Oil
6
Structure
40
Total System Weight
2732 Pounds
Single-Point Mission Weight
2732 Pounds
Multi-Point Mission Weight
2732 Pounds
Powered Required
Single-Point Mission
—
Multi-Point Mission
125,1 HP
Figure 29 -18 Configuration.
CLUTCH -REVERSER UNIT
CLUTCHES
TABLE XVII
SUMMARY
SINGLE-POINT PLUS FOUR-POINT
LOAD SUSPENSION CONFIGURATIONS
Config,
Number
Description
Total
System
Weight*
(lb)
Drive
Hoi
Typ
-1
Single-point hoist mechanically driven. Hydrau¬
lically powered zero moment hoists for the four-
point system.
4396
Mech.
Con
-2
Single-point hoist hydraulically powered by one
or two motors. Hydraulically powered zero-
moment hoists for the four point system.
4287
Hyd.
Con
-3
Single- and four-point hoists mechanically
driven.
4213
Mech.
Cor
-4
Single-point hoist mechanically driven. Two
mechanically driven, dual drum hoists with
cables reeved over hydraulically powered
traction sheaves for the four-point system.
4430
Mech.
Con
-5
No single-point hoist. Hydraulically powered zero-
moment hoists for the four-point system with a
frame, lockable to the aircraft, to provide
single-point capability
3733
—
-6
No single-point hoist. Hydraulically powered
zero-moment hoists with cables joined to a
common hook for single point capability and
reeved over hydraulically powered traction
sheaves for the four-point system.
2991
-7
No single-point hoist. Hydraulically powered
2753
-
-
zero-moment hoists for the four-point system,
t Cables joined by a master hook carried by one
of the four-point hoists to provide single¬
point capability.
*Add 103 pounds for single-point hoist with
Add 362 pounds for single-point hoist with
87
FOUR-POINT
FIGURATIONS
Single Point
Four
Point
Cable
System
Cable
Total
System
Hoist
Length
HP
Weight*
Hoist
Length
HP
Weight
Drive Type
(«)
Reqd
(lb) Drive
ift)
Reqd
(ft)
Mech.
Conv.
80
94.2
2396
Hyd.
Zero
Mom.
50
84.8
2436
Hyd.
Conv.
80
132.5
2287
Hyd.
Zero
Mom.
50
84.8
2327
Mech.
Conv.
80
94.2
2261
Mech.
Conv.
50
58.0
2233
Mech.
Conv.
80
94.1
2450
Mech.
Dual
Drum &
Trac¬
tion
Shelves
50
64 #9
2470
-
—
-
•m
3733
Hyd.
Zero
Mom.
80
171.8
2561
-
-
-
-
2991
Hyd.
Zero
Mom.
80
179.1
2751
2753
Hyd.
Zero
80
171.8
2753
Mom.
hoist with 100 feet of cable
hoist with 150 feet of cable
£>.
TABLE XVIII
SUMMARY
SINGLE-POINT PLUS TWO-POINT
LOAD SUSPENSION CONFIGURATIONS
m
Config.
Number
Total
System
Weight*
ilb)
Hoist
Drive Tyne
m
-11
Single-point meciianically driven. Hydrauli¬
cally powered zero-mamsnt hoists for the two-
point system.
4430
Mech. Conv.
-13
Single- and two^soint hoists mechanically
driven.
4344
Mech. Conv.
-14
Single-point hoist a dual drum type mechani¬
cally driven. Cables reeved over hydrauli¬
cally powered traction sheaves for the two-
point system.
2762
Mech. Dual
Drum
-15
No single-point hoist. Hydraulically powered
zero-moment hoists for the two-point system
with a beam, lockable to the aircraft, to
provide single-point capability.
3613
-17
No single-point hoist. Mechanically powered
conventional hoists with cables joined to a
common hook to provide single-point capability.
2680
mm m,
-18
No eingle-point hoist. Mechanically powered
conventional hoists with cables reeved over
hydraulically powered traction sheaves and
joined to a master hook for single-point
capability.
2732
Hyd. Trac¬
tion
Sheavi
*Add 103 pounds for single-point hoist wit]
Add 362 pounds for single-point hoist wit!
89
II
WO-POINT
FIGURATIONS
Single Point
Two Point
Hoist
Drive Type
Cable
Length HP
(ft) Read
System
Weight*
(lb) Drive
Hoist
-Sa?.s_,
Cable Total
Length HP
(ft) Read
System
Weight
(lb)
Mech. Conv, 80 94.0 2426
Mech.
Conv.
80
94.0
2300
Mech.
Dual
Dm
80
90.6
2762
-
-
-
3613
-
-
-
-
2680
Hyd.
Trac-
2732
tion
Sheave8
Hyd.
Zero
50
86.3
2470
Moo.
Mech.
Conv.
50
61.1
2384
Hyd.
Trac¬
tion
Sheaves
-
63.8
2570
Hyd.
Zero
80
174.8
2507
Mom.
Mech.
Conv.
88
123.9
2488
Mech.
Conv.
92
125.1
2732
nt hoist with 100 feet of cable
nt hoist with 150 feet of cable
B.
LOAD ACQUISITION AND RELEASE
SINGLE-POINT MODS
The single-point mode offers the most versatile and safest method for ac¬
quiring and releasing external loads. It is equally adaptable to ground
and hovering type pickups. Since the cargo hook is connected to the cable
by means of a swivel assembly, there is little resistance to rotation of
loads being lifted. Because of this feature, bulky loads will have to be
carried below the main landing gear; less bulky loads may be snugged up as
close as in any of the multi-point systems to permit higher forward speeds.
The effect of oscillating loads on the stability of the aircraft is mini¬
mized, since the hoist is located close to the center of gravity of the
aircraft. This is more fully discussed in the AIRCRAFT - LOAD INTERACTION
section. Hovering pickup, by any of the methods requiring the use of a
beam to convert the multipoint system to single point (as in the -5 and
-15 configurations), will present inherent hazards to both ground crew and
to vehicles during both pickup and release. Therefore, the use of these
systems is not recommended. The physical size of the hook, which weighs
192 pounds, will require that the vehicle sling be carried to the hook.
A short leader line fran the sling will facilitate this type of hookup and
enable the ground crewman to hook yp without having to climb to the top of
the vehicle to make the connection.
In-flight release of cargo will be possible by using the electrical hook
release. Use of tandem-dual cable cutters (see page 54) will ensure that
cargo can be jettisoned in the event of a malfunction of the electrical
release. Normal load release will be made after the cargo is put on the
ground. Four methods of load release are possible: two for the pilot and
two for the ground crew. The pilot can open the hook either by the elec¬
trical release, or, in the event of emergency, by shearing the cable with
the tandem-dual cable cutters. The ground crewman can open the hook with
the manual release knob or can slide the load ring off the hook by re¬
tracting the keeper on the hook in the event of a malfunction of the
manual load release. An automatic touchdown release, whereby the hook
automatically opens when the cargo is put on the ground, can be incor¬
porated. This feature adds another release mode and provides greater re¬
dundancy.
Towing by the single-point hoist requires the use of special equipment of
the type used on the CH-54A (see Figure 30, page 92) if cable loads are
expected to exceed 18,000 pounds. This limitation is detailed in the
AIRCRAFT-LOAD INTERACTION section, page 96 .
TWO-POINT MODE
The two-point mode is adaptable to either ground or hovering pickups, de¬
pending primarily upon the type of terrain for the method to be used.
Ground pickup offers the safest method and should be used whenever circum¬
stances permit.
91
' • - j -riT "nnwfwr.
Figure 30. Special Tow Gear
Two-point hovering pickups are inherently more risky them those made in
the single-point mode. This is because of the possibility that upsetting
loads could be transmitted to the aircraft if it should drift and cause
one cable to become tight before the other was attached. This risk can be
largely eliminated by reeling out enough cable to put the hooks on the
ground with adequate slack to permit attachment to the cargo. Since the
hooks weigh 87 pounds,leader lines from the cargo slings, or vehicle
attachment points, will greatly facilitate hookup. These lines can be
attached to the hooks instead of requiring the hooks to be carried to the
attachment points. Lifting cargo with a C.G. located midway between pick¬
up points from a hovering attitude will be performed with the load sensi¬
tive control energized. When cable loads are equalized, the synchronized
lift control will be engaged and the load can be snugged into position.
Beeping (individual control of hoists) is available to allow control of
cargo that has an asymetrical center of gravity. These control systems
have been fully described in the HOIST SYSTEM AND COMPONENTS DESIGN section
page 12.
In-flight trimming of loads to compensate for the effects of the aero¬
dynamic drag can be attained by engaging the load sensitive control. The
advisability of making such in-flight adjustment at other than hovering or
at very low forward speed conditions is questionable. Preliminary analysis
indicates that multi-point loads will assume a stable aerodynamic position
for reasonably adjusted cable lengths at any given forward speeu. In¬
flight changeo in cable lengths may affect aircraft stability and tend to
produce ptiching oscillations. Further evaluation utilizing wind tunnel
tests is desirable to obtain qualitative data upon which the limitations
or advisability can be based.
Normal load release will consist of synchronized lowering of the load to
the ground from either a hover or the landed position of the aircraft.
After the cables are slackened, the hooks can be opened electrically by
the pilot or manually by the ground crew. In the unlikely event of both
electrical and manual release failure, a ground crewman can slide the
load ring off the hook load beam by retracting the spring loaded keeper on
the hook. In the event that no ground crewman is on-site, it will be
necessary to keep a slight tension (a cable tension indicator is provided)
on the cables to permit the load to pull off the hook in the electrical
release mode. This is necessary since tho hook load beam is spring loaded
for automatic relatching. As in the single-point mode, a tandem- dual
cable cutter on each hoist will provide emergency release by shearing both
cables.
The addition of a cockpit controlled, manual hook release, if practical,
will not permit safe in-flight jettisoning of loads by hook release. Only
by the combination of electrical and manual hook release motion in the
cockpit could this be considered as a possible method to be used. Even
this combination of motions should not be considered as rn acceptable
means of in-flight jettisoning, since malfunction of the hook unlocking
mechanism could still occur. Only by use of the tandem-dual cable cutters
mounted on the hoists can a safe in-flight load jettison be performed.
93
1
f
"i
§
K
I
*'A
j wa w i 1 .
Tewing is feasible within the 24,000-pound capability of the aft hoist.
It is not feasible to use both hoists for greater tow capabilities because
of the difficulties of obtaining equal cable loads.
FOUR-POINT MODE
The four-point system is equally adaptable to ground or to hovering pickup.
Ground pickup offers the safest method and should be used whenever circum¬
stances permit. As in the two-point mode, it should be standard procedure
during hovering pickups to have the hooks on the ground with adequate
slack in the cables to permit their attachment to cargo. Since the hooks
weigh 59 pounds, it is possible to attach the hooks to vehicle pickup
points without the use of leader cables which are required for both the
single- and two-point systems. Use of leader cables may be desirable,
however.
Lifting cargo with a C.G. located midway between pickup points from hover¬
ing attitude would be performed with the load sensitive control energized.
When cable loads are equalized, the synchronized lift control will be en¬
gaged and the load will be snugged into position. Beeping (individual con¬
trol of hoists) is available to allow control of cargo that has an asym¬
metrical center of gravity. Lifting cargo with a C.G. that is not sym¬
metrically located with respect to the pickup points could result in the
cargo assuming an extreme angle with the ground (see Figure 12, page 46).
In-flight trinming of loads to compensate for the effects of aerodynamic
drag of the cargo can be attained by engaging the load sensitive control.
The advisability of making such an adjustment, as more fully discussed on
page 111, is questionable. Normal load release will consist of synchro¬
nized lowering of the load to the ground from either a hover or the landed
position of the aircraft. After cable slack is observed, the hooks can be
opened electrically by the pilot or manually by the ground crew. As in
the two-point system, a ground crewman can slide the load ring off the
hook load beam by retracting the spring loaded keeper on the hook in the
event of failure of the other release systems. If no ground crewmen are
available, it is necessary to keep some tension (a cable tension indicator
is provided) on the cables to permit the load to pull off the hook when
released from the cockpit. As in the single- and two-point systems, tan¬
dem-dual cable cutters provide emergency release.
To achieve maximum safety, it is recommended that in-flight load release
be accomplished by shearing the cables. It is not considered safe with
present stage of the art release systems to attempt an electrical hook re¬
lease in flight because of the inherent risk if one or more hooks fail to
open.
Should one or more of the hooks fail to open, the entire load would be
transferred to the cables supporting these hooks and could cause the air¬
craft to become uncontrollable. Since many loads to be carried may be
well within the ultimate (failing) strength of one or more of the cables,
it is not possible to assume cable failure as a backup release system.
94
In-flight release of any multi-point load should therefore be accomplished
only by use of the tandem-dual cable cutters.
Limited towing (12,000 pound maximum) can be accomplished by using the
left rear hoist. Any method of towing requiring that two or more hoist
cables be joined presents inherent load sharing problems.
95
AIRCRAFT - LOAD INTERACTION
STABILITY OF SLUNG LOADS
The stability of the load Is the major limitation on forward speed when
carrying slung loads. An oscillating or spinning load can transmit periodic
forces to the aircraft which are detrimental to performance and handling
qualities. High density spherical or cube shaped loads* such as a cargo
net filled with ammunition* are generally aerodynamically stable and do not
Impose limitations on the aircraft. Leer density* nonsynmetrical loads* such
as a helicopter fuselage* are aerodynamically unstable and require some type
of stabilizing device. Stabilization can be accomplished by multi-point
suspension or by use of a small parachute attached to the load through a
swivel Joint.
The major advantage of a multi-point suspension system is the restoring
moment it generates when the load is displaced in yaw. With either the
two- or four-point suspension system* pods can be pulled snug against the
aircraft* thus completely eliminating the yaw divergence problems. This
stability contribution of multi-point systems deteriorates as cable length
increases. Figure 31* page 97, shows the change in static directional re¬
storing stability, , with cable length for both the two- and four-
point systems with a typical load of 25*000 pounds.
- W ft-lb per degree (23)
where
W is the load* pounds
x is the longitudinal distance between cable attachment
points* feet
y is the lateral distance between cable attachment
points* feet
L is the vertical distance between the load and cable
attachment point* feet
At yaw divergence speed* the restoring torque of the system Just equals the
unstable aerodynamic moment of the load. Figure 32* page 98, shows the
variation of static directional stability of a helicopter fuselage with
forward epeed when slung at various distances below the fuselage. It can
be seen that the four-point system provides a benefit of 5 knots in for¬
ward speed over that of a two-point system. Beyond 60 knots* a stabilizing
device would be mandatory*
96
CABLE LENGTH (FT)
Figure 31•
Restoring Moment for Two-
snd Four-Point Suspension
with 25j OOO-Pound Load,
AIRCRAFT - LOAD INTERACTION
STABILITT OF SLUMS LOADS
The stability of the load is the major limitation on forward speed when
carrying slung loads* An oscillating or spinning load can transmit periodic
forces to the aircraft which are detrimental to performance and handling
qualities* High density spherical or cube shaped loads, such as a cargo
not filled with ammunition, are generally aerodynamically stable and do not
impose limitations on the aircraft. Low density, nonsynnetrical loads, such
as a helicopter fuselage, are aerodynamically unstable and require some type
of stabilising device* Stabilisation can be accomplished by multi-point
suspension or by use of a mall parachute attached to the load through a
swivel joint.
The major advantage of a multi-point suspension system is the restoring
moment it generates when the load is displaced in yaw. With either the
two- or four-point suspension system, pods can be pulled snug against the
aircraft, thus completely eliminating the yaw divergence problems* This
stability contribution of multi-point systems deteriorates as cable length
increases. Figure 31, page 97, shows the change in static directional re¬
storing stability, Nw,, with cable length for both the two- and four-
point systems with atypical load of 25,000 pounds.
» W 3 ?) ft-lb per degree (23)
where
W ie the load, pounds
x is the longitudinal distance between cable attachment
points, feet
y is the lateral distance between cable attachment
points, feet
L is the vertical distance between the load and cable
attachment point, feet
At yaw divergence speed, the restoring torque of the system just equals the
unstable aerodynamic moment of the load* Figure 32, page 9&, shows the
variation of static directional stability of a helicopter fuselage with
forward speed when slung at various distances below the fuselage* It can
be seen that the four-point system provides a benefit of 5 knots in for¬
ward speed over that of a two-j oint system* Beyond 60 knots, a stabilizing
device would be mandatory*
96
CABLE LENGTH (FT)
Figure 31 •
Restoring Moment for TWo-
and Four-Point Suspension
with 25,000-Pound Load.
Figure 32. Typical Yaw Divergence of
25,000-Pound Helicopter Fuselage*
98
The loads considered as typical are listed in Table XIX, page 101.
Although actual wind tunnel data are not available for these vehicles, it
is felt that high density loads, such as the personnel carrier and the
self-propelled mortar, will need some additional stabilisation in the
single-point mode and none in either of the multi-point modes* Based on
wind tunnel tests of store* made for the S-60 Flying Crane, the 5-ton
wrecker will require a drag parachute for stabilisation at speeds of 100
knots on all suspension systems. The 155 mm howitzer will require added
stabilization in the single-point system and little, or none, on either
of the multipoint systems.
CENTER-OF-ORAVITY SHIFT
None of the loads evaluated will present any C.G. problems for either the
single- or tandem-rotor aircraft. For the single-rotor type, all C.G.'s
are within the F.S. 526 to 574 allowable limits. With the exception of
the 5-ton wrecker on the multi-point systems, all C.G.'s are at, or near,
F.S. 550. In this case the ovarall C.G. is at F. S. 569 within allowable
limits. Similarly, for the tandem-rotor type, all C.G.'s are within the
F.S. 527 to 589 allowable limits. Except for the 5-ton wrecker on the
multipoint systems,all C.G.'s are at F.S. 557. In this case the overall
C.G. is at F.S. 576, which is within allowable limits.
TOWING CAPABILITY
Towing characteristics of the single-rotor aircraft were calculated with
the aid of a computer program. Figure 33, page 10Q shows trim control
settings for a zero skew angle (aircraft plane of symmetry parallel to and
coincident with the direction of the tow force). Experience in towing with
the RH-3A, and in particular pilot's consents, dictates maxi mum pitch and
roll attitudes of -22° and + 10° respectively. Figure 33 indicates that
there will be no difficulty in meeting the roll attitude requirement but
that the pitch attitude restriction limits the maxima tow cable tension
to 16,000 pounds. In this situation the cable angle, relative to the
earth, is 6°, tow cable length is 475 feet, and altitude is 50 feet.
Lateral and longitudinal control positions show adequate control margin;
trim positions are such that the pilot can execute a recovery in case the
tow cable is suddenly released.
An increase in tew capability would be possible if special towing gear
(see Figure 30, page 92) were fitted to the aircraft. This gear will move
the tow cable reaction point further aft and is similar to that used on
the CH-54A.
99
«>•**
kUlwJHRaliiliP
TOW FORCE - THOUSANDS OF LBS
Figure 33 • Low-Speed Towing Characteristic!
(Groat Weight 38 f OOO Pound#, C.G.
at Sta. 550, Zero Skew Angle).
100
AIRCRAFT CONTROLLABILITI
Neither the single- nor the tandem-rotor aircraft should experience any
trim difficulties during load acquisition or normal load release with the
single or multi-point systems. Figure 34, page 10^ shows trim control
positions for the single-rotor type both with and without external loads.
A parasite drag correction was made to the computer program for each load.
No pitching moment corrections were made, as the lines of action of moment
contributing forces are at, or very near, the aircraft C.G. The parasite
drag corrections are in reasonable agreement with wind tunnel tests;
estimated parasite drag corrections are shown in Table HI,
TABLE XIX
ESTIMATED PARASITE DRAG
Item No. *
_Vehicle_Parasite Drag (ft 2 )
38
155 MM Howitzer
18
49
Personnel Carrier
16.7
83
5-Ton Wrecker
49
86
Self-Propelled
33
Mortar
♦Item number in Appendix I
As shown in Figure 34, significant changes from the basic aircraft trim are
required for some loads. If the load should be jettisoned, the aircraft
would respond as if a sudden input were applied to the controls. Figures
35 end 36, pages 103 and 104, are time history relationships to the aircraft
equations of motion for the more adverse trim situations in Figure 34*
The aircraft will rise rapidly after the self-propelled mortar is jetti¬
soned from a hovering attitude and at 60 knots forward speed, as shown in
Figures 35 and 36. In both situations, a reduction in collective pitch is
necessary to prevent an excessive rate of climb.
Figures 37 and 38, pages 105 and 106, show that the most critical situations
in pitch for jettisoning of the 5-ton wrecker are in hover and at 60 knots
forward speed. As seen from these time histories, the pitch response is
controllable with the automatic stabilization equipment engaged. Due to
• the basic instability of the aircraft in pitch, jettisoning of the load
with the automatic stabilization equipment off should be accompanied by a
corrective control input.
101
CONTROL POSITION-DEGREES CONTROL POSITION-DEGREES
40000 50000 60000 70000 80000
GROSS WEIGHT-LB
40000 500'''' 60000 70000 80000
jROSS WEIGHT-LB
Figure 34 • Trim Characteristics vs Gross
Weight - Single Rotor.
102
4
Pigure 35.
8 12 16 20
TIME-SECONDS
Vertical Response to Release of
Self-Propelled Mortar in Hover*
103
Uti
4
8 12
TIME-SECONDS
16
Figure 36.
Vertical Response to Release of
Self-Propelled Mortar at 60 Knots.
_ NO CONTROL INPUT
_CONTROL INPUT AFTER
SECOND
TIME-SECONDS
Figure 37. Pitch Response to Release of
5“Ton Wrecker in Hover,
Flgur* 38. Pitch Response to Release of
5-Ton Wrecker at 60 Knote.
106
■fj
Neither single- nor tandem-rotor aircraft should exhibit any adverse trim
changes when reeling in, or out, the entire length of cable on the single¬
point hoist if cable travel limitations are kept within proper limits. In
order to keep resulting stick motion equal to or less than that corre¬
sponding to values on the CH-54A, it is necessary to require the use of a
double layer drum hoist if a 150-foot cable length is required.
VHtTICAL OSCILLATION
Divergent vertical oscillation (vertical bounce), potentially inherent in
all heavy lift helicopters carrying external cargo, is dynamically a
forced response of the aircraft's first fuselage mode and coupled cargo
moae at one times main rotor (lp) excitation frequency. Although the air¬
craft's basic fuselage mode frequency may be well removed from its lp
operating speed,the attachment of a relatively large load though a flexible
suspension cable can shift the fuselage bending mode to within the lp
ranje, as shown in Figure 39, page 108.
Depending on the ratio of the load mass to that of the fuselage, the input
parameters, and the amount of inherent system damping, the resulting fuse¬
lage response can vary from small, convergent, uncomfortable cockpit levels
to amplitudes divergent to aircraft structural integrity. Therefore, it is
necessary to analyze the coupled load suspension/fuselage bending mode
characteristics and to incorporate positive control to decouple the first
fuselage and load suspension modes.
Dynamic decoupling is achieved in the heavy-lift helicopter by incorporat¬
ing an isolator with variable stiffness characteristics as a function of
load, as shown in Figure 40, page 109. The isolator provides essentially
constant first fuselage and load suspension vertical frequencies, both well
separated from lp frequency excitation throughout the load application
range (see Figure 41, page 110). These results were analyzed by using a
Sikorsky Aircraft free vibration program on the IBM 7094 conputer using
70 degrees of freedom, as shown in Figure 42, page 112. Vertical, longi¬
tudinal, and pitch motions are included in the programing analysis.
The heavy lift helicopter isolator requirements of Figure 40 are similar
to those of the CH-54A. The isolator on the CH-54A has been flight-tested
and proven to be effective in providing the required dynamic decoupling.
Therefore, a similar isolator configuration is planned for the heavy lift
helicopter.
i
i
\
\
f
POD JETTISON
In the event of an in-flight emergency, such as loss of one or more engines,
it may be desirable to jettison a cargo loaded pod in order to increase the
probability of effecting a safe emergency landing. Such a procedure is
possible, when the pod is supported by the multi-point hoists, by shearing
the cables. Preliminary analysis indicates that there should be no re¬
sulting pod pitching problems and that the pod should clear the tail cone
107
0 100 200 300 400 500
£Um - CYCLES PER MINUTE
Figure 39• Vertical Bounce Mode Frequencies
vs Cable Length Without Decoupler.
108
MAXIMUM SPRING RATE - THOUSANDS OF POUNDS PER IN
CABLE LENGTH-FEET
EXTERNAL LOAD RANGE
10,000 LB TO
40,000 LB
on the single-rotor heavy lift helicopter. Whether the pod will clear the
main landing gear is dependent on the landing gear design and aerodynamic
forces and cannot be determined at this time. The interference effects
between the pod and the fuselage are the major unknown factors and will
have to be evaluated by wind tunnel testing before the use of in-flight
jettisoning can be considered safe.
A pod used to carry personnel requires the use of fixed fuselage pod locks
to ensure that accidental in-flight jettisoning cannot occur. If the pod
is used to carry cargo, it is desirable to permit in-flight jettisoning.
Thus it is necessary to include provisions in the locks for an explosive
bolt release. The use of replaceable nonexplosive bolts when carrying
personnel and explosive bolts when carrying cargo would introduce a serious
human factors problem. For this reason the use of explosive bolts in the
fuselage pod locks would be undesirable, hence precluding ary possibility
of in-flight jettisoning of the pod.
IN-FLIGHT ADJUSTMENT OF MULTI-POINT HOISTS
All multi-point systems have the capability of in-flight adjustment of one
or more of the cable lengths. The advisability of making such in-flight
adjustment at other than hovering or very la/ forward speed conditions is
questionable. Preliminary analysis indicates that multi-point loads will
assume a stable aerodynamic position for reasonably adjusted cable lengths
at ary given forward speed. In-flight changes in cable lengths may affect
aircraft stability and tend to produce pitching oscillations. Further
evaluation utilising wind tunnel tests is desirable to obtain qualitative
data upon which the limitations and/or advisability can be based.
*
111
'
f VERTICAL JEGREE OF FREEDOM
P PITCH DE6REE OF FREEDOM
H.L.H. PuMlAge MathflMtic*! Modal
PROBLEM AREAS AND PROPOSED SOLUTIONS
MECHANICAL LOAD RELEASE FROM COCKPIT
One of the performance objectives of this study is to provide for two
methods of cockpit controlled load release, electrical and mechanical.
The mechanical release objective presents a major problem area.
One approach is the incorporation of a hydraulic line in the central core
of the cable. The electrical conductors would then be used to replace one
of the outer strands of the cable. The conductors would be suitably pro¬
tected by wire braiding and would not support any of the loads. There
would be no loss in strength of the cable, since, in the standard non¬
rotating construction, the outer layer of strands have substantially
greater strength than the inner layer. Unfortunately, the size of the
hydraulic line in the central core would, even with 3000 psi oil supply,
give a relatively small output force. Leakage problems would have to be
eliminated, and the added complexity required would probably negate the
advantages expected by having a redundant hook release method.
Another possibility is the use of a push-pull mechanical cable in the core
with the electrical conductors woven into the outer strands as described
above. Unfortunately, the smallest available size of the cable is 3/8
diameter. Also, this type of cable does not lend itself to operation if
it is forced to maintain a helical position. Considerable design and
development work would have to be done in order to solve these two basic
problems.
A third approach is the use of a mechanical release line supported on a
separate, constant tension cable drum. A hydraulic or electrical motor
drive would provide the tension required for hook release. The mechanical
release line would be attached to the hook. The primary problem to be
expected is that of this line winding around the primary load suspension
cable. A secondary problem is that of the line becoming entangled in the
equipment to be hoisted. An added weight penalty would also be incurred,
and system complexity would be increased.
A fourth solution is the design of a cable strip-off feature. This fea¬
ture requires Incorporation of a clutch to let the load pull the cable off
the drum. In addition, some protective device iB required to prevent the
free falling load from overspeeding the gear train. While this approach
offers the most feasible method of providing a redundant method of load
release, controllable from the cockpit, it results in the lose of both
cable and Hook. It also requires design of a clutch which can be released
under full load.
Although it does meet the specific requirement for pilot operated mechani¬
cal hook release, a system utilizing radio control was also investigated.
In this system radio signals are used to operate a battery Dowered release
mechanism in the hook. Two or more separate and distinct radio signals
must be used to trigger the hook release to prevent operation by random
radio signals. In addition, the hook batteries require regular recharging
to ensure proper operation.
In summary, the development of a cockpit controlled mechanical hook re¬
lease for any hoist system is considered to be a major problem area,
whether or not it is integrated with the primary suspension cable. No
problems exist with the design and fabrication of the cable sizes required
for any of the hoists if only electrical conductors are required to pro¬
vide the power for hook actuation.
Several of the configurations under study require a separate electrical
conductor cable to provide the electrical power required to open the hook
of the single-point hoist. Such a cable would be suitably protected by a
braided wire jacket and wound on a reel. The use of a separate electrical
‘juiiductor cable eliminates the need for conductors in the core of load sus¬
pension cables when a beam is used to convert from multi- to single-point
hoist mode, as in the -5 and -16 configurations. It is more a matter of
opinion than of fact that use of a separate electrical conductor cable
offers advantages over that of conductors buried in, and suitably pro¬
tected by, the load suspension cable.
IVo alternate solutions for the requirement of a separate mechanical cock¬
pit controlled load release are proposed.
The first alternative is simply reliance on the manual ground controlled
hook release as the backup method of release. The use of a tandem-dual
cable cutter controlled from the cockpit will serve as a secondary
(emergency) release*
The other solution is to provide a cable strip-off feature in addition to
the systems described above. Although this concept can result in loss of
cable and hook (as well as load) if used under emergency conditions, it
does provide a release mode entirely independent of the normal electrical
release system for additional redundancy.
WEIGHT
All of the systems evaluated which offer separate functions for the single
and multi-point hoists (the -1, -2, -3, -4, -11, and -13 configurations
sumnarized in Tables XVII and XVIII, pages 37 and 39 ) weigh slightly more
than tho 4000-pound goal. However, the combined function systems (the -5,
-6, -7, -14, -15, -17, and -18 configurations) will still provide an appre¬
ciable weight savingB.
This separate function system weight penalty must be balanced against the
redundancy offered by having two separate, independent systems available
as well as the ability to reduce the aircraft empty weight for any specific
mission by the removal of the major components of the system not to be used.
This capability reduces the system weight well below 4000 pounds, and even
below the weight of the combined function systems. All systems which re¬
quire the combining of functions offer a total system weight below 4000
pounds. They offer no redundancy, however, and cannot be reduced in weight
114
to a noticeable degree by removal of major componente,
SYNCHRONIZATION OF MULTI-POINT HOISTS
Synchronization of the cable travel (hook position) has been considered
by several investigators to be a major problem area in a hydraulically
powered multi-point hoist system. The inclusion of an electrical feed¬
back system will limit the maximum variation in hook position to within
7 inches in a total cable excursion of 50 feet. In addition, the use of
a relatively simple one-time check-out procedure will "zero out 1 ' most of
the instrument error and further reduce the cable length variation to
approximately 1-1/2 inches (see ERROR ANALYSIS, page 148).
115
COMPARATIVE RELIABILITY AND MAINTAINABILITY ANALYSIS
A study to compare the reliability, maintainability, and unavailability
of the various configurations (single-point plus multi-point and multi¬
point alone) of the subject system has been made. Table XX gives the re¬
sults of this preliminary study. The results are valid on a relative
basis; however, on an absolute basis they are subject to considerable
variability because of the limited data available at this time.
For each configuration considered, a failure or malfunction rate is given
for both a single-point and a multiple-point mission. These rates are not
to be interpreted as abort rates (unable to complete mission) but rather
as rates of unscheduled maintenance actions. The failure rate for the
single-point missions is the rate of malfunction expected for executing a
single-point mission only. For example, with the -1 configuration, in
1000 hours of single-point mission flying, an average of 5*59 failures
would be expected. Similarly, for the multi-point mission, the failure rate
is associated with the type of mission only.
Also included in Table XX are comparative values for the maintenance man¬
hours per flight hour and unavailability for the various configurations.
Unavailability is the complement of availability. These numbers are for
the cargo handling system only.
Table XX also includes estimates of mission reliability data for the 13
external cargo handling system configurations. These data are presented
in three columns as follows:
1. The column headed "Single-foint Only" gives the abort rate
for the single-point system only. For example, in 1000
hours of single-point mission flying, the -1 configuration
would experience an average of 1.66 aborts.
2. The column headed "Multi-Point Only" gives the abort rate
for the multi-point system. For example, in 1000 hours of
four point missions, the -1 configuration would experience
an average of 3*46 aborts.
3. The column headed "Combined System" gives the abort rate for
the cargo system as a whole for a 1.76-hour and a .463-hour
mission. This assumes that either the single-point or the
multi-point system could be used for any mission. Where
there is no redundancy, the lowest abort rate for single or
multipoint suspension is applicable.
TABLE XI
RELIABILITY/MAINTAINABILITT COMPARISON
H.L.H. EXTERNAL CARGO HANDLING SYSTEMS
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*****
EVALUATION PROCEDURE
INTRODUCTION
As an aid in selecting the external cargo handling system configuration
that best meets the 40,000-pound payload requirements of the heavy lift
helicopter, an evaluation procedure employing both qualitative and quanti¬
tative factors pertinent to the system design was employed.
Productivities or costs are calculated for each of the competitive cargo
handling designs. These results are combined with qualitative factors
and are evaluated in a comparison matrix. The matrix attempts to quantify
the qualitative factors in relation to productivity or cost by assigning
relative weighting values to each of the parameters. These relative
weighting values are based on judgment. Each of the cargo handling sub¬
systems is then scored. Systems within 5 pet to 10 pet of each other are
considered to be equivalent.
DISCUSSION
This section presents a discussion of the methodology used by Sikorsky Air¬
craft to select an optimum cargo handling subsystem for the heavy lift heli¬
copter system. The classic steps in devising a good selection process are:
1. Acquire complete understanding of the operational concept
and the functional requirements of the system and subsystem.
2. Establish criteria of effectiveness and/or cost as a basis
for system selection.
3. Identify all relevant factors that are pertinent to the cost
and effectiveness of the operation.
4. Classify all relevant factors into qualitative and quantitative
categories.
5. Quantify all the factors that are possible.
6. Construct a mathematical simulation model relating all quan¬
tifiable factors to the criteria of selection. This model
could consist of a simple equation or it could consist of a
complex set of equations requiring computer processing.
7. Exercise the simulation model to determine effectiveness
and/or cost.
8. Examine all qualitative factors.
9. Evaluate and relate all factors and select optimum system.
118
One of the most important steps in the evaluation process is to acquire
couplets understanding of the system operational concents and mission
requirements. A comprehensive missions analysis is required to develop
the technical requirements for the heavy lift helicopter and its cargo
handling subsystem. These missions and operational analyses are presently
in progress at Sikorsky Aircraft, but as in all concepts, such analyses
are difficult to perform because of the uncertainty of so many of the
basic parameters* Some of the more complex factors of such an analysis
include consideration of the types of loads to be transported and their
frequency distribution, frequency distribution of operating ranges, new
loads packaging concepts and methods of suspension, logistics and main¬
tenance problems of field conversion of single-point and multi-point sus¬
pension equipment, etc.
These factors are difficult to define at this time because the crane con¬
cept is still relatively new. Even though the U.S. Amy has had over 2
years of field experience with the CH-54A in the Continental U.S. and in
S. Viet Nam, new operational concepts are formulated almost daily. This
is complicated by the fact that the heavy lift helicopter configuration
has not been fully defined.
For the purposes of this study, the mission profiles described in the
BASIC DATA section (pages 3 and 4) were used, assuming equal distribution
of single-point and multi-point suspensions loads and equal distribution
of long and short range operation. Future missions and operational analy¬
ses may modify these assumptions, and therefore changes to some of the
conclusions of the evaluation may result.
The criteria of effectiveness used to evaluate the cargo handling sub¬
system can be either productivity of the heavy lift helicopter crane
system or total system cost. The effect of the cargo handling subsystem
on the performance or cost of the total heavy lift helicopter crane system
is important, not the absolute differences of values for the cargo han¬
dling equipment alone. For example, one cargo handling subsystem may br>
more reliable but heavier than another subsystem. The more reliable sub¬
system may not be the better one, since it is possible that the gain in
system productivity due to the higher reliability may be more than offset
by the loss in productivity due to the increased weight.
Relevant factors that are considered in the evaluation process include
weight, reliability, power required, maintainability, stability, safety,
load acquisition and release time, design compromise to the heavy lift
helicopter airframe, versatility, logistics problems due to conversion,
development problems, and cost. Most of these attributes can be quanti¬
fied into productivity or cost relationships. However, factors such as
versatility, airframe design compromise, and logistics problems are non-
quantifiable at this time. These attributes will be analysed qualitatively
and will be considered on the basis of past experience and good judgment.
** -“-rriiTrfflniiiHi
119
DESIGN OBJECTIVES
As mentioned earlier, in order to evaluate the various cargo handling sub-
systems effectively, all factors pertinent to the performance effectiveness
and cost of the operations must be identified. The interrelationships and
the sensitivities of these factors are determined by combining them into a
single effectiveness paramster of either system cost or productivity (ton-
miles/hour). It was felt that system productivity was as accurate as cost *
and was much easier to use, so productivity was selected as the criterion
of effectiveness.
A discussion of the relevant factors is presented in the following para¬
graphs to ensure that quantitative ratings of these parameters as used in
the productivity formula are considered in their proper perspective and
are not accepted as unequivocal ratings.
System Weight
Calculations have been made to determine the overall system and special
mission weights for each of the alternative hoist systems under considera¬
tion. The mission weight is defined as the weight of equipment required
to accomplish a specific mission, such as a single-point load, and is
lower than the system weight in most configurations because major compo¬
nents of the single- or multi-point system could be removed prior to per¬
forming a mission.
There are several other design considerations which must be taken into
account while minimizing system weight. For example, a design for mlnlmnn
weight might not be as satisfactory as a design which permits interchange-
ability of moving parts because of the savings in maintainability and lo¬
gistics. Extra design features for ease of maintainability such as acces¬
sibility or quick disconnects can become more important than an associated
weight penalty.
System weight plays an important part in the determination of the produc¬
tivity of a configuration. Although each alternative hoist system is de¬
signed for a 20-ton capability, the aircraft performance is penalized by
system weight. This penalty may affect system effectiveness in terms of
reduced productivity if a constant gross weight heavy lift helicopter is
assumed; or if the gross weight is increased to maintain a 20-ton payload,
the penalty will be increased procurement cost.
Another consideration which should be taken into account when calculating
system weight is the type of supporting structure within the airframe.
The single- and the tandem-rotor aircraft configuration will be affected
in different ways by the many possible hoist system configurations. Ex¬
ternal placement of the motors, pumps, and hoists will affect the drag,
while internal placement will disrupt the location of fuel cells and pas¬
senger or cargo compartments; thereby, additional structural weight pen¬
alties will be required.
120
These comments have been made to show that the best system is not neces¬
sarily the one having the lowest system weight. There are structural
weights associated with each system which have only been estimated. In
addition, there are other design factors which demand a degree of sophis¬
tication resulting in slightly Increased weight.
Safety
The factor of safety is one which must be expressed in qualitative terms
rather than in quantitative terms. The design of a system may be reviewed
and rated as either safe or unsafe. Naturally, those criteria responsible
for an unsafe rating must be corrected before the design is acceptable.
One of the most critical factors considered within the category of safety
is the stability of the load. As the aircraft speed increases, the lead
stability changes and this in turn affects the controllability of the air¬
craft. Speed is one of the more sensitive items in the productivity study,
so maximum speed allowable within the limits of safety is desirable, and
that system which allows the greatest speed within safety limitations is
regarded as the best.
Each of the suspension systems is capable of carrying any type of load by
using various combinations of bridles. However, the manner in which the
load is supported has a direct bearing upon its in-flight stability.
Oscillations of a load affect control of the helicopter because of the
changing drag factor and the change in location of the center of gravity
of the load. Suspension of a load by a four-point system provides the
most stable means of support and hence allows the greatest forward air¬
speed. Precise control of the load during flight permits even greater
load stability throughout the changing attitudes of the aircraft. This
can provide the safest system capable of the highest airspeed. However,
load stability is also dependent upon its density. High density loads can
be carried satisfactorily from a single-point system; high drag-low density
loads such as downed aircraft cannot be stabilized even through the use of
multi-point hoists, and drogue chutes are required to prevent in-flight
oscillations. In these cases, safety of flight demands low airspeed;
hence, productivity is a poor measure for comparing alternate hoist con¬
figurations.
Reliability
Reliability is generally regarded as the likelihood that a given system
will function normally. With a system as sophisticated and expensive as
a heavy lift helicopter, it is important that the cargo handling subsystem
be as reliable as possible. Failure of the cargo handling system could
result in grounding the entire helicopter system.
Reliability is measured in terns of failure rates or unscheduled mainte¬
nance requirements. Thus, a multi-point system would be considered less
reliable because, with a larger number of components, there are more
chances that something will fail. A review of the type of failures should
be made because many would occur to those components which have a direct
121
association with the raising and lowering of a load. If this is true* it
might still be possible to lock the cable drums in position and utilize
the extended cable or cables as a form of bridle. This would allow utiliza¬
tion of the helicopter to perform a mission with performance somewhat de¬
creased due to load instability.
In several of the configurations studied there are duplicate systems,
single and multi-point, so that failure of one system would not neces -
sarily result in an aborted mission. Through the use of bridles or slings,
both the single and multi-point syste&s can handle all types of loads.
Thus it is necessary to evaluate the penalty of greater system weight and
higher maintainability of both e. single and multi-point system against
higher mission reliability.
Maintainability
The factor of maintainability is most easily represented by cost. The
maintenance man-hours per flight hour for each system can be measured and
converted into dollars by using standard labor rates. However, when the
factor of maintainability is to be included in the productivity formula,
it is represented as a function of availability. , The availability of the
heavy lift helicopter is penalized by the downtime due to maintenance.
This may not be completely true, however, because the scheduled mainte¬
nance and sometimes unscheduled maintenance of the cargo handling system
can be done simultaneously with maintenance of the aircraft itself. The
fallacy of this method is that in comparing two closely rated systems, the
factor of time attributed to maintainability receives the same emphasis
as other time factors such as loading time which have a direct effect on
mission accomplishment.
There are other factors to be considered in the area of maintainability
which may not appear in the availability figure b. Standardization and
conmonality of parts is important when considering the storage and han¬
dling of spares. Accessibility and vulnerability of components is impor¬
tant, particularly for those components which may be removed for weight
saving purposes for daily missions. Check-out of the system without start¬
ing the main engines and rotor blades is important, since it permits the
check-out to be completed by the ground crew instead of the pilot. Char¬
acteristics of this type are not always taken into account by quantitative
measures and as availability, but thsy should be considered in the selec¬
tion of the recommended system.
Size
The physical, size of the systems does not appear in the productivity analy¬
sis because it does not directly affect the size or weight of cargo which
cam be transported. It is important to consider the physical size of the
system when considering its installation within an airframe. Size is less
significant when the system is to be mounted on a crane type helicopter
than if it is to be mounted within am existing cabin. It is assumed that
the configurations will all be competitive in size and that it would be
feasible to install ary of the systems in an aircraft.
122
Power Requirement
It is difficult to determine the sensitivity of the power required for
several alternative systems, especially in a productivity model. If the
basis of comparison is cost, dollars per horsepower are readily assigned
and provide a relative rating. If the comparison ie to be made on the
basis of productivity, differential power required would be translated
into pounds per horsepower and the total pounds would be subtracted from
payload.
System Cost
Initial system cost is considered to be important but by itself can be
very misleading. Not only should lifetime cost be considered, but lo¬
gistic cost due to maintainability and reliability must be considered.
All of the above-mentioned design objectives can be measured as a function
of cost and can be related to the cost of the overall system operation.
In addition, there are several other qualitative factors which are not
easily assigned a dollar figure but do affect operating costs of the
system.
Qualitative Selection Factors
The preceding paragraphs discussed the design objectives assigned in the
cargo handling work statement provided by the Amy. There are other de¬
sign criteria which may be considered of less significance but which
should have some bearing on the system selection. Examples of some of
these are as follows:
Ground Handling Time
Minimum ground handling time is important during load attachment
because it has a direct effect on helicopter hover time. If the
assumption is made that this eystem would be used to carry Any
Division equipment between 10 and 20 tons, 30 pet of that equip¬
ment must be attached while the helicopter is in a hovering posi¬
tion. Hookup time then becomes a factor which could influence
the choice of a system. The time required for ground preparation
of the load, such as the bridle attachment prior to arrived of the
aircraft, should not be included. This time should perhaps be in¬
cluded with maintainability, since it directly affects availability.
Versatility
The current study has shown that the one- , two- , and four-point
systems are each capable of handling any type load through different
combinations of bridles. The degree of effectiveness with which the
different systems can handle all loads is not the sans. Therefore,
further information on the type of loads to be carried could also
influence the selection of a system.
123
The else of pods to be used for hospitals, troop transports* or
cound posts should be standardized in height and width and perhaps
in multiples of length. These could then be transferred to or from
railroad and highway vehicles. As the super transports such as the
C-5A become a reality, much Arny equipment will be packaged in con¬
tainers being specially designed for efficient loading and unloading.
These containers will remain intact from shipper to destination so
that compatibility with a helicopter handling system is important.
Similarly, there is growing interest in the Fast Deployment Logistic
(FDL) and in programs which include Amphibious helicopter Assault
Ship (UiA) plans for unloading ships by helicopter.
In addition to the pod or container, pallets must be accommodated
for aerial flight. Pallets can be used not only for loading small-
sized containers but for transporting multiple vehicles or trailers
with their associated towing vehicles. Power suspension is extremely
important in these situations to prevent loss of load during flight.
Another element of versatility to be considered in selection of a
system depends upon the desirability of splitting loads among several
destinations. A multi-point system could be used to make more than
one drop without requiring assistance from the ground for rearrang¬
ing bridles or making other adjustments. Limit ations for this type
of operation are established by helicopter center-of-gravity limita¬
tions and by the weight of the cargo being carried.
Hin—n
After considering the advantages of multi-point systems due to the
greater flexibility in manipulating the load, investigations should
be made to determine that the pilot or other crewman has the means
of controlling the load. This includes means of knowing the position
of the load and individual tensions on the cables. Knowing this in¬
formation is only part of the problem. The remainder of the problem
is providing the means of independently operating the cables so as to
reposition the load. Even with adequate visual facilities to sight
the load, displays are required in the aircraft due to the difficulty
of depth perception when the ends of the cables are more than 20 feet
below the aircraft.
Structural Design Compromises
At the outset of the study, the philosophy was to define a cargo
handling system without reference to a particular type of airframe.
It was felt that this would ensure the best type of cargo system
which then could be attached to any type of airframe. As different
rating methods were tried and different mission profiles examined,
no single factor or group of factors seemed to give any of the
alternative systems significant advantages. As various criteria
were considered, the impact of the method of installation within an
airframe became more relevant. The effect of system weight was
briefly mentioned but there are other considerations in both the
124
single- and tandem-rotor configurations which could influence
selection of a system. In addition to the structural arrangement
of these types of helicopters, each has inherent handling charac¬
teristics which will tend to influence the relative merits of the
systems. The criteria of compatibility should be considered at
least as strongly as the other qualitative criteria.
Productiv ity Arm^yw-ta
Productivity is the rate of payload delivery for a given mission. This is
the major quantitative parameter that will be used to compare the perform¬
ance effectiveness of the various cargo handling designs. The best con¬
figuration enables the heavy lift helicopter to deliver the greatest
amount of payload over a specified distance in the shortest period of time*
Productivity is a function of payload, radius of operation, availability,
mission reliability, and total time to accomplish the mission. The weight
differences of the various cargo handling designs are accounted for by
modifying the payload to penalize any cargo handling design that is heavier
than the base subsystem. The reliability and maintainability of the equip¬
ment are reflected in availability and in mission reliability. Lead
stability are reflected in cruise speed, acceleration, and deceleration of
the heavy lift helicopter. Conplexity of cargo hookup and release is re¬
flected in loading and unloading time. The equations and the assumptions
used to determine productivity are described below:
Productivity »
where
Payload
Avail.
Mis. Rel. -
t
n
h
Payload x Distance x Ava^a hllity x Mis. Rel. , .
+ iza;
capacity of the aircraft less the difference in
weight between the configuration in question and
the llghteat configuration (lightest configuration
has 20-ton payload).
assumed base availability of the aircraft less
difference in unavailability between configuration
in question and least available configuration.
Assume aircraft has .65 availability.
percentage of missions which, once initial, will
not be aborted due to cargo handling system
failure.
t 9 + n (t ? + t x + t^ + t 5 + tg + t 2 + + t$)
number of round trips
time to accelerate to cruise speed with load
t£ ■ time to accelerate to cruise speed without load
t^ " time at cruise speed with load
t^ “ time at cruise speed without load
t^ ■ time to decelerate with load
tg - time to decelerate without load
tq - time to load
tg ** time to unload
t^ - time to warm up
Hie cargo handling systems have been evaluated for single and multi-point
cargo suspension for the following missions:
12-ton transport: One round trip of 100-nauti cal-mile radius
carrying 12 tons outbound only.
20-ton transport: One round trip of 20-nautical-oile radius
carrying 20 tons outbound only.
The assumptions made in the productivity analysis are shown in tabular
fora:
1 + 4 Pt.
4 Pt.
1 + 2 Pt.
2 Pt
Time to warm up, min
3
3
3
3
Acceleration without load, n.m./aec
4
4
4
4
Cruise speed without load, kn
130
130
130
130
Angle of climb
30°
30°
30°
30°
Deceleration without load, n.a/sec
4
4
4
4
Hoist speed, fpra
(lower and raise 20 ft)
60
30
60
30
Time to load, min
1/2
2
1/2
1
Acceleration without load, n.m./eec
3
3
3
3
Cruise speed with load, kn
60
60
60
60
Deceleration with load, n.m./sec
3
3
3
3
Time to unload, min
1 + U Pt. 4 Pt. 1 + 2 Pt. 2 Pt.
1/2 2 1/2 1
Notes For loads carried single point and assuming the hookup
is made with the least number of hooks available.
The same productivity formula applies for loads carried on the multi-point
suspension and assuming the hookup is made with the largest number of
hooks available with the changes in assumptions listed below:
Hoist speed, fpm
Time to load, min
Cruise speed with load, kn
Time to unload, min
1 + U Pt. L, Pt. 1 + 2 Pt. 2 Pt.
30
2
100
2
30
2
100
2
30
1
80
1
30
1
80
1
The results of the productivity calculations for each configuration are
given in Table XU, page 129. An average productivity for these data is
also presented.
QU ALITATIVE EVALUATION
As has been previously indicated, many of the factors governing the selec¬
tion of an oxternal cargo handling system cannot accurately be quantified
and included in the previous evaluation method. These factors must there¬
fore be ranked solely on the basis of judgment and past experience. These
factors are then combined with productivity in a comparison matrix to
select the best system.
It is difficult at this point in the study to establish accurate cost in¬
formation. Therefore, to simplify preparation and evaluation, the 13 con¬
figurations have been given a relative ranking for prototype hardware and
development costs. The following factors are included in the qualitative
comparison matrix:
FACTOR
RELATIVE WEIGHTING
Productivity 20
Safety 10
Cost 4 development effort 5
Versatility 5
127
FACTOR
RRTJTTTC WTSTCHTTMt
Airframe compatibility 5
Productivity ie given a weighting of 20 because it is a function of weighty
reliability, and maintainability, as well as load stability in flight and
acquisition tine.
Safety, which includes evaluation of potential safety hazards in all opera¬
tional phases including load acquisition, flight, and load release has
been given a weighting of 10. In ranking the various configurations, the
single-point load suspension has been considered the safest for flight
conditions, since triple redundancy is provided (tandem-dual cable cutters
and cable strip-off) for emergency jettisoning. Two-point and four-point
suspensions are considered to have approximately equal safety characteris¬
tics. While the four— point has a lower overall cable cutter reliability
than the two-point, a four-cable suspension is considered to offer suffi¬
cient redundancy to offset this factor.
The cost of prototype hardware and development effort, versatility, and
the compatibility of the system to the airframe have been given weighting
factors of 5.
A summary of the results of the qualitative comparison matrix for the
heavy lift helicopter is presented in Table XXII, page 130. Configurations
with scores within 5 pet of each other aim considered to be approximately
equivalent.
128
130
V
SUMMARY
All 13 configurations studied during Phase I have been evaluated with re¬
spect to the following parameters:
1. Weight
2. Reliability
3. Safety
4. Maintainability
5. Physical Sine
6. Power
7. System cost
8. System development effort
9. Technical confidence
10. Redundancy
11. Versatility
12. Airframe compatibility
On the basis of this Phase I investigation, an external cargo handling
system incorporating a mechanically powered single-point hoist and hydrau¬
lically powered multi-point hoists is considered to best meet the require¬
ments of a 40,000-pound-payload heavy lift helicopter.
For a single- plus four-point system the -1 configuration
is recommended.
For a single- plus two-point system the -11 configuration
is reconmended.
The major factors leading to this selection are as follows:
1. While the -1 and -11 separate function configurations are
heavier than the combined function systems and exceed the
4000-pound goal by approximately 10 pet as a total system,
the capability of major component removal results in a
single function weight of approximately 60 pet of the de¬
sign goal.
2. Separate function systems such as the -1 and -11 configurations
provide better aerial cargo system availability, since the mal-
131
function of one system does not necessarily down the aircraft.
The single and multi-point systems of these two configurations
have entirely separate power sources. This provides greater
drive system reliability than that of the separate function
systems.
3. A single-point load suspension is considered the safest for
flight conditions, since the normal electrical release mode is „
backed up by an emergency system with triple redundancy. Even
total malfunction of ell release modes does not inherently pro¬
duce aircraft instabilities, as in the case of partial failure
of any of the multi-point systems combined to perform single¬
point missions.
A. The -1 and -11 configurations require the least development
effort to achieve the desired goals of this program. In addi¬
tion, the proposed clutch-reverser unit for the mechanical drive
single-point hoist can be replaced by either the variable speed
drive unit or a hydraulic motor if these systems prove to be
advantageous as the state of the art is advanced.
5. The -1 configuration was selected rather than the -2 primarily
because it utilises different power sources for the singlo-
and the multi-point hoist systems. A hydraulic motor, pump, or
clutch failure would not prevent use of the aircraft for a cargo
handling mission. In addition, both the pump and the motor pro¬
posed for the -2 configuration require further developmental
effort.
6. The -1 and -11 configurations can be adapted to either the single-
or the tandem-rotor aircraft. They are easily adapted to the
single-rotor type because of the proximity of the accessory gear¬
box and the single-point hoist. Addition of a gearbox between the
drive shafting connecting the forward and aft rotors is required
to provide a power source for the clutch-reverser unit on the
tandem-rotor aircraft.
The -1 configuration utilising a single-point hoist and four multi-point
hoists is considered the better of the two, especially for the single¬
rotor aircraft, because of its versatility and compatibility with air¬
frame and structure.
1. For short-range missiona, the single-point load suspension is
the most versatile. It requires lower hookup and release times
and has higher hoisting speeds. Analysis indicates that the
four-point system has only a slightly better aircraft cruise *
■peed advantage over the two-point system for missions of
longer range. However, the four-point load suspension has
better versatility, since no special bridle or rigging is re¬
quired to carry a wide variety of vehicles, as is the case for
any two-point system.
132
The -1 configuration is more compatible with the airframe
structure, elnce it requires only one well for the single-
point hoist installation. The single-point hoist is partic¬
ularly adaptable to the single-rotor helicopter, since it is
located directly under vjk! utilizes the main transmission
support structure. The four-point hoists are located out¬
board of the fuselage structure on davit type structures
with suitable aerodynamic fairings.
The -11 configuration with a single- plus two-point load
suspension requires three airframe wella. In addition to
a small airframe weight penalty, these wells utilize the
crane-type fuselage cavities normally used for fuel tankage.
PHASE II
PRELIMINARY DESIGN
DISCUSSION
The external cargo handling system selected for the preliminary design
phase of this study is a single- plus four-point arrangement utilizing a
mechanically driven single-point hoist and four hydraulically driven zero-
moos nt hoists as shown in Figures 68 and 69, pages 251 and 253*
The single-point hoist is capable of raising and lowering a 40,000-pound
load 100 feet at a rate of 60 feet per minute. The single-point hoist
drive train consists of a clutch-reverser unit, two angle gearboxes, and
several sections of shafting. On the single-rotor heavy lift helicopter,
the clutch-reverser unit is mounted on the accessory gearbox (Reference 3>
Figure 58, page 245) permitting operation on the ground from the auxiliary
power plant with the rotors stationary as well as from the rotor system in
flight.
The four-point hoists are rated at 11,550 pounds at a speed of 30 feet per
minute and contain 50 usable feet of cable. These units are hydraulically
driven by a pump mounted on the accessory gearbox and utilize a hydro
electric feedback control system to ensure synchronized load lifting. The
four-point hoists can be operated on the ground without the rotors turning
when the accessory gearbox is driven by the auxiliary power plant.
Both the single-point and the four-point hoists are readily removable from
the aircraft when missions requiring minimum empty weight are to be uhdsr-
taken.
For the tandem-rotor heavy lift helicopter, the single-point hoist has
been located with the drum axis parallel to the aircraft longitudinal
centerline (B.L. 0 as shown in Figure 70, page 255), This arrangement was
selected to accomodate the somewhat smaller lateral trim moment capability
of the tandem. Since the hoist is driven from the interconnecting shaft¬
ing, it cannot be operated from the APP on the ground with the rotors
locked.
The investigation conducted herein, as well as the experience gained on
the CH-54A Sky crane, has clearly demonstrated to the authors that the basic
preliminary design of ary external cargo handling system is interrelated
and should be conducted concurrently with that of the If die airframe.
This is particularly true for the mechanically driven dingle-point hoist
system selected. This system is undoubtedly more suit „d to the single-
rotor heavy lift crane type helicopter (where the authors' experience lies)
than to the tandem. Several other single-point hoist designs, although
somewhat more complex, might be more appropriate for the tandem in view of
its lateral trim limitations. Included among these are:
134
Large diameter, narrow width multi wrap drum
Capstan type, zero-moment
Traversing drum
Two part, doubled reeved
If rotor locked ground operation is mandatory, the use of hydraulic power
will most likely result in an appreciable weight savings at a slightly
lower mission reliability (see Table XX, page 117)•
The hydraulically powered four-point hoist system is equally adaptable to
both tandem- and single-rotor helicopters. The structural and hydraulic
system design problems associated with hoist mounting drive and control
are similar for both aircraft. The four-point hoist installation for the
tandem-rotor heavy lift helicopter is shown in Figure 70, page 255*
%
SINGLE-POINT HOIST SYSTEM
HOIST
The single-point hoist. Figure 71, pegs 257, is a one-part, single-reeved
type. The cable is wrapped in a single layer on a grooved drum, which is
coated with a 0,25-inch-thick polyurethane rubber jacket and is attached >
to one end of the drum by a pin and clamp tjpe fitting. Even winding of
the cable on the drum is maintained by a level wind mechanism consisting
of a ball screw driven bellmouth assembly. The level wind structure is
free to rotate about the drum axis so that towing loads, with the cable
swing 60° aft, will not result in high loads being reacted through it.
In addition, the level wind contains scrub rollers which provide the power
required to pull the cable off the drum without the need for an added
weight on the hook. The scrub rollers are adjustable to compensate for
wear.
Limit switches are also mounted on the level wind. AH gearing required
to power the ball screw and scrub rollers is supported in one level wind
support arm. The bellmouth assembly is mounted on the ball screw nut and
a reaction pipe. The reaction pipe provides the restraint required to
transmit the rotary action of the ball screw into linear motion of the
bellmouth-nut assembly. It also reacts the moment created when lateral
cable swing is encountered. The bellmouth assembly contains the cable
cutters which are used to shear the cable in the event of an emergency.
A split, easily removable wear liner is also contained in the bellmouth
aase^>ly.
A Weston brake (Reference 6) is provided to prevent the load from dropping
in the event of a power failure. It functions automatically and is lo¬
cated between the first and second stages of the hoist reduction gearing.
A hydraulically powered free reel clutch assembly is built into the hoist
gearbox to permit the hook and cable assembly to be jettisoned in the
event of a hook release malfunction. A manual control valve actuated by
a push-pull control cable from the cockpit diverts flow from the aircraft
utility system to the free reel clutch assembly. It is reusable and re¬
sets automatically when the hydraulic pressure is removed. The Weston
brake is located between the input shaft and the free reel clutch assembly
so that it does not affect proper free reeling operation.
A completely sealed slip ring assembly it provided to permit transfer of
electrical control and power signals to the electrical conductors in the
core of the main suspension cable from the e rcraft's electrical system.
A quick disconnect is provided to permit fleui removal of the cable with- «
out affecting the slip ring assembly.
A conventional gear reduction system with an overall reduction ratio of
443 to 1 is used. The first and second reduction stages of 4*190 to 1
and 3*513 to 1 are conventional spur gears. The third stage of reduction,
30.111 to 1, is a compound planetary tystem. All internal gearing is
136
splash lubricated and completely sealed. Oil fill, drain, and level plugs
are provided.
Auxiliary drives are conventional spur gears. These gears are coated with
a dry film lubricant and in addition are lightly coated with grease. Tbs
feears are enclosed by weathertight fiber glasB shields which are easily
removable for inspection and servicing. A shear joint is provided to pro¬
tect the ball screw assembly in the event that excessive side loads should
tend to jam the level wind assembly. Replacement of the shear pin, follow¬
ing the removal of the cause of jamming, is possible without disassembly
of the hoist.
A cable length potentiometer is provided. It is easily accessible both
for maintenance and/or replacement. An anti-backlash cover is provided
to prevent the cable from jumping off the drum in the event a load is air
dropped.
CLUTCH-REVERSER UNIT
A clutch-reverser unit mounted on the accessory drive gearbox provides the
power to operate the main hoist. It is a conventional reversing gear
similar to the type comnonly used in marine applications. The unit con¬
sists of five spur gears, two oil actuated wet plate clutches, and a bevel
gear set as shown in Figure 72, page 259* The cil used to engage the
clutches and to provide clutch plate coolant when the hoist is inoperative
is supplied by a gear pump through appropriate solenoid operated valves.
This clutch control-lubrication pump is mounted on the accessory gearbox.
The clutches are spring loaded in the disengaged position. To hoist the
load, oil at 250 psi pressure is supplied to clutch number 1. Clutch
number 2 remains disengaged. To lower, the sequence is reversed. When
hoist operation is not required, bouh clutches numbers 1 and 2 are dis¬
engaged. The hoist drive shafting is therefore stationary except when
loads are being raised or lowered. Smooth clutch engagement is attained
by a modified form of pressure modulation accomplished by use of an orifice
between the accelerator and force cavities.
When the clutch is engaged, pressure oil enters the accelerator cavity.
Since the accelerator cavity displacement is small, the pressure drop is
. momentary. The clasping force is then generated by a controlled pressure
buildup in the force cavity created by the metering of a small amount of
oil required through the orifice in the accelerator piston. A schematic
showing this operation is shown in Figure A3.
» The mechanical variable speed drive unit of Appendix II, page 2A5, was in¬
vestigated as an alternate means of providing power to and reversing di¬
rection of the single-point hoist.
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237
INPUT
CABLE
The single-point hoist cable is stainless steel and is of 18 x 19 non¬
rotating construction. It has an outside diameter of 1.39 inches and a
guaranteed minimum breaking strength of 150,000 pounds. The individual
wires are .060 inch in diameter and are made from heat treated type 302
stainless steel. The 18 x 19 construction, 18 strands with 19 wires per
strand, gives a flexibility greater than that attained in the 18 x 7
cable presently used in the CH-54A main cargo hoist.
The core of the cable contains seven electrical conductors of stranded
wire construction. They are helically wound and encased in a tough, re-
silent, plastic jacket. Five of these conductors are required to trans¬
mit the electrical power required to operate the actuating solenoid in the
cargo hook. Two conductors are spares; thus, field level maintenance per¬
sonnel can "wire out" defective conductors without removing cable. This
"wiring out" is possible by using the following procedure:
(a) Remove hook from swivel assembly and slip ring assembly
from swivel.
(b) Remove slip ring assembly from hoist by pulling only far
enough out of housing to service.
(c) Check continuity to determine good conductor.
(d) Connect good conductor to proper terminals after
removing defective corductor.
(e) Reinstall slip ring in hoist.
(f) Reinstall slip ring in swivel assembly and hook into
swivel.
(g) Check for normal hook operation.
firibedding the conductors in the core of the cable protects them from
damage due to rough handling and from adverse environmental conditions.
Three design features have been incorporated in the single-point hoist
design to provide good cable fatigue strength. Abrasion type wear, which
results from cable slippage on the drum during the starting cycle, will be
appreciably reduced by the use of a hard rubber jacket molded onto the
aluminum drum; an added advantage is that drum wear is also reduced. In
addition, a drum diameter greater than the minimum permissible based on a
standard wire diameter to drum ratio is used. This results in a stress
level in the cable, duj to winding on the drum, which is 42 pet lower than
that which would result if the minimum permissible diameter of 24 inches
were used. By the use of a single layer design on the drum, the abrasion
wear which would occur at cable crossovers if a multiple layer design were
used is completely eliminated.
139
The estixui^ed stretch in the cable, based on preliminary tests conducted
on a 5-foot sample length of 18 x 19 cable, is equal to 1.53 x 10~° inches
per foot of cable extended per pound of load or at maximum load and cable
length:
Cable stretch - (1.53 x 10" 6 ) (100) (40,000) - 6.22 inches (26)
MTST.KTJ.AHraxjs COMPONENTS
Hook-Swivel Assembly
The hook-swivel assembly provides the means for attaching cargo to the
single-point hoist cable. It permits free rotation of loads about the
cable centerline and transmits electrical control and indicator signals
from the hoist cable to the hook.
The assembly consists of a swivel assembly and a hook assembly. These
assemblies are integral units and can be quickly separated for either
maintenance or replacement.
Swivel
The swivel asseobly consists of a main housing assembly and a slip ring
assenfciy. The main housing is threaded to accept the cable and contains
a sealed grease packed thrust bearing. Two lugs permit attachment to the
hook assembly. The slip ring assembly, which is a completely sealed unit,
is bolted into the lower part of the housing. An electrical conduit, with
suitable shielding to prevent wear due to rough usage, connects the slip
ring to the hook assembly. An AN 4064 type dehydrator, suitably protected
from abuse, is also installed to indicate excessive amounts of moisture
contamination without the need for disassembly. The complete assembly
weighs approximately 48 pounds.
Hook
The hook assembly is of the open throat or self-loading design. A toggle
linkage is used to lock the load beam in position. A rotary solenoid pro¬
vides power to open the locking linkage which allows the load to open the
load beam. The load beam is restrained, following load release, by a re¬
placeable rubber bumper. The beam is returned to the locked position by
an integral spring. Integral microswitches are incorporated to provide
the signal required to indicate load beam position. A compression spring
mounted at the load beam pivot point and a microswitch provide the means
to initiate the automatic touchdown release mode. While the load is being
supported on the load beam, the spring is compressed and the microswitch is
not actuated. When the load on the beam is reduced below 150 pounds (by
putting the load on the ground), the spring extends and the microswitch is
actuated. This signal causes the solenoid to open the locking linkage and
thereby release the load.
The solenoid is completely enclosed and sealed in a weathertight housing.
140
All micro switches are completely sealed. An AN 4064 type dehydrator, suit¬
ably protected from abuse, is installed to indicate excessive amounts of
contamination without the need for disassembly. A manual release control
knob is mounted on the outside of the hook. It pemits manual opening of
the same locking linkage as the solenoid and provides a redundant hook re¬
lease method. A knob design is used instead of a lever, since it is less
subject to accidental fouling from slings and other load attachment type
equipment.
Two handles, one on each side, ire provided to facilitate ground handling
of the hook. A rugged keeper is spring loaded to permit load rings to be
slid onto the load beam with ease and yet to prevent them from sliding off
under adverse loading conditions. The complete hook assembly weighs 102
pounds.
Decoupler
The single-point hoist decoupler (or isolator) is a nonlinear spring for
which the spring constant increases as the load increases. By varying the
stiffness of the spring with load, it is possible to maintain a constant
natural frequency for the load and cable system. This system frequency is
sufficiently removed from the one per revolution of mean ratio excitation
frequency to eliminate any tendencies of vertical oscillation (vertical
bounce).
A liquid spring is used in this application. It is provided with a 10,000-
pound preload and low friction seals to minimize the operational coulomb
friction (less than 5 pet of the applied load) and breakaway force (less
than 250 pounds).
The incorporation of a reentrant short pipe type orifice in the liquid
spring retards the return stroke of the isolator when a load is air dropped.
Ibis eliminates the use of shock struts and causes a consequent reduction
in both weight and complexity.
An integral load cell permits cable loads to be measured. A charging cyl¬
inder, pressurized by the aircraft's utility hydraulic system, compensate*
both for temperature induced pressure changes and any leakage that may
occur.
CONTROLS AND INDICATOR SYSTEM
Single-point hoist and hook controls are available to both pilots and to
the hoist operator (aft-facing pilot).
The master control for the hoist is located on the console between the
pilots. It consists of a master switch, which energizes either the single-
or four-point hoist system, and a station selector switch, which permits
hoist operation by either pilot, copilot, hoist operator, or all three. A
three-position rocker switch located on all three collective pitch sticks
allows the hoist to be raised, lowered, or stopped. This switch is spring
loaded to the off position. A guarded cable shear switch located in the
overhead console between the pilots permits the cable to be cut in the
event of an emergency.
The free reel release lever located in the center console on the floor be¬
tween the pilots permits the cable and hook assembly to be jettisoned in
the event of an emergency. The master switch ensures that only the appro¬
priate hoist system will be capable of free reeling.
A cable cutter test panel is located in close proximity to the shear
switch. It can be used to permit preflight checking of the firing circuit.
A similar shear switch and a test panel are located on the bulkhead to the
right of the hoist operator. A free reel release lever is also provided.
Cable length and cable load indicators are provided for both the pilot and
hoist operator.
The master control for the cargo hook is located in the center console be¬
tween the pilots. It contains a station selector switch which permits
hook operation by either pilot, copilot, or hoist operator, and a mode se¬
lector switch which can be placed on ELEC. REL., AUTO. REL., or SAFE.
Pv«h- V ' , ’ttr" switches on all three cyclic control sticks permit the hooks
to be opened.
Lights in the main advisory panel and at the hoist operator's station in¬
dicate when the hook is in the AUTO REL. condition or when it is in the
(FEN position.
142
FOUR-POINT HOIST SYSTEM
HOIST
The four-point hoists (Figure 73 , page 26l) are universally mounted and
are of the one-part, single-reeved type. The universal mounting, as
shown by Figure 69, page 253* permits the hoist to be pivoted through a
cone with an included angle of 60°. This enables attachment to loads of
the wide variety of physical dimensions that are within the load carrying
capability of the H.L.H. Figure 3 , page 13, shows the space envelope for
the load attachment points that are within the hoists capability without
exceeding the 60° included cone angle.
Since the hoist is mounted in a universal joint, attwII side loads will be
reacted in the level wind support structure. Figure 44 illustrates this
condition. To reduce wear of both cable and bellmouth, two nonpowered,
polyurethane rubber coated rollers are integrated into the bellmouth
assembly.
Figure 44. Four-Point Hoist Attitude at
Cable Extremes.
143
The “able Is wrap pad In a single layer on a grooved drum which ie covered
by a 0.25-inch-thick Molded rubber jacket which is attached to the drum
by a pin and clamp fitting. A level wind mechanism, consisting of a ball
screw driven roller bellmouuh assembly, ensures jven distribution of the
cable across the drum. Scrub rollers are used to provide the tension re¬
quired tj pull the cable off thj drum and are adjustable to compensate
for wear.
The power required to drive the scrub teller and the ball screw is supplied
by a chain drive from a drum mounted sprocket. The roller bellmouth assem¬
bly is supported on a ball screw nut and a reaction pipe. The reaction
pipe provides the restraint necessary to transmit the rotary motion of the
ball screw into a linear motion of the roller bellmouth-nut assembly. The
cable cutters, which are used to shear the cable in the event of an emer¬
gency, are located in the rjller bellmouth assembly.
A Weston brake is provided to prevent the load from dropping in the event
of a power failure. It fmotions automatically and is located between the
first and second stages of the main drive gear train.
A hydraulically actuated fiee riel clutch similar to that used in the
single-point hoist (see page 136) is incorporated in the hoist.
A completely sealed slip ring ass-uibly is provided to transfer electrical
signals from the aircraft to the main suspension cable on the hoist. A
quick disconnect fitting permits hoist cable renjval without affecting the
slip ring assembly.
Microswitches are provided to limit cable travel, and a cable length poten¬
tiometer is provided. A feedback control system potentiometer and clutch
assembly is also integrated in the hoist drive train.
An anti-backlash cover prevents the cable from Jumping off the drum in the
event that a load is air dropped.
The power gear train consists of four reduction stages with an overall re¬
duction ratio of 514.4 to 1. The first and second reduction stages, 4*52
to 1 and 6.27 to 1, are conventional spur gears. The third and fourth
stages, 4.90 to 1 and 3*70 to 1, are conventional planetary gear sets.
All of this gearing is splash lubricated and completely sealed. Oil fill,
drain, and level plugs are provided.
HYDRAULIC jYSTEM
The hoist pump is a yoke type in which the inclination of the yoke estab¬
lishes the angle of the swashplate, to which the displacement of the pump
is proportional. When the yoke is rotated across the center, or no-flow,
position the direction of flow through the puap can be reversed. The pump
displacement is controlled by a single stage servo positioner. The signal
144
to the valve is electrical, and control power is supplied by the replenish¬
ment hydraulic system. If either the control signal or control pressure
is interrupted, the pump returns to the no-flcw position. The magnitude
and polarity of the electrical control signal determines quantity and
direction of pump flow. The control system is designed to permit seven
delivery positions: 1/4 flow, 1/2 flew, and full flew in both directions
and zero flow. The pump displacement is 1.6 cu in./rev. It is rated at
a flow of 45 gpm at 6000 rpm and weighs 20 pounds.
Motors
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The motors are of the fixed displacement type whose direction of rotation
depends on the direction of flow. They have a displacement of 0.95 cu in./
rev and an output power, at 3000 psi, of x7.5 horsepower. They weigh 10
pounds each.
Hydroelectrical Feedback System
The basic hydraulic system is a pressure demand type in which the direc¬
tion and rate of flow are established by control of the hoist pump and the
pressure developed is determined by line losses and the load. This type
of system eliminates the heat generation of a system which uses a pressure
compensated pump at anything less than full load. The system is closed
and does not contain a reservoir. Losses from the system through pump and
motor leakage are made up by a separate replenishment system. The pump
supplying the replenishment system also provides control pressure tc the
hoist punp controller and cooling flow to the hoist pump.
Synchronized operation of all four hoists is obtained by use of a hydro¬
electrical feedback system which utilizes servo controlled flow divider
valves to control the flow to individual hoist motors, as shown by
Figure 45, page 146.
The division r f flow between forward and aft and between port and star¬
board hoists is established by servo controlled flow divider valves. The
signal to the flow dividers is derived from a comparison of the signals
from rotary potentiometers on each hoist. The output signals from the
potentiometers on the forward pair of hoists supply two amplifiers. The
output of one amplifier is proportional to the algebraic sum of the signals,
ti*e output of the other is proportional to the algebraic difference between
the signals. The potentiometers on the aft pair of hoists supply two
similar amplifiers. The outputs from the sunning amplifiers of both for¬
ward and aft pairs are supplied to another amplifier, the output o. which
is proportional to the difference between the two supply signals. The
output of this amplifier controls the flow divider valve which determines
the distribution of flow between the forward and aft hoist pairs.
Division of a flow between port and starboard hoists, foxvard or aft, is
effected by a flow divider valve which is controlled by the signal from
the differencing amplifier of the appropriate hoist pair.
During lowering of a load the signal to the flow divider valve to correct
145
CLUTCH
an error must be of the opposite sign to that to correct an error when
raising a load. To achieve this, the polarity of the potentiometers is
reversed when the direction of movement is reversed.
Operation of the hoists individually is accomplished by supplying bias
signals to the two relevant flow divider valves to direct flow to one
hoist only. During this beeping operation, the output of the pump will
be l/4 of its full flow delivery to prevent overspeeding of the hoist
motor.
The forward or aft hoists may be operated in pairs by supplying a bias
signal to the forward/aft division flow divider and a centering signal
to the other approp .ate flow divider valve. During this dual*beeping
operation, the output of the pump will be 1/2 full flow delivery te pre¬
vent overspeeding of the hoist motor.
Magnetic clutches between the potentiometers and the hoists will dis¬
engage during beeping and reengage for collective operation. Thus the
eyetem can be trimmed for any load shape and the established cable lengths
will be maintained during collective operation.
The system will be somewhat load sensitive. A heavily loaded hoist will
lag initially until the bias on the flow divider valve is sufficient to
compensate for the unequal loads. The hoists will then operate at the
same nominal rate. If collective operation is stopped at this stage and
errors are recovered by beeping, the hoists will continue to operate at
the eame rate and the load will remain level when collective operation ie
resumed. This occurs because, during beeping, the potentiometer settings
and hence the signals to the flow dividers were unchanged and the resis¬
tances of the circuits remained balanced.
System Components
The flow divider valves provide the means of varying the resistance of
each hydraulic circuit to achieve equal flow ratee to each hoist regard¬
less of the load. The valve ie comprised of two stages. The first stage
is a torque motor operated, flapper-nozzle type which establishes pilot
pressures in response to the differential signals in the coils. The sec¬
ond stage ie two spool type throttle valves which are positioned by pilot
pressures from the first stage. In operation, an electrical signal to the
valve causes the second stage spool to move from the center position and
provide increased restriction to flow in one leg. The increase in re¬
striction is proportional to the magnitude of the electrical eignal.
Two sizes of valves are required, one handling a total flow of 45 gpm to
control the division of flow between forward and aft hoist pairs, the
other to control the division of flow between port and starboard hoists
and to be capable of passing 22.5 gpm total flow. Two of the latter type
are to be used. The three valves will be mounted together to make one
compact unit and to eliminate connecting plumbing.
147
The potentiometers are high precision 10 turn type with a total resistance
error of + l/2 pet, a linearity error of ,Q3 pet, and a resolution error
of .09 pet.
Electromagnetic clutches are used to drive the potentiometers. Since they
are of the magnetic particle type, they will not be affected by tempera¬
ture and humidity. Wear and adjustment requirements will be negligible,
since there it no contact between the driver and driven elements.
The main system lines will be of stainless steel with a 1-inch O.D. and a
.065-inch wall thickness. Lines to the individual hoists will also be of
stainless steel with a 3/4-inch O.D. and a .049-inch wall thickness.
ERROR ANALYSIS
A hydroelectrical feedback system (page 145) is need to provide synchro¬
nous operation of the four-point hoist system to minimize the variation in
hoist operating speeds due to differences in load, motor leakage, mechani¬
cal and hydraulic efficiencies, and errors introduced by the flow con¬
trol components. Errors are Introduced by the sensing elements, by hy¬
steresis in the control components, and by drua and cable diameter toler¬
ance.
Difference in hoist positions can be caused by the errors in the control
elements (steady state errors) or by bias in the system necessary to
accomodate different loads, leakages, or efficiencies (dynamic errors).
The latter can be removed by beeping the offending hoist (or hoists) until
the load is level (or at the desired attitude) and then continuing collec¬
tive operation. As this dynamic error can be eliminated, it will not be
coi sidered in the error analysis. The steady state error is that differ¬
ence in hoist position which cannot be sensed by the potentiometers
(resolution error), or is insufficient to actuate the flow divider valve,
and that difference in hoist position which is necessary to condensate to
linearity and total resistance tolerances of the potentiometers. The
specification of the potentiometers will permit a maximum total resistance
error of + l/2 pet, a maximum linearity error of + .03 pet, and a maximum
resolution error of + .03 pet. The flow divider valve will have a hy¬
steresis of not greater than 1 pet of its rated signal, i.e., the signal
to direct all flow to one port, and the control system is so designed that
a discrepancy of 6 inches in hoist position will produce this rated signal.
The steady state error is composed of two components: one part is depend¬
ent upon the amount of cable travel; the other is independent of cable
movement. The total resistance tolerance produces an error dependent upon
cable travel; other control system errors are independent of cable move¬
ment. In assessing the steady state error, it is unnecessary to consider
the linearity of the flow divider valves or amplifiers, as in the steady
state these components will always be operating at the same point of their
characteristic. The errors made possible by the linearity and resolution
tolerance of the potentiometers is a function of the total hoisting capac¬
ity;!. e., the error introduced by the linearity error of .03 pet is .18 inch.
146
The tolerance on total resistance produces an error of .5 pet of the move¬
ment of the hoisting (i.e., after L feet of hoisting the error could be
.005 L feet).
Total errors in the system are sumaarized below:
Error in a
ft
Percent
50-Foot Syst«
Errors
(inches)
Errors Dependent on Cable Travel:
ft
Cable drum diameter tolerance
-.045
±.006L
Cable diameter tolerance
±.020
±.0Q3L
Pot. total resistance
±.50
±.06QL
tolerance
Errors Independent on Cable Travel:
Pot. linearity tolerance
±.03
±.18
Pot. resolution
±.03
±.18
Flow divider valve dead band
-
±.06
where: L is the distance traveled in feet
With the data shewn above, it is possible to derive a formula describing
maximum cable length discrepancy between any two hoists:
Error, inches « .138 x cable travel in feet + 0.84 (27)
These data are shown graphically in Figure 46.
The total system error may be effectively reduced in service b7 the use of
a calibrated, parallel connected, trimming potentiometer* on each hoist.
A one-time check-out procedure, which requires extending the cables a full
50 feet and then reeling in to the in-limit stops, is required. The dif¬
ference in cable lengths extended after the "fastest' 1 hoist reaches its
limit stop is measured and the trimming potentiometer is adjusted to com¬
pensate for the error of the out-of-time hoist(s). This adjustment may
compensate for errors due to potentiometer total resistance tolerance and
could theoretically reduce the 7-3/4-inch error in 50 feet to approximately
3 inches. Practical considerations, such as the desire to reel in at no
load, would probably reduce the actual error to a value slightly greater
than 4 inches in 50 feet. This approach will require further investigation.
CABLE
The four-point hoist cables are stainless steel and are of 18 x 19 non¬
rotating construction. They have an outside diameter of 0.79 inch and a
guaranteed minimal breaking strength of 49*000 pounds. The individual
wires are .035 inch in diameter and are made from heat treated type 302
149
Figure U>, Difference in Cable Length m Cable
Trarel, Peur-Polnt Holst Stjrstaou
150
%
stainless steel. The IS x 19 construction, 18 strands with 19 wires per
strand, gives a flexibility greater than that attained in the 7/8 diameter
18 x 7 cable presently used in the CH-54A main cargo hoist.
The core of the cable contains seven electrical conductors of stranded
wire construction. They are helically wound and encased in a tough, re¬
silient, plastic jacket. Five of these conductors are required to trans¬
mit the electrical power required to operate the actuating solenoid in the
cargo hook. Two conductors are spares; thus field level maintenance per¬
sonnel can "wire out" defective conductors without removing the cable.
This "wiring out" procedure is identical with that described for the
single-point hoist cable on page 139. Embedding the conductors in the
core of the cable protects them from damage due to rough handling and from
adverse environmental conditions.
Three design features have been incorporated in the four-point hoist de¬
sign to provide good cable fatigue strength. Abrasion type wear, which
results from cable slippage on the drum during the starting cycle, will be
appreciably reduced by the use of a hard rubber jacket molded onto the
aluminum drum; an added advantage is that drum wear is also reduced. In
addition, a drum diameter greater than the minlMum permissible based on a
standard wire diameter to drum ratio is used. This results in a stress
lsvel in the cable duo to winding on the drum, which is 58 pet lower than
that which would result if the minimum permissible drum diameter of 14
inches were used. By the use of a single layer design on the drum, the
abrasion wear which would occur at cable crossovers if a multiple layer
design were used is completely eliminated.
The estimated stretch in the cable, based on preliminary teats conducted
on a sample length of 18 x 19 cable, is equal to 4.75 x 10~° inches per
foot of cable extended per pound of load or at maximum load and cable
length:
Cable stretch - (4.75 x 1CT 6 ) (50) (11,550) - 2.75 inches (28)
UTSTKT.TJMEinus COMPONENTS
Hook-Swivel Assembly
The hook-swivel assembly provides the means for attaching cargo to the
hoist cable. It permits free rotation of loads about the cable centerline
and transmits electrical control and indicator signals from the hoist
cabls to the hook. The use of a swivel assembly permits individual loads
to be carried on the four-point hoists. This feature will permit individ¬
ual loads, such as fuel bags, to be carried to and off loaded at separate
sites. The assembly consists of a swivel assembly and a hook assembly.
These assemblies are integral units and can be quickly separated for either
maintenance or replacement. Figure 74, page 263, shows the hook-swivel
assembly. The relative sizes of both the 40,000 pound and the 11,550
pound capacity hooks are also shown.
.. .
151
Swivel
The swivel assembly consists of a main housing assembly and a slip ring
assembly. The main housing is threaded to accept the cable end fitting
and contains the sealed, grease packed thrust bearing. Two lugs permit
attachment to the hook assembly. ▲ slip ring assembly, which is a com¬
pletely sealed unit, is bolted to the lower part of the housing. The
complete assembly is identical in design to the single-point hoist swivel
assembly described in detail on page 140. The swivel assembly weighs 21
peunds.
Hook
The hook aseeitoly is of the open throat, or self loading, design. It is
similar in design to the 40,000-pound-capacity hook used for the single¬
point hoist which is described in detail on page 140. One feature, the
automatic touchdown release, has been eliminated in the 11 , 550 -pound-capac¬
ity hook as a safety feature. Appendix III describes a typical four-point
mission and serves to illustrate why the elimination of the automatic
touchdown release is considered a safety feature. The complete hook
assembly weighs 31 pounds.
Isolator
The four-point hoist isolator (or decoupler) is a nonlinear spring for
which the spring constant increases as the applied lead increases. The
isolator consists of two air-oil accumulators, a servo valve, a housing
containing two pistons, and a load cell. Any relative motion of small
amplitude between the load and the aircraft is absorbed by the two accumu¬
late re. The acctmulators receive and release hydraulic fluid into the
chamber below the lower piston as it moves up and down relative to the
housing. The resulting spring rate, therefore, is a function of the pneu¬
matic characteristics of the accumulators. Sudden load release is damped
out by the movement of the tapered pin which is attached to the upper
piston through a fixed orifice.
The load position is maintained hydraulically by the action of the serve
valve and isolator feedback linkage.
The _ cvrulatore contain gages to permit check-out for proper precharging
and a service valve to permit recharging as required.
A separate lead cell is attached between the isolator and hoist attachment
fitting, thus permitting field replacement without disassembly of the
isolator. ?5grre 75* page 265, shows the isolator.
The complete isolator assembly weighs 38 pounds.
152
CONTROLS AND INDICATOR SYSTEM
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Four-point hoist and hook controls are available to the pilot and copilot
and to the aft-facing pilot (hoist operator). In addition, a special con¬
trol box with a coil cord extension is provided for use b 7 a dismounted
loadmaster. The master control for the hoist systems is located on the
console between the pilots. It consists of a master switch, which ener¬
gizes either the single- or four-point system, and a station selector
switch, which permits hoist operation by either pilot, copilot, aft-facing
pilot (hoist operator), loadmaster, or all four.
As in the single-point hoist system, the three-position, rocker-type switch
located on all three collective pitch sticks allows the hoists to be
raised, lowered, or stopped. The four-point hoist selector control is
located as an extension of the pilots'and the aft pilot's collective pitch
stick. The control consists of seven selector buttons. These buttons are
marked to Dermit operation of hoists number 1, 2, 3> and 4; numbers 1 and
3 (forward;; numbers 2 and 4 (aft); or all four hoists simultaneously’.
The buttons are located in precise visual orientation with the heists to
be operated, as shown in Figure 47, page 155, which is the view of the aft-
facing pilot when operating the hoists.
The cable length indicators are of the tape type and are spatially orien¬
ted with the appropriate hoist.* Cable lead indicators are of the dial
type, to avoid confusion with the length indicators, and are also located
so that they are oriented, by the eye, with the hoist whose load they in¬
dicate.
The master control for the four-point hooks is located in the center con¬
sole between the pilots. It contains a station selector switch and a hook
selector switch. The hook selector switch allows hook numbers 1, 2, 3,
and 4 or all hooks to be operated. The actuation of the hook(s) is con¬
trolled by push-button switches on the cyclic sticks.
Lights in the pilot's advisory panel indicate when the hook load beam is
open. The aft pilot is provided with similar indicator lights.
The free reel release lever (described on page 142) permits all four cable
and hook assemblies to be jettisoned in the event of an emergency. ▲
guarded cable shear actuation switch is located in the overhead console
between the pilots and permits all four cables to be cut simultaneously in
the event of an emergency. A hoist selector switch spring loaded to the
ALL position is included in the same panel. It also permits selection of
any of the individual cables for shearing in event of an emergency when
carrying individual hoist leads. A cable cutter test panel is also pro¬
vided to permit preflight checking of the firing circuit. Shear actuation
controls are located on the bulkhead to the right of the aft-facing pilot*
J
'i
I,*
*This type of instrument correlates mere naturally with cable length
than the dial type*
153
The dimounted loadmaster's control, shown In Figure 47» is a hand held
control for both hoist and hook actuation. It consists ef seven hoist
control selector push buttons and is similar to that available to the
pilot and copilot. A toggle switch pemits the desired hoist(s) to be
raised, lowered, or shut off. Four toggle switches, spring loaded to the
OFF position, permit any of the four hooks to be opened. If all four are
to be released at once, the "gang bar” is used to ensure simultaneous re¬
lease.
J
154
CONTWL
ENGINE TORQUE METER
SINGLE - POINT
CABLE LOAD INDICATOR
SWGLE- POINT
CABLE LENGTH INDICATOR
MOOt SELECTOR
9 9
O aFT M
'°o» -o
5S O'- »?
HOOK controls
LOADMASTER
CONTROL
Figure 47* Controls and Indicators, Single- and Four-Point Hoists
CABLE LOAD ifOCATO*
CABLE LENGTH INDICATOR
HOOK OPEN INDICATOR
POUR - POINT
INSTRUMENT DISPLAY
HOIST SELECTOR BUTTONS
FOUR-
POINT (
HOIST V
CONTROL
CONTROL
CONSOLE
LOAD AND STRESS ANALYSIS
INTRODUCTION
The hoist components- evaluated herein have been designed to achieve a mini
mum service interval of 3600 cycles or 200 hours of continuous operation.
One of the design objectives as set forth in the contract was that the
system shall have a minimum service interval of 1200 cycles. Defining a
cycle as reeling in and reeling out all the available cable at full load,
the minimum service interval of 1200 cycles is equivalent to 66.7 hours
of operation for both the single- and four-point hoists. The hoist bear¬
ings are selected on the basis of a minimum B-10 life of 200 hours or
static nonbrinell capacity, whichever is more critical. For the most part
bearings on the cable load side of the VJeston brake are selected on the
basis of static capacity, while those on the input side are selected on
the minimum B-10 life requirement.
The 200-hour service interval corresponds to 3600 cycles of operation as
defined above. In actual operation, the load and cable travel will often
be less than the maximum. The true service life of the hoist components
should therefore exceed 3600 actual service cycles, and 200 hours should
be used as the governing factor for overhaul.
The selection of materials for the hoist components evaluated in this
study is based on Sikorsky Aircraft's extensive test and production air¬
craft experience. All materials considered for the hoist are currently
used in similar applications on production aircraft.
DESIGN CRITERIA
Gearing
Material AMS 6260 (SAE 9310) Steel
Spiral Bevel Gears
Bending Stress, F^ =
Compressive Stress, F c =
Spur Gears
Bending Stress, F^ =
Bending Stress, F b =
Compressive Stress, F c =
30,000 psi
200,000 psi
36,000 psi (one-way bending)
32,000 psi (reversed bending)
150, OCX) psi
157
"*• ‘ *■' ■- »e
Shafting
Material AMS 6260 (SAE 9310) Steel
F^ u * 136,000 psi
AMS 5000 (SAE 4340) Steel
F^ u »= 200,000 psi
Bending Stress = 20,000 psi
Torsional Stress = 30,000 psi
Drums
erial 7075-T6 Aluminum
Compressive Stress, F ru = 72,000 psi (ultimate load)
Housings
Material AZ 91C Magnesium Casting
ZK 60A Magnesium Forging
7075 Aluminum Forging
Planetary Carrier Plates
Material 6AL - 4/ Titanium
Bending Stress, F^ = 40,000 psi
Slope, 0 « .0010 inch per inch
Bearings
All bearings used in the single and multi-point hoist components are an¬
alyzed by the methods of References 1, 2, and 13. The bearings are design¬
ed for a B-10 life of 200 hours or for static capacity, whichever is worse.
Static capacity is checked with the cable at limit load. Under these con¬
ditions, the allowable load ratings of the bearings are as follows:
p
Rotating bearing, allow = 1.25 Co
(29)
p
Monrotating bearing, allow =- 3 Co
(30)
where
Co * Basic static load capacity of bearing
Level wind drive bearings are designed for 200-hour B-10 life using loads
derived from the power necessary to drive the level wind ball screw when
the cable is at 30° side angle. Static capacity is checked using limit
load conditions and 30° side angle.
158
SINGLE-POINT HOIST
4 -
t
The sain drive train of the single-point hoist (Figure 71, page 257) con¬
sists of three stages of gearing. The 4.190/1 ratio first stage of gear¬
ing is a spur gear set whose pinion is driven through a Thomas coupling
from drive shafting connected to the accessory gearbox. Torque flows
through the input pinion and the input idler to the input gear. The out¬
put of the first stage gear drives the input side of the Weston brake.
The Weston brake ie arranged to raise the load with all components driving
as a unit, to lower the load in the reverse drive direction with the plates
slipping, and to hold the load with the power off. The output of the Wee-
ton brake is splined to a quill shaft connected to the input side ef the
free reeling clutch. This clutch has drive plates that are preloaded by
Belleville washers and contains a hydraulic cylinder that can compress the
washers in the event of an emergency, thereby releasing the drive train
and allowing the load and cable to strip off the drum. The output of the
free reeling clutch is a quill shaft connected to the second stage 513A
reduction ratio spur pinion and gear set.
The third reduction stage consists of a 30.111 A ratio compound planetary
with one fixed ring gear ana one output ri:t 2 g gear. The driving sun gear
of this planetary is integral with the second stage gear shaft. Since the
planetary plates carry no load, they are used only to position the plane¬
tary. The main drum is driven directly by the output ring gsar of the
compound planetary. All other components, such as the level wind, slip
rings, and cable length potentiometer are driven off the main drum through
spur gears.
The normally fixed input housing is isolated from the mounting structure
by bearings allowing the reaction torque to be absorbed by a liquid spring
load isolator.
Design Data
Normal cable load
Limit cable load
Held cable load
Ultimate cable load
Cable diameter
Useful cable length
Mean drum diameter
Cable angle - static
Cable angle-dynamic
Input speed
Overall reduction ratio
1st stage spur ratio
2nd stage spur ratio
3rd stage confound
planetary ratio
Pitch line velocity
40,000 lb
100,000 lb
115,000 lb
150,000 lb
1.39 inches
100 feet
34 inches
± 30 c (starboard,
+ 15° (starboard,
3000 rpm
443.218 to 1
4.190 to 1
3.513 to 1
port) + 30 ® (forward)
- 60° (aft)
port) + 15® (forward)
- 30° (aft)
30.111 to 1
60.2 fpm
159
Figure 48 is a schematic of the basic drive gearing arrangement
Figure 48. Gearing Schematic,
Single-Point Hoist.
Drum Design
In the Phase I Design Analysis section of this report the, the single-point
hoist width and diameter were conservatively chosen at 15 inches and 39
inches respectively for 100 feet of cable to maintain the single-rotor
heavy lift helicopter's lateral cyclic stick movements within desirable
limits, see page 18. A more detailed aerodynamic trim analysis, based on
the 3-inch permissible control stick travel of Reference 10 ,* indicates
that the allowable lateral C.G. shift is 9 inches to the right and 8.4
inches to the left. Figure 49 shows the C.G. shift due to hoist loads.
*Note: The longitudinal control displacement of 3 inches from the
initial trim position is considered to be applicable to
lateral trim conditions in the absence of a specific limita¬
tion for lateral stick displacement limits.
160
CO of AIRCRAFT
EXTERNAL LOAD
Figure 49. Center-of-Gravity Shift Due to
Hoist Load At Maximum Limits of
Cable.
From Figure 49,
it - Aircraft gross weight /ot)
1 2 External load
For an external load of 40,000 pounds, the single-rotor heavy lift heli¬
copter of Reference 4 has a gross weight of 79,071 pounds. Using the mini¬
mum permissible C.G. shift equal to 8.4 inches (to the left) and solving
for 1^ we obtain
161
IfiWW
h
h
8.4
79.071
40,000
16.6 inches
On the basis of this analysis, a drum diameter as small as 24*0 inches can
be utilized. A drum mean diamste* of 34 inches has been chosen, however,
to provide for good cable life and adequate space for the hoist gearing
and housings.
With this size drum, 12 cable wraps are required for 100 feet of cable.
With the 12 active and 3 "dead" wraps spaced at a 1.5- inch pitch, the
drum width is 22.5 inches. The total cable travel is 16 indies (or 9
inches to right or left of the aircraft centerline).
Utilizing the drum analysis of page 30, the drum thickness for a 34-inch
7075 aluminum forged drum is 1*44 inches.
Major Structural Members
Figure 50 shows the location of the major structural bearings in the
single-point hoist.
Figure 50. Drum and Support Structures,
Single-Point Hoist.
i
I
162
Cable elde loads are transmitted from the bellmouth, through the level
wind ball screw, and into the level wind arm. On the main structure, the
Timken bearing drum support reacts all the thrust loads from the level
wind am. The end view shown in Figure 51 locates the load isolator and
fore and aft maximum cable limits*
P c 30° *H>
60 ° APT
Figure 51. Load Isolator and Support
Structure, Single-Point Hoist.
Table XXIII sunsarises the bearing reactions. A, B, C, and D, for various
load conditions.
163
a
Figure 52 ie a sketch of the input housing sheering the critical section
and bearing reactions for the worst case load (cable out, 30° forward).
1
Figure 52, Input Housing, Single-Point Hoist.
At critical section a-a,
*i
Z
Z
- 23
- 22.625
JT A*
32 d 0
i (23.00(A - 22.625 4 )
32 23.000
165
z
m
76.02 in. 3
(32)
V
m
9.9 R cr
"br
m
9.9 (57,820)
(33)
**bT
m
572,420
m
9.9 RcH
“bH
m
9.9 (36,680)
(34)
%
m
363,130
«
i
i
“b
-
*bv
1
!
I
*b
-
572,420 -H* 363,130
(35)
*b
-
677,880 in.-lb
Using equation (6), we obtain
f . . 6^880
b 76.02
f. - 8,920 pai
Figure 53 shows the bearing reactions and critical section for the worst
load condition (cable in, 30° forward) on the drun and support housing.
Figure 53* Drum and Support Housing,
Single-Point Hoist,
,
! 166
At critical section a-a,
d 0 - 14.2
di - 12
Using equation (32), we obtain
z
" 32 14.2
z
- 137.7
*br
- 11.9 8^
“by
- 11.9 (93,510)
*bv
- 1,112,770
*bH '
- 11.9 8jjn
"bH '
- 11.9 (55,870)
**bH
- 664,850
ng equation (35X we obtain
■b 1
- 1,112,770 4* 664,850
"b 1
- 1,296,250 in.-lb
Using equation (6), we obtain
- 9*410 psi
. i&Lm
137.7
f b - 9*410 pal
Meanting attachment loads
Figure 54 is a scheaatic drawing of the hoist housings showing the
■ounting feet locations.
From a statical equilibrium analysis of the system, the Mounting foot
loads are found to be
R a « P c [ Y 2 2 ?8 9 ^ U-56 4 cos 0 - 1.632 sin 0) - .0820 (38)
% = P c [ X 7 C 8 ° s e $ (cos 0 4 1.046 sin 9 -1) 4 .Q340 (39)
Rp - P c (1.56 4 cos 0 - 1.632 sin 0) - 1.2360 (40)
Rp “ P c [ ^ --- y^g 00 - 0 (cos 0 4 1.046 sin 0 -1) 4 .5170 (41)
Foots A and C contain 2 bolts in each foot; hence, the bolt load ie found
by dividing the reactions at A and C in half. It is assumed that foots A
and B share the shear loads. The foot shear loads are given by
S A - Sg - | J (sin 0 cos f6 4 sin 40?) 2 t (sin 0 sin 0) 2 (42)
The worst case of bolt tension occurs in bolt D for cable out, zero side
angle, 30° forward angle. Under these conditions the applied bolt load
is 101,550 pounds. The total bolt load is given by
P - KP. + Pi m
where
P a - applied load * 101,550 lb
Pj_ “ initial bolt preload
K - pet of applied load felt in the bolt
For a flange thickness of 1.25 and a bolt diameter of .875,
K - .295
Pi - 30,000 lb (180,000 F tu bolt, 7/8 diameter)
Using equation (43)» we obtain
P - (.295) (101,550) 4 30,000
P “ 59,960 lb-Maximum bolt load
P allowable “ 86,100 lb
169
* "mil BH*
The level wind ball acre* and loads an transmitted into the rigid drive
half of the level wind am. The two level wind an bearings react all
the moment caused by side cable loads and transfer the moment directly
into the hoist side mounting plate. The worst case of stress in the
level wind am occurs under ultimate cable load conditions and 30° side
angle. Figure 55 shows the level wind side am and critical section.
Figure 55. Critical Section, Level Wind
Mechanism, Single-Point Hoist.
170
*
»
t
At critical section AA,
x - 3.461
5 - 317
Z - 57.3
% - P (sin 0) 1
Mb - (150,000) (sin 30°) (12.2)
Mb «* 915,000 in.-lb (44)
Using equation (6), we obtain
fv . j l S iOO O
b 57.3
f b - 15,970 psi
For a 7079-T6 aluminum forging, F^ u ■ 71,000 P»i
F tu
M.S.
a
fv
-1
MS ■» 71«000 -I
M-S - 15,970 - 1
M.S. - + 3.45 (45)
The bending moment on the hoist centerline is given by equation(44):
- (l50,000)(sin 30°) (21.5)
- 1,612,500 in.-lb
This moment is taken out by the two level wind ball bearings.
p . !L
brg i
brg
1.612 . 500
11.5
171
P brg " HO,200 lb (46)
The static capacity (Co) of these bearings is 82,000 pounds; therefore.
P brg
- 1.71 Co
82,000
(47)
M.S. -
** allowable
© A
(48)
*brg
Substituting
in equation (48),
*
d
M.S.
-J-Co,
1.71 Co
M.S. - + .75
The level wind bellmouth utilizes a screw to keep the cable tracking in
the grooves of the drum. A ball screw is used to reduce friction and in¬
crease efficiency of the system. Reaction torque of the ball screw is
provided by a fixed reaction pipe which also serves to stiffen the support
structure. The basic ball screw data are given below:
Screw diameter
2 - 1/2
Ball diamster
3/8
ThresH./inch
2
Lead
.5
No. of turns of balls
Operating load
Maximum static load
7
35>280 lb (Reference 11)
196,000 lb (Reference 11)
Since the drum pitch is 1.5 inches, the ratio from the drum to the screw
is
RR
L screw
*tlrum
RR
1
3
(49)
*
f
172
Hence the ball screw must turn three times faster than the drum. The
axial loads felt by the screw are given by
P a (normal) - P(normal) sin 0
P ft (normal)
« 4000 (sin 30°) = 20,000 lb
(50)
P a (max)
= Pmax sin 0 ““
P a (max)
~ 150,000 (sin 30°) » 75,000 lb
(51)
These loads are below the allowable loads for the ball screw as given in
Reference 11 for 1 million inches of ball travel. For 12 wraps of cable,
1 million inches of travel is equivalent to 27,800 cycles of hoist opera¬
tion.
The screw must also be designed to avoid buckling as a column.
L
= 29.5 (distance between supports)
d^ (effective)
0
« 2.31
d i
- 1.75
I
- .937 in. 4
The critical buckling load for pinned ends is given by
L 2
* 2 (30 x 10 6 )(.937)
(29.5 ) 2
318,800 lb (52)
The torque required to turn the screw against the axial load P a is given
by
T
The efficiency
*
:k
*a 1
2 W7)
(53)
rf is conservatively assumed to be 90 pet (Reference 8):
173
20*000 ( .??
2 t(.9)
1,770 in.-lb(at level wind ball screw)
The scrub roll is a rubber coated roller that pulls the cable off the drum
when lowering and is free wheeling when ra' ling the cable. The cable is
loaded against the scrub drive roller by the scrub roll pulley which is
adjustable to compensate for wear and to obtain the proper initial tension.
The scrub roll pulley is mounted on the bellmouth and travels along the
ball screw, keeping the same relative position with the cable as shown in
Figure 56.
^normal
T
1 normal
Figure 56. Bellmouth, Scrub Roller,and Ball
Screw Assembly, Single-Point Hoist.
To provide a minimum tension of 50 pounds in the cable at all times, the
normal force required on the pulley is found by
n
T
Assuming that the coefficient of friction is .3, we obtain
(54)
174
This force produces a torque on the scrub roller given by
T
F -i.
» 2
T
167
T
230 in.-lb(torque required to drive scrub rolls)
Weston Brake
(55)
The Weston brake is used to raise the load with all units locked, to hold
the load with the input power removed, and to lower the load at the same
speed as the motor. The important design considerations are adequate heat
dissipation in the discs when lettering the load, the proper lead ar^le of
the unlocking device to prevent self-locking if the angle is too low, and
failure to lock if the angle is too high. For heat dissipation, the plate
pressure must be lett.
» 9.00
d i
- 6.00
n
= 7 friction surfaces
T
- 6430 in.-lb (40,000-lb cable load)
The brake discs are SAE 1095 steel against high friction bronze. For
these materials operating in oil, the coefficient of friction is 0.07.
P a
8 T
r /i.d i (d Q 2 - dj 2 ) n
8(6430)
P a
1r (,07)(6.00)(9.00 2 -
6.00 2 )(7)
P
a
«= 96.3 psi
(56)
P (allowable) -» 150 psi
175
Screw Data
5 - 1 ^ triple thread
lead = 1 = 2.25
0 * " 4 ‘ 25
The lead angle is given by
a <= arc tan (— 7 ——)
a = arc tan — rf^rr
ir ( 4 . 25 )
a ** arc tan .16851
a *= 9° 34'
For proper operation (Reference 7)»
6°< a < 12 0
*
(57)
m
Heat Riae in Weston Brake
When the Weston brake is used to lower the load, all the energy of the
load must be absorbed by the brake. This energy is the work done per
revolution times the number of revolutions required to lower the load.
Assuming that a 40,000 pound is being lowered 100 feet, we obtain
where
t = time to lower load, min
L = length of cable, ft
V = cable velocity, fpm
( 59 )
1
100
"5o72
1.661 min
176
Heat generated is given by
t
Btu (60)
= energy expended in ft-lb
\ = (2 tr T)(rpm)(t) (6l)
where
T - torque * 536 ft-lb
rpm *= 716
\ = 2 tt (536)(716)(1.661)
\ - 4,010,000 ft-lb
H - 4 ,010.«0 0 0
778
H * 5,150 Btu per Weston brake lowering operation
The temperature rise after one lowering operation is
= Jk
778
where
At
H
Ob W b + Co W Q
( 62 )
where
At « temperature rise, °F
= specific heat of brake parts = .12 Btu/lb/°F
W b = weight of brake parts that will heat up, lb
Co = specific heat of oil ** .55 Btu/lb/°F
W Q *= weight of oil surrounding brake, lb
While the drive train and surrounding castings weigh 589 pounds, it has
been conservatively assumed that 150 pounds of metal and 25 pounds of oil
which surround the Weston brake are the effective heat sink during a low¬
ering operation. Using equation (62Xwe obtain
177
i
I
1.
At - _ 5150
A (,12)(150) + (,55)(25)
At « 162 °F
The temperature rise is 162°F in one braking operation. Between opera¬
tions, the plates will cool by convection and radiation.
Free Reeling Clutch
The free reeling clutch is normally engaged by the axial load caused by
the compression of a stack of Belleville washers. Since this clutch never
slips, it is designed on static torque capacity and can have a high plate
pressure. In addition, there will be very little oil on the plates and
the friction torque will be governed by the static coefficient of friction.
The larger Timken bearing carries the axial clutch load during normal
operation but has no relative rotation between inner and outer races. The
ball bearings used to prevent rotation of the hydraulic cylinder rotate
during normal operation but carry no load. When the free reel clutch is
disengaged, these bearings must withstand the axial load created by the
hydraulic cylinder necessary to compress the Belleville washers and free
the clutch plates.
During free fall of the load, high rotational speeds are produced, causing
dynamic tensile stresses in rotating parts. An analysis is made showing
the effect of inertia on final speed after free fall and also the effect
of stresses in critical parts due to the high rotational speed.
T. . * T x 3.75 x F.S.
design normal
Assuming a factor of safety of 1.25, we obtain
(63)
design
design
6430 x 3.75 x 1.25
30,140 in.-lb
6.00
4.00
11 friction surfaces
ft •= .25 (static)
Using equation (56), we obtain
n
8 (30.140) _
(.25)(4.00)6.00 2 - 4.00 2 ) 11
178
V
p a « 348 pai
The axial clutch load necessary to produce this pressure is
„ it Pa d i( d o “ d i)
F a " - 2 -
F « Z (348)(4«00)(6«00 - 4,00)
f a 2
F * 4370 lb (64)
cl
Belleville Washer Design (Reference 15)
The Belleville washer must provide a preload of 4370 pounds on the clutch
plates for proper torque. This load must be produced before the spring
becomes flat so that further compression can take place in order to re¬
lease the clutch plates. The stack of Belleville washers is shown in
Figure 57.
f
*
Figure 57. Belleville Washers, Free Reeling
Clutch, Single-Point Hoist.
179
(65)
R
r
t
n
a
- 3.00
» 2.00
- .120
■= .156
c x C E tA
R 2
where
ie a function of 8/t and h/t and is given by a
curve in Reference 15.
C ie a function of R/r and is given by a curve in
Reference 15.
With the values of R/r - 2*00 “ ^ and c ** 1»9A substituted in
equation (65), we obtain
C - A370 (3.00) 2
1 1.9A(30 x 10 6 )(.12r
C ± - 3.26
For three springe in parallel,
C 1 - ^ - 1.09
For this and ~ - 1.3, ^ - .7,
t t
8 - .7t - .08A
For two springs in series,
A «28
A - (2)(.7)(.120> - .168
The amount left for compression to release the clutch is
clutch plate clearance * 2h - A
( 66 )
*
180
clutch plate clearance «= 2( .156) - .168
clutch plate clearance - .144 (67)
To find the force required to flatten the springs and release the clutch,
$ = h = 1.30 (68)
t t
C x - 1.42
For three springs in parallel,
C x - 3(1.42)
C x = 4.26
With these conditions the force required to flatten the springs is given
by equation (65):
r _ 4.26(1.94)00 X 10 6 )(.12) 4
* (3.00) 2
F - 5710 lb - Force required to flatten springs
Drua Speed After Free Fall of Load
The isolated drum and load are shown in Figure 58. The mass polar moment
of inertia, J, of the drum includes all rotating internal parts from the
drum to the free reeling clutch. These parts are related to the speed of
the drum by the ratio of the speed squared.
Figure 58. Free Falling of Load,
Single-Point Hoist.
181
Zf
ma
(69)
W-T
-
W
— a
g
(70)
ZM
J a
(71)
TR
“
Ja - J i
(72)
Here, the
acceleration is due to gravity: a ■ g
Also,
a
-
R W T
(73)
S
-
(74)
Solving for u> and eliminating time,
t, and cable tension, T,
U)
„ 1
2 V/ S
(75)
J
r 2 ITT.
8 R 2
30 1 2 W S g
final drum speed after fall
(76)
1 H“ttrum
* J W R 2 + g J
of weight V/ through distance
S.
J
B
4736 in.-lb/sec 2
for all rotating parts of the single■
point hoist from the drum to the
drum to the free reel clutch and re¬
flected to the drum speed.
O O
for R = 17, g 15 386 in./sec , and J ■* 4736 in.-lb/sec ,
30
2 W S (386)
rpm drum T
N
V (17) 2 + 386 (4736)
•
r-nm q CIO 1 2.67 w S
r P“drum 7.549^1 w + 6326
9
Figure 59 shows a plot of drum speed increase factor versus distance of
fall for various weights.
If J = 0, the equation for RPM of the drum reduces to
182
rpadrum
rp*Arm = 9.549 J 2.671 £ (77)
As can be seen in Figure 59* when heavy loads are on the hoist and the free
wheel unit is released, the drum inertia has little effect on the final
rotational speed over that of a free falling body.
The stress due to high rotational speed on a cylinder of constant thick¬
ness (Reference 6) is
f
r max
0 + *)
32
(78)
l t max
p U) 2
16 g
(3 + v ) do 2 + (1 -
(79)
The free reeling clutch shaft is normally turning at 716 rpm. After free
fall of 100 feet with a 40,000-pound cable load, the Bhaft is turning at
53*104 rpm. The free reeling clutch output side reaction plate is the
most highly stressed member under these conditions.
For this plate,
d Q = 6.00
=0
p * .263 lb/in. 3
v ■= .3
ui = rpm x —
30
« - 53,104 -*q-
ut = 5,561 rad/sec
Using equation (78), we obtain
f
r max
, 0
32
( ,263)(5561 ) 2
36S
(6.00 - 0) 2
(80)
♦
t
184
f
max
84,160 psi
Using equation (79i we obtain
ft max - [ (3 + • 3)<6 - 00)2 + (1 - ' 3) (C)2
f t max * 168 ' 300 P si
The free reeling clutch output plate is made from AMS 5000 steel whose
ultimate tensile strength is 200,000 psi and whose yield strength is
176,000 psi. The maximum tensile stress produced in this part by high
speed rotation is therefore below the yield strength.
Gear Design
The face width of the spur gears in the single-point hoist may be governed
by bending stress or Hertz stress. The formulas for these stresses are as
follows:
1.5 W t h
X F
21 x 10 6 W t
cf t a
sin 2 0 F 'dp ~ d &
The tangential tooth load, W^, is given by
)
(81)
(+ for external gear)
(- for internal gear) V '
W*
2 T
d
(83)
The torque, T, is found by dividing the torque at the drum by the appro¬
priate gear ratio (from the drum to the gear in question):
T
P d
c m
2 RR
(84)
Table XXIV summarizes the bending and compressive stresses for'all the
spur gear teeth on the single-point hoist. Tho tangential tooth loads on
the level wind gears are derived by finding the torque required to turn
the level wind feed screw to overcome the axial load caused by a 30° side
cable angle. Since the gear tooth ultimate tensile strength is greater
than 3.75 x bending stress allowable, the lowest margins of safety occur
during norma] operating conditions; hence the ultimate bending stresses
have been omitted.
Shaft bearing reactions caused by gear loads may be found by well-known
• 185
TABLE XXIV
- SINGLE-POINT HOIST
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186
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methods, as shown in Reference 3. All bearing loads, ratings, and lives
are tabulated in Tables XXV and XXVI for* normal and static conditions.
Drive System Shafting
The drive train shafts may be subjected to bending stresses, torsional
stresses, or a combination of both. In bending, the hoist shafts are
critical under normal load fatigue conditions, since the endurance limit
multiplied by the ultimate load factor is less than the ultimate tensile
strength of the material. In torsion, however, the shafts are critical
under ultimate load conditions. The single-point hoist shafting has been
analysed by the methods of Reference 3 and the results are given in Table
XXVII.
Cable
The cable for the single-point hoist .is designed for a 150,000-pound ulti¬
mate load. This load is derived by the following method.
ult - P limlt x F-S '
(85)
limit " 1 <° Factor)
(86)
For the single-point hoist, the G Factor is 2.5 and the factor of safety
for ultimate load conditions is 1.5.
P limit * 40,000 (2.5) - 100,000 lb
P ult “ l 00 * 0 ^ (1.5) - 150,000 lb
The cable breaking load is given by
P «= F tu KA (87)
where
K - Area Factor ■ .71
A “ Total Area of all Strands
For the single-point hoist, the wire diameter is .060 in. and the num¬
ber of wires is 342 (18 x 19 construction)
A - No. of wires ^
4
167
TABI£ XXV
OF BEARING LIVES AND LOADS - SINGLE-POINT HOIST
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188
Notes: 1. Static loads are not felt from the input to the Weston brake
2. Ary bearings not shown carry no load (or negligible load).
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n.
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k
A = 18 x 19 x -f- (.060) 2
4
A - .967 in? (88)
F t - 250,000 psi
P «= (250,000)(.71)(.967)
P ® 171»600 pounds - cable breaking strength
SINGLE-POINT HOIST DRIVE SYSm
Clutch-Reverser Unit
«
The clutch-reverser unit is the power source and directional control for
the single-point hoist. Power is taken from the accessory gearbox to
which the clutch-reverser unit is mounted. The 1.364 to 1 reduction gear¬
box consists of three mutually geared shafts, two of which contain hy¬
draulic clutches as shown in Figure 72. One clutch is used for raising
the load while the other is used for lowering the load. The hydraulic
control system prevents both clutches from simultaneously engaging.
Design Data
Input rpm
Output rpm
Reduction ratio
Input torque
Horsepower
5995.2
4396.5
1.364 to 1
957 in.-lb
91
(40,000-lb cable load)
The gear tooth bending and compressive stresses are found by methods simi¬
lar to those used in the single-point hoist and are sunnarized in Table
XXVIII, page 193.
Bearing loads and lives are found by methods similar to those of Refer¬
ence 3* AH bearing loads, ratings, and lives are tabulated in Table XXIX,
page 194,for the clutch-reverser unit.
Clutch Analysis
Both the load lifting and load lowering clutches of the clutch-reverser
unit are identical in design. The load lowering clutch can have a much
small er capacity than the load lifting clutch, since it must only over¬
come friction in the system. The units are made identical because of
cost and assembly reasons, since their weight is small.
The clutches are hydraulically actuated by a supply pressure d 250 psi.
The axial load is given by
i
I
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7
■ U
I
191
p, ■ pa (89)
where
P - pressure ■ 250 psi
A « piston area - 7.903 in. 2
F a - 250 (7.9Q3)
F a - 1980 lb
The torque capacity is given by
T “ -jr" (d 0 + d i )n (90)
where
d Q - plate outside diameter = 3.82
d^ “ plate inside diameter « 2.00
n * number of friction surfaces *• 16
fi ■ coefficient of friction ■ .07 dynamic (Reference 5)
T - . Ct°7J (3.82 + 2,00)(l6)
4
T - 3220 in.-lb
On the low-speed (4396.5 rpm) sliaft,this torque is equivalent to 224 horse¬
power which is well above the 94 horsepower design criteria even when using
the very conservative coefficient of friction of .07.
Using equation (56), the plate pressure at a torque of 3220 in.-lb is
8 (3220) _
ir (, 07 )( 2 . 00 )( 3 . 82 2 - 2.00 2 ) 16
P a - 366 psi
The clutch plates are lubricated by a constant oil supply and only slip
during acceleration of the load.
192
Upper Angle Gearbox
The upper angle gearbox (Figure 72) is mounted directly on the output of
the clutch-reverser unit. The 1.172 to 1 reduction ratio le accomplished
by a single set of spiral bevel gears which turn the drive shaft downward
and sideward toward the hoist angle gearbox.
Spiral Bevel Gear Data
0
*
"p
<*P
F
*P
Shaft angle = 139° 30'
Pressure angle - 20°
Spiral angle » 35°
Diametral pitch « 6
Number of teeth in pinion ■= 29
Pitch diameter of pinion - 4.833
Number of teeth in gear «= 34
Pitch diameter of gear « 5.667
Face width - .86
Contact ratio ■ 1.35
Table XXX summarizes bearing lives for the four tapered roller bearings
in the reverser output angle gearbox for 40,000-pound cable load.
%
table XXX
BEARING LIVES AND LOADS-
UPPBt ANGLE GEARBOX
Location
Loads(Pounds)
Basic
Radial
Rating
Life
(Hours)
Thrust
Radial
Radial
Equivalent
Input Pinion, Gear End
483
1070
1455
1800
662
Input Pinion, Input End
0
238
711
865
700
Output Gear, Gear End
216
1110
1356
1240
1020
Output Gear, Output End
0
400
400
510
5250
195
Lower Angle Gearbox
The lower angle gearbox (Figure 72) ie Mounted on the airframe near the
hoist input. A drive shaft connects the input to the output of the rovers-
er angle gearbox, while another drive shaft connects the output to the
hoist input pinion. The 1.25 to 1 reduction ratio Is accomplished by a
single set of spiral bevel gears.
Spiral Bevel Gear Data
a
0
*
%
"p
Shaft angle
Pressure angle
Spiral angle
Diametral pitch
Number of teeth in pinion
Pitch diameter of pix.ion
Number of teeth in gear
Pitch diameter of gear
Face width
Contact ratio
= 80°
«=■• 20 °
- 35°
= 6
*= 20
»= 3.667
- 25
■= 4.167
= .88
- 1.40
*
Table XXXI summarizes bearing lives for the four tapered roller bearings
in the hoist input angle gearbox for 40,000-pound cable load.
table xxii
BEARING LIVES AND LQADS-
. Wm AMPLE GEARBOX
Location
Loads (Pounds)
Radial
_Thrust Radial_Equivalent
Basic
Radial
Ratine
Life
(Hours)
Input Pinion, Gear End
895
1625
2852
12,020
540
Input Pinion, Input End
0
705
1646
7,870
800
Output Gear, Gear End
875
1.810
3140
11,910
476
Output Gear, Output End
0
873
1329
6,140
903
196
FOUR-POINT HOIST
The main drive train of the four-po:.nt hoist consists of four stages of
gearing as shown ii, Figure 73, page 261. A hydraulic motor drives the
first stage spur pinion, idler, and gear of reduction ratio 4.517 to 1.
The first stage ge-r shaft drives the input section of the Weston brake.
As in the single-point hoist, the Weston brake turns as a unit when rais¬
ing the load, slips at the same speed as the input when lowering the load,
and holds the load with the input power off. A hydraulically powered free
reel clutch similar to that used in the single-point hoist permits the
hook and cable assembly to be jettisoned in the event of a hook release
malfunction. It is located between the Weston brake and the second stage
spur pinion. The second stage of gearing is a 6.273 to 1 reduction ratio
spur pinion gear. A conventional planetary (sun gear driving, rin^ gear
fixed, cage output) third stage of 4.9Q3 to 1 reduction ratio is driven
from the second stage gear. The output cage of the third stage planetary
drives the sun gear of the 3.702 to 1 reduction ratio fourth stage con¬
ventional planetary whose output cage is bolted directly to the drum.
The level wind ball screw is rotated by a chain drive whose drive sprocket
is attached to the drum side plate. The main support is a two-way swivel
arrangement which does not allow the transmission of ary moment loading
into the airframe. A hydraulic load isolator is mounted between the air¬
frame and hoist to dampen load oscillations.
Design Data
Normal cable load
=
11,550 lb
Limit cable load
m
32,300 lb
Yield cable load
m
37,200 lb
Ultimate cable load
m
48,500 lb
Cable diameter
B
.79 in.
Useful cable length
-
50 ft
Mean drum diameter
B
22.19 in.
Cable angle (static)
B
+ 30 ° (any direction)
Cable angle (dynamic)
B
+15° (ary direction)
Input speed
■
2750 rpm
Overall reduction ratio
B
514.356
1st Stage spur ratio
B
4.517
2nd Stage spur ratio
■
6.273
3rd Stage planetary ratio
■
4.903
4th Stage planetary ratio
B
3.702
Cable pitch line velocity
-
31.06 fpm
Drum
The drum for the four-point hoist has a mean diameter of 22.19 inches and
a length of 10.5 inches. The analysis of this component is summarized
in Table VIII, page 30.
197
Major Structural Members
i
Figure 60 shows the location of the major structural bearings in the four-
point hoist.
Figure 60* Major Structural Members,
Feur-Peint Heist*
Cable loads are transmitted to the side plates as tensile loads. No
thrust can be transmitted because of the zero-moment mounting.
From a consideration of static equilibrium,
Ra
«b
Ra
Rb
- .0667P e ) „ .. T
- 1.0667P e ) Cable **
■ ^AAAAPj j Cable Out
(91)
(92)
The minimum cross section on the side of bearing 6 is shown in Figure 61 •
%
«
198
2.75
Figure 61. Side Plate, Four-Point Hoist.
At section a-a,
base ■ 6.5
height « 1.5C
for P,
- 2.438
- 2.75 Hg
- 48,500 lb (ultimate load)
- 2.75 (1.0667) (48,500) (cable in)
“b
From equation (6)
142,300 in.-lb
. 142x200
2.438
- 58,370 psi
199
The level wind mechanism In the four-point hoist has axial forces Induced
due to the angle of the cable when it is in its limit positions. This
Figure 62. Induced Axial Loads in Level Wind
Ball Screw, Four-Point Hoist.
From Figure 62,
tan 0 - | (97!
P a - P c sin 0 (98]
For small angles sin 0 ~ tan 0 ,
P- - P r | (99]
I - 3.5 in.
T - 35.35 in.
P a
35.35
P a
.099 P c
P a normal “
.099 (11,550)
P a normal
1140 lb
p a ult
.099 (48,500)
p a ult
4800 lb
Ball screw data
Ball circle diameter
Ball diameter
Lead of thread
Turns of balls
Operating load
Max static load
- 1.5
- 5/32
- 1/4
- 5
* 5240 lb (Reference 11)
= 31,500 lb (Reference ll)
The normal load is less than the operating load for 1 million inches of
travel, as shown in Reference 11. One million inches of travel in the
four-point hoist corresponds to 71,430 cycles of operation (3830 hours).
The ratio from the drum to the ball screw is given by equation (49).
Hence the ball screw must turn 3.5 times as fast as the drum to keep the
bellmouth of the level wind in the same relative axial position as the
cable takeoff point on the drum.
If we assume that the ball screw shaft is singly supported, the critical
buckling load can be found from equation (52).
L
*i
15
1.4 (effective)
.90
201
I - .1564 in. 4
P . iP. xio 6
(15) 2
P » 206,OOO-lb screw buckling load
The torque required to turn the screw against the 1140-pound normal load
is given by equation (53).
V
■ .90 (Reference 8)
T
. 1140 (.25)
normal
2 ir (.90)
T normal
« 50.4 in.-lb
The scrub roll is a rubber coated roller whose pitch line velocity is 9.8
pet greater than the cable pitch line velocity when the cable is reeling
out. A one-way clutch disengages the roller when the cable is reeling in.
An adjustable pulley, which rides with the bellmouth and ball screw, is
used to load the cable against the scrub roll rubs on the cable, keeping
a slight tension in it to keep ths cable tracking in the drum grooves when
the hook is unloaded.
Weston Brake
The Weston brake in the four-point hoist is similar in design to the
Weston brake in the single-point hoist. The lining and screw data are
given below:
d Q - 5.25
d A - 3.25
n « 6 friction surfaces
/a - .07 *
T - 1130 in.-lb (11,550-lb cable load)
Prom equation (56),
p ._8 1U 10J_
8 tt (.07)(3.25)(5.25 2 - 3.25 2 ) 6
P a - 124 psi
202
Screw data:
2£ - 3 triple thread
lead *= 1 * .3333
d* - 1.917
The lead angle, a , is calculated using equation (57).
a =■ arc tan 2 . (?23 . 233 2
a - arc tan .166Q3
a » 9 ° 26 *
This value meets the requirement for proper operation as given by
equation (58).
Heat Rise in Weston Brake
When the Weston brake is used to lower the load, all the energy of the
load must be absorbed by the brake. This energy is given by the work done
per revolution times the lumber of revolutions required to lower the load.
For a 11,550-pound load being lowered 50 feet at 31.06 fpm,the time to
lower is calculated using equation (59).
»
*
Heat generated is given by equations (60) and (61).
H . 2 ir (94.2)(608.6)(l.61)
778
H - 746 Btu
It has been conservatively assumed that 50 pounds of metal and 10 pounds
ef oil which surround the Weston brake are affected by heat when lowering
the load. The temperature rise after one lowering operation is calculated
using equation (62).
At
At
.12 (50) + .55 (10)
65°F
203
Free Reeli ng Emergency Release
As in the single-point hoist, an emergency free reel release has been pro¬
vided. This unit is a clutch whose plates are axially loaded by Belleville
waehers. A hydraulic piston compresses the washers to release the clutch
plates. The clutch is designed to carry the ultimate cable loa' 1 with a
factor of safety of 1.25.
^design
P d*
2 RR
x F. S.
^design
48.500 (22.19)
2 (113.86)
x 1.25
1 design
5,910 in.-lb
( 100 )
o
d i
n
3.00
2.00
14 friction surfaces
.25 (static)
Using equation (56), we obtain
P _ 8 (W10)
ir (.25)(2.00)(3.00 2 - 2.00 2 ) 14
434 psi
The axial clutch load necessary to produce this pressure is given by
equation (64).
T.(4?AK 2 t°Q) ( 3,00 - 2.00)
1360-lb axial load required by Belleville washers
The Belleville washers design for the four-point hoist is similar to the
Belleville washer design for the single-point hoist. The same stacking
arrangement is used: two sets in series of three washers in parallel for
each set. Using the nomenclature of Figure 57, page 179, the data and
results are as follows:
204
9
«
R - 1.88
r = 1.06
t = .069
h = .103
C - 1.57
C x « 1.50
8 - .050
A * .100
Clutch plate clearance = .106
V.'ith these values substituted in equation (65) the force required
to flatten the springs is 1450 pounds.
Drum Speed After Free Fall
By substituting appropriate values in equation (76), the curves shown by
Figure 63 have been plotted to show the ratio of rpm before and after fall
versus distance of fall for various cable loads.
As was the case in the single-point hoist, in the four-point hoist the
inertia has little effect over that of a free falling weight, when the
weight is large.
The free-reeling clutch shaft is normally rotating at 608.8 rpm. After
free fall of 50 feet with a cable load of 11,550 pounds, the free reeling
clutch shaft will be rotating at 65,780 rpm. At this high speed, rota¬
tional stresses will be induced into the outer clutch plate holder (larg¬
est member on shaft). For this part,
d 0 *= 3.90
d A *= 3.25
p = .283 lb/in. 3
v ■ .3
Using equation (80),we obtain
u> - 65780 -
6886 rad/sec
Solving equation (78), we
obtain
f = (L+ .t 21
(.283)(6888)
r max 32
38<i
f rmax ' 1520 psi
(3.90 - 3.25) 2
205
RPli
DISTANCE OF FALL, INCHES
Figure 63. Load vs Free Fall Velocity,
Four-Point Hoist.
Solving equation (79), we obtain
*t max
f
t max
[(3 + .3)(3.90) 2 + (1 - .3 X 3.25) 2 ]
125,200 psi
The free reeling clutch output plate is made from AMS 5000 steel whose
ultimate tensile strength is 200,000 psi and whose yield strength is
176,000 psi. The maximum tensile stress produced in this part by high¬
speed rotation is therefore below the yield strength.
Gear Design
The bending and compressive stresses in the gear teeth of the four-point
hoist are found by methods similar to those used on page 185 for the
single-point hoist. The level wind drive for the four-point hoist is a
chain and sprocket arrangement and is covered in another section. Tabie
XXXII summarizes the bending and compressive stresses for all the spur
gear teeth on the four-point hoist.
Planetary Plate Design
The planetary carrier plates of the third and fourth stages of gearing in
the four-point hoist are subjected to steady bending stresses. 6AL-4V
titanium is used because it is lighter than steel plates. The plates may
be designed for maximum slope or maximum stress. The allowable stress
and slope is given below:
f b allow " 40,000 psi
0 allow " .001 inch per inch
Plate stress is given by
*b
where
T
12 T L
n d s 1 ^0 " d i " 1,2 d )
■ sun gear torque
- geometry factor
1
2
■8d
* sun gear diameter
( 101 )
( 102 )
TABLE XXXII
GEAR SUMMARY - POUR-POINT HOIST
I
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208
I
t
1
n
d P
d
g
t
d o
*i
= (d s + dp) sin I
*= number of pinions
«= pinion diameter
*= O.D. of inner race of pinion bearing
= distance between plates
*= thickness of single plate
** outside diameter
*= inside diameter
(103)
Plate slope is given by
0 -_16 TL3
n d„ E r (d_ - d, )
(* +_t)
(104)
where the symbols are the same as those used for the stress formula.
For the 3rd stage planetary,
g *= 1.50 in.
t ■= 7330 in.-lb ( 11 , 550 -lb cable load)
d B - 2.583 in.
d p * 3.75 in.
n *= 4 pinions
d *=2.31 in.
1 *= 4.478 in.
L *= 1.315 in.
d 0 *= 9.84 in.
d A - 2.75 in.
t *> .28 in.
E ■= 16x10^ psi
Substituting in equation (101),we obtain
209
(12 )(7330)( 1.315)
(u)(2.5^)(i:mW.& -2.75 - 1.2 X 2.31)
(1.50 + 28)
(.28)2
f b -
13,140 psi
Substituting in equation (104) we obtain
d m
(16)(7330)(1.315)3
r
(4)(2.583)(16 x 10 6 )(4.478) 2 (9.84 - 2.75)
t -
.00092 inch per inch
For the 4th stage planetary,
g
- 1.83 in.
T
- 35,960 in.-lb (11,550-lb cable load)
d s
» 4.700 in.
"P
4.000 in.
n
- 6 pinions
d
- 2.18 in.
1
= 4.350 in.
L
- 1.303 in.
d o
- 12.375 in.
d i
~ 5.000 in.
t
» .36 in.
E
■ 16 x l(fi psi
Substituting
in equation (I0l),we obtain
(12)(35,960) (1,303)
r b
(6)(4.70)(4.35X12.375 - 5.0 - 1.2 x 2.18)
f b "
16,280 psi
Substituting
in equation (104), we obtain
d
_(16)(35,9&Q)(I«3<S) 3 _
(6)(4.70)(16 x 10 6 )(4.350 ) 2 (12.375 - 5.00)
xi*aa ± 86)
(.28)3
1 1, 63 ± ^361
(.36) 2
LL& t .36)
(.36)3
210
0 « .00095 inch per inch
Level Wind Drive Chain
The level wind drive chain ie designed for a normal torque of 50.4 inch-
pounds. The chain and sprocket data are given below:
Nunber of Teeth in Drive Sprocket
-
63
Number of Teeth in Driven Sprocket
-
16
Diameter of Drive Sprocket
«=
7.560
Diameter of Driven Sprocket
American Standard Chain #35
r=
2.160
Chain Pitch
as
.375
Average Tensile Strength
=
2100 lb
Rated HP for 17-Tooth Sprocket
K
.061
Correction Factor for 18-Tooth Sprocket
XZ
1.05
RPM of Driven Sprocket
»
18.7
%
HP -
Txro
63 , 025 '
hp - IQAl M ifl
63,025
HP - .015 (105)
This is well belcw the allowable HP of .061 from Reference 12 for 15,000
hours of chain life. The maximuii tensile load in the chain is given by
max
2 T normal / p cable ult \
J ' p '
r cable normal
‘max
2 (50.4)
2.160
( 46,500 )
(11,550)
P max " 196 lb (106)
This is well below the average tensile strength given in Reference 12.
Bearing Design
Shaft bearing reaction caused by gear loads may be found by well-known
methods, as shown in Reference 5. All bearing loads, ratiigs, and lives
are tabulated in Table XXXI V for normal and static ultimate conditions.
Drive System Shaft Stresses
The drive train shafting of the four-point hoist may be subjected to bend¬
ing stresses, torsional stresses, or a combination of both. In bending,
211
?
.'fc
n-
S
J
m
the hoist shafts are critical under normal load fatigue conditions, since
the endurance limit multiplied by the ultimate load factor is less than
the ultimate tensile strength of the material. In torsion, however, the
shafts are critical under ultimate load conditions. The four-point hoist
shafting has been analyzed by the methods of Reference 3> and the results
are given in Table XXXIII.
TABLE XXXIII
CRITICAL SECTION SHAFT STRESSES - FOUR-POINT HOIST
Location
Critical Section
T^rpe of
Critical
Stress
Stress
(psi)
M.S.
Input Pinion Shaft
Output Pinion Brg. Rad.
Bending
5,770
2.47
Input Idler Shaft
Bearing Radius
Bending
8,030
1.49
Input Gear Shaft
Spline Undercut
bending
1,100
17.18
Free Reel Clutch
Input
Torque Shaft
Torsion
21,070
3.13
Free Reel Clutch
Output
Torque Shaft
Torsion
57,610
.51
2nd Stage Pinion
Shaft
Bearing Radius
Bending
8,930
1.24
2nd Stage Gear
Shaft
Webb Connection
Bending
12,280
.63
Note: Torsional stresses are critical under ultimate load
conditions.
Cable Breaking Strength
For the four-point hoist the G factor is 2.8 and the factor of safety for
ultimate load conditions is 1.5. Using equation (85) to determine the re¬
quired cable breaking strength, we obtain *
P ult " 11,550 ( 2 *8)( 1 »5) = 48,500 lb
The cable to be used has 342 wires (18 x 19 construction) that are .035-in. '
diameter. Solving equation (88) to determine cable area,we obtain
A = 18 x 19 x -j- x .Q35 2 = .329 in. 2
4
Solving equation (87) to obtain cable breaking strength, we obtain
P => 250,000 (.71)(.329) •= 58,400 1b
212
(TABLE XXXIV
SUMMAHI OF BEARING LIVES AND LQADS-
FOUR-POINT HOIST_
Special Load
Brg. Loa
Cable
(P ■ 32 ■'
_location_
Input Pinion, Gear End *
condition_
2750
inrust
Input Pinion, Outboard End *
2750
-
Input Idler, Outside End *
2750
-
Input Idler, Inside End *
2750
-
Input Gear, Outside End *
606.8
-
Input Gear, Weston Brake End *
608.8
-
Free Reel Clutch Main Timken
0
1360
Free Reel Clutch Piston Isolator**
Free Reel Piston On
608.8
1450
Free Reel Clutch Washer Preload**
Free Reel Piston On
608.8
1450
2nd Stage Pinion, Input End (Roller)
606.8
0
2nd Stage Pinion, Outboard End (Roller)
608.8
0
2nd Stage Gear, Input End
97.1
0
2nd Stage Gear, Planetary End
97.1
0
3rd Stage Planetary, Pinion Bearing
53.2
0
4th Stage Planetary, Pinion Roller Brg.
17.0
0
Level Wind Ball Screw Thrust Bearing
Cable Max. In or Out
18.7
3200
Main Drum Support Bearing (Brg. B Fig, 60)
Cable In
5.3
0
Secondary Drum Support Brg. (Brg. A Fig.60)
Cable Out
5.3
0
Note: Any bearings not shown carry no load (or negligible load
*Static loads not felt from input to Weston brake.
**Free reel bearings loaded only when free reel piston act:
A *
TABUS HOT
SUMKARX OP BEARING LIVES AND LOADS-
_FOUR-POINT HOIST_
Brg. Load 9 Limit
Cable Load
Static Cap.
1.25 Co
Brg. Load 9 Normal
Cable Load
Special Load
(3 Co for
■rcZPfvvSVulK
Dynamic
Life
2750
-
-
0
495
1,740
263
2750
-
-
-
0
165
1,740
7,110
2750
-
-
m
0
286
965
233
2750
-
-
-
0
286
965
233
606.8
-
-
-
0
152
5,090
1,030,000
608.8
-
-
-
0
178
5,280
714,400
0
1360
0
25,700
1360
0
-
Free Reel Piston On
608.6
1450
0
14,900
1450
-
Free Reel Piston On
608.8
1450
0
11,400
1450
0
-
606.8
0
2570
4,150
0
920
5,420
9,860
606.6
0
2570
4,150
0
920
5,420
9,860
97.1
0
3240
4,780
0
1,160
4,890
12,860
97.1
0
1900
2,570
0
680
2,300
6,640
53.2
0
7940
8,000
0
2,840
5,520
1,330
17.0
0
14,260
17,850
0
5,100
19,100
43,480
Cable Max. In or Out
18.7
3200
170
7,650
1140
60
5,270
19,290
Cable In
5.3
0
34,450
53,400
0
12,320
14,290
4,860
Cable Out
5.3
0
17,950
23,600
0
6,420
16,200
50,030
t shown carry no load (or negligible load).
t felt from input to Weston brake.
ngs loaded only when free reel piston activated.
B.
I
WEIGHT ANALYSIS
A weight analysis based on the preliminary design (layout) drawings of
Appendix III has been made of all the components comprising the 40,OOO-
pound-capacity external cargo handling system. A detailed weight break¬
down for the mechanically driven single-point hoist with a usable cable
length of 100 feet and the hydraulically powered four-point hoists is
presented in Table XXXV. For comparison with the weight estimates of
Phase I, the weight for an 30-foot single-point hoist is included in
parentheses.
The calculated weight of the complete single- plus four-point system is
4974 pounds, including controls wiring and aircraft supporting structure
peculiar to the hoist system. Of this, approximately 2200 pounds (single¬
point hoist and input drive shaft) is readily removable when missions re¬
quiring minimum aircraft empty weight are to be undertaken utilizing four-
point suspension. Similarly, if only the single-point hoist is to be used,
the four-point hoists c*n be removed, providing a weight reduction of 2236
pounds. The weight of the cargo handling system chargeable to aircraft
empty weight is then:
Single-Point Mission (four-point removed) - 2738 lb
Four-Point Mission (8lngle-^>oint removed) = 2704 lb
TABLE XXXV
WEIGHT SUMMARY-
AO. OOP-POUND EXTERNAL CARGO HANDLING SYSTEM
Weight
(lb)
Item Component _ Assembly _ System
Single-Point Hoist System 2406
(2133)*
Single-Point Hoist
2270
(I960)*
Drum and Bearings
353
Gearing and Housings
589
Supports and Bearings
180
Level Wind Assembly
Potentiometer, Cable
257
Cutters
14
Anti-Backlash Cover
58
Lubricating Oil (4 gal)
28
Free Reel Unit and Controls
71
Cable
440
Hook, Swivel, & Slip Ring
150
Decoupler (Isolator)
130
Drive Train
120
Clutch-Reverser
46
(173)*
Upper Angle Gearbox
21
Lower Angle Gearbox
25
Shafting and Bearings
28
Control Unit, Display Wiring, Etc.
16
(0)*
Four-Point Hoist System
Four-Point Hoist
559
Drum and Bearings
89
Gearing and Housings
184
Level Wind and Supports
Potentiometers, Potentiom¬
48
eter Clutches, Switches,
Cable Cutters
13
Anti-Backlash Cover
16
Lubricating Oil (2 gal)
14
Free Reel Unit
19
Cable
76
Hook, Swivel, & Slip Ring
52
Isolator
38
Hydraulic Motor
10
*
I
4
216
TABLE XXXV (continued)
r
Weight
(lb)
Item
Component
Assembly
Hydraulic Subsystem**
111
Hydraulic Pump
20
Lines, Fittings, Fluid
90
Filters
18
Flow Divider /alvea
12
Relief Valve
2
Shutoff Valves
4
Free Reel Control
36
Display, Control Boxes, & Wiring
35
Structure (2 Davits)
150
System
Single- Plus Four-Point Cargo Handling System 4974
^Weights in parentheses are those estimated in
Phase I for hoist with 80 feet of cable.
**Hydraulic pump, filters, relief valve, and a portion
of the lines and fittings will also be used for
engine starting system. Weight chargeable to engine starting
is 35 pounds.
\
217
MAINTAINABILITY AND RELIABILITY
INTRODUCTION
A comprehensive maintainability and reliability study was made of the
heavy lift cargo handling system. This study was based on the system de¬
scription of pages 134 through 156 and the drawings of Appendix III. The *
reliability and maintainability characteristics indicate that a combined
system reliability of .969 can be expected, with an estimated .0718 main¬
tenance man-hours per flight hour. A failure mode and effect analysis was
also completed.
i
Considerable effort was also expended in a study of the safety aspects of
the system; primary emphasis was placed on capability of jettisoning a
load carried on the four-point suspension system. A reliability block dia¬
gram of the cable cutter system was prepared. This diagram shows that the
least redundancy occurs at the explosive charge and cutter and indicates
that primary emphasis should be placed on adequate testing of these com¬
ponents to achieve maximum safety.
RELIABILITY AND MAINTAINABILITY CHARACTERISTICS
The selection of a mechanical drive for the single-point hoist and a hy¬
draulic drive for the four-point hoists greatly increases the reliability
and maintainability of the cargo handling system. Detail design, based on
field experience with the cargo handling system on the CH-54A,has served
to maintain these qualities, aB shown in Table XXXVI, page 220.
FAILURE MODE AND EFFECT ANALYSIS
The results of a failure mode and effect analysis are sumnarized in Table
XXXVII, page 222. Failure modes considered were only those that the re¬
liability analysis indicated were the most probable. In all cases, the de¬
sign is such that these failure modes are minimized, or features are in¬
cluded in the basic design of the system, to avoid detrimental effects on
the mission.
♦
SAFETY CONSIDERATIONS
Considerable time has been spent studying the safety aspects of the cargo
handling systems. The single most critical consideration is the jettison¬
ing of a four-point load. Should one of the attachment points fail to be
released, the results could be catastrophic. Because of this, tandem-dual
cable cutters at each hoist are used. The wiring is redundant to each
cable cutter, and further, the redundant wiring is routed through the air¬
frame in such a way as to reduce the vulnerability of the cutter charge
ignition circuit significantly. This redundancy is required to ensure the
proper level of safety.
218
WIRING
Reliability Block Diagram of Cable Cutter System.
TABLE XXXVI
RELIABILITY AND MAINTAINABILITY CHARACTERISTICS
Single -
Point
System
Four-
Point
System
Combined
System
Mission Reliability *
.995
.986
.9999**
System Reliability *
.985
.933
.969***
MTBUMA, Hours
32.5
7.2
15.8
Mean Time to Repair, Hours
.383
.364
.377***
Maintenance Burden, MMH/FH
.039
.148
.0718***
Inherent Availability, Percent
99
96
97.8**»
♦Based on a 30-minute mission and two conplete cycles.
♦♦Assumes complete redundancy of systems,
♦♦"Based on a 70-30 distribution of single- to four-point missions.
The ignition circuit is provided with a built-in test circuit which is
used on the preflight check to ascertain that all circuits are functional.
Figure 64 is a reliability block diagram of the cable cutter system pro¬
vided for each hoist. This diagram shows the redundancy employed in each
of the systems. Note that the electrical source and the control switch
are shared for each hoist.
In this diagram each parallel path provides a successful operation for
that particular function. The diagram for the successful cutting of all
the cable in the four-point system would shcrv: four of the diagrams of
Figure 64 in series.
From the diagram it can be seen that the least redundancy is at the charge
and cutter. Therefore it is recommended that adequate testing of these
components be undertaken to ascertain with a high degree of confidence that
their reliability is consonant with the overall reliability desired of this
system.
INSTALLATION AND REMOVAL
Singla-fpint Hoist
Removal is accomplished simply by disconnecting the input shaft, loosening
220
6 bolts, and disconnecting the electrical lines at a quick disconnect
fitting. The hoist is lowered from the aircraft by means of a special
support equipment bridle and cable attached to the lifting points on the
hoist support structure. The cable passes through airframe mounted pulleys
and can be attached to a truck or wheeled vehicle or vehicle winch to
lower the hoist to the ground.
Four-point hoist removal is accomplished by loosening 4 bolts and dis¬
connecting with hydraulic and electrical quick disconnects. The four-
point hoists are lowered by means of the same cable bridle system, which
is supported over a similar pulley arrangement, as that utilized for the
single-point hoist.
221
! ' ■ X JT-ttOtHW VI IjitU trik
0)
©
I
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224
8ingle-point hoist.
I
*
*
GROWTH POTENTIAL
Both the mechanically driven single-point hoist system and the hydrauli¬
cally driven four-point hoist system have potential for increased capac¬
ity. Growth to 50,000 pounds for these systems can be accomplished in
incremental stages of modification and payload increase. The following
paragraphs summarize the stages of improvement or "break-points", the modi¬
fications necessary, and the estimated weight increase.
SINGLE-POINT HOIST SYSTEM
Design Data :
Capacity (pounds)
Ultimate Load Factor
Usable Cable Length (feet)
Cable Speed (ft/min)
Power Requirements (HP)
Stage 1 Growth
40,000
3.75
100
60
94.2
Capacity « 45,000 lb
The first "break-point" in the single-point hoist system growth is in the
ultimate (breaking) strength of 171,600 pounds for the single-point hoist
cable (see page 187). For an ultimate load factor of 3.75, the maximum per¬
missible cable load is
A
cable
171.600
3.75
45,800 lb
The power required at 45,000-lb capacity is 106 HP. At this load and
power level, the only modifications necessary in the single-point hoist
system are as follows:
1. Drum (p. 30)
Increase thickness from I .44 to
1.65 inches.
2. Planet Pinion Bearings
(p. 188)
Increase bearing size from 114
series to 212 series for increased
static capacity.
3. Level Wind Ballscrew
Bearings (p. 189)
Increase taper roller bearing size
for increased static capacity.
The estimated weight increase for the stage 1 growth is 62 pounds.
Stage 2 Growth Capacity *= 50,000 lb
To obtain a capacity of 50,000 pounds at the 3.75 load factor, a 1.52-
inch-diameter 18 x 19 cable is required. Maintaining the same drum diam-
4
i,
225
ater and width reduces the useful cable length to 90 feet. The power re¬
quired is 118 HP.
The necessary modifications to the hoist are
1. Cable (p. 187)
2. Cable Drum (p. 30)
3. Planet Pinion Bearings
(p. 188)
4. Level Kind Balls crew
and Nut (p. 172)
5. Level Wind Ballscrew
Bearings (p. 189)
The estivated weight increase for the
New 18 x 19,1.52 inch diameter.
Increase thickness to 1.85 inches
and change pitch to 1.625 inches.
Same as Stage 1 change.
Change pitch to .541 inch.
Same as Stage 1 change.
above modifications is 170 pounds
*
FOUR-r PINT HOIST SYSTEM
Design Data:
System Capacity (pounds)
40,000
Hoist Capacity (pounds)
11,550
Ultimate Load Factor
4.2
Usable Cable Length (feet)
50
Cable Speed (ft/min)
30
Power Requirements (HP)
Hydraulic Requirements
84.8
Pressure (psi)
Flow (fpm/hoist)
3500
11.1
Stage 1 Growth System Capacity
Hoist Capacity
48,000 lb
13,900 lb
Ae was the case with the single-point hoist system, four-point capacity
growth is paced by the ultimate (breaking) strength of the cable. The
♦79-inch cable nas a breaking strength of 58,400 pounds (see page 212),
giving a dynamic capability of 13,900 pounds. The modifications to the
system to achieve this capability are
1. Drum (p. 30) Increase thickness from 0.81 to
0.97 inch.
2. 1st Stage Planet Pinion Increase bearing size from 1911 to
Bearings (p« 213) 110 series for increased static
capacity.
k
226
»
f.
3. 2nd Stage Planet Pinion
Bearings (p. 213)
Increase bearing size from 110 to
208 series for increased static
capacity.
4. Planetary Carrier Plate
(r. 207)
Increase plate thickness to .31 to
reduce deflector.
5. Hydraulic System
Increase pressure from 3500 to 4000.
The estimated weight increase for the above modifications is 25 pounds per
hoist.
Stage 2 Growth System Capacity = 50,000 lb
Hoist Capacity * 14,450 lb
To achieve a four-point hoist system capacity of 50,000 pounds at the 4.2
load factor, a cable with a breaking strength of 60,700 pounds is required;
inasmuch as this is such a minor increase over the actual cable strength
of 58,400 pounds,a small reduction in ultimate load factor is recommended.
The load factor at 50,000 pounds is
Ultimate Load Factor * " 4.03
14,450
In addition to increasing the hydraulic system pressure to 4000 psi (as is
required for Stage l), it will also be necessary to increase the hydraulic
motor displacement by 12 pet by reboring the existing motor.
227
•
• tafrv*
; SSR^'r*'
I
f
1
COMPONENT AMD SYSTEM DEVELOPMENT PLAN
DISCUSSION
The following paragraphs outline the development effort suggested for the
heavy lift cargo handling system of this report. For the endurance test
phase of both the single- and four-point hoists, the construction of a
special test facility that can be used for both the single- and fo\ir-
point hoist systems will be required. No attempt has been made to provide
a schedule for the fabrication of such a facility, although its estimated
costs are presented on page 230.
While the plan presented is directed toward the systematic development of
the full-size aircraft hardware, several areas are suitable and are rec¬
ommended for development effort in advance. Among these are the investi¬
gation of synchronous operation of ths four-point hoist systam (where the
hydroelectrical feedback system can be developed in model fora) and the
reliability of hook release of both the single- and four-point hoists.
TEST PROGR AM
The following test programs are recommended to develop and qualify the
heavy lift external load handling system.
Mechanically Driven Single-Point Hoist
All components of drive train and single-point hoist will be utilized in
all phases of the test program except the static load test.
1. Static Load Test:
Conduct a static load test on single-point hoist
to limit load (100,000 lb)
2. Endurance Test:
Conduct a 3600-cycle endurance test at the following
load spectrum:
Cycles
Distance (ft)
Load (lb)
Cable Position
800
100
40,000
vertical
2400
100
30,000
vertical
200
100
20,000
vertical
100
100
0
vertical
50
100
30,000
15° lateral
228
Cycles Distance (ft) Load (lb) Cable Position
50 100 30,000 15° aft
3. Environmental Test:
Conduct an environmental test on the single-point
hoist and drive train.
High Temperature Test
Low Temperature Test
Sand and Dust Test
Rain Test
Note: All tests to be in accordance
with MIL-E-5272C
Hydraul ^ ^lly Driven Four-Point Holsts
1. Static Load Test:
Conduct a static load test on a four-point
hoist to limit load (29,000 lb).
2. Endurance Test:
Conduct a 3600-cycle endurance test on a four-point
hoist at the following load spectrum:
Cycles
Distance (ft)
(lb)
800
50
11,550
13
11,550
2400
50
8,650
13
8,650
200
50
5,800
13
5,800
200
50
0
13
0
3* Environmental Test:
Conduct an environmental test on the four-point hoist,
control system, hydraulic pump, motor, and lines to
the following requirements:
High Temperature Test
Low Temperature Test
Sand and Dust Test
Rain Test
229
Note: All teste to be In accordance
with HIL-E-5272C
4. System Test:
Conduct a 200-cycle test to demonstrate synchronous
operation of the four-point hoist system and the
adequacy of the hydroelectrical feedback control
system at varying load conditions.
Hook Swivel Assemblies
1. Qualification and Environmental Tests: i
Conduct a 3600-cycle release test on both the
single- and four-point hook assemblies.
Conduct separate environmental tests at the
conditions outlined for hoists.
2. Functional Test:
Conduct a hook functional test in conjunction
with respective hoist endurance test.
ESTIMATED DEVELOPMENT COSTS
Based on preliminary data available at this time, it is anticipated that
the configuration recoamtended can be developed for the estimated prices
set forth below:
Single-Point
Four-Point
Engineering Design
$270,000
$240,000
Prototype Hardware
225,000
250,000
Tooling
100,000
40,000
Ground Tests
50,000
70,000
Test Facility (100-ft height) $300,000
230
I
SCHEDULE
Single-Point Hoist
Design
Fabrication
Development Tests
Endurance
Environmental
Static
Months
Four-Point Hoist
Design
Fabrication
Development Tests
Endurance
Environmental
Static
System Test
Hook & Snivel Assemblies
Design
Fabrication
Qualification Test
CONCLUSIONS
1. A separata function astern that incorporates a mechanically
driven single-point hoist and four hydraulically driven
multi-point hoists best meets the external cargo handling
requirements of a 40,000-pound-pay load, single-rotor heavy
lift helicopter on the basis of reliability, safety, main¬
tainability, redundancy, versatility, and technical confidence,
2. The single- plus four-point system described above weighs
4974 pounds. For missions requiring single-point load sus¬
pension, the four-point hoists are readily removable, making
the cargo handling system weight chargeable to aircraft weight
empty 2738 pounds. Similarly, for four-point missions re¬
moval of single-point components results in a cargo system
weighing 2704 pounds.
3. The mechanically driven single-point plus the hydraulically
powered four-point hoist system has a capability of growth
to 50,000 pounds. All major components of both hoists except
the cable drum have adequate margins of safety to accommodate
a 25-pet growth. The single-point system can be up-rated to
50,000 pounds for 170 additional pounds. The weight increase
for the four-point system is 100 pounds.
4. With the use of a hydroelectrical feedback control system, the
total cable length error (difference between individual cable
lengths) in 50 feet of travel of the four-point system ie
7-3/4 inches. This error may be further reduced to 4 inches
by a one-time calibration procedure, A study should be under¬
taken to determine if these limits are within acceptable limits
for equipment in the U.S. Arny inventory,
5. The hydroelectrical feedback system proposed for the four-
point system lends itself to evaluation and development in
model form. Further effort in this area is re commended.
6. Thin investigation has led the writers to believe that acquisi¬
tion and release of loads carried by four-point suspension will,
for the most part, be accomplished with aircraft cn the ground
or from a "wheels light" hover. It is therefore concluded that
an analysis of mission requirements is in order to determine if
some other four-point cable length might better meet the heavy
lift multi-point requirement. The limited investigation per¬
mitted within the scope of this contract has led the investi¬
gators to believe that a major portion of heavy lift helicopters
should be equipped with powered four-point hoists with a maximum
capability not exceeding 20 feet. Further study in this area
is definitely required.
232
4»ir
7. The use of manually actuated control valves which divert fluid
from the aircraft utility hydraulic eyetem to provide the power
to operate the cable free reel clutchee in the e ingle- and the
four-point hoists meets the requirement for a pilot actuated
manual load release system.
8. In-flight release of loads under emergency conditions is more
safely accomplished in the single-point mode (for an equal
number of release components) where electrical, free reel,
auto touchdown, and two explosive release systems are pro¬
vided. For the four-point system, only the use of the tan¬
dem-dual cable cutters is recommended for in-flight release
of loads under emergency conditions. The use of the free
reel system to release loads in the four-point mode would be
practical if the primary release system failed at the time that
the aircraft wae hovering just off the ground at the drop
site, however.
9. The evaluation of current production cargo hook-swivel
assenfcliee design indicates that further efforts are required
to achieve the degree of reliability required* The swivel
design and method of sealing the slip ring as proposed herein
should add appreciably to the reliability of theee components.
A comprehenisve program to study and evaluate the reliability
of pilot controlled load release would be invaluable to the
aerial cargo handling system designer.
BIBLIOGRAPHY
1. AFBMA Standards. Methods of Evaluating Load Ratings fo r Ball
Bearings, Section No, 9. The Anti-Friction Bearing Manufacturers
Association, Incorporated, New York, New York, Oct. I960.
2. AFBMA Standards, Methods of Evaluating Load Ratings for Roller «
Bearings, Section No, 11 , the Anti-Friction Bearing Manufacturers
Association, Incorporated, New York, New York, July I960.
3. Burroughs, Lester, R., Power Transmission Studies for Shaft -
Driven Heavy Lift Helicopters . USAAVLABS Technical Report 65-40. A
U.S. Army Aviation Materiel Laboratories, Fort Bustis, Virginia,
Oct. 1965.
4. Dutton, Walter, J., Paramatr <« Anni vm |b and Pra Ha^n*™ - Dgs^gfl
of a Shaft Driven Rotor System for a Heavy Lift Helicopter.
USAAVLABS Technical Report 66-56, U.S. Army Aviation Materiel
Laboratories, Fort Eustis, Virginia, February 1967, p. 14.
5. Engineering Information Bulletin No. 500. Raybestos - Manhattan,
Inc., Stratford, Connecticut, 1954 pp. 28 - 40.
6. Faupel, J.H., Engineering Design. John Wiley and Sons, Incor¬
porated, New Jerk, New York, 1964, pp. 605 - 608.
7. Harter, Russel W., "New Equations Simplify Design of Mechanical
Load Brakes”, Product Engineering . Voluas 19, September 13, 1965,
pp. 140 - 143.
8. Hope, John, A., "How to Design With Precision Ball Screws”,
Product Engineering. Volume 9, April 24, 1961, pp. 59 - 64.
9. MIL-HDBK-5A, M aterials and Elements for Aerospace
Vehicle Structures. February 8. 1966.
10. MIL-H-8501A, Helicopter Flying and Ground Handling Qualities :
General Requirements for . April 3. 1962. p. 3.
11. Saginaw B »T\ Ber ing Screws and Ball Bearing Splines. Saginaw *
Steering Gear Div., General Motors Corp., Saginaw, Michigan,
p. 14.
12. Shigley, Joseph, E., Machine Design . McGraw-Hill Book Co.,Inc.,
New York, New York, 1956, pp. 447 - 452. *
13. The Timken E ngineering Journal. Section 1. The Timken Roller
Bearing Company, Canton, Ohio, 1964.
234
14* U.S. Arssy Combat Developments Command, Essential Elements of
Analysis to Support QMR for a Heavy Lift Helicopter. Study:
EEA No. 6 Annex A, September, 1965, pp. A-l- A-5.
15. Wahl, A.M., Mechanical Springs , Second Edition, McGra*41111 Book
Company, Inc., New York, New York, 1963, pp. 155 - 162.
APPENDIX I
SURVEY OF MILITARY VEHICLES
The 94 vehicles lifted in Table XXXVIII represent the U.S. Army equipment
considered in the design of the cargo handling system of this study. It
is based on a survey conducted by the U.S. Arny Combat Developments
Command (Reference 14)«
An item number has been assigned to each piece of equipment to provide a
more convenient form of reference within the report. This procedure was
necessaxy because the official "line number" of all the equipment was not
known.
The equipment is listed in order of increasing weight. To prevent the
table from becoming unwieldy, a special set of appreviations was devised.
These abbreviations, listed below, apply only to Table XXXVIII.
List of Abbreviations
ABN.
airborne
ARB.
ambulance
AMMO.
ammunition
AMPH.
amphibious
ARMD.
armored
ASLT.
assault
AUX.
auxiliary
A VLB.
assault vehicle launch bridge
CARR.
carrier
C-C
combat weight
CGO.
cargo
CMD.
command
DRVN.
driven
DSL.
diesel
ENCR.
engineers
EftUIP.
equipment
EXP.
experimental
FT.
full track
GAL.
gallon
GEN.
generator
H
height, inches
HON.
howitzer
IND.
industrial
236
»
I
*
L
length, inches
LT.
light
LWB.
long wheel base (trailer)
mm
millimeter
MED •
medium
MTD.
mounted
MTZD.
motorized
MULT.
multiple
OBS.
observation
PERS.
personnel
PROP.
propelled
RD.
road
RECONN.
reconnaissance
R/M
reduced height
RKT.
rocket
SP.
self propelled
SPD.
speed
STLR.
semi-trailer
SUP.
supply
SVC.
servicing
SWB.
short wheel base
T
ton
TIB.
trailer
TRACT.
tractor
TRANSP.
transporter
TRK.
truck
TRKD.
tracked
VEH.
vehicle
W
width, inches
WHID.
wheeled
WKR,
wrecker
WPH.
weapons
WWN.
with winch
XIWB.
extra long wheel base
ID.
yard
337
Figure 65* Typical Military Vehicle.
tabiz mrai
1
M-100
TLH. AMPH. COO.
1/4-T
0.283
0.533
109
57
42
2
M-14
CART RKT.TRANSPT.
318 m
0.340
0.340
221
58
31
58
3
H-274
CARR.LT.WPH.
l/2-T
0.398
0.985
117
46
28
-
4
-
PROP.,EQUIP.
AUZ. HCW.
0.449
-
-
-
-
-
5
H-91
LAUNCHER, MULT.
RET. 115 an
0.625
0.625
152
90
67
-
6
XM-34
LAUNCHER, RET.
318 an
0.659
1.117
126
62
50
03
7 M-101 TIZ.,CG0. 3/4-T 0.670 1.420 147 74 85 53
6 OH-13 HELICOPTER, OBS. 0.859 1.400 365 90 143 112
TAELS mm IsanUnuall
Weight
Item
No.
Model
No.
Designation
Net C-C
(Tone)
L
W H
(inches)
R/H
9
M-151
TRK. UTILITY
1.137
1.417
132
63
71
53
10
M-332
TLR., AMMO.
1.175
2.225
153
86
56
-
11
T-3
TRANSPORTER,
LIQUID 1000 GAL.
1.200
4.200
155
102
64
-
12
M-200A1
CHASSIS TLR. GEN.
2-1/2-T
1.205
4.705
165
93
40
-
13
M-149
TLR., WATER
1-1/2-T
1.300
3.948
157
80
73
-
14
M-105A2
TLR., CGO.
1-1/2-T
1.325
2.700
166
83
98
58
15
M-38AIC
TRK., UTILITY
1/4-T
1.333
1.733
109
57
42
-
16
-
TLR., BASIC
UTILITY 2-1/2-T
1.350
4.850
198
98
44
-
17
M-170
TRK., AMB. 1/4-T
1.482
1.767
155
61
79
-
18
XM-102
HON., 105 nan
TOWED- SP.
1.530
1.530
-
-
-
-
19
M-242
HCW., 105 nan
TOWS)
2.490
6.350
236
85
62
-
20
UH-1B
HELICOPTER,
UTILITY
2.328
3.600
461
103
157
90
21
UH-1D
HELICOPTER,
UTILITY
2.328
3.600
481
103
157
90
22
M-37
TRK., CGO. 3/4-T
2.850
3.600
185
74
90
64
23
M-37
TRK., CGO. 3/4-T
WWN.
3.000
3.725
190
74
90
64
24
M-146
STLR., VAN SHOP
3.400
6.000
276
95
136
-
25
-
TRK., PORK LIFT
2-T
3.400
3.400
-
-
-
-
26
M-118A1
STLR., STAKE 6-T
3.570
9.570
276
95
104
38
239
TABLE mvin (cQliUnufld)
Item
No.
Model
Mo.
Designation
Weight
Net C-C
(Tone)
T~
Dimensions
W H
(Inches)
27
M-43
TRK., AMB. 3/4-T
3.585
4.275
206
75
92
-
28
M-119A3
STLR., VAN,
CGO. 6-T
3.590
9.590
275
96
135
-
29
M-405
HANDLING UNIT,
762 ran RKT,
4.315
4.315
338
96
130
81
30
-
TLR., LCW BED
8-T
4.915
12.915
281
102
58
-
31
M-9
BULLDOZER
5.000
5.000
146
36
78
-
32
0V-1
AIRPLANE, OBS.
5.328
6.659
525
no
156
-
33
M-273
TRK., TRACT.
2-1/2-T SWB.
5.590
5.590
226
94
98
81
34
M-345
TIR., FLAT BH)
10-T
5.630
15.630
536
98
56
-
35
M-48
TRK., TRACT.
2-1/2-T IWB.WWN.
5.921
5.921
254
94
98
82
36
M-35
(XM-410E)
TRK., CGO, 2-1/2
-T LWB.
6.233
11.408
262
96
H2
86
37
M-56
GUN FT., 90 mm
ABN.
6.250
8.750
241
86
88
-
38
M-114A1
HCW., 155 ran
TOWED
6.350
6.350
288
96
81
a»
39
M-35
TRK., CGO 2-1/2-
T DWB-WWN.
6.440
11.790
276
96
112
86
40
M-129
STLR., VAN SUP.
12-T
6.750
18.750
345
96
140
-
a
M-49C
TRK., TANK, FUEL
SVC. 2-1/2-T
6.978
6.978
262
96
98
90
42
-
SCRAPER, TOWED
7-1/2 ID.
7.060
7.060
-
-
-
-
43
M-49C
TRK., TANK, FUEL
SVC. 2-1/2-T WWN.
7.200
7.200
277
99
130
-
TABLE XXXVIII (continued)
Item
No.
Model
No.
Designation
Weight_
Net C-C
(Tone)
_ D iqengtanp.
L W H
(Inches)
r/h
44
-
LOAD EE, SCOOP TYPE 7.200
1-1/2 YD3
7.200
209
84
81
-
45
M-127A1
STLR., STAKE 12-T
7.200
19.400
345
97
109
58
46
M-131A3C
STLR., TANK, FUEL
SVC.
7.400
7.400
353
98
no
-
47
M-172A1
STLR. LOW BED
25-T
7.430
32.430
141
115
68
-
48
MT 2D
GRADER, RD» MTZD •
DSL. DRVN.
7.460
7.460
264
81
85
-
49
M-114A1
ARMD. RECONN.
CARRIER -FT.
7.500
7.500
169
92
80
-
50
M-313
STLR. VAN, EXP.
6-T
7.500
323
98
134
-
51
M-220
TRK., VAN SHOP
2-1/2-T
7.543
10.043
267
96
131
-
52
M-342
TRK., DUMP 2-1/2
-T
7.583
10.083
273
96
101
83
53
M-342
TRK., DUMP 2-1/2
-T WWN.
7.790
10.290
273
96
100
100
54
M-109
TRK., VAN, SHOP
2-1/2-T WWN.
7.823
10.291
277
99
130
-
55
•
TRACTOR, FT., LCW
SPEED, DSL. DRVN.,
LT.
7.988
7.988
175
99
78
*
56
CH-47A
HELICOPTER, CGO.,
MED •
8.000
16.500
600
145
222
-
57
M-129
STLR., VAN SUP.
12-T
8.010
30.010
345
96
140
-
58
-
TRACTOR, WHLD. IND
DSL. DRVN. LT.
.8.050
8.050
194
96
90
-
59
-
TRK., FORK LIFT
3-T
8.400
8.400
-
-
-
241
TABLE XXXVIII (continued)
Weight
Dimensions
Item
No.
Model
No.
Net C-C
Designation (Tons)
L
W H
(inches)
“r7h
60
M-270A1
STLR., LOW BED 8.750
WKR. 12-T
20.750
597
97
121
80
61
M-52
TRK., TRACTOR 9.200
5-T SWB.
9.200
258
97
107
87
62
M-292
TRK., VAN EXP. 9.500
2-1/2-T
12.000
329
97
139
-
63
M-54
(XM-656)
TRK., CGO. 5-T 9.616
IWB.
14.616
299
97
116
86
64
M-113
CARRIER, PERS. FT. 9.878
11.308
192
106
80
-
65
M-108
TRK., WKR., CRANE 9.893
2-1/2-T WWN.
1C.143
303
96
100
-
66
M-54
(XM-656)
TRK., CGO., 5-T 9.973
LWB, WWN.
14.973
314
97
116
86
67
M-577
CARRIER, CMD.POST 10.700
LT. TRKD.
11.650
192
106
106
-
68
M-51
TRK., DUMP 5-T 11.333
WWN.
16.333
282
98
111
88
69
XM-106
MORTAR, SP.,FT. 12.538
4.2 in.
192
106
80
mm
70
M-60
TRK., WKR„LT. 11.980
2-1/2-T
12.980
303
96
101
-
71
M-78
HEAT & TIEDOWN 12.Q32
UNIT 762 mm RKT.
12.032
370
96
95
-
72
-
LOADER, SCOOP TYPE 12.200
DSL. DRVN.1-1/2
YD3
12.200
248
105
97
MB
73
GRADER, RD.,MTZD.,12.390
DSL. DRVN.
12.390
311
96
111
93
74
M-139
TRK., STAKE 5-T 13.400
BRIDGE TRANSP.
13.400
369
114
114
-
75
M-123C
TRK., TRACTOR 14.200
10-T
14.200
280
114
113
92
242
TABLE XXXVIII (continued)
1
I
Weieht
Dimensions
Item
No.
Model
No.
Designation
Net C-C
(Tons)
L
V h
(inches)
fc/Jl
I.
76
CL-60
AVIB.
14.300
14.300
338
158
73
-
V
p.
1
f
77
LOADER, SCOOP
TYPE DSL. DRVN.,
2-1/4 YD3
14.414
14.414
248
105
97
t
l
I
1.
1
78
M-551
ARMD. RECONN./
ABN. ASLT.
VEHICLE
15.000
15.000
252
115
95
-
1
i
79
M-115
HOW. 8-in.
TCWED
15.288
15.288
432
112
108
-
l
f
1
80
—
TRACTOR, FT.
DSL. DRVN. LOW
SPD., MED.
15.666
15.666
196
116
88
-
\
1
1.
81
M-162A1
STLR. LOW BED
60-T
16.348
76.348
441
144
81
-
i
i
i
82
M-246
TRK., TRACT.
XLWB. WWN.5-T
16.415
16.415
352
98
132
89
$3
M-62
TRK., WKR., MED.,
5-T WWN.
16.700
16.700
310
97
103
-
84
M-386
LAUNCHER, 762 mm
TRK., MTD.
17.291
17.291
389
114
105
-
i
85
M-572
HANDLING UNIT,
318 mm TRK. MTD.
19.850
19.850
339
96
97
-
1
i
86
M-84
(XM-106)
MORTAR, SP., FT.,
4.2 in.
20.561
20.561
221
129
109
-
i
1
j
87
M-15A2
STLR. TRX.,
TRANSPORTER
21.300
71.300
462
146
105
-
*
f
i
•
88
(XM-551)
TANK COMBAT,
FT. LT. GUN
76 mm (ARMD.
RECONN./ABN.
ASLT. VEH.)
25.400
25.900
280
126
122
108
!.
89
M-44
(M-109)
HOW. SP., FT.,
155 nm
29.000
29.000
325
140
134
127
TABLE XXXVIII (continued)
Item
No.
Model
No.
Designation
Weight
Net C-C
(Tons)
L
Dimensions
-R— ~
(Inches)
~r7h
90
M-55
(M-110)
HCW.GP. FT.
8 in.
45.000
45.000
325
140
146
117
1
91
M-60
TANK, COMBAT
FT. 105 mm
47.150
47.150
366
143
126
-
92
M-60
LAUNCHER, A VIE
47.650
47.650
274
143
127
-
1
93
M-102
COMBAT, ENGR.
VSH.
51.800
55.000
337
148
122
-
94
M-88
TANK RECOVERY
VEH. MED.
54.000
56.000
326
135
127
mm
*
244
APPENDIX II
MECHANICAL VARIABLE SPEED DRIVE
INTRODUCTION
A mechanical variable speed drive concept, designed and developed by the
Lycoming Division, AVCO Corporation, was investigated as an alternate to
the clutch-rever3er unit of page 137 as the drive for the mechanically
powered single-point hoists.
This concept provides an infinitely variable bidirectional output rotation
mechanism. The unit proposed for the hoists of this study is a modified
form of the traction mechanism employed in the constant speed drive units
for A.C. generating systems used on the Navy A4E. Over 1000 of these
units have been produced and approximately 500,000 operational hours have
been accumulated.
DESCRIPTION
The proposed actuator combines an epicyclic gear differential coupled to
an infinitely variable ratio traction transmission. This combination re¬
sults in a unit capable of infinitely variable, stepless, bidirectional
output rotation.
The variable ratio section consists of two flywheel members, called to¬
roids, which are concentric with the drive shaft. Their dished surfaces
form a toroidal space and contain four rolls, mounted in yokes and fas¬
tened to a fixed cage. A precalibrated load bolt squeezes the toroids
against the rolls and provides the traction for power transmission. Speed
ratios are changed by varying the angular position of the rolls with re¬
spect to the drive axis. The control rod motion can be linearized with
respect to output speed if desired. The differential section is an epi-
cyclic gear train and is so designed as to permit equal and opposite out¬
put speed. The planet cage rotates as a function of the position of the
rolls in the ratio change section.
Table XXXIX gives the applicable design data of the unit. Figure 67,
page 249 . gives the output power and torque versus rpm relationship.
It should be noted that added cooling is required for this unit. It will
be provided by an electric motor driven blower/heat exchanger unit of the
type used on the CH-3C and CH-53A. This unit will be interlocked so as to
function only when the variable speed drive unit is in operation. Figure
66 is included to she* the physical dimensions of the complete unit, less
heat exchanger.
It has been estimated that it would require approximately 16 months to
complete the design, fabrication, and prototype developmental testing
required prior to delivery.
TABLE XXXIX
DESIGN DATA,
MECHANICAL VARIABIZ SPEED DRIVE
Input Speed
7000 rpm
Output Speed
Plus or minus 1000 rpm
Power Capability
225 HP maximum (lifting load)
115 HP maximum (lowering load)
Control Power (max)
.01 HP at maximum acceleration
rate of 1300 rpm/sec
Control Force
45 lb at control rod for
maximum acceleration
Efficiency
93 pet
Lubrication
MIL-S-81087 (Weps) Type 1
Oil Flow
20 lb/min
Cooling Requirements
500 Btu/min
Operating Temperature
-65°F to 350°F
Weight (Estimated)
Dry weight with
integral oil
reservoir and pump
107 lb
Lubricant
6 lb
V
246
APPBJDH III
CARGO RANDLItC STSTB< DRWIHCS
I
I
VIA IN GEARBOX
PULLEY -HOiST
removal/ installation
CABLE - HOIST
RFV’VAL/ installation
n
BEARING SUPPORT
Figure 68. Single-Point Holst Installation, Single-Rotor H.L.H.
4
251
%
INCHES
0 12 3 4 5
i I I I _I
S'ALE
DECOUPLER / ISOLATOR
M-'ES
0 5 10
SCALE
•—- C.OU°LMG (TrPiCAL)
BEAP-NG SUPPORT
VISCOUS DAMPED
CiuICH PEVFRSEP UNi I
LiPPER ANGLE BOX
•OPJVE SHAFTING I
B
CLUTCH REVERSER UNIT
BEAP'KiG SUPPORT
VISCOUS DAMPED
✓
SCALE
DECOUPLER / ISOLATOR
10
.E
INCHES
0 12 3 4 5
U- J i I_I_I_I
S'ACE
253
HOIST REMOVAL PULLEY
FOUR POINT HOIST
return to ^
UTILITY HYDRA'
SYSTEM
f
&
EMOVAL PULLEY
FOUR POINT HOIST
V
PUSH-PULL CABLE
TO AFT FACING
PILOT
V.
I si
RETURN TO /yt ■ — ^
UTILITY HYDRAULIC
SYSTEM
i
5 p
‘Ls
FROM A/C UTILITY HYDRAULIC SYSTEM
—FREE REEL MODE SELECTOR VALVE
- - 10 SINGLE POINT HOIST
MANUAL CONTROL VALVE
MANUAL CONTROL VALVE
TO FREE REEL CLUTCH
INCHES
0 50 100
lihlllihl
SCALE
CLUTCH-REVERSER UNIT
INCHES
0 5 10
Liiiihl n 111
SCALE
Figuro 70. Singlo- Plus Four-Point
- HOIST INPUT DRIVESHAFT
Installation, Tsnrti-ftotor H.L.H
INTERCONNECTING
SHAFTING
FOUR-POINT HOIST
HOIST SUPPORT
SINGLE-POINT HOIST
SINGLE--POINT HOIST
—TO MODE SELECTOR
VALVE
'H r"7
PUSH-PULL CABLE
TO AFT FACING
I I / PILOT
• I / I
PILOT M -MANUAL CONTROL
I I / R/ VALVE
-CJ t CJ 1
l , \ I!
k \\^y / / TO FREE REEL-l—RETURN TO A/t
\ \ CLUTCH UTILITY HYDRAULIC
\\ \ / A \ SYSTEM
i
*4
*
l
K
REACTION PIPE —
BALL SCREW
WEAR LINER-
_LI
INPUT SHAFT
DECOUPLER
REACTION ARM
FREE REEL —
CLUTCH ASSY
WESTON BRAKE
CABLE
- DUAL - TANDEM
CABLE CUTTERS
BELLMOUTH ASSY
BALL SCREW
REAL I ION Pll
Figure 71* Single-Feint Holst.
—TO MODE SELECTOR
VALVE
SECTION A-A
ANTI - BACKLASH COVFR
' —CABLE LENGTH POTENTIOMETER
-SLIP RING ASSY
259
■COOLING OIL
FITTING (TYP)
/—HIGH PRESSURE
OIL FITTINGS (TYP)
CLUTCH Nal
UPPER
ANGLE GEARBOX
COUPLING
OUTPUT-TO
SINGLE - POINT
HOIST
DRIVE SHAFT
SECTION A-A
OUTPUT-TO
LOWER ANGLE
GEARBOX
*9»*J^*W*'W <*•
-COaiNG OIL
FITTING (TYP)
-HIGH PRESSURE
OIL FITTINGS (TY$
INPUT - FROM-
UPPER ANGLE
GEARBOX
OUTPUT-TO
LOWER ANGLE
GEARBOX
SCALE
' i> ii t ii 111'
-Mi) - —
LiLL 1 ION L C
Fliuri 73, FaurW’ilnt Holvt*
261
I T »
ANTI-BACKLASH COVER
CABLE LENGTH
POTENTIOMETER I
FEEDBACK —. ]-J
CONTROL
I SLIP RING -
ASSY
l^vLT SCRUB -
JBk R0 ^er
-M—Jl ^
\
rr
5_
i
x,
3
QETAILA
W/O FREE REEL CLUTCH
\ ^— DETAIL A
ROLLER BELLMOUTH
ASSY
, CABLE
DRUM
]>— FREE REEL
7 CLUTCH
FROM FOUR- g
POINT HOIST
MANUAL CON- ^
TROL VALVE
WESTON BRAKE
L - BALL SCREW
- REACTION PIPE
10 FREE REEL CLUTCH
DETAIL A
FREE REEL
CLUTCH
FROM FOUR
POINT HOIST
MANUAL CON¬
TROL VALVE
<3
<3
WESTON BRAKE
WIPER-SCRAPER
SEALS
xloJ_I_I
SCAlI
Figure 74. Cargo Hook, 12,OOO-Pound Capacity
HOIST CABLE
r
11,550-LB-CAPACITY
HOOK-SWIVEL ASSY
SIZE COMPARISON
40,000-LB-CAPACITY
HOOK-SWIVEL ASSY
INCHES
0 12 345
/-HOUSING
/-TAPERED PIN
/-UPPER PISTON
I OWER PISTON
r~ LOAD CELL
INCHES
APPENDIX IV
TYPICAL SEQUENCE OF OPERATIONS
INTRODUCTION
To provide a better understanding of the operation of the proposed single-
and four-point cargo handling systems, typical missions involving both
systems are described below. Two missions are described for the four-
point system. The second mission is included to show methods that can be
used to assure safe operation under adverse loading conditions.
No in-flight emergencies are described. Standard procedures for in-flight
emergencies for the single-point mission, in all cases, should be the
Jettisoning of the load by the hook release method. If a malfunction did
not permit the hook to release the load, the cable would then be sheared
by the tandem-dual cutters or would be free reeled off the drum. In a
multi-point mission no hook release would be attempted. Instead, the
cables would be sheared by using the tandem-dual cable cutters.
Prior to all missions the cargo handling system should be checked out by
the crew chief, with the APP providing the power, before the pilots enter
the aircraft.
SINGLE-POINT MISSION
Mission : Fly to pickup area for bulldozer which is to be transported
to clear area for observation post.
Terrain : Bulldozer in level field but 70-foot trees surround drop
area on mountain top.
Load : Prerigged in a sling with sling legs attached to a nylon
ring at apex.
Sequence :
1. Aircraft flown to pickup area, hover over bulldozer.
2. Reel out cable, ground crew slides nylon ring on
load beam of hook.
3. Load is lifted off the ground by the aircraft..
4. Pilot checks aircraft controllability. If satisfactory,
he signals aft pilot (hoist operator) to reel in on
single-point hoist to cable length required for best
flying qualities.
5. Aircraft is flown to drop site and hovers above trees.
6. Hoist is reeled out and hook is placed in auto touch¬
down mode.
7. When load is placed on the ground and the cable tension
drops to 150 pounds, the hook opens, releasing the load.
267
8. Hook control is placed in safe ar.d hoist cable is
reeled in.
9. Aircraft departs drop site when hook is reeled in
sufficient distance to assure clearance with tail rotor.
10. Aircraft returns to base.
Note: In the event of a malfunction of the automatic
touchdown release, the electrical release would
be used.
FOUR-POINT MISSION - STANDARD LOAD
Mission : Fly to pickup area for self-propelled mortar which is ^
to be transported to forward area.
Terrain : Vehicle located in level field, to be put down on
relatively rough terrain.
Load : Rigged for four-point pickup; no single-point sling
available.
Sequence :
1. Aircraft flown to pickuo area and landed near vehicle.
2. Vehicle driven under aircraft; hookup is made by
ground crew.
3. Hoists reeled in until load is a foot off the ground.
A. Cable load indicators are checked to ensure that load
falls within the C.G. limits of the aircraft.
5. Aircraft is lifted off into a hover; flight controls
checked out as satisfactory.
6. Aircraft flies to drop site, which is found to be too
uneven to permit landing.
7. A low altitude hover is established.
8. Hoists are reeled out until vehicle is several feet
below the wheels of the aircraft.
9. Hover altitude is slowly reduced until load is on
the ground and all four cables are slack.
10. Electrical hook release is actuated and all four hooks
open.
11. Aft pilot (hoist operator) confirms that all hooks have
released and aircraft hovering altitude iB slowly in¬
creased until it is confirmed that all hooks are free.
12. Hoists are reeled in until a safe length is reached.
13. Aircraft returns to base. \
Notes: (1) Step 10 requires the use of electrical release of the
hooks. If the auto touchdown release were provided
and were to be used under these conditions, the load
release could result in adverse loads being felt by
the aircraft*
268
These loads would result if one side of
the vehicle touched the ground first.
This would cause the hooks on this side
to open, and the resultant loss in load
on the aircraft would cause it to roll
about the hooks that had not released.
For this reason, the automatic touchdown
release is not provided for the four-
point hoist hooks.
(2) If one or more hooks fail to open, a ground
crewman must be available to climb up on
load and manually release the hooks. If
no crewman is available, or hook(s) cannot
be opened, the hoist cable(s) can be free
reeled off the drum(s) or sheared with the
tandem-dual cable cutter(s).
FOUR-POINT MISSION - NONSTANDARD LOAD
f
1
Mission ? Fly to pickup area for bulldozer which is to be transported
to forward area.
Terrain : Vehicle located on rough terrain, to be put down on a road
in forward area.
Load ; Rigged for four-point pickup, but pickup points not symmetri¬
cally located about C.G. of vehicle.
Sequence :
1. Aircraft flown to pickup area.
2. Rough terrain and unknown condition of pickup points,
or reasonable suspicion of same, results in a 15- to 20-
foot hover being established over vehicle.
3. Hoists reeled out until hooks are on the ground and the
cables are slack.
A. Hookup is made by ground crew.
5. Vehicle is slowly lifted off the ground by the aircraft;
hoists are not reeled in.
6. Vehicle swings forward, as vehicle C.G. is too far for¬
ward relative to the pickup points.
7. Pilot corrects for load swing with azimuth control
(cyclic control stick) but feels that too much forward
stick is required to permit forward flight.
8. Pilot requests aft pilot (hoist operator) to trim load
by reeling in on aft hoists.
9. Aft hoists are reeled in but vehicle assumes an extreme
noae-down attitude (or maximum cable load is reached
and hoists stall).
269
10 .
11 .
12 .
13.
14.
15.
16 .
Hoist operator informs pilot that he has run out of
trim control with the hoists.
Pilot rechecks cyclic control and decides that not
enough improvement has been made to warrant an attempt
at forward flight. , .. ,
Pilot informs hoist operator that it is a no go and
asks that load be leveled up.
Vehicle is leveled up by lowering aft hoist cables.
Hover altitude is slowly reduced until load is on
ground and all four cables are slack.
Electrical release is actuated and all four hooks open.
If tine permits, the pickup points on the vehicle are
repositioned and another attempt is made or a sling is
rigged to permit single-point lifting. A single-point
sling, with adjustable length legs, could be quickly
set to compensate for the nonsymmetrical C.b. ol tne
load so that it could be carried level.
«
<
270
DISTRIBUTION
US Army Materiel Command
US Army Aviation Materiel Command
Chief of R&D - DA
Director of Defense Research and Engineering
US Army R&D Group (Europe)
US Army Aviation Materiel Laboratories
US Army Human Engineering Laboratories
US Army Ballistic Research Laboratories
US Army Research Office - Durham
US Army Test and Evaluation Command
US Army Electronics Command
US Army Combat Developments Command, Fort Belvoir
US Army Combat Developments Command Transportation Agency
US Army Command and General Staff College
US Army Aviation School
US Army Armor and Engineer Board
Air Force Flight Test Center, Edwards AFB
US Army Field Office, AFSC, Andrews AFB
Air Force Materials Laboratory, Wright-Patterson AFB
Systems Engineering Group, Wright-Patterson AFB
Naval Air Systems Command, DN
Office of Naval Research
US Naval Research Laboratory
Marine Corps Liaison Officer, US Army Transportation School
Testing and Development Division, US Coast Guard
NASA Scientific and Technical Information Facility
NAFEC Library (FAA)
US Army Aviation Human Research Unit
US Army Board for Aviation Accident Research
US Naval Aviation Safety Center, Norfolk
Federal Aviation Agency, Washington, D.C.
US Army Medical R&D Command
US Government Printing Office
Defense Documentation Center
U nclassified _
Security Classification
This report presents the results of s two-phase feasibility and preliminary design
study of load suspension configurations capable of meeting the external cargo
handling system requirements of a 40,000-pound-payload heavy lift helicopter
In Phase 1, Design Analysis, both separate function configurations (those that
incorporate individual single- and multi-point hoists) and combined function configu¬
rations (multi-point hoists used to perform both single- and multi-point missions)
have been Investigated for both single- plus two-point and single- plus four-point
load suspensions. This phase was primarily concerned with investigation of hoist
types; methods of power transmission to the hoists; and selection of mechanical,
hydraulic, and electrical components. A comparative evaluation of 15 system arrange¬
ments was made on th? basis of weight, power, reliability, in-flight safety, versatil¬
ity, and productivity.
The single- plus four-point system was determined to meet the requirements best and
was recommended for the Phase IJ, Preliminary Design. This phase Included the
preparation of layout drawings, load and stress analysis of major components, a
maintainability and reliability analysis, and the preparation of a component develop¬
ment plan. The single- plus four-point system weighs 4974 pounds for a hoist capacity
of 40,000 pounds. The system has been designed such that the hoists of both systems
are readily removable for missiors ’•equiring minimum empty weight. For single-point
operation (four-point hoists removed), the system weighs 2758 pounds; for four-point
1 4.
KIV NONOB
Helicopter - Heavy Lift
Cargo Handling System - Helicopter
External Load Handling System - Helicopter
Hoist - Helicopter
none* or chakks in classification,
D IS T RI B U T I ON AND AVAILAHLUT
69-16 15 snrmBKR 1969
AD -828 ?di
Iio Foreign without
No limitation
United Aircraft Coi p.,
approval of Army
Stratford, Conn.
Aviation Materiel
Sikorsky Aircraft
Labs., Fort Eustis,
Div.
Final rept.
Kept. no. SER-50M1,
USAAVLABS-TR-67-^6
Nov 67
Va.
Contract DA-Mi- 177-
AMC-fc67(T)
*
USAAML notice,
1U Jul 69
Live Music Archive
Librivox Free Audio
Metropolitan Museum
Cleveland Museum of Art
Internet Arcade
Console Living Room
Books to Borrow
Open Library
TV News
Understanding 9/11