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I

731

Marine Engines and Boilers

Ubeit Besign anb Construction

A HANDBOOK FOR THE USE OF STUDENTS, ENGINEERS,

AND NA VAL CONSTRUCTORS

BASED ON THE WORK

BERECHIQUG URD KORSTRUKTIOR DER SCHIFFSIIiSCHIHEN UHD -KESSEL

By

DR. Gr BAUER

ENGINBER-IN-CHIEP OP THE VULCAN WORKS, STETTIN

TRANSLATED FROM THE SECOND GERMAN EDITION BY

E. M. DONKIN

t

AND

S. BRYAN DONKIN

ASSOCIATE MEMBER OF THE INSTITUTION OP CiViL ENGINEERS

EDITE;p BY

LESLIE S'/'rOBERTSON

»

SECRETARY TO THE ENGINEERING STANDARDS COMMITTEE
MEMBBK OF THE INSTITUT^ION OF CIVIL ENCINEEES, MEMBER OF THE INSTITUTION OF MECHANICAL FJs'CINEERS

MEMBER OF THE INSTITUTION OF NAVAL ARCHITECTS, ETC. KIXT.
AUTHOR OF "water TUBE BOILERS," AND EDITOR OF M. BERTIN's "MARINE BOILERS "

TDlitb over 550 ^lludtmtiond an^ flumeroud Sables

LONDON

CROSBY LOCKWOOD AND SON

7 STATIONERS' HALL COURT, LUDGATE HiLl

1905

\AU Rights Reserved]

Printed at Thr Dariex Prbss, Britta P/ace, Bdinhtrgk,

PREFACE TO THE ENGLISH EDITION.

When invited by the Publishers to undertake the editing of the
Engh'sh Edition of Dr Bauer's valuable work on " Marine
Engines and Boilers," it was with extreme hesitancy that I
contemplated undertaking the work on account of the many
heavy calls upon my time in another direction. I was, how-
ever, ultimately persuaded to do so on the Publishers securing
the collaboration of Messrs E. M. and Sidney Donkin, to whose
lot fell the duty of the original rough translation and the
conversion of the formulae into English measures. It would
have simplified our labours enormously had we been able to
adhere to the original metrical figures without giving their
English equivalents. Though most English engineers are con-
versant with the metrical system, they are, however, still
accustomed to think and make their mental comparisons in
pounds, feet, and inches ; and, as the book is primarily intended
for young engineers and draughtsmen, we felt that it would be
more useful, and many a mental exercise avoided, if English
units were used throughout. I trust, therefore, that the work
in its present form may be of value and assistance to the
younger generation of marine engineers and naval architects
of this country.

As a work of this nature involves an enormous number of
conversions, some errors, in spite of the precautions taken to
eliminate them, will, we fear, inevitably have crept in. We

137128

VI PREFACE TO ENGLISH EDITION.

should, therefore, deem it a favour if our readers would call our
attention to any cases coming under their notice.

Thanks are due to Mr E. H. Sprague, of University College,
London, for checking many of the calculations, to Engineer-
Lieutenant H. J. Loveridge, R.N., of the Admiralty, and to Mr
Charles Dresser for kindly running through the proofs.

LESLIE S. ROBERTSON.

28 Victoria Street,

Westminster, S.W.
February 1905.

DR BAUER'S PREFACE TO THE FIRST EDITION.

The present work owes its origin to an oft-felt want of a
condensed treatise, embodying the theoretical and practical rules
used in designing marine engines. It contains also drawings
of the more usual types of engines, and tables of the various
data required in ordinary practice.

The need for such a work has been felt by most engineers
engaged in the construction and working of marine engines,
by the younger men, as well as by those of greater experience.
Probably many of those connected with the building and driving
of stationary engines will also welcome the book, as an addition
to the literature in which they are more specially interested.
To them the chapters on Calculation of Cylinder Dimensions,
Turning Moment, Balance of the Moving Parts, Details of the
Engine, Piping, Pumps, &c., will be found more particularly
useful.

In accordance with the general character of the book, and
the necessity of limiting the range of subjects treated, only the
most modern types of engines — those of the vertical type used
in screw steamers — and of boilers in ordinary use, have been
dealt with. Special types have not been dealt with. The
greater part of the book is based upon results obtained in
actual practice ; but little has been taken from other writers on
the subject, and then only from specially trustworthy sources.

The compilation of the work and of the drawings has been
greatly assisted, or rather rendered possible, by the kind help
afforded by a number of the principal firms in Germany and
abroad, who have very liberally placed a large quantity of
material at the author's disposal. To all of these he desires to

VllI PREFACE TO FIRST GERMAN EDITION.

express his grateful thanks, but most of all to the Stettin er
Maschinenbau-A.-G. " Vulkan," who have permitted him to
make use of a large number of working drawings and photo-
graphs.

The work has been compiled as follows: — Mr Ludwig
undertook — Part I. Section 5, Sections 49 to 91, and Sections
141 to 153; Part II.; Part IV. Sections 2 to 5 ; Part V.
Sections 2 and 7; while Mr Boettcher undertook — Part I.
Section 2 ; Part V. Section 1 ; and Part VI. The whole of
the remainder was written and compiled by the undersigned,
with the constant help of Mr Foettinger.

In judging the book, it must be remembered that it is the
fruit of the few hours of available leisure snatched from the
midst of a strenuous professional life. Hence it will often be
found to lack the necessary finish, and to bear evidence of
various shortcomings. It is however hoped that as the book
has been written in response to an actual daily felt need, it
may, for that very reason, be of the greater service.

In conclusion, the Author would be glad if his readers would
bring to his notice any errors and omissions so that he may
be able to rectify them should, in the future, a new edition be
called for.

Dr G. BAUER,

Engineer to the Vulcan IVorkSy Stettin.
Stettin, March 1902.

DR BAUER'S PREFACE TO THE SECOND EDITION.

As the first edition was exhausted in little. more than a year —
much sooner than I had ventured to hope — I was unable, much
to my regret, to undertake more than a small part of the revisions
and alterations which appeared to me to be necessary in the
preparation of a second edition. Nevertheless I have found it
possible to make good various omissions, and to considerably
enlarge the chapters on "Arrangement of the Main Engines"
and " Water Tube Boilers," and also to add throughout the
book Tables giving data of engines which have been actually
constructed. I attach particular importance to the addition
of these Tables, because they are of far greater assistance in
determining the most suitable dimensions for new engines and
machinery, than any number of comparative values or empirical
formulae.

Additional value is given to these Tables because His Excel-
lency the Secretary of State for the Imperial Navy has allowed
me to use data compiled from various ships belonging to the
Imperial German Navy, and I desire to avail myself of this
opportunity of tendering my best thanks to His Excellency.

I desire also to express my indebtedness to the Stettiner
Maschinenbau-A.-G. "Vulkan," who have again placed very
valuable material at my disposal in order to complete the
illustrations of the book.

My thanks are also due to the other firms, in Germany and

*

elsewhere, who have supplied me with material for the second
edition of the book, as also to my publishers, who have contri-
buted greatly to the success of the work by their excellent
workmanship in the first edition.

X PREFACE TO SECOND GERMAN EDITION.

In spite of the great, not to say fantastic, expectations
entertained in some quarters in regard to the future of steam
turbines for ships, I have* not ventured at present to add a
chapter on the subject to this edition. Although the Parsons
turbine has had considerable success on land^ and has already
been fitted in a number of ships employed in running regular
services, reliable results from these latter, especially as regards
the consumption of coal, are not yet available.

In this second, edition, as in the first, Messrs Ludwig
and Foettinger have rendered me valuable assistance, while
the chapters contributed by Mr Boettcher have not required
alteration.

I would esteem it a favour if the reader will bring to my
notice any discrepancies he may discover, as was kindly done
upon the appearance of the first edition.

Dr G. BAUER,

Engineer-in-Chief of the Vulcan Works ^ Stettin,
Stettin, January 1904.

TABLE OF CONTENTS.

PART I.— THE MAIN ENGINES.

SECTION I.— DETERMINATION OF CYLINDER DIMENSIONS.

PAGE

§ 1. Horse-power --..... 3

S 2. Measurement of Indicated Horse-powcr ... 4

§ 3. Measurement of the Actual Work exerted by the Engine - 6

S 4. Indicator Diagrams and Steam Distribution - - - 8

S .">. Compound Expansion -...-- 9

§ 6. Work of the Steam in the Cylinder . . . . 9

S 7. Clearance ..-..._ 14

S 8. Calculation of Cylinder Dimensions for a given Horse-power - 16

§9. Determination of the Actual Mean Pressure (;>,„) for a New

Engine ....... 17

§ 10. The Theoretical Diagram and Efficiency - - - 17

J5 11. Combining the Diagrams ----- 18

S 12. Designing Engines, and Approximate Calculations - - 19

S 13. Number of Expansions, Cylinder Ratios, and Cut-off in each

Cylinder ....... 2I

S 14. Total Cut-off and Cylinder Dimensions - - - - 22
§ 1.'). Example of Method of Designing Triple-expansion Engine

for a Screw Mail Steamer ... - - 24

jS 16. Receivers ....--. 27

}i 17. Construction of a Theoretical Indicator Diagram from the

Diagram of Volumes - ..... 27

SECTION II.— THE UTILISATION OF STEAM IN THE

ENGINE.

S 18. The Fundamental Principle of the Mechanical Theory of Heat 32

§ 19. Losses by Throttling or Wire-drawing during Admission - 36

xu

CONTENTS.

PAGE

S 20. Direct Loss of Work due to Clearance - - - - 36

21. Indirect Loss of Work due to the Influence of the Cylinder

Walls ' - 37

§ 22. Object of the Steam Jacket ----- 37

§ 23. Influence of Multiple Expansion - - - - 37

§ 24. Heating the Receivers ------ 39

§ 25. The Condenser ------- 39

SECTION III.— STROKE OF PISTON, NUMBER OF REVOLU-
TIONS, TURNING MOMENT, BALANCING OF THE
MOVING PARTS.

S 26. Stroke, Number of Revolutions, and* Piston Speed - - 41

^5 27. Tables 9 to 16, Particulars of Vessels - - - - 42

S 28. The Crank .... ... 58

§ 29. Moving Parts of the Steam Engine - - - - 60

§ 30. Tangential Pressure on the Crank Pin, and Turning Moment

of the Multiple-crank Engine - - - - 63

§ 31. Variation in Crank-pin Velocity ----- 69

§ 32. Example - - - - - - - - 71

55 33. Explanation of Figures ------ 74

§ 34. Variations in Torsional Strains of the Shafting - - 74

Balance of the Moving Parts.

j5 35. General Remarks ------ 82

§ 36. Balancing the Moving Parts of a Single-crank Engine - - 83

§ 37. Balancing the Moving Parts of a Two-crank Engine - - 83

§38. Balancing the Moving Parts of a Three- crank Engine - - 84

§ 39. Balancing the Moving Parts of a Four-crank Engine : Schlick

System .-..-_- 80

§ 40. To Determine the Balance of the Moving Parts - - 88

S 41. Study of the Valve Gear- ----- 88

§ 42. Remarks -------- 89

g 43. Most favourable Arrangement in an Engine to secure Perfect

Balance of the Moving Parts ----- 90

§ 44. First Example: Calculation of the Balance of the Moving Parts

in an Existing Engine ----- 92

§ 4."). Second Example : Calculation of the Balance of the Moving

Parts in the Engine of a Fast Mail Steamer - - 100

§ 46. Effect of the Length of the Connecting Rod on the Balance of

the Moving Parts ------ 103

§ 47. Critical Number of Revolutions, and Effect of the Position of

the Engine on the Vibrations of the Hull - - 104

CONTENTS.

XlU

SECTION IV.— ARRANGEMENT OF MAIN ENGINES.

ji 48. Arrangement of the Cylinders and Cranks
§ 49. Longitudinal Bracing of the Cylinders - - - .

ji 50. General Remarks on the Arrangement of the Main Engines -
S 51. Starting the Engine -..-.-

i»ac;k
106

122

123

125

SECTION v.— DETAILS OF MAIN ENGINES— THE

CYLINDER.

§ 52. General Remarks - - - - - - 128

§ 53. Velocity of Steam - - - - - - 128

§ 54. Thickness (S) of the Cylinder Liner - - - - 130

S 55. Thickness (Sj) of the Cylinder Jacket or Outer Shell - - 131

S 56. Thickness (S^) of the Walls of Cylinders without Liners - 131

§ 57. Method of Fixing the Cylinder Liner - - - - 131

S 58. Cylinder Cover Studs ------ 133

§69. Cylinder Flanges -.--.. 133

§ 60. Cylinder Bottom ------- 134

§ 61. Cylinder Feet - - - - - - - 134

§ 62. Cylinder Passages and Valve Casings - - - - 134

'i 63. Calculation of Flat Surfaces in Valve-chests, Passages, and

Covers ------- 134

§ 64. Piston-valve Liner and Ports ----- 138

§ 65. Water Tests for Cylinders ----- 138

§ 66. Rules for Construction - - - - - - 139

§ 67. Cylinder Fittings 139

§ 68. Description of Figures 135 to 143 - - - - 144

§ 69. Cylinder Covers ------- 152

§ 70. Stuffing-boxes - - - - - - - 156

§ 71. Metallic Packings 158

Valves.

§ 72. General Remarks

§ 73. Thickness of Piston Valve Liners

§ 74. Ports of Valve Face

§ 75. Symbols used in connection with Slide Valves

§ 76. Stroke of Valve - - - -

S 77. Amount of Eccentricity -

§ 78, Principal Dimensions

161
165
165
165
166
166
166

XIV CONTENTS.

PAGE

§ 79. Cut-off -------- 166

§ 80. Linear Lead -..--.. 168

§ 81. Exhaust Lead ....... 168

§ 82. Compression -.-.-_- 169

S 83. Valve Diagrams - - - - - - 169

§ 84. Miiller-Reuleaux Diagrams for ordinary D Slide Valves • 169

§86. Zeuner*s Valve Diagram - - - - - - 171

S 86. Variations in the Cut-off- - - - - - 172

S 87. Stephenson's Link Motion - - - - - 174

Various Types of Valve Gear.

S 88. Klug Valve Gear - - - - - - - 176

§ 89. Marshall's Valve Gear --...- 179

§ 90. Joy's Valve Gear --..--- 179

§ 91. Heusinger Valve Gear - - - - - - 181

Piston Rods.

S 92. Maximum Load - - - - - - - 181

Pistons.

§ 93. General Remarks -.-..- 182

§ 94. Cast-steel Pistons - - - - - - 182

§ 94a. Cast-iron Pistons ... - - 183

§ 96. Piston Packing - - - - - - - 184

J5 96. Clearance between the Piston and the Top and Bottom Covers

of the Cylinder ...... 189

§ 97. Thickness of Junk Rings ----- 190

§ 98. Remarks .--.--.- 190

J5 99. Piston Rods -.-.--- 191

Connecting Rod and Crosshead.

§ 100. Length of the Connecting Rod .... 193

§ 101. Connecting-rod ------- 194

§ 102. Connecting-rod Fork ------ 195

§ 103. Crank-pin Brasses - ----- 196

S 104. Connecting-rod Bolts ------ 198

§ 105. Crosshead and Guide ------ 199

^ 106. Crossheads Forged in one with the Rod - - - 204

S 107. Pressure on the Guides . . . . . 205

i5 108. Guides -------- 205

CONTENTS. XV

Crank Shafts.

PAGE

5^ 109. Crank Shafts -....-. 208

i§ 110. Lloyd's Rules for Determining the Sizes of Crank Shafts of

Screw Steamers - - - - - - 210

J5 110a. Board of Trade Rules for Shafts - - - 212

S 111. Crank Pin ------- 214

S 112. Built-up Crank Shafts - - - - - - 214

1$ 113. Crank Shafts with Crank Pin and Web forged in one piece - 216

S 114. General Remarks on Crank Shafts - - - 217

S 115. Crank Shaft Couplings - ----- 219

.^ 1 16. Materials for Crank Shafts - - - - - 220

Valve Gear Rods.

i5 117. Power required to drive the Valves - - - - 220

J^ 118. Valve Rods 222

J5 119. Diameter of Rods ... - - . 226

§ 120. Valve Levers and Quadrants ----- 226

S 121. Stephenson's Link Motion ----- 227

J$ 122. Eccentric Rods . - - - - - - 228

Ji 123. Eccentrics and Eccentric Straps - . - . 229

S 124. Eccentric Straps ------ 232

)$ 125. Concluding Remarks .-.-.- 234

IJed-plates.

S 126. The Bed-plate ------- 234

)§ 127. Holding-down Bolts ------ 239

i^ 128. Longitudinal Bearers ...... 240

^ 129. Main Bearings .-.-.-- 240

§ 130. Main Bearing Caps ------ 243

.^ 131. Main Bearing Bolts -..--- 243

>^ 132. Dimensions of Main Bearings ..... 244

S 133, Thickness of Caps - - - - - - 245

Engine Columns.

S 134. Arrangement of the Columns ----- 246

§ 135. 1. Engines for Small Merchant Vessels - - - 246

S 135A. 2. Heavy-built Engines for Large Merchant Vessels - 246

1$ 136. 3. Engines for Modem Fast Steamers and Large Warships 247

§ 137. 4. Engines for Warships in General - - - - 248

XVI CONTENTS.

PAGK

i5 138. 5. Very Light Engines ----- 249

S 139. Stresses in the Columns and Framing - - - - 249

J$ 140. Fixing the Columns ------ 251

Reversing and Turning Gear.

S 141. Reversing Shaft and Lever ----- 251

S 142. Method of Handling the Reversing Gear - - - 256

S 143. Direct-acting Reversing Engines . - - - 256

S 144. All-round Reversing Gear ----- 261

S 145. Principal Dimensions of Reversing Engines - - - 264

S 146. Turning Gear ------- 266

5§ 147. Calculation of the Dimensions of the Wheels - - - 269

Condensers.

i. surface condensers.

J5 148. General Remarks .---.- 272

§ 149. Cooling Surface ...... 273

§ 150. Tubes and Tube Plates ----- 276

§ 151. Condenser Shell .-..-- 277

§ 152. Fittings and Connections ----- 280

2. JET CONDENSERS.

J5 163. Jet Condensers ------- 280

PART IL— PUMPS.

Air Pumps.

§ 154. General Remarks ------ 285

§ 156. Principal Dimensions ------ 285

§ 156. Air-pump Valves ------ 286

§ 157. Suction and Delivery Pipes ----- 286

§ 158. Pump Body ----- - - 289

§ 159. Pump Bucket ------- 292

55 160. Pump Rod 292

§ 161. Separately Driven Air Pumps ----- 292

Circulating Pumps.

§ 162. General Remarks . - - - - 294

CONTENTS. XVn

Reciprocating Circulating Pumps.

PAGE

55 163. General Remarks - - - - - 294

^ 164. Pump Valves 297

S 165. Suction and Delivery Pipes ----- 297

S 166. Pump Body 297

J$ 167. Plunger and Pump Rod ----- 297

Centrifugal Circulating Pumps.

Ji 168. General Remarks - - - - - - 298

^ 169. Suction and Delivery Pipes ----- 298

Ji 170. Pump Vanes - - - - - - - . 299

Ji 171- Centrifugal Pump Spindle ----- 303

S 172. Pump Casing ------- 306

H73. Engines for Driving Centrifugal Pumps - - - 306

>§ 174. Particulars of Surface Condensers - - - - 307

Fked Pumps.

«

S 175. Classification ------- 310

5i 176. Amount of Feed Water required . . - . 310

1. pumps driven direct from the main engine.

5i 177. General Arrangement ------ 310

S 178. Size of Feed Pumps 311

S 179. Barrels and Valve Boxes - - - - - 311

S 180. Pump Valves - - - - - - - 314

55181. Velocity of the Water - ----- 314

2. INDEPENDENT FEED PUMPS.

§ 182. Steam Pumps - - - - - - - 314

S 183. Duplex Pumps ------- 315

5$ 184. Simplex Pumps - ----- 316

§ 185. Weir Pumps - - - - - - -317

S 186. Blake Pumps - - ----- 318

Auxiliary Pumps.

55 187. Bilge Pumps Driven by the Main Engine - - - 321

§ 188. Sanitary Pumps ------ 323

^ 189. Method of arranging the Pumps . - 323

15 190. Separate Steam-driven Pumps - - - - - 323

xviil CONTENTS.

Pump Rods.

PAGE

§ 191. General Remarks ....-- 324

§ 192. Different Pump Arrangements taken from actual practice - 324
§ 193. Lloyd's Rules for Pumps and Pumping Arrangements - 329

PART III.— SHAFTING, RESISTANCE OF

SHIPS, PROPELLERS.

SECTION I.— SHAFTING.
Thrust Shaft and Thrust Block.

§ 194. Axial Thrust ------- 335

S 195. Thrust Shaft - - - . - - . 330

§ 196. Thrust Block ....... 336

§ 197. Thrust Blocks in Small Ships ----- 339

Tunnel Shafts and Plummer Blocks.

§ 198, Intermediate or Tunnel Shafts ----- 340

§ 199. Plummer Blocks or Bearings - - - . . 341

ij 200. Bulkhead Stuffing Boxes ----- 341

§ 201. Shaft Brake ------- 342

Shaft Couplings.

J5 202. Detachable Shaft Couplings ----- 343

§ 203. Muff Couplings ----.-. 344

§ 204. Disconnecting Couplings - . . . . 344

§ 205. Tail or Propeller Shaft ------ 345

Stern Tube.

§ 206. General Remarks ------ 346

§ 207. Construction of Stem Tubes for Cargo Boats - - - 347

§ 208. Method of Construction for Light Warships - - - 352

§ 209. General Remarks on Shafts ----- 352

CONTENTS.

XIX

SECTION II.— RESISTANCE OF SHIPS.

.^ 210. Froude's Method -...--

S 211. Calculation of the Resistance of Ships, and Power required
for the Engines in Screw Steamers - - - -

§ 212. Approximate Method for Determining the Horse-power of an
Engine ---._--

PAGE

353
355
360

SECTION III.— THE SCREW PROPELLER.

^213. Introduction -..--.- 361

S 214. General Remarks ------ 362

S215. Number of Blades -..-.. 367

§ 216. Different Forms of Blade ----- 367

S 217. Speed of the Screw, Stream-line Wake, and Slip - - 368

Ji 218. Propeller Efficiency ------ 370

S 219. Other Formula? - 374

S220. Remarks -...-.- 376

)$ 221. Taylor's Method for Calculating a Ship's Screw - - 376

§ 222. Taylor's Theoretical Formulie ----- 376
)i 223. Example of Taylor's Method of Calculating the Dimensions

and Shape of the Screw ----- 38I

Strength of Propeller Blades.

§ 224. Stress in the Propeller Blade due to Thrust and Tangential

Forces .-...-- 384

S225. Working Calculations .----- 387

>i 226. Stresses in the Blades due to Centrifugal Force - - 388

§227. Example I. — Effect of Centrifugal Force - - - 390

J5 228. Example II.— Effect of Centrifugal Force - - - 391

S 229. Thickness of Tip of Blade - - - - - 393

)^ 230. Material used for Blades ----- 394

Construction of the Screw.

1$ 231. Moulding and Casting the Screw
S 232. Explanation of the Drawings of Screws
§233. Propeller Boss -----
§ 2.34. Machining the Surface of the Blades -

395
396
401
406

XX CONTENTS.

PART IV.— PIPES AND CONNECTIONS.

SECTION I.— FLANGES, VALVES, ETC.

PACK

§ 236. General Remarks ...... 409

S 236. Pipe Connections ...... 409

S 237. Flanges .---..-- 409

55 238. Jointing - ... - - - 410

S 239. Bulkhead Fittings - - - - - - 411

§ 240. Extract from Regulations of German Lloyd^s respecting

Valves, Cocks, Pipe Connections, and Pumps - - 411

S 241. Valves - - - - . - - - 412

S 242. Sluice Valves - - - - - - - 417

S 243. Plug Cocks - - - - - - 417

SECTION II.— UNDER-WATER FITTINGS.

S 244. Under- water Fittings - - - - - - 418

S 245. Discharge Valves ...... 421

SECTION III.— MAIN STEAM, AUXILIARY STEAM, AND

EXHAUST PIPING.

1. Main Ste.\m Piping.

I? 246. Main Steam Piping ...... 423

S 247'. Draining of Steam Pipes - . . . . 423

S 248. Diameter of the Steam Piping ..... 424

S 249. Expansion due to Heat - - - - 425

§ 250. Arrangement of the Main Steam Pipes - - - 428

S 251. Thickness of Steam Pipes ----- 428

S 252. Lagging ...... 433

2. Auxiliary Steam Piping.

S 253. Auxiliary Steam Piping ..... 434

3. Exhaust Steam Piping.

S 254. Main Exhaust -----.- 435

§ 255. Auxiliary Engine Exhausts . . . - . 435

j5 255a. Diameter of the Exhaust Pipes .... 435

§ 255b. Thickness of Exhaust Pipes - - - . - 435

CONTENTS. XXI

SECTION IV.— FEEU-WATER PIPES.

PAGE

Ji io6. Boiler Feed Pipes ...... 436

j$ 267. General Arrangement --.... 436

^ 258. Feed-water Filter ...... 439

§ 259. Float Tank ....... 441

55 260. Diameter of Suction and Deliver>' Pipes - - - 441

§ 261. Thicknesses of Copper Delivery Pipes - - - 442

S 262. Feed Pipe Bends ...... 443

jS 263. Feed-water Heaters - - - - 445

SECTION v.— BILGE PIPES, BALLAST . PIPES,

CIRCULATING PIPES.

S 264. Bilge Pipes -..-.- 447

^ 265. Ballast Pipes ....... 448

j$ 266. Diameter and Thickness of Ballast Pipes - - - 449

)$ 267. Circulating Water Pipes ..... 450

PART v.— STEAM BOILERS.

SECTION I.— FIRING AND THE GENERATION OF STEAM.

S 268. "General Remarks ...-., 453

§ 269. Process of Combustion --.--. 453

5$ 270. Incomplete Combustion ..... 455

5^271. Losses by Excess of Air - - - - 456

S 272. Grate Area ....--- 457

^ 273, Natural Draught ..-.-. 458

§ 274. Artificial Draught ...... 460

§ 275. Centriftigal Fans ...... 4(50

>$ 276. Dimensions of the Fans ..... 46O

S 277. Example I. - - 461

J5 278. Example II. - - - - - - 461

5^ 279. Form of the Vanes ...... 462

.^ 280. Number of Blades ...... 466

5$ 281. High Temperature in the Combustion Chamber - - 466

§ 282. Mixing of the Gases of Combustion .... 467

XXU CONTENTS.

I'Ar.K

S 283. Useful Heat of Combustion . .... 467

S 284. Generation of Steam ------ 468

S 285. Efficient Transmission of Heat from the Gases of Combus-
tion to the Water ------ 468

S 286. Heat transmitted to the Contents of the Boiler - - 468

S 287. Formation of Steam in the Water - - - - 470

S 288. Efficiency of Steam Production - - - _ 471

S 289. Transference of Steam from Boiler to Engine - - - 472

S^ 290. Percentage of Water in Steam - . . . 473

SECTION II.— CYLINDRICAL BOILERS.

S 291. General Remarks ------ 474

j5 292. Selection of Heating Surface and Grate Area - - 474

S 293. Furnaces and Grates -.--.- 476

<^ 294. Boiler Tubes ----..- 480

S 295. Manholes ...---. 483

S 296. Thickness of Material Used . . - . - 483

S 297. German Lloyd's Rules ----- - 488

S 298. Hamburg Standard, 1898 ----- 493

J5 299. Extract from Rules of the " Bureau Veritas " - - - 498

§ 300. Extract from Lloyd's '' Regulations for British and Foreign

Shipping ..-.--- 502

SECTION 111.— LOCOMOTIVE BOILERS.
S 301. Dimensions of Locomotive Boilers - - - - 510

SECTION IV.— WATER-TUBE BOILERS.

J5 302. General Remarks - - - - - - 512

J^ 303. Belleville Boiler - - - - - - 513

S 304. Durr Boiler ..-.-.- 520

S 304a. Dimensions of a Diirr Boiler ----- 522

S 305. Yarrow Boiler ------- 526

S 306. Normand Boiler ------ 531

S 307. Small Water-Tube Boilers ----- 539

S 308. " Daring " Type Thomycroft Boiler - - - - 539

S 309. " Speedy " Type Thornycroft Boiler - - - - 539

S 310. Thornycroft Boiler ---... 540

§ 311. Recent Thornycroft Boilers ----- 543

CONTENTS. XXlll

SECTION v.— SMOKE BOX, FUNNEL, AND BOILER LAGGING.

PACK

S 312. Smoke Box 549

§313. Funnel 549

^ 314. Fixing of Funnel ...-.- 650

S 315. Funnel Dampers ...... 552

§ 316. Uptake and Funnel for a War Vessel . - - - 552

i^ 317. Boiler Lagging ------- 554

SECTION VI— FORCED DRAUGHT.

§ 318. General Remarks ------ 555

S 319. Induced Draught ...... 555

§ 320. Howdcn's System of Forced Draught - - - - 559

S 321. Closed Stokehold System _ ... - 564

SECTION VII.— BOILER FITTINGS AND MOUNTINGS.

S 322. Boiler Safety Valves ----- - 569

§ 323. Load on Valves ...... 570

§ 324. Safety Valve Casings - ----- 572

§ 325. Steam Stop Valve ...... 573

§ 326. Feed Check Valves ..-.-. 575

§ 327. Water Gauges ------- 575

§328. Pet-Cocks or Valves ...-.- 676

i^ 329. Density Cocks or Valves ----- 676

§330. Blow-off Cocks or Valves .. ... 576

§ 331. Scum Cock --..... 577

§ 332. Boiler-emptying Plug --.--- 577

§ 333. Apparatus for improving the Circulation of Water in the

Boiler 677

§ 334. Summary of Remarks ------ 577

§ 335. Regulations affecting Marine Boiler Fittings - . - - 577

PART VI.— MEASURING INSTRUMENTS.

§ 336. Pressure Gauges ------ 5^5

§ 337. Thermometers --.--.. 585
S5 338. Analysis of the Flue Gases ----- 588

XXIV CONTENTS.

PACK

S 339. Draught Gauge ...-.-- 589

j5 340. Determination of the Heating Value of the Coal - - 589

§ 341. Determination of the Amount of Moisture in the Steam :

Dryness Fraction ------ 590

§ 342. Indicators and their Use - - - - - 592
i^ 343. Study of the Indicator and its Accessories : Preparations for

Indicating --.-.-- 593

8 344. The Driving Gear ...-.- 595

§ 345. Putting on the Paper ------ .597

Ji 346. Planimeter ....... 597

§ 347. Schlick's Pallograph 597

J5 348. Instrument for Measuring the Uniformity in the Turning

Moment of an Engine . , - . . 599

§ 349. Fottinger's Torsion Indicator - - ... 599

PART VII.— VARIOUS DETAILS.

j§ 350. Bolts, Nuts, and Screw Threads, &c. . - - - 605

§ 351. Screw Spanners ------ 6(>9

§352. Platforms 613

S5 353. Edging Plates - - 613

S 354. Gratings - - - - - - 613

^ 355. Ladders - - - - - ■ - - 61 4

§ 356. Balusters and Handrails - - . . . (U4

S^ 357. Lifting-gear over the Engines - - - - - 614

Ji 358. Lifting-gear for Engines of Warships - - - - 615

S 369. Engine Foundations - - . - - 616

S 360. Construction of the Engine F'oundation - - - 616

S 361. Boiler Seatings ------ 618

S5 362. Lubrication of the Steam Spaces - - - - (j20

S 363. Lubrication of other Parts . - - - - 620

S 364. Ash Hoists ------- 622

S 365. Ash Ejectors ------- 622

J5 366. Ventilation of the Engine and Boiler Kooms - - - 624

g 366a. Area of Engine-room Ventilators . . - - 624

S 367. Ventilation of the Engine and Boiler Rooms - - - 626

§ 368. German Lloyd's Rules for Spare Gear for Engines and Boilers 62
§ 369. Lloyd's Rules for Spare Gear - , •

629

CONTENTS. JcXV

PART VIII.— VARIOUS TABLES.*

TABLE PAGE

I. Squares, Cubes, Square Roots, Cube Roots, Reciprocals,
Natural Logarithms, Circumferences, Areas of Circles,

from 1 to 1,000 ------ 634

11. Common Logarithms from 1 to 100 - 674

III. Sines and Cosines .--.-. 677

IV. Tangents and Cotangents . . . - - 679
V. Various Equivalents - - 681

VL Cos (tf + A, cos 2(1) --.-_. 681

VIL Inches and Millimetres - - 682

VIII. Square Metres and Square Feet .... 686

IX. Square Feet and Square Metres - ^ . . 687

XI. Pounds and Kilogrammes ----- 688

XII. Kilogrammes and Pounds ... - - 689

XIII. Pounds per Square Inch and Kilogrammes per Square

Centimetre ------ 690

XIV. Kilogrammes per Square Centimetre and Pounds per

Square Inch ------ 691

XXI. Comparison of Thermometers . - . - 692

XXII. Properties of Saturated Steam - - . . 694

XXIII. Expansion of Rigid Bodies by Heat - - 698

XXIV. Melting Points of Various Materials - - 698

XXVa. Specific Gravity of Woods 699

XXVb. Specific Gravity of Metals - - 699

XXVc. Specific Gravity of Various Materials - - 700

XXVd. Relative Weights of Coals 700

XXVe. Specific Gravity of Fluids 700

XXV F. Specific Weights of Gases at 30 Inches of Mercury and

32'Fahr. 701

XXVI. Strength and Elasticity of Various Materials - - 702

XXVII. Strength and Elasticity of Manganese Bronze - - 704

XXX. Moments of Inertia " I " and Internal Moments of Resistance
or Moduli of Section "w" for Circular Sections of

Diameter "rf" ..-.-. 705

XXXI. Bending Moments - ----- 707

XXXI I. Torsional Strength 711

* The missing Tables refer to data which it has not been thought necessary to
include in the English Edition.

As the German Plates have been used for Tables L to XXI. the decimal points are
indicated by commas.

C

XXVl

Contents.

TABLE tAGfe

XXXIII. Strength of Struts 712

XXXIV. German Lloyd's Rules for Iron and Steel for Boilers - 713

XXXV. German Lloyd's Rules for Steel and Cast Steel for Parts of

Engines ------- 714

XL. Weight of Machinery - - - - - 716

XLI. Weight of Boiler Equipments Compiled from the German

Navy ------- 716

XLII. Weight of Cylindrical Boilers - - - - 716

APPENDIX.

Final Report (June 1904) to the Lords Commissioners of
THE Admiralty of the Committee on Naval Boilers -

71'

INDEX

733

EXPLANATION OF SHADING USED IN BLOCKS.

I.

n.

in.

IV.

I. Cast Iron.
IL Steel or Wrought Iron.

III. Bronze.

IV. White Metal.

V. Material other than above.

LIST OF PLATES.

TO PACE
PLATE PAGE

I. Figs. 13, 16, 17, 18, 19, 20, 21. Combined Indicator Diagram
and Construction of Indicator iagrams from Diagram
of Volumes ------- 18

II. Figs. 77, 78. Compound Engine of Small Freight Steamer - 108

III. Fig. 88. Triple-expansion Engine of Imperial Yacht " Hohen-

zoUem" - - - - - - - 116

IV. Fig. 90. Triple - expansion Engines of Small Armoured

Cruiser .--.-.. Hg

after Plate III,

V. Fig. 96. Four-cylinder Triple-expansion Engines of Japanese

Armoured Cruiser " Yakumo " - - - - 120

VI. Figs. 97, 98. Four- cylinder Triple-expansion Engines of

" Kaiser Wilhelm der Grosse " - - - - 120

after Plate V.

VII. Fig. 100. Four - cylinder Triple - expansion Engines of

** Kaiserin Maria Theresa ^ - - - - 120

after Plate VI.

VIII. Fig. 101. Quadruple-expansion Engines of Twin-screw Mail

Steamer ------- 120

after Plate VII,

IX. Fig. 102. Quadruple - expansion Engines (Four Cranks, Six

Cylinders) of " Deutschland '^ - - - - 120

after Plate VI I L

X. Figs. 106, 107. Quadruple - expansion Engines of " Kaiser

Wilhelm II." 124

xxviii CONTENTS.

TO FACE
PLATE PAGE

XI. Figs. 132, 133, 134. Aspinall Governor - - • 142

XII. Fig. 142. Arrangement of Cylinders in Destroyer with Triple-
expansion Engines and Two l.p. Cylinders - - 160

XIII. Fig. 143. Arrangement of H.P. and l.p. Cylinders of

" Deutschland " .---.- 150

after Plate XII,

XIV. Figs. 308, 309. Duplex Pump - - - - - 316
XIVa. Fig. 461a. Yarrow Boilers for Chilian Battleship 530

XV. Fig. 479. Uptake and Funnel-seating of a Passenger Steamer 550
XVI. Figs. 493, 494. Combined Stop and Safety Valve for a Warship 572

PART I.

THE MAIN ENGINES.

SECTION I.
DETERMINATION OF CYLINDER DIMENSIONS.

§1. Horse-power. — The unit employed to measure the rate at
which work is done in a steam engine is the " horse-power," />., the
power exerted in the performance of 33,000 ft. lb. of work per minute.

A distinction is made between what is called indicated horse-poiver
(lh.p.) and the actual or brake horse-ptnver (b.h.p.).*

Let — T— s A be the area in square inches of a steam piston having a

diameter d ; let /„ be the mean pressure upon the piston in pounds per
square incht during one revolution; r= radius of the crank in feet;
2r=x, the stroke; n = number of revolutions per minute ; then the power
developed in a double-acting cylinder in indicated horse-power will be —

I.H.P. ^^^

and the mean speed of the piston will be —

r=2j« feet per minute.

Therefore ••«-r=^||^'

Hence by indicated horse-power is understood the work done per
minute by the steam on the piston of the engine. Part of this is lost
by friction in the working parts of the engine and the shafting. If
these frictional losses are deducted from the indicated horse-power, we
get the actual work available, or the brake horse-power^ which is the rate
at which useful work is done in turning the screw or propeller, or in
running the shaft with an artificial brake on.

The brake horse-power in very large engines is rather less, and in
small engines considerably less than the indicated horse-power.

If we take b.h.p. = ^x i.h.p., the following table (calculated from

* The term nominal horse-power is no longer used as a unit of measurement, and
therefore no mention is made of it here. If the power of a marine engine is given, for
brevity sake, as H.P., indicated horse-power is always meant,
t The unit of steam pressure is taken at 1 lb. per square inch.

4 MARINE ENGINES AND BOILERS.

Middendorf (Schiffstviderstand und Maschinenletsfung) gives the values
of i;, or what is known as the efficiency : —

Table No. 1.

I.M.P.

V = Efficiency.

1

i i.n.r.

1

f? = Effiiipncy.

•

0-68

1

I.H.P.

*/ = Efficiency.

Below 10

0-58

400- 500

1000-2000

0-79

10- 50

0-59

500- 600

0-69

2000-3000

0-85

50 100

0-60

600- 700

0-71

3000-4000

0-88

100-150

0-61

700- 800

0-72

4000-5000

0-89

150-200

0-62

800- 900

0-73

5000-6000

0-90

200-300

0-64

900-1000

0-74

6000 and

0-91

300-400

0-66

1

over

1
1

To measure the actual work done by means of a brake has hitherto
been very difficult in the case of marine engines, owing to the large power
to be absorbed, and the difficulties in fitting the brake ; until recently the
above approximate values were alone available. The latest trials made
according to the method described in Part VI. have given the following

values : —

In a 1630 i.h.p. engine 7; = 0-885
1640 ., 77 = 0-91

1940
2370
2690
4500

77 = 0-911
77 = 0-92
77 = 0-911
77 = 0-935

§ 2. Measurement of Indicated Horse-power— This is done

,by means of an instrument called an iW/-
cator* with the construction and mode of
working of which the reader is supposed to
jy A^ be familiar.

[^ >sv *^s the abscissae of the indicator diagram

are proportional to the volume passed through
by the piston, and the ordinates proportional
to the steam pressures, the area of the dia-
gram will give the work done during one
stroke.

Let F be the area of the indicator diagram
in square inches, m = the scale of the indicator
spring (so adjusted that if the scale =, J^yth, the indicator pencil moves

Atmos Line

/

A

Fig. 1.

* For the indicator,, see Part VI,

THE MAIN ENGINES. i)

one hundredth of an inch for a pressure of 1 lb. per square inch in
the steam cylinder), and / the length of the indicator diagram in inches,
then the mean pressure in the cyKnder during the stroke, in pounds per
square inch, will be —

Ai = 7--^'« (see Fig. 1).

Separate diagrams are taken for the top and the bottom of the
cylinder, and the mean of both mean pressures is used to calculate the
work done.

The atmospheric line is the line traced by the indicator pencil
when the indicator piston is not connected to the steam cylinder, but
is open to the air. Upon each diagram should be marked, whether
it was taken from the top or from the bottom of the cylinder, the
atmospheric line, the number of the engine, the date and time of the trial,
the number of revolutions, the number or allocation of the cylinder (if
the engine has more than one), the cut-off, the scale, and lastly the mean
pressure and the corresponding indicated horse-power (see Fig. 2).

&f^ine.Mi595

LPc^

't //^^

a.m

Mean Press ■ ISSlhs abs.

Seah;Sh in^^llb.
IHP =Z200

Fig. 2.

The simplest way of determining the mean pressure is by means of
a planimeter (see Part VI.). If, however, none is available, the best
way is to use the " Rule of Mean Ordinates," or, if greater accuracy is
required, what is known as " Simpson's " formula.

Rule of Mean Ordinates, — Divide the diagram into ten equal parts by
lines at right angles to the atmospheric line, and measure the distance in
the centre of each division between the top and bottom lines forming
the diagram. The mean height of the ten divisions, measured in inches

MARINE ENGINES AND BOILERS.

iqual to the n

and divided by the scale of the spring m
in pounds per square inch (see Fig. 3)—

Simpson's Formula (see Fig. A). — Divide the diagram into ten equal
parts as before, then if the dividing lines be lettered as shown, and we
take —

^0 + Ai„ = Hi,

Aj + Aj + ^j + Aj + Afl = Hj,

and Aj + ^4 + Ag + :4g = Hj,
the mean pressure (in pounds per square inch) will be —

It is important to note that the number of the spaces must be even.
Ta^U of Constants. — The work done in the cylinder in i.h.p. is—

33,000"

aa.ooo""

The constant c is a " characteristic " of each cylinder. If the product
« . C be calculated for any number of revolutions, and arranged in the
form of a table, the i.h.p. can be quickly obtained. Such tables are
frequently employed, especially on trial trips.

In calculating the area of cylinders, the sectional area of the piston
rod is sometimes deducted, but in large engines the work is usually
calculated without this deduction. If the rod is to be taken into
account, and it passes through both cylinder covers, the mean sec-
tional area above and below the piston is usually taken, as it is only
the mean power developed in the top and bottom of the cylinder
which is used as a basis of calculations.

g 3. Measurement of the actual Work exerted by the

Eagine. — In larger marine engines it has only been possible quite

THE MAIN ENGINES. 7

recently to do this. Use is made of the fact that all shafting twists, />.,
is distorted under the action of the turning moment exerted by the
engine. Two lines which were originally parallel in two different
sections of the shaft, in consequence of the torsional pressure, lie at an
angle to one another, forming what is known as the torsional angle. So
long as the stress upon the material of which the shaft is composed
is within the limit of elasticity of the material, this angle .is proportional
to the turning moment exerted. Let s be the arc of deflection of the
angle, measured at a distance k from the centre of the shaft, m the turning
moment transmitted to the shaft in inch pounds, l length of the shaft
in inches, and i the moment of inertia of the section ot the shaft in
inches**, g the modulus of elasticity of the material of the shaft in inch
tons, then —

s= = constant x m.

G. I

For the material generally used in shafting (Siemens-Martin steel of

about 28 tons per square inch tensile strength, and 20 % elongation),

(;, according to the latest experiments, is = 5,250 inch tons. As the

198,000 N , 1 J •

mean turning moment m= — , where N = work done m h.p.,

TT n

« = revolutions of the engine per minute, we get —

198,000 N __i:^R _ 198,000 n ^ ' *^-

TT ' n' 5,250 X 2,240 1 5,250 x 2,2407r ' ;/ * d^ , J-
' ^ ' ' 32

= mean arc of deflection during one revolution.

In turbines and electric motors^ the turning moment m, and hence
the arc of deflection j, is a constant, and equal to the corresponding
mean value.

In reciprocating engines the turning moment passes, at each revolution,
through a range of fluctuations ; as the arc of deflection j = constant x m,
it varies in proportion to the turning moment. If the arc of deflection be
measured at different positions of the crank, the corresponding actual
turning moment can easily be determined from it ; the cur\'e of the arcs of
deflection plotted above the developed circle of the crank represents, on
a different scale, the curve of actual turning moments, or tangential
pressures.

The arcs of deflection may be detertnined experimentally in either of
the following ways : —

1. In long shafts (from 60 to 100 feet in length), by utilising the

5 =

8 MARINE ENGINES AND BOILERS.

instantaneous action of the electric current (by Frahm's,* Professor
Denton's, and Fottinger's methods t).

2. By the very latest method in which a torsional indicator records
automatically the torsional deflections of the shaft, and draws a diagram
giving the actual tangential pressures.

For further details, see Measuring Apparatus, Part VI.,

also Shafting, Part III.,
and Torsional Vibrations, page 74.

Let s be the mean arc of deflection obtained from the diagrams,
and n the mean number of revolutions of the engine, then from the
foregoing equation we get —

Mean turning moment m = ^ . — ^— inch pounds,
and the actual work of the engine —

B.H.P. = ^-oTToTi = s .n. = s.ft , constant.

63,020 •0546 . L . R

To obtain the b.h.p. quickly it is advisable to draw up a table giving
the values of « x constant for the various speeds at which the engine is
likely to run.

§4. Indicator Diag^rams and Steam Distribution (see Fig.

5). — The steam enters the cylinder at a pressure somewhat less than

that of the boiler p, and fresh steam is
admitted while the piston travels along
part of its stroke a, which is called the
period of admission ; the steam then ex-
pands along the part of the stroke e,
which is called the period of expansion ;
before the end of the stroke, while the
piston is passing from e to ^, the exhaust
opens, and the steam is discharged while
the piston travels back along part of the
'' Y\rr, 5. " stroked — this is ih.e period of exhaust ;

the exhaust closes, and the steam re-
maining in the cylinder is compressed along the part of the stroke r,
which is the period of compression ; a little before the piston reaches
the dead point, at v^ fresh steam is admitted— a^/ww/V?«.J

For description of the manner in which this distribution of the

* Zeitsckrift des Verehus Deutscher Ingenieure^ 1902.

\ fahrbtuh des Schiffbautechnischen Geselischafiy 1903.

X In future the letters a, e, g, s, &c., in the diagram Fig. 5, will be used to denote
the corresponding cylinder volumes. Thus s signifies the volume of the entire stroke
in the cylinder, a volume at cut-off, &c.

THE MAIN ENGINES.

steam is obtained, and of the way in which it should be carried out,
see " Valves," Section 72.

§ 5. Compound Expansion. — Single-cylinder engines are those
in which the whole work of the steam is performed in one cylinder.

Twin-cylinder engines are those in which each cylinder works in
the same way as a single-cylinder engine ; the steam passes into both
cylinders direct from the boiler, and out of both direct to the condenser,
or to the atmosphere.

Compound engines are those in which the steam works successively
in several cylinders placed close to each other. It passes from the boiler
into the high-pressure (h.p.) cylinder, from thence —

(<?.) In a two-cylinder compound engine to the low-pressure (l.p.)
cylinder, and so on to the condenser, or to the atmosphere.

{b,) In triple expansion engines to the intermediate cylinder, thence to
the low-pressure cylinder, and then to the condenser, or atmosphere.

{c.) In quadruple expansion engines to an intermediate cylinder, then

to a second larger intermediate cylinder, and thence to the low-pressure

cylinder, and so on to the condenser.

r As the steam decreases in pressure in its passage through the various

* cylinders, and increases correspondingly in volume (see Table XXII.,

Part VIII.), the sizes of the cylinders, from the high-pressure cylinder

-onwards, must be larger according to the degree of expansion employed.

For constructive reasons the same degree of expansion may, in large
engines, be sometimes divided between two cylinders — either high or low
pressure — which are placed side by side, as in " twin-cylinder " engines.
Thus it often happens that in a triple expansion engine there are five
cylinders, namely, two h.p., one inter-
mediate, and two l.p. (see "Arrange-
ments of Main Engines," page
106). The cylinder dimensions are
calculated as if there were only one
cylinder for each degree of expan-
sion, equal in volume to the com-
bined volumes of the two cylinders.*

ConipKJund engines are calcu-
lated in precisely the same way as
single-cylinder engines ; the reason-
ing is the same as if all the work
of the steam were done in the low-pressure cylinder.

§ 6. Work of the Steam in the Cylinder.

(1.) Admission. — Assuming that there is no clearance, and that

Fig. 6.

• For the reason why compound expansion is employed, see page 37.

10 MARINE ENGINES AND BOILERS.

during the period of admission the pressure in the cylinder is the same
as that in the boiler p^ the work done by the steam during admission
(see Fig. 6) is —

Work of admission — a,p.

The mean pressure during this period, in relation to the whole
stroke, is thus —

The quotient - = c is called the cut-off, which is represented either as

s

a fraction, or as a percentage of the volume of the cylinder s. The

J 1 .
reciprocal value - = - is called the degree of expansion,

a €

In compound engines, the term total cut-off is understood to mean
the ratio that the volume of steam admitted to the high-pressure cylinder
bears to the volume of the low-pressure cylinder; by the term total
expansion is meant the reciprocal of this ratio.

If m be the ratio of the volume of the low-pressure cylinder to that
of the high-pressure cylinder, and €,, denote the cut-off in the h.p.
cylinder, we get —

Total cut-off

€

1_
€

_- ^

m
m .

Total expansion

•

m . s

fn,s_
a

1

For remarks on the selection of the proper proportions of the
cylinders, see page 21.

For information respecting the proper total degree of expansion,
see page 22 et seq,

(2.) Expansion, — After the valve has cut off the flow of fresh steam

to the cylinder, the steam in the cylinder begins
to expand, and drives the piston before it.

A study of actual diagrams shows that the
expansion line usually resembles very closely a
hyperbola,* and hence this curve is usually taken
as the basis for calculating the work done during
expansion, the more so as it is easily calculated.
^ — The equation (see Fig. 7) is —

/, . ^ = constant,
/>., the product of the volume and the pressure of the expanding steam
at each point of the stroke is a constant.!

* For the actual nature of the expansion curve, see page 33.

t Care must be taken not to confuse this purely theoretical curve of expansion with

THE MAIN ENGINES.

11

Construction of the rectangular hyperbola used as the curve of
expansion (Fig. 8). Let ab=/, the initial pressure; BC = tf, the volume
up to cut-off; AG = x, the

volume at the end of the B C If. £

stroke.

To get the final pres-
sure after expansion, draw
the diagonal line ae; then
through the point k, the
intersection of cj and ae,
draw the line kf parallel
to AG. The line fg gives
the required final pres-
sure. To get the pressure
at any given point of the
expansion curve, say for
the volume bd = ah, draw
the diagonal line ad, then
LM parallel to ag; the
point M where it intersects
the line dh is the required
point in the expansion curve,
the line mh giving the pres-
sure corresponding to the
volume AH.

To ascertain what volume
the quantity of steam a at the
pressure/ would occupy if it
were compressed to the pres-
sure /|, draw the diagonal line
AP (Fig. 9), then qo parallel to
NA. The line no gives the
volume required.

Work of Expansion. — This ^
is equal to the area cfgj, be-
low the expansion curve (Fig.
10).

We have cfgj = / p^dx.

Since the product of the

l*ig. 8.

Fig. 9.

the isotheimal expansion curve of a perfect gas, which is also a rectangular hyperbola.
Expansion in a steam-engine cylinder is not isothermal, as the temperature of the steam
£ills as it expands.

12

MARINE ENGINES AND BOILERS.

volumes and pressures is a constant for each point of the expansion
curve —

p^ . X —p . a, and /, =<-!—

\

Fig. 10.

Therefore cfgj = / ^-^—, dx^pa \q%^x =p , a loge -

•^ a JC L Ja a

The mean pressure during the work of expansion in relation to the
stroke s is therefore —

/e=/.-loge -
s a

(3.) Mean Theoretical Pressure of Admission and Expansion, — This
value is frequently used as a basis of calculation when determining the
dimensions of cylinders. It is taken from what is called a theoretical
diagram (see Fig. 10). The admission pressure is assumed to be the
boiler pressure in pounds per square inch absolute. The hyperbola is
used for the expansion curve, and perfect vacuum for the exhaust
pressure. The mean pressure of such a diagram is —

A=A+A=/ .€+/.€ loge - -/. € I 1 +l0ge-J

If the expansion takes place in a compound engine, c signifies the
total cut-off, - the total expansion.

f) 1

Table No. 3 gives the ratio — for different values of c and -

THE MAIN ENGINES.

13

Table No. 3.
Mean Theoretical Pressures of Steam,

2-25

2-3

2-4

2-5

2-6

2-7

2-75

2-8

2-9

3-0

31

3-2

3-25

3-3

3-4

3-5

3-6

3-7

3-75

3-8

3-9

40

41

0-752
0-714
0-667
0-625
0-588
0-571
0-556
0-526
0-500
0-476
0-455
0-444
0-435
0-417
0-400
0-385
0-370
0-364
0-357
0-345
0-333
0-323
0-313
0-308
0-303
0-294
0-286
0-278
0-270
0-267
0-263
0-256
0-250
0-244
0-238
0-235
0-233
0-227
0-222

A
P

1

€

0-9657

0-9546

0-937

0-9188

0-9003

0-8911

0-882

0-8641

0-8465

0-8294

0-8129

0-8048

0-7968

0-7814

0-7665

0-7521

0-7382

0-7315

0-7249

0-712

0-6995

0-6875

0-676

0-6703

0-6648

0-654

0-6436

0-6335

0-6238

0-6191

0-6145

0-6054

0-5965

0-588

0-5798

0-5757

0-5717

0-564

0-5564

— I.

4-6

4-7

4-75

4-8

4-9

5-0

51

5-2

5-25

5-3

5-4

5-5

5-6

5-7

5-75

5-8

5-9

6-0

61

6-2

6-25

6-3

6-4

6-5

6-6

6-7

6-75

6-8

6-9

7-0

7-1

7-2

7-25

7-3

7-4

7-5

7-6

7-7

7-75

0-217
0-213
0-211
0-208
0-204
0-200
0-196
0-192
0-190
0189
0-185
0-182
0-179
0-175
0-174
0-172
0-169
0-167
0-164
0161
0-160
0-159
0-156
0-154
0-152
0-149
0-148
0147
0-145
0-143
0141
0-139
0-138
0137
0-135
0-133
0-132
0-130
0-129

A

0-5491

0-542

0-5385

0-5351

0-5284

0-5219

0-5155

0-5093

0-5063

0-5033

0-4975

0-4917

0-4862

0-4808

0-4781

0-4755

0-4703

0-4652

0-4604

0-4555

0-4532

0-4509

0-4463

0-4418

0-4374

0-4331

0-431

0-4289

0-4248

0-4208

0-4169

0-4131

0-4111

0-4093

0-4056

0-4019

0-3984

0-3949

0-3932

1

e

A

6

/

7-8

0128

0-3915

7-9

0127

0-3882

8-0

0-125

0-3849

8-1

0-123

0-3817

8-2

0-122

0-3785

8-25

0-121

0-377

8-3

0 120

0-3755

8-4

0-119

0-3724

8-5

0-118

0-3694

8-6

0-116

0-3665

8-7

0115

0-3636

8-75

0-114

0-3622

8-8

0-114

0-3608

8-9

0-112

0-358

9-0

0111

0-3552

9-1

0-110

0-3526

9-2

0-109

0-3499

9-25

0-108

0-3486

9-3

0-108

0-3473

9-4

0-106

0-3447

9-5

0-105

0-3422

9-6

0-104

0-3396

9-7

0-103

0-3373

9-75

0-103

0-3361

9-8

0102

0-3349

9-9

0-101

0-3326

10-0

0-100

0-3302

10-1

0-099

0-3279

10-2

0-098

0-3257

10-25

0-097

0-3246

10-3

0-097

0-3224

10-4

0-096

0-3213

10-5

0-095

0-3191

10-6

0-094

0-3170

10-7

0093

0-315

10-75

0-093

0-314

10-8

0093

0-3129

10-9

0-092

0-3109

110

0091

0-3088

14

MARINE ENGINES AND BOILERS.

Table No. 3 — continued.

1

€

A

1

€

A

1

€

A

€

0-3065

€

/

c

14-5

/

IM

0-090

12-75

0-0784

0-2781

0-0690

0-2534

11-2

0-089

0-3049

12-8

00781

0-2773

14-6

00685

0-2521

11-25

0-089

0-304

12-9

0-0775

0-2757

14-7

00680

0-2509

11-3

0-0885

0-3030

13-0

0-0769

0-2741

14-8

0-0676

0-2496

11-4

0-0877

0-3011

13-1

0-0763

0-2726

14-9

0-0671

0-2484

11-5

0-0870

0-2994

13-2

0-0758

0-2712

15-0

00667

0-2472

11-6

0-0862

0-2974

13-25

0-0755

0-2705

15-1

0-0662

0-2460

11-7

0-0855

0-2956

13-3

0-0752

0-2697

15-2

0-0658

0-2448

11-75

0-0851

0-2947

13-4

0-0746

0-2683

15-3

00654

0-2436

11-8

0-0847

0-2939

13-5

0-0741

0-2668

15-4

0-0649

0-2425

11-9

0-0840

0-2921

13-6

00735

0-2654

15-5

0-0645

0-2413

12-0

0-0833

0-2904

13-7

0-0730

0-2640

15-6

0-0641

0-2402

12-1

0-0826

0-2887

13-75

0-0727

0-2633

15-7

0-0637

0-2391

12-2

0-0820

0-287

13-8

00725

0-2626

15-8

00633

0-238

12-25

0-0816

0-2861

13-9

00719

0-2613

15-9

00629

0-2369

12-3

0-0813

0-2853

14-0

0-0714

0-2599

16-0

00625

0-2358

12-4

0-0806

0-2836

141

00709

0-2586

16-25

0-0615

0-2331

12-5

00800

0-2821

14-2

0-0704

0-2573

16-50

0-0606

0-2305

12-6

0-0794

0-2804

14-3

0-0699

0-256

17-0

0-0588

0-2255

12-7

0-0787

0-2789

14-4

0-0694

0-2547

For instructions as to the use of this table, see page 19.
For fuller information on exhaust^ compression^ lead^ and release, see
" Valves."

§ 7. Clearance. — Before the incoming steam can force the piston
out, it has to fill the space between the piston and the valve face, known
as the "clearance" space. This space has an injurious effect on the
working and economy of the engine, because it is filled alternately with
hot admission steam and with exhaust steam of much lower tem-
perature, and therefore the large superficial area of ports between the
piston and cover causes considerable losses by condensation during
admission. For further disadvantages, see page 36.

It prejudicially affects expansion, because it raises the terminal pres-
sure, and also affects compression, because it reduces the final pressure
of compression.

In Fig. 1 1 the expansion curves are drawn with and without taking
the clearance o- into account ; if it be included the point a' forms the
starting point for the expansion line, because the steam in the clearance
is also expanded.

\

THE MAIN ENGINES.

15

The work of expansion is increased by the clearance —

JCF'g > JCFG.

(jCFG being work of expansion without the clearance.)

It can thus be proved that the compression, necessary to obtain a
given terminal pressure, must be increased, if there is a clearance space.

B' S

A

^

, a ,

y ^

y

f

F'

F
G

J

A

>4

/i

Fig. 11.

The clearance o- is measured in fractions, or as a percentage, of the
whole cylinder volume of which it forms a part.

The mean pressure^ including admission and expansion, and taking
account of the clearance, is, according to the reasoning on page 12 —

/.=/.l[«-H(«+.)iog.(£±5)]

How far the clearance affects the mean " theoretical " pressure, is
shown in the next table (No. 4). Clearance is always considerable in
marine engines. In cylinders with slide-valves it is —

For large cylinders, from 8 to 14 °/^.
For small cylinders, from 10 to 15 "Z^.

The higher values may be used for quick-running engines, with large
ports. In large low-pressure cylinders of mercantile steamers, fitted with
flat slide-valves, the clearance is generally from 8 to 10 7o of the cylinder
volume.

In cylinders with piston valves the clearance may be taken as —

12 to 18 "/^ for small cylinders with short straight ports.

15 to 19 7o for very large cylinders with long ports.

18 to 30 "l^ for small and medium sized cylinders (h.p. and inter-
mediate cylinders of war-vessels), with long and large ports. The higher
v:alues may also be used for quick-running engines.

16

MARINE ENGINES AND BOILERS.

Table No. 4.

Values of E^ {Mean Theoretical Pressure^ taking Clearance into
P Account),

(From Haeder, Die Dampfmaschinen,)

Admission

Percentage of Clearance Volume.

or Cut-off

€

0 7o

6 7o

!

8 7o

0-21

10 7o

12 7o

14 7o

16 7o

0-00

0-00

0-17

0-24

0-27

0-29

0-32

0-05

019

0-30

0-32

0-35

0-37

0-39

0-41

0-06

0-22

0-32

0-34

0-37

0-39

0-41

0-42 .

0-07

0-25

0-34

0-36

0-39

0-41

0-43

0-44

0-08

0-28

0-36

0-38

0-40

0-42

0-44

0-46

0-09

0-31

0-38

0-40

0-42

0-44

0-46

048

0-10

0-33

0-40

0-42

0-44

0-46

0-47

0-49

012

0-37

0-44

0-45

0-47

0-49

0-50

0-52

0-14

0-42

0-47

0-49

0-50

0-52

0-53

0-54

0-16

0-46

0-50

0-52

0-53

0-55

0-56

0-57

0-18

0-49

0-53

0-55

0-56

0-57

0-59

0-60

0-20

0-52

0-57

0-58

0-59

0-60

0-61

0-62

0-25

0-60

0-63

0-64

0-65

0-66

0-67

0-68

0-30

0-67

0-69

0-70

0-70

0-71

0-72

0-73

0-40

0-77

0-78

0-79

0-80

0-80

0-80

0-81

0-50

0-84

0-86

0-86

0-87

0-87

0-87

0-88

0-60

0-90

0-91

0-91

0-92

0-92

0-92

0-92

§ 8. Calculation of Cylinder Dimensions for a g:iven Horse-
power.— The power developed by an engine in indicated horse-power
is (see page 3) —

_p^ x2sxAxn
' 33.000

Where a =

'jrD'

If the stroke, number of revolutions, and cut-off be assumed, the
diameter of the cylinder can be calculated as soon as the actual mean
pressure ^^ has been determined. The calculation of the cylinder
dimensions thus depends wholly on the determination of /„,-

In compound engines the whole work of the steam is assumed to be
carried out in the l.p. cylinder. The mean pressure is determined for a

THE MAIN ENGINES. 17

corresponding single-cylinder engine, and is inserted in the formula for
the I.H.P., in which d denotes the diameter of the l.p. cylinder.

§ 9. Determination of the Actual Mean Pressure (/„,) for

a New Eng^e. — This may be done by determining the mean pressure
/t from the theoretical or empirical diagram, and reducing it to the value
/,„ by comparing it with results obtained from similar engines.

-.E

§ 10. The Theoretical Diagram and Efficiency (>^).— This

diagram shows how much work the steam admitted at each stroke from
the valve chest of the h.p. cylinder could do, if it were expanded from
it^ initial to its final volume. Here it is assumed —

1. That the whole clearance is filled at each stroke.

2. That the expansion curve is hyperbolic.

3. That the steam exhausts from the l.p. cylinder into a perfect
vacuum.

4. That there is no compression, no lead, and no premature opening
to exhaust.

5. That no losses by condensation, leakage, &c., take place.

The steam pressure at the point of cut-off in the h.p. cylinder is
always lower than in the boiler, but in this diagram during the whole
period of admission it should
be taken at full boiler pressure. r^ r ^' ^ C

The construction of the dia- fg'
gram is based on the assump-
tion that the admission is at
the pressure p. The diagram
is obtained by lengthening the
hyperbola c'f' along the line
to c (Fig. 12). The distance
Bc = ii represents the desired
reduced period of cut-off. The
volume a^ + cis the initial vol-
ume before expansion. The

quantity of steam ^/j +cr at the pressure p is the same as the quantity
a + ir at the initial pressure of admission in the h.p. cylinder. Fig. 12
shows the theoretical diagram of a single-cylinder engine, with the
actual diagram of the engine sketched-in inside. On the above
assumptions the area a'b'cf'g of the theoretical diagram represents
the energy contained in the quantity of steam o- 4- a^

llie theoretical diagram of a compound engine is determined in
the same way. The initial volume, as already stated, is equal to the

11

Fig. 12.

18 MARINE ENGINES AND BOILERS.

clearance in the h.p. cylinder + its volume up to point of cut-off; the
final volume is equal to the clearance in the l.p. cylinder + the total
volume of the l.p. cylinder.

Fig. 13 (Plate I.) shows the theoretical combined diagram of a
triple engine. The area a'b'cf'g is the total area of the diagram ; the
distance b'c is the volume of the steam at cut-off, reduced to the pressure
of the boiler (during the period of admission), + clearance in the h.p.
cylinder. The diagram of the actual engine is sketched in inside the
other. The figure obtained by drawing the actual indicator diagrams
inside the theoretical diagram is called a combined or Rankine diagram,

§11. Combining the Diagrams. — ^When the two are thus

shown together, the clearance belonging to each cylinder is set off from
a'b' \ and the expansion curves of the separate diagrams will then lie
correctly as they are affected by the clearance (see above). The lengths
of the diagrams are to each other as the volumes of their respective
cylinders. The ratio of the areas of the actual diagram Fj + Fg + Fg to
the area of the theoretical diagram a'b'cf'g is called the efficiency k.

If /„ denotes the sum of the mean pressures in the diflferent
cylinders, reduced to the l.p. cylinder, then —

Efficiency >&=!LtJVtZ?=A" • ^-^An

A B CF G A • "^ A

If, when designing an engine, the mean pressure /t be determined
with a planimeter from the theoretical diagram, it is only necessary to
multiply it by the efficiency ^, to get the requisite mean pressure /,„
of the actual engine. Approximate values for k may be obtained, as
already stated, by drawing the actual indicator diagrams inside the
theoretical diagram. When choosing an efficiency, that of a known
engine of the same type, and, as far as possible, of the same dimen-
sions, should be selected.

To construct a theoretical diagram for a new engine, it is necessary
to know the fall in pressure of the steam before expansion begins.
This depends on the size and length of the steam pipes, their position,
loss by radiation, the number of revolutions, arrangement of the passages
in the h.p. cylinder, and the type of valve gear.

The following table gives the data for determining this fall in
pressure : —

3

■

THE MAIN ENGINES.

19

Table No. 5.

Fcdl in Pressure of the Steam bettveen the Boiler and the beginning of

Expansion (between c and c').

Small single-cylinder and \ 12 to 17 lb. per square inch for low

twin-cylinder engines - ( boiler pressures and low speeds.

/=70 to 100 lb. per square X 17 to 22 lb. per square inch for high

inch - - - - ; boiler pressures and high speeds.

Compound engines - - \ 22 to 30 lb. per square inch for low

( boiler pressures and low speeds.

^=100 to 150 lb. per square ( 30 to 45 lb. per square inch for high

inch - - - - * boiler pressures and high speeds.

Large triple and quadruple
expansion engines -

/=at or above 160 lb. per
square inch

Large triple and quadruple
expansion engines -

/=at or above 220 lb. per
square inch

50 to 60 lb. per square inch.

52 to 65 lb. per square inch.

§ 12. In Designing Engines and for Approximate Cal-
culations the mean theoretical pressure /t is not taken from the com-
bined diagram, but simply from Table No. 3 on page 13. To get the
actual mean pressure, /, is multiplied by the efficiency k. Fig. 14 shows
how illogical this process is. The efficiency is, as we have said —

abcf'g

If the method of calculation
given above be used, it works
out as though it were —

ABDFG

The reason why this method
gives fairly satisfactory results
is that the shaded areas a'b'ba
and cdff' are, under ordinary
conditions, of practically the
same size.

Fig. 14.

The mean theoretical pressure will in future be denoted by (/»),

20 MARINE ENGINES AND BOILERS.

whether obtained by the more exact method, or only by the approxi-
mate one.*

For an example of method of designing a Triple Engine^ by means of
the theoretical diagram, see page 24.

Table No. 6.

Efficiency k^ for Condensing Engines,

The values for k given below hold for the usual number of total
expansions and cylinder ratios.

Expansion in a Single Cylinder.

I^rge slow-speed engines - - /t = 0*70 to 0*75

Small high-speed engines - - - >t = 0'65 „ 0-70

Expansion in a Two- Cylinder or Compound Engine,

Large engines up to about 100 revolutions per

minute - . . . . ^ = 0-60 „ 0*67

Small engines, with a higher number of revolu-
tions , - . - - /t = 0-55 „ 0*60

Triple Expansion in Three Cylinders.

War-vessels, with a high number of revolutions - >& = 0-53 „ 0*54
Mercantile ships, up to 100 revolutions per minute ^ = 0*56 „ 0-61 1

Triple Expansion in Four or Five Cylinders,

High-speed engines . - . . /t = 0-50 „ 0*52

Mercantile steamers, up to 100 revolutions per

minute _ - - . _ >t = 0-54r

Quadf uple Expansion in Four or more Cylinders,

Large mercantile steamers - - - >& = 0'52 „ 0-53

N,B, — The mean of the indicator diagrams, taken from the top and
bottom of the cylinders, has been used to draw up the above table.

For non-condensing engines the mean pressure, calculated from the
above data, must be reduced by about 15 lb. per square inch, on account
of the back pressure.

* The method here described for constructing the theoretical diagram is only
to be used for obtaining comparative results. It is inaccurate to assume that the
clearance of each cylinder is filled afresh at every stroke. This inaccuracy has
purposely been allowed to stand, to avoid further complications of the diagram, and
because much greater inaccuracies would inevitably arise in selecting the coefficient
/' for new calculations. The method has also been retained, because it is in general
use among marine engineers, and gives fairly good results.

t Rarely over 0-58.

THE MAIN ENGINES.

21

§ 13. Number of Expansions, Cylinder Ratios, and

Cut-off in each Cylinder. — Expansion can only be carried to
such a point that the final or terminal pressure in the l.p. cylinder
is equal to the exhaust pressure. This lower limit can easily he
determined from the theoretical diagram, but it is never realised in
actual practice. Expansion is limited in this direction by the question
of economy, which falls off if the expansion is carried too far. In
engines for warships and
small light craft, the maxi- ^ y. n. n
mum power is based upon ^

a comparatively small num-
ber of expansions, the rea-
son being that in the former
type of vessel the highest
power has only to be de-
veloped occasionally.

In § 14, page 22, the
normal number of expan-
sions for different types of
engines are given. The
theoretical diagram. Fig.
15, gives a starting point
for determining the ratio
of the cylinder dimensions
and the cut-off^ in the intermediate and l.p. cylinders.

If the area of the diagram abcfc; represents the actual work done in
a compound engine, the cylinder dimensions and cut-off corresponding
to the' work in each cylinder would be determined thus :—

2/3 = volume of the l.p. cylinder; and

— = total admission or cut-off being assumed : —

ABCFO is divided into three equal areas, f^ f., and Fg by drawing
parallel lines to ag.

Then v^ = volume of the h.p. cylinder.

— = cut-off or admission in the h.p. cylinder.

77 = volume of the intermediate cylinder.

Fig. 15.

^* = cut-off or admission in the intermediate cylinder.

22 MARINE ENGINES AND BOILERS.

iff

-2 = cut-ofror admission in the l.p. cylinder.

1 : -2 . _ 8 — ratio of cylinders.

The areas Fj, Fjj, Fg do not in any way represent the actual diagrams.

If account be taken of the clearance (page 14), the back pressure in
the condenser, compression, early admission, and other similar data,
the fall in pressure of the steam on entering the h.p. cylinder, and
also between the various cylinders, the diagrams Fj, Fg, Fj may be so
modffied as to resemble very closely the actual diagrams. From their
shape or form the dimensions of the cylinders may be determined.
The simplest w^ay is to take the cylinder ratio and the cut-off from
engines which have given satisfactory results.

In marine engines an endeavour is nearly always made to distribute
the power equally on each crank, but for constructive reasons it is not
always possible to do this, or to obtain a uniform turning moment
(see page 63) or reciprocation of the moving parts (page 82).*

The cut-off in the separate cylinders may be varied within such wide
limits, by adjusting the valve-gear while the engine is running, that the
division of the work between the cylinders may be considerably altered.
If, for instance, in a triple expansion engine the cut-off in the inter-
mediate cylinder be increased, the exhaust pressure in the h.p. cylinder
falls, and more work is thrown upon the h.p. cylinder. As the total work
in the three cylinders does not vary considerably so long as the cut-off in
the H.p. cylinder remains the same, the work done in the intermediate
cylinder will be less as cut-off is increased. In the same way, by
making the cut-off in the l.p. cylinder less, more work will be done in
that cylinder, and less in the intermediate, where the back pressure will
be increased.

An ordinary triple expansion three-cylinder engine may be cited as
an example where the work was —

With a cut-off of 70 7o in the h.p. cylinder 690 i.h.p.^

i.H.p. -2,170 i.H.p.

„ 71 „ M.p. „ 700

„ 55 „ L.P. „ 780 I.H.p. j

and with a cut-off of 70 **/„ in the h.p. cvlinder 690 i.h.p.

„ „ 71 „ M.p. „ 730 I.H.P. ^2,138 I.H.P.

60 „ L.P. „ 718

I.H.p. j

I.H.P. y2,

I. H.P.J

g 14. The following may be taken as the total cut-offand cylinder
dimensions for various types of engines : —

* In practice it is hardly ever possible to obtain the same fall or range of tempera-
ture in the different cylinders.

THE MAIN ENGINES. 2?>

1. Single-cylifider Engines. — These are hardly ever used now except
as twin-cylinder engines for small light boats, steam pinnaces, &c.

Cut-off € = 60 to 80 7^.

Auxiliary engines, such as steam cranes, reversing and turningengines,
circulating engines, &c., generally work with a very late cut-off, and
sometimes even with full admission. This is always the case if the
reversing engine is worked by rotary or reciprocating slide valves (see
Reversing Engines).

2. Compound Engines. — These are now only used for small freight
and passenger boats developing up to 300 h.p. The total rate of ex-
pansion depends upon the space available and the weight of the engine.

Light Engines,— p = 100 to 140 lb. per square inch.*

M- =^ = 3-2 to 3-8.

L.P.

Heavy Engines, — p-^0 to 100 lb. per square inch.

w = 4 to 4-6.
Cut-off in H.p. cylinder 50 to 70 7o» which corresponds to a total

expansion of about - = 5 to 8.

For the table giving data from actual compound engines see page 42.

3. Triple Expansion Engines, — These are made in many types and
sizes. For various arrangements of cylinders see page 106. The
cylinder dimensions and rate of expansion vary correspondingly within
a wide range.

(a.) Engines for Torpedo-boats and Torpedo-boat Destroyers,— p=^\^0
to 220 lb. per square inch, and occasionally up to 250 lb. per square
inch. Ratio of the cylinders about 1:21: 4*4 up to 1 : 2*2 : 5. Cut-
off in H.p. cylinder about 65 '/^ at maximum power. Total expansion
6*8 to 7*7 times the original volume.

(b.) Engines for Cruisers^ 6*^.— /=160 to 220 lb. per square inch

* Here and in future the pressure al)ove the atmosphere is given at the stop-
valve of the engine and called (/). In water-tube boilers steam is often generated at
from 2nO to 3fJ0 lb. per square inch, and the pressure is reduced before entering the
engine, by means of a reducing valve, to a pressure of from 180 to 225 lb. per
square inch.

24 MARINE ENGINES AND BOILERS.

(above atmosphere), sometimes up to 250 lb. per square inch. Ratio
of the cylinders about 1 : 2-3 : 5-5. Cut-off in h.p. cylinder about 70 7o
at maximum power. Total expansion about 7*5 to 8 times the original
volume.

(c.) Engines for Ironclads y Fast Steamers^ and Mail-boats, — -/=160
to 200 lb. per square inch. Ratio of cylinders about 1 : 2-4 : 6 up to
1:3:7. Cut-off in h.p. cylinder 70 to 75 **/^ at maximum power.
Total expansion about 8 to 10 times the original volume.

(d.) Engines for Large Slow Freight and Passenger Steamers, and
for Cargo-boats, — /= 150 to 220 lb. per square inch. Ratio of
cylinders about 1 : 2-6 :6'8 up to 1 : 3*2 : 7*2. Cut-off in h.p. cylinder
60 to 70 7o fitt maximum power. Total expansion about 9-5 to 12
times the original volume.

For table giving data from actual triple expansion engines see page
44 et seg.

4. Quadruple Expansion Engines. — These are only used where a
high rate of expansion can be obtained. On small light ships this is
not possible, as weight and space must be economised. Quadruple
expansion engines are therefore chiefly found in fast mail steamers and
large freight and passenger vessels. /=190 to 220 lb. per square inch
(above atmosphere). Ratio of cylinders 1:2:4:8 up to 1 : 2.2 :
4-4 : 9*2. Cut-off in h.p. cylinder 65 to 72 7o ^^ maximum power.
Total expansion about 10 to 13 times the original volume.

For table giving data from actual quadruple expansion engines see
page 56.

For arrangement of cylinders of such engines see page 120.

§ 15. Example of Method of Designing Triple Expansion
Engine for a Screw Mail St^^mtT.— Assuming : i.h.p. = 6,300;
/r = 75 revolutions per minute; ^=786 feet per minute; /= 184*8 lb.

per square inch (absolute); we have stroke J = ^7- = 5 fee 3 inches.

In

Under normal conditions the engine is supposed to work with a high

rate of expansion, and to be capable of developing still more power

by adjusting the valve gear of the h.p. cylinder. For the normal power

developed (6,300 i.h.p.) let a total expansion of - = 1M, € = 0*09, be
assumed.

1. Using in the first instance the shortened method referred to on
page 19.

THE MAIN ENGINES. 25

According to Table No. 3, € = 0-09, A = 0-3065; therefore A =

/
0-3065 X 184-8 = 56-6 lb. per square inch. ^

For an efficiency ^^ = 085 (see Table No. ik) the mean pressure
(calculating from the l.p. cylinder) will be /„» = 0-58 x 56*6 = 33 lb.
per square inch.

The diameter of the l.p. cylinder is worked out from the equation —

I.H.P. =/,„ X

^d2

33,000

, 6,300 X 33,000 , , « aaa

whence a = . ' — --^ — = about 8,000 sq. m.

.5.3 X 7oo

Therefore d = about 100 inches.

The diameter of the h.p. cylinder is thence obtained by assuming a
cut-off for that cylinder.

Assuming €^ = 0*6, we have —

Area of the l.p. cylinder _
Area of the h.p. cyhnder

and (from S 6) - = w . -1; therefore w = 1 1 -1 x 06 = 6*7.

8 000
Area of h.p. cylinder = -f-—- = about 1,190 sq. in.

6-7

Diameter of the h.p. cylinder = 39 inches.

The diameter of the intermediate cylinder is obtained by comparing
the engine here under consideration with similar engines which have
worked satisfactorily.

According to § H the proportional volumes of the cylinders will be
from about 1:3:7 to about 1 : 2-4 : 6, &c. The ratios chosen will
therefore be as 1 : 2*7 : 6*7. Thus the area of the intermediate cylinder
= 1,190x2-7 = 3,210 square inches, and the diameter of the inter-
mediate cylinder = 64 inches.

2. Calculation from tlu Theoretical Diagram (page 17).— (Compare
the theoretical diagram of an actual engine, a'b'cf'g, Fig. 13, where the
cut-off in the h.p. cylinder is 57 7o instead of 60 °/^.) Assuming the
work done, number of revolutions, and piston speed, to be the same
as before, the total expansion being taken as =11-1; cut-off in h.p.

HP

cylinder =60 7o> ^^i^ of cylinder-volumes -^—^ as 1 : 6-7 ; also clear-

ance in h.p. cylinder = 16 7o, in l.p. cylinder = 8 7o (see page 14); fall
in pressure between boiler and cut-off (cc') = 35 lb. per square inch (com-

26 MARINE ENGINES AND BOILERS.

m

pare Table No. 5). The cutoff will then be b^^^^^"^ ^^^^ ^^^'^

= 61-4 7^. Total volume after expansion a'g = (100 + 8)x 6*7 7, of
the volume of h.p. cylinder = 724 7o-

The mean pressure in the theoretical diagram can be obtained with
a planimeter, or by calculation from Table No. 3. Selecting the latter
method, w^e have —

Initial volume b'c 61*4 ^^q^
Final volume a'g 724

For this cut-off Table No. 3 gives -^^ = 0*295 ; therefore A = 0*295 x 184 S

= 545 lb. per square inch.

According to the other (shorter) method, /t was = 56*6 lb. per
square inch. By this second method we get rather larger diameters
for the cylinders. The process of calculation of these latter is the
same as for the previous method.

3. In the corresponding engine as actually built the total i. h. p. = 6, 320 ;
« = 75; ^=786 feet per minute; j = 5 feet 3 inches; h.p. cylinder dia-
meter =39*4 inches; intermediate cylinder diameter = 64 inches; l.p.
cylinder diameter = 102 inches.

Ratio of the cylinders, 1 : 2-66 : 6*7. Cut-off in h.p., 57 7o > i" inter-
mediate, 55 7o ; in L-P- cylinder, 50 *'/^. Work done, in h.p., 1,970 i.h.p.;
in intermediate, 2,050 i.h.p.; in l.p., 2,300 i.h.p.

Mean pressure in the theoretical diagram a'b'cf'g = 52*5 lb. per
square inch=/i. Mean pressure reduced to area of the l.p. cylinder,
in H.p. = 9*95 lb. per square inch; in intermediate = 10*4 lb. per
square inch; in l.p. cylinder =11*65 lb. per square inch. Therefore
/„ = 9*95 + 10*4 + 1 1 -65 = 32 lb. per square inch.

Therefore the efficiency will be —

>fe = An ^ !jl+ ^2 +.^8 ^ _^ ^ 0*609.
p^ ABCFG 52*5

4. Had the engine hsidfive cylinders^ with the intermediate cylinder
working on the centre crank, and an l.p. and an h.p. cylinder working
on each of the two outside cranks, then the calculation (compare the first
method) would be as follows : —

Diameter of each h.p. cylinder = 38*5 ^ = 27*2 inches.
Diameter of each l.p. cylinder = 100 ^/i = 70*7 inches.

If the efficiency had been taken at 0*54 instead of 0*58 when
using five cylinders, the cylinders would be of somewhat larger dia-
meter than given above.

THE MAIN ENGINES. 27

§ 16. Receivers. — Marine engines have no special receivers ; firstly,
in order to diminish the weight; and secondly, because the exhaust
passages, connecting pipes, and valve chests provide sufficient receiver
volume. The following are the usual values given to receivers : — Be-
tween H.p. and M.p. = 1-8 to 3*8 x volume of h.p. cylinder; between
M.p. and L.P. = 1-3 to 2'3 x volume of m.p. cylinder. The effect of the
receiver upK)n the steam distribution can be ascertained by constructing
a diagram of volumes (see § 17).

§17. Construction of a Tlieoretical Indicator Diagram
from the Diagram of Volumes. — If for a given engine the cylinder
dimensions, clearance and receiver volumes, steam distribution, and
positions of the crank are known, the volume occupied by the steam in
its passage through the engine may be graphically determined at any
point of the stroke.

Assuming (1) that the steam is expanded and compressed according
to the law Pressure x Volume = constant ; ('2) that, in combining a
volume zfj at a pressure ^i with another volume 7'., at a pressure/.,, the
final pressure obtained is —

I

then the pressure corresponding to the volume at any given point in the
stroke may be calculated, and a theoretical indicator diagram drawn.

This kind of diagram is not suitable for calculating the work done
by an engine, because the losses by condensation are not taken into
account, and the results obtained are too large. But it gives particulars
of the steam distribution, fall in pressure, and any peculiarities that may
be looked for in the actual indicator diagrams, and it is therefore as
well to draw it, when getting out the calculations for a new engine.

Example. — Construction of a "combined indicator diflgr(H«-r^=-er
"diagram of volumes" for a triple expansion engine. See Figs. 16 to
21, Plate I.

The following data are assumed : —

28

MARINE ENGINES AND BOILERS.

Data.

1

H.P.

Cyl.

Inter-
mediate
Cylinder

= i.p.

L.P.

Cyl.

First
Receiver

1

Second
Receiver

Cylinder diameters
Ratio of volumes -
Volumes as plotted in

Fig. 16 (inches)
Clearance -
Clearance (as plotted),

in inches -
Cut-off

Cut-off, in inches -
Exhaust lead
Exhaust lead, in inches
Compression
Compression, in inches

38-58"
1

1-97

•315

118

20 7o
•394

5 7,

•098**

64"

2-7
5-32

11 7o

•585
56 7
2-97'

i«7o

•955

12 7o

•638

100"
6-7

13-2

8 7

/o

1-055

•54 7o
713

16 7o
211

12 7.
1-58

2 5
4-92

4^05
7 95

Draw a diagram, the abscissae of which give the volume swept through

by the piston, and the ordinates the angles of the crank, calculated

from the upper dead point of the h.p. cylinder. If the volume of the

H.P. cylinder be taken as 1-97 inches, then the clearance in this cylinder

1-97 X 16
will be ' — = '315 inch. The volume of the intermediate cylinder

== 2'7 X 1^97 = 5*32 inches, and so on.* In order that the diagram may
not be unnecessarily long, receiver i is shortened by 3*94 inches, re-
ceiver II by 5*91 inches, and the lower part of the l.p. cylinder volume
is drawn in on the left of where it ought to be.

It is assumed that the connecting rods are infinitely long, and an
ordinary sine curve is used (compare page 59) to represent the path of
the piston. The cranks are set at an angle of 120° to each other.
In Fig. 21 (Plate I.) the volume of the stroke ef of the h.p. cylinder is
indicated for a given angle of the crank (75"), The distribution of the
steam at different points of the stroke being drawn, the volume of
steam at any given point can be measured from the diagram.

* The diagram, Fig. 16, was originally drawn to the scale volume of u.v.
cylinder = 1 '07 inches -50 mm., but has here l)een reproduced at one-fourth of the
actual size.

THE MAIN ENGINES. 29

Determination of Steam Pressures,

Assumed, — Initial pressure in h.p. cylinder = 185 lb. per square inch.
Hence, from the drawing, Plate I., the final pressure after expansion in
H.p. cylinder will be= 146*5 lb. per square inch (see Fig. 17).

Assumed, — Exhaust pressure from l.p. cylinder to condenser = 2*82
lb. per square inch. Hence the final pressure of compression in l.p.
cylinder =7-11 lb. per square inch (see Fig. 19). The other pressures
are deduced from the following equations for the pressures at different
periods of the stroke.

I. Combining the pressure and volume of the h.p.t (146*5 lb. per
square inch) with the volume h.p.b + Rj + i.p.t,* having the unknown
pressure p^

Combined pressure —

II. Expansion in h.p.i cylinder + h.p.r + Rj + i.p.t until point of com-
pression in H.p.u.

206*1
Final pressure /g = -—— x/., = 0*945/.,.

Jlo

III. Expansion in h.p.t + Rj + i.p.t until point of cut-off in i.p.t-

Final pressure p^ = ;^^ ^ x/g = 0*733/.,.

IV. Compression in h.p.t + Ri till admission begins in i.p.r.

177-7
Final pressure /g = 0*733/2 x ill-' = 0*895/2.

The pressures in h.p.t + Rj must next be combined with the clearance
of LP. p. The slight fall in pressure is neglected ; therefore the pressure
in H.P.T + Rj -I- i.P.B is /g.

V. Compression in h.p.t + Ri + i.p.b until exhaust begins in h.p.b.
This must be at the same pressure as given above, where H.p.,j-f-R^-»-
i.p.T was combined with h.p.t; therefore —

Final pressure /i= ^r^7>- xA==^*^^A•

* As in all cylinders the period of exhaust is greater than the period of compres-
sion, the steam sometimes passes from one side of the piston to the other — from the
U{>per (Top) to the under (Bottom). Suffix T denotes top and suffix B denotes bottom.

i> »»

30 MARINE ENGINES AND BOILERS.

From Equations I. and V. we get —

/j = 103 lb. per square inch.
A = 113-5

Hence ^3= 107 lb. per square inch.

A = 83
/»5= 101-5

The H.p. diagram (Fig. 17) can now be drawn.

The pressures for the intermediate and l.p. cylinders are obtained in
the same way.

In the i.p.T the pressure at the end of Period III. is/4 = 83 lb. per
square inch.

Final pressure after expansion in this cylinder /g = 59'5 lb. per
square inch.

VI. Exhaust from i.p.t to R2+i.p.b + l.p.t at the unknown pres-
sure /y.

Combined pressure />8= ^^^^^^-o^^ "^^^ =1*33 + 0-683/,.

VII. Expansion in i.p.t + i.p.b + Ro + l.p.t until compression in i.p.b.

Final pressure /g =^3 -^^ = 0-984/s.

VIII. Expansion in i.p.-, + R2 + L-P«t till the cut-off ends in l.p.t.

371*9
Final pressure /iq=/9 ^ =0*669/g.

ti47

IX. Compression in i.p,t + R.2 until l.p.b opens.

Final pressure /ii=/io 4^-2 + 202-5 =^*^^^A-

X. Combining i.p.t + R2 with the clearance of l. p. „ where the final
pressure of compression is 7*11 lb. per square inch.

n u' A ^ 250-7;),, + 26-8x7-11 n-jiivj^ ^i\.r<
Combmed pressure /^^^ \^^.' =0-818/g + ()-ob.

XI. Compression in i.p.t + R., + l.p.„ until exhaust begins in i.p.„.
The same final pressure must be obtained as in equation VI., where
I.P.T was combined with i.p.„ + Ro + l.p.t.

Therefore the final pressure /. = |I(^^ . /^o = 0-836/« + 0-049.

THE MAIN ENGINES. 31

From Equations VI. and XI. we get—

/7 = 38*6 lb. per square inch.

whence Py--i4:-b lb. per square inch.

/,« = 30

/>!.. = 37-6

The intermediate and l.p. cylinder diagrams (Figs. 18 and 19) can
now be drawn.

The diagrams thus obtained show the characteristics of actual dia-
grams, but their mean pressures are naturally much higher than they
would be in actual practice.

SECTION II.

THE UTILISATION OF THE STEAM IN THE ENGINE.

j;!^. The Fundamental Principle of the Mechanical Theory
of Heat is that " Heat and Work are equivalent to one

another." It forms the basis for determining the efficiency of the
utilisation of steam in the cylinder (1 Britifih Thermal Unit = 772 ft. lb.,
and 1 calorie = 424 mkg.). In an engine working with a given pressure
of admission /, and of exhaust p.^ the total work obtained per unit
weight of steam is given by the difference between the toul heat of
admission and that of exhaust.

If the influence of the walls be neglected, it follows that, the higher
the temperature of admission, and the lower the temperature of exhaust,
the more work will be available per pound of steam. \Vith saturated
steam the amount of heat contained in the steam is proportional to its
pressure ; therefore to obtain the best results our aim should be to work
with high admission and low exhaust pressures. In practice the present
limit of boiler pressure is about 300 lb. per square inch, and for the
exhaust or condenser pressure 07 1 lb. per square inch absolute. As

all external radiation of heat is a loss, the cylinders and receivers, and all
connecting pipes and passages, should be carefully lagged. To add heat

THE MAIN ENGINES. 33

during expansion (that is, to heat the steam after it has entered the
cylinder) is not an economical utilisation of heat, because to make the
best use of it, it should be added at the highest pressure, ue, during
admission. (For the reason why steam jackets are nevertheless useful,
see below, § 21.) Hence in a theoretically perfect engine the steam
should expand adiabatically, — that is, without heat being either added or
subtracted during expansion.

Taking 1 lb. as the unit of weight, and

1^1 = volume of admission in cubic feet (taken from the table of

saturated steam, Part VIII.) ;
v^ = final volume of expansion (calculated from the equation for adia-

batic expansion, see below) ;
p^ and /g *^^ corresponding pressures of steam (absolute) in pounds

per square inch \

we get the following relations (see Fig. 22) : —

Work during admission ^x= p^x ft. lb.

The equation for adiabatic expansion, according to the mechanical
theory of heat, is /.«;* = constant, in which >^= 1-135 for dry steam (see
Zeuner's Technische Themwdynamik^ 1890, vol. ii., p. 75).*

z, = ., (A)*-

Hence , . .

Work during expansion Lg = // • dv

= -^— y(/iz;i-/3z/2)ft. lb.

Work during exhaust A3 —p.p^ ft. lb.
Total work done by 1 lb, of steam in passing through the engine —

= ;^(A^i-/27'2) ft- lb-
Table No. 7 is calculated from this equation, and gives the theoretical
work in foot pounds for various pressures of admission and exhaust. It
furnishes a ready means of ascertaining to what extent an engine is
utilising the steam supplied to it.

* For low steam pressures this curve varies somewhat from the adiabatic expansion
curve.

C

;u

MARINE ENGINES AND BOILERS.

Table No. 7.

Theoretical Work done during the adiahatic expansion of 1 lb, of Steam
adiabatisally between pj and p.,, including the work done during
admission and exhaust,

k

w =

k-\

■ (/i^i ~p2^2) ^^- ^^- P^^ pound of steam.

[p in pounds per square inch, v in cubic feet per pound.]

^=1-135 for dry steam.

Vi=--

4-39c. ft. perlb.

1

3-Olc. ft. perlb.

' 2-30c. ft. perlb.

1

1 -85 c. ft. per lb.

A-

/i = 1001b. per
square inch.

/i = 150 lb. per
square inch.

/i=-2001b. per
square inch.

j/i = 2o01b, per
square inch.

lb. per sq. in.

ft. lb.

ft. lb.

' ft. lb.

ft. lb.

1

1

3
6

213,100
169,600
141,700

233,300
192,500
163,700

247,900

, 207,700

179,200

257,300
217,800
190,000

9

12
15

124,700
111,400
100,500

146,400
133,500
123,500

162,250
149,400
139,000

173,400
160,700
150,650

20
25

87,180
75,680

109,600
98,700

125,900
115,000

137,200
126,300

30

66,600

89,360

105,800

117.600

35

58,120

81,850

98,450

110,000

40
45

52,070
45,410

74,710
68,900

91,300
85,120

103,200
97,120

50

60

1

38,750
29,060

62,730
53,040

79,560
70,360

91,680
82,100

70
80

19,980
12,110

45,530
36,930

61,510
54,010

73,860
66,600

90

6,055

30,030

47,710

60,190

100

0

1

24,100

1

41,410

54,250

THE MAIN ENGINES. 35

Example. — An engine works with an admission pressure of 150 lb.
per square inch absolute, and an exhaust pressure of 3 lb. per square
inch absolute. The consumption of feed- water per i.h.p. hour
amounts to 13-5 lb. It is required to determine the ratio between the
indicated work and the work theoretically available.

1 I.H.P. hour = 33,000 x 60= 1,980,000 ft. lb.

w according to Table No. 7 is about 192,500 ft, lb. per pound of
steam.

Since 13'5 lb. per hour are required, this corresponds to 2,598,750
ft. lb. per hour.

Consequently the work yielded per pound of steam expressed as a
percentage of the work theoretically available

2,598,750 '^^

With the help of this table it is also possible to ascertain how the
steam is utilised in the different cylinders. To determine the work l in
each cylinder, the highest pressure shown in the diagram for p^ and the
lowest for/g are taken.

In marine engines the comparison is generally expressed as a per-
centage or efficiency (see § 10), and this is a practical way, provided the
engines to be compared are of the same type and size. But if the com-
parison is extended to engines of different sizes and types, working
under different conditions (with wet or dry steam, with or without a
condenser, with or without steam jackets, at high or low speeds), this
method of comparison by means of the efficiency must be used with
great caution, otherwise considerable errors may creep in.

For instance, from Table No. 6 the efficiency of triple-expansion
engines is less than that of single-cylinder engines ; nevertheless steam
is more efficiently utilised in triple-expansion engines than even in com-
pound engines. Hence we see that the efficiency does not afford a
reliable standard for comparing the economical utilisation of steam in
an engine.

The maximum work obtainable from the steam, calculated according
to the equation on page 33, is reduced in practice by losses of various
kinds. The reasons for these losses will now be briefly discussed.

§ 19. Losses by Throttling or Wire Drawing during Ad-
mission.— The cause of this is that marine engines work with a late
cut-off, and the valve gear does not permit of a sufficiently large opening
for the quantity of steam required to be admitted, nor does it cut off
with sufficient rapidity (see Fig. 23). Another reason is that the engine

•JO MARINE ENGINES AND BOILERS.

may have too small a h.p. valve chest, so that the steam expands into
the cylinder without being replaced with sufficient rapidity by fresh
boiler steam, or the steam
pipes may be too smalL
The loss of work due to
this cause is shown by the
area of the shaded triangle,
a 6 i:, in the
{Fig. 23).

§ 20. Direct Loss of
Work due to Clear-
ance.—At each dead
point the piston leaves a
space, from the piston to
the slide valve, which must
be filled with fresh steam
at each stroke. (For Clearance, see page H.) Thus a part of the
work of the unit weight of steam {which forms the basis of the diagram),
represented by the area a k i k, Fig. 23, is lost, a h representing the
volume of the clearance.

This unavoidable loss of work must not be confused with the loss due
to the influence of the cylinder walls (see § 21). To reduce its injurious
effect, a portion of the steam is compressed into the clearance, in order
that the latter may not have to be filled with entirely fresh steam at each
stroke. This does not result in as great an advantage as is usually sup-
posed; for, as shown by Fig. 24, the loss due to incomplete expansion
is greater than before, and the gain is merely the difference between

the areas smut and q' rp o'. Under certain conditions it may even
be a negative advantage, and compression would then be a disadvantage
contrasted with the filling of the clearance space with fresh steam.

THE MAIN ENGINES. 37

Since the curves w«, op^ qr^ all have the same exponent k^ and the
distances s m and o q are the same, the two areas smnt and oqrp are
equal to each other, and therefore the work per unit weight of steam is
less with a clearance and compression than without a clearance, because
expansion is not complete. Hence to utilise the steam effectively it is
usual to make the clearances as small as possible.

§21. Indirect Loss of Work due to the Influence of

the Cylinder Walls. — The question here is not one of losses due
to radiation, for the cylinders are always carefully covered with non-
conducting materials, in the same way as the pipes. We are dealing
with phenomena which take place on the inner side of the cylinder
walls, and their action cannot be observed externally.

The temperature of saturated steam varies with its pressure. During
one revolution the temperature of the steam will pass through the whole
range lying between the temperature corresponding to the limits of
pressure; for instance, in a single-cylinder condensing engine it will
pass from 338' F. to 140** F., corresponding to 115 lb. per square inch
admission and 3 lb. per square inch exhaust pressure. The walls of the
cylinder and clearance spaces follow these variations of temperature, as
far as they can. This is only possible by the incoming steam giving
up some of its heat to the walls of the cylinder and clearance spaces,
and this heat is given back by the walls later in the stroke, when the
steam is at a lower temperature. In other words, heat is withdrawn
from the steam at high pressure, and restored to it at a lower pressure.
This occasions a loss of heat, as compared with an engine in which
heat is not withdrawn by the walls, when the pressure is highest.

§ 22. Object of the Steam Jacket.— It is only the innermost
layers of the cylinder walls which participate in all the fluctuations of
temperature taking place in the cylinder. The variations of temperature
in the outer layers will be less. Each concentric layer has a mean tem-
perature which, diminishing towards the exterior surface of the walls,
approximates to the surrounding temperature. The higher the latter,
the less far will the variations of temperature extend outwards through
the walls, and the smaller will be the exchanges of heat during one revolu-
tion. This explains in a few words the real value of the steam jacket.

§ "IZ, Influence of Multiple Expansion.— Expanding the

steam m several cylinders has the advantage of diminishing the losses
due to ati^ clearance mentioned in § 20. A comparison may be made
by means o( the diagram. Fig. 25.

The area a ^r^ represents the diagram of a single-cylinder engine.

38 MARINE ENGINES AND BOILERS.

Compression is assumed to be up to the initial pressure ; the volume of
steam admitted per stroke is ai. Let the cylinder volumes/^, Ai, ki,
represent the equivalent triple expansion engine, the steam being com-

pressed in each cylinder to the initial pressure, and ihe admission volume
a'l>' = ab, the admission volume for cylinder ii will then be mn = l\K
exhaust volume of cylinder i. In the same way the admission volume
for cylinder in is c^ = the exhaust volume of cylinder ii. As all the
curves have the same equation, it is clear that the advantage of mul-
tiple expansion in this case is equal to the difference between the areas
crsd&nA ult.

Dividing the expansion between several cylinders has the further
advantage of reducing the so-called " initial " condensation, by reducing
the temperature range in each cylinder, vw represents the temperature
curve for saturated steam. For a single-cylinder engine the fluctuations
would, in the example chosen, be equal to t. By dividing up the ex-
pansion they are much reduced for each cylinder, Tj, t^ t,. It must
be noted that the temperature range in cylinder I is much less than that,
for example, in cylinder in.

To divide the total range of temperature into three equal pans
would be a mi.stake, not only because it would give unsuitable areas in
the diagrams, but also because steam is more susceptible to cooling at
high pressures than at low. In muhiple expansion engines the cooling
surfaces of the receivers have to be added to those of the cylinders,
and the former, as a result of the variations of temperature (not of radia-

THE MAIN ENGINES.

39

lion, which is assumed to be wholly prevented), affect the process of
cooling in a similar way to the cylinder walls (see § 21).

§ 24. Heating the Receivers. — The transferences of heat from the
steam to the walls, and vice versA^ mentioned in § 21, may be consider-
ably reduced, if the heat capacity of the walls be diminished. This is
best done by heating them. As regards the exchanges of heat, we must
note that this heating is not intended to heat the ^team itself, but to
produce by external action a storing up of heat in the walls, the object
being to limit the exchanges of heat between the steam and the
walls to as thin a stratum as possible, reckoning from within outwards,
/.^., to reduce the quantities of heat participating in the exchange.

In heating the receivers, it must be especially noted that to do so
by means of a system of pipes heated with boiler steam is wrong in
principle. Either the heat is transmitted from the heating steam to the
working steam through the metal (and this, according to § 18, is not so
effective as if the heating steam were actually used in the cylinder), or if
this direct transference of heat does not take place, the metal walls of
the pipes would only increase the surfaces affected
by the fluctuations in temperature of the steam, and ^{TTN. ^t^
thus augment instead of diminishing the quantities
of heat transferred.

>$ 25. The Condenser.— According to § 18,

the lower the exhaust pressure, the better is the utili-
sation of the heat supplied to the engine. A con-
siderable improvement can be effected by arranging
a special apparatus into which the engine exhausts
(see Fig. 26), as compared with an engine exhausting
into the open air.

The exhaust pipe a is led into a vessel a which
by an outlet pipe b is so connected to an open over-
flow tank c so that the total fall h is more than 30 feet.
The vessel a is kept at a constant low temperature
by a cooling coil ; the exhaust steam is rapidly con-
densed, and falls to a pressure corresponding to the
temperature in a. According to Table XXII.,
Part VIII., a temperature of 144° F. in the condenser
will give a pressure of 1 lb. per square inch absolute.
The .steam thus condensed to water runs off through the outlet ^,
which, through the pressure of the external atmosphere, is always full
of water up to the level h. The apparatus in this shape is unsuited to
marine purposes, and must be modified to meet the conditions existing

Fig. 26.

40 MARINE ENGINES AND BOILERS.

in vessels. This is effected by substituting for the waste pipe a suction
pump, which lifts a quantity of water corresponding to the steam con-
densed. So far the presence of air has not been considered. If the
pump worked too slowly and pumped too little water, the vessel a
would fill slowly, the action of the cooling coil would be affected, and
the vacuum would become defective. If the pump sucked too much
water the condenser would be pumped dry, and a mixture of steam
and water would be pumped out. In the first case the apparatus would
be practically useless, in the second the pump would have to be forced.
The ideal conditions would be reached when exactly as much water
was pumped as steam condensed. The gain in work by the use of
such an apparatus, as compared with an ordinary non-condensing engine,
would be equal to the difference between that part of the area of the
diagram of steam lying below the atmospheric line, and the pump dia-
gram. As the upper and lower limits of pressure for the areas of both
diagrams are the same, the quantities of work will be directly propor-
tional to the volumes. The ratio of these volumes for equal weights
of steam and water is about 1,700 : 1 ; in other words, the loss of work
due to the pump, as compared to the gain in work by the use of a
condenser, is a negligible quantity.

In actual engines air always finds its way from the feed water, and
through leakages in the stuffing boxes, flanges, &c., into the condenser,
and reduces the vacuum. Therefore the pump must be of such ample
dimensions that it can draw off the air as well as the condensed steam,
hence it is usually termed an air pump. Theoretical formulae for the
dimensions of the air pump cylinder cannot be given ; in designing new
plants, experience alone can serve as a guide. In practical work all risk
of leakages must be particularly avoided, and attended to at once if they
occur. Particular attention should be paid to the arrangement of the
so-called snifting-valves (see Pumps). These should only be used in
cases of necessity to deaden the shock of the water (if the pump is
working very rapidly). Wherever they have to be in constant use there
is something defective in the engine.

If the cooling surfaces, quantity of condensing water, &c., are care-
fully determined from results of actual experience (see Part II., Air
Pumps), and there is no leakage, a vacuum of 90 to 95 % of the perfect
vacuum may be obtained.

SECTION III.

STROKE OF PISTON— NUMBER OF REVOLUTIONS—
TURNING MOMENT— BALANCING OF THE MOV-
ING PARTS.

% 26. Stroke, Number of Revolutions, and Piston Speed.—

By piston speed we understand the mean speed of the piston during
one stroke. This is —

r= -7»7x- or = -— - feet per second.
For ordinary values of n, s, and c, see Table No. 8.

Table No. 8.
N'umber of Revolutions^ Stroke^ and Piston Speed,

Type of Engine.

Torpedo-boats and torpedo-
boat destroyers

Pinnaces, ships' boats

Small tugs

Small passenger steamers -

Large tugs and steam
trawlers

Unannoured cruisers

Armoured cruisers

Ironclads - - - -

Fast steamers -

Large cargo and passenger
steamers

Small cargo boats

Large cargo boats

• I

//

• 300 to 400

250

fv

380

180

)f

250

150

}}

200

1

100

^^

160

, 120

^^

180

100

,t

150

100

n

120

75

>>

95

70 „ 90
95 „ 130
70 „ 85

inches.
15 to 20
6 ,. 8

8
11

12
24
36
36
60

50
26
35

12
20

28
36
42
60
72

60
36
54

feet per sec.

16 to 19-5

5 .. 8

5
6-5

6-5
11-5
13
13
13

11-5

10

11-5

8
10

11-5

16

16

14-7

15-7

14-7
12-5
13

For further details of stroke and number of revolutions, see following
tables : —

42

MARINE ENGINES AND BOILERS.

i^ ^7. Tables of Particulars of Vessels.

Table
Compound

Name of
Ship.

Tyijc.

01)scrvations.

Ems*

City of
Chester*

I.H.P.

The small Mul-

' tiple represents

the number of

Engines in

the Ship.

Fast
steamer

Nether-
lands

Mail
steamer

Ice-breaker

Passenger
steamer

Small

passenger

steamer

Small

passenger

steamer

Steam
trawler

Pinnace

— Ice-breaker

Built by Elder of Glasgow for
the North German Lloyd.
Speed, 1 6 '4 knots.

Inman Line. 1885. 18^ knots.
(Busley, Schiffsmaschine.)

Built 1895.

Hoboken Ferry. Engines by

Fletcher. 12*4 knots.
Engineerings i., 1894, p. 224.

For river traffic.

1 X 6000

1 X 4600

For river traffic.

Built at Vulcan Works, Stettin,
1895.

For an armoured cruiser.

1 X 1600

Ix 740

Ix 180

Ix 125

Ix 200

Ix 50

Ix 350

* The engines of these ships are only inteicsting historically, as such large conv
pound engines are no longer made.

THE MAIN ENGINES.

43

No. 9.
Engines.

Xo.of
Revs.

Boiler

Pressure

above

Atm.

lb. ^r
sq. in.

95

78

Diameter of Cylinders.

The small Multiple

represents the numoer of

Cylinders.

H.P.

ft.

5

in.
If

5 8

95

114

119 100

150 100

175

100

130

114

380

120

142

106

2 5

1 6

1 2

0 11-8

1 3

0 6^

1 6i

L.P.

ft. in.
2x

7 ^

10 0

5 1-4

3 2

2 1

1 8-8

Stroke.

ft. in.
5 0

6 5J

3 2

2 4

1 4^

1 2

2 2i

1 11

0 Hi

3 2

0 7|

1 7J

Ratio of
Cylinder
Volumes.

1 : 4

Total Expansion
for a given Cut-
off" in H.P.
Cylinder.

Cut-off".
8 for 50 7,

1 : 31

10 „ 31 7^

1 : 4-4

1 : 4-5

1 : 3-4

1 : 81

1 : 3-2

1 : 3-5

1 : 4-5

7-3 „ 60 7.

7-4 „ 60 7„

57 „ 60 7,

4-4 „ 70 7,

8 « 40 7.

5 „ 70 7.

7-5 „ 60 7.

44

MARINE ENGINES AND BOILERS.

Table No. 10.

Triple-Expansion Engii

■

1

I.H.P.

Name of
Ship.

Type.

Oljservalions.

The small Mul-
tiple represents 1
the number of
Engines in ^
the Ship.

2x240

Poseidon

Scouting
steamer

Built 1901 by the Bremer
Vulcan Vegesak for the
German Home Office.

Flat-
bottomed
gunboat

Built by Thornycroft & Co.

2x260

Santa Fe

Destroyer

Built by Yarrow for the

Argentine Navy.
Engineerings ii., 1896, p. 122.

2 X 2000

Sword-Fish

Destroyer

Built by Armstrong for the
British Navy. Engines
by Belliss.

Engineeringy'u,, 1896, p. 122.

2 X 2200

1

Hart

Handy

Hunter

Destroyer

Built by Fairfield Co. for

British Navy, 1896.
Engineerings i., 1896, p. 245.

2 X 2250

1

Janus

Lightning

Porcupine

Destroyer

Built by Palmer's Shipbuild-
ing Co. for British Navy.
EngineeringSx,, 1896, p. 142.

2x1900

Satellit

Torpedo
cruiser

Built by Schichau, Elbing,

for Austrian Navy.
EngineeringXy 1893, p. 846.

2 X 2300

1

1

Torpedo-
boat

Built by Schneider, Creusot,

for French Navy.
Engineerings ii., 1 898, p. 257.

1 X 1500

THE MAIN ENGINES.

45

Torpedo^oats and Destroyers.

Boiler

Pressure

above

Atm.

lUper
sq. in.

114

Diameter of Cylinders.

The small Multiple represents the
number of Cylinders.

H.P.

ft. in.
1 2

135

150

192

1 0

1 6

1 ^

210

1 6i

210

180

1 6

1

9^

213

411

M.P.

. I

ft. in.

2 2

I., p.

ft. in.
2 2

1 8

3 %\

2 34

2x
2 4

2 3^ ' 3 6

2 3J

2 lOJ

3 6

4 U

2 0

2 lOi

Stroke.

ft.

in.

1

8

0

11

1

G

1

6

1

6

1

6

1

711

1

Ratio of
Cylinder
Volumes.

1 : 3-36

1:2-7

1:2-1 :4-8

1:21 :4-5

1 : 2-27 : 5-8

1 : 2-34 : 5-4

1 : 2-55 : 5-24

1 : 2-06 : 4-19

! Total Expansions
with Cut-off 707,
inH.P. Cylinder.

Speed
of the
Ship.

Knots.

with

—

60 7o
5-6

with

.^__

60 7o
4-5

6-8

26J

«*4

27

7-6

27

7-8

27

7-4

m

5-9

46

MARINE ENGINES AND BOILERS.

Table No. 11.

Triph'Exp

"mansion Engin

I.H.P.

Name of
Ship.

Type.

Observalions.,

The small Mul-
tiple represents ^
the number of
Engines in
the Ship.

S.42

Torpedo-
boat

Built by Germania, Tegel,
for Imperial German Navy.
Schifibau, i., No. 18, p. 553.

1 X 1440

Torpedo-
boat

U.S. Navy.

Engineerings ii., 1898, p. 819.

2 X 1500

Shirakumo
and

Destroyer

Built by Thornycroft & Co.
for the Japanese Govern-

2 X 3600

Asashio

ment.

Ardent
Boxer

Destroyer

Built by Thornycroft & Co.
for the British Navy.

2 X 2200

Bruiser

—

Torpedo-
boat

Built by Thornycroft & Co.

3000

M

Small
cruiser

Building for the Imperial
German Navy, 1903.

2 X 5000

Bogatyr

Armoured

Built at Vulcan Works,

2 X 9750

cruiser

Stettin, 1901-2, for the
Russian Navy.

Eber

Gunboat

Built at Vulcan Works,
Stettin, 1903, for the
German Navy.

2x650

THE MAIN ENGINES.

Torpedo-boats, Destroyers, and Cruisers.

Diameter of Cylindeis.

be tmall Mulliple rcivcunlA ll

DDmbcr oTCrliiultrs.

2 4|

2 H

2 3 2x

^2 7

2 5 2x
2 6

4 5-4' 6 9-4

5 0

2x

Ratio <>r
Cylinder
Volumes.

14 1 : 2-02 : 4-04

186 1 ii

48

MARINE ENGINES AND BOILERS.

Table No. 12.

Name of
Ship.

Minne-
apolis

Victoria
Luise

Hai-Yung

Hai-Shen

Hai-Shew

Arethusa

Buenos
Ayres

Brooklyn

Powerful

Type.

Protected
cruiser

Protected
cruiser

Protected
cruiser

Torpedo
cruiser

Cruiser

Large
cruiser

Armoured
cruiser

.Armoured
cruiser

Triple-Expansion Mngint

Observations.

I.H.P.

The small Mul-f '
I tiple represents! j^'q
the number of I .,
Knin|;es in i *^C
the Ship.

Built by Cramp, Philadel-
phia, for U.S. Navy, 1894.

Built by Weser, Bremen, for
the Imperial German Navy.
Schiffbauy i.. No. 18.

Built at Vulcan Works,
Stettin, for Chinese Navy,
1897-8.

Built by Orlando, Leghorn,

for the Italian Navy.
Engineerings ii., 1 893, p. 756.

Built by Armstrong, 1895,
for Argentine Navy.
Engines by Humphrys,
Ten nan t, & Co.

Engineerings i., 1896, p. 708.

Building 1903 for the
Imperial German Navy.

Built by Cramp, Philadel-
phia, 1898,forU.S. Navy.

Built for British Navy by
Naval Construction Co.,
•Barrow, 1895.
Engineerings ii., 1 896, p. 693.

2 X 6800

13:

3x3800 14C

2 X 3800 1 78

2 X 2200

265

2 X 7000

154

8 X 6838

120

4 X 4500

136

2x12,500 114

THE MAIN KNGINES.

49

Cruisers.

Boiler

Pressure

above

Attn.

Diameter of Cylinders.

The small Multiple represents the
number of Cylinders.

II. P.

M. \\

'** ^' ft

aq. in. II

156

in.
6

L.r.

Stroke.

Ratio of
Cylinder
Volumes.

ft. in. ft. in.
;4 11 '7 8

185 I 2 6-3

3 7-7! 2x
4 8

185

2 6

170

155

3 8.^: 2x
4 0

1 Hi

3 4

3 ^-^

4 61

5 0

207 ! 3 0-6

155

In boiler

260

at engine

290

2 8

3 9

4 8-3

3 11

5 10

2x
5 6

7 G

6 0

2x

6 4

ft. in.
3 6

2 5^

2 3A

.2o.g

-^ *:

^ *w*

Speed
of the
Ship.

1 : 1-97: 4-79

1 :2IO:5-76

1:2-3 :506

1 61 1 : 2-42 : 543

3 0

3 4

3 6

1 : 2*25 : h^h

1 : 2-37: 612

1:2-47: 516

4 0

1 : 2-42 : 570

6-9

8-2

7-2

7-8

7-8

8-7

7-4

8-2

Knots.
23

20

23-2

21-9

21-8

D

50

MARINE ENGINES AND BOILERS.

Table No. 13.

Name of
Ship.

Infanta

Maria

Theresa

Iowa

Renown

Yashima

Majestic

Type.

Formidable

Russell

Kaiser

Friedrich

III.

Preussen

Armoured
cruiser

Battleship

Battleship

Battleship

Battleship

Battleship

Battleship

Battleship

Battleship

Triple-Expansion £»x

Observations.

Built for Spanish Navy by
Astilleros del Nervion,
Bilbao.

Engineerings i., 1894, p. 806.

Built by Cramp, Philadel-
phia, for U.S. Navy, 1897.

I.H.P.

Tbe small Mul-

tiplereprescDtM

the number of i

Engines in

the Ship. I

2 X 6800

2x5ii00

Built by Maudslay, London, 2 x 6000

for British Navy.
Engineerings i., 1896, p. 79.

Built by Armstrong for
Japanese Navy. Engines
by Humphrys.

Engineerings ii., 1898, p. 850.

British Navy. Engines by
BarrowShipbuildingCo.,
1895.

Engineerings ii., 1898, p. 830.

British Navy. Engines

by Earle, Hull, 1899.
Engineerings ii., 1 898, p. 830.

New British Battleship.
Engineerings ii., 1898, p. 830.

Built at Imperial Dockyard,
Wilhelmshaven,1900, for
Imperial German Navy.

SchiffbaUs i.. No. 18. .

Building 1903 for the
Imperial German Navy.

2 X 6750

2 X 5700

2 X 7500

2 X 9000

3 X 4870

3 X 3530

THE MAIN ENGINES.

51

Ironclads,

Boiler

Pressure

alx>ve

Atm.

Diameter of Cylinders.

The small Multiple represents the
number of Cylinders.

II. P.

sq. in. j It- in.

143 3 6

160 3 3

155

155

Ido

lo boiler

300

at engine

250

In boiler

300

atenKtoe

250
171

192

3 4

3 4

3 4

2 1\

2 9it

2 10*6

8 0

M.P.

L.P.

Stroke.

ft in.
5 2

4 7

4 II

4 11

4 11

4 H

4 6J

4 6-3

4 9

ft. in.

7 8

7 1

7 4

7 4

7 4

7 0

2x
5 3

7 2

7 4

ft. in.
3 10

4 1

4 3

3 9

4 3

Ratio of
Cylinder
Volumes.

1 : 2-18: 4-79

1:1-98: 4-75

1:2-16:4-8

1 : 2-16: 4-8

1:2-18:4-8

4 3

4 0

3 OJ

8 3§

Total Expansions
with Cut-off 70 7,
in H.p. Cylinder.

Sjjeed
of the
Ship.

Knots.

6-8

20-2

6-8

17

6-9

18

6-9

6-9

17-5

1 : 2-68 : 71

1:2-65:71

1 : 2-49 : 6-24

1 : 2-45 : 59

10

10

8-9

8-5

18

19

52

MARINE ENGINES AND BOILERS

•

Table

No. U.

Triple-Expansion Eugi

Name of
Ship.

Type.

Fast
Steamer

Observations.

I.H.P.

The small Mulj

tiple represents

the number or

Engines in

the Ship.

1

Augusta
Victoria

Built for Hamburg-Ameri-
can Line, at Vulcan
Works, Stettin, 1889.

2 X 6000 '

Spree

1

1

Fast
steamer

Built for North derman
Lloyd, at Vulcan Works,
Stettin, 1890.

1 X 12,750

Campania
Lucania

Fast
steamer

For Cunard Line, by Fair-
field Co., 1893.
Engineerings i., 1893, p. 480.

2 X 15,000

Trave
Saale

Fast
steamer

North German Lloyd. New
Engines at Vulcan Works,
Stettin, 1895-97.

1 X 8700

Fiirst
Bismarck

Fast
steamer

Hamburg -American Line,
at Vulcan Works, Stet-
tin, 1891.

2 X 8200

Kaiser

Wilhelm

der Grosse

Fast
steamer

North German Lloyd, at
Vulcan Works, Stettin,
1897.

Engineerings i., 1898, p. 364.

2 X 14,000

Nile

Mail
steamer

Royal Mail Steam Packet
Co., by J. ^r G. Thomson.
Engineering, ii., 1893, p. 370.

1 X 7700

Majestic

Fast
steamer

White Star Line. Harland
& Wolff, 1890.

2 X 8500

City of
Paris

Fast
steamer

Inman Line. Thomson,
Clydebank, 1889.

2 X 9200

Prinz-

Regent

Luitpold

Freight and

passenger

steamer

North German Lloyd, by

Schichau, Elbing.
Engineering^x.y 1895, p. 338.

2 X 2800

THE MAIN ENGINES.

5:3

Fast Steamers.

Boiler

Pressure

above

Diameter of Cylinders.

The small Multiple represents the
number of Cylinders.

Stroke.

Ratio of
Cylinder

xpansions
Cylinder.

Speed
of the

Volumes.

Total E
with Cu
in H.P.

Ship.

Atm.

H.P.

M.P.

L.P.

lb. per
sq. in.

150

ft. in.
3 6-3

ft. in.

5 7

fl. in.
8 10

ft.
5

in.
3

1:2-62: 6-6

9-4

Knots.

18

(mean)

156

2x
3 U

6 2f

2x
8 ^

5

10 J

1:20 : 6-92

9-9

18i

(mean)

165

2x
3 1

6 7

2x
8 2

5

9

1 : 2-28 : 701

9-9

21

(mean)

163

3 8

5 10

9 0

6

0

1 : 2-53 : 603

8-6

18

156

3 7

5 7

8 10

5

3

1:2-38: 602

8-6

19-5

178

4 4

7 5|

2x

8 0^

5

^

1:3 : 6-9

9-9

22

(mean)

160

3 2

5 0

7 10

5

6

1:2-49: 6-12

8-9

17-25

180

8 7

5 8

9 2

5

0

1:2-5 :6-54

9-3

19

(mean)

150

3 9

5 11

9 5

5

0

1 : 2-41* : 63

9

19
(mean)

175

2 ^

3 10

5 10-8

3

11

1 : 2-65 : 625

8-9

15-5

54

MARINE ENGINES AND BOILERS.

Table No. 15.

Triple-Expansion En\^

I.H.P.

Name of
Ship.

Type.

Freight
steamer

Obscn-ations.

North German Lloyd, at
Vulcan Works, Stettin.

The small Mul

tiple represeni

the number oi

Engines in

the Ship.

Krefeld
Aachen

1 X 1750

1 screw.

Iberia

Passenger

and cargo

steamer

Built at Felton Works,

Liverpool. 1 screw.
Engineerings ii., 1893, p. 206.

1 X 4500

Kherson

Passenger

and cargo

steamer

For the Russian Volunteer
Fleet, by Hawthorn Leslie.
2 screws.

Engineerings ii., 1896, p. 800.

2 X 6650

Stephan

Cable
steamer

Built at Vulcan Works,
Stettin, for the North
German Ocean Cable
Works, 1902-3. .

2x1200

Giralda

Pleasure
yacht

By Fairfield Co., Glasgow.

2 screws.
Engineerings i., 1895, p. 11.

2 X 4250

Columbia
Alma

Cross -chan-
nel passen-
ger boats

By Thomson, Clydebank.
Engineerings i., 1895, p. 209.

2 X 1870

Speedy

Pleasure
yacht

By Ramage & Ferguson.
Engineerings ii., 1 896, p. 241 .

2x 200

Hermes

Steam
trawler

By Hall, Aberdeen.
Engineerings i , 1894, p. 352

Ix 418

Delaware

Oil steamer

Dunlop & Co.
Engineerings'i-s 1894, p. 209.

1 x 2680

Sylvania
Carinthia

Cattle
steamers

Cunard Line, by London
and Glasgow Engineer-
ing Co.

Engineerings ii., 1895, p. 539.

2 X 2725

THE MAIN ENGINES.

55

Various Shi/fs,

Diameter of Cylinders.

n The small Multiple represents the

i'ressure number of Cylinders.

above 1

Attn.

lb. ^gtr
s(|. m.

178

H.P.

M.P.

L.P.

ft. in.

1 lOJ

ft. in.

3 2i

ft. in.
5 Of

180

2 9

4 6

In boiler

250 3

at engine

lt>0

185

4 9

1 7-6

170 2 1

2 7-0

160

1 7

180

1 0

160 ' 1 0

3 4

2 5

7 4

7 8

4 3

160

2 3

170 I 1 lOi

1 7

1 7

3 1\

3 Oil

2x
3 9

2x

2 H

2 6
2 %\
5 10
5 0

Stroke.

Ratio of
Cylinder
Volumes.

ft. in.
5 0

4 1-2

4 0

3 4

2 3

2 6

1 9

2 0

4 3

4 0

1 : 2-^5 : 7*05

1:2-67: 711

1 : 2-51: 6-53

1 : 2-56 : 676

1 : 2-56 : 648

1:2-33:603

1 : 2-51 : 624

1:2-51:7-21

1:2-6 : 672

1 :2-63:7ll

H-5.S

101

10-2

9-3

9-7

9-3

8-9

8-9

10-3

9-6

101

Speed
of the
Ship.

Knots.
13

16

19-5

20-9

19-3

13

10-5

12-3

15-3

56

MARINE ENGINES AND BOILERS.

Table

: No. 16.

Quadru^

pie-Expansion Jl*

'%« e

i.H.r.

Thesinall Mul-

n

r^ame ol

Type.

()l.)servalions.

tiple represent

No. ol

Ship.

/ r *

the number of
Engines in

Revs.

I'orpedo-

the Ship.

Cushing

Built by Herreshoff, Bristol.

1 X 1720

boat

5 cranks. Inst. Am. Soc.
Naval Engineers, 1890.

Torpedo-
boat

BySchichau,Elbing,1891.
Busley, " Entwickelung
der Schiffsmaschine."

1x1714

320

Northwest

Passenger
steamer

For the American I^kes.
The Engineer, i., 1 895, p. 1 14.

2 X 3500

120

Victoria

Yacht

Hamburg-American Line,

2x2100

123

Luise

by Blohm <V Voss, Ham-
burg, 1900.

St Louis

Fast
steamer

By Cramp, Philadelphia.
Engineering, i., 1895, p. 800.

2x8000

85

Deutsch-

Fast

Hamburg- American Line,

2x17,000

78

land

steamer

at Vulcan Works, Stet-
tin, 1900.
Engineering,\\., 1900, p. 662.

Emperor

Fast

Built at Vulcan Works,

2 X 20,000

80

William

steamer

Stettin, for North Ger-

II.

1

man Lloyd, 1902-3.
Engineering, ii., 1903.

Patricia

Passenger

and cargo

steamer

Hamburg-American Line,
at Vulcan Works, Stettin,
1899

Schiffbau, 1900, No. 18.

2 X 2800

74

Friedrich

Passenger

North German Lloyd.

2 X 3500

75

der Grosse

and cargo

Vulcan Works, Stettin,

Konigin

steamer

1896.

I.uise

Kensing-

Passenger

International Navigation Co.

2x4150

86J

ton

and cargo
steamer

T.&G.Thomson,Clydebank.
EngineeringSx., 1894, p. 199.

Singapore

Cargo
steamer

By Fleming & Fergusson,
1889. Busley.

1 X 1600

Fonar

1

Cargo
steamer

By Wigham Richardson,
1889. Busley.

1 X 1690

THE MAIN ENGINES.

07

Various Ships,

Diameter of Cylinders.

Tb« small Multiple represents the number
of Cylinders.

H. \\

1st
M.r.

2nd

M.P.

L.r.

ft. in. ft. in. I ft. in. ft. in.

0 Hi 1 4 ,1 lOJ, 2x

1 lOJ

1 4-9 2 0 2 9 3 6J

I

2 1 3 0 4 3J 6 2

17 2 3J

2x

3 3 4 9

4 7 6 5 2x
•2 4J 6 5

2x 6 1-6 8 7-9 2x
3 0-61 8 10-3

2x

2x
3 1-4 4 1

2x
6 3

1 11 2 9J 4 0

2 1 3 2 4 4-3

2 IJ 3 U

2 0 2 6 3 4 5 0

4 4i

2x

9 4

5 10

6 3J

Stroke.

ft. in.
1 3

1 6

3 6

3 0-2

5 0

6 0-8

5 10

4 7

4 7

6 2 ,4 6

3 6

19 2 5

3 6

5

0 3 6

Ratio of Cylinder
Volumes.

1:2-04: 4-03: 817 ! 11-7

1 : 201 : 3-81 : 63

1 : 207 : 4-25 : 878

1 : 2-08 : 4-25 : 8-9

1 : 1-86: 3-66: 7-3

1:21 : 4-16: 8-45

1:1-78:4:9

1 : 2-15 : 446 : 9-46

1:2-3 : 4-3 : 9

1 : 2-16 : 42 : 8-4

1 :l-8 :2-8 : 6-4

1:1-9 :4 : 816

Total Expansions
with Cut-off 70 7,
in H. p. Cylinder.

Speed
of the
Ship.

Knots.

11-7

22-5

9

12-5

17-5

12-7

10-4

19-5

12

23

for

75 7o
12

22^

13-5

13-5

12-9

15

12

15-8

91

11-7

]

58 MARINE ENGINES AND BOILERS.

§ 28. The Crank.— We shall in future use
p to denote the pull or pressure on the piston rod (pounds).

pull or pressure on the connecting rod (pounds),
pressure on the guides (pounds per square inch),
tangential force acting on the crank circle (pounds),
length of the connecting rod (feet),
radius of the crank (feet).

ratio of crank radius to length of the connecting rod.

momentary speed of the piston (feet per second),
momentary circumferential speed of the crank (feet

per second),
momentary angular velocity (feet per second),
mean piston speed (feet per second),
angle formed by the connecting rod and piston rod.
angle of the crank at any given point reckoned from

the upper dead point.

The following observations hold good for vertical engines : —
From Fig. 27 \ht pull or pressure on the connecting rod is p* =

pi

ij

K

i»

T

n

/

>i

r

>»

_r

~ I

>»

V

»»

w

>»

€

♦»

C

1)

a

»>

OJ

>i

cos a

Pressure on the guides^ k = p tan a,

K is greatest when the connecting rod is
in its most oblique position (assuming that
p is constant throughout the stroke). This

occurs when oj = 90". Then tan a^ - , . =

As ^\—X' is almost equal to

x/l - X2'

unity, the approximate formula K„ax = P-y

= p . A. . is generally used to calculate the
greatest pressure of the guides.

The tangential force in the crank circle is

T = pi sin (a + co) =

cos a

sin {a + to))

If the connecting rod be of infinite
Fig. 27. length, then « = o ; therefore t = p sin co.

Circumferential Speed and Angular Ve-
locity of the Crank.— The first is understood to mean the arc of
the circle traversed in a unit of time. Let dtm be the angle travelled
through in an infinitely short space of time dt, then rd<a is the

THE MAIN ENGINES. 59

arc of the crank circle passed through in the time dty and there-
fore w = — -- will be the varying circumferential speed of the crank.

The mean circumferential speed is —

_ 27rr . « _ n

60 9-55

By angular speed c is meant the circumferential speed in the circle
having a radius equal to unity. This is —

_dia _ 27r« _ n _

^~~di' ^'" " "60" " 9^5' ^"'-^'"•''

Piston Speed. — The table on page 41 gives the usual values for the

mean speed of the piston —

s , n

Here the value assigned to the variable piston speed v for the angle of
rotation of the crank w must be determined. The travel of the piston
for a "connecting rod of finite length " (see Fig. 27) is —

X = r(l - cos w) + /(I ~ cos a).
As—

/sin a. — r sin w, or sin « = -^ sin w = X sin w,

then cos a = Jl - k- sin*-^ w.

If this be worked out we get —

cos a = 1 - JA2 . sin^ w.
and by inserting the values for x we get —

X = r(l - cos w) + ^A.2 sin^ cd

The piston speed at a point situated at a distance x from the upper
dead point thus becomes —

dx dx dui dx w
dt d<a dt dta r

Differentiating this we get —

V = (r sin CO + i — 2 sin ai cos (u)— = w(sin <i> + JA. sin 2w).

For " connecting rods of infinite length " A = o and v = w sin a> — that is,
the piston speed is greatest at the middle of the stroke (w = 90") and is
equal to the circumferential velocity of the crank for this value of co,
while at the upper and lower dead points it is = 0.

60

MARINE ENGINES AND BOILERS.

Acceleration of the Piston, — By this expression is meant the change
in the piston speed during a given time <//, and it is —

d\ dw dm dw, .

k .

W

B = -j- = ---X-=- = -p (sin w + -. sin 2o)) + — (cos w + A cos 2o)).
at a(a at dt J* T

dw
If the circumferential speed w of the crank is uniform, we get — - = 0, and

therefore —

B = — (cos w + A. cos 2co).

r

W2

For "connecting rods of infinite length" b = — cos co — that is, the

r

acceleration of the piston is greatest at the dead points, and is then

= — , while it is = 0 in the middle of the stroke.

r

Acceleration of the Rotating Masses, — The acceleration of a non-uni-
formally rotating crank pin consists of —

, Radial acceleration — and

' r

Tangential acceleration -— ;

and it may be divided into a verti-
cal and a horizontal component.
According to Fig. 28 the vertical
component is —

^ . . w^

-3- sm w + — cos w,

dt r

Fig. 28.

and the horizontal component is-

w*^

-;- COS w H sm CO.

dt r

d\v

If the rotation of the crank pin is uniform -j- = 0, because
w = constant, and therefore the vertical component w^ill be — cosw,
and the horizontal component — sin &>.

S 29. Moving Parts of the Steam Engine.— It is convenient

to divide these into classes {a) those moving to and frOy or having a
reciprocating motion^ and (b) those having a rotating motion. This
division gives rise to no inconvenience, except in the case of the con-
necting rod. In marine engines the ends of these rods, viz., the crank-
pin end and the crosshead, are so heavy, that the weight of the rod forms

THE MAIN ENGINES.

61

only a small part of the total weight. We may therefore practically class
the head and lower half of the rod with the rotating, and the fork and
upper half of the rod with the reciprocating parts of the engine.

Fig. 29.

The weight of the connecting rod is generally divided up as follows : —
Let s be the centre of gravity of the rod, then the weight of the

h a

rotating part = Gy, and of the oscillating part = G- . Therefore, in future,
let the reciprocating or oscillating parts be understood to be the

> = Mo = total mass

= M, = total mass

Piston

Piston rod

Crosshead

Fork of the connecting rod

Rods of the air pump, &c.

and the rotating parts to be the

Head of the connecting rod

Crank

Crank shaft and shafting

Propeller, &c.

In the same way the masses of
the valve gear may be divided into
reciprocating and rotating parts, and
they will, where necessary, be in-
cluded under Mo and m,.

Those parts, the centres of
gravity of which lie off the centre
of the shaft, have a considerable in-
fluence in the calculations employed
in the balancing of the moving parts.
They are called the eccentrically rotat-
ing parts. For instance, the eccentrically rotating parts of the crank

Fig. 30.

62 MARINE ENGINES AND BOILERS.

shaft consist of the crank pin and the shaded part of the crank cheeks
in Fig. 30. The eccentric rotating parts of an engine are —

(1) The eccentric rotating parts of the

crank shaft v = m

(2) Connecting-rod head and end of

the crank shaft

To these we must add the eccentrically rotating part of the valve gear
and reversing link, and the lower part of the eccentric rod.

Reduction of the Moving Parts to the Circle described by the Crank. —
To simplify the calculations, it is usual to reduce all these masses to the
crank circle.

(1.) Reduction of the Masses relatively to their Radial Acceleration,
— The radial acceleration (centrifugal force) of a mass rotating at a
uniform distance r from the centre of the shaft is (see page 60) —

= = Mfie''

r r

If it rotates at the distance ^, its radial acceleration is —

MW,2 M^jM

L = = Mtfc^

a a

The radial acceleration of the rotating masses varies as the distances of

their centres of gravity from the centre of the crank shaft. In the same

way the radial acceleration of the oscillating masses varies as the radii

of the cranks upon which they are worked. The reduction is made by

multiplying the mass by the ratio of the radius of rotation of the centre

of gravity to the radius of the crank.

The reduction of the reciprocating masses to the crank circle is

arrived at in the same way. If, for instance, a pump rod with stroke s^

is driven from a crosshead with stroke j, through a lever, the mass of

the pump rod is reduced to the circle described by the crank by

s
multiplying it by J.

(2.) Reduction of the Masses in relation to their Kinetic Energy,
— The kinetic energy of a rotating mass m is —

^w2= -r2€2=L:€2
2 2 2

where i^ = Mr^ is the moment of inertia of the mass m at radius r. If
the mass rotates at the distance a from the centre of the shaft, its kinetic

I £2

energy = — , and i, will be its moment of inertia for radius a. But

THE MAIN ENGINES. do

if the mass rotates so that its centre of gravity is in the centre of the

shaft, its kinetic energy = -^, and i is its polar moment of inertia

about that axis.

The reduction of the masses to the circle described by the crank is
effected by dividing their moments of inertia by r^, and the quotient
gives the desired result.*

§ 30. Tangential Pressure on the Crank Pin, and Turning
Moment of the Multiple-crank Engine.— The turning moments

of all the cranks upon the shaft of the engine are generally made as
uniform as possible during one revolution, in order to diminish the
variations in the turning moment on the shaft, and to reduce, as far as
possible, any torsional strains that may arise. To determine the con-
ditions necessary for a uniform turning moment, experiments are made
on actual engines; in other words, they are tested to determine the
energy of gyration in the crank circle.

Graphic Method of Determining the Turning Moment in Engines
(for exannples see page 65 and Figs. 32-35). — The steam pressure on the
piston during one revolution (differences of pressure between the top and
bottom) is obtained from indicator diagrams, and is plotted as abscissae
on a base representing the travel of the piston. The radial acceleration
or centrifugal force of the reciprocating masses of the connecting rod is
added graphically. This latter pressure (see page 60) is

= - - (cos CO + A. cos 2(1)),

where m -= - = the mass of the-

[ Piston

Piston rod

Crosshead

Upper part of the connecting rod
^ Air-pump rods

reduced to the circle described by the crank.

This simple formula is used to determine the centrifugal force,
because to assume that rotation is not uniform, and to take count of
frictional resistance, would make the study of the turning moment
too complicated. As the vertical and horizontal components of the

* Care must be taken to distinguish between the diflferent kinds of rotation of, for
instaiice, the crank pin and the connecting-rod head. The moment of inertia of the
first consists of the product \\^ and of the polar moment of inertia round the centre
line of the crank pin ; that of the connecting-rod head is only equal to the product
Mf^, as it has approximately no rotation round its own centre of gravity.

64

MARINE ENGINES AND BOILERS.

Mg = the weight of the

rotating parts combine to produce a radial force, the. vertical com-
ponents are here neglected.

The weight* of the crank cheeks, which is a positive quantity as
they descend and a negative as they rise, must be added to the steam
pressure and to the radial acceleration. Thus the total weight to be
considered is —

TPiston

Piston rod

Crosshead

The whole of the connecting rod

Air-pump rods

.Eccentric portion of the crank shaft

reduced to the circle described by the crank. The weight is reduced
to the crank circle in the same way as for the radial acceleration. The
frictional resistance, weight, and radial acceleration of the valve gear are
not usually taken into account in these calculations, as their influence
upon the turning moment of the engines is small.

If the sum of the steam pressures, centrifugal forces, and weight
of the rods (plotted as abscissas over the travel of the piston) be
drawn as a line representing the vertical force p for the different

,. , , , . , , , . sin (tf + w)
cyhnders, then the tangential pressure on the crank pm r = p >-

(see Fig. 27, page 58).

Graphic Determination of t w/ten p is given for any position of the

Fig. 31.

Crank (Fig. 31). — Draw ab parallel to the connecting rod, and bc
parallel to the crank. Plot the pressure on the piston p from b along bc,

* Strictly speaking, the tangential pressure due to the weight of the crank and
the connecting-rod head should be separately estimated, because the path it follows is
always a sine curve, and therefore T, constructed as in Fig. 31, is not accurate.

THE MAIN ENGINES.

65

SO that BD= p.. Draw de at right angles to ac. Then according to the
law of sines de is the required tangential force t. t must be drawn for
each separate cylinder, and plotted over the developed crank circle.
The curves of t thus obtained for each cylinder are added together
graphically, and the resultant curve represents the total variations in the
turning force. The product t . r is the total turning moment of the
engine at each instant, during one revolution.

The total turning moment t for one revolution is represented by the
area below the curve t between
the ordinates ab and cd, Fig. 32.
Work of one revolution = area abcd

-/ o

Trdm = T^ . 27rr

:*■

Fig. 32.

Here t„ is the mean turning

force on the crank pin, and is

found by taking the area abcd

with a planimeter, and dividing

the result by 2irr. The value

T. . T is called the mean turning moment of an engine. The values

Taa,.r and Tmi^.r are respectively the maximum and minimum

turning moments. The mean turning moment, determined graphically,

must be equal to that calculated direct from the work shown by the

indicator diagrams and the number of revolutions (/;), according to

the well-known formula —

Tm./'=

I.H.P.

n

X 5,252 (t„ in pounds, r in feet).

As a rule ^^ = 1-5 to 25, l2!i; = M to 1-5.

-m

The total turning moment is expended in overcoming frictional and
propeller resistances.* If the mean value of the sum of these two
resistances for one revolution be denoted by q^ we get the equation
Tni = Qni. (For the fluctuations in propeller resistance see "variation
in crank-pin velocity," page 69.)

Example. — To determine the tangential pressure of a three-crank
engine of about 6,650 i.h.p. Diameter of cylinders, h.p. 3 feet 3|
inches, m.p. 5 feet 4 inches, l.p. 8 feet 5 inches. Stroke, 5 feet 3 inches.
r=3Ii inches ( = 2*625 feet). « = 75 revolutions per minute.

* Propeller resistance is the resistance offered by the propeller to turning.

E

66

MARINE ENGINES AND BOILERS.

Weight G for H.r. cylinder, 24,861 lb. ' Weight MgforH.P. cylinder, 9,41611).

,, ,1 M.P. ,,
L. P.

i» »»

>i

26,840 „

28,820 „

»» »»

Yl l»

M.P. „ 11,440 ,,
I., p. „ 13,420 „

Mean speed of crank pin, w,„ = '^-n^ = 20*60 feet per second. The
indicator diagrams are given in Fig. 33.*

Ta

(I /d. per sq. in. = '137 itt^h.)

MS

fa/ m 10 nttf

(1 /d. per sq, in, = "027^ inch. )

fat. 30mm

(1 ib, per sq. in, = '083 itich,)

Fig. 33.

Fig. 34.

Determination of the Vertical Force p (Fig. 34). — All the weight
pressures are calculated and reduced to pounds per square inch c
area of the l.p. cylinder, cc gives the difference in pressure abov
below the piston plotted from the axis a.\ ; gg gives the weights p
from AA (h.p. 24,861 lb. = 3-04 lb. per square inch reduced to tb

* The scale of the springs refers to the original diagram and has he
reduced to one-third.

THE MAIN ENGINES.

67

of L.P. cylinder, &c.). mm gives the accelerating forces due to inertia
calculated from the formula m' = * (cos w + X cos 2o)) with A. = - = i

^ttfoJ p99p itt)^

Fig. 35.

^uicjpp^p d»j^

plotted from gg. The abscissae 0, 2, 4, 6, 8, &c., correspond to angles
of the crank C, 30\ 60% 90% 120', &c., for connecting rods of given
lengths. The ordinates between the lines cc and m'm' give the required
vertical pressure p. The tangential pressures t for the h.p., m.p., and

68 MARINE ENGINES AND BOILERS.

L.P. cylinders (Fig. 35) can be determined from the values of p obuined
from Fig. 31. These may be combined in the tangential pressure

curve (Fig. 36), regard always being had to the proper sequence of
the cranks. Hence we have ^=1-7. :^ = l-25, t„,= 177,40O lb,
T„ may also be calculated direct from the indicated work according to

THE MAIN ENGINES.

6d

the formula t„ = li^ x 5,252 x 1 = ^4^ x 5,252 x -i— ; whence t„. =

n r Id 2'62o

177,4001b. as above.

§ 31. Variation in Crank -pin Velocity.— If w„„ be the

highest, w„in the lowest, and w^ the mean value of the velocity of the
crank pin during one revolution, the degree of irregularity will be —

W

m

Determination ofh, — The resistance of the propeller q varies very
little during one revolution, if the shaft is not subjected to great varia-
tions of torsional strain. It may therefore generally be taken as a
constant, and (^ = (^,^^. The areas f^, f.^, F3 between the curve t

Figs. 37 and 38.

and the line Q (Fig. 37) thus represent the positive or negative work
of the turning force t in relation to q. This work serves to increase
or diminish the kinetic energy of the moving parts. Hence the follow-
ing equation holds : — ^Work between w^ and w = variations in the kinetic
energy between w„ and w, or —

- (w2 - w2„) = / (t - cordis} = area rstu - ru x q
2 -^ w„

0)

The curve of crank-pin velocity (Fig. 38) will thus change its direc-
tion when the maximum or minimum of the turning force is developed,

70 MARINE ENGINES AND HOILERS.

and will show a maximum or minimum speed when the curve t cuts
the line q. The greatest crank-pin velocity will be at the point where

the algebraical sum of the successive areas f^ + Fj, + Fg + reaches

its highest positive value = x, and the lowest speed where it attains its
highest negative value = y. The change in kinetic energy between its
maximum and minimum values corresponds to the difference between
the greatest positive and negative areas. Therefore —

Note that y must have a negative sign before it. The degree of irregu-
larity agrees with the definition —

W

We may take approximately —

From Equations (2), (3), and (4) we deduce —

m5w2,„ = X - y (5)

6 can be calculated from Equation (5). If 8 be known, w„„ and w„,i„
may be found from Equations (3) and (4). The velocity at any given
moment (corresponding to the angular velocity w) can be obtained from
Equation (1). In place of an arbitrary speed at starting, w^, the known
speed w„«x may be substituted. To obtain this —

— (w2 - w2„,j,) = / (t - Q)rd(i) = L = difference in the work of

T and Q betw^een the ordinates rs and ut. Hence the

speed required w = ^w2„,„ + — (C)

With the help of this equation the curve of crank-pin velocity may be
deduced from the diagram of tangential forces.

By mass m is understood to mean here all the rotating parts reduced
to the crank circle, including the screw, shafting, &c. (see pages 60 and
63) ; also those parts of the reciprocating masses which contribute to
the turning moment of the engine. To calculate all these exactly would
be too long and complicated a process. We may assume for them the
following values : — !! of the reciprocating mass of one connecting rod, in
engines with three cranks at a less angle than 120** ; i of the reciprocating

THE MAIN ENGINES. 71

mass of one connecting rod, in engines with four cranks at a less angle
than 90^, the masses of the connecting rods being the same.

In vessels with long shafts of small diameter, the regulating or
steadying effect of the screw does not extend to the engine end of the
shafting, and the degree of irregularity is greater near the engine than
close to the screw. This irregularity 8 is seldom more than 6 % at the
screw, and 12 % at the engine. In all cases where great irregularity in
running is noticeable, it is caused by the great variations in torsional
strains on the shaft. (Compare on this point 2^itschrift des Veretnes
Deutscher Ingenieure^ 1902, dJ\A /ahrbuch des Schiffbauttchnischen Gescll-
schaft^ 1903, page 403.) In engines with short, thick shafts (fast
steamers, ironclads, ice-breakers) it is the same throughout the whole
length of the shaft, and varies between 4 7o and 7 %. In ships,
therefore, with long, slender shafting, the mass of the screw must not be
included in m, when calculating w. If the curve t be regular, and m
relatively large, as in ships with a heavy screw and built-up crank shaft,
Q may be taken as a constant when calculating w from Equation (6). In
other cases the curve of speeds is determined, assuming a constant resist-
ance of the propeller. As this, according to the latest experiments, is
approximately proportional to the 3*8 power of the circumferential
velocity, q =» ^tw^"^, the value of q may be found by means of the first
approximate value of w, and being thus obtained, the correct curve
of speeds can be determined (compare Fig. 39),

§ 32. Example. — To determine the mean crank-pin velocity curve,
the mean curve of tangential forces given on page 68 must be used.

The ship is supposed to have comparatively short and thick shafting,
therefore all the rotating parts may be understood as included under m.

The total mass m reduced to the crank circle is

G 110,000 , , ., ,.,^
M = - = — i^- — = about 3,437
g 32

corresponding to a reduced weight of 110,000 lb.

The areas Fj, Fg, Fg, &c. (marked i, ii, &c., in Fig. 36) are : —

ft. H).

i = Fi= - 1352-5 I ft. lb.

11 = ^2= + 6762-8 ' Fi + Fj, =+ 5410-2

iii = r3=- -5^7-7 , F1 + F3 + F3 =+ 2842-5

IV = F4= +13019-4 I Fi + Fg+Fj+F^ =+15861-9

Fi + F.,+ F3+F4+F^ . . . =-43086-9=Y

Fj + F?, ... +F^j . . =+ 43492-0 = x
F1 + F2 . . . . +F7 . = -31875-8
VIII = F8= +31875-8 F1 + F.2 +F8= 0

Therefore x - y = 43,492-0 + 43,086-9 = 86,579 ft. lb.

v = r5= -58948-9
vi = Fg= +86579-0
vii = F^= -75367-8

72

MARINE ENGINES AND BOILERS.

7

¥

^

i

T

■ B ■

V

'4«'^p»p«r^;^

^ fc •* ***
^ ♦* «^ •*'

V tf ■

^y V^

5

ULQmtUff

d^-fuvt^i^-^

i

M

Fig, 39.

THE MAIN ENGINES.

73

T

d^ <iytnn«9'^)^

« •• « o •

mn

^ trt" «*r iH* Cr

• • • « •
V V V V V

MM ml

^m

^
^

/jgj^^^l^

w w %i- Sr w

S «^ ^ ^- — i5~

& /«l

STi

t

d^'^yrwe^^if)^

Fig. 40.

•itei ►* ^» S

MARINE ENGINES AND BOILERS.

B can now be calculated, and is according to Equation (5) —

8 = ^^ = „_^^^5!1__ , = 0059 = 5-9 %, say '06 or 6 %.
M.w?„ 3,437 X 20-6- ^''' ^ ^°

From Equations (3) and (4) —

" IIUU "ill

W •
" iniii

M+_j=21-2 feet per second.
= w,„n - ,^) = -0*0 feet per second.

The remaining speeds are obtained from Equation (6), by inserting
the areas of work, f- (vii), f^ + Fg (vii + viii) successively, in place of l.
For example —

w- = , A 1-22 - '^1^^ = 20-2 feet per second.
' V 3,437 ^

In this calculation the frictional and propeller resistance q is assumed
to be a constant throughout the revolution. To calculate the speed
curve more accurately, the above curve of w should be used to deter-
mine the curve of resistance q according to the formula q = kw^\ The
factor k may be approximately obtained from the formula —

III

The corrected speed curve is calculated, as before, from the positive
and negative areas between the curves of t and Q.

§ 33. Explanation of Figures. — Fig. 39 shows (for a fast steamer
with two four crank engines, balanced on the " Schlick " principle) the
curves of turning moment and of the calculated resistances Q, as well as
the speed curve, calculated from various approximations. In Figs. 39
and 40 the crank-pin velocities are denoted by v instead of w, and the
mean of these speeds by c. Fig. 40 shows how the calculated speed
curves and those obtained by experiments agree, in a cargo and pas-
senger steamer with two four-crank balanced engines. Curve No. II.
has been calculated as described ; curves 5, 6, 7, and 8 are taken
from experiments.!

§ 34. Variations in Torsional Strains of the Shafting.—

The points considered in §§ 31 to 33 hold, on the assumption that the
shaft, together with the masses of the crank shaft and of the propeller,

* The value for the exponent of wj* requires confirmation by further experiments,
t For further details see Journal of the Schiffbcaitechnischen Gesellschaft^ vol. i. ,
1899, p. 311, &c.

THE MAIN ENGINES. 75

is not subject to varying torsional strains. By this term is understood
oscillations which succeed each other in the same way as in the balance
wheel of a watch ; the masses of the crank and propeller respectively
playing the part of the escapement, the elastic shafting, subjected to
torsional strains, representing the hair-spring.

Assuming the engine to be stationary, the propeller fixed, and a
turning moment p . r to be transmitted from the engine to the shaft, the
latter will first twist, />., the cranks will turn through a certain small
angle relatively to the fixed propeller. This angle may be calculated as
follows : —

T . , .P.r.L ., jP.r.L 180
In circular measure 9 = or m degrees 9 = . —

G.I G ,1 TT

p . r being the turning moment in inch-pounds, G the modulus of elasticity
of the shaft in pounds per square inch, i the polar moment of inertia of
the section of the shaft in inches '*. If p . r be suddenly released, the
shaft will fly back, and the crank-shaft masses will then oscillatOj setting
up torsional strains about the centre of gravity.

As the elastic force increases in proportion to the deflections from
the central position, the oscillations follow the laws regulating the swing
of an ordinary pendulum. The duration of one complete oscillation to
and fro is thus —

= 2 V'

where /// is the mass of the crank reduced to the crank circle, and k is
the force which must be developed at the crank circle to produce a twist
amounting to 1 inch of arc, measured on the crank circle.

K is determined from the formula for <^ in the following way : —

XT ^ \ f« I

<A . r=s — ^ — '— . r= 1 inch, whence k= V— i" pounds,
G.I r^. L

r being the radius of the crank in inches.
If these values are used, we get —

=V"-t^=^'V

m . L
G . I

For practical use it is better to calculate the number of oscillations to
and fro per minute, namely —

n' =^ = ^ /
^ T TrrV

G. I
m . L

76

MARINE ENGINES AND BOILERS.

Such specific oscillations of the shaft would be produced if the propeller
were fixed.

The number of specific oscillations may be calculated in the same
way, if the propeller is not fixed. According to the law of action and
reaction, a backward swing of the propeller masses must correspond
to a forward swing of the crank masses, and vice versa. As the front
part of the shaft moves in the same direction as the crank masses, and
the after part in the same direction as the propeller masses, there must,
at some place between the two, be a node or point of no movement.
Its position is determined by the condition that, if the external forces
cease to act, the opposing momenta of the propeller and crank masses,
reduced to the crank circle, must, at every instant, be equal to each other.

The momentum of each = mass x speed. But the speed of the
oscillating masses is proportional to the crank radius r, and to the dis-
tance of their fixed points from the node. For it is easy to see that

mm

M

te^-j

m

;«

4

Kv

Ubmd.

Fig. 41.

from the node, both towards the cranks and the propeller, the amplitude
of the oscillations, and therefore the speed of the oscillations, must
increase uniformly.

The momentum of each being the same, we have —

mv^VLV
or by substitution m{kl^ = m(>^/^,), where —

/j is the distance of the node from the centre crank.
4 }} „ „ from the propeller.

^ is a varying factor.

It follows that —

/j _ i(' . M _ M
/., k . ;;/ m

that is, the lengths of the shaft up to the node are inversely proportional

THE MAIN ENGINES. 77

to the masses at either end, reduced to the crank circle. The following
is obtained by a simple transposition —

■^ — ^ = -= whence /j = l . , and therefore '2 = l .

/j /l M * W + M - W + M

Thus the position of the node depends wholly on the ratio of the
masses m and m, reduced to the crank circle. If /// = m, it will be in the
centre : if not, it will be nearer to the larger of the masses.

The number of oscillations per minute, in the case of a freely- working
propeller, is most easily obtained by deducing it from the formula for n^

already given. In the first case the propeller, and in the second case
the node, is stationary ; therefore, instead of l, the lengths l^ and l^ re-
spectively must be inserted, and thus for the forward end of the shaft
we get —

«k

^30 /g'. I

or, if the value l . — ' — be substituted for /,-

w + M *

_30 /g.i.(w + m)
Trr V L . w . M

and the same for the after end of the shaft —

30 / G.I 30 /o . I . (m + m1

Hence n^^n^; i,e., the number of oscillations per minute of the crank
masses (wt) is the same as those of the propeller masses («p). The oscilla-
tions thus described are the actual torsional oscillations of the shaft, and
would be produced if no external forces, especially if no friction or other
retarding force, acted upon the oscillating parts. The propeller and
crank-shaft masses oscillate in opposite and equal phases, t\e,, they reach
their middle positions simultaneously, but swing in opposite directions.

On account of the unavoidable friction, and the impeding action
of the water on the propeller, the oscillations produced would rapidly
become smaller, and ultimately cease. But if there is an external
oscillatory force^ which tends to produce oscillations, these may continue
for any length of time. If the oscillations are of a certain amplitude, a
kind of equilibrium is established, because the external force does as
much work on the shaft, as is annulled by the impeding action of the
friction, &c. These two counteracting forces cause the node itself to
vibrate and describe a sine curve, but the vibration of the latter compared
with the vibrations of the crank and propeller masses is always very small.

The amplitude of the oscillation of the propeller and cranks is, as in

78 MARINE ENGINES AND BOILERS.

all artificially produced oscillations, proportional to the maximum value
of the producing force, and inversely proportional to the impeding forces,
and is further dependent on the variations per minute in the producing
force. If these are considerably larger or smaller than the periodic
oscillations «k = ^p of the shaft, a strong producing force may only
produce small torsional deflections of the oscillating masses. If the
recurrent action of this force be nearly equal to the periodic oscillations
of the shaft «k = «p, the amplitude of the torsional deflections of the
propeller and crank rapidly increase, until they reach their maximum
value ; this corresponds to the critical number of oscillations^ which is
attained when the recurrent action of the force is = «k = «p-

To determine the torsional oscillations produced in actual workings
let it be assumed that all the oscillating parts — viz., the masses of the
crank, shaft, and propeller — are made to rotate at the same number of
revolutions n as the engine, while they are subjected to the above-
described retarding forces and artificially produced oscillations. As a
result of the uneven forces acting on the cranks and propeller, these
move very unevenly and jerkily ; under certain conditions the piston rods
may even appear to stand still at a certain point during each stroke,
though the engine continues running.

The forces which produce torsional deflection when the engine is at
work are the variations which occur in the tangential force (see page 63).
Every tangential diagram, however irregular, may be resolved into a
constant mean force and a series of harmonic oscillations, which describe
complete cycles for each revolution of the engine 1, 2, 3, 4, . . . . The
above method of resolving the diagram is known as Harmonic AncUysis,
On the methods employed see Fischer-Hinnen, Electrotechnische Zeit-
schrift^ and I-orenz, Dynamik der Kurbelgetriehe, The number of
oscillations of these harmonic forces is = «, 2//, 3w, 4« . . . . where //
is the number of revolutions of the engine. If one of these series of
oscillations chances to be equal in number to the periodic oscillations
of the shaft «k = «p, strong torsional oscillations will be produced. The
engine then runs at its so-called critical speedy and according to whether
«, 2«, 3//, &c. = «k = ''p» this critical number of revolutions is said to be
of the first, second, or third order.

The critical number of revolutions of the shaft is independent of the
critical number of revolutions during which the vertical or torsional
oscillations of the shifs body attain their greatest amplitude. The re-
tarding forces at work are the friction of the engine and shafting, and
the rapidly increasing resistance of the propeller, when any alteration in
speed takes place. On the laws governing this resistance see page 71.

When calculating the number of periodic oscillations «k = «p for
practical purposes, the following points should be noted : — One-half of

THE MAIN ENGINES. 79

the oscillating masses, and one-third of the mass of the forward part of
the shaft, up to the node, is to be included in the crank masses. One-
third the reduced masses of the shaft aft of the node is to be added to
the propeller masses ; and from 25 to 30 % must be added to the pro-
peller masses, to allow for the water entrained by it. The reduction of
the masses must be in the ratio of the square of their actual radii to the
radius of the crank circle r.

As the shafting consists of lengths of varying diameters, while the
formula for n^ = w^ requires the introduction of a definite moment of
inertia i, the different shaftings are reduced to the diameter of the
smallest, />., of the tunnel shafting. The torsional angle must be the
same for the reduced as for the actual lengths of the shaft. If /^ is the
actual length, d^ the actual diameter of the shaft, d^ the diameter of the
tunnel shaft, and /the reduced length of shaft, then —

^1 =_ • whence /=/^ =/ (zs]

The reduced value of flanges, thrust shafts, and cranks can only be
estimated approximately. The modulus of rigidity for the material of
which shafts are usually constructed (Siemens-Martin steel of 25 to
30 tons per square inch tensile strength, and above 20 % elongation)
varies very little, and may, according to the latest experiments, be taken
at a mean value of —

G^ 5,260 tons = 11,782,400 lb. per square inch.

Example, — To calculate beforehand the number of torsional oscilla-
tions in an engine of 2,000 i.h.p. —

n = about 75 : stroke = 50 inches : r = 25 inches.

From the working drawings of the engine, the polar moment of inertia
of the rotating parts of the crank shaft works out at 63,510 lb. ft.^

Therefore the weight of the same reduced to the crank circle

= 63,510(i?y= 14,630 lb.

To this must be added : —

The three bottom ends of the connecting rods, totalling = 4,600 „
And half the oscillating masses, including the pump

rods (reduced) - - - - - = 6,600 „

Therefore crank masses (without the shaft) - = 25,830 lb.

80 MARINE ENGINES AND BOILERS.

Similarly the polar moment of inertia of the propeller = 188,700 lb. ft.-
Therefore weight of same reduced to the crank

circle - - - = 1 88,700 x ^i|y= 43,4501b.

To which add 25 % for the entrained water - = 10,860 „

Therefore propeller masses (without the shaft) = 54,310 lb.

Polar moment of inertia of the whole shafting = 13,740 lb. ft.^
Whence weight of same reduced to the crank circle

= 13,740 X ^13y = about 3,166 lb.
To the crank masses must be added

and to the propeller masses must be added

Thus with all additions the crank masses reduced to the crank

25,830 + 715 26,545 ^^ ^
386 386

and the corresponding propeller masses

54,310 + 340 54,650 , ., .
''=^ 386 =":386^-^^^'^'

Here the acceleration due to gravity is taken at 32-2 ft. sec.2 = 386
inches.

Diameter of smallest shaft (tunnel shaft) = 12*5 inches.

Polar moment of inertia of section of the shaft, having a diameter
of 12*5 inches

= 1 = ^ = 2,396 in.^

Length of the shaft reduced to the diameter (12*5 inches) of the
tunnel shaft = 1,592 inches.

We have now all the data for calculating the oscillations. These

work out at

^ ^30 /g . I . (^ + m)
** Trr V L.m.U

30 711,774,160 X 2,396 X (68-7
"7r.25V 1,592x68-7x141-

;= about 236 double oscillations per minute.

X 141-6)
6

THE MAIN ENGINES. 81

Therefore the critical number of revolutions of the shaft are —

First order = ^ = 236

Second,, =?|?=118

Third „ =HJ5 = 78-7

Fourth „ =1_ = 59, and so on.

4

As the engine usually runs at about 75 revolutions per minute, it is
only orders in and iv of the critical number of revolutions which need
be taken into account. If « = 80, very strong torsional oscillations
would be recorded in actual running.*

* Compare Lorenx, Dynamik der Kurbelgetriebe ; GUmbel, TransacUom of the
InstUution of Naval Architects, 1902 ; Frahm, ZeitschHft des Vereims Deutscher
IngMieure, vol. xxxvi., 1902; Foettinger, Jahrbuch der Schiffbautechnischen
Gesellsckafi^ 1903.

K

Balance of the Moving Parts.

§ 35. General Remarks. — The steam admitted to the space
between the piston and cylinder cover exerts upon them both the pres-
sure p. This pressure may be divided into
the radial acceleration Pj and the actual pres-
sure on the piston rod Pg, so that p = Pj h- p.,.
The force p acting upwards on the cover
is transferred through the framing of the
engine to the bed-plate, which it tends to
1^^ "^^ lift. To balance it there is only the pres-

L I I I sure Pj, as the rest of the steam pressure

on the piston p^ is absorbed in accelerating
the piston and connecting rods. The
force Pj is transmitted through the rods
to the bed-plate. During the down stroke
the unused force Pj is exerted to raise
the framing of the engine. In the same
way it can be proved that during the
up stroke the weight of the rods tends
to force the framing of the engine down-
wards. The horizontal components of the
forces seek to move the frame horizontally,
\ in a direction perpendicular to the longi-
tudinal axis of the engine, while the centri-
fugal force of the crank endeavours to drag
Fig, 42. the bed-plate along with it

In the following considerations it is
assumed that —

1. The connecting rod is of infinite length, and that the movement
of the reciprocating masses is continuous, similar to that of the crank

* and the crosshead.

2. The rotation of the crank pin is uniform, and therefore the

acceleration of the reciprocating masses follows the law b= — cos w

(see page 60). (w = angle of the crank calculated from the upper dead
point.)

THE MAIN ENGINES. 88

§36. Balancing the Moving Parts of a Single-crank

Engine. — The radial acceleration of the rotating masses can be com-
pensated for by a counterweight on the crank, the weight of which is
equal to that of the rotating masses. If the counterweight be made
equal in weight to the weight of the rotating and reciprocating masses,
the vertical pressure of the latter disappears. A new horizontal force is
however introduced, which is equal to the radial acceleration of the
reciprocating masses. The reason of this is that the vertical com-
ponents only of the counterweight, which act upon the reciprocating
parts, are absorbed by the latter. By the use of such a counterweight
the radial acceleration of the reciprocating parts is deflected through
an angle of 90°.

§ 37. Balancing the Moving Parts of a Two-crank

Eng^e. — For marine engines, only those which have their cranks at
an angle of 90° need be considered, as with any other crank angle the
engines work very unevenly.

Let M^ and M2 be the reciprocating, and a^ and
Ao the rotating masses of the h.p. and l.p. cylinders
respectively, then the value of the vertical components
will be given by the equation —

w^
Pj = (M| +»,) — cos 0) for the h.p. cylinder.

w- \]

P.2 = (Mo + ».>) — cos (w ± 90) for the l.p. cylinder — ' ^

(see Fig. 42), the plus and minus values depending fig- 43-

on whether the l.p. crank is in advance (w + 90) or
behind (w — 90) the h.p. crank. The value of the horizontal components
is given by the equation —

w-
p^ = /Bj — sin w for the h.p. cylinder.

iiaiiiiiii I

m

r

w

2

p.2 = ».> — sin (w ± 90) for the l.p. cylinder.

r

The vertical components of the moving parts tend to raise and lower
the framing of the engine with a force Pi + p.^, the values of which are
shown in Fig. 44 for every angle of the crank w. At the same time the
forces Pi and p.> tend to twist the frame in the vertical plane. The
horizontal components tend to move the framing in a horizontal
direction at right angles to the longitudinal axis of the engine, with
a force P^ + Po* ^^^ so to twist it at the same time in the horizontal
plane. -

84 MARINE ENGINES AND BOILERS.

The horizontal coniponcnts can be wholly overcome by balance
weights on the cranks ab, and »., ; if balance weights equivalent to
«! + M, and BBj + M., be used, the vertical components disappear, but in
their place horizontal components appear, namely^

p'j = M, — sin (iB + 180), and p', = M^ — sin (u. + 180± 90).

Fig. «.

g 36. Balancing the Moving Parts of a Three-crank
Engine. — As a rule no attempt is made to balance the effect of the
reciprocating parts of a three-crank engine, as the components of the
moving parts cannot be eliminated without introducing practical
difficulties.* In small engines the effect of the rotation can be
balanced by corresponding counterweights, as aheady explained, but
this method is not without drawbacks.

In ordinary three-crank engines there still remains a. free vertical
force, as well as vertical and horizontal moments to be dealt with. As
the axis (about which the components of the moving parts tend to turn
or tilt the frame) changes its position every instant, the moments above
mentioned are usually transferred to the plane of one of the outer
cranks (see Reciprocation in a four-crank engine).

If a and i are distances between the cyhnders (Fig. 45) a and /3, the
angles between the cranks (Fig. 4(i), then the vertical force will be —

P = i'i + Pi-fP3 = (Mi + »,) — cos<u-t-(Ma + »^) — cos (a + w)-!-
{Mg + »j) — COs(a + ^ + u)
* Compare \'arrow's suggesiioii for employing so-called " Ixib-weighls." £'ip'

THE MAIN ENGINES.

85

and the horizontal force will be —

w2

w2

p = Pi + p2 + p3 = »i — sinco + »2 — sin(a + a))-f

w^

A3 — sin (a + P + io).
The vertical components of the moving parts, reduced to plane i

w2

= a (Mo + Ag) — cos (a + w) +
(a + 3) (M3 + Aj) — cos (a + )3 + to).

The moment of the horizontal component
of the moving parts, reduced to plane i

= a »2 — sin (a + o>) +

I

ZXID

3

w

2

(a + ^) »8 — sin (a + )8 + o>).

•*•

(g

UI

Fig. 46.

These formulae are for " connecting rods of
infinite length," but for connecting rods of
finite length cos co-f X cos 2a), &c., must be
substituted for cos w. In ordinary three-crank
engines the static moments are very consider-
able, and if there are no perceptible vibrations
in many of these engines it is because the
engines are well placed in the ship,* or because

the vibrations of the ship do not synchronise with those of the engine
(see page 104), or because the masses of the ship's hull are very large
in proportion to the moments acting on them.

To secure the balance of the moving parts (apart from con-
siderations of economy of space, weight, &c.), the following is a good
arrangement of a three-crank engine : —

1. The L.P. cylinder is placed in the centre, between the h.p. and
the M.p. cylinders.

2. The crank angles do not exceed 1 20''.

3. The cylinders are placed as near together as possible, the valves
being placed off the centre-line of the engine.

^ 39. Balancing the Moving Parts of a Four-crank Engine :

System, t — Complete balance can be obtained in four-

* See 2^itschrift des Vereines Detitscker Ingtinieure^ 1884, p. 1091. O. Schlick.
f See Schlick, Transactions of the Inst, of Naval Architects, 1893, 1894; also
Loreiu, Die Dynamik der Kurbelgetriebe,

/

86

MARINE ENGINES AND BOILERS.

crank engines, on the assumption of infinitely long connecting rods.
The columns and bed-plate can be cast in the form of a stiff girder, in
order to take up the vertical components p in the vertical plane, and
the horizontal components p in the horizontal plane. The framing
remains at rest if, under the action of the above forces, it is in
equilibrium at each instant during a revolution of the engine. This
happens, if at each instant —

1. The sum of the vertical components p, and of the horizontal
components p = 0 ; and

2. The sum of the turning moments of the components p and p,
referred to any plane at right angles to the axis of the shaft, = 0.
These moments are generally referred to one of the planes of either of
the outermost cranks.*

Equilibrium of the engine framing can also be obtained, when

the sum of the turning moments

f^ «r about two different planes = 0 : — if,

j I for instance —

{

T

X-J.

Y\g. 47.

1. The sum of the turning moments
of p and P referred to the plane of
the outer right-hand crank = 0, and

2. The sum of the turning moments
of p and P about a plane through the

outer left-hand crank = 0. These two last conditions are generally used
to determine the balance of the moving parts. If we call

a, ^, c (Figs. 47 and 48), the distances between _

the four cranks, that is, the distance between the
centres of the cylinders, and a, )8, y the angles
between the four cranks—

Mj, M2, M3, M^t the masses affecting the vertical
components of the moving parts p ; and

»i, »2, »3, »4t the masses affecting the hori-
zontal components of the moving parts P

Fig. 48.

• It is a question of the statical moments of the components of the moving parts,
relative to a given plane. In future, for brevity sake, instead of writing ** moment
relative to plane, &c.," the usual expression, "moment about plane, &c.," will be
used, although of course a force cannot produce rotation about a plane, but only
about an axis.

t The symbols Mj, M2, &c., used here have the same significance as the corre-
sponding letters in § 37, 38.

THE MAIN ENGINES.

87

Then the two equations for the vertical moments of the reciprocating
parts may l>e written —

(1) Mg cos (c«> + a + /3 + y)xa + M8 COS <ox(tf + ^) +

M4 COS ((u + a+/i^) X (dr + ^4-r)«0.

(2) Mg COS Ctfxr+Mg COS (c«>4-a + /3-»-y)x (^ + ^) +

Mj COS (ctf + a) X (a + ^ + r) = 0.

The equations for the horizontal moments of the reciprocating parts

are-
(3)

Ao sin (a> + a + iS + y)xdr + iBg sin o) (a + ^) +
«4 sin (a) + a-|-/i^)x(tf + ^ + ^) = 0.

Ag sin a)x^+«2 sin ((^> + a + )S + y)x(^4•r) +
AJ sin (cu + a) x {a-\-b-\-c) = 0.

These equations must be determined for every value of w between 0*
and 360\

A

(a^lj'N^

The common factor — disappears, showing that in a balanced engine

the balance is satisfactory for any given crank-pin velocity, and therefore
for any number of revolutions, whether running ahead or astern.

It must be specially noted that, if the above conditions of equilibrium
are fulfilled, they will only produce a balance, and prevent vibration in
the ship, when the columns and bed-plate of the engine are cast in
the form of a stiff girder. In order to make the balance perfect, it is
therefore desirable to connect the upper ends of the columns or the
cylinders firmly with each other in a longitudinal direction. Light
engines should also be braced
diagonally in the same direction.
If the columns and bed-plate are
not sufficiently stiff, they may
bend under the weight of the
cylinders, and cause vibration in
the hull.

In calculating the reciproca-
tion of the moving parts of a given
engine, the conditions necessary
to secure a good balance apply
equally to the rods of the valve
gear. In a four-crank engine fitted
with Stephenson's link motion

there are 4 + 8 = 12 cranks (four cranks and eight eccentrics) which
must be balanced. Generally, however, the valve gear is combined into

Figs. 49 and 50.

88 MARINE ENGINES AND BOILERS.

one resultant factor or figure (see below). This resultant must be fur-
ther reduced to the crank circle, before combining it with the cranks
(see page 62).

!$ 40. To Determine the Balance of the Moving Parts

in actual practice, graphic methods, based upon the four above-named
equations, are employed. If the values of am.,, {a + ^)m3, and (« + ^ + ^)m^
are plotted to any scale from the centre along the cranks (compare Figs.
48, 49), the sum of the vertical components must, according to Equa-
tion 1, § 39, give u) = 0. It may be proved that such is the case when
the polygon formed by the projected moments of aMo, {a + b)}A^ and
{a-\-b-\'c)yL^ on their respective cranks, is closed (see Figs. 49 and 50).
The line {a-^b-\'C)M^ must end at the final point 0. If the conditions
obtaining in the second equation, § 39, are fulfilled, the polygon made
up of the values ^Mj, {b + r)M.,, and (a + ^ + c)yi^ will be closed. Precisely
the same holds for the horizontal balancing of the moving parts. These
may be worked out for every angle of rotation w by drawing a closed
polygon with the values of iCo^, ab^{a-^b\ stJi^a^b-^-c), and another
with the values of A3 . r, tt^ib + c\ and ab^{a -f <^ + r), regard being had to
the angle of the crank.

A complete vertical and horizontal balance of the moving parts is
obtained if all the four polygons, drawn as above, are closed. These
polygons are used to determine graphically the positions of the crank,
and the weights required to balance the moving parts. The values
tf, b, and Cy from centre to centre of cylinders, the approximate weight
of the connecting rods of the two middle cranks, and the angle between
them, are generally determined by structural conditions.

For balance in the vertical direction we then get —

1. Position and weight u^ of crank i, from the polygon reduced to
the plane of crank iv, as given by Equation 2, § 39.

-. Position and weight m^ of crank iv, from the polygon reduced
to the plane of crank i, as given in Equation 1, S ^^^
For the horizontal balance in the same way we get —

3. Position and weight /©j of crank i, from the polygon as given by
Equation 4, § 39.

^4. Position and weight ct^ of crank iv, from the polygon as given by
Equation 3, § 39.

To obtain vertical and horizontal bakincing, the positions and
weights of the crank must be made to agree (see Example, page 92).

§ 41. Study of the Valve Gear. — As we have already said,
the moments due to the moving parts of the valve gear are

THE MAIN ENGINES.

89

Fig. 51.

reduced to one moment about plane iv, and to one about plane i.
Let OA, OB, oc, &c. (Fig. 51), be the moments due to the aforesaid
pressure of the valve gear relative to plane iv, plotted respec-
tively in the direction of their eccentrics. They may be combined
into one polygon of moments as in Fig. 52.
Then the line OiH will be the resultant
moment of the eight moments oa, ob, &c.
In the same way the moments of the pres-
sure due to the masses of the moving parts
of the valve gear may be combined into a
polygon relative to plane i. The length of
the line closing the polygon gives the value
of the resultant moment, and the direction
of this line gives the position of the ideal
or hypothetical crank.

To determine the position and mass of
crank i for the vertical balancing of the
moving i>arts with reference to the valve gear^
the lines M2(^ + ^), m/, and o^h, starting
from c. Fig. 53, must be successively plotted
to form a polygon.* For determinating o^h
the moment of the vertical pressure due to
the valve-gear masses alone must be used.
The closing line ch, produced in the direc-
tion of the arrow, gives the moment of
pressure due to the masses of crank i, and
its direction. To determine the mass of
crank i for the vertical balance, divide the
moment ch by the distances « + ^ + r. The
mass and direction of crank iv for the
vertical balance, and of cranks i and iv for
the horizontal balance, are obtained in the
same way. The masses and directions of
the cranks differ somewhat for the vertical
and horizontal balances, and a mean of the
two should therefore be taken (compare
Example II., page 100). The vertical

balance is often used alone, without combining it with the hori-
zontal, as it is of more importance in preventing vibration than the
latter.

§ 42. Remarks. — l. If the valve gear is outside one of the end

* The direction of the line OiH should be the same in Figs. 52 and 53.

90 MARINE ENGINES AND BOILERS.

cranks, the moment about this crank due to the pressure of its masses
must be plotted in the opposite direction in the polygon, as compared
with the moments of the valve gear when inside the cranks.

2. If some of the valves are worked by a double-armed lever, the
masses on the further side of the fulcrum must be treated as negative.

/ / /

! I

I i

t

N <1L

i i I r M IT

V

i i i. ^ .i, ^ .i.

Fig. 54. Fig. 55.

I
J

3. If parts of the valve gear are worked by an eccentric in different
lateral planes — say e and s, Fig. 54 — the calculation of the moment
of the masses relative to plane i must be made by multiplying the
moments in plane e by ^, those in plane s by s. Also, to determine
the direction of the two moments in the polygon, the direction of the two
actuating eccentrics must be taken into account.

4. When a slide valve is fitted with an auxiliary relief piston, the
weight of both must naturally be taken into account in the calculation.

5. If the slide-valve rods or the pump rods are not on the centre
line of the valve gear, a special vertical balance of the moving parts
about this centre line ought, strictly speaking, to be drawn out for
them. This is not usually done, because these masses are always
small, and the rods are treated as though they were on the centre line
of the valve gear.

§ 43. Most favourable Arrangement in an Engine to
secure Perfect Balance of the Moving Parts.— 1 he equations

in § 39 show that balancing is most effectually carried out when —

1. The cylinders with the heaviest connecting rods are in the
centre.

2. The distance between the two middle cylinders is as large as
possible, as compared with the distances between the two outside
cylinders —

Q=l-8to2)

If these conditions cannot be secured in practice, as sometimes
happens, because the advantages of some other arrangement cannot be

THE MAIN ENGINES.

91

sacrificed merely to secure more effective balancing, the outer connect-
ing rods must be artificially lightened by means of countenveights, and
the connecting rods of the middle cylinders artificially weighted, by the

L 4
use of heavy pistons. Besides this, in engines having a ratio --<^ the

angles of the two outer cranks are apt to be rather small. This has an
injurious effect on the manoeuvring of the engines, and the regularity of
the turning moment (see pages 125 and 63). But if the heavy rods be in

the middle and -= about 2, the best crank angles will be obtained,

the smallest being about 65** to 70**, and perfect balance can be secured
without adding extra weights (see following example). In triple expan-

lA

MP

28 t-97

HP -LA.

f8z -St 3f raz

MwAs

H

r

fsr-z

^f^

JJ

As\ JihAa \Ah

AsaM.

^

1 ^-^r/^ ■!, ^2rfti

E

M

I

Z^r^ ,|

JT

Fig. 56.

hLP

miA

sion engines good balancing is most easily obtained by having two l,p.
cylinders, each doing purposely less work, and having lighter connecting
rods than the h.p. and m.p. cylinders,
outside which they are placed (Yarrow-
Schlick-T weedy system ). An engine with
four cranks, in which the forward pair
of cranks is at right angles to the after
pair, and the cranks of each pair are at
an angle of 180° to each other, gives
an almost perfect balance, if the dis-
tance between the two pairs of cranks
is relatively large in proportion to the
distance between the individual cranks
constituting each pair, and when the
masses of the forward pair are equal
to the masses of the after pair. In

engines running at high speeds, the balance is considerably affected by
the relatively heavy parts of the valve gear, and their long stroke. A

'a-'KP

i-HP

Fig. 57.

92

MARINE ENGINES AND BOILERS.

better balance is obtained in such engines, by placing the valve gear
outside the end cranks, and carefully proportioning the steam and
exhaust ports of the slide valve.

§ 44. First Example : Calculation of the Balance of the
Moving Parts in an Existing Engine. — This engine was built on

the Yarrow-Schlick-Tweedy system, the balancing of the moving parts
being specially considered. The distance between the two middle
cylinders is relatively great. The connecting rods of the l.p. cylinders
are specially light, in view of the less amount of work to be done in
these cylinders. Fig. 56 shows the arrangement of the engine. Fig. 57
the positions of the cranks, Fig. 58 positions of the eccentrics.

B^HP

sr-NP

I -LP,

incAi»4

\Ah As\

Fig. 58.

Weights of the Conmcting Rods,

Details.

Reciprocating Parts —
Piston

Piston rod and crosshead -
Connecting rod (top end)
Pump rods

Total

Rotating Parts —
Connecting rod (bottom end)
Crank pin and checks » -

Total

'iv=No. 2

L.p.

Cylinder.

lb.

588-47

427 07

184-47

-50-25

1150-26

319-58
780-21

1099-79

Total weight of the reciprocating and
rotating parts - - - .

225005

III = M.P.

Cylinder.

lb.
606-10
606-10
295-33

II = H.P. I

Cylinder.

lb.
312-96
606-10
295-33

1507-53

418-76

1214-39

418-76

1055-71 1055-71

i = No. 1

I..F.

Cylinder.

lb. '

588-47
427-57
184-47

1200-51

319-58
780-21

1474-47

2982-00

1474-47 ! 1099-79

2688-86

2300-30

THE MAIN ENGINES.

93

•5P

i = No. 1 L.P.
Cylinder.

Astern.

lb.

35-04
55-10

90-14

140-39

2512

127-40

f-H

q>

CM

Oi
CM

p

00
CO

r-H -^ CM O

'3 r-H O O f-H
n *****

5 .a 00 o ^ o

■5 ^ Oi CO CM »o

612-27

147-68

2512

137-52

CM

CO

•

o

f-H

CO

Ip

CM
CM

Cylinder.

Astern.

CM CM

CM CO
* . . « •

Xi . '^ "T-H

^ - lO -00

lb

CO

f-H

191-75

38-13

172-57

cf^l

o

m

CO

a;

.
to.
im

II

Ahead.

r-t C<1 «0 CM
IC CM T-H CO

a • • • •

-Q t- '^ ** T-H

"^ 1:* »0 CO 00
CM

447-41

198-80

38-13

184-69

421-62

869-03

Cylinder.

Astern.

CM CM

CM CO
. • • . •

^ ' lO -00

135-54

191-75

38-13

172-57

CM

O

537-99

•

ft.

■

;i

1

<

00 CM ^ CM
t^ <M i-H CO

■ ■ ■ • •

.a It* '^ -rj^ r-H

•^ CO lO CO 00

00

•

O

198-80

38-13

184-69

«M
CD

•

r-H
CM
**

o

r-H

a

CM

T-H

l-H

•

itu

>

m

C

Im

lb.

35-04
55-10

90-14

140-39

2512

127-40

292-91
383-05

Ahead.

f-H "TtH CM O t^
^ f-H p p r-H C^l

.a oD >b 4hi »h CM

"" 0> CO CM ITS f-H

147-68

25-12

137-52

CM

CO

•

o

r-H
CO

cfi

CM

Details.

Reciprocating Farts —

Slide valve, valve rod, and link block
Half of reversing link
2 X J reversing rod - - - -
Eccentric rod (top end)

Total

Rotating Farts —

Eccentric strap

Eccentric rod (bottom end)
Eccentric

Total

Total weight of the rotating and recipro-
cating parts

94

MARINE ENGINES AND BOILERS.

Gi

to

>

<
>

O
H .

hi 2

o

<

<
PQ W

<

O

N

t^

O

u

<

(3A

THE MAIN ENGINES.

95

Vertical Balance— Movement of Valve Gear about

Plane IV.

96

MARINE ENGINES AND BOILERS.

Horizontal Balance — Movement of Valve Gear about

Plane IV.

/

Fig. 62.

In the following calculations the actual weights are used, instead of

w^ 1
the masses of the moving parts, because the factor — - is common

to all the latter (see page 87). The weights of the moving parts of
the valve gear are first reduced to the circle described by the eccentric,
• and the resultant reduced to the circle described by the crank.

We will work out first the resultant moment of the moving parts of
the valve gear, as follows : —

1. For the vertical balance about plane i.

2. For the vertical balance about plane iv.

3. For the horizontal balance about plane i.

4. For the horizontal balance about plane iv.

THE MAIN ExXGINES.

17

Moments of the Moving Parts of the Valve Gear about Plane i.

Vertical Balance.

lb. ft. ft. lb.
922-59 X ( - 2-25) = - 2075 8
383-05x(-l-97)= - 759-4

86903x4-8 = 4173-4
537-99x511 -^ 2755-7

Horizontal Balance.

1 Ahead
Astern

lb. ft.
310-32x(-2-25):
292-91 x(- 1-97) =

ft. lb.
= -697-98
= -578-64

II Ahead
Astern

421 -62 x 4-8
402-45 x511

421-62x6-08
402-45 X 6-39

=2025-24
=2061-40

=2567-71
=2574-94

HI Ahead
Astern

112910x608 = 6872-3
537-99x6-39 = 34429

922-59x13-42 =124118
383-05x1316 = 50486 '

IV Ahead
Astern

310-32x13-42
292-91 X 1316

= 4173-44
=3862-42

Moments of the Moving Parts of the Valve Gear about Plane iv.

1

Verttcai Balance

Horizontal Balance.

1

I Ahead
1 Astern

lb. ft.
922-59 X 13-42 =
38305 X 1316 =

ft. lb.
12411-8
5048-6

lb. ft.
310-32x13-42
292-91 X 1316

ft. lb.
= 4173-44
= 3862-42

II Ahead
1 Astern

869-a3x6-38
537-99x6-07

5547-7
3260-3

421-62x6-38
402-65x607

=2690-67
-2444-75

1

1 III Ahead
Astern

1 12910 X 510
537-99x4-79

5750-2
2574-9

421-62x5-10
402-65x4-79

= 2148-20

= 1923-97

1

IV Ahead
Astern

1

922-59x( -2-25)= -
383-05x(- 1-97)=-

-2075-8
- 759-4

310-32x(-2-25)
292-91 x(- 1-97)

-: 697-98
= 578-64

These moments are combined in the polygons shown in Figs. 59 to '
62, r^ard being had to the positions of the eccentrics. The lines o iv r

are the resultant moments of each polygon, />., the moment o iv r re-
presents the resultant action of the different moments of which it is
composed.

We are now able to plot the polygon of moments of the cranks and
valve gear about planes i and iv, to arrive at the horizontal and
vertical balance. If the lines thus produced form a closed figure,
the balance of the parts is complete.

G

98

MARINE ENGINES AND BOILERS.

Moments of the Connecting Rods about Plane i.

1

Crank

Vertical Balance.

, II

1

lb. ft. ft. lb.
2688-8 X 2-97= 802139

1
III

2982-0 X 8-20 = 24462-0

IV

2250-2x11-17 = 25170-8 ;

Horizontal Balance.

lb. ft. ft. lb.

1474-4 X 2-97= 43776
1474-4 X 8-20=12115-2
1099-8x11-17 = 123250

Moments of the Connecting Rods about Plane iv.

Crank

I

II
III

Vertical Balance.

lb. ft. ft. lb.
2300-9 X 11-17 = 25749-4

2688-8 X 8-20 = 2-2075-1

2982-0 X 2-97= 88893

Horizontal Balance.

lb. ft. ft. lb.

1099-7x11-17 = 123-25-0
1474-4 X 8-20=12115-2
1474-4 X 2-97= 4397-6

1. Vertical Balance: Moments reduced to Plane i. — Fig. 63 shows
the polygon combining the moments of the connecting rod for vertical
reciprocation, as reduced to plane i, with the corresponding sum of the
moments o iv r of the valve gear, o iv r = 8,509 ft. lb. (reduced to the
radius of the eccentric) must be reduced by the ratio of the radius of
the eccentric (3"24 inches) to that of the crank (14-99 inches) in order
to combine it with the moments of the crank. The moment of the

valve gear, thus reduced, is iv 0=1,844 ft. lb. Fig. 63 shows that the
polygon of moments o ii in iv o forms a closed figure, and therefore the
moment of the vertical balance of the moving parts, reduced to plane i,
disappears.

2. Vertical Balance : Moments reduced to Plane iv (Fig. 64). — o iv r =
21,337 ft. lb., moment of the valve gear, reduced to the radius of the

eccentrics, mo = 4,629 ft. lb., is the same, reduced to the circle de-
scribed by the crank. As the polygon o i ii iii o also forms a closed
figure, the vertical balance is therefore complete.

3. Horizontal Balance: Moments reduced to Plane i (Fig. 65). —

k ^ - . '- .

I '

Vertical Balance
Moments reduced
to plane I.

I

Fig. 63.

Vertical Balance

Moments Reduced

lb plant IV.
I

2/537 Ft lbs.
Fig. W.

Horizontal Balance
Moments reduced

Fig. 65.

Horizontal Balance
Moments reduced
to flane IV.

j Sa56Ftlhs.
Fig. 66.

100

MARINE ENGINES AND BOILERS.

o IV R = 3,884 ft. lb., moment of the valve gear, reduced to the radius of

the eccentric, iv r = 839 ft. lb., is the same, reduced to the radius of
the crank. Here the polygon of moments o iiiii iv r is not closed ;
there is thus an incomplete balance of the moving parts = or = 578
ft. lb. (reduced to the crank circle).

4. Horizontal Balance: Moments reduced to Plane iv (Fig. 66). —

o IV R = 9,836 ft. lb., moment of the valve gear, reduced to the circle
described by the eccentric, in r = 2,133 ft. lb., is the same, reduced to
the crank circle. Here also the polygon o i ii in r is not a closed figure,
and the balance is incomplete by or = 1,519 ft. lb. The engine has thus
perfect vertical and almost perfect horizontal balance. The small dis-
crepancies in the latter may be neglected. The balance is obtained
without counterweights, and without weighting the piston.

§ 45. Second Example : Calculation of the Balance of the
Moving Parts in the Engine of a Fast Mail Steamer.— The

arrangement of the cylinders and
/ // EI 17 valve gear are given, and for con-

U\P f^P t\p / \d structive reasons must be adhered

Hr rfr Lr. ^A ^^^ although it does not give the

best balance. It is usually stipu-
lated that the different parts of the
crank shaft shall be interchangeable.
The holes for the coupling bolts
must therefore be bored in such a
way that the angles of the crank are

multiples of the angles of the bolt holes. For the arrangement of the

cylinders see Fig. 67.

r
I

L.

15 2' I 98'

lii

5Z'8' '■

Fig. 67.

^

Weights and Moments of the Crank Rods, as determined by

Constructive Conditions,

Weights for calculating the vertical bn lance -
Weights for calculating the horizontal balance
Moments for calculating the vertical balance

about plane i

Moments for calculating the vertical balance

about plane iv

Moments for calculating the horizontal balance

about plane I

Moments for calculating the horizontal balance

about plane iv

H.P.

MP.

No. 1

I..V.

No 2
L.r.

tons

18-8

20-7

20-7

20-7

>»

110

no

11-0

11-0

ft. tons

• • •

203

455

• • •

»)

• • ■

455

2a3

• • •

n

• • «

108-4

253

• « »

>»

• • •

253

108-4

... '

THE MAIN ENGINES.

101

hHP

jr.iPt

Fig. 68.

A cursory examination shows that in this case the crank positions
given in Fig. 68 are the best, and they also satisfy the condition for the
interchangeability of the coupling bolts, if fourteen bolts are put into each
flange. The positions of the crank
being thus fixed beforehand, the resul- M^LP

tant moment of the valve gear can be
determined. This is done in exactly
the same way as in Example I., and
only the results obtained are given
here ; the moments of the valve gear JWff
about planes i and iv for the vertical
balance Si"" and s/ are drawn out in
Fig. 69, and for the horizontal balance
Sj** and s^ of the valve gear in Fig. 70.
As already observed, this arrangement

of the cylinders is not favourable to good balancing, so it is necessary
to increase the weight of the centre pistons, and to fit counterweights
on the outer cranks. The amount of additional weight required is
deduced from Figs. 69 and 70.

1. Horizontal Balance: Determination of the Counterweights (Fig.
70). — The moment of crank i^ must annul those of cranks ii** and iii^
about plane iv. o \^ is thus = the resultant of the moments o \\^ and

0 lIIl^ But crank ^ must further annul the moments of the valve gear
about plane iv. Therefore to the resultant o \^ we must add the resulting
moments of the valve gear for the horizontal balance in plane iv, namely,

1 J** I2'* = 0Sl^ If balance were complete, o io'' = 226 foot tons would
be the moment of the masses of crank i actually rotating about plane iv.
For constructive reasons (compare the Table of Weights) the moment of
crank i** about plane iv has to be o 1** = 361*4 foot tons. Thus there is a
discrepancy, i.^** i**, in the balance of the parts. This may be eliminated
by a counterweight on crank i, the dimensions and direction of which
are i*" \^, The moment of this counterweight about plane iv is og^ =
I*" i2** = 132*3 foot tons, and the weight of the counterweight is therefore

132*3

—^ = 4*0 tons. The dimensions and direction of the counterweight

on crank iv are found in the same way.

Moment G4 = iv** iv2*' = 83*8 foot tons, weight 04=2*56 tons.

The addition of these weights makes the horizontal balance of the
parts complete.

2. Vertical Balance : Determination of the Weight required on the
Piston (Fig. 69). — oii'=the resultant of the vertical moments of cranks
II and III about plane iv (pi\{ and o iiii'); o iv^'' also = the resultant

102

MARINE ENGINES AND BOILERS.

jo^V=206 II I'

Fig. 69.

JT

^Z .Ot r s 206 -

Fig. 70.

THE MAIN ENGINES. 103

of the vertical moments of cranks ii and iii about plane i (o 11/ and
o 11I4''). The moments of the valve gear -about plane iv and plane i
respectively must compensate for cranks i and iv. These moments
are oSi"" and 084% and are added to i^'' and iVj*, so that os^'' = i^'' ig""
and os/ = iVi' iv/. The moments of the cranks or and oiV' are
determined by constructive conditions; the counterweights required
for horizontal balance must be deduced from them. If they are
added to r and iv* respectively, we get the points I5'' and iv^". If
this gave the required vertical balance, o ig"" and o ig^ and o iVg* and
01V5'' would be respectively equal to each other. This is not the
case; and since the counterweights cannot be altered, because that
would affect the horizontal balance, the two centre pistons must be
weighted. For piston iii the weight must be such that its moments
about plane iv = ig* ij^ and about plane i = iv./ iv^'' ; for piston 11 the
moments of the weight about planes iv and i must be = 13'' i^'' and
IV,* and IV/ respectively.

With the pistons thus loaded there are still discrepancies in the
vertical balance, but they are not important ; they consist of the result-
ing moments i^* i^* and iv^' and iVg*, which cannot be eliminated by
weighting the pistons, because if the load on piston in is increased the
moment iv^'' iv^'' is reduced, but the moment 1/ I5* is augmented. This
very small discrepancy might be completely avoided if the angles of the
cranks and the counterweights could be slightly varied.

The weight on piston iii is —

C V - 1 V 'V - 290 _ 67-7 _ ., Qj. .
9-8 22^9 9^ " 22-9 " " ^'' *''"^-

The weight on piston 11 is —

W »ViV 90-3 38-7 «o. ,
2V9- -W^ =22^9 = -9:8 ='^^^^^"^-

§ 46. Effect of the Length of the Connecting: Rod on the
Balance of the Moving Parts. — The above deductions for
the vertical balance are valid for "connecting rods of infinite
length." In practice, for connecting rods of a given length, the vertical

M W^

balance of the moving parts does not follow the law cos w, but

approximates to the law (cos a> + A. cos 2(y). See page 62.

If in the sum of the moments for vertical balance (Equations 1 and 2,
§ 39) we substitute cos w + X cos 2a», cos (a -i- oj) + A cos 2(a + w), &c., for

104 MARINE ENGINES AND BOILERS.

COS o>, cos (a + o>), &c., these totals will not be = 0,* even in engines with
complete balance, and connecting rods of " infinite length."

These totals may then be written thus : (1) Moments about plane i —

MoOL cos (w + a + /J + y) + M^(a + d) COS to + lij^a + ^ + r) cos (w + a + )8) +

M.ja . A. cos 2(ftf + a + j8 + y) + hi^{a + lf)k cos 2<u -i-

u^{a + ^ + r)A cos 2(<o + a + /^)

and (2) moments about plane iv —

MgT cos (0 + M./^ + C) cos (w + a + /J + y) + Mi(a + ^ + r) COS (oi + a) -J-

Mjj^rX COS 2cu + M.,(/^ + c)X COS 2(w + a + )8 + 7) +
Mi(a + ^ + ^)X COS 2(<o + a).

In every engine where the vertical balance of the moving parts is
complete, and with connecting rods of " infinite length," the three first
terms of each of these equations disappear, and the three last terms
may be combined into one resultant moment. These moments attain a
maximum and a minimum twice in every revolution, and therefore they
can only produce vibrations, the number of which is equal to twice the
number of revolutions of the engine.

No vibrations to any considerable extent, produced by these mo-
ments, have been detected in vessels, although the existence of the
vibrations has often been proved by means of the " Pallograph." (See
Part VI.) They are caused by the length of the connecting rod, and
to reduce them within small limits a special modification of the balanced
parts has been introduced by Schlick ; but want of space forbids a
description of it.

§ 47. Critical Number of Revolutions and Efifect of the
Position of the Engine on the Vibrations of the HuU.t—

The ship's hull may be considered as an elastic beam or girder
supported throughout its whole length upon an elastic medium. Such

a beam will "vibrate" to its greatest

A C -D extent, if it does so at all, in the manner

r^-'^'^'^ ^---^^^^ shown at Fig. 71, the curves of vibration

* "^ ^^ forming nodes at the two points a and

Fig. 71. B. The engine produces the greatest

vibratory effect when the number' of its
revolutions (and therefore of vibratory impulses) coincides with the
natural vibrations of the ship. The number of revolutions which

* In four-crank engines only one of these totals can be = 0 with connecting rods
of a given length.

t See Schlick, Zeitschrift des Vcreines Dcutscher Ingenieure^ 1894, vol. ii., p. 1091,

THE MAIN ENGINES. 105

produces the maximum vibratory effect of the first order is called the
**m//V»/" number of revolutionsy and is approximately* —

In this formula t signifies the moment of inertia of the mid-ship section,
to get which the cross-sections must be given in square feet, and the
arm in feet ; d the displacement, in tons ; l the length of the ship
on the water line, in feet ; and >& is a coefficient having the following
values —

k = 34050 for torpedo-boat destroyers ;
= 31200 for large, fast steamers with fine lines; and
= 27800 for " full " cargo vessels.

Position of the engine in the ship. An engine with free moment of
the moving parts would tend to accentuate vibrations of the nature shown
in Fig. 71, when it is at the points a or b, and would not materially
influence them when it is in the centre at c.

An engine with free resultant of the balance of the moving parts
7vould tend to increase the vibrations shown in Fig. 71, when it is in the
centre at c, and has no similar tendency when it is at the points a or b.

* See Schlick, IVam. of the Inst, of Naval Architects , 1894, p. 350.

SECTION IV.

ARRANGEMENT OF MAIN ENGINES.

§ 48. Arrangement of the Cylinders and Cranks.—!. Twin-
cylinder engines (sometimes used instead of compound engines for small
and very light boats, such as pinnaces) have two cylinders, either cast
in one or else solidly bolted together, and two cranks at an angle of 90**.

2. Compound engines (for small freight steamers and passenger
steamers for river traffic, &c., with engines from 30 to 500 i.h.p., or
even up to 1,500 i.h.p.). These are now only made with one h.p. and
one L.P. cylinder, and with two cranks at an angle of 90*. Formerly
in compound engines with more than two cylinders several different
arrangements were adopted, and they were often used for higher powers
with one h.p. and two l.p. cylinders, and three cranks at an angle of
120°. (Fast steamers " Ems " and " Elbe " with engines of 6,000 i.h.p.)

Figs. 72, 73, 74, 75 give four views of the engine of a ship's pinnace.
I.H.P. about 50; number of revolutions per minute = 300. Diameter of h.p.
cylinder, 160 mm. (6*29 inches); l.p. cylinder, 300 mm. (11*81 inches);
stroke, 200 mm. (7*87 inches); boiler pressure, 142 lb. per square inch.
The reversing gear is worked by a hand lever, altering a single adjust-
able eccentric for both cylinders. Air, bilge, and feed pumps are also
driven by one eccentric.

Figs. 76, 77, and 78 show a rather heavier engine of a small freight
steamer, i.h.p. about 200; revs. = 130. Diameter of h.p. cylinder, 380
mm. (14-96 inches) ; l.p. cylinder, 680 mm. (26*77 inches) ; stroke, 500
mm. (19-68 inches); boiler pressure, 114 lb. per square inch. Stephen-
son's link motion, the reversing link being moved by a screwed spindle
and hand wheel. Air, bilge, and feed pumps are driven direct from the
L.P. crosshead by means of a lever.

3. Triple-expansion engines are now almost universally used for all
kinds of marine engines developing more than 300 i.h.p. The types
most generally employed are —

{a.) Three-crank engines y with one h.p., one m.p., and one l.p.
cylinder^ the cranks being at 120". This arrangement is employed on
all kinds of cargo steamers, passenger boats, warships, and torpedo-
boats, but it cannot be used for ver)' high powers, because the dimen-

THE MAIN ENGINES. 107

sions of the l.p. cylinder become too large and unwieldy. The risks and
difficulties of machining and fitting such heavy castings are so great, that

Figs, "2 and 73.
the diameter of the l.p. cylinder is seldom more than from 8 feet 10
inches to 9 feet 2 inches.

MARIXE ENGINES AND BOILERS.

^1

7^

Plate II.

Fig. 78.

[To face page 108.

J

MAUIXE ENGINES AND BOILERS.

Plate II.

Fig. 78.

\To face page 108.

I

J

THE MAIN ENGINES.

109

Fig. 79.

110

MARINE ENGINES AND BOILERS.

Figs. '82 to 84 show the engine of a steam trmvler, i.h.p. about
400. Cylinder diameters, 305, 483, 819 mm. (12 inches, 19 inches,

32 J inches) respectively; stroke,
610 mm. (24 inches); revs. = 140;
boiler pressure, 164 lb. per square
inch. Stephenson's link motion
with crossed eccentric rods. Air,
circulating, bilge, and feed pumps
are worked by a lever from the m.p.
crosshead.

Figs. 79 to 81 show the engines
of a battleship with twin screws.
I.H.P. = 2 X 6,000; revs. = 100.

•

Diameter of cylinders, 1,016, 1,498,
2,235 mm. (40 inches, 59 inches,
88 inches); stroke, 1,295 mm. (51
inches); boiler pressure, 156 lb. per
square inch. Stephenson's link
motion ; condenser at the side of
the engine. Each cylinder is sup-
ported by four cast-steel columns.

Figs. 85 to 87. Engines of the
twin-screiv mail steamer " Furst Bis-
marck." I.H.P. = 2 X 8,200 ; revs. =
85. Diameters of cylinders, 1,100,
1,700, 2,700 mm. (3 feet 7J inches,
5 feet 7 inches, 8 feet 10 inches);
stroke, 1,600 mm. (5 feet 3 inches):
boiler pressure, 156 lb. per square
inch. Stephenson's link motion.
Each cylinder is supported on two
cast-iron A frames, one of which
rests on the cast-iron condenser.

Fig. 88, Plate III. Engines of
the Imperial yacht " Hohenzollem."
I.H.P. = 2 X 4,500. The reversing
gear is on the Klug system; con-
denser at the side of the engine.
Each cylinder is supported on two
cast-steel columns.
Fig. 80. (^.) Three-crank engines with five

cylinders^ namely, two h.p., one m.p.,
and tivo l.p. cylinders. Angle of the cranks, 120". The centre crank

THE MAIN ENGINES.

Ill

is driven by the m.p. cylinder ; each of the two outer cranks is driven by
one of the l.p. cylinders, upon the cover of one of which an h.p. cylinder
is placed tandem. This arrangement has been carried out in several
large mail steamers, such as the " Campania " and " Lucania," which
have two engines each indicating 14,000 h.p. ; also on the "Spree" and
the " Havel," each of which has a single engine indicating 12,500 h.p.

Fig. 81.

(c.) Four-crank engines with one h.p., one m.p., and two l.p. cylinders.
This arrangement gives long, narrow engines; the cranks are usually
set at an angle of 90*, or balanced in accordance with the Schlick
system. The advantages of this arrangement are many ; the division
of the L.P. cylinder, necessary in large engines, is obtained without having
to place one cylinder above another ; and in small war ships, where the
engines must not project above the armoured deck, large cylinders can
be avoided, and the two engines put side by side. When the Schlick

112

MARINE ENGINES AND BOILERS.

^

CO
1^

3?

THE MAIN ENGINES.

113

system of balancing is used, a good even turning moment is obtained; ''^
and lastly, the strain on the crank shaft is not so great, because the pres-
sure of the connecting rods is more evenly distributed over four, instead
of over three crank pins. This arrangement is used in torpedo-boats
and destroyers, warships of all sizes, mail and fast steamers.

Figs. 254, 255 (see page 252) show the engines of twin-screw destroyers
built by Thomycroft & Co. Diameter of cylinders, h.p. 22 inches,
M.p. 29 inches, l.p. 2 x 30 inches. Stroke, 18 inches. Boiler pressure
above atmosphere, 225 lb. per square inch. Number of revolutions,
390. i.H.p. of each engine = 3,000. The engines are mounted on light

Fig. 84.

columns, the bed-plate is also made as light as possible, so that practi-
cally the engine foundations are formed by the engine seatings and
framing of the ship.

Fig. 89, page 116, shows the engine of a destroyer with twin screws,
built by Thomycroft for the Japanese Navy. Each engine is of 3,600
I.H.P., and runs at 390 revolutions per minute. Cylinder diameters,
H.p. 22 inches, m.p. 30 inches, l.p.j 31 inches, l.p.j 31 inches. Stroke,
19 inches. Boiler pressure, 250 lb. per square inch. The forward pair
of cranks is at 90*" to the after pair, and the individual cranks of each
pair are opposite each other. Compare Balance, page 90.

* See Ix>renz, Dynamik dtr Kurbelgetriebe,

H

MARINE ENGINES AND KOILERS.

^ k

THE MAIN ENGINES.

115

Fig. 113, page 126, belongs to the same type of engine. Fig. 90
(see Plate IV.) gives a similar engine for a small armoured cruiser with
twin screws, i.h.p. 2x3,750, « = 178. Boiler pressure above atmo-
sphere, 210 lb. per square inch. The two low-pressure cylinders have
one valve chest between them. The cylinders are supported on steel
columns in front, and on cast-steel framing at the back. Stephenson's
link motion, separate condenser, and Schlick system of balancing.

Figs. 93, 94, 95, and the photo-
graphs Figs. 91 and 92, show one of
the two engines of the Russian crui-
ser " Bogatyr," which was completed
in the summer of 1902 at the Vulcan
Works, Stettin.* The engines are
built on the Schlick system, each of
them developing 10,000 i.h.p. at a
speed of 150 revolutions per minute.
Dimensions : — Diameter of h.p.
cylinder = 40 inches, m.p. cylinder ==
60 inches, of both l.p. cylinders 70
inches. Stroke, 36 inches. Admis-
sion pressure in h.p. cylinder, 240 lb.
per square inch. The columns and
bed-plate are of cast steel. The l.p.
cylinders have balanced slide valves.
(Compare Fig. 153, page 162.)
The reversing gear is an all-round
gear with two cylinders 6 J inches
diameter x 5 inches stroke; the turn-
ing gear has a single cylinder of the
same dimensions. Each engine has
an independent air pump of the Weir
type ; each of these has two steam
cylinders 12^ inches diameter, two
pump cylinders 37 inches diameter,
and stroke 15 inches. Each engine
has a circular condenser with a cool-
ing surface of 10,760 square feet,
and a circulating pump with two vane wheels of 47 inches external
diameter.

To a similar type of vessel belong the engines of the Japanese
armoured cruiser " Yakumo," shown in Fig. 96, Plate V. At about 140

Fig. 87.

Compare Zeitsckrift des Vereines Detttscher IngMeure^ 1902, and Engineering.

116 MARINE ENGINES AND BOILERS.

revolutions per minute, each of the two engines indicates 8,000 h.p.; they
are balanced on the Schlick system. The bed-plate is of cast sleel, and
each cylinder rests on four cast-steel ribbed columns. This arrange-
ment gives increased stiffness, but at the sacrifice of accessibility and
facility of inspection, as compared with the type described above.

Figs. 97 and 98, Plate VI., and Fig. 99, showihe engines of the twin-
screw fast steamer "Kaiser Wilhelm der Grosse," built at the Vulcan

Fig. 89.

Works, Stettin, 1896. i.h.P. 2 x 14,000,n = 78. Diameter of cylinders, 52
inches, 90 inches, and two of 96 inches. Stroke, 69 inches. Boiler pres-
sure above atmosphere, 210 lb, per square inch. The main engines are
balanced on the Schlick system. Each cylinder is supported on four
cast-steel ribbed columns ; the bed-plate is of cast iron. The reversing
gear is Stephenson's link motiori. The h.p. cylinder has one piston
valve, the m.p. cylinder two ; the l.p. cyhnders have flat slide valves.
Each of the engines has a condenser of sheet copper with a cooling sur-

THE MAIN ENGINES. 117

face of 17,750 square feet, and 5,530 tubes. Each of the two Blake
air pumps has two steam cylinders 18 inches diameter, and two pump
cylinders 42 inches diameter and 24 inches stroke. The circulating
pump for each condenser is driven by a compound engine with h.p.
cylinder 11 inches diameter, l.p. cylinder 20 inches diameter, and 12

inches stroke. To either end of the shaft of the circulating engine is con-
nected the vane wheel of a centrifugal pump, 4« inches external diameter
and '20 inches diamet(;r of suction, so that for the two main condenseis
there are in all four centrifug.'xl pumps. Each main engine is fitted
with a Brown's steam reversing gear, the steam cylinder of which is
26 inches diameter, the hydraulic cylinder 16 inches diameter, and

118 MARINE ENGINES AND BOILERS.

stroke 20 inches; also a two-cylinder turning engine, each cylinder
being S inches diameter and S inches stroke.

Particulars may also be given of the engines of the twin-screw fast

mail steamer, "Kaiserin Maria Theresa," built at the Vulcan Works,
Stettin, in 1S19S. Each engine develops about (<,000 i.h.p. at 92 revolu-
tions per minute. The cylindrical cast bronze condenser is separate

THE MAIN ENGINES.

120

MARINE ENGINES AND BOILERS.

from the main engine. The latter has a bed-plate of cast iron, and is
mounted on eight cast-iron columns of equal length. (See Fig. 100, Plate
VII.) The engines are balanced on the Schlick system, and although
they run at a high number of revolutions and high piston speed, there
is not the slightest vibration in the hull of the vessel.

(d.) Triple-expansion engines with two cranks and three cylinders are

now only used where local con-
ditions do not allow of three
cranks, as for instance when
compound engines have to be
tripled. The h.p. cylinder is
then generally placed above
the M.p. cylinder. The turning
moment of such engines is
always more irregular than that
of three-crank engines.

4. Quadruple - expansion
Engines. — These have very
varied arrangements of
cylinders.

{a.) The most usual method
is to place, the one h.p., the
first M.P., the second m.p., and
the one l.p. cylinder side by
side, with the four cranks at an
angle of 90**, corresponding to
the Schlick system of balanc-
ing. The larger cylinders are
generally in the middle. A
similar engine is shown in Fig.
101, Plate VIII. It is that of
a twin-screw Imperial Mail
Steamer, each engine develop-
ing 4,500 I. H.p. Columns and
bed-plate of cast iron ; Stephen-
son's link motion. Balanced on
the Schlick system. Thesecond
M.p. and the l.p. cylinders
are in the middle, the h.p. cylinder forward, and the first m.p. aft.

(A) With very large engines the l.p. cylinder must for reasons given
above be divided. The engine is then built as a three-crank engine
with five cylinders, or as a four-crank with six cylinders. An example
of the latter arrangement is shown in the engines of the " Deutsch-

Fig. 99.

I-

JI

-E-

I

THE MAIN ENGINES. 121

land " (Fig. 102, Plate IX.). i.h.p. = 2 x 17,500 ; revs. = 78. Diameter
of cylinders, 2 x 930 mm. (two of 3 feet 6 inches), one 1,870 mm.
(6 feet 1 J inches), one 2,640 mm. (8 feet 11 inches), and 2 x 2,700 mm.
(two of 8 feet 10 inches) ; stroke, 1,850 mm. (6 feetl inch) ; boiler pres-
sure, 214 lb. per square inch. Reversing gear, Stephenson's link motion.

Of the four cranks which are placed opposite each other, on the
Schlick system, the foremost is driven by No. 1 m.p. cylinder. Then
come the two l.p. cylinders, and No. 2 m.p. cylinder works the after-
most crank. An h.p. cylinder is mounted on the cover of each of the
L.P. cylinders.* Each cylinder is supported on four cast-steel columns
resting on a steel bed-plate. The two condensers are placed in the
wings of the ship, and each has 5,320 condenser tubes, with ^,500
square feet of cooling surface. The air pumps, of which there is one to
each condenser (and hence one to each engine), are Blake air pumps.
Each has two steam cylinders 18 inches diameter, two pump cylinders
44 inches diameter x 24 inches stroke. Each condenser has also two
circulating pumps, the vane wheels having an over-all diameter of 47
inches, and 30 inches diameter of suction on either side of the pump.
The pumps are coupled direct to a compound engine. Diameter of h.p.
cylinder of the latter 11 inches, of l.p. cylinder 20 inches x 12 inches
stroke. The engines driving the turning and reversing gear are of the
same dimensions as those of the " Kaiser Wilhelm der Grosse." The
engines of the fast steamer " Kronprinz Wilhelm " are constructed on
the same system as those of the " Deutschland."

(r.) The engines of the " Kaiser Wilhelm II.," also built at the Vulcan
Works, are of a special type; the arrangements are shown in Fig. 105,
and Figs. 106, 107, Plate X. Each of the two crank shafts has six
cranks in all. The arrangement of the cylinders by which these cranks
are driven is as follows, beginning from the after end : — First an l.p.
cylinder, then No. 2 m.p. cylinder, next No. 1 m.p. cylinder with an
H.p. cylinder on top, then No. 1 m.p. cylinder with an h.p. cylinder on
top, next No. 2 m.p. cylinder, and lastly an l.p. cylinder. Thus the
three after and the three forward cranks are each, as it were, driven
from a quadruple-expansion engine. A watertight bulkhead divides the
after from the forward three-crank engine. The sequence of the cranks
is shown in Figs. 103, 104. The port and starboard engines together
develop about 40,000 i.h.p. at 80 revolutions per minute, with admis-
sion pressure of 245 lb. per square inch above atmosphere. The
dimensions are as follows : —

* In the phottigraph, Fig. 1()2, the two ii.i'. cylinders which are mounted on the
rovers of the L.v. cylinders, arc not in place. For their arrangement see "Steam
Cylinder" Section, and for complete description see Zeitschrift des Vcrcincs Dcutscher
InqinUure, 1900.

122 MARINE ENGINES AND BOILERS.

Diameter of h.p. cylinder, 38 inches. No. 1 m.p. cylinder, 50 inches.
„ No. 2 M.p. cylinder, 75 inches. l.p. „ 112 „

Stroke, 70 inches.
Diameter of crank shaft of the three forward cranks, 20 inches.
>i ♦! >i » alter ,, -^o ,,

„ crank pin of all six cranks 25 „

Total length of crank shaft from the forward end
to the aftermost flange coupling
The condensers, of which there is one for each aggregate of three

i 72 feet.

! I I

i ! . !

I

I — h-T 1 * I

j I i i i i

IP MPJl MPJ MPl MPtt IP

HP MP

Fig. 103.

cranks, and therefore two for each of the port and starboard engines,
have each a cooling surface of 11,700 square feet. Each condenser
is fitted with an independent Duplex air pump, and a single-cylinder

circulating pump. The port and starboard
ahea4i engines have each two Brown's reversing engines,

the valves of which are connected by a rod,
and worked with the greatest ease through a
single hand lever. The crank shaft is of nickel
steel. Each cylinder is supported on two hollow
cast-steel columns with large openings in them.
The bed plate consists of cast-steel cross girders,
with flanges for bolting them together longi-
tudinally.

g

49. Longitudinal Bracing of the

Cylinders. — Most engineers endeavour so to
connect the cylinders, that the whole engine forms one stiff structure.
This is a necessary arrangement, to ensure the stability of each cylinder,
and to prevent vibrations in the ship. It is obtained by the use of
strong bed-plates and engine seatings, and also by bracing the cylinders
or the upper ends of the frames together as firmly as possible. In
connecting the cylinders of large engines, allowance must be made for
expansion, /,^., the tie-bars or connecting supports should be joined in
such a way, that each cylinder can expand without throwing the one
next to it out of line. (See Figs. 108 and 109.) The arrangement in

THE MAIN ENGINES, 123

Fig. 108 is not so good as in Fig. 109, which allows for free expansion.
Similarly the arrangement in Figs. Ill and 112 is preferable to that
shown in Fig. 110.

In small engines the dimensions of
the cylinders are less, and their expansion
is not of so much importance, so that
there is no disadvantage in having all the
cylinders cast in one, or rigidly connected
through their valve chests. .^.

§ iJO. General Remarks on the
Arrang:ement of the Main Engines.

— The arrangement of the cylinders de-
pends in great measure on the choice of the
valves. In triple-expansion engines piston
valves are generally used for the h.p. and
often for the m.p. cylinders, and flat slide
valves fortiiei..p. ; in quadruple-expansion
engines the second m.p. cylinder is some-
limes fitted with flat slide valves. For fear
of excessive wear, large flat slide valves are
generally avoided when the steam is at
any considerable pressure, say 55 to ^5
lb. per square inch. For large powers
two piston valves or two flat slide valves
are placed side by side in the same valve
chest- If it is desired to make the
engines very short, a valve gear is selected
which will allow of the valves being Kig. loj,

placed at the side of the engine, such as

the Heusinger or the Klug valve gear. (See pages 17G, 181.) If the
lengths of the crank shaft have to be the same, the distances between
the cylinders are practically fixed, because only two long or two short
sections, or one long and one short section can be coupled together.
(See "Crank Shaft.") But this need not be considered in light engines
for warships, the crank shafts of which are forged in one piece.

In balanced engines the valve chests are so placed that both the
distances between the cylinders, and the position of the eccentrics, arc
favourable to balancing. (Seeg35.) If the engines are not balanced the
cylinders are sometimes crowded closely together, the object being to
diminish the twisting of the engine in the longitudinal direction as
much as possible.

To make the most of the available space, and esp>ecially to reduce

124

MARINE ENGINES AND BOILERS.

the height of the machinery-, Thornycroft has built engines for torpedo-
boats and destroyers, with inclined columns, the piston rods of which
therefore are at acute angles to each other. This class of engine has
been fitted in H.M. destroyers "Ardent," "Boxer," and "Bruiser."

Fig. 110.

i.H.p. per engine 2,200, at .190 revolutions per minute. Diameter of
cylinders, h.p. 19 inches, M.p. 27 inches, two i..p. 27 inches. Stroke 16
inches. Boiler pressure 200 lb. per square inch. The two forward cranks
are at right angles to the two after cranks, and the individual cranks of

s8

THE MAIN ENGINES.

125

each pair are at such an angle that if the cylinder axes were in the same
plane, they would be at ISC'. (Thus one piston of each set is always at
the upper dead point when its fellow piston is at the lower dead point, and
vice versa.) Under certain conditions this type of engine is well balanced.
(Compare page 90.) Although it is desirable, as far as balancing the
engines is concerned, to place the large cylinders in the centre,* yet to
facilitate the working and overhauling of the engines, it is better to place
the casing containing the large flat slide valves at either end of the engine,
so that the covers of the slide-valve chests can be easily taken off, and

3iE^^3[-

Fig. 112.

the valves removed for inspection. If a flat slide valve is between two
cylinders, care must be taken to leave sufficient space between its cover
and the adjacent cylinder so as to be able easily to remove the valve.

§ 51. Starting the Engine. — As all marine engines, and espe-
cially those for warships, must be rapidly manoeuvred, it is necessary
that they should be easily and quickly started. This must be taken
into account when determining the position of the cranks. Nearly
all marine engines have starting or bye-pass valves (see page 143) to

* In some engines with Schlick balancing the slide valves of the l. p. cylinder
are quite outside, and the connecting rods of the outer cranks are consequently made
as light as possible (see Fig. 114).

126 MARINE ENGINES AND BOILERS.

admit live steam into either the M.p, or the l.p. receiver, or to the
top or boliom of the m p. or l.p, cylinder. The diagram. Fig. 115, for a
three^rank engine, shows where the auxiliary steam should be intro-
duced. The cut-offs are marked in the concentric circles, of which the
outermost represents ihe l.p., the middle the m.p., and the inner the

H.P. cylinder. The h.p., m.p., l.p. crank shaft is made to rotate on its
axis through any position. If one of the cranks falls within the arc or
section representing the cut-off of its corresponding cylinder, the engine
can be started if live steam is admitted to that cylinder on the proper
side. It must not be fot^otten, however, that, up to about 20° beyond
the dead point, the crank cannot exert sufficient turning n

THE MAIN ENGINES.

127

the engine in motion. The diagram shows at what positions of the
crank the engine will not start, and to top or bottom of which cylinder,
or to which receiver, the live steam should be supplied. If for any
given position of the crank shaft, none of the cranks are in their

Bottom Dewa Point.
Fig. 115.

corresponding arcs representing the cut-offs, the engine will not start
merely by supplying steam to the receiver. If live steam cannot be
supplied to the top or bottom of a cylinder, the valve gear must be
reversed, the engine started backwards, a better position of the crank
obtained, and the engine re-started in the right direction.

SECTION V.
DETAILS OF MAIN ENGINES— THE CYLINDER.

§ 52. General Remarks. — Steam cylinders are almost universally
made of the best fine-grained cast iron ; gunmetal or bronze being only
used in exceptional cases, and then only for small and specially light
engines. If this letter metal be employed, it is usual to make the piston
and piston rings of steel or cast iron, because bronze working upon
bronze does not wear well.

The cylinders are made either with single or double walls. Single
or unjacketed walls are used in the main engines of small warships,
where it is necessary to have them extremely light, or in the engines
of ordinary small or medium-sized cargo or passenger steamers, and
in all auxiliary engines — in other words, wherever economy of weight,
cheapness, and simplicity of construction have to be considered.
Double ivalls are used either when steam jackets are used, or when
the liners may need renewing.

§ 53. Velocity of Steam. — The mean velocity of the piston, cor-
responding to the normal speed of the engine, is always used as the
basis on which to calculate the different cross-sectional areas of the
passages, receiver pipes, &c., and is —

c— -■— =feet per second.

Here ^= piston speed in feet per second; j = stroke in feet; « = revolu-
tions per minute.

The mean velocity of the steam (t/) in ordinary engines is as follows : —

1. Main steam pipe. «/= 100 to 130 feet per second. If the steam
pipe is very long the speed should be a little less.

2. In steam passages of the h.p. cylinder, t/ = 80 to 100 feet per
second; m. p. cylinder, z^= 100 to 120 feet per second; l. p. cylinder,
?;= 120 to 140 feet per second.

3. In exhaust passages and receiver pipes of the h.p. cylinder, «; = 65
to 80 feet per second; m. p. cylinder, «; = 80 to 95 feet per second;
L.p. cylinder, z/ = 95 to 110 feet per second.

In very light quick-running engines, where a saving of weight and
space is of more importance than economy, the steam velocities given
above may be increased by 10 to 20 °/^,

THE MAIN ENGINES.

129

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130

MARINE ENGINES AND BOILERS.

The cross-sectional area of the passages is found from the equation —

X f 4.U y A velocity of the piston
/=area of the cylinder x — - — ^^ — - - — ^

velocity of the steam

In Table No. 16 the cross sectional areas of steam passages and
the diameters of main steam pipe are given for varying piston speeds,
assuming respectively the area of the cylinder and the diameter of the
H.p. cylinder to be unity.

§ 54. Thickness (5) of the Cylinder Liner— The thickness

of the cast-iron cylinder liner may be determined from the following
empirical formula, which gives fairly correct results for vertical cylinders
for steam pressures above 85 lb. per square inch-^

6 =

+ '4 inch

5,125 + 10/

in which ^/= internal diameter of the cylinder in inches ; /= pressure of
steam in pounds (above atmosphere) ; 8 = thickness of the cylinder liner
in inches. Table No. 17 is calculated from this formula. The thick-
nesses of liner here given may be used for ordinary single-cylinder engines,
and for the high-pressure cylinder of multiple-expansion engines for pas-
senger and cargo steamers. The thickness of liners may be made the
same for the m.p. and l.p. cylinders as for the h.p. cylinder, without
taking their diameters or the steam pressures into account.

Table No. 17.
Thickness of the h.p. Cylinder Liner in next larger yV ^^^^-

Diameter of the

For 150 lb.

Cylinder.

per square inch.

4 inches

\ inch

8

12

fi

1?

15

20

7

24

16

TIT

27

1

32

u

36

ifV

39

n

43

If

48

1 "^

51

55

1 6

59

1 ^ ^

For 175 lb.
per square inch.

For 200 lb.
per square inch.

i inch

tV inch

a

1 1

1

7? »

rs

99 1

:i

:j

4 "

4

1 ;J

1

Tff »

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1 5

1 R

TT >»

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1 „

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91

-F

THE MAIN ENGINES. 131

In light engines the thicknesses of the liners may be slightly less
than those given above, and the same applies to their m.p. and l.p. cylin-
ders, provided they are first tested under water pressure. If a mai^n
has to be left for reboring, the
thickness must l>e increased by
about ' to yV inch. For liners
of \vrought steel the thickness
in the table may be reduced 30

g 55. Thickness <S,) of the
Cylinder Jacket or Outer
Shell.— The jacket is generally
made of the same thickness as
the corresponding liner ; but the
extra thickness of the liner, to
allow for subsequent boring, need
not be considered. The space
between the jacket and the liner,
if it is to be used as a steam
jacket, must not be less than
^ inch, and in large cylinders
should be as much as 1^ to I.'
inch.

S 56. Thickness (S.^) of the
Walls of Cylinders without
Liners. ^ — The walls of such
cylinders are from 10 to 15 "/,
thicker than those fitted with
liners. To strengthen the cylin-
der, it is usually stiffened ex-
ternally with circumferential rings

or webs about 0-8S^ in thickness, tig, ii8_

and rather less in depth than the

cyligj^er cover flanges. These webs are usually spaced about ten to
fifteen times the thickness of the cylinder walls apart.

S 57. Method of Fixing the Cylinder Liner.— In small cylin-
ders the liner is pressed into the cylinder casting and firmly secured at the
lop and bottom. The joint is made with asbestos cord at the bottom,
and by a turned copper ring at the top, or by a kind of stut1!ing-box with
asbestos and iron wire packing. Longitudinal movement is prevented by

MARINE ENGINES AND BOILERS.

Fig. 117.

a few tap-bolts at the side, or by lengthen
ing the liner till it touches the flange
of the cylinder cover, (See Fig. 118.)

In large cylinders where more Space
is available, the cylinder liners are held
to the bottom by a steam-tight flange
and tap-bolts or studs. (See Figs. 116,
117, 119.) In engines with a longer
stroke than about 3 feet, the liner has
often a fitting ring in the middle, in
which, if the jacket is heated, a few cross

grooves are provided. The diameter d-^ of the tap-bolts or studs is
about = S. The bolts, which are made with collars and squared heads,

THE MAIN ENGINES. 133

are of the best wrought iron, or Siemens-Martin steel. Bronze screws
are not to be recommended. (See Part VII.)

The width of tJu holding-down flanges should be as small as pos-
sible, and their thickness a about = 1 '35, and the spacing t of the bolts
is about —

/= ^d^ for the h.p. cylinder

/= 5*5^^1 for the m.p. „

/=7^, for the l.p. „

and where possible they are made of the same size for all the cylinders.
The joint under the liner flange is best made with red lead and a
thin copper wire, while the upper end is made steam tight with a copper
ring w^ell caulked into a slightly dovetailed groove. Asbestos and iron
wire may also be used for this purpose instead of a copper ring. The
joint is generally still further secured by means of a wrought-iron ring,
held in place by ^ to |^ inch set screws. (See Fig. 116.) To avoid
all shoulders on the wearing surface, the length of the actual working
surface of the cylinder or liner is such that the piston rings are either
flush with it, or overrun it by about \ inch at the top and \ inch at
the bottom, according to their width. At each end, beyond the work-
ing surface of the liner, the metal should be cut or bevelled away, its
diameter being increased by J to ^ inch. (See Fig. 116.)

§ 58. Cylinder Cover Studs. — As a rule the studs for securing
the cylinder cover are made of the best wrought iron or steel. Their
overall diameter d^ may be approximately taken as about = S, or up
to 1'258. The stress on the screws, produced by the maximum steam
pressure on the cover of the h.p. cylinder, should never be more than
from 4,700 to 6,500 lb. per square inch. The spacing / of the studs
varies with the steam pressure, and should be as follows : —

For pressures up to 50 lb., /=from T^d.^ to ^d^,

„ „ from 50 to 100 „ /= „ 4-5//., „ 5//.,.

„ 100,, 170 „ /= „ 3-54 „ 4-5^2.
„ 170 „ 200 „ /= „ 2-7//2„ 3-5^2.

» >»

It is desirable to make the studs the same size for the covers of all
the cylinders, and their nuts should be about 1 ^d^ to 1 'f>d^ deep, so
that the thread of the screw may not wear out unduly quickly with the
constant screwing and unscrewing which is apt to take place when
shorter nuts are used. In the best work the nuts are case-hardened.

§ 59. Cylinder Flang^es. — The thickness of the flange should be
about l-35i to l-46i, ^^ ^^ ^'*^^^ -'^^2 ^^ ^'^^2- ^'o strengthen the

134 MARINE ENGINES AND BOILERS.

flange it is desirable, in large cylinders, to fit radial ribs between each
pair of bolt holes.

§ 60. Cylinder Bottom. — If cast-iron pistons are used, the bottom
of the cylinder is generally flat, and must therefore be well stiffened with
ribs, which serve at the same time to brace the supporting brackets
or feet to the bottom and wall of the cylinder. The thickness of the
bottom should be about = 8 or = I'lS, and that of the ribs = S ; the depth
of the ribs should be at least 55 to 66. The ribs should be as many
in number as are required to afford sufficient support to the flat surfaces
between them. If the cylinder bottom is double, the thickness of each
wall should be 0*98, and the distance between them, measured inside,
about 58 to 68, while the radial ribs bracing top and bottom together
should be 0-88 to 0*858 thick. If conical pistons are used, the thick-
ness of the walls may be from 09 to 0*95 of that given above. It is
of special importance to make the circular centre rib, which forms the
opening for both the piston rod and the boring bar, very strong.

§ 61. Cylinder Feet. — To connect the cylinders with the A frames
or columns, they are cast with feet or faced brackets. In the design of
these feet care must be taken that the strains coming upon them are dis-
tributed over as large an area of the cylinder shell or bottom as possible.
The thickness of metal of the feet, if they are cast hollow or ribbed, is
generally about 0*858, and that of the flanges 1 58 to 1 '68. The bolts
securing the feet to the columns are of such dimensions that, at the maxi-
mum pressure on the piston, the stress on them will not be more than
from 3,000 to 6,000 lb. per square inch. Their diameter being taken at
1*58 to 1*88, the number required can be easily determined.

§ 62. Cylinder Passages and Valve Casings.— If piston

valves are used, the valve casings are cast in one with the cylinder,
and are, where possible, cylindrical in shape. In small cylinders the
valve casings are cast in one piece with the cylinder, but with larger
cylinders for merchant vessels they are often cast separately, and fitted
with an independent cover, so that the valve may be easily removed.. In
warships, to economise weight, and because the valve cannot be taken
out from above, on account of the armoured deck, the valve chest is fre-
quently cast in one with the cylinder, and is provided with a large cover
at the side.

Thickness of valve-chest walls, 0*9 to 0*9.58.

!^ 63. Calculation of Flat Surfaces in Valve Chests, Pas-
sages, and Covers. — The flat walls of the valve chest and cover

/

THE MAIN ENGINES.

135

must be stiffened with ribs to afTord sufficient strength to the surfaces
between them.

1 . Empirical DeUrmination of Thickness of Walls and Pilch of Rib\.
— A flat cast-iron surface is strong enough if —

where 8 = thickness of the surface in inches ; b = smallest distance be-
tween two contiguous ribs in inches ; fi = steam pressure in pounds
per square inch (above atmosphere) on the surface in question.

Generally speaking, however, the thickness of the walls of the sur-
faces, between the ribs, is made e^ual lo Ihe thickness of the cylinder liner
itself The stiffening ribs must be so calculated that they will safely
bear the steam pressures upon the surfaces a x b (see Fig. 121). b is
here the smaller side of the rectangular area ; the bending stress may
be taken at about 4,500 lb. per square inch. About 6S should be
allowed, on either side of the rib, for the width of the flange of the
T-shaped cross section. The height of the ribs should be about
4 to 58 ; thickness the same as that of the flat wall ; and the distance
of the transverse ribs from each other is

= about 12 to 148 for pressures of 50 lb.

19

9)

99

10

99

99

99

128

108

88

99

99

99

50 to 100

99

100 „ 160 „

160 „ 200

99

99

Fig. 120.

Fig. 121.

. The ribs should, if possible, be placed on the side exposed to the
pressure, as the metal is then better distributed to withstand the pressure
than if they were on the other side, the larger area of the metal being
thus irr'tensioni Such ribs certainly offer considerable resistance to

136

MARINE ENGINES AND BOILERS.

the incoming steam, esjjecially if they are transverse to its direction of
entry, and they should therefore be allowed for, when determining the
cross sectional area of the steam passages.

2., Calculation of Flat Surfaces in Valve Chesty Passages^ and Covers^
on Bach's Method, — From a large number of experiments it has been
found that plates having corners, split across the diagonal line when
subjected, under hydraulic pressure, to a uniform strain. Such covers
should therefore be calculated for a bending stress along the diagonal.
(See Fig. 122.)

Sxcamp^eMun^

Fig. 122.

Load on the cover, p = ax^x/, / being the pressure per unit of
surface. Half of the force p may be considered as acting at the centre
of gravity s of half the cover; then the bending moment about the

diagonal* is „ x o- "^^^^ force is counteracted by two forces which

act along the sides a and d. The resultant moment of these forces
about the diagonal is —

p c P C PC

""=2 '^2-2'^ 3 = 12

If c be expressed in terms of the sides a and b^ and p = a x ^ x/, then
we get —

p^r_, /.^'V/-
12 "^12 7^2 ^^'-^

M,

THE MAIN ENGINES.

137

The coefficient <^ is introduced to allow for the initial pressure ; it varies
from ri25 to 0*75, but is generally taken as= 1.

The maximum stress in the cover is obtained from the equation —

T, ,. ^ bending moment

Bendmg stress = — r— ^ : —

modulus of cross section

l:98S

Fig. 123.

■

Example. — ^What is the stress in a cover, of the form shown in
Figs. 123, 124? Working pressure, 185 lb. per square inch. Material,
cast steel.

The moment of inertia of the section of this cover on the line ab is
1^1-478 inchest Distance of the centre of gravity from the outside
fibre is ^ = 1 '09 inch.

Therefore the modulus of cross section

is z= =1*35 inches^.
e

trSOAi.perO-'

jUt^UU

^w%^^^^^

^ — 'rr' ,r4-u

/r-s

0uImJ'^

^putaide

Fig. 125.

Fig. 124.

The bending moment is —

p aW 185 13-8=2x9-82
^'»>=r2 ^ 'j^^^Ti " 713^8'-' ?9.8-^ = ^^''"^ ""• ^^-
The stress in the cover is thus —

s = ^1" = l!^-i?P? = 12,480 lb. per square inch.
The diagram, Fig. 125, shows the distribution of the stresses.

138

MARINE ENGINES AND BOILERS.

§ 64. Piston-valve Liner and Ports. — In smaller engines a
piston valve is often allowed to run direct in the bored casting of the
valve chest, and a flat slide valve to work direct upon a valve face cast
in one with the cylinder. In larger cylinders the piston valves work
in specially hard cast-iron liners, fitted into the valve chests. (For
dimensions see S "^3.) These liners have to be accurately fitted and
firmly bolted in position. (See P'ig. 126.) In the same way for flat
slide valves of larger cylinders separate valve faces of hard cast iron are
screwed on to the cylinder faces so that they can be renewed when
necessary. (See Fig. 127.)

The valve faces, when cast with the cylinder, are made about 0*9 to
riS in thickness. The removable valve faces are made from 8 to I'lS
thick, and are fastened to the cylinder by wrought- iron or steel set screws,
with round or squared countersunk heads, or, in small cylinders, with

!^>^AV1

Fig. 126.

Fig. 127.

cheese heads. The pitch of these screws is usually about six to nine
times their diameter. On account of the wear on the valve face, the
he&ds of the screws should be sunk to such a depth that their upper
edge is ^ inch or so below the valve face. (See Fig. 127.) The liners
and valve faces should be of the best hard close-grained cast iron.

§ 65. Water Tests for Cylinders. — Before the steam cylinders
are covered with a non-conducting material, they should be tested by
water pressure for strength and tightness. The pressures used in these
tests, where p = boiler pressure in pounds per square inch, are —

1. For compound engines, h.p. cylinder about 1*3/.

„ „ L.p. „ 0-45/.

2. For triple-expansion engines, h.p. cylinder about 1*3 to 1*4/.

„ „ „ M.p. „ 0*7 to 0-9/.

, „ „ L.P. „ 0*25 to 0-3/.

THE MAIN ENGINES. 139

3. For quadruple-expansion engines, h.p. cylinder about 1*3 to 1'4/.

„ 1st M.p. „ 0-75 to 0-9/.
„ 2nd M.p. „ 0-4 to 0-5/.

„ L.p. „ 0*2^.

If the cylinder has a liner, the space between it and the shell must
also be similarly tested by hydraulic pressure. The valve chest and
cover are tested with the cylinder, but the exhaust passages are often
submitted to a separate test

§ 66. Rules for Construction. — Care should be taken that no
undue local thickening occurs in the material of which the cylinder is
made, as such places almost always become porous and leak. The ribs
in the cylinder passages sometimes crack as the metal cools down after
casting, and it is desirable to strengthen those which are exposed to
strains by stay-bolts. Large cylinders are thus frequently strengthened
by stay-bolts and ties at dangerous places. The opening in the cylinder
bottom for the separate stuffing-box casting, or the hole for the piston
rod, must be made of such a size that the boring bar can be passed
through it. To join two cylinders by steam-tight flanges is only cus-
tomary in small or medium-sized engines ; with large cylinders it should
be avoided. These should either be made quite independent of each
other, or if bolted together, the connection should not be exposed to
steam pressure. (See page 124.) The covers of the valve chests should
be so arranged that the valves are easily accessible. In the same way
care should be taken that the stuffing-boxes can be examined without
difficulty while the engine is running. All flanges for connecting pipes
and external fittings must be outside the cylinder lagging. The latter gene-
rally consists of fossil meal (a mixture of infusorial earth, asbestos fibre,
and some binding or cementing material) for the h.p. cylinder, and felt
with layers of asbestos, or felt only, for the l.p. cylinder. This is covered
with sheet iron from about 20 to 14 S.W.G. in thickness, and should be
as simple as possible, as elaborate sheet-metal lagging is expensive.

§ 67. Cylinder Fittings.— These are—

1. Regulating valve.

2. Throttle valve and governor.

3. Starting valve.

4. Cylinder and valve casing drain cocks.

5. Relief valves.

6. Indicator connections.

7. Connections and fittings for steam to jackets.

1. Regulating Valve, — This is fitted directly to the h.p. cylinder,

140 MARINE ENGINES AND BOILERS.

and is so arranged that the engineer can open and shut it conveniently
and quickly from the platform by a liand lever or wheel.

Single seated valves are as a rule only used in small engines, as when
fully open they do not close quickly enough. To close lac^e valves of
this kind, when the steam enters below the seat, and to open them
when the steam pressure is above the seat, requires too much power,
and such valves are therefore only used when specially balanced.
(See Fig. 128.)

In this type of valve the top of the valve is shaped like a piston and
iits into a cylinder. There is a small valve and seating in the middle
of the large valve which, when the valve is shut, is closed by the
spindle. The steam first enters the outer part of the valve box, and as

KiK. 1-28.

the piston does not fit tightly, it finds its way into the space above the
valve and presses the latter firmly against its seat. As the spindle rises,
it opens the smaller valve, which has a lift of about yV to \ inch,
the steam above it escapes, and the lai^e valve is then easily raised by
the spindle. To avoid any chattering of the valve when it is open, the
spindle is often fixed to it, in which case there is a small separate
bye-pass valve on the valve cover, which can also be worked by the
engineer from the platform. Through this bye-pass valve, the steam
above the larger valve can pass into the H,r. cylinder and thus relieve
the main valve. The spindle of the main valve is worked by the
engineer from the starling platform either directly or by means of gear-
ing, or sometimes by means of a special auxiliary engine.

THE MAIN ENGINES. 14l

Butterfly valves, instead of mushroom valves, are often used in large
engines developing up to 2,000 h.p. These are also worked by hand levers
from the engineer's platform. The angle of travel of the valve is about 90°.

In large engines double-seated valves are in favour, worked either
direct or through an auxiliary cylinder, as shown in Figs. V2^ to
131. The valve body, valve, separately fitted valve seat, and the
spindle are of bronze. The upper valve
is only from J to J- inch larger in dia-
meter than the lower one, so that both
valves can be drawn out at the top, and at
the same time the difference of pressure
on the two valves is reduced to a mini-
mum. The valve spindle is carried down-
wards through a stuffing-box, and has a
rod for operating the valve att3.ched to its
lower end. When the engine is completely
shut down the valve can be closed by
means of a hand wheel and screw acting
on the upper end of the spindle, which
projects through the top cover of the valve
casing. The main steam pipe is connected
through an expansion joint forming part of
the stop valve. (See "Main Steam Piping,"
page 423.) A throttle valve is frequently
placed either outside this valve, or between
it and the h.p. cylinder. This can be
operated by the engineer either by hand
or by a governor (of the Aspinall or some
similar type), to prevent the racing of the
engine in rough weather.

2. Mention may here be made of the
Aspinall Governor (Figs. 132, 133, 134,
Plate XI.). It consists of a weight w,
hinged to a bracket, which acts upon two y^„ 129.

pawls p, p,, and is carried on a frame,

bolted to the air-pump lever or some other reciprocating part. If the
normal speed of the engine is exceeded by about 5 %, the weight w, in
conformity with the laws of inertia, lags behind on the downward stroke
of the governor, and is held fast by a dutch, causing the lower pawl p
to project, and the upper pawl to be brought back. ^Vhen the governor
reaches its lowest position, the lower pawl catches under lever h, and
carries it with it up to the highest position, thus shutting off steam by
closing the throttle valve. On the return stroke of the governor, the

MARINE ENGINES AND BOILERS.

latch D strikes against lever h, after the upper of the two pawls has
already slipped past it ; w is released, and the two pawls again take up
their normal position, the upper one in front, the lower behind. So

THE MAIN ENGINES. 143

long as the speed of the engine exceeds the normal, this action is
repeated when the governor is in its highest position, and the throttle
valve is held closed. When the speed becomes normal, the upper pawl
moves the lever h back into the lowest position, and the throttle valve
is again opened. The emergency gear only comes into operation when
the speed is abnormally increased, such as in cases of fracture of shaft,
loss of the propeller, &c., in which cases the small weight a is left
behind. This brings weight w into the position for closing the throttle
valve, where it is held fast, and the valve remains closed until the
weight is released by hand.

3. Starting or bye-pass valves are used to supply the m.p. or the l.p.
valve chests or cylinders with live steam, in order that the engine may
start away more easily. (See § 51.) If the cut-off in the cylinder is
not less than 60 */^, it is sufficient to introduce the steam into the valve
chest. With earlier cut-offs it is better so to arrange the bye-pass
valves, that steam can be admitted direct into the cylinder, on whichever
side of the piston it may be required. In small engines, instead of an
ordinary bye-pass valve, a dead-beat valve is often fitted, which is kept on
its seat by the pressure of the steam, but can be raised at will by a lever.
The diameter of the auxiliary steam pipes to each valve chest should be
from \ to \ the diameter of the main steam pipe. If the steam is
admitted direct to the cylinder, the diameter of each pipe may be rather
less. This auxiliary steam should be taken from the main steam pipe
to the engine, and controlled by a separate stop valve.

4. Cylinder drain cocks in vertical steam cylinders are only needed
on the bottom, and should be so arranged that they can be easily
worked by the engineer from the platform ; and in larger engines they
must therefore be fitted with the necessary gear for this purpose.
Their diameter is from ^\ to -^^ the diameter of the cylinder; the
lower value should be taken for large cylinders. Drain cocks of more
than 2 inches in diameter are not usual, and instead of one large cock
two smaller ones are then used, the discharges from which are connected
to each other. The drain cocks should be placed below the cylinder
bottom, and as low as possible in the passages. It is desirable, especially
in the h. p. cylinder, to have a drain cock also on the cover side, because
the water collecting at the top of the piston is only carried off very
slowly with the exhaust steam. The drain pipes are generally led to
the condenser, but they may also be carried to the hot well. In the
latter case they must, however, be fitted with non-return valves, to
prevent the water being sucked back. The drain cocks for the valve
chests should be of the same size as those for the corresponding
cylinders, and arranged in precisely the same manner.

144 MARINE ENGINES AND BOILERS.

5. Relief valves should be fitted both on the cover and on the
bottom of the cylinder, to prevent the risk of fracture of the cylinder or
rod, and allow water to escape, should it suddenly collect and " water-
hammer " occur in the cylinder. Their diameters should be —

In the H.p. cylinder about y.j to yV ^^ diameter of the cylinder.

MP i' I

» M.r. „ y^ „ Yjj ,, „

T P 11

The relief valves should be surrounded with a casing, to prevent any
one being scalded by the escaping water. In warships a small copper
pipe is frequently led from this casing to the bilge. The springs of the
valve should be set to about 12 times the maximum pressure which can
occur in the valve chest of the corresponding cylinder. Relief valves are
also generally fitted to the valve casings, when the auxiliary steam is led
direct to the valve casing of the m.p. or l.p. cylinder, and should be of
the same diameter as the relief valve on the corresponding steam cylinder.

6. Indicator connections and fittings are now provided for all main
engines, and in warships also for some of the auxiliary engines. The
connections to even the smallest cylinder should not he less than \
inch, and the pipes leading to the indicator should be of the same
size. In large engines the indicator connections should be of larger
diameter, at least from I^ to 2 inches, and may then be gradually
reduced to \ inch as they approach the indicator cock. As a rule there
is only one indicator cock to each cylinder. The two sides of the
piston can be put into communication with the indicator cock, by means
of the connecting pipes and a three-way cock. The latter should be
either in the centre of the cylinder or at the upper end ; the second
arrangement is the better, because the top end is more likely to be
free from water. The indicator drum should, if possible, be worked
off an air-pump or similar lever, and, failing this, through a light
lever from the crosshead. The gear should be as simple as possible.
(See Part VI.)

7. \i jacket-Jieating is required, the steam is led into the upper part
of the cylinder jacket, and the condensed water drawn off from the
lowest part, and led either direct or through a steam trap to the con-
denser. The diameter of the steam supply pipe to each cylinder should
be \ to 2 inch and of the drain pipe to the condenser f to ^ inch. A
reducing valve is often fitted to each cylinder, and set to reduce the
pressure in the jacket to that obtaining in the cylinder. It is advisable
to have a small relief valve on each cylinder jacket.

§ 68. Description of Figures 136 to 143 (drawings of actual

THE MAIN EN(;INES. 145

cylinders). — Fig. 13o shows the arrangement of cylinders for a com-
pound engine of about 2O0 i.h.p., working with an initial pressure of
/— 100 lb. per stjuare inch. As usual in small engines, the cylinders
have no liners. The h.p. cylinder has expansion gear and Rider valves,
and the l,p. cylinder has flat d slide valves, with single ports. The h.p.

cylinder is fitted with a hardened valve face. The two cylinders are
bolted together by means of a steam-tight joint. The slide valve of the
L-p. cylinder can be taken out through the cover placed above it. The
backs of the cylinders rest upon upright cast-iron frames and the fronts
upon steel columns. (Compare Figs. 77, 78.)

Figs. 136 and 137 show the second M.r, cylinder, 2,640 mm. (8 feet

146

MARINE ENGINES AND BOILERS.

Fig. 136.

Fig. 137.

THE MAIN ENGINES.

147

H inches) diameter, of the quadruple engines of the ss. " Deutschland."
The cylinder has a liner, two piston valves with steam on the outside,
separate valve liners, and auxiliary steam cylinders to take the weight of
the valves. The flat walls of the passages are strengthened with screwed

X

mm*

stays. The valves can be removed from above. The cylinder is cast
with four feet, which are bolted to the cast-steel framing. Both the
bottom and cover of the cylinder are made double, and each has a man-
hole and relief valve.

Figs. 138 to 141 show the l.p. cylinder, about 1,700 mm. (5 feet 7

148

MARINE ENGINES AND BOILERS.

inches) diameter, of an armoured cruiser. The cylinder has a liner,
but the bottom and cover are single, the latter of cast steel. The
stuffing-box is arranged for Philadelphia packing. The bottom is
strengthened by radial ribs, and the ports by stay bolts, which also serve

Fig. 139.

to secure the liner a. The cylinder rests on two cast-steel frames,
to which it is bolted by two ribbed feet cast on to it. The weight
of the overhanging valve chest is taken by two wrought-iron columns.
On the cylinder bottom there are bosses, A for lifting eyes, g for
draining the cylinder, m for draining the jacket Relief valves s are

THE MAIN ENGINES.

149

fixed to the cylinder cover and cylinder bottom. The valve is a double-
ported slide valve, which works on an independent valve face bolted to
the cylinder. The weight of the valve and valve rod is taken by a
piston working in an auxiliary cylinder. The space above this balanced
piston c is connected to the condenser. The pressure and therefore the
friction of the valve on the face of the cylinder is relieved by a cast-iron
ring of rectangular section placed at the back of the valve and connected

LL^iU'

Fig. 140.

to the valve-chest cover by means of a copper diaphragm, which forms a
flexible steam-tight joint. The ring is kept up against the back of the
valve by springs. The space between the valve and the valve-chest cover,
which is enclosed by the diaphragm, is connected to the condenser. The
valve chest is of cast iron, and the cover of cast steel. Steam is admitted
on both sides of the valve chest, but there is only one exhaust pipe.

Fig. 142, Plate XII., shows the arrangement of cylinders in a
destroyer with triple-expansion engines and two l.p. cylinders. To

150

MARINE ENGINES AND BOILERS.

economise weight, none of the cylinders have liners, and all are
made with single covers, top and bottom. The covers are of bronze.
The H.p. cylinder has a centrally fed piston valve without a separate
liner. The exhaust from the h.p. is led, by means of a pipe running

Fig. 141.

over the m.p. cylinder, to the m.p. valve casing. The m.p. piston valve
takes steam at both ends, has a special liner and piston rings. The l.p.
cylinders have double-ported unbalanced flat slide valves, working direct
on the valve faces, which are cast in one with the cylinders. Each l.p.

THE MAIN ENGINES.

cylinder exhausts through a separate exhaust pipe at the side. AH the
piston-rod stuffing-boxes are arranged for United States packing. The
cylinders are supported by sleel columns. The engine is balanced on

152 MARINE ENGINES ANT) BOILERS.

the Schlick system, hence the thick pistons in the m.p. and first l,p.
cylinders. The crosshead guides are bolted to the columns and to lugs
cast on the cylinder bottoms.

Fig. 143, Plate XIII., shows the l.p, cylinder of the twin-screw
steamer " Deutschland," on the cover of which the h.p. cylinder is fitted.
The L.p. cylinder is entirely jacketed both at the sides and on the top and
bottom. There are two double-ported slide valves side by side, each
having an auxiliary steam cylinder and a balanced ring at the back
of the valve. The h.p. cylinder is secured to the cover of the l.p.
cylinder by a casting and two strong cast-iron feet. The h.p. cylinder
also is entirely jacketed, and is fitted with a piston valve working in hard
finegrained cast-iron liners forced into the valve chest. The two
stuffing-boxes between the cylinders, as well as the stuffing-box of the
, L.p. cylinder, are fitted with metallic packing. The flat surfaces of both
cylinders are stayed.

§ 69. Cylinder Covers. — These are made either single or double.
Single covers are used either for the sake of cheapness and simplicity, or,
as in warships, to effect a saving in weight. Double cylinder cavers are
employed in large merchant vessels, where it is intended to steam-jacket
the covers as well as the cylinder walls. They generally have radial
ribs in which openings are provided to allow the water to drain away, as
well as to connect the cores. (See Figs. 144, 156.) In single cylinder
covers the ribs often have a T-shaped section. It is very important that
the cylindrical or cup-shaped recess, provided in the centre of the cover
for the piston-rod nut, should be very solid. (See Fig. 1 45.) The inner
surface of the cylinder cover must be made as far as possible of the
same shape as the piston, leaving a uniform clearance of from \ to ^
inch, according to the size of the cylinder. This is to reduce the
clearance to a minimum, but as irregularities in the castings cannot
be wholly avoided, less space than that mentioned above should not
be allowed.

Cylinder covers are usually made of cast iron in the mercantile
marine, of cast steel in warships, and in the case of very small and light
engines generally of gunmetal. As the strain on the cylinder cover can
only be calculated approximately, and its shape is usually complicated,
the thickness a must be largely determined by actual experience. For
cast iron this thickness is —

1. For flat single covers a = from 0-9 to 0-958

2. For conical single covers « = „ 0*8 „ 0*855

3. For double covers a = „ 0*75 „ 0*855

5 being the thickness of the cast-iron cylinder lining. (See Table No.
17, page 130.)

THE MAIN ENGINES.

15,*^

Cylinder covers of cast steel are generally made single, and about 60
to 65 "j^ the thickness of a similar cast-iron cover.

Fig. 145.

The height h of the radial ribs as well as the clear space between
the two walls of a double cylinder cover is generally about h = la to

154

MARINE ENGINES AND BOILERS.

Fig. 146.

9a ; thickness of the ribs d = about 0-9S. The thickness and width

of the flange of the cylinder cover depend on the diameter d., of the

cover studs or bolts, as follows : —

Thickness of the flange / = from M5 to I'lody
Width „ „ = „ 2-6 „ 3-3^

For diameter of cylinder cover studs or bolts see § 58, page 133.

THE MAIN ENGINES. 155

If the piston rod passes through the top cover so as to form a guide
to the piston, it is either made steam-tight by a stuffing-box, or it runs
in a closed cast-iron casing bushed at the bottom with gunmetal. A
lubricator should be fitted at the top or side of the casing.

Large cylinder covers of above 4 feet diameter are generally pro-
vided with a manhole and cover, so that the piston or cylinder can be
examined, without having to lift the large cylinder cover. (See Figs.

156

MAKINE ENGINKS AND BOILERS.

1*3, 144, and 146.) The manhole should be al>out 15 to 16 inches
in diameter, or, if elliptical, 11 by 15 inches. In order to take off the
cylinder or manhole cover easily, from two to five tapped holes should
be provided, into which strong tap-bolts can be fitted, for prizing up
the covers. Two strong eye-bolts on the covers, exactly opposite each
other, should also be provided, to which the lifting gear for raising the
covers can be attached.

General Remarks.— ^h.^ covers of all cylinders are lagged with a
non-conducting composition, over which is placed a covering of sheet
iron or of chequer plate, if the cover be large and much trodden on. If
the covers are steam -jacketed the core-holes must be carefully closed-

Fig. 148.

Kig. Ufl.

This is generally done with slightly conical screw plugs with a fine
thread, firmly screwed home. If the holes are merely rimered out and
plain taper plugs simply driven in, the covers are liable not to be
quite so steam-tight.

§ 70. Stuffing-boxes.— Soft packing is generally used — for steam,
asbestos or " Tucks " packing, with interwoven wire ; and for water,
greased hemp rope. The depth and thickness of the packing art;
determined by the pressure to which the stuffing-box is exposed. Tablu
No. 1« gives ihe usual dimensions of these stuffing-boxes for pressures of
140 to 170 lb. i)er square inch. For higher pressures the depth of the
packing space musi be increased. In stuffing-boxes for rods up to li

THE MAIN ENGINES.

157

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158

MARINE ENGINES AND BOILERS.

rod is about «] q^ the diameter of the rod.

inch diameter the gland is generally of gunmetal ; but for rods of larger
diameter the gland is made of cast iron bushed with gunmetal. The
clearance between the neck bush and rod, and between the gland and

The thickness of the flange
of the stuffing-box collar is
^=1-25 to l-SdT^, d^ being
the diameter of the gland
bolts. The distance of the
latter from the centre of the
rod should be as small as
possible, in order to avoid
bending the flange of the
gland. The distance c (Fig.
1 48) from the centre of the
bolt to the outer face of
the gland is generally about
l-O^/j to 1-25^1.

In order that the nuts
of the stuffing-box may be
turned equally they are fre-
quently shaped like cog
wheels, and engage in a
common circular rack,
thus enabling them to be
turned at the same time.
(See Fig. 147.)

§71. Metallic Pack

ing^s. — In modern practice
stuffing-boxes are frequently
packed with metallic pack-
ings^ of which there are a
large number of different
kinds. Fig. 149 shows
the ^^ Katzensiein packing"
which consists of a number
of conical rings divided
into two or three j>arts.
The rings next the piston
rod are of white metal, those between them of bronze. In order that
each ring may exert an elastic pressure, one or two turns of common
hemp packing are laid over the top ring, and held in place by the gland.
This packing does not allow of any lateral movement of the rod. In the

V ' . y^y ///. . . ' <»«^ U4^ .ii^

Fig. 150.

THE MAIN ENGINES.

"■Si/ieUing packing" shown in Fig. 147, the outer metallic rings are in

one piece, the inner, of white metal, in three or four parts. The bottom

of the packing space and theuppermost

metallic ring are turned smooth. A

slight lateral motion is allowed 10 the

piston rod, as the space round the outer

rings gives a small amount of play.

This space isconnecied Ihrougha small

cock with the condenser, to draw off

condensed steam. Two pieces of soft

rope are laid over the metal rings,

and pressed down by the gland.

Another kind of " Schelling
packing" is shown at Fig. 15Ia. The
rod is enclosed within a ring made
in two halves and dovetailing into
one another. The ring is coniimsed
of an alloy of copper which fuses at
a red heat. Soft packing is fitted at
the liack of this ring. The ring and
the packing are held in position Fig. ISIa.

a.\ially by several other turns of soft

packing and an adjustable gland, the object being to allow the rod
sufficient lateral play.

160 MARINE ENGINES AND BOILERS.

Figs. 150 and 151 show the " United States packing." It consists
of two metallic rings, each in four parts, ground to fit the one upon the
other. Two of the parts opposite each other are filled in with white melal,
while the two other parts are solid, and fit into them exactly. The two
rings are placed at an angle of
90° to one another. They are
pressed against the rod by
lateral springs contained in a
casingenclosingihe inner rings.
At the bottom the rings bear
upon another massive ring,
which is made steam tight on
the side next the gland by
means of a spherically shaped
cover-piece. The rings are held
together at the lop by springs
resting against the bottom
of the stuffing -box. This
packing also allowrs of slight
lateral movements of the piston
rod. The gland of the stuffing-
box has a drain cock, to carry
off the condensation water to
the condenser. Of late what
is known as electro-deposited
p. jj.^ metallic paper packing\as,\itx,\i

used for superheated steam.
(See Kig. l')2.} It consists of a large number of rings made of a very
soft electrically deposited metallic preparation, packed one above the
other. A piece is cut out of the rings, so that when they are fitted into
the packing space they assume a conical shape. To keep the lubricant
in the stuffing-box, one or two turns of soft packing are laid on the top
of the metallic rings.

. Valves.

§ 72. General Remarks. — Slide valves are now almost exclusively
used in marine engines. Unbalanced flat slide valves are generally
employed when the steam pressure does not exceed 100 lb., but piston
valves are, as a rule, used for higher pressures. As ordinary flat slide
valves require a considerable amount of power to drive them, they are
frequently fitted with an arrangement designed to relieve the pressure of
the slide valve on the face of the cylinder (Fig. 158). The space between
the valve and the cover, shut ofi" by the packing ring, is, as a rule, con-
nected to the condenser. (See Cylinder, Figs. 138 and 143.) The desire
to relieve the pressure on the back of the slide valve has led to the
adoption of the so-called covered valves (Fig. 153). The valve, the two
slide faces of which are exactly alike, is made to slide with as little play
as possible between the actual valve face and the cover fitted over it,
the working surface of which is exactly the same as that of the valve
face. As may be seen from the figure, the valve has double steam and
exhaust ports. The valve cover has two sides resting upon the valve
face, and it must be sufficiently strong not to bend or sag from the
pressure of the steam on the back. A strong spring holds it down,
and two lugs prevent any movement longitudinally. The usual types
of slide valves are shown in Figs. 154 to 159.

When flat slide valves are used care must be taken that the valve
and valve face do not bend under the pressure of the steam, and that the
bearing surfaces are large enough to take up the pressure of the valve
without undue friction. To relieve the valve to some extent, and lubri-
cate it at the same time, the valve faces or the slide valves frequently
have grooves cut in them, from f to f inch wide, and from yV to
y inch deep, or smooth fiat-bottomed holes drilled in them from f
to J inch in diameter, to allow freer access of the steam between the
working surfaces. Piston valves, owing to their shape, need no balancing,
but they have the disadvantage of not being as steam-tight as slide
valves. In piston valves fitted with spring rings, or with rings held out
by means of springs, the valve liners are apt to wear unevenly. It has,
therefore, been the practice of late to make piston valves with solid rings.
These, though not so steam-tight as spring rings, have been found to
wear well. The rings are often cut, and the ends firmly fastened

L

MARINE ENGINES AND BOILERS.

THE MAIN ENGINES.

164

MARINE ENGINES AND BOILERS.

Fig. 155.

THE MAIN ENGINES. 165

with screws. This enables the diameter of the rings to be increased
by inserting thin strips of sheet metal as may be required to take up
the wear. (See Fig. 155.) The adjusting screws are often so arranged
that the edges of the rings can be brought together, and the opening
entirely closed. To prevent the rings being forced too far apart, and
thus becoming jammed in the cylinder, lock nuts are provided.

The material both of the valve and of the valve face is generally cast
iron. When flat slide valves are used the valve faces are frequently
separate, and bolted to the cylinder; they are held in place by set
screws of steel, wrought iron, or ** Delta " * metal, the heads being
countersunk. (See Fig. 127.) Piston valves generally work in separate
cast-iron liners, in which the ports are cut. (See Fig. 126.) The liner
is occasionally simply pressed into the cylinder or valve-chest casting,
but it may also be fixed in the way shown at Fig. 126.

§ 73. The Thickness (s^) of Piston Valve Liners is

about = — + '39, D being the internal diameter of the valve liner.

The webs of the valve, which are generally at an angle of 60" to each
other (see Fig. 126), should have a thickness j = from 1 to l'2s^.

§ 74. Ports of Valve Face. — The ports of the valve face are
usually as long as circumstances will allow, and their height correspond-
ingly small. The length of the ports is as a rule from 0*9 to 0*95
the diameter of the cylinder. The effective cross-sectional area of the
openings in the valve face is made about 5 to 10 % greater than the
cross -sectional area of the port, especially in double-ported valves. In
piston valves the effective area of the openings in the valve face are
generally about 10 to 20 % more than the area of the port, as the
various webs across the openings offer more resistance to the steam
than do the smooth walls of the port. The effective circumferential
area of the valve face is about 65 to 75 % of the inner circumferential
area of the valve liner. The exhaust port should be made so large
that, oven when the valve is at the dead point, the actual area of the
opening is not less than the cross section of the port.

^ 75. Symbols used in connection with Slide Valves.—

In the following paragraphs —

r= amount of eccentricity = half travel of eccentric.

S = angle of lead.

a = height of the port in the valve face.

e = outside lap.

/ = inside lap. (See Fig. 156, &c.)

* Bronze screws, of whatever kind, offer little resistance to torsional stress, and
break off easily, if the strain on them, when putting them in, is too great.

166 • MARINE ENGINES AND BOILERS.

§ 76. Stroke of Valve. — As the friction of the valve varies
approximately with its stroke, the latter is made as short as practical
conditions will allow. To diminish the travel of the valve, the larger
cylinders have double or treble ported valves (Figs. 158 and 159). The
latter, however, are only used under exceptional conditions, as they
require too great a length of valve and valve chest. To get a quick
admission and a large opening for steam, "Trick" valves (Fig. 157) are
often used instead of the ordinary D slide valves. The steam is ex-
hausted in the same way as in the ordinary D slide valves.

§ 77. The Amount of Eccentricity, or half the stroke of the
valve, is usually r = from 1 to 1 •4a. It is generally so chosen that the
largest steam port opening (mean value of top and bottom) is as
follows : —

In the H.p. slide valve about 0*8 "j

„ M.p. „ „ 0*75 V of the area of the port.

„ L.P. ,, ,, U'/ )

In piston valves the largest mean steam port opening is generally made
somewhat greater than that given above, on account of the various webs.

§ 78. Principal Dimensions.— 1. Ordinary D slide valve (Fig.
156)—

Thickness of the walls of the valve \ — from 0*5 to 0*6 the thickness

of the cylinder walls.
Thickness of the slide face /= from 0*8 to 0*9 the thickness of the

cylinder walls.

2. Trick valve (Fig. 157)—

a^'^ a-^-b (a being = height or width of the port for an ordinary I)
slide valve). (See Fig. 156.)

^1 = T to -^- ^ = i to } inch, ^o = ^u • ^u = ^o-
4 2

5j = 0'5 to 0*6 thickness of cylinder wall.

/ =0-8 to 0-9

3. Double-ported slide valve (Figs. 158 and 159) —

a^ — r-e, b — \X.o\% inch. ^== r - /. r = 2r 4- ^.
\ = 0*5 to 0*6 thickness of cylinder wall.
/=0-8to0-9
82= 1-0 to 1-1

§ 79. The Cut-off can be taken at the following values for valves
having a full stroke or travel : —

THE MAIN ENGINES.

In the H.p. cylinder about 60 to 75 %.
» M.P. „ „ 55 „ 70%.

The work in the different cylinders can be equalised, as required, by
adjusting the respective links and gear.

Figs. 158 and 159.

In compound engines, with cranks at 90°, it is advisable to have
a maximum cut-off of at least 60 %, because the engine can then be
started by admitting live steam to the receiver. If the cut-off is less,

168 MARINE ENGINES AND BOILERS.

it is necessary to admit this steam to the top and bottom sides of the
low-pressure piston. (See § 51.)

§ 80. Linear Lead. — By this term is meant the port opening (in
inches), when the piston is at the dead point ; this varies from yV in
small to lj\ inch in large engines. The lead is generally twice as large
at the bottom as at the top end of the cylinder, in order that the cut-off
may be as uniform as possible at both ends. The length of the con-
necting rod would otherwise cause it to be from 1 to 2 % less at the
bottom than at the top end. In a valve working at its full stroke, the
following values may be taken for the lead : —

In the H.p. cylinder, top end, — . to ^

14 o

99 99 91

bottom „ ^ to^g

>»

M.P. .. top „ ^ to ^3

,, ,, ,, •^Vfk.t.VFai* ,,

bottom .. ^5 to Jig

" L.P- » top » {2 ^°V8

I) )) 99

bottom „ F-o ^o o-b

5-8 3-3

The following angles a between the position of the crank at which
the admission port begins to open, and the dead point, correspond
approximately to the linear leads given above : —

In the H.p. cylinder 4" to 10"* at the top end.

w

»»

S'^ „ 13"

bottom

M.P.

91

5- „ ir

top

>9

»

9' „ IS"*

bottom

L.P.

»

6"^ „ 12^

top

»

9)

lO'' „ 18"

bottom

It should be noted that the lower values refer to slow-running engines with
speeds of from 70 to 80 revolutions per minute, and the higher to those
running at 350 revolutions per minute and upwards. For engines running
at other speeds, values between those given above should be selected.

§ 81. Exhaust Lead. — This may be smaller with slow speed than
with high speed engines, and is usually —

7 to 14 % of the stroke in the h.p. cylinder.

" 99 1" /o >» »> M.P. ,,

11 „ 22 % „ „ L.P. „

It is also desirable that this lead should be 15 to 20 ^ more at the

» '

THE MAIN ENGINES. 169

bottom than at the top end of the cylinder, so that the steam, during
the rapid change at the top of the stroke, can get away at the bottom
quickly enough.

§ 82. Compression. — The degree of compression is so arranged
that the final maximum pressure shall not be greater than the pressure
in the valve chest. It varies from —

4 to 8 % of stroke in the h.p. cylinder.

7 ,, 14 % „ „ M.P. „

10 „ 20 % „ „ L.P. „

It may be taken as somewhat higher in compound than in triple engines,
because in the former the difference of pressure in each cylinder is
greater. Like exhaust lead, or inside lap, compression should be greater
at the bottom than at the top end of the cylinder. On each side of the
piston the one is directly dependent on the other, a small exhaust lead
or inside lap producing a high degree of compression. The two must
therefore be made to correspond with each other, for any given working
conditions.

§ 83. Valve Diagrams. — To determine the dimensions of the
gearing driving the valves by means of eccentrics, the Miiller-Reuleaux
or Zeuner valve diagrams are generally used. A description of these
diagrams will now be given. It should be noted that it is not
necessary to consider the length of the eccentric rod, as this is generally
made so long in proportion to the stroke of the eccentric, that the effect
of the length of the rod may be neglected.

• •

§ 84. Muller-Reuleaux Diagram for ordinary D Slide

Valves. — Here—

a denotes breadth of port in valve face.

tfo and ^u denote outside lap at top and bottoq[i respectively.

'o » 'u » mside ,, ,, „ ,,

^o »» ^'u » linear lead

*o n ^u )9 cut-off in the cylinder „
As a rule, when designing the valve gear, the breadth of port a, the linear
lead of the valve top and bottom, and the cut-off are given. The amount
of eccentricity is then determined, and taking it as the radius, the valve
circle is described from the centre o. (See Fig. 160). Let aob be the
travel of the valve. Draw circles round the two dead points a and b with
radii equal to the lead of the valve at the top and bottom of the cylinder
respectively, and plot the cut-off ac as a percentage of the stroke of the
valve for one side of the piston, say the top side. Then if an arc be
drawn through c with radius equal to the length of the connecting rod,
bearing the same ratio to the amount of eccentricity as the length o*f the

>> >> j>

170

MARINE ENGINES AND BOILERS.

main connecting rod does to the crank, it will cut the circle of the
valve at point 4. o4 will then represent the position of the crank
corresponding to the cut-off on the top side of the cylinder. Through
point 4 draw a tangent to the lead circle at the top of the stroke, and
draw a parallel line touching the lead circle at the bottom of the stroke.
If a third parallel line xv be then drawn through the centre o, and o

TOP

B O T T O

Fig. 160.

joined to the points where the first two lines cut the valve circle, oil
will represent the position of the crank corresponding to the cut-off at the
bottom end, and the angular lead S = the angle adv. The outside lap of
the valve is represented by e^ and e^. The inside lap can be measured
in the same way from the central line xy. If the inside lap is positive,
it will be on the opposite side to the outside lap for the same end of the
piston ; if it is negative, it will be on the same side.

THE MAIN ENGINES. 171

Principal Posiiions of the Crank, (See Diagram, Fig. 160.)

Position 1. Admission, top end.

,, 2. Crank at top of stroke, port open by amount equal to
the lead = v^,

3. Maximum port opening, top end.

4. Cut-oif, top end.

5. Beginning of exhaust, top end.

7. Valve in middle position.
14. Beginning of compression, top end.

8. Admission, bottom end.

9. Crank at bottom of stroke, port open by amount equal
to the lead = v^,

10. Maximum port opening, bottom end.

11. Cut-off, bottom end.

1 2. Valve in middle position.

13. Beginning of exhaust, bottom end.

6. Beginning of compression, bottom end.

99
•»

H
>»

If

The diagram also shows —

AC : 2r = Co = cut-off, top end.
BD : 2r=€u= „ bottom end.
BG : 2r = exhaust lead, top end.
AE : 2r= compression, „
AF : 2r= exhaust lead, bottom end.
BH : 2r = compression, „

§ 85. Zeuner's Valve Diagram (Fig. 161).— This diagram is
constructed in the following way : — Let ab be the direction of the travel
of the valve, and o the centre of the diagram. Describe a circle with
any given radius about o. Plot off from a the cut-off ac = €„ x ab, and
draw the crank position o3 corresponding to this cut-off for the top end,
taking the length of the connecting rod into account. Crank position
ol is then drawn, representing the point where the port opens on the
top side. The angle lo3 is bisected. The line bisecting it forms with the
line MN (diameter of the circle ab) the angle of advance 3 = angle mo2.
With half the stroke of the eccentric as radius, the eccentric circles are
then drawn through o, having their centres on the radii o2 and 06 respec-
tively. For any given position of the crank ox, the line ov cut by the
valve circle will be the distance of the valve from its middle position. The
portions op = OQ of the lines representing the crank positions o3 and
ol are respectively the outside lap, and the distance Rs represents the
lead of the valve at the top end. Let bo be the exhaust lead at this

172

MARINE ENGINES AND BOILERS.

end, then the crank position obtained from it (taking the length of the
connecting rod into account) will be oi, and the corresponding inside
lap /'o = OT = ou.

Crank position o5 is that in which compression begins at the top
end. If the inside lap ox = ou of the valve (top end) is positive, the
circle drawn with centre o, taking the inside lap as the radius, must
lie in the lower valve circle (as shown) ; if negative, it will lie in the
upper valve circle, as indicated by the dotted line. In the latter case,

N

Fig. 161.

crank position o iv will correspond to the exhaust lead, and crank
position o v to compression on the same side of the piston. The
diagram for the bottom end is drawn in the same way.

§ 86. Variations in the Cut-oflf.— To vary the cut-off in a
cylinder with link motion, the link itself is generally used when the
cut-off is not less than about 45 */^, and the variation required is only
from 15 to 20 7o* ^^ ^ smaller cut-off is required, a separate expansion
valve is generally provided (Meyer's or Rider's), but these valves are

THE MAIN ENGINES.

173

174

MARINE ENGINES AND BOILERS.

seldom used in marine engines, and need not therefore be further
considered here.

§ 87. Stephenson's Link Motion.— In this form of motion
there are two eccentrics, one for the ahead and the other for the
astern gear. The amount of eccentricity and the angle of lead are

TOP

I:
O T T O M

Fig. 167.

determined from the valve diagram for the maximum cut-off, and are
as a rule the same both for the ahead and the astern gear. As
regards the method of connecting the eccentric rods to the link, a dis-
tinction is made between open rods (Figs. 162 and 163) and crossed
rods (Figs. 164 and 165). When both eccentrics are so placed that
their eccentricity is towards the link, the difference between the two
methods is clearly seen.

THE MAIN ENGINES.

175

If the gearing lies fully over to one side, only that eccentric which
has the upper end of its rod in line with the head of the valve rod need
he considered. If the link is only partly over to one side, both eccen-
trics act, their effect being to reduce the cut-off, and vary, the lead,
compression, and exhaust lead. With open eccentric rods these will be
increased, and with crossed eccentric rods they will be diminished.
In both cases, if the link motion is only partly over, the maximum port
openings will be reduced, but, for the same positions of the link, the
port openings will be larger with open than with crossed rods.

The various functions of the link motion, between the two full-
over positions, may be approximately determined by graphical methods
as follows : —

In Fig. 167 let ab be the path of the valve rod, od the amount of
eccentricity, cod = 3, the angle of advance. Through d describe the arc
of a circle df having its centre on the line ab, and a radius x which can

be calculated from Fig. 166 by means of the equation ^ = — '- — - — '—,

2ef

If the eccentric rods are open (not crossed), the concave side of this arc

is turned towards the central point o as shown ; if they are crossed,

the convex side will be turned towards o, as shown by the dotted lines.

Upon the arc df the lengths dDj and DDg calculated from dd^ = dd.2 =

ES

DF X — are then marked off. od, will then be the amount of eccen-

EF

tridty, and 5j the angle of advance of an eccentric under these conditions,
which corresponds to the farthest point s reached by the motion of the
valve rods with open eccentric rods. oEj will be the position of the
crank when expansion begins. The lead, exhaust lead, and compression
are determined by the method already described. If the rods are
crossed, ODg will be the resultant amount of eccentricity, and Sg ^^^
corresponding angle of advance.

The corresponding positions of the crank are : —

^m _

. _ . .

._

"

Full Gear.

Link Motion.
Open Rods.

OGj

Link Motion.
Crossed Rods. '

' Admission

OG

OG2

. Beginning of expansion

OE

OEi

OE2

„ exhaust lead -

OH

OHj

OH2

„ compression -

1

OK

OKj

OK2

Various Types of Valve Gear.

The following systems of valve gear, as well as Stephenson's link
motion, are frequently used in marine engines : —

§ 88. Klug Valve Gear.— The motion of the valve is obtained
by an eccentric lying in the same direction as the main crank (see Fig.
168). The centre line of the eccentric rod motion ou is as a rule laid
down at right angles, or nearly so, to the piston and valve rods. If it is
necessary, from want of space, to arrange it otherwise, the eccentric
must be at the same angle as the crank, and the motion must be trans-
mitted from the eccentric rod to the valve rod by a suitable arrange-
ment of levers.

In Fig. 168, o is the crank shaft, u the reversing shaft, ok the crank,
OEj :*= r amount of eccentricity. The usual proportions of the parts of
the valve gear are —

ou = from 5 to 7r.

Eccentric rod EjU = „ »Jt^ + (ou)^ to ou.

AiU = about 0-5 EjU.
cu = y^ ^ to ^r.

The angle of throw a of the reversing shaft from its centre is generally
about 15° to 20*". By making this angle smaller the cut-olf is diminished,
and compression and exhaust lead increased ; by reversing the process,
opposite conditions will be obtained.

The lead of the valve will be constant if e^u = sjt^ -H ou^.

The distribution of steam can best be determined by the use of a
small model ; but, failing this, it can be worked out graphically.

The curve described by the end point a of the eccentric rod may
be drawn for different positions of the reversing lever uc. The circle
described by the eccentric is then divided into eight to twelve equal
parts, and for each of these positions the relative position of point a is
determined. By joining these various points, the desired curve for any
given position of the reversing lever uc is found. If from positions Aj
and Ag of the end of the eccentric rod, corresponding to the dead points
of the crank, the leads v^ and v^ respectively be plotted parallel to the
path of the valve rod, and an arc drawn through these points, with ad

THE MAIN ENGINES.

177

e

CX^tcn^t'

Fig. 168.

as the radius, then, as m corresponds to the middle position of the
valve—

^o will be = outside lap at the top end.

'a »

»

i9

bottom end.

M

178

MARINE ENGINES AND BOILERS.

V-—

Fig. 169.

Fig. 170.

The inside lap of the valve is determined from the positions assigned to
the crank or piston at the beginning of the exhaust lead, at the top and
bottom ends respectively. Further, the points on the eccentric circle

THE MAIN ENGINES. 179

corresponding to the positions at the end of the eccentric rod are as
follows : —

Ag beginning of admission at the top end.

A^ „ expansion „ „

Ajo „ admission „ bottom end.

A^2 M expansion ,, „

From these points, by reversing the construction of the diagram, the
corresponding positions of the crank and of the piston respectively can
be found. For instance, position of the eccentric e^, at which exhaust
begins at the upper end of the cylinder, corresponds to position a^ of
the end of the eccentric rod. If an arc be drawn through a^ with ad
as radius, it will cut the curve on the other side at the point a^j, corre-
spK)nding to position e^ of the eccentric, at which compression begins
at the top end. The distance between the two arcs drawn through m
and A4 gives the inside lap /'o- The inside lap i^ for the bottom end is
found in the same way. It should also be noted that the inside lap is
negative, if it is on the same side as the outside lap for the same end of
the cylinder. The valve motion is reversed by moving the shaft, and
the lever connected to it, from position uc to position uf.

§ 89. Marshall's Valve Gear.— This is very similar to the Klug
valve gear, except that the valve connecting rod is not at the end of the
eccentric rod, but comes in between points b and e. (See Fig. 169.) The
eccentric forms an angle of 0° or of 180° with the crank. The steam
distribution is determined in the same way as has already been fully
described for the Klug valve gear.

§ 90. Joy's Valve Gear. — Motion is here communicated to the
valve from a point on the main connecting rod. (See Fig. 170.) The
movement of the valve is regulated in the same way as in the Klug or
Marshall gear. It is best determined from a small model, or by drawing
the curve described by the point f. The usual proportions of the
parts of this valve gear, as shown in Fig. 170, are —

^1^-0-20 to 0-23

K^K

A= = 1.5to

KW

1-65

KW

„ 1-5

5_^. = 0-2

KW "

0-3

«^ -0-5

AC

„0-6

^^-0-8 „

KW

1-0

CD _ 1.4
KW

„ 1-5

KW

2-5

180

MARINE ENGINES AND BOILERS.

At the top and bottom of the stroke, points e and m coincide, fro?)
which it follows that the lead of the valve is constant for each position

I

Fig. 171.

of the reversing lever mg. By increasing the angle of throw of this
lever the cut-off is increased, and the exhaust lead and compression

THE MAIN ENGINES. 181

diminished, and vice vers&. The engine is reversed by shifting the
lever mg into the position mGj. The angles a and a^ may be about 15**
to 20\

§ 91. Heusing^er Valve Gear. — In this valve motion the valve is
also driven by an eccentric on the crank shaft, and by a lever connected
to the crosshead of the engine. (See Fig. 171.) The eccentric works on
one end of a slot link which can rotate or swing about its centre. The
radius of the link being equal to the length of the valve connecting
rod, the link is in its central position when the piston is at the top and
bottom of the stroke. The upper end of the valve connecting rod is
fixed to a lever, connected to the valve rod, and also at the other end,
through a short link, to the main crosshead. The motion of the engine
is reversed by moving the end of the valve connecting rod from one end
of the slide link to the other. This is done by a reversing shaft common
to all the cylinders, which shifts the link block by means of a reversing
lever and link, in the same way as in the ordinary link motion. The
cut-off may be varied by moving the link block nearer to the centre of the
reversing link. The lead is constant for all positions of the slide block.

Piston Rods.

§ 92. Maximum Load. — The so-called maximum load is generally
taken as a basis for calculating the strength of the following parts,
vizi, the piston, piston rod, crosshead, connecting rod, and crank shaft.
This load p = area of the h.p. piston x boiler pressure. The rods of all
the cylinders in compound engines are calculated on this basis. They
are all made alike, to avoid the multiplication of spare parts, and to
promote interchangeability.* If the rod is driven by an l.p. cylinder
with an h.p. cylinder above it, the maximum load is taken at p = area
of H.p. cylinder X pressure in boiler (above atmosphere), +area of l.p.
cylinder x maximum pressure of steam entering l.p. cylinder in pounds
per square inch (absolute). The highest admission pressure in the l.p.
cylinder in the triple-expansion engines here considered may be taken
at 50 lb. per square inch (absolute), and in quadruple-expansion engines
at 30 lb. per square inch (absolute).

Example, — In a quadruple-expansion engine let each of the two
middle cranks be driven by an l.p. and an h.p. cylinder placed "tandem,"

* The same interchangeability is provided for when the h.p. cylinder rests upon
one of the other c}'linders.

182 MARINE ENGINES AND BOILERS.

while the m.p. cylinders are arranged over the two outside cranks. Dia-
meter of the H.p. cylinder (800 mm.) = 30 inches; of the l.p. cylinder
(2400 tnni.) = 90 inches. Pressure in boiler (15 atmospheres), 200 lb. per
square inch. The maximum load on the rods in pounds will then be —

p = (-78 X 30! X 200) + (-78 x 90= x 30) = 337,940 lb. = 1 50-8 tons.
This load p forms the basis for calculating all the four sets of rods, &c.,
including those of the m.p. cylinders.

§ 93. General Remarks. — In most lai^e engines, and in ihe main
engines of all large warships, the pistons are made of cast sled. In

.-^

vessels where weight is no object (cargo steamers, &c.) they are made
of cast iron, and in very light ships (torpedo-boats) oi forged steel.

S 94. Cast-Steel Pistons. — Typical shapes are shown in Figs. 172
to 174. Thickness of boss, d= 1-5 to Vlk. Xlk should be the value
used for small pistons, and those of engines where weight is no object ;
\-hk for large pistons and light engines. Height of boss, h= \\k.

The thickness of the steel piston in the middle t is obtained from
the formula i = KXi-. The value of k is given in Table No. 19; the
coefficient c is taken at —

c=\ for flat pistons, or those in which the inclination

measured on the inside is very slight (angle
0° to 6°).
cmO-85 to 0-95 ,, slightly coned pistons (angle 6' to 1 8°).
i:=0-75 „ 0'85 „ medium coned pistons (angle 18° to 28°).
t = 065 „ 07.^ „ strongly coned pistons (angle 28° to 35°).

THE MAIN ENGINES.

183

Saftfy

Nut

The thickness a near the outer circumference is obtiuned from the
equation a = 0-45i to 0'55i, It is best not lo make the piston quite flat,
but slightly conical. In engines with several cylinders the dimensions
/, A, and * (Fig. 172) are
generally made the same
for the H.P., M.P., and
L.p. cylinders. If an
inclination of I : So to
1 :6-5 be taken for the
internal slope of the
largest piston, normal
proportions for all the
pistons will be obtained.

In Table No. 19 the steam pres-
sures are to be taken as absolute
admission pressures, and the values
given cannot be used for pistons
which have an internal slope of
more than 35° to the horizontal. Fig- 173.

Table No. 19.

Thickness of Steel Pistons. Values ofy,.

(Pressure in pounds per square inch absolute.)

Pnunni,

Pkuius,

P™hu«.

_

PreBUKS,

tMimcletof lb. p«

lb.™'

ib.^'

lb.p.r

[b.p«'

tb.ptr

Cylinder. Kt- inch.

•4. inch.
■Dtotft.

S3','°^:

^■A

ir,-,Sk

,a,'"Si.

.siii.

^■r4

InchB. ' Incho.

iDChu

Incho

Incbo.

iDChu

!nch«.

IlKbu.

liKb..

1 15 to 23 1 1

"t ■

"ig ■

li

v

"2"-

3i

2i

a „ 81 1

H

IJ

2

2*

2*

3

31 ,

.11 ,. 39 1

i|

2

2i

SI

H

3i

4

38 „ 47 1

■2

28

3

31

4

41

47 „ 53 2

2*

2i

H

S!

41

-35 ,. 63 1 2B

■2\

^

H

4i

In""' %
' 79" 8«

3
3*

11

4

4*

*i

■■■

3

31

4j

»6„ 94

3

4

»*„102

3*

..

102 „ no

3B

3|

_'U

g94A. Cast-iron Pistons.— For typical form, see Figs. 175, 176,
179, and 180.

Thickness of boss, d= 1 -5 to 1 -Ik.
Height of boss, k = I -Ak.

184

MARINE ENGINES AND BOILERS.

The thickness of the boss is the same as for steel pistons, because
it is strengthened by the upper and lower wall.

Depth of piston = about height of boss.

Kig. 174.

Thickness 8 of the upper and lower walls and of the ribs should be

from about ^77 + '4 inch to about 77: + *4 inch. The higher values are
60 40

for the H.p. and m.p., the lower for the l.p. piston. The number of ribs

are as follows : —

z =

4

for 12

to

24 inches diameter of

cylinder.

z =

6

»

24

>»

40

>>

))

>i

z =

8

>}

40

»>

60

»

>>

»>

z =

10-12

>>

60

>i

80

i»

»

n

]c^^

i

/

^3:

Fig. 175.
ting the rod into the piston, see " Piston Rods," page 191.

The walls or surfaces
between the ribs or webs
must be sufficiently strong to
resist the maximum absolute
pressure to which the piston
is exposed. If the strain is
too great, the above number
of ribs should be increased.
For different methods of fit-

§ 95. Piston Packing^. — 1. Ramsboitam Rings may be used for all
pressures. The rings are almost always of cast iron. The inner and

THE MAIN ENGINES.

185

outer circumferences are eccentric in regard to one another. The rings
are cut across diagonally at the thinnest part, so that no scoring of the
cylinder can take place where the ends meet. The rings are turned
in the first instance of somewhat larger diameter than the cylinder;
in most cases a cylindrical casting is first made, from which the rings

t

-I:

-M-

Fig. 17tl. Fig. 177.

Tadle No. 20:

Ranubottotn Phlon Rings. (See Figs. 176 and 177.)

Diameli-r

/

20H
25

I{i ,.1

2 to 3
2 ., 3

36A

186

MARINE ENGINES AND BOILERS.

are parted off. A piece is then cut out of the ring, and the ring is
pulled in with a band or soldered together, and turned to fit the
bore of the cylinder. Larger rings are turned in the first instance to
the correct size of the cylinder, and after cutting out the piece, are
fitted into the cylinder, and carefully hammered to a true circle. For
dimensions of the rings, &c., see Table No. 20.

In H.p. cylinders of smaller diameter four rings are used, which need
not be made as deep as the dimensions given in the table. With small
pistons a number of small steel Ramsbottom rings are often used ; say,
about three rings from \ to yV ^"^^ deep, and about fV to j inch
thick for pistons of 7 to 10 inches diameter.

In piston^ with Ramsbottom rings care must be taken that the
piston does not run too far into the coned portion at either end of the
cylinder, so as to allow the outer ring to become jammed.

2. A ring very generally used for l.p. cylinder pistons is the Buckley
J^ing {Fig, 178). A flattened helical steel spring is so placed, behind the
inner slanting surfaces of two packing rings, that it presses them out

Table No. 21.
Buckley Piston Packing. (See Fig. 178.)

a

b

c

d

e

Diameter of
Cylinder.

Depth of
Ring.

Inches.

Thickness of
Ring. 1

1

Length of

Piece
Cut Out.

Thickness

of Ring at

Bottom.

Inches.

!
Inches.

Inches.

Inches. '

20J to 22J

1 7

5

H

15

14

22i „ 2H

2

8

lA

1

24* „ 274

^^

1 1

1 11
lift

ItV

27| „ 30i

21

1 1

"nr

1 13

U •

30i „ 33i

11

TIT

1 7

If

ifV

33i „ 36i

•>1

11

1 15

U

36| „ 39>

•^'y^

11

16

0 1

--Tir

lA

39| „ 43i

n

2^

U

431 ., *7|

•>7

:)
T

21

liV

47| „ 51

3

03

-F

1 1
l¥

51 „ 55

3^

:)
T

*7 '

1 «

55 „ 59

H

91

1 *

59 „ 67

4

2|

lii !

67 „ 75

3i

i

2f

i|

75 „ m\

7

21

lU

82i „ 90i

3f

1

3

U

90i „ 98^

Q7

7

Ql
6^

115

98i and upwards

4

7

H

Ql

0

1

THE MAIN ENGINES.

against the walls of the cylinder, and also up and down against the junk
ring and the body of the piston. The thickness of the rings c is the
same throughout. For dimensions of the Buckley ring see Table No, 21,

>

188

MARINE ENGINES AND BOILERS.

3. Ihe Ptck piston packing is often used for the l.p. and m.p.
cylinder pistons. It consists chiefly of two cast-iron rings, in several
sections, lying one above the other, which are pressed outwards against
the sides of the cylinder by springs. One section in the upper and one
section in the lower ring are riveted together, so that the steam cannot
blow directly through the joint. Behind these sections lie bronze rings,

against which the springs bear. (Compare Figs. 179, 180, lefthalf.) Above
the cast-iron rings is a thick covering ring with turned grooves. Thick-
ness of cast-iron segments, yV to J inch. This packing is used chiefly
for pistons of more than 4 feet diameter. In the latest types of the Peck
piston, the springs are replaced by a self-adjustiug cast-iron ring.

For both high and low pressures the follow-
ing packings, which are not self-adjusting, are
used : —

1. Packing Ring, with Single Hal Springs.

—Behind the flat, cast-iron piston ring (of

which there are sometimes two) lies a flat steel

spring, which is made thinner towards either

end. This type of packing is much used for

donkey engines. (See Figs. 181 and I8:i.)

2. Packing Ring, with Multiplt Flat Carriage Springs. — This old

method of pwcking the piston, which is very convenient for adjusting

and for repairs, has the disadvantage, that the springs are in constant

Fig. 182.

THE MAIN ENGINES.

189

IE

J

motion while the engine is running, and therefore apt to wear them-
selves, as well as the body of the piston, away. The height of the
cast-iron rings is about the same as that of the Buckley ring ; for high
pressures the thickness of the rings should be somewhat more. (See Figs.
179, 180, right half.)

General Remarks on Piston Packings, — To prevent steam leaking
past them, the smaller and thinner rings are often cut, as shown in Fig.
183. larger rings are fitted with small plates overlapping the joint
(Fig. 180, right hand). These are generally riveted to one end of the
ring, and then fitted on to the other end. To reduce the wear and risk
of firacture to a minimum, the piston ring
should project as little as possible beyond
the body of the piston. With small
pistons of about 3 feet diameter, it is
sufficient to turn the body of the piston
about -^ to ^^ inch smaller than the

bore of the cylinder ; but with pistons of larger diameter, the allowance
should be from iV '^ f i^^ch.

§96. Clearance between the Piston and the Top and
Bottom Covers of the Cylinder. — As a rule this space is made larger
between the piston and the bottom cover, than between the piston and
top cover, because the wear of the bearings causes the piston, in course
of time, to drop slightly. A larger clearance is allowed for unmachined
cast-iron pistons than for turned cast-steel pistons. Sometimes the
covers (especially if of cast steel) are turned on the inside; in any case it is
advisable, if the cover and body of the piston are not turned, to check
or test them in the lathe, in case there should be any undue thickening
on one side, which would necessitate their being machined during
erection, in order to give the necessary clearance.

Fig. 183.

Table No. 22.
Piston Clearances.

Diameter of Cylinder.

15
24
39
60

78

Inches,
to

»

»9

}>

»9

24
39
60

78
98

Above 98 inches

Clearance at Bottom.

1
T

A

1

9

11
TF
1.1

Inch,
to

»>

n

»

n

}i

■ff
1

Tir
11

TV
l:»

r
■ff

Clearance at Top.

(Thickness of Packing for

Jointing Cylinder Cover

mcluded. )

Inch.

to

1
4

1

:)

4

F

\

i

u

6

H

IT

s

1 1

¥

1«

')! ^Q 111

„ „ 39 and over, „ li „ U „

190 MARINE ENGINES AND BOILERS.

§ 97. Thickness of Junk Rings.— This should be such that
the junk ring, when removed (Fig. 182), is strong enough not to bend or
break. If the junk ring is made of cast iron, a thickness of 1|^ to 1?
inch is sufficient in pistons of less than 3 feet in diameter ; in larger
pistons the thickness should be 1 J to 2^ inches. If the ring is of steel,
the thickness may be less ; but as the bolts are generally made with
countersunk heads, there is not much saving in weight by using steel.
Holes should be tapped in these rings for starting pins and lifting bolts.

Junk Ring Pins, — These are generally made in the form of tap
bolts. The thread of the screw is as strong as possible, and the heads
are either square or hexagonal. Cheese-headed bolts must not be used,
as the frequent screwing up and unscrewing spoils the slits.

Size of Junk Ring Pins.

For cylinder diameters of from 15 to 24 inches, about |^ to 1 inch.

w4 ,, 39 „ ,, 1 „ 1^

39 and over, „ li „ H

Their pitch should be —

For the h.p. cylinder = 5 to 7 times their diameter.

,, M.P. ,, = 0 ,, o ,, ,,

,, L.P. „ = I ,, L\J ,, ,,

To prevent the set screws of the junk ring rusting into the body of
the piston, they are often screwed into gunmetal plugs, which are them-
selves screwed into the piston with a fine thread. Some simple means
must be used to lock the bolts, such as a screw let in close to the
head, to prevent the latter slacking back. Sometimes a brass washer is
placed below the head of the set screw, which, as it expands with the
heat, keeps it tight.

§ 98. Remarks.— All set screws and lock nuts on the piston must
be very strong, and of such dimensions that they do not wear with the
constant screwing and unscrewing. The piston rings must be a perfect
fit between the junk ring and the body, especially if the surface they
present to both is small. The rings must also be a steam-tight fit
against the walls of the cylinder. To draw the piston off the rod, two
large tapped holes are made in the body. Studs are fitted into these,
by means of which a bar or clamp can be screwed down on to the
top of the rod, and the piston slacked off and lifted from it. Pistons
which fit very tightly can often only be drawn off from the rod by
putting a stout pipe between the piston and the crosshead nut, and
heating it. As it expands the piston is forced from the cone. To lift

THE MAIN ENGINES. 191

the junk rings from the inside of large cylinders, which are accessible
throiigh manholes, tapped holes are sometimes made in the cylinder
covers for eyebolts. This is especially necessary if, above the larger
cylinder, there is a high-pressure cylinder, which would otherwise have
to be dismantled every time any repairs had to be carried out in the
lower cylinder.

§ 99. Piston Rods. — In large engines the crosshead is not, as a
rule, forged in one piece with the piston rod ; in small engines, however,
and where it is necessary to economise weight, the two are forged in
one. In auxiliary engines the piston and the rod are also sometimes
forged in one piece.

Separate piston rods are usually fitted into the crosshead and piston
with coned ends, and held in place by a nut. At the piston end
the cone or taper is 1 in 5 to 1 in 7 when it is not necessary to take out
the rod frequently (merchant ships, &c.) ; 1 in 3 to 1 in 4 when the rod
has to be removed from the piston to examine the latter (engines below
the armoured deck of an ironclad, top piston of tandem engines, &c.,
and where the piston rod and crosshead are in one). Under these
conditions a taper of 1 in 3 to 1 in 4 is used at the crosshead end,
because it is at this end that the piston rod is most frequently removed.
Cylindrical ends are also sometimes met with ; if they are very accurately
fitted, the result is satisfactory.

The diameter of the rod itself is made slightly larger than the largest
diameter of the tapered ends, that the latter may not be injured during
erection or tearing down. At the largest end of the tapered part there
is usually a small collar, against which the piston rests, and which
prevents the boss from splitting, or the piston from eventually becoming
loose, should it be strained owing to water in the cylinder. (See
Fig. 184a.)

Fig. 184a.

The nuts of the piston rod are hexagonal, or are sometimes circular with
two slots cut in them. A strong split pin, or a small set screw fitted
against one side of each nut (see Figs. 180 and 172a) prevents it from
slacking back. To prevent the piston turning on the rod, the larger end
of the cone is generally fitted with a small key or set pin.

Dimensions of Fiston Rods, — The diameter below the thread at either

192

MARINE ENGINES AND BOILERS.

end of the rod must be such that the maximum stress on the area at
bottom of thread shall not exceed the following : —

5,500 to 7,000 lb. per square inch for cargo boats.
7,000 „ 8,500 „ „ mail steamers.

8,000 „ 10,000 „ „ warships.

10,000 „ 12,500 „ „ light cruisers and torpedo-boats.

The material used is assumed to be Siemens-Martin steel. In the
rod itself the stress should only be about half the above figures, and, in
this case, the breaking stress need not be taken into account. The
diameter of the rod is about —

^=^i + yi for small rods.

d=d^x—^foT large rods, as an allowance for future turning down.

Tail J^ods, — In large pistons the piston rod is frequently carried
through the cover, in order that the piston may be well guided. The
diameter of the tail rod is generally made slightly less than the diameter
below the thread on the piston rod.

Piston rods are occasionally secured to the piston by a flange. In

Fig. 184b.

such cases the coupling bolts must be very accurately fitted, and the nuts
must be well locked. The advantage of this arrangement is the ease
with which the piston can be taken off, against which must be set the
difficulty of preventing the bolts from slacking back. (See Fig. 184b.)

Material. — The material used for the piston rod is almost exclu-
sively Siemens-Martin steel, and sometimes crucible or nickel steel (for
warships). It should be fairly hard, so that the rod does not wear
quickly, or get ridged at the ends.

Connecting^ Rod and Crosshead.

S 100. The Length of the Connecting Rod is almost always
made as great as possible, so as to reduce the pressure on the guides
to a minimum. The ratio of the length of the connecting rod to the

radius of the crank is scarcely ever less than 4 to 1 ; but the restricted
height of marine engines seldom allows of a larger ratio than 4-5 to 1,

194

MARINE ENGINES AND BOILERS.

The most usual types are shown in Figs. 185 and 186. In engines for
merchant vessels, and frequently in those for warships, the crosshead
bearings are on the connecting rod, the pin being fixed to the crosshead
itself. If the crosshead and piston rod are forged in one piece, the
bearing is generally in the crosshead, and the pin forms the upper

Fig. 186.

end of the rod. (See Fig. 186.) This latter arrangement is frequently
used in warships, and in small engines for steam pinnaces, auxiliar)*
engines, &c.

§ 101. Connecting Rod. — The diameter of the rod just below
the fork is generally taken as equal to the diameter of the piston rod.
The diameter of the shaft produced to the centre of the crosshead pin
(Fig. 187) will therefore be 5^ = about 0*75 k, k being the diameter of
the piston rod. The shaft is made thicker at the bottom end, firstly,
to make the change of form from the large bottom end less sudden ; and

THE MAIN ENGINES.

195

secondly, to secure greater strength. Fairly correct values for the larger
diameter 5^ are obtained if 8a = about O'^d^,, Connecting rods are occa-

Fig. 187.

sionally made hollow, but this method of construction may prejudicially
affect the solidity of the fork.

§ 102. Connecting-rod Fork. —
This is generally the weakest and most
dangerous part of the connecting rod,
and its proportions should at least be
calculated approximately as follows : —

The cross section at y (Figs. 187
and 188) has to take the greatest pres-
sure and bending strain. The greatest
pressure to which it is exposed will be —

p
s = .ix

+

2 y^g y^xg ''iyxg

(■-?)

Fig. 188.

The first equation is the direct

tensile or compressive stress produced

p f

by the maximum load - ; the second equation gives the greatest tensile

or compressive stress produced in the outer edges of the cross section y

p
by the moment - xp. The stress on any cross section .r of the cross-

196 MARINE ENGINES AND BOILERS.

p
head fork may be calculated in the same way. The maximum load -

p' • . p'

yields a component — , which gives rise to a bending moment of — x/

IS A

about the section x^ and at the same time sets up a tensile or compres-
sive stress in this section. The greatest tensile or compressive stress
upon the section x is thus —

p'

^"2 ^^x^"*"^2~^

The shape of the fork must be determined by calculating the section at
various points. If Siemens-Martin steel is used, and s = 5,500 to 11,000
lb. per square inch, the usual dimensions for the fork will suffice, the
higher stresses being used for very lightly built engines. With small rods
s is generally = 7,000 to 10,000 lb. per square inch where weight has not
to be economised, but there are rods in use both in very large and small
engines, in which the combined compressive and bending stresses are as
much as 14,000 lb. per square inch.

Crosshead Pin, — Length of pin bearing = about l^xits diameter.
The dimensions may be determined from the pressures given in Table
No. 23. The pin must of course also be sufficiently strong to resist
bending stresses. It may be" made hollow, if saving of weight is of
importance. It is fitted into the fork by hydraulic pressure, or shrunk
into it while the fork is hot. It is usual to key the pin in the fork by
means of a strong set screw, as shown in Figs. 186 and 187. To fit
the pin in more easily, it is made from ^^^ to -J inch larger at one end,
where it is held by the fork. The dimensions of the fork are as follows
(Fig. 187) :-

/// =

about 1 '%d to

2</.

/o =

»

l-5rf „

■2d.

^o =

»)

t)-6a'.

The width g is slightly greater than the diameter of the shaft at the same
place ; the thickness y is calculated to resist the tensile, compressive,
and bending stresses as given above.

§ 103. Crank-pin Brasses. — The dimensions of these brasses are
determined by the maximum allowable pressure on the bearing surfaces.
(See page 181.)

THE MAIN ENGINES.

197

Table No. 23.
Maximum allowable Pressure on the Crank-pin and Crosshead Brasses,

Type of Vessel.

Crank- pin Brasses,

lb. per

square inch.

Crosshead Brasses,

lb. per

square inch.

Cargo boats - . - -
Mail steamers, &c.
Fast steamers, ironclads
Small cruisers, &c.
Torpedo-boats and destroyers

*

350 to 500
425 „ 570
570 „ 700
700 „ 850
850 ,,1000

700 to 1000

850 „ 1000

1000 „ 1250

1250 „ 1700

1700 „ 2100

The form shown in Fig. 185 is the one most commonly used for the
connecting-rod head. For pins of large diameter the brasses are made
of cast steel or bronze (the latter now only in warships) with white metal
linings, and for small connecting rods they are often of gunmetal without
any white metal, especially in the case of the crosshead bearings.

In Fig. 185, let

^ = fxl-3to>4 = fxl-4
2 2

for cast-steel brasses lined with white metal, or for gunmetal brasses
where no white metal is used ; and

^ = ^xl-35to/4 = fxl-5
2 2

for gunmetal brasses lined with white metal, the higher values being
taken for smaller, and the lower for larger pins.
The thickness of the white metal should be —

w = 1 J to J inch if ^=: 20 to 24 inches.

_ 6

r ^

I 1

>» iir

= 15

))

20

_ 0

= 12

9)

15

= i

1)
» Iff

= 8

n

12

.. i

= 6

• •

8

= i

= 3

)}

6

_ »

» i

= less than 3

To prevent the loosening of the white metal, circumferential dove-
tailed grooves are left in the blocks, and similar shallow recesses round
each edge keep the white metal from working out. The inner surfaces
of the blocks are tinned before the white metal is run in. Packing

198

MARINE ENGINES AND BOILERS.

pieces from ^ to 1^ inch thick, according to the diameter of the pin,
are placed between the brasses, and very thin strips of metal from 25
to 35 S.W.G. thick are also inserted, for* adjusting the brasses. The
"play" allowed when fitting them is about yinny ^^ iftf ^^^ ^^-
meter of the pin. The crank-pin brasses are allowed a lateral play
of from ^f to ^ inch on each side. In order to prevent the brasses

from moving, they are secured in the head of the
rod and in the cap by strong central pins, or
(which is less desirable) by a slot parallel to the
centre line of the pin (Fig. 189).

Circular-shaped Brasses. — These are often
employed for the heads of connecting rods (see
Fig. 186), and are rather lighter than the brasses
described above, and shown in Fig. 185. If
made of gunmetal their thickness should be as

above, h = --. When round brasses are used, the

V

Fig. 189. packing strips between them must be very strong,

to prevent the brasses from turning. If there are
set pins for this purpose in the crown of the brasses, it is very difficult
to fit them accurately into the connecting rod.

§ 104. Connecting-rod Bolts. — These are placed as close to-
gether as possible, and in order not to weaken the bolt a fine thread is
generally used. (See Table No. 71, Part VII.) The nuts have the
Penn locking device (Table No. 72, Part VII., page 608). For small
rods lock nuts are used with split pins. To afford greater facility in
fitting up and dismantling, the nuts are always placed at the top, and
the head of the bolt at the bottom or cap end.

The maximum stress on the bolts at the bottom of the thread should be

5,000 to 7,000 lb. per square inch for cargo steamers.

7,000 „ 8,500 „ „ passenger ships and ironclads.

8,500 ,,11,000 „ „ light war vessels.

A feather prevents the bolts from turning; those for the crank-pin end
have tapped holes for eyebolts, so that they may be more easily removed.
The crank-pin bolts are made a tight fit in the head of the rod and in
the cap, but are given a little play in the brasses. The nuts are gene-
rally stamped with graduations, so that it may be seen at a glance exactly
how far they have been screwed up.

Bearing Caps, — The usual dimensions for these can be arrived at if
they are considered as a beam having an equally distributed load between

THE MAIN ENGINES.

199

the two bolt centres. Let p be the maximum load, equally distributed,
then the bending moment will be (Fig. 190) —

M =

P. 2a p . fl

Modulus of cross section —

z =

d,s^

p . a 1

X -

4 s

The safe stress is taken at s = 5,500 to 8,500 lb.

per square inch, the higher value being adopted

for light engines. Width of the cap (see Fig. 185) <^u = about 8^ for

the bottom end, and A, = (1*5 to 1-7) xy for the top or fork end.

If the crank pin be very long, ^u may be made wider, to avoid

weakening the head of the rod.

§ 105, Crosshead and Guide. — In large engines for merchant
vessels the crosshead consists, as a rule, of a square steel block with
forged pins at each side. The
guide blocks are generally made
separate of cast iron or cast steel,
and are secured in position by
suitable bolts^ the piston rod being
connected by means of a cone and
nut. In small engines, and those
in which weight has to be econo-
mised, the piston rod, crosshead,
and guide shoe are forged in one
piece, and the latter faced with
gunmetal or cast iron.

Where the crosshead and
piston' rod are not forged in one
(Fig. 191), 2 = 2 X >ti = double the
diameter of the screwed end of
the piston rod over the threads.
The height h and the width /
are determined from the bending
strain at the centre of the cross
section. The bending moment
at this point is —

Fig. 191.

p // . s\
2^2 ■'2) =

__P(/+3)

200 MARINE ENGINES AND BOILKKS.

THE MAIN ENGINES.

201

m m.

n\ . i„ ^^ -^P-n

!S

ta
(^

-7"^^ w/ ui w ui w w

202

MARINE ENGINES AND BOILERS.

The maximum stress upon the central section will therefore be-

p(/+s)

s =

(f-k,W

Fig. 196.

The value of s may be taken at 7,000 to 10,000 lb. per square inch,
according to the type of engine.

As a rule /should not be made less than TH x ^j ; the height h being

THE MAIN ENGINES.

203

taken as large as possible. The length of the pin at each side is gene-
rally made equal to its diameter, />., /=^. The pin has no fillet on the

-^ V

N.^^

N

Fig. 197.

outer end, because as little play as possible is allowed between the
crosshead bearings and the body of the crosshead. The pins are
slightly flattened on the sides, so as to form a recess to retain the

204

MARINE ENGINES AND BOILERS.

lubricant, and thus ensure more efficient lubrication at the top and
bottom of the pins. Very large engines having four guide columns are
usually fitted with four slipper blocks. This is a good arrangement, as
it facilitates inspection (Figs. 192 to 195).

•

Method of fitting the Slipper Blocks to Crossliead, — The shoes must be
capable of being easily taken down, and yet be so well fitted that there
is no possibility of their working loose. As the faces wear, the slippers
must be capable of being easily adjusted. This is usually done by fitting
thin strips of metal between the body of the crosshead and the slippers,
or by providing a wedge-shaped piece between them, which can be
drawn up. The arrangement shown in Fig. 196 is often used. The
slippers are slipped in from below, and secured by plates and a few
strong set screws. They are also held at the sides by set screws. The
slippers are made of cast iron or cast steel. The bearing surfaces con-
sist of flat strips of white metal held in position by slightly wedge-shaped
grooves, or of plates of white metal cast on, and secured to the slipper
by lateral dovetailed grooves. Of the two, the first arrangement is
preferable. (See Fig. 192.) In large engines side plates are generally
fixed to the slippers, to guide them laterally. These plates have the
further advantage of preventing the white metal from being forced out.
They are generally secured in position with countersunk set screws. (See
Fig. 196.)

Fig. 198.

Fig. 199.

g 106. Crossheads Forged in one with the Rod.— These

are employed in nearly all small marine engines, and in large engines
where weight has to be economised. The bearing is generally in the
crosshead, and the pin is held in the forked end of the connecting rod.
The crosshead has generally only one slipper guide on the ahead side,
and plates overlapping the sides of the shoe take the thrust when going
astern. The shoe and crosshead are generally forged in one piece, and

THE MAIN ENGINES. 205

the slipper is made of cast iron or guntnetal, screwed on to the shoe with
a shoulder above and below, so that the two faces can be accurately fitted
tt^ether. As a rule the slippers
are faced with while metal.
The back of the shoe is often
provided with fitting strips, to
prevent any cutting of the guide
plates when running astern.
Figs. 198 to 200 show an ar-
rangement of crosshead which
is often used. The dimensions
of the brasses and cap bolts are
similar to those given for the
connecting rod in § 104. The
width at X (Fig. 199) is made
as large as possible, in order
that, when the inner brass is
driven into position, it may not
throw the piston rod out of line

with the slipper face. The Fig. 2(K).

brasses are made of gunmetal

lined with white metal, and both these are made the same thickness as
for the connecting rod. (See S 100.)

§ 107. Pressure on the Guides. — If ^ is the ratio of the radius

of the crank to the length of the connecting rod, the pressure on the
guide, which is taken as a basis for calculating the dimensions of the
latter, will be —

Pressure on the guide = ^ x maximum load.

§ 108, Guides- — The surface of the guides may be determined
from the following permissible pressures. That is —

5-5 to G5 lb. per square inch for cargo and slow-running passenger

steamers.
65 to 80 lb. per square inch for mail steamers.
70 „ 85 „ „ ironclads and large cruisers.

85 „ 120 „ „ small cruisers and torpedo-boats.

The width of the guide face depends on the arrangement of the sup
porting columns. As a rule —

206

MARINE ENCIIfES AND BOILERS.

Cast iron is the most suitable material for the guides ; for this reason,
in all engines which have no cast-iron columns, special cast-iron guides
have to be fixed to the columns. It is only in smalt engines that the
front of the column itself can be utilised as the guide. An engine can
be more easily erected, when separate guides are provided.

In large engines the guides are almost always internally cooled with
water, and are therefore made either as separate chambers (Figs. 197 to

Kig. ail.

r%. 2ia

201), or as the front wall of a hollow recess in the column (Fig.
302), They are secured by bolts or set screws with countersunk
heads. A few oil-ways are cut diagonally across the guides, and there
is a lubricator at the top and an oil receiver at the bottom, out of
which a comb on the slipper picks up the oil, and smears it over
the guides. The water for cooling them is generally admitted at the
bottom, and discharged at the top. Ihe guides are often fitted in

THE MAIN ENGINES.

207

such a way ihat the front and back faces, when cold, are rather nearer to
each other at the top than at the bottom, to allow for expansion of the
cylinder when hot. In order to hold the piston and crosshead in their
highest position when the connecting rod is being laken down, two
large tapped holes are made in the guide, and set screws fitted in to
hold an iron plate which supports the crosshead. The arrangement of
the guides usually adopted in engines entirely supported on steel or
wrought-iron columns is shown in Figs. 1203 and '204. In all engines

^-•S^-

Fig. 2C«.

in which weight ha.<; to he minimised, back plates are fitted, instead of a
separate astern guide. These back plates are of cast iron in small, and
of steel or gunmetal in larger engines. Either the crosshead or the back
plates are faced with bronze or white metal, so that steel does not run
upon steel, as this is likely to cause abrasion.

The guide plates are always fitted so that they can be removed from
either side as the .strips wear, without having to take down the whole of
the crosshead, piston, and rod. New gunmetal strips can then be fitted
on to the slipper.

208

MARINE ENGINES AND BOILERS.

P

If - is the pressure upon the working faces of the guide plates,

the strain upon the bolts holding them will be ? x ?-l- (see Fig. (204).

As the guide plates are not made to resist much
bending strain, there should be as many set
screws or bolts, per unit length of the slipper,
throughout the length of the guide, as will
prevent the permissible strain being exceeded.
The method of fixing the guide plates to the
guides must be such that, in taking down the
plates, the guide itself is not completely dis-
connected from the framing of the engine.
Some bolts in addition to those holding the plates

to the guide should be provided, to secure the main guide to the

framing.

Fig. 204.

Crank Shafts.

§ 109. Crank Shafts. — The turning moment of the crank,
corresponding to the maximum load on the engine, is generally
taken as the basis for calculating the torsional stresses in the crank
shaft. If the crank shaft is built up of several sections, each section
must be of the same strength as the end section, which has to take
the total torsional stress.

The mean turning moment (m) of the shaft, in inch pounds, is —

M = L!Lf: X 63,000

where i.h.p. is the indicated horse-power and n is number of revolutions
of the engine. From this moment, the diameter (d) of the shaft, if
solid^ is calculated as follows : —

0^* = — X— and d = 1w2 x^/ —
TT s V s

and for hollmv shafts —

the diameter of the bore being = d.

THE MAIN ENGINES. 209

The allowable torsional stress is usually as follows ; —

s = 4,000 to 4,500 lb. per square inch in engines for cargo and

passenger steamers,
s = 5,000 to 5,500 lb. per square inch in engines for heavy war vessels,
s = 5,500 „ 6,500 „ „ „ light „

These allowances for stress may appear small, but should not be
exceeded, in view of the various and exceptional strains upon the crank
shaft, such as bending, if the bearings do not wear uniformly, racing of
the screw in a rough sea, &c. If the maximum turning moment (as
shown in the tangential diagram of forces at page 63) is taken as the
basis for calculation, the above stresses can be increased in the ratio
of I to from I "2 to I -5, the higher figure being used for engines with
one crank, and the lower for two, three, or more crank engines.

Holiaiv Shafts. — Shafts are made hollow to lighten them, and to
distribute the material more efficiently. It is usual to take - = about

D

0* 1^ to 0*6. After the shaft is erected, the mouth of the bore should
be closed up, so that the inside of the shaft may not become rusty, and
also to prevent water reaching the bore, and getting between the
coupling flanges and corroding them.

o

§ 110. Lloyd's Rules for Determining the Sizes of Shafts

(July 1902).

" Rule 60. The diameters of intermediate shafts are to be not Jess
than those given by the following formula : —

For compound engines with two cranks at right angles, diameter of
intermediate shaft in inches =

(•04a + OOGd + -028) X Vp

For triple-expansion engines with three cranks at equal angles,
diameter of intermediate shaft in inches =

(•038a + -0098+ -0020+ -01 65s) x Vp

For quadruple-expansion engines with two cranks at right angles,
diameter of intermediate shaft in inches =

(•034A-f01lB-»--004c + -0014D-»-'016s)x Vp

For quadruple-expansion engines with three cranks, diameter of
intermediate shaft in inches =

(•028a + -01 4b + •OOec + -001 7p + -01 5s) x ^/p

For quadruple-expansion engines with four cranks, diameter of inter-
mediate shaft in inches =

(•033a + OIb + -0040 4- "OOISd + 01558) x Vp

Where a = diameter of high-pressure cylinder in inches.

B = diameter of first intermediate cylinder in inches.
c = diameter of second intermediate cylinder in inches.
D = diameter of low-pressure cylinder in inches.
8 = stroke of pistons in inches.

p = boiler pressure above atmosphere in pounds per square
inch.

The diameter of crank shaft, and of thrust shaft under the collars, to
be at least f J^ of that of the intermediate shaft. The diameter of thrust
shaft may be tapered off at each end to the same size as that of the
intermediate shaft.

THE MAIN ENGINES. 211

The diameter pf the screw shaft to be equal to the diameter of
intermediate shaft (found as above) multiplied by ( '63 + J, but in

no case to be less than 1*07t, where p is the diameter of propeller
shaft, and t the diameter of intermediate shaft, both in inches.

This size of screw shaft is intended to apply to shafts fitted with
continuous liners the whole length of the stern tube, as provided for in
paragraph 33 below (rule for material). If no liners are used or if two
separate liners are used, the diameter of the shaft should be f ^ that
given above.

The diameter of screw shaft is to be tapered off at the forward end
to the size of the crank shaft.

Note, — These rules are intended to apply to two-cylinder compound
engines in which the ratio of areas of low and high pressure cylinders
does not exceed 4*5 to 1 ; to triple-expansion engines in which it does
not exceed 9 to 1 ; to quadruple-expansion engines in which it does not
exceed 12 to 1 ; and in all cases, as regards the stroke, in which the
length of stroke is not less than one-half the diameter, or greater than
the diameter of the low-pressure cylinder. Engines of extreme propor-
tions beyond these limits must be specially submitted to be dealt with
on their merits."

Lloyd's Rules for Materials of Shafts (July 1902).

" Rule 31. Shafts, — All shafts are to be made of good material and
are to be examined when rough turned and when finished. In the
case of screw shafts, scrap steel is not to be used. It is recommended
that these be made of ingot steel, or forged from blooms made from
rolled iron bar of good fibrous quality.

A tensile and a bend test are to be made on pieces cut from one
end of each ingot steel shaft forging, the piece from which they are cut
being of the same' size as the body of the forging. In the case of built
crank shafts, the tests are to be taken from the material of the crank
pins and journals, not from the webs. If more than one piece is forged
from one ingot, one test only will be required from the ingot. The
tensile strength is not to be less than 27 tons per square inch, nor to
exceed 32 tons per square inch. The elongation is not to be less than
30 per cent, in a length of 2 inches, measured on a plain portion turned
not more than J inch diameter. The bend test piece is to be made
1 inch square and must be capable of being bent cold, without fracture,
through an angle of 180° over a radius not greater than \ inch." Cast
steel is practically never used for crank shafts, not even for the cheeks.

" Rule 32. Gauges of an approved description for testing the truth

212 MARINE ENGINES AND BOILERS.

of the crank shafts are to be supplied with all new engines, and adjusted
in the presence of the Surveyor."

" Rule 33. The length of the stern bush is to be at least four diameters
of the shaft. It is recommended that the shaft liner should be continuous
the whole length of the stern tube, and that the after end should be
tapered in thickness, and made watertight in the propeller boss. If the
liner is made in two pieces, the joint should be burned. If the liner
does not fit tightly at the part between the bearings in the stern tube, the
space between the shaft and the liner should be charged or * forced '
with a plastic material insoluble in water, and non-corrosive. If two
liners are used, it is recommended that they be tapered in thickness at
the ends, and that the shaft should be lapped or protected between the
liners. In this case, and also if no liners are used, the diameter of the
shaft should be rl J- of that required for a shaft with a continuous liner."

§ 110a. Board of Trade Rules for Shafts.

"Rule 144. Size of Shafting, — Main, tunnel, propeller, and paddle
shafts should not be passed, if less in diameter than that found by the
following formulae, without previously submitting the whole case to the
Board of Trade for their consideration. It will be found that first-class
makers generally put in larger shafts than those obtained by the
formulae.

For compound condensing engines with two or more cylinders, when
the cranks are not overhung —

s =

; c X P X D-

^ CXD2V W

Where s = diameter of shaft in inches.

d' = square of diameter of high-pressure cylinder in inches, or sum
of squares of diameters when there are two or more high-
pressure cylinders.
D- = square of diameter of low-pressure cylinder in inches, or sum
of squares of diameters when there are two or more
low-pressure cylinders,
p = absolute pressure in pounds per square inch — that is, boiler

pressure + 15 lb.
c = length of crank in inches.
/= constant from following table (Table No. 23a).

THE MAIN ENGINES.

213

For ordinary condensing engines with one, two, or more cylinders,
when the cranks are not overhung —

_ a/cxpxD*^
"V 3x/

3x/
. p_3x/xsg

C X D^

Where s = diameter of shaft in inches.

D*^ = square of diameter of cylinder in inches, or sum of squares* of
diameters when there are two or more cylinders,
p = absolute pressure in pounds per square inch — that is, boiler

pressures- 15 lb.
c = length of crank in inches.
/= constant from following table.

Table No. 23a.

For Two Cranks,
. Angle between Cranks.*

1

For Crank, Thrust, and
Propeller Shafts, t

For Tunnel Shaft.

90''
lOO**
110^
120'
130**
140"*
150**
160**
170^
180*"

1047 +
966
904
855
817
788
766
751
743
740

1221
1128
1055
997
953
919
894
877
867
864

For three cranks,
120''

1110

1295

* When there is only one crank the constants applicable are those in the table
opposite 180*.

t The constants in this column should l)e reduced by 15 7o when dealing with
propeller shafts of new vessels. The portion of the propeller shaft which is forward
of the stem gland, and all the thrust shaft, with the exception of the part enclosed in
the thrust bearing, may be of the same diameter as the intermediate tunnel shafting.

X For paddle engines of the direct -acting type, multiply constant in this column
suitable for angle of cranks by 1*4."

e

v/aM/////t

V//M y/ZA

I

'ZIM/jYAWA .

*— t!K^^5!^T^

I

§ 111. Crank Pin. — This is usually made of the same diameter as
the crank shaft ; or sometimes rather larger, in view of the pressure on

its surface. The length of the
crank pin is determined from
the pressures given for " connect-
ing rods" on page 197. Long
crank pins should, if possible, oe
avoided, as they diminish the
rigidity of the crank shaft. In
the first place, the bending stress
on the crank pin and crank shaft
is greater (Fig. 205) ; and, in the
second place, the torsional stress
on the after crank cheek a (Fig.
206), which is carried round with
the shaft and the forward cheek
B, is increased. In very quick
running engines (for torpedo-
boats) the pins are generally very
long ; for, although the force of the piston working on the rod is great,
the diameter of the shaft, on account of the high speed, is relatively
small. The length of the pins can be reduced if they are made some-
what thicker than the shaft, and weight can be saved by making them
hollow. The ratio of the length to the diameter of the crank pin is —

Fig. 205.

A , E

4ii>/-T — L J-

3-

Fig. 206.

d

- = 1*2 to 0-9 in engines for merchant vessels.

1-3
1-4
1-6

j»

T)

J>

M
1-2
1-4

»»

n

M

heavy war vessels,
light cruisers,
torpedo-boats.

§ 112. Built-up Crank Shafts (see Figs. 207, 211, and 212).—
These are used in all cases where weight is not the first consideration,
and therefore nearly always in merchant vessels, but they are seldom
employed in warships. Width of the cheeks (Fig. 207), d = 1*9 to Id
(occasionally d = 1 '^d) ; and thickness of cheeks, w — 0*6 to 0-7//. The
end of the shaft, where it fits into the cheek, is generally strengthened

THE MAIN ENGINES.

215

by making the diameter //^ so that space may be left for a small fillet
where the shaft rests in the bearing, and a gradual transition made
between the cheek and the shaft —

^1 = about </+77r
The distance (z) between the crank pin and the shaft should not be

Fig. 207.

less than 0*45 to 0*50^, otherwise the pin and. the shaft might wear
loose. The stresses produced in the material round the pins by driving
them in are doubled in the section 7v x z, so that if z is too small, the
material between the pins may be unduly stretched. The ends of the
crank pin, which are shrunk into the cheeks, are left of the same diameter,
or only very slightly thickened. The crank pin and shaft are fitted into

k

3A-

Fig. 208.

the cheeks while the latter are hot ; the cheeks are therefore bored out
previously to a diameter oi d^- ^J-- . Sometimes the pin and shaft are

forced into the cheeks by hydraulic pressure, and in this case also, the

d
pin, before being driven in, is made ,- Ax larger than the hole. The shaft

ends are keyed to the cheeks by flat keys or round pins ; it is better to

216

MARINE ENGINES AND BOILERS.

have two small rather than one large key or pin, and to put them off the
centre line (Fig. 207), so as not to weaken the section between the crank
pin and shaft. They are driven into the cheeks to a depth of about 07 d ;

the diameter, if pins are used and there is only one, is = — + 0*39 inch

for shafts of 4 inches diameter and upwards. The crank pins are
sometimes keyed in a similar manner, but it is unnecessary.

In some crank shafts the crank pin and cheeks are forged in one
piece, and only the shaft ends are fitted in. The disadvantage of this
arrangement is the difficulty of securing a sound forging and of handling
it, but with it a saving in weight and space is obtained. It is unusual to
have crank shafts in which the coupling flanges are let into the cheeks
(Fig. 209). The space for the flanges can only be spared at the cost

!

\

^ » . ^

^!rn

~r ~ -
1

1

-U — -

1

—

_ .

1

.J

«-

Fig. 209.

of considerably weakening the cheeks. At the most, this arrangement
is permissible for the front crank and the one next to it ; because the
torsional stress upon them is less than that upon the after crank.

§ 113. Crank Shafts with Crank Pin and Web forged in

one Piece. — These are considerably lighter than built-up crank shafts,
but more difficult to manufacture, and are at the .same time less solid,
because the complicated shape of this class of forging is not conducive
to strength. On account of its lightness, this kind of shaft is exclusively
used in engines for war vessels, yachts, steam pinnaces, and other
lightly-built engines. In torpedo-boats and ships' launches, where the
weight has to be reduced to a minimum, crank shafts are sometimes
forged with all the cranks and eccentrics in one piece ; but the cost of
constructing shafts in this way is enormous, and the manufacturing risks
very great. The strength of the crank pin is the same as that given
above; the thickness of the cheeks is also the same, viz., 7€' = 0-6 to

0-7 if. Width of the cheeks, fi = about ^-f 2 m2 + — V This gives the

THE MAIN ENGINES.

217

depth of the fillet a = about '2 + -^77. The cheeks are generally rounded

off top and bottom, while in the lathe, with centres m or c (Fig. 210).
The outer sides are frequently tapered off as far as the centre of the
crank pin, or of the shaft, but care must be taken not to cut off too
much, and thus weaken the connection between the crank pin or shaft
and the cheek. In the forged crank shafts of warships, both shaft and
crank pins are nearly always made hollow.

Fig. 210.

i^ 114. General Remarks on Crank Shafts.— As a rule in

merchant vessels separate lengths of crank shaft are made for each
cylinder, and bolted together by means of flange couplings. In warships
the whole crank shaft, consisting of two or three cranks, is often forged
in one piece. If the engine has four cranks, they are forged two and
two together. To forge two cranks in one piece has no advantage,
beyond the saving of weight and space due to the absence of the
couplings. The spare shaft is dearer and more complicated to make,
more cumbersome and difficult to stow away. The separate sections of
the crank shaft in the engines of merchant vessels should all be alike
and interchangeable, necessitating only one spare piece of shafting.
Even if the cylinders are at varying distances from each other, this
arrangement can easily be carried out, by having one long and one
short length of shaft in each crank section. (See Fig. 214 for a three-
crank engine.)

For angle of the cranks refer to " Balancing the Moving Parts " (§ 37),
and ** Turning Moment of the Multiple-crank Engine " (S 30). As a
rule the cranks, with two-crank engines, are set at an angle of 90°, with
three-crank engines at 120", and with four-crank engines at 90°, or they
are disposed in accordance with the Schlick system of balancing. (See
"Arrangement of Main Engines," page 106.)

MARINE ENGINES AND BOILERS.

! I

■-S

W—

^

THE MAIN ENGINES. 219

S 115. Crank-Sbaft Couplings.— The various sections of shafting
are boiled together by means of flange couplings. In warships sometimes
easily disconnected couplings are fitted between the aft crank shaft and the
tunnel shaft (see below), but these are now seldom used. The diameter
of the flanges depends on the number and size of the bolts. Thickness
of the flange i- 0-2.5 to 028i^. (See Fig. 207.) The coupling tiolts are

u ^ ,y -^ 111 -^ .1. ^ ^, '^ .1. -^ ^1

_+ 2ji J

IP ^ ^

Y\g. 2U.

generally turned with a taper of 1 in V> to 1 in 25. The mean diameter
is used in calculating the shearing stress. Mean turning moment of
shaft (see page 208)—

M = '"-^. 63,000. in. lb.

\t r is the radius of the bolt circld, the shearing stress will be —

and the area corresponding to the mean diameters of the bolts-

The stress s is assumed to be the same as the torsional stress on the
crank shaft, (See above.) If a large number of small bolts are used,
the flanges may be smaller, and this is the usual arrangement in war-
ships. If there is a small number of larger bolls, larger flanges
must be employed, and this is the general practice for merchant
vessels. As a rule the diameter of the bolts in the middle of the taper

= about - to .. The bolts are tightened up by means of nuts; the

diameter over the thread being made a good deal smaller than the
diameter at the end of the taper, in order that the nuts may not be too

220 MARINE ENGINES AND BOILERS.

large. If thin nuts are used, they are usually fitted with a locking device
and split pin. The holes for the coupling bolts must not only be
accurately bored, but also carefully rimered out after both flanges
have been fitted together. // should he possible to draw t/ie holts from the
couplings without having to take down any other part of the engine^ but it
is often impossible to avoid removing the cap of the nearest bearing.
The coupling flange runs into the shaft with a curve, the radius of which

is about — ; the bolts are set so closely to the body of the shaft that

clearance for the nuts has generally to be made by recessing into the
root of the flange. Concentric recesses are often turned in the face
of the flanges, for inserting centring discs.

§ 116. Materials for Crank Shafts (compare Lloyd's Rules,
page 210).

1. Wrought iron is used only for small, cheap cargo steamers.

2. Siemens-Martin steel, tensile strength 26 to 32 tons per square
inch, with not less than 20 % elongation in 8 ins,, for warships and
merchant steamers.

3. Nickel steel, tensile strength 35 to 40 tons per square inch, with
not less than 20 % elongation in 8 ins., for fast merchant steamers, iron-
clads, &c.

4. Crucible steel for torpedo boats and destroyers, where the greatest
care in construction is required. Tensile strength, 28 to 32 tons per
square inch, and 20 to 25 % elongation in 8 ins.

Valve Gear Rods.
§ 117. Power required to Drive the Valves. — It is diflficult to

obtain accurate knowledge as to the power required to drive the valves,
and the strain that the rods will have to stand. In flat slide valves the
strain depends upon the coefficient of friction, and the varying steam
pressures. In piston valves it depends on these factors, on the pressure
exerted by the packing rings of the valves, and on the fitting of the
valves. In multiple cylinder engines each valve gear is calculated to
transmit the power required to work that slide valve which has the
greatest load on it, the dimensions of the valve rods of the other
cylinders being made the same.

The pressure p (above atmosphere) upon the valve is obtained from
the following table (/j being equal to absolute boiler pressure) : —

THE MAIN ENGINES.

991

>j>j J.

Compound.

Triple.

Type of Engine.

Quadruple.

H.P.

L.P.

crchant vessels "j |

and large war r/i I 0'3^i
vessels I I

Small war vcs- "j
scls and tor- !-
• pcdo-boats j

H.P.

M. P.

L.P.

H. P.

M.P.,

M.P 5

I., p.

Px

to
U-5A

012/,
to

0-2A

/.

()-4.Vi

(»16/i

0-08/ J

Px

0*55/j

0-2o/i

• • •

■ • ■

The coefficient of friction should be taken at —

/=0'15 for the l.p. cylinder.

/=0-2 „ M.P. and h.p. cylinders.

The total area of the back of the valve must be taken as the area
on which the calculation is made, without taking into account any
balancing arrangement which may be fitted, as the rods must be strong
enough to work the valve, should the balancing arrangements for any
cause fail to act. To calculate the strength of the rods a maximum load
Q is taken, which is arrived at as follows : —

L and B being the overall length and breadth of the valve, /and p being
respectively the coefficient of friction and the effective pressure as given
above.

Example, — Let the slide valve of the m.p. cylinder of a triple expan-
sion engine be 50 inches wide and 40 inches long. Boiler pressure
(absolute) = 1 50 lb. per square inch. According to the above table
p = 0'4/j = 60 lb. per square inch.^ Therefore —

Q = 0-2 X 40 X 50 X 60 = 24,000 lb.

If an engine is fitted with pis/on valves only, the dimensions of the
rods are calculated from the diameter at the bottom of the thread of
the actual valve rod, which can only be arrived at empirically. This
diameter d for piston valves is ^=c x n, d being the diameter of the
valve, and

c = 0"ll to 020 in h.p. cylinders — 115 to 215 lb. per square inch

pressure (absolute).
c = 0*10 to 0*17 in M.P. cylinders — 45 to 115 lb. per square inch

pressure (absolute).
c = 0'07 to O'Oll in l.p. cylinders — 15 to 45 lb. per square inch

pressure (absolute).

222 MARINE ENGINES AND BOILERS.

The higher values of c here given are for higher pressures, smali
valves, and big heavy engines ; the tower values for lower pressures, large
vaipesCll to 48 inches diameter), and light engines. The
load Q, used to calculate the remaining parts of the gear for
the particular type of engine in question, is obtained by
multiplying the allowable stress on the area at the bottom
of the thread of the valve rod by this area. (Sec below.)

g 118. Valve Rods.— The valve is supported at the
bottom by a strong collar, and at the top by lock nuts.
Between these two it should be able to work horizontally
to and fro, so that the pressure of the steam may, if It is a
flat D valve, hold it truly against the valve face; or so that the
valve, if a piston valve, may work true in the liner. In order
to get this adjustment the upper nut is fitted with a lock
nut and pin, or a distance piece, composed of a piece of
pipe, is fitted over the rod (Fig. 215).

It should be possible to draw out the rod from below
through the stuffing box, when the rod and tail are foiled
in one piece, or when it cannot be drawn out from above,
as in the case of a ship with an armoured deck. Instead
of the collar mentioned above, therefore, a ring is often
used, fitting tightly on to a cone on the rod (Fig. 'iV)). As,
when setting the valves, this ring may have to be turned down,
to lower the valve, it is advisable to fit a flat ring above it.
The conical ring can then remain intact, and any adjustment
made on the flat ring above it. Care must be taken that
the upper surface of the coned ring, as well as the collar of the
valve bearing on it, are exactly at right angles to the axis of
the valve rod, otherwise the valve will leak and wear un-
evenly. The allowable stress in ilu valve rod at l/u bottom
of the thread (diameter at bottom of thread = i/) produced
by the load Q is —

s = 1,700 to 3,500 lb. per square inch for merchant vessels.

s = 4,500 to 5,500 lb. „ „ warships.

The lower portion of the rod, whicli works in the stuffing
box and guide, is made very thick, partly on account of the
coned ring just described, partly to facilitate the withdrawal
Fig. -215. of ,he rod.

Diameter of rod in the stuffing box is 17 to ISrf.
„ guide is 1-8 „ 20rf.

The taper in the ring below the valve is about I in 6.

THE MAIN ENGINES.

223

In small engines y in which the eccentric rod is connected directly to
the valve rod, the head of the latter consists of a simple eye bushed
with gun metal, which forms a bearing for the pin in the forked end of
the eccentric rod. (See Fig. 215.) This bush is sometimes made in two
parts, which can be adjusted by means of a cotter.

In large engines, and where the head of the valve rod works directly
on the link block, it is similar in shape to the head of a connecting rod.
The brasses are generally of gunmetal, round or polygonal outside, with
strips of metal between, and as a rule without any white metal lining.
Generally the head is forged in one with the rod, but it is sometimes
fixed to it by means of a tapered sleeve and cotter. The allow-
able pressure on the effective surface of the bearing is p = 570 to 1,000
lb. per square inch. With link motion the width of the bearing is
reduced as much as possible, in order that the fork of the eccentric rod
may not have too large a gap. The valve rod should have a strong guide

Fig. 216.

Fig. 217

close to the valve chest, and as near as possible to the link and its
reversing lever, &c. This guide is generally a simple casting of iron
or steel vrith gunmetal bearings, and is sometimes made in two parts.
This latter mode of construction is safer, and enables the rod to be
more easily taken down at sea. At the top end, the valve rod generally
works in and is guided by a gunmetal hood (Fig. 215).

Balance Cylinder Pistons (see Figs. 153, 154). — A balance piston
working in an auxiliary cylinder often acts as the upper valve rod guide.
The diameter of the piston is such that, at the mean working pressure
in the casing, it just supports the weight of the valve. By this means
the strain on the valve rod is nominally the same for the up and down
strokes. The space above the piston is connected to the condenser.
It is usual to fit balance cylinders in connection with very heavy piston
or flat slide valves. The pistons of these balance cylinders are solid
cast-iron discs, fitted either with two or three small steel or cast-iron
Ramsbottom rings, or with one broad packing ring (Fig. 216).

224

MARINE ENGINES AND BOILERS.

To ensure smooth and quiet working of the valves, and to diminish
the strain on the rods, ^^Jofs^^ assistant cylinder is sometimes used.

Like the balance cylinder, it is fitted
to the top of the valve chest. Shortly
before the valve, and with it the bal-
ance piston, have reached their upper or
lower extreme positions respectively,
fresh steam from the. main steam
pipe is admitted above or below the
balance piston. This forms a cushion,
while the strain in the valve rods is
reversed, and helps to work the valve after the change of stroke. The
exhaust steam from the assistant cylinder is led into the valve chest to

Fig. 218.

Fig. 219.

which it is fitted. To balance the weight of the valve, the steam is
admitted sooner to, and discharged later from, the under than from the

THE ^^Ar^■ kngines

upper side of the piston. The assit<Eant cylinder is fitted with a guide
bush having turned circumferential grooves, made as good a fit as

possible on the rod, so as to prevent any steam leaking from the
cylinder into the casing.

226

MARINE ENGINES AND BOILERS.

§ 119. Diameter of Rods. — In valves which may have to be
worked by hand, such as the reversing valves of steering engines, the
valves of reversing gear engines, &c., care must be taken (compare
Fig. 222) that the diameter of the guide rod is the same as that of

B^-J

Wi/.*<i'/4L

I^^rei

Fig. 222.

the valve rod. Otherwise the steam pressure on the difference of the
area between d^ and d.^ acts in the direction of the valve rod, and this
may make any valves that have to be worked by hand difficult to move.

§ 120. Valve Levers and Quadrants.— It is often necessary

to alter the line of action of the motion for working a valve, either
parallel to itself or through a right or an acute angle from the longi-
tudinal vertical axis of the engine ; and this is done by means of two-
armed levers of cast or forged steel. If such levers are used, the
following points must be noted : —

1. The valve rods and levers must be so adjusted that the mean
position of each arm of the lever is perpendicular to the line of motion
of the valve or rod (Fig. 217).

2. The bearing block in the head of the rod, if so fitted, must slide
freely, to allow for the swing of the lever (Fig. 218).

3. The valve rod should be well guided, and, as near as possible to
the point where the lever joins it, there should be a small crosshead
with a guide shoe, or simply a bushed guide block. (See Fig. 219.)

4. If the motion is transmitted by levers* from the reversing link to
the valve rod through an acute angle to the vertical axis of the engine,
care must be taken that the centres of all the pins lie in one plane,
in order to avoid any twisting strains upon the gear.

5. The pin a (Fig. 218), on which the valve rod works, is subjected to
a load Q ; the pin b, to a load Q x t ; and pin c, to a load q( I + ? ) • It fre-
quently happens that two piston valves or two slide valves are worked
from the same reversing link, or that there are two valve rods for one
slide valve. In such cases the rods are connected by a so-called

THE MAIN ENGINES. 227

" cross " beam. This should be most carefully guided, and a little play
given in the eye round each rod, to allow for expansion in the cylinder,
and also for any little irregularities in the construction. The material
for the crossbeam may be either wrought iron, steel, or cast steel. (See
Fig. 220.)

§ 121. Stephenson's Link Motion (Fig. 221).— This consists, in

principle, of two curved links, the mean radius of which = length of the
eccentric rod. Both the links are of rectangular section, and their ends
are held apart by distance pieces ; the link block slides between them,
and the pins for the eccentric and reversing rods are placed outside
the bars.

Material Used, — This may be cast steel, or better still, forged steel.

Distance between centres of eccentric rod pins, <i = 5 to 6 x amount
of eccentricity.

The length of the rubbing surfaces of the quadrant and the space
between the two distance pieces are made to correspond with the full
travel of the quadrant, from the full ahead to the full astern position.
The travel is arrived at by drawing out the path described by the
blocks at the extreme position of the links. The pins for attaching
the reversing links are generally fixed on the quadrants at the " ahead "
position (Fig. 221) in mercantile vessels; but in warships the point of
attachment is frequently at the centre of the quadrants, so that the steam
may be evenly distributed ahead or astern. The exact position on the
quadrants of the point of attachment of the reversing lever is often
dependent upon the actual design of the levers.

The link block consists of a pin for the head of the valve rod, and
two slide blocks, the whole being generally forged in one piece. The
slides are fitted with gunmetal faces above and below (Fig. 221). The
pin itself is kept as short as possible, that the forked ends of the
eccentric rods should not be too wide.

Allowable pressure on the link block pin = p = 550 to 1,000 lb. per
square inch.

Allowable pressure on the faces of the link block = p = 220 to 350
lb. per square inch for merchant, 350 to 600 lb. per square inch for war
vessels.

Each of the two quadrant link bars is of such strength that,

when half the load of the valve rod ^ is thrown upon its centre, the

bending strain is within moderate limits. (See Fig. 223).
The section of the sliding block is worked out from —

s

228

MARINE ENGINES AND BOILERS.

The following are the values of permissible stress : —

s = 3,500 to 8,500 lb. per square inch for merchant vessels.
s = 5,500 „ 10,000 „ „ war

Also ^ = 2*5 to 3^, from which, with the above data, the values of h
and b may be determined.

Eccentric Rod Fins, — These are generally forged in one with the
link bar, and made without outer collars. The pressure on these pins
should be from 650 to 1,050 lb. per square inch. Diameter of pin =
about 1*2 to 1*3 x length of pin. The pin or pins for the reversing links
are placed either in the centre of the quadrant, or may be simply an
extension of the eccentric rod pins, but fitted with outside collars. The
pressure upon them is calculated from the power transmitted through
the reversing levers and links. (See Fig. 220, and page 251.)

^:

g >k '^

- i

h

1 "

•

i

Fig. 223.

Fig. 224.

So/id Slotted Links, — These are still largely used for the engines
of small screw steamers, and almost universally for engines of paddle
steamers. The dimensions are determined in the same way as those of
the double-bar quadrants. It will be apparent from its construction
that the slot link is not suitable for transmitting heavy loads. The pins
for attaching the reversing links may be fitted to the middle or to one
end of the link in the same way as described above. (See Figs. 224
and 225.) The link block is generally formed of a single forging, the
sliding surfaces of which are sometimes faced with white metal, or fitted
with special strips.

§ 122. Eccentric Rods. — The bearings are generally similar in
design to those for the connecting rod, and bushed with gunmetal,
but without any white metal. The brasses may be square or round
(Fig. 226). The fork of the "ahead" driving rod is usually made
symmetrical, and that of the "astern" driving rod more or less one-
sided. Sometimes the "astern" rods are joggled just above the

THE MAIN ENGINES.

229

eccentric straps, thus making their forked ends also symmetrical. For
calculating the dimensions of the forks see S 102, page 195. When-
ever it is possible, the ahead and astern running eccentric rods are both
made the same size, to avoid the necessity of carrying more than one
spare rod of each sort. The rod is secured to the eccentric strap by a
flange and two set screws or pins ; in large engines the rod and the
upj)er part of the strap are seldom forged in one piece, owing to their
size and the difficulty of erecting them. Round rods are simplest, but
flat rods are generally used in the engines of warships.

The allowable stress in the eccentric rod immediately below the
fork is 1,500 to 3,500 lb. per square inch.

The cross section at the bottom of the rod = about 1 '8 to 2 x cross
section at the top (for long rods). Cross section at the bottom of the
rod = about 1 '4 to 1 -6 x cross section at the top (for short rods).

Fig. 225.

Length of the Eccentric Rods, — They should be as long as possible,
as there is usually plenty of space for guiding the valve rod above
the link.

'J'he allowable stress in the fork bearing bolts, and in the set screws
attaching the rod to the strap, is usually 2,000 to 5,500 lb. per square
inch, according to the type of ship.

§ 123. Eccentrics and Eccentric Straps.— As a rule the

eccentric or eccentric sheave is made as small in diameter as possible,
and placed, if it can be arranged, on the shaft ; but often it is impossible
to avoid putting some or all the eccentrics on the shaft couplings.
The ahead and astern ecceptrics are either made of the same width,
or the latter may in some cases be made slightly narrower. In other
respects the same rule holds good as for the other parts of the valve
gear, viz., that all the parts should as far as possible be of the same size,
because they are cheaper to make, and fewer spare parts need be carried.
In smaller engines the ahead and astern eccentrics are often cast in one
piece. The eccentric sheaves are nearly always in two halves, and the

MARINE KNGINES AND ROILF.RS.

I

bolts bolting them together are made as strong as possible, and at least
as large as the eccentric strap bolts. The bolts are secured by nuts, or,
if room for these Is wanting, by cotters. (See Figs. 226 and 227.)

THE MAIN ENGINES.

2:51

Materials for Eccentric Sheaves, — In cargo boats both halves are
made of cast iron, especially if the eccentrics are large, and the shaft
small. This happens especially in engines where some of the eccentrics
are fitted on the couplings and others on the shaft, but where all the
straps are of the same size. In other cases the larger half is made of
cast iron, the other half of cast or forged steel. Cast steel is now often
used for both halves.

Fig. 227.

Size of Eccentric at Minitnum Cross Section (see Fig. 230). — s y,b —
a X cross section of both eccentric bolts.

For cast iron a = 3*5 to 5.
cast steel a = 1*8 to 2 '5.
forged steel « = 15 to IS.

*>

11

As a rule the two halves of the eccentric have a groove and tongue on
the jointing face, that they may be turned with accuracy. The eccen-
tric must be most carefully fitted to the shaft, to preclude any possi-
bility of its working loose. The key is made very wide, and is generally
fitted in the crown or centre of the larger half of the eccentric. A set
screw is added, to prevent the eccentric from moving sideways. Some-
times, with the same object, the ahead and astern gear eccentrics are
bolted together, so that the key of the one eccentric acts as a slop to
the other eccentric. This arrangement also distributes the strain over
both keys.

232 MARINE ENGINES AND BOILERS.

The diameter of the eccentric having been determined from the cross
section at the crown, the width is determined from the maximum pres-
sure on the bearing surface, which may be taken at
from 70 to 140 lb, per square inch, according to the
weight and space available.

The greater the circumferential speed of the
eccentric, the lower the value selected for the allow-
able pressure between the surfaces, other conditions
• being equal. The higher the speed of the working

Fig, 228. surfaces, the greater the friction per unit of time, and

consequently the greater the chance of overheating.
1 To prevent the eccentric strap slipping sideways,

,, f , the face of the sheave is either made conical from

I I I either side inwards, or projecting edges are turned

either on the eccentric or on the strap. The draw-
back to the latter arrangement is that the edges are
apt to wear away, and the eccentric or strap then
becomes useless (Fig. 229).

g 124. Eccentric Straps. — In small vessels these

I are often made of gunmetal ; in large vessels almost

Vie. 229. universally of cast steel with white metal linings, and

occasionally of cast iron with gunmetal linings. The

latter are hea^y, and not durable. In very light warships the straps

are often made entirely of wrought iron or steel, and sometimes the

rod and top strap are forged in one piece, as in locomotives. For

the connection between the top strap and the rod, see S 122, page 228.

The top half of the strap, where it is bolted to the rod, must be

thick enough in the middle not to bend or pull away from the sheave.

A, Fig. 230, should iherefore be as deep as possible.

Moment of resistance per unit stress at A (Fig. 230)^

^i,xA,_Q^^ 1

S being = 4, 2.50 to 8,-500 lb. per square inch for wrought iron or steel
or cast steel, and= 1,140 to 4,2-50 lb. per square inch for cast iron. In
some engines S is as much as 14,000 lb. per square inch for cast steel,
but with such a stress the eccentrics are apt to run hot, because the
side or fork of the strap tends to close in. If the upper half of the
strap is strong enough to resist this action, the lower half need only act,
more or less, as a. connecting band. As a rule this half is made thicker
at jj, Fig. 230, near the crown, than at the sides.

Eccentric straps of cast steel, whose cross-sectional area in the lower

THE MAIN ENGINES.

Pig 23<l

half is double that under the thread of the two strap bolts, will be found
strong enough. Generally—

I, X ^j = 2 to 2'5f for cast steel.

J, X (*,=.! -8 to 2f for wrought steel.
F being the area under the threads of the two strap bolts.
The eccentric strap bolts are put as close together as p

2U

MARINE ENGINES AND BOILERS.

i

JH

^

though it entails an increase in their length. The bolts should be large

enough to withstand the load of the valves, q, and the resulting stress

must not exceed 2,100 to 5,500 lb. per square inch
(according to the type of vessel). They should be
fitted with lock nuts or ring nuts and pins, a fine
thread being advisable. (See Table 72, page 609.)

Gunmetal fitting strips are fitted in the joint be-
tween the top and the bottom halves of the eccentric
strap. They are usually made so that they can be
taken out from the sides, the metal inside the bolt
holes being cut away for this purpose (Fig. 231).
Steady pins should also be provided.

As with all bearings fitted with white metal, the

white metal must be held in position at each side by a projecting edge,

and also by dovetailed grooves or slots.

§ 125. Concluding Remarks. — Only Stephenson's link motion
has been here considered; but the different parts of other systems of
valve gear may be designed in the same way, the load upon the valves
Q being taken as a basis.

Fig. 231.

Bed-plates.

§ 126. The Bed-plate consists of as many transverse girders as there
are bearings, these transverse members being connected by longitudinal
members. In smaller engines (up to about 1,000 i.h.p.), for cargo or

Fig. 232.

passenger steamers, tugs, pinnaces, &c., the whole bed-plate is cast in
one piece, even when there are several cranks ; the cross section of the
transverse members is 7" shaped, _J or ^_J L shaped, and the

THE MAIN f^NGINES.

235

cross section of the longitudinal members | or J |^_ shaped.

Thickness of the metal of the bed-plate (5), if of cast iron, is —

8 = - + -5 inch {d being the diameter of the shaft in inches).

0\J

Fig. 233.

Fig. 234.

In engines for large merchant vessels a separate bed-plate is made
for each crank. The bed-plates each contain two transverse members,
and are bolted together by means of flanges. (See Figs. 232 to 234.)

236 MARINE ENGINES AND BOILERS.

The cross section of the transverse and longitudinal members is
channel-shaped thus _] \__. The thickness of the metal, if of cast
iron, is, as before —

B = ^ + -5 inch.

The longitudinal members on one side are generally replaced by the
condenser, which is bolted to the bed-plate. (See S 131, page 277.)
The thickness of the flanges for cast iron, allowance being made for

planing the casting afterwards, is —

S, = 1-9 to 28.

The flanges are connected by strong fitted bolts.

In the engines of modern quick-running steamers the bearing frames
are often made of cast steel. As it is not desirable to have very com-
plicated castings in steel, these transverse members are made separately
and joined by distance pieces. The cros.s section of both is made

1^ shaped. In the case of steel castings, large openings are

provided in the sides, to enable the cote to be rapidly removed after
casting, on account of the rapid shrinking of this material. (See Figs.
102, 105, 107.) For thickness of cast-steel bed-plates, see below.

In modtrn warships the framing to carry the bearings is always
made of cast steel in separate castings, and these are joined either by
distance pieces, or, preferably, as is gener-
ally don^ , by two side pieces running from
end to end, having a section shaped thus
L or thus C (Figs. 235 to 239). The
engine of a small cruiser shown at Fig.
90, has a bed-plate of the first-named
type. The photograph is taken in the
erecting shop; the cast-iron supports
which are placed below the cast-steet
cross girders, only serve for the erection
of the bed-plate, and are replaced in
the ship by the actual engine seating.
The longitudinal girders at the sides are
of cast steel, or, in torpedo-boats or
Fig. 235. Qjher light craft, of wrought or rolled

steel. The cross section of the trans-
verse members for the bearings is shaped thus r~| . and made
much deeper in the middle than on the outside ; firstly, because of
the extra strength thus secured ; secondly, because, when cramped for
headroom for the engine, the shaft must lie as low as possible in the
ship ; and lastly, because the plating forming the engine seating under

THE MAIN ENGINES.

237

5^

238

MARINE ENGINES AND BOILERS.

ex

THE MAIN ENGINES. 239

the side bearers should he as large, strong, and deep as possible (see
S 359, page 616).

Thickness of the cast-steel bed-plates, 5 = = ; + 6 inch.
„ „ flanges, 6, = 1GB.

Id merchant ships the height of the centre of the shaft above the under
suit of the bed'Plate, in steel and cast-iron bed-plates, is such that in its
lowest position the cap of the connecting rod does not come below the
bottom of the l>ed-plate. The latter either rests on fitting strips of hard
wood, and the under side is then left rough, or it rests upon cast-
iron or steel fitting strips, in which case the under side of the hed-plate
is usually planed tip.

The foundation bolls generally jjass through the fitting strips, the
spaces between the Utter being filled in with hard wood.
Diameter of foundation bolts —

J to 5 inch for sha^s iinder 4 inches diameter,
y „ 1 „ „ from 4 to 8 inches diameter.

1 .. H ■■ ,. „ 8 „ 16 „

U „ Ij ■■ „ „ 16 „ 24 „

In bed-plates with flanges running all round them the distance of the
Ijolis from each other is about ten times their diameter.

§ 127. Holding-down Bolts. — The holding-down bolu must be
sufficiently strong to withstand the upward vertical component of the

moving parts (see page 60), allowance, of course, being made for the
weight of the engine. They must also hold the bed-plate in position
firmly enough to withstand the horizontal component of the moving

240 MARINE ENGINES AND BOILERS.

parts, and the tendency to slide, due to the weight of the engine, when
the ship rolls up to say 15** from the vertical. (See Fig. 240.) The
holding-down bolts are usually roughly and carelessly taken out, and
should therefore be as strong as possible. In engines whose centre
of gravity lies very high up, it may so happen that when the ship rolls
the vertical line dropped from this centre falls beyond the bed-plate,
and the bolts must then be strong enough to prevent the engine tilting
over on one edge of the bed-plate. In the engines of torpedo-boats,
side projections are sometimes provided on the bed-plate to counteract
this tendency. (See Fig. 2ol.) If the bed-plate is immediately above the
water tanks in the double bottom, the joint between the bolts and the
top plate must be made watertight. Either the head of the bolt pro-
jecting into the tank must be carefully packed, or a nut must be fitted
in between the top plate and the flange of the bed-plate (Fig. 241).

Definite rules cannot be laid down for the thickness of the longi-
tudinal and transverse members of the bed-plate, because this depends
in the main upon the solidity of the engine foundation. The transverse
members must always be of sufficient strength, so that, when the maxi-
mum load is thrown upon their centre, the bending strains set up in them
are very small. As a rule their height is such that, as mentioned above,
the cap of the connecting rod in its lowest position does not come below
the bottom of the bed-plate.

§ 128. The Longitudinal Bearers, together with the longitudinal

plates of the engine seating, must be strong enough to withstand the
turning moment of the engine without appreciable distortion. (See
j5 35, page 82.) Sufficient strength to resist these forces can only
be obtained by having a strong foundation, and a good connection
between the upper ends of the columns. For this reason, in warships
especially, the longitudinal members of the bed-plate are low and small,
and serve principally, during the erection of the engine, as distance
pieces for staying the transverse members, the engine seating itself
affording the requisite longitudinal rigidity.

§ 129. Main Bearings. — For the diameter of these, see § 109,
page 208. The length is determined according to the allowable
pressure upon the bearing surfaces, due to the maximum load.

The allowable pressure on the main bearings should not exceed
the following : —

Ordinary freight steamers - - - 200 to 225 lb. per sq. in.

Passenger boats and quick-running steamers - 225 „ 300 „ „
Ironclads, large cruisers - - - 250 „ 350 „ „

Small, light cruisers - - - - 350 „ 400 „ „

Torpedo-boats, steam tugs, &c. - - 400 „ 550 „ „

THE MAIN ENGINES.

241

In engines with several cranks it is assumed that the maximum load
of the H.p. cylinder acts upon each crank, and the total length of the
main bearings is such, that the maximum load, divided by the total pres-
sure on their surfaces, does not exceed the allowable pressure given in
the above table. The main bearings are generally all made of- equal
length. The ends of the brasses approach the cheeks of the crank
as near as possible, but sufficient play must be allowed to prevent the

Fig. 242.

cheeks fouling the brasses as the thrust block wears or the crank shaft
expands with heat.

Brasses, — As a rule the lower brass is made round at the base, and
the upper one square on the top. In large cargo and fast passenger
steamers the lower brasses are frequently made with square bottoms, while
occasionally in warships both upper and lower brasses are made round.

Square Top and Bottom Brasses (Figs. 242 and 243). — Both brasses
should be fitted perfectly tight into the bearing block. If this is done

o

242

MARINE ENGINES AND BOILERS.

there is less fear of their wearing loose than if rounded brasses are
used. If the bearings are large the brasses are made hollow, and
water is circulated through them.

The material of which they are made is usually cast iron or gun-
metal,, lined with white metal held in place by dovetailed grooves and
recesses. The brasses are tinned before the white metal is run in.
Sometimes the white metal is only fitted into the brasses in longitudinal

Fig. 243.

strips. Between the two brasses intermediate packing pieces or liners
of gunmetal are placed, and held in position by set pins, and under
these again are laid a few thin pieces of metal plate, which are removed
as the wear on the bearing is taken up. To get all the bearings
exactly in line, they are machined up in place after the bed-plate has
been bolted together. The white metal in the bearings is left in the
rough, and both ends of each bearing are fitted w^ith a wooden disc,

I

THE MAIN ENGINES.

243

in the centre of which is a piece of thin metal plate with a very small
hole in it. The metal centres are shifted until a light, placed opposite
the hole in the centre of the shaft at one end of the engine, can be
sighted through all the little holes simultaneously. The hole in the
thin metal plate then forms a centre, from which a circle can be struck
on each brass, for boring out the white metal.

Round brasses have the advantage that they can be taken out with-
out removing the shaft. For instance, if a bearing is working hot, the
lower brass can be turned round and drawn out, the surface scraped,
and the oil channels made deeper. To effect this, the shaft is slightly
lifted, and the brass drawn out with a bent hook (Fig. 245). If both
the upper and lower brasses are round, the packing pieces or liners
between them must be held tight by lugs on the cap, to prevent the
brasses turning round.

Fig. 244.

Fig. 245.

§ 130. Main Bearing Caps. — If the upper brass is round, the cap
is made of cast iron or cast steel ; if square, a thick wrought-iron or
steel plate or slab is generally used. For dimensions see page 245. A
hole about 4 J x 3 inches, large enough to admit the hand, is made in
the cap and the upper brass, through which grease can be applied to
the bearing, and its temperature felt. Over this hole an oil box is
usually fitted with two divisions, one having an opening direct to the
hole, and the other having a worsted syphon-feed arrangement. The
cap and upper brass are often joined by special screws, so that both can
be removed together. Tapped holes are also provided in the cap and
in both the brasses for eyebolts.

§ 131. Main Bearing Bolts.— Of these there are two for small,
and four for large bearings. Where it is possible, bolts are used with

244

MARINE ENGINES AND BOILERS.

nuts at either end, and the lower one is often a square-headed nut,
which cannot be turned (Figs. 242, 243). Less often the bolts have
heads at their lower ends ; in any case they should be as easily remov-
able as possible. Bolts with nuts above and below are held tight in
the block by a collar countersunk in the upper surface under the block.
Care should be taken that the latter is well stiffened where the bolts go
through it. (See Figs. 242, 243.)

With smaller bearings, studs up to about 3 inches diameter are used,
but they have the disadvantage that they are difficult to replace, especially
if they break off below the collar. Both bolts and studs are made with
fine threads (see Tabl.e No. 71, page 608), and in larger engines ring lock
nuts are used for the cap, and double nuts in smaller engines. Ring nuts

Fig. 246.

should be graduated so that the^ can be screwed up and carefully set.
Tapped holes should be provided in the top of the bolts for eyebolts.

§ 132. Dimensions of Main Bearings (see Fig. 246).— Distance
of the bolts from centre of the bearing —

tf = 0*85 tor0'9d, if the brass and bed-plate are of cast iron.

a = 0*75 to 0*85</, if the brasses are of gunmetal and the bed-plate is

of cast steel.

d

Width of the brass
Depth of the brass

- // = 1 -05^ to g.

Thickness of the white metal tv = -20 -h -— . inches.

35

Thickness of the distance pieces s= "20 -I- -— inches,

15

THE MAIN ENGINES.

245

Thickness of round brasses, omitting the thickness of the white
metal —

r=0-07 to 0'09</+ -125 inch for bronze.
r=0-ll „ 012</+-20 „ cast iron.

9)

Thickness of bed-plate below the brasses —

J = 0*20 to 0-28df for cast iron.
j = 0-12 „ 0-16^,, cast steel.

Strength and Size of Main Bearing Bolts, 6r»r. — If the maximum
load on one crank be divided between the bolts in the two adjacent
bearings, the stress in each should not exceed —

s = 3,000 to 4,250 lb. per square inch in merchant vessels.
s.= 4,250 „ 6,500 „ „ warships.

s = 6,500 „ 7,750

>}

>>

torpedo-boats.

If there is one bearing between two cranks, it is generally fitted with
four bolts ; if there are only two bolts, they are made proportionally
stronger.

§ 133. Thickness of Caps. — These are generally taken as beams,
supported at points represented by the centres
of the bolts, with a uniformly distributed load,

the total of which is equal to half the maximum

p
load= -. The bending stress is generally

somewhat less than that allowed for the con-
necting-rod cap. If there is only one bearing
between two cranks, the cap is made Trom f to
^ times as strong as in the other bearings.
Width of the cap ^ = about 0*6 to 0*9 x length
of the bearing.

The depth of the cap is obtained from the above values of s by the
formula —

- X 2tf X 1
_ Bending moment _ 2 ^

s = ■- .- -

Modulus of section

^x^

whence h

Engine Columns.
§ 134. The Arrangement of the Colunms of an engim:

considerably affects its construction, and is characteristic of the type of
the engine.

§ 13.5. 1. Engines for Ordinary Small Merchant Vessels.—

The columns^ are of cast iron, either open and ribbed, or hollow and in
one piece with the guides. Two columns are provided for each cylinder,
or turned columns in front, and cast columns behind. The columns on
one side are frequently cast in one with the condenser, and also some- .
times with the bearings for the reversing gear and for the air-pump
levers. With jet condensers/ even the air pump is sometimes cast
with the columns. The typical form is shown in Figs. &2 to 84.

Thickness of columns, whether cast hollow or ribbed S = .,^ + "5 in.

„ , flanges o^ = 2 io2'26.

„ guide plate &,= 1*5 to 1*76.

d being the diameter of engine shaft.

§ 135a. 2. Heavy-built Eng^ines for Large Merchant

Vessels. — Hollow cast-iron columns of square section are almost
universally used, two' to each cylinder. The back columns are almost
always placed on the condenser, provided they fall within its length.
For typical shape of column see Fig. 101. The guides are bolted on,
and are generally water-cooled. (See § 107, page 205.)

Thickness of columns S = . -■ -i- 5 inch.

oK)

n

flanges S^ = 2 to 2-25.

The flanges are strengthened with strong ribs or webs.

In large engines the tall columns are often forked at the bottom in
shape like an inverted Y. This method of construction increases the'
stability of the engine, and the columns can be placed nearer to the
shaft, because the crank can work freely between the two legs of the
column. For typical shape, see Figs. 83 to 85, and 248, 249.

For facility in machining and in erecting the engine it is not desir-
able to cast the bearings for the reversing-gear or air-pump link gear,
&c., in one piece with the columns. It is better practice to leave the
requisite facings for these as well as for the draincock gear, reversing
gear, &c. The columns are frequently fitted at the top with brackets

THE >rAIN ENGINKS. 247

or flanges, to which the distance pieces are bolted for bracing the
columns together. The hollow casting of the column is often utilised
as an oil tank, and not infrequently, in the case of the l.i> column,
the exhaust steam is led through it direct to the condenser.

§ 136. -i. Engines for Modern Fast Steamers and Large
Warships- — In these the condenser is generally separate, and is
either of cast gunmetal or of sheet copper or brass, and is generally
placed in the wing of the vessel. There are usually from two to four

248

MARINE ENGINES AND BOILERS.

long hollow or ribbed cast-steel columns to each cylinder. This is
the type of construction adopted in H.M.S. "Powerful" and the new
North-German Lloyd and Hamburg-American liners. (See Figs. 96,
97 to 99.)

Thickness of cast-steel columns

3 = - + -5 inch to ^ + '^ inch.
4U «5o

„ flanges for cast steel 6j = 2*2 to 2*85.

d being the diameter of crank shaft in inches.

Fig. 250.

Fig. 251.

§ 137. 4. Engines for Warships in general (except Tor-
pedo-boats).— These usually have on the front of the engine two
wrought-iron, or steel columns for each cylinder, and on the side of
the ahead guide, either one or two cast-steel columns, to which the cast-
iron guide is attached. The reversing-gear bearings are either bolted

THE MAIN ENGINES. 249

to the back columns or to the cylinder, or are fitted into forged pro-
jections on ihe front columns. The columns are fitted with top and
bottom flanges. The thickness of the columns is the same as that
given in § 136.

S 138. 5. Very Light Engines. — In torpedo-boats, yachts,
and steam pinnaces, the cylinders are usually placed on wroughtsteel
columns. These must be properly braced together in such a way as to
absorb the strains coming on the engine framing. In larger engines of
this type there are generally four columns to each cylinder, the two
front columns being firmly connected to the corresponding back columns
by diagonal ties. The columns are also braced diagonally in the
longitudinal direction. For types of this arrangement see Figs. 72 to
75, and 250, 251. For drawings of engines supported on columns
see Fig. 89, also "Arrangement of Main Engines," p. 106. For engines
supported on inclined columns see Fig. 113, also Figs. 254, 255.

The columns generally have collars and screwed ends top and
bottom. These are fixed with nuts into strong lugs on the cylinder
and bed-plate. Sometimes flanges forged on the columns are used,
instead of screwed ends and nuts. Even in large engines for merchant
vessels the cylinders are sometimes supported on wrought-steel columns.
This arrangement has the advantage of rendering the engine more
accessible and easier of inspection.

§ 139. Stresses in the Columns and Framing.— These are

produced by the following : —

1. Weight of cylinders.

2. Maximum thrust of connecting rod.

3. Pressure on the guides.

4. Pressure exerted by the reversing shaft.

5. Strains due to the expansion stuffing boxes on the receiver pipes

and main steam pipe.

6. Strains set up by the rolling of the ship.

It is usual, for the sake of simplicity, to calculate the size of the
columns only to withstand the tensile stress due to the maximum load
on the piston. The stresses allowed are correspondingly smaller, and
may be as follows : —

350 to 640 lb. per sq. in. for cast-iron columns for merchant vessels.
1,280 „ 1,420 „ „ cast-steel „ fast steamers.

1,420 „ 1,850 ., ,, cast-steel „ warships.

1,700 „ 2,125 „ „ wrought-steel „ „

„ 7,000 „ „ „ „ torpedo-boats.

250

MARINE ENGINES AND BOILERS.

The latter stress is taken on the area at the bottom of the thread.
In engines with cylinders supported entirely by columns the stresses
can be more accurately calculated.

Example, — Calculation of the stresses set up in the columns and

Fig. 2o2.

Fig. 2j-2a.

diagonal bracing of a destroyer due to the pressure on the guides. For
diagrammatic arrangement see Fig. 250. Let the maximum load on the
piston be 76,000 lb., then the greatest horizontal force on the guide will
be (if length of connecting rod = 4 x radius

of the crank) q = '^^'^^^= 19,000 lb. This

occurs when the crosshead is at about the
middle of its stroke. The forces on the
upper and lower guide supports will there-
fore be = ^, and the force? = 4,750 lb. will
2 4

be transmitted to each side column at the
upper and lower end of the guide (points i
and II in Fig. 253). The horizontal force
at the upper end of the guide cannot be
taken by the member in, as the latter
would bend. On the other hand, the
cylinder transfers the thrust to the point iv.
The tie ii v, however, cannot bear any great
strain, or the column iv v vi would bend.

Thus, to take up the horizontal thrust q, there remain only the links
shown in Fig. 252. The stress in the different members is repre-
sented by the polygon of forces. Fig. 252a. Assuming the direction of

Fig. -253.

THE MAIN ENGINES. 251

the forces to be as there shown, the stresses produced by the thrust on
the guide will be as follows : —

Member ii iv, compression - 6,480 lb.
„ IV V and v vi, tension 4,408 „
„ II III, compression - 12,200 „
„ II VI, tension - - 12,300 „

If the engine is reversed, the direction of the thrust on the guide,
and therefore of all the forces, is reversed.

§ 140. Fixing the Columns. — The columns are secured above
and below by fitted bolts. If studs are used (and they cannot always
be avoided in warships) a few accurately turned set pins must be used,
to keep the columns in their places. The allowable tensile strain on
the bolts at the bottom of the thread is from 3,500 to 5,600 lb. per square
inch, according to the type of engine. It is best to make the bolts at
the bottom end much stronger than those at the top, because they have
to absorb the strains set up at right angles to the direction of the piston
rod (thrust on guides, &c.). The steel columns, and also the horizontal
distance pieces between the A-frames, are as a rule made somewhat
short, and a fitting strip about j\ to § inch is used, which is accurately
fitted while the engine is being erected.

Reversing and Turning Gear.

§ 141. Reversing Shaft and Lever.-— The reversing shaft is

parallel to the crank shaft, and carries a lever for the valve motion of

each cylinder. The movement of this lever is transmitted by a link or

links to the quadrant. The main lever for actuating the shaft is also

fitted on to the reversing shaft, and is worked direct by the reversing

gear, or by the reversing engine. This lever should be placed at about

the centre of the shaft, so as to reduce the twisting stress upon the

latter to a minimum. The bearings carrying the reversing shaft should

be as close as possible to the various levers. This cannot always be

managed, and the shaft is then exposed to bending as well as twisting

stresses. Both must be allowed for when determining its diameter.

For calculating the latter, the power required to reverse one valve

motion, when the valve is exposed to the maximum steam pressure,

should be taken as a basis.

. ^ , . * r * • -4. /stroke of eccentric\ .
Let r be the amount of eccentricity or ( 1 m

inches.
R „ length of reversing lever in inches.

:MAkIXE ENGINES AND BOILERS.

THK ^MAIN KNr.INES.

a.^! MARINE ENGINES AND BOILERS.

IjCI a be the distance l)Ctween the eccentric-rod pins on the quadrant

in inches.
Q „ maximum load on the ralve rod in pounds (see

page ill).
p „ maximum load required to reverse one t'alre motion in

pounds.

Then p = 21q -. This power is divided between the two links con-
necting the lever to the quadrants.

The stress in the weakest part of the links is generally small, viz.,
about 1,200 to 1,500 ib. per square inch, and the stress in the reversing
shaft-bearing bolts about 3,500 to 5,600 Ib. per square inch. The pins

of the links, the length of which is greater than their diameter, should
only be allowed to lake a pressure of from 425 to S50 Ib, per square
inch. The diameter of the reversing shaft is therefore —

^=1-72 'Am?

Here c= \ for engines with one crank.

f r= 1 -3 „ two cranks at an angle of less than 90°.

c=l-**n „ three cranks „ „ 120°.

c= 2-4 „ four cranks „ „ 90'.

The amount of twisting stress allowable k is generally about 3,500 to
4,250 lb. iwr square inch. If the shaft is subjected to a bending as well
as a twisting stress, the above stresses should be somewhat reduced.

THE MAIN ENGINES.

255

If weight has to be economised, the reversing shaft is usually made
hollow.

In order to keep down the stress upon this shaft, it is desirable to
make the levers working the links as short as possible. This may
be done, for any given length of link, by increas-
ing the angular travel of the reversing shaft ; the
latter, however, should not generally exceed 90°.
(In engines of over 100 i.h.p. each reversing lever
generally has a slot in which a slide-block works,
and to which the reversing links are attached.
The position of the block can be regulated by
means of a screw.) (See Fig. 256.) By this
arrangement the most economical cut-off in each
cylinder can be easily obtained, and the power
developed varied within certain limits (say about
12 to 16 "IJ. The slot is so designed that it is
parallel to the valve rod when the engine is in
full backward gear, so that the position of the
valve gear is unaffected by the position of the
adjusting block in the slots. (See Fig. 258.) If
the angular travel of the reversing shaft or lever is 90^ the slot is at
right angles to the valve rod, when the valve gear is in full forward
gear. Hence the travel of the block in the slot is equal to the travel of
the quadrants. The chord of the angle through which the lever travels,

Fig. 257.

^

«?

I

I

Xink,

; ^ ill' mid jpo^iUott^

Fig. 258.

s, is generally equal to or somewhat less than the distance between the
eccentric-rod pins on the quadrant. The reversing shaft and adjusting
block are made of wrought iron or steel, the levers of cast iron, wrought *
iron, but more generally of cast steel.

256 MARINE ENGINES AND BOILERS.

If d is equal to 1, the usual dimensions for cast steel are given in
Fig. 256. If cast iron be used, the dimensions of the levers should be
. increased by 10 to 15 'Z^, but the main reversing lever should always be
considerably stouter than the auxiliary levers. If the levers are of forged
steel (as is nearly always the case in engines for warships), they are gene-
rally of the shape shown in Fig. 257, and are either forced on to the
reversing shaft under pressure, or shrunk on, and are secured by strong
keys in both cases. To prevent any movement of the reversing shaft
while the engine is running, some kind of locking gear is usually provided.
If the reversing gear is worked by hand, a screw clamp may be used. If
steam gear is used, such an arrangement is not possible, because if the
ship has suddenly to be brought about, the clamping device takes too
long to release. It is better to have a stop, against which one of the
reversing levers, or a lever specially provided for that purpose, bears
when the engine is in forward gear.

§ 142. Method of Handling the Reversing Gear.— Engines

up to 150 H.p. may be reversed by means of a hand lever keyed direct
to the reversing shaft. In larger engines, up to 500 h.p., the reversing
shaft is generally worked by a hand wheel with either worm gear or
screwed spindle ; engines above 500 h.p. are generally provided with
steam reversing gear, and a hand gear in addition, as a stand-by in
case of need, the two being connected up to one another. Reversing
engines are either "direct-acting" or what is known as "all-round"
reversing engines.

§ 143. Direct-acting Reversing Engines (Brown's reversing
gear). — These have a steam cylinder, and a brake or dash-pot cylinder
filled with oil or water, in line with one another (Figs. 259-262). The
brake or dash-pot cylinder is necessary to prevent a jerky or too rapid
motion of the piston in the reversing engine, and therefore of the valve
gear. In the reversing engine shown in the figures, the crosshead
is connected to two levers on the reversing shaft by means of two
links or connecting rods, the engine itself remaining stationary. A
simple arrangement is to make the upper end of the piston rod pro-
ject through the top of the brake cylinder, and attach the pin of the
main reversing lever to it by means of an eye. The reversing engine
does not remain stationary in this case, but can swing, at its bottom end,
round a pin solidly connected to the bed-plate of the engine. The
steam cylinder is worked by an ordinary D slide valve (with verj- small
outside and inside lap), and the brake cylinder by a piston valve. (See
Figs. 261 and 262.) The two sets of valves are interconnected and
worked by a hand lever from the engine platform. As soon as both

THE MAIN ENGINES.

25'

258

MARINE ENGINES AND BOILERS.

Fig. 260.

THE MAIN ENGINE;

valves have moved, and steam enters the cylinder, the piston rod of the
reversing engine begins to work, and carries with it the crosshead, to

which the connecting links of the reversing gear are attached. To this
crosshead is also connected an arm carrying a nut which is prevented
from turning. Ii travels over a vertical spindle having a thread with

260 MARINE ENGINES AND BOILERS.

a very coarse pitch, and it is therefore able to rotate the spindle. The
lower part of this spindle has also a fine thread, which works in a nut
attached to the lever working the valve
rod. The engineer having moved the
reversing lever, the valve rod moves
because the fine-threaded nut directly
transmits the motion. The valves now
being open, the piston and crosshead
begin to move. The lever being now
stationar)', the valve spindle, caused
to rotate by the coarse-threaded nut,
travels as it turns in the fine-threaded
nut which is fixed to the lever. The
valve attached to the spindle travels
with it, and thus comes into its mid
position. This object may also be
effected by means of a series of levers,
in place of the screwed spindle. This
is illustrated diagram matically in Fig.
263. The slide valve of the steam
cylinder is moved in the first instance
by the hand lever ; this sets the piston rod in motion, and as the |K)int
A is stationary, the valve has to return to its mid position.

The brake cylinder may be used without separate valve gear, in which
case a bye-pass, closed partly or wholly by a valve or cock, connects
the two ends of the cylinder. Occasionally the brake piston has a small
hole in it, through which the water can pass from one side of the piston
to the other. The piston of the reversing engine is generally designed

in the same way as an ordinary cast-iron piston. The brake cylinder
piston has either the usual metal packing rings, or a double leather bucket
packing (see Fig. 264) between which is a turned ring of white metal.

THE Main engines. 261

The diameter of the brake cylinder is about 0*6 to 0*7 that of the
steam cylinder. The maximum pressure in the brake cylinder is there-
fore about 2 to 2*8 times greater than the maximum steam pressure in
the steam cylinder. The steam cylinder is generally made with a piston
rod passing through its upper end only, while the hydraulic or brake
cylinder has a rod passing through both covers, and is of the same
diameter throughout. This obviates the necessity for an expansion
chamber, into which the fluid is forced, and withdrawn, which would be
necessary if the piston rod were not carried through, and the volume
above and below the piston were different. The reversing engine
should be placed as close as possible to the main reversing lever
which it operates. The hand gear, for working the reversing gear
by hand, should also be brought down to the engine platform with
the fewest possible number of joints, levers, and shafts.

The dimensions of the hand gear should be so proportioned that
the power required to work the engine is not too great, nor the twisting
and bending strains thrown upon the rods so considerable that they
bend, or fail to act satisfactorily. The steam for working the engine
should be taken both from the main and the auxiliary steam pipes.
The exhaust is usually led to the condenser.

§ 144. All-round Reversing Gear.— This generally consists of
a single or double cylinder steam engine, on the crank shaft of which
a worm is fitted driving a worm wheel. The arrangement of a two-
cylinder reversing engine of this kind is shown in Fig. 265. A crank
on the worm-wheel shaft, or a crank pin fixed direct to the worm wheel,
actuates the reversing shaft by means of a connecting rod. In Fig. 266
a diagram is given of an auxiliary reversing engine, which is used in very
large marine engines as a stand-by. The reversing gear is actuated by a
rack and pinion, the latter being connected to a worm wheel by means of
a clutch coupling. An engine drives the worm wheel through a worm.

The reversing of both these small engines can be effected either
by a revolving slide valve, as shown at Figs. 267, 268, or by a change-
over valve as in the small two-cylinder engines. The revolving or
reversible slide valve is so arranged that it acts on one side like an
ordinary D valve, and on the other like a so-called " E slide valve,"
which in its mid position keeps all the ports closed. The valve is
turned by means of a hand lever, the valve rod being square in section
beyond the valve chest, and also in the slide valve itself. The slide is
worked by an eccentric set at an angle of 90" with the crank. If the
reversing is efTected by a change-over valve, the distributing slide valve
is made like an ordinary piston valve, and is operated as in the other
case by an eccentric set at 90' to the crank. The change-over valve is

"* — •

"^5^ .A^.

I

I

I

^^- 3Sii

THIC MAIN ENGINKS.

264 MARINE ENGINES AND BOILERS.

in a se])arate valve chest, and is worked by hand from the platform ; it
admits steam into and discharges it from the piston valve chest. Steam
is thus admitted to the inner or to the outer side of the piston
valve, according to the iX)sition given to the change-over valve. (In
both the above tyjjes of engines, the valves have a very small inside and
outside lap.)

A flywheel is generally placed on the crank shaft of the reversing
engine, which acts at the same time as a hand wheel, to throw over the
gear of the main engines by hand. The diameter of this wheel varies
from 2 feet 6 inches to 5 feet 3 inches, according to the size of the
engine. In the smaller sizes it is made wholly of cast iron ; in the
larger it has a wrought-iron rim and arms, with a cast-iron boss." Projec-
tions on the outside of the rim should be avoided. If of cast iron, the
cross section of the rim is elliptical j if of wrought iron, it is generally
round, and its diameter from 1 J to 2 inches. In lighter-built ships it
consists of a bent tube.

§ U5. Principal Dimensions of Reversing Engines.—

These, arrived at from the following equations, are as follows : —

(1) /xs = Ci ^ — ^^^ direct-acting reversing engines.

/

Ox/*

(2) /x s = c,yX^ -~ for all-round reversing engines.

' px n

Here Q = the maximum load on one main valve rod in pounds.

r=half stroke of main engine eccentric in feet.

/ = absolute steam pressure in the main steam pipe in pounds
per square inch.

/= area of the steam piston of the reversing engine in square
inches. If there are two cylinders, / is the sum of the
areas of both pistons.

5 = stroke of piston of reversing engine in feet.

«^ number of revolutions required in an all-round reversing
engine, to reverse the main valve gear from ** full ahead "
to " full astern."

^1 and ^2 = constants given in the following table.

THE MAIN ?:NGINES. 265

Table No. 24.

Coefficients for Calculating the Dimensions of Reversing Engines,

'i

^a

Type of Main Engine.

2-7

6-7

Single-cylinder engine.

3-8

9-3

Compound engine with two cranks at an angle of 90°

5-4

13-4

Triple „ „ three „ „ 120"

7-6

18-6

Quadruple „ „ four „ „ 90'

In determining the dimensions of reversing engine details, the full
boiler pressure must be taken as a basis. The stress on the pins, rods,
&c., if they are of wrought iron or steel, should be from 4,250 to 7,000
lb, per square inch ; the stress in the teeth of the worm wheels, if of cast
iron, from 2,850 to 3,550 lb. per square inch; and from 3,550 to 4,250 lb.
per square inch if of bronze or steel, on the assumption that the whole
oad is taken by two teeth. This corresponds to the twisting moment

(vn*^) ^" *^^ main reversing shaft. According to the size of the

engine the worm wheels have from 20 to 60 teeth, with from If to 3
inch pitch. The worms are either of mild steel forged in one with the
shaft, or of bronze, and keyed on to the shaft. It should be noted that a
bronze worm works best in a cast-iron or cast-steel worm wheel, and
a steel worm in a cast-iron or gunmetal worm wheel. In order to
avoid backlash, the teeth of the worm wheel should be milled out by a
cutter. If this is done, the worm must be accurately adjusted to the
wheel, so that it may be in proper alignment.

The diameter of the pitch circle of the worm is about 1*8 to 2*5/ if
the w^orm is of steel ; if of bronze and separately cast it is 2*5 to 3*5/,
where / denotes the pitch. For convenience in making and fitting, the
slope of the worm thread is made straight, whereas the teeth of the
worm wheel are slightly hollowed. Length of the worm from 3 to 3*5/.
Width of rim of worm wheel 0'6 to 0*8 times the diameter of the pitch
circle of the worm. To diminish the wear of the teeth, care must be
taken to ensure sufficient lubrication. The best way to secure this is to
make the worm or the wheel work in an oil bath.

266 MARINE ENGINES AND BOILERS.

§ 146. Turning Gear. — Engines with a stroke of not more than
H inches can usually be turned by means of a hand lever. This may be
applied either at the forward end of the shaft, or in holes specially bored
in the circumference of a coupling flange, or of a small separate wheel
fitted especially for that purpose. Larger engines have a worm wheel
on the crank shaft, which is turned by a worm and hand lever. If the
power required to turn the worm is too great to be worked direct by
hand, it is turned through toothed gearing, or by special worm gear and
a small steam engine. Instead of having a separate engine, the reversing
engine or centrifugal pump engine is often utilised for this purpose.

The turning moment m^ on the crank shaft required to turn the
main engine is —

N
M, = c- D ft. lb.

n

Here n is the i.h.p. of the main engine.

n „ number of revolutions per minute of the main engine.
D „ diameter of the crank shaft in feet,
c is a coefficient, which is about 280 for engines of very light
build, and about 560 to 670 for heavily built engines.

The required turning moment Mj may be more accurately deter-
mined from the following equation —

Mi = ^(2Gi + G2)ft. lb.

Here g^ is the total weight of the pistons, piston rods, crossheads, and
connecting rods in ]X)unds.
G2 is the total weight of the shafting (including the screw) in
pounds.

Further, referring to Fig. 269 —

Sj is the number of teeth in the worm wheel a on the crank shaft

of the main engine.
Z2 is the number of teeth in the worm.
/ is the cylinder area of the turning engine in square inches. (If

there are two cylinders of equal size,/ is the sum of the

areas of both cylinders.)
s is the stroke of the turning engine in feet.
fi is the boiler pressure (pounds per square inch) above

atmosphere.
/ = total reduction in gear from the crank shaft of the main

engine to that of the turning engine.

THE MAIN ENGINES.

267

Assuming that the worms s^ and s^ are single threaded, then —

/= z, X So, and /x s = 150-^—^-

The total reduction in gear / varies from 1,500 to 4,000, and as the
turning engine generally runs at 250 to 400 revolutions per minute,
from 4 to 16 minutes are required to turn the main engine through one
revolution. The diameter of the pitch circle of the worm wheel on the
crank shaft is from 1*2 to 1*6 times the stroke of the main engine.

The materials used for the worm wheel are generally cast iron or
cast steely seldom bronze. The corresponding worm is made of bronze,
cast iron, or mild steel, and so arranged that mild steel works upon
bronze or cast iron, and bronze upon cast iron or cast steel. Mild steel
upon cast steel is not to be recommended, as they do not wear well.

Fig. 269.

If the worm is made of mild steel, it is generally forged in one piece
with the spindle ; if of cast iron or bronze, it is held on the steel spindle
by a taper cotter, or preferably by a key and nut. To enable the
turning gear to be thrown out of gear when the engine starts, either the
worm s^ (Fig. 269) must move axially along its shaft, or it must be possible
to throw both worm and shaft out of gear from the worm wheel. The
latter is the arrangement adopted in the turning gear shown in Figs.
270 to 274. In Figs. 271 to 273 the worm is thrown out of gear by
means of a hand wheel, which turns a crank shaft by means of a worm
and worm wheel. The crank pin is fitted with a swivelling bearing,
which can slide on the worm shaft, and at the same time forms the
bottom bearing in which the worm shaft rotates. In the turning gear
shown in Fig. 274 the worm is thrown in and out of gear direct by
hand. The worm wheel on the main crank shaft should be fixed on

268

MARINE ENGINES AND BOILERS.

a coupling, and to facilitate fitting is usually made in two halves, so
that the thrust shaft or after crank can easily be taken out and
replaced.

THE MAIN ENGINES.

269

to

^ 147. Calculation of the Dimensions of the Wheels.—

The following are the symbols employed in regard to Fig. 269 : —

Pj Tangential force in the pitch circle of the large worm wheel a
in pounds.
Diameter of the pitch circle of the large worm wheel a in feet.
Pitch of the large worm wheel a in inches.
Width of the large worm wheel a in inches.
F., Tangential force in the pitch circle of the small worm wheel
in pounds.

MARINE ENGINES AND BOILERS.

Fig. 274

Diameter of ihe pitch circle of the small worm wheel ii
Pitch of the small worm wheel in inches.
Width of the small worm wheel in inches,
and fj are constants.
Number of teeth in the la^e worm wheel A.
Number of teeth in the small worm wheel.
6m,

Thenp, = 2^; *,x/, = £l; f
a single thread worm gear s) ; b^t..

s,rf.

(with an efficiency = J and

0^91

THE MAIN ENGINES. 271

The constants c^ and r^ may be used on the assumption that the
strain is taken by two teeth simultaneously, and in the accompanying
Fig. 275,/= 0-6/; /=0-65/. The con-
stants ^1 and c^ are — m ^ ,i

For cast iron or bronze c^ = 640
„ „ ^2 = 355

For cast steel r, = 895

The stress in the teeth, if made of Fig. 275.

cast iron or bronze, is taken at 3,500

lb. per square inch; if of cast steel, at 5,000 lb. per square inch.
As the thread of the worm wears out much quicker than the teeth
of the worm wheel, it is advisable, especially if the worm is made of
bronze, to make the teeth of the worm considerably thicker than the
teeth on the wheel. If the worm wheels are milled by a cutter, the
clearance between the thread of the worm and the teeth of the worm
wheel need only be from 0*05 to 003/.

Thickness of the rim of the worm wheel inside the teeth is about
0*05/ ; thickness of the arms and ribs about 0*4 to 0*45/.

Width of the rim of the worm wheel and of the teeth is as follows: —

^1 = 2-4 to 2-8/i ; A^ = 2'4 to 2-8/2.

The boss is generally as wide as or rather wider than the rim of
the wheel.

Radial thickness of boss, if the wheel is of cast iron, is about 1*2 to 1*5/.
„ „ „ cast steel, „ 0*75 to 0*95/.

The diameter of the pitch circle of the worm is —

From 1*5 to 2*5/ if the worm and shaft are forged in one piece.

„ 2*5 „ 3*5/ „ is fixed on separately.

Radial thickness of the metal below the thread about 0*5 to 0*6/.
Length of worm should be at least from 3 to 3*5/.

1. Surface Condensers,

§ 148. General Remarks- — Modem vessels are invariably fitted
with surface condensers. The weight, cost, and space occupied are
considerable, but the recovery of the feed water outweighs these dis-
advantages. The condenser tubes are generally horizontal or slightly
inclined, and only in very exceptional cases vertical. The steam to be
condensed generally enters from above, and circulates outside the con-
denser tubes, and care must be taken that
as far as possible it is equally distributed
over the whole length and breadth of the
tubes. To effect this a sheet of galvanised
iron pierced with holes is fixed immedi-
ately beneath the steam inlet (see Fig.
279), and this serves at the same time
to prevent the steam impinging directly
on the tubes. These baffle plates must be
stiff enough to stand ihe current of steam
striking against them without deflection,
as otherwise they might injure the upper
most rows of tubes. In the upper part of
the condenser some of the tubes are some-
times omitted in order to facilitate the
passage of the steam into the centre of
the nest of tubes. (See Figs. 276 and 281.)
The circulating water is generally led
twice through the tubes, that is, it is led
Fig. 276. forward through half of them, and back

through the other half. Whether the
water is admitted at the bottom and discharged at the top, or vice
versd, does not materially affect the vacuum produced. Surface con-
densers in which the circulating water passes outside the tubes, and
the steam to be condensed passes inside them, are hardly ever made
in modern practice. If the circulating water is admitted at the bottom,
the temperature of the condensed steam will be lower than if the water

THK MAIN ENGINES.

273

is p^assed through the condenser in the reverse direction, because in the
former case the condensed steam comes in contact with the coldest
condenser tubes last of all.

The air-pump suction is connected to the lowest part of the con-
denser ; if the air pump is separate from it, the two must be connected by
a suction pipe of sufficient diameter. For the size of the air-pump suction
pipe see page 286. For dimensions of inlet and outlet for the circulating
water see pages 294- and 298. If a plunger pump is used for the circu-
lating water, it is desirable to form an air chamber on the cover of the
condenser, on the side at which the circulating water enters. By this
means water-hammer and the resulting injurious strains to the con-
denser are avoided. When placing the condenser in the ship, care must
be taken that sufficient space is left at one end to enable the tubes to
be withdrawn and replaced. This space should not be encroached on
by any parts of the engine, piping, &c. Sometimes a hole, over which a
plate is fitted, is left in the nearest bulkhead, through which the tubes
can be drawn.

§ 149. Cooling Surface. — The surface required per i.h.p. can be
taken at the following values —

1-5 to 1'7 square feet in compound engines.
1 „ 1*5 „ triple engines.

The external surface of the cooling tubes is reckoned as the cooling
surface, and the tubes are arranged as shown in Fig. 277. The following

Fig. 277.

Fig. 278.

Table No. 25 gives the amount of cooling surface per linear foot of tube
and per cubic foot of nest of tubes, for different diameters and pitches
of tubes.

274

MARINE ENGINES AND BOILERS.

THE MAIN ENGINES.

275

276

MARINE ENGINES AND BOILERS.

Table No. 25.

Number

in a

square

fool.

Pitch of

Tubes / in

inches.

External Diameter of Tubes in inches.

1 inch.

i inch.

3 inch.

I inch.

I inch.

172

150

137
128
121
116
110

99

1

1

liV

u

1 s

11

lA

sq. ft.
'22-515
19-635
17-933
16-855

sq. ft.

28-139

24-54

22-413

2f)-94

19-795

sq. ft.

29-445
26-893
25 126
23-752
22-77

sq. ft.

31 -386
-29-325
27-721
26-575
25-201
22-68

•28-798
25-918

Cooling sur-
face (in sq.
ft.) per cubic
foot of nesi

of tubes.

1

Cooling surface in
square feet per
foot run of 1 lube

•1309

-1636

•1963

-2291

•2618

sq. ft.

Note. — Table No. 26 has been incorporated in the above table.

§ 150. Tubes and Tube Plates.— The /udes are of solid-
drawn brass, containing about 60 to 70 °/^ copper and 30 to 40 7^ zinc
(an alloy of 70 % copper, 29 % zinc, and 1 % tin has been found to
answer well), and are often tinned inside as well as out. Their external
diameter generally varies between | and ^ inch, and their thickness
between 19 B.W.G. and 16 B.W.G., the figure generally used being
18 B.W.G. (0-049 inch). The maximum length of the tube between the
tube plates is about 20 feet. If the length is more than 80 to 100 times
the external diameter of the tubes, they must be supported by special
intermediate plates of sheet brass, from -^^ to ~ inch thick. The pitch
of the tubes (/), /.^., the distance of the centres of the tubes from each
other, is not less than / = ^+0-35 inch, d being the external diameter
of the tube. Table No. 25 gives the cooling surface per foot run of
different tubes, as well as the number of tubes in a square foot of tube
plate, for different pitches of tube.

Tude plates are now seldom made of cast bronze, but generally of
rolled brass or " Muntz " metal (60 % copper, 40 % zinc), from \ to
1 inch thick. The packing and securing of the tubes in the tube
plates is usually done by means of small stuffing boxes and a screwed
ferrule or gland, packed with soft cord, as shown in Fig. 278. The
ordinary dimensions of these stuffing boxes (see Fig. 278) are —

0'85//, D = ^+-25 inch, ^i = i/--094 inch, /=0-95^,
b = ^^ to /tt inch, e^-^\.o\ inch.

a

THE MAIN ENGINES. 277

The number of threads to the inch is about 16 to 18. When the tube
plates are very lai^e, they must be braced to one another by tie rods to
ensure their rigidity, the stays being spaced about 15 to 20 inches apart.
(See Figs. 276 and 279, 281 and 382.) These stays are generally thick
brass tubes, accurately fitting between the end plates, and secured from
the outside by a set screw passing through the plates, and screwed into
the tapped bore of the tube.

g 151. Condenser Sfaell. — This is either cast in one with the
engine frame, or made separately. In the former case it is of cast iron,

and rectangular in shape. With separate condensers the form is generally
cylindrical, although in warships they are often oval, the flattened sur-
faces being stayed. They are made of cast iron, cast gunmetal or
bronze, galvanised steel, copper, or brass plates, and tinned inside. If
niade in one with the framing, the bottom of the condenser is so
shaped that the condensed steam can flow easily to the air-pump suction
pipe. (See Fig. 279.) With the same object, separate condensers are
usually slightly inclined towards the end connected to the air pump.

278

MARINE ENGINES AND BOILERS.

THE MAIN ENGINES.

279

The thickness s of cylindrical condenser shells is —

d

j =

180
d

+ -393 inch for cast iron.

!='*-- + '196 „ wrought iron or mild steel.
400

s = - + -196 „ cast bronze or gun metal.
s= — - + -039 „ rolled copper or brass.

oO\)

To stiffen the shell and assist in maintaining its form, a few circum-
ferential ribs are fitted. If the condensers are of copper, brass, or iron
plates, these ribs are generally T-shaped in section, and riveted on.

The thickness of condenser shells, cast in one with the engine
framing, and with flat walls, may be taken from Table No. 27. The
flat walls are stiffened by ribs placed at a distance of about 20x apart,
the height of the ribs being about 4j. In large condensers of this
kind the two vertical longitudinal walls are also sometimes braced
together by stay rods passing between the two groups of condenser
tubes (Figs. 279 and 280).

Table .

No.

07

1

Coolinj; Surface of Condenser,

Thickness s of Condenser Shell having

in square feet.

Flat Walls, in inches.

Above 5400

M81

From 3200 to 5400

M06

„ 2150 „ 3200

0-984

„ 1600 „ 2150

0-866

„ 1000 „ IGOO

0-787

540 „ 1000

0-708

' Below 540

0-511 to 0-708

In cylindrical condensers the water chambers at each end are made
of cast iron, bronze, or of copper plates. They have to be entirely re-
moved to reach and overhaul the tube plates and stuffing boxes, unless
they happen to be fitted with a large inspection door which can be easily
removed. The ends of large condensers are therefore always fitted with
manholes and manhole doors, to allow the tube stuffing boxes to be more
easily examined. The tube plates are fixed to the condenser shell and
ends in the manner shown in Figs. 279 and 283.

280 MAkINK KNXINKS AND BOILERS.

To test the condenser and packing for leakage, they are subjected
to a water pressure of from 30 to 40 lb. per square inch ; first
the shells only, with the covers
but without the tubes, are tested,
and then the steam space, after
the tubes have been put in,
to make sure that the tubes,
stuffing boxes, and shell joints
are tight.

§ 152. Fttting:s and Con-
nections.—Besides the opening
for the main steam exhaust, inlet
and outlet of the circulating
water, and the air-pump suction,
the following connections are
generally provided : — Auxiliary-
engine exhaust, electric light
engine exhaust, reversing en-
gine exhaust, and the drains
from the steam cylinders and
valve chests.

On the top of the condenser
Fig. 38!!. shell a manhole is usually fixed,

and at the bottom one or
more mud-holes. The bottom of the condenser has also a valve and
pipe connection from J to ^ inch internal diameter, to allow of its
being boiled out with steam. At a suitable place there is also a cock
~ to 1 inch internal diameter, through which the make-up feed water
is added. In modern practice this pipe is usually connected to
the reserve feed tank only. A soda cock, about J inch internal dia-
meter, fitted with funnel or hose, is also provided. At the lowest
point in the bottom of the condenser there is another valve from IJ
to 3 inches internal diameter, according to the size of the condenser,
through which, if the air pump is not working, the condenser can be
kept clear by a steam pump. Drain cocks are placed as required
in the lower part of the condenser, and generally also a special valve,
through which the condenser can be filled direct from the sea. To
protect the inside of the condenser, where it is in contact with the sea
water, against corrosion, zinc slabs are often fitted.

-. Jet Condensers.
g 153. Jet Condensers. — These are now used only in river

THE MAIN ENGINES. 281

Steamers, and consist of an arrangement by which a jet of water taken
from the river is injected into the exhaust pipe, or into the hollow
column of the engine through which the exhaust is carried. In case of
need the injection water can also be drawn through a special cock from
the bilge (so-called auxiliary injection). The quantity of injection water
is regulated by a cock or valve placed in the admission pipe, which
can be easily manipulated by means of levers from the engine platform.
The area of the injection pipe is as follows : —

•025 to '031 square inch per i.h.p. in compound engines.
•0186 „ -025 „ „ triple-expansion engines.

The diameter of the holes in the rose of the injection pipe is about ^
to tV inch, and their total area is about 0*5 the area of the injection pipe.
If there is any danger of the small holes getting stopped up, the water
is admitted over a weir. An automatic valve is placed at the bottom
of the condenser, which is held shut by the pressure of the atmosphere
outside. The object of this valve is to discharge the water automatically
from the condenser when the engine is not running.

PART II.

PUMPS.

I

Air Pumps.

§ 154. General Remarks. — The air pump is generally fixed to
the main engine, and driven either direct or by levers from one of the
crossheads, the latter being the most usual arrangement. In modern
practice the air pump, especially in large engines, is often entirely
independent of the main engines. It then forms a separate engine,
and is sometimes combined with the circulating pump, as in the Blake,
Weir, Worthington, and other systems. In the mercantile marine, if the
air pump is to be driven from the engine, there is generally only one
pump, while in warships two are frequently provided for each main
engine. In the latter case the suction and delivery pipes of the
pumps should have suitable screw-down or sluice valves, so that each
air pump may be worked independently of the other, if necessary.

Air-pump barrels are generally vertical, and the pump should be
placed as low as practicable, but the foot valves should at least be
at the same level as the bottom of the condenser. The amount
of clearance space in the air pump has not a very serious effect on the
vacuum produced, because it is filled with water.

i^ 155. The principal dimensions of single-acting air pumps are
determined from the equation —

I.H.P.

/xs = c

n

Here y= sectional area of the pump cylinder in square inches.
s = stroke of the pump cylinder in inches.
I.H.P. = indicated horse power of the main engine.

n ==■ number of double strokes of the air pump per minute.
c = constant, equal to volume delivered by the air pump per i.h.p.
per minute.
The coefficient c = 86 to 111 in surface condensers of triple or quad-
ruple expansion engines, with separately driven air pumps.
(See § 161.)
c= 185 to 245 in surface condensers of triple or quadruple expan-
sion engines, the air pump being driven by the main engine.
c = 300 to 365 in surface condensers of compound engines, the air
pump being driven by the main engine.

286 MARINE ENGINES AND BOILERS.

If jet condensing is used as well as surface condensing, or if the
former alone is used, c^ 610 to 730. If, instead of one air pump, two
pumps are fixed to and driven by the main engine, the volume swept
through per stroke in each is about 0*6 of that given above.

Piston Speed. — In cargo steamers the mean speed of the air-pump
piston varies from 200 to 350 feet per minute, and in warships from 300
to 500 feet per minute.

§ 156. Air-pump Valves. — Air pumps have, as a rule, suction,
plunger, and delivery valves, although, generally speaking, either suction
and plunger valves, or plunger and delivery valves should be sufficient.

In the Edwards' patent air pump, shown in Fig. 286, there are
delivery valves only. Here the water collecting below the bucket is, as
the bucket descends, forced through openings at the side of the barrel
on to the top of the bucket. As the bucket rises, the openings at
the side are closed, and water and air are forced through the delivery-
valves into the hot-well tank. The net sectional area of the valve
openings should be such, that the mean velocity of water through them
does not exceed 13 to 16*5 feet per second. The radial speed of dis-
charge at the circumferential ports or openings should not be more than
20 to 33 feet per second.

The valves are of various types. Those in most frequent use are
flexible rubber flap valves (Fig. 285) from ^ to ^ inch thick. "Beldam "
valves (Fig. 284) are also often used, consisting of a corrugated brass
plate about -jV ^"^^ thick, and a valve guard corresponding in shape.
" Kinghorn" valves (Fig. 287) consist of three loose plates of sheet brass
from about ^V to about -|\ inch thick, placed one over the other, the
two lower having holes about ^ to yV inch diameter drilled in them
at diflerent radii. Ordinary rubber flap valves are fixed at the
centre, and a little " play " is allowed, the lift at the circumference due
to their elasticity being about | to ^ inch. Beldam and Kinghorn
valves sit loose on the valve spindle, and work up and down on it, the
lift being only from about \ to ^^ inch. The greater the number of
strokes of the bucket per minute the smaller should be the lift of the
valves. The clear openings between the ribs of the valve seatings
should not be more than about ly\ x ly\ inch, while the width of the
ribs in the guards should be at least f^ inch. The valve seats, guards,
and spindles are made of gunmetal. It is best to screw the spindles
into the seat, and rivet them over on the under side. The valve guards
are screwed on, and secured by lock nuts and split pins.

§ 157. Suction and Delivery Pipes.— With surface condensers
the sectional area of the air-pump suction pipe, or the connection to the

PUMPS.

28^

Fig. 284.

Fig 285.

288

MARINE ENGINES AND BOILERS.

condenser, should be such that the mean velocity of air and water in the
pipe, calculated from the volume of the effective strokes of the air pump,
is from 10 to 16 feet per second. The velocity in the delivery pipe should
be about 26 to 33 feet, sometimes it may even be as much as 50 feet, per
second, and the sectional area of the pipe must be determined accord-
ingly. With jet condensers the velocities in the suction and delivery
pipes should only be about 50 '/^ of the velocities for surface condensers.
When it is not required to use a surface condenser on emergency as a
jet condenser, it has been usual of late years not to fit an overboard

Fig. 5286.

air-pump discharge valve, but to leave the end of the delivery pipe
open, and let it discharge the air direct into the engine-room, while the
water delivered by the air pump either flows over into a hot well tank,
or is sucked by the boiler feed pump direct from the delivery chamber
of the pump. When an overboard discharge valve is fitted on the
discharge pipe, a small branch air pipe is taken from underneath the
valve, and is made to discharge either as high as possible above the
water line, or straight into the engine-room. Through this pipe the
air discharged by the pump can escape, in case the large discharge

valve has to be kept closed, owing to the roHing of the vessel, lis
diameter is about 02 to 0'3 the diameter of the delivery pipe. Close
to the delivery chamber of the air pump an air vessel is placed, the

cubic contents of which are about 0'3 to 0 5 that of the pump cylinder. •
Sometimes the delivery chamber of the air pump forms practically the
tank from which the feed pumps suck, and is known as the "Hot Well."

§ 158. Pump Body. — In merchant vessels the body of the air pump

290

MARINE ENGINES AND BOILERS.

is made of cast iron, the pump barrel itself being of gunmetal in the form
of a liner. In larger pumps a solid gunmetal barrel is sometimes bolted
by flanges between cast-iron suction and delivery valve chests. (See
Fig. 289.) In warships the pump barrel and suction and delivery valve
chests are all of gunmetal, and generally cast in one piece. (See Figs.

285, 290, 292.) To make the
bucket and foot valves in large
air pumps easily accessible,
the barrel of the pump is made
with a round manhole at the
side, from 10 to 12 inches in
diameter, the cover of which
must, of course, form part of
the cylinder wall. (See Figs.
285 and 290.) In the same
way hand holes with covers
are made in the sides of the
delivery valve chest, so that
these valves also can be easily
reached.

The thickness of the walls
s of the gunmetal barrel of
the pump is j = 0'015d + -25
inch, D being the diameter of
the air pump in inches. The
thickness of the walls of a
cast-iron pump body is s^ =
0-01 5d + -39 inch. The thick-
ness of metal in the suction
and delivery valve chests of
the air pump are, in warships,
from iV ^^ 1 *^^^ thinner than
the pump barrel. The foot-
valve seating is fixed to the
Fig. 290. lower part of the pump either

by several set screws round the
rim, or by a strong bolt through the centre. (Compare Figs. 285 and
292.) The head or delivery valve seating is generally held and made
tight against the pump rod by the main stuffing-box in the centre,
and at the rim either by set screws or by a few^ long bolts screwed
right through from the upper valve-chest cover. The thickness j.,
of the valve seat is j., = ri to l*2x, strengthened by radial ribs under-
neath.

PUMPS.

291

Fig. 292.

292 MARINE ENGINES AND BOILERS.

§ 159. The Pump Bucket is made of gunmetal, and the thick-
ness J3 of the flat portions of the piston is jg = 0"02D + 0-23 inch. It is
stiffened with radial ribs, about O'bs^ thick, placed between the openings
for the valve seats. The bucket is usually packed with cotton or hemp,
which in small buckets is wrapped round a deep rectangular groove on
the outside of the bucket or piston, while in the larger sizes it is packed
into a similar space, and pressed down by a junk ring. (See Figs. 284
and 291.)

The ring is not intended to adjust the packing, and is screwed down
firmly against the piston by means of studs, so that it cannot work loose.
The depth of the packing space >4 = about 0*1 0 + 1*38 inches. The

thickness of the packing space (or the thickness of the packing) = - to -.

§ 160. The Pump Rod is made either wholly of phosphor bronze,
or of steel, in which case it is fitted with a bronze sleeve. It is fitted
into the bucket and crosshead by means of coned ends with nuts, in
the same way as an ordinary piston rod. If the rod is of steel, the nut
is generally a cap nut, to prevent the water reaching the thread. It is
advisable in large pumps to provide a boss in the centre of the suction-
valve seat, to support the pump-rod nut, and to prevent the plunger
striking against the valves, when detaching the pump rod from its cross-
head. If from want of space it is impossible to have a guide and con-
necting rod above the air pump, the bucket is made as a trunk piston,
the connecting rod being attached to a pin at the lower end of the
trunk. (See Figs. 285 and 292.)

The diameter of the plain part of the air-pump rod is about
</=012d + *5 inch, and the diameter at the bottom of the thread at
either end is ^i = 0-085d to O'Id.

§ 161. Separately Driven Air Pumps.— These pumps, which
are entirely separate from the main engine, are made by the Blake
& Knowles Steam Pump Co., G. & J. Weir of Glasgow, the Worthing-
ton Pump Co., and others, and have been much used of late years.
Number of double strokes per minute = 12 to 25.

Fig. 294 shows a Blake-Knowles two-cylinder air pump. The two
air pumps are single-acting, and each is driven by a separate steam
cylinder, the piston rod of which is coupled direct to that of the air
pump. To equalise the work, the two piston rods are further inter-
connected by a beam and connecting links, and therefore move in
opposite directions at the same time. The valve gear of the steam
cylinders is worked by levers from the beam shaft.

Fig. 293 shows a similar air pump, as made by Messrs Weir. The
air pump is often combined with the circulating pump, and both are

PUMPS. 293

worked at the same time by one steam piston. The Blake-Knowles
Co., Worth ington Co., and others make pumps of this type. In the
Blake-Knowles combined pumps the two pumps and the steam cylinder
are all placed in a line, on the same axis, the steam cylinder being
in the middle. The Worthington pumps are made on the Duplex
system, with the two sets of cylinders side by side.

The Blake & Knowles Steam Pump Co. have lately constructed
vertical single-cylinder double-acting air pumps, in which, as in the
Edwards pump, the pump cylinders are only fitted with delivery valves.
(See SchiffbaUy 1903, p. 951.) Two such pumps, with cylinders of
dififerent diameter, can be coupled together through the exhaust of the
small cylinder, and be worked as compound pumps.

Circulating Pumps.

§ 162. General Remarks. — To provide the cooling water for the
surface condensers, either reciprocating or centrifugal pumps are used.
Reciprocating circulating pumps are usually worked from the main
engine, and driven, together with the air pump, bilge and feed water
pumps, by means of a rocking lever. Sometimes, however, the cir-
culating pump is separate from the main engine, and is then usually
combined with an independent air pump, as in the Blake, Worthington,
and similar systems.

The quantity of cooling water required per i.h.p. per hour is: For
triple or quadruple expansion engines, about 440 to 550 lb. ; for
compound engines, about 550 to 650 lb.

Reciprocating Circulating Pumps.

§ 163. General Remarks. — This class of pump may be made
either single or double acting, and is generally driven, in conjunction
with the air and feed pumps, by a beam from one of the crossheads of
the main engine. The stroke of the pump is usually half that of the
main engines, and the piston speed may reach 490 feet per minute.

The principal dimensions of a double-acting circulating pump may be
arrived at as follows : —

^ 100x«

and those of a single-acting pump —

/xj =

Q X I.H.P.

50 x«

where /= area of the pump piston in square feet.
J = stroke „ „ feet.

n = number of double strokes per minute.
Q = quantity of cooling water per i.h.p. per hour in cubic feet,
I.H.P. = the indicated horse power of the main engine.

The quantity of cooling water required per i.h.p. per hour is —

8-8 to 10*6 cubic feet in compound engines.

7 „ 8 "8 „ triple and quadruple expansion engines.

The above equations allow an efficiency of 83*3 */^ for the pump
itself.

1

296

MARINE ENGINES AND BOILERS.

B&

LVwk^'.WA.' ■

=^

T

Fig. 294.

.

I •

>

PUMPS. 297

§ 164. Pump Valves. — The suction and delivery valves should be
of such dimensions that the mean velocity of the water through the net
sectional area of the valve-seat openings is about 8 to 10 feet per second.
The lift of the valve should not, if possible, be more than y\ to ^ inch,
and the mean radial velocity of discharge at the circumference of the
valve not more than 13 to 15 feet per second. The valves and valve
seats are of the same type as those used in air pumps.

To avoid water-hammer, snifting valves are placed between the suc-
tion and delivery valves, and an air vessel, the capacity of which is equal
to the volume of the air pump, is fitted on the discharge side of the
delivery valve.

§ 165. Suction and Delivery Pipes.— The mean velocity of
the water in the suction and delivery pipes of the circulating pump
should be about 8 to 10 feet per second, and should be less in relatively
long and small pipes than in short large ones. In order that the circu-
lating pump may be used, in case of emergency, for pumping water
out of the ship, the suction pipe from the sea is provided with a branch
leading to the bilge, and in German warships to the main bilge suction
pipe. This connection must be provided with a non-return screw-
down valve. The diameter of this auxiliary bilge suction pipe is about
0-6 to 0*8 that of the main suction pipe.

§ 166. The Pump Body of the circulating pump is made of cast
iron, of about the same thickness as that of the air pump, and is
provided with a gunmetal liner, the thickness of which is about —

j = 0-015d + -25 inch.

D being the diameter of the pump barrel in inches.

The circulating pump generally forces the water direct through the
condenser, and is placed close to the air pump, so that both pumps may
be driven off the same beam. Sometimes, however, the air and circu-
lating pumps are driven from different crossheads of the main engine.

§ 167. Plunger and Pump Rod. — In double-acting circulating
pumps the plunger is made solid, and either packed, like that of the air
pump, with hemp, or provided with grooves. In the latter case the
depth of the plunger is from two to three times that of the air-pump
plunger. The pump rod is made similar to and of the same size as that
in the air pump. It is good practice to make the pump rods of the air
and circulating pumps of the same size^ because then only one spare
pump rod need be carried for both pumps.

298

MARINE ENGINES AND BOILERS.

Centrifugal Circulating Pumps.

§ 168. General Remarks. — Of late years centrifugal pumps have
been much used to pump the cooh'ng water for the condensers of large
engines on merchant vessels. In warships they are employed almost
exclusively. They have the advantage of working without shock, and of
delivering a uniform supply of water. With these pumps dangerous
strains upon the castings, pipes, &c., such as sometimes occur with
reciprocating pumps, are entirely avoided.

§ 169. Suction and Delivery Pipes. — The velocity v of the

water through these pipes is generally taken at 500 to 600 feet per
minute. In relatively short, large pipes, a higher velocity of the water
may be taken than in long small pipes.

In warships a higher velocity of water through the pipes is often
allowed, at maximum power, than in merchant vessels, because it is only
in exceptional cases that the maximum power in the former is developed,
and weight of both water and material must be economised. The requi-
site sectional area /in square inches per i.h.p. is given in Table No. 28.

Table No. 28.
Sectional Area of Pipes.

Velocity of the Water

in feet

per minute.

700 1

Quantity of Watei

• per I.H.P. per houi

r in pounds.

450

TwO

650

034 sq. in.

•044 sq. in.

051 sq. in.

600

•029 „

•035 „

•044 „

500 i

1

•024 „

•031 „

•037 „

The total sectional area (f) is obtained from the formula —

F=/X I.H.P.

Both suction and delivery pipes should be made as short, and with
as few bends, as possible. If there are two circulating pumps to one
main engine, each is generally provided with a separate valve to the sea,
and a separate sluice valve on the discharge pipe to the condenser, so that,
should one of the centrifugal pumps be damaged, it may be shut off; or
blank flanges inserted in the pipes will answer the same purpose. In
twin-screw steamers each engine and condenser generally has its sei)arate
centrifugal circulating pump. The delivery pipes are often connected
together, and fitted with a sluice valve. In case of need one pump may
be made to supply water to both condensers.

PUMPS.

299

As the centrifugal pump is also intended, should the necessity arise,
to pump out the bilges, the suction pipe is made with a branch leading
either to the engine-room bilge, or, as is customary in warships, to the
main drain or emergency bilge suction, through a valve which can
be either a sluice or a non-return foot valve. The inside diameter of
this connection is about 0*6 to 0*8 that of the delivery pipe of the
pump. When pumping from the bilges, the pump only gives about half
the usual discharge.

§ 170. Pump Vanes. — The inner diameter d (see Fig. 295) of the
wheel is generally —

d= 1*1 to I'idi (dy being the diameter of the suction and delivery pipes).

The overall diameter d = 2 to 2-6^.

It should be noted that the
higher values of d must be

chosen, if the head of water is id \^ Xk ^!

considerable. The overall dia-
meter of the vane wheel is de-
termined with respect to the
number of revolutions per minute
at which it is designed to run, and
to the shape of the vanes. (See
Fig. 296.) The clear width 3,
Fig. 295, at the inner circum-
ference of the wheel is so cal-
culated that the radial speed at
this point is about 3-3 to 5*0 feet
per second. It varies generally
from B = 0-23 to O'U. The clear
width b at the periphery of the
wheel is either made equal to b, or

J ^ • ^„L^d Fig. 295.

reduced m proportion to ^ = B-.

The water may enter the wheel either from one or from both sides.
(See Figs. 295 and 297.) If it enters at one side only, the wheel exerts
an axial thrust which must be taken up by carefully adjusted collars,
or some similar arrangement.

The vane wheel is generally of gunmetal, the vanes being cast with
it in one piece. Sometimes the vanes are made of sheet copper, and
side discs are riveted and soft soldered on to them. These side discs
should fit very closely, at least at their inner circumference, to the pump
casing, with only about j}^ to yV ^^^^ clearance at each side, so that as

-i

•>i ©

300

MARINE ENGINES AND BOILERS.

little water as possible flows back from the delivery to the suction
chamber. The discs are often made to fit into the pump casing with
very little clearance at the side of their outer, as well as of their inner
circumference.

The shape of the vanes may be varied considerably. The angle a^
(Fig. 296), formed by the vane at the inner circumference of the vane
wheel, should be so calculated that the water enters with the least
possible shock. Therefore it must be —

tan tt, = —

Here e^j = circumferential speed of the wheel at its inner periphery.
Vr = radial velocity of the water on entering.

Fig. 296.

To simplify their construction the vanes are generally built up of
arcs of circles whose centres lie along the line ab (Fig. 296) at right
angle to the blade ii.

To obtain a given difference of pressure between the suction and
delivery pipes, the curved blade i requires the maximum, and the curved
blade iv the minimum circumferential velocity, or number of revolutions
per minute. The blades ii and iii lie in between these values. The
circumferential velocity 7'., in feet per second at the outer edge of the
wheel is —

Vo =

10 /

H

l+sin(a + ffl
sin (a - (i)

The angles a and fi are assumed.

H = total head against which the water is pumped, in feet.
=■ (actual head of water + head due to friction in the pipe).

Fig. 297.

The radial velocity i', (and at the same time the width of the cross
section of the wheel at its inner periphery) is assumed ; the angle a^ is
most easily determined by constructing a velocity parallelogram. The
angles o and P must be so chosen that the radial velocity r,' of the water

302

MARINE ENGINES AND BOILERS.

leaving the wheel case is about 3-3 to 5*0 feet per second, and that the
number of revolutions of the wheel «, when circulating the water through
the condenser, is from 140 to 360 per minute. The circumferential

Fig. 298.

velocity v^ is taken at about 25 to 40 feet per second, and the
equivalent head due to friction h^ under ordinary conditions at about
^j = 5 to 8 feet.

If the circulating pump discharge is below the water line, the

PUMPS.

303

centrifugal pump in ordinary work has only the head due to friction
to overcome. In this case h = >4i = 5 to 8 feet. If the discharge
is at the height h^ above the water line, then in ordinary working
the head of water will be equal to A„ and therefore h = ^^ + ^2 ^*^^^- ^^
Ag denotes the height of the level of the water outside the ship above
the bilge-suction inlet of the circulating pump, then when pumping from
the bilges —

Hj = Aj + ^3 if the discharge is below the water line.
Hj = Aj + ^2 + ^3 ^^ ^^^ discharge is above the water line.

Fig. 299.

§ 171. The Centrifugal Pump Spindle is made either wholly
of phosphor bronze, or of steel cased in gunmetal, to resist the action of
the sea water. The wheel is usually fitted on to the spindle by means
of a cone, key, and nut; but it is frequently keyed on to a parallel
shaft, and held by a nut against a collar at the other side. (See
Fig. 297.) If the water enters at one side only, the wheel is generally
overhung with one side bearing in the casing. But if, as is usual in

304

MARINE ENGINES AND BOILERS.

i

IT / >- 1 '-N\jXr
7 / I TP--^

_-_- — ^ — ^ ' t —

^

zz

Fig. 300.

Fig. 301.

PUMPS.

305

larger centrifugal pumps, the water enters on both sides, the spindle is
sup|x>rted by bearings at either side, with stuffing boxes where it passes

Fig. 302.

through the casing. (See Figs. 297 and 301.) The pump spindle is
generally made separate, and bolted by means of a flange coupling to
the crank shaft of the engine driving it.

u

306 MARINE ENGINES AND BOILERS.

§ 172. The Pump Casing of the centrifugal pump is made of
cast iron in merchant vessels, and of brass in warships (or occasionally
but not often, partly of brass and partly of copper sheeting). In larger
pumps it is divided in half across the spindle (Fig. 298), while in
smaller pumps, where the water enters at one side only, it has a circular
side cover, by the removal of which the wheel can be lifted in and
out. The shell of the casing is so shaped that the sectional area of the
delivery space is gradually enlarged, until the full area of the delivery
pipe is reached. The thickness of the cast-iron casing and of the so-
called "vortex chamber" is from i to f inch, according to the size
of the pump. If made of brass, the thickness is about y\ to ^% inch.

§ 173. Engines for Driving Centrifugal Pumps. — To deter-
mine the dimensions of these engines, the work done by the pump when
pumping from the bilges is generally taken as a basis for calculation,
because the power required under these circumstances is much greater
than that needed for the ordinary work of the pump.

The indicated horse-power of these engines is approximately —

Q being the quantity of water pumped per minute in pounds, and
H the total head of water in feet when drawing from the bilge. The
above work should be obtained with about 0-75 of the normal
boiler pressure. The required number of revolutions n^ of the centri-
fugal pump, when pumping from the bilge, is determined from the
circumferential velocity necessary to overcome the total head of water
Hp including all frictional losses.

For small pumps these engines are generally made single cylinder,
and for large pumps, compound. In the latter case there are frequently
two pumps. The principal dimensions are —

^ . , ,. , J. 39,760 X i.H.p. ^ 51,120 XI.H.P.*

For smgle-cyhnder engmes / x s = — ^— to —

py.n py.n

, . ^ 102,240 X i.H.p.
For compound engmes fx^s^ —

Ratio of the cylinders is about — h.p. to l.p. = 1 : 3*5 to 1 : 3.

* Circulating pumps of still larger dimensions are not infrequently met with.
In these the steam supply has to be considerably throttled, and consequently the
pumps are much larger than is necessary, and their steam consumption is correspond-
ingly high.

PUMPS. 307

Here f— the area of the piston in square inches (in compound engines
the area of the low-pressure cylinder).
s = stroke of the piston in feet.

/ = absolute boiler pressure in pounds per square inch.
« = number of revolutions per minute when pumping from the
bilges (/.e., against max. head).
i.H.p. = the indicated horse-power of the engine when pumping from
the bilges (/>., against max. head).

J5 174. Particulars of Surface Condensers.— The following

table gives particulars of surface condensers taken from actual
engines : —

308

MARINE ENGINES AND BOILERS.

Table No. 29.

Particulars of

Type of Ship.

Small
Cargo Boat.

Medium'

sized

Cargo BoaL

Laree
Cargo boat.

Mail
Steamer.

Number and i.h.p. of main engines

1x700

1x2000

2x2100

2x4500

Number of revolutions per minute -

75

70

85

90

Number of condensers and - \

sq. ft.

sq. ft.

sq.ft.

sq. ft.

Cooling surface (in sq. ft. ) for etuh -

1 X 1076

1x9012

1 X 3120

1x6025

main engine - - - -

1

1

/Number of vane wlieels

per mam engine

—

1

1

1

•

External diameter of

•:3 £

each wheel

—

aOin.

43-3 in.

51 in.

&§.

Internal do. do.

2 X 13-7 in.

2x15 in.

2 x 19-6 in.

ntrifu
ting]

Number of circulating
engines per main

engine

—

1

1

1

Cylinder diameter of

•S

circulating engines -
Stroke of circulating

~~"

8-6 in.

9*8 in.

— ^

\ engine

7 in.

7 in.

•■^

g» /Number per engine

1

. ^

„^^

£go||ll (

Diam.

•Q-z: -g g { Diameter of cylinder )

9-8 in.

4^ S a g. and stroke - - }

Stroke

« o \ (

15*7 in.

%

/Type and number per J

1 attached
to main

1 attached
to main

1 attached
to main

1
Blake

mam engme - - i

■

•

•

I

engines.

engines.

engines.

Diameter of steam

1

cylinder -

—

—

2 x 12 in.

Diameter of pump

■

cylinder -

15-7 in.

26-7 in.

23-6 in.

2x25in.

B

3

Stroke

15*7 in.

23 -6 in.

19-6 in.

18 in.

c^ Volume swept through

1

h4

per I.H.P. per mmute

« = 17 '

cubic feet -

190

268-3

202

39 1

Ratio

21 1

15-5

18-5

33-8

I., p. cylinder

(for each

Air-pump cylinder

pumpcyl.)

\

PUMPS.

309

Surface Condensers.

•♦DcuLsch-
land."

"Kaiser
Wilhelmll.'

2 X 16500
75

sq. ft.
1x21520

47 in.
2x19-6 in.

11 X 19*6in.
irSin.

4x10000
80

sq. ft.
1x11700

1
Blake

2 X 18 in.

1 X 44 in.
24 in.

»= 15

.%
35-3

(for each

pump cyl.

referred to

both I-. P.

cyls.)

1

51 in.
2 X 19-6 in.

1
12 in.
9-8 in.

Russian

Cruiser

"Bogatyr."

2x10000
150

sq. ft.
1 X 10760

Building for the German Navy in 1903.

Gunboat.

1
Weir

2 X 10 in.

2 X .^3 in.
31 in.

» = 15

31-4

391

(for each

pump cyl.)

47 in.
2x23-6 in.

1

2x7-8 in.
9-8 in.

2x650
180

sq. ft.
2 at 807

Small
Cruiser.

2x5000
150

sq. ft.
2x5380

1

19-6 in.
2x7*5 in.

4-3x7-8 in.
4-3 in.

1

39-3 in.
2x19-6 in.

.c

7-5x11-8

i in.

Large
Cruiser.

3 X 6330
120

sq. ft.
3 X 7530

Battleship.

3x5330
115

sq. ft.
3x6260

1

43-3 in.
2x19-6 in.

7 X 12 in.
7*4 in.

1

Weir

2x12-5 in.

2 X 37 in.
15 in.

/i = 15

28

17-0

(for each

pump cyl.

referred to

both I.. P.

cyls.)

2 attached
to main
engines

12 in.

7-8 in.

295

both pumps
9-8

(for both
pump cyls.

together)

2 attached
to main
engines

18-5 in.
lo in.

139-4

both pumps

20-6

(for both

pump cyls.

together)

1

47-2 in.
2x23-6 in.

6-3x11 in.
9-8 in.

Duplex

2 X 12 in.

2x31 -5 in.
19-6 in.

« = 15

42

16-5

(for each

pump cyl. )

1
Weir-
Duplex

2 X 10 in.

2 X 31 in.
14 in.

w=20
46

22-8

(for each

pump cyl. )

Feed Pumps.

§ 175. Classification. — Boilers are fed by pumps driven from the
main engine and by steam pumps, and in small vessels even injectors
or hand pumps are used. It is now made compulsor}' to have at
least two independent feed systems for each set of boilers, each
capable of providing the full amount of feed water required to work
the boilers. In practice, however, there are often three feed-supply
systems entirely independent of one another, and in large ocean-going
steamers even more, so as to render impossible, as far as may be,
any breakdown in the feed arrangements. Where steam is required for
electric lighting, winches, &c., when the main engines are not running,
two independent feed systems must be provided. In such cases, two
independent feed pumps are required ; or for small and medium-sized
boilers, one pump and one injector are sufficient.

§ 176. Amount of Feed Water required. — Practice has shown
that the number of cubic feet of feed water per i.h.p. per hour ^, which
must be supplied to the boiler, may be taken at —

q = 0-32 cubic feet or 20 lb. for compound engines.
^ = 0*24 to 0*26 cubic feet or 15 to 16*5 lb. for triple and quadruple
expansion engines.

These quantities include not only the steam required for the main
engine, but also that for the auxiliary engines and steam heating, as
well as losses due to radiation, &c.

1. Pumps Driven Direct from the Main Engine.

§ 177. General Arrangement— Engines up to 100 i.h.p. have
frequently only a single feed pump. Larger engines have generally
two feed pumps of equal size, each capable of supplying the full
amount of feed water required for the boilers, when the ship is going
at full speed. The pumps are generally placed close to the air and
circulating pumps, and are driven from the crosshead in common
with the other pumps. In engines running at more than 200 revolu-
tions per minute, the feed pumps are often driven by a worm and
worm wheel from the crank shaft. They are then placed horizontal or
inclined, and parallel to, or on the longitudinal axis of the engine. If

PUMPS. 311

there are two feed pumps, each suction and delivery is as a rule fitted
with a valve, so that either pump may be worked independently. To
prevent abnormal strains occurring, due to excessive rise of pressure
in the pumps, relief valves are fitted, either on the pump itself, or on
the delivery pipe between the pump and delivery valve. The spring
of the relief valve should be set so as to allow the valve to lift at
about 1*3 times the boiler pressure.

S 178. Size of Feed Pumps. — Each separate pump must be of
such dimensions that it can easily deliver the full quantity of feed water
required for the boiler. Experience has shown that the pumps are
large enough if they are of such dimensions that, with an efficiency
= unity, each pump can deliver from one-and-a-half to twice the amount
of feed water required.

Pumps driven from the main engine have plunger pistons, and hence
are single acting. The volume of each pump works out at —

where /= area of the plunger in square inches, s — stroke in inches,
q = quantity of feed water per i.h.p. per hour in pounds.

In twin-screw steamers the pumps for each engine are designed to
deal with from 1*3 to 1*5 times the amount of feed water required for
each engine.

J5 179. The Barrels and Valve Boxes of the feed pumps are

generally of cast iron in cargo boats, and always of gunmetal in war-
ships. As a rule the barrel and valve boxes are cast separately, and joined
by flanges, in order that, if one part is damaged, it alone has to be
replaced. If the internal diameter of the pump is larger than that of
the plunger, the valve box must be connected to the highest part of the
clearance space thus left, in order that no air cushion may be formed.

The thickness of the barrel S is taken at S= ,7! + *28 inch for cast iron,

5= J- -J- '16 inch for gunmetal, d^ being the internal diameter of the

pump cylinder. Pump rods in cargo boats are made of steel, and
frequently cased with gunmetal, the cases being held in position by
cap nuts. The pump rod is made tight with an ordinary stuffing
box packed with hemp or metallic packing. Each feed pump driven
from the engine has an air vessel, the capacity of which is about 2 '5
times the volume swept through by the plunger.

312

MARINE ENGINES AND BOILERS.

Fig. 303.

Fig. 304.

PUMPS.

313

314 MARINE ENGINES AND BOILERS.

§ 180. The Pump Valves, in smaller pumps up to about 2 J inches
diameter, are usually of the mushroom or ball valve type ; in larger
pumps they are of the Kinghorn or of a similar type. If mushroom
valves are used for larger pumps, there should be several small instead
of one large valve, the object being to have sufficient sectional area of
valve opening, while reducing the lift (about ^ to ^.j inch) and weight
of valve to a minimum. To make the valves close quickly at ihe change
of stroke, they are sometimes fitted on the back with helical springs of
" delta " metal. The valve seats are made separately, and fitted into
the valve chests, so that they may be easily taken out and renewed. If
solid valves are fitted the valve seatings are made about ^ to g\ inch
wide. Between the suction and delivery valves a small snifting valve
is usually fitted, and a small air cock is sometimes placed in the lower
part of the air vessel, the cock being connected to the hot well. This
not only allows the air collected there to escape, but at the same time
prevents any water being wasted from a too free use of this cock.

J5 181. Velocity of the Water. — The net sectional area of the
openings in the valve seat is calculated to allow a mean velocity of the
water through it of about 6 5 to 8*25 feet per second. The radial
velocity of the discharge at the circumference of the valve should be
about 20 to 26 feet per second, and the lift given to the valve I to ^^.r
inch, to prevent shock or jarring at the change of stroke.

Suction and Delivery Pipes, — The diameters of these are so calculated
that the mean velocity of the water in the suction pipe is about 7 feet
per second, and in the delivery pipe about 10 feet per second.

Remarks. — In calculating the different areas such as the bore of the
pipes and the net openings of the valves, the mean speed and sectional
area of the pump plunger must always be taken as a basis, and it is
unnecessary to consider whether the pump is single or double acting.

2. Independent Feed Pumps.

§ 182. Steam Pumps are generally used, in small engines, as
auxiliary feed pumps only, while in larger engines they are frequently
employed to pump the whole of the feed water required. In the
latter case the working of the pump may be controlled automatic-
ally, and its speed made to correspond to the quantity of feed
water required at any given time. The type of steam pump most
generally used is either the Duplex or Simplex. These pumps, which
may be either vertical or horizontal, are double-acting, and have no
rotating parts. The steam slide valve is driven either directly or
indirectly by the piston rod. The diameter of the steam cylinder is
about r4 to 1*7 that of the corresponding pump cylinder. For the

PUMPS. 315

steam cylinder cast iron is used, and for the pump cylinder either cast
iron with a gunmetal liner and valves, or the pump and valve box are both
made of brass, and cast in one piece. It is best to have the pump
plunger and pump rod of phosphor bronze, or some similar material.

§ 183. Duplex Pumps.— Figs. 303 to 305 show a horizontal,
and Figs. 306 and 307 a vertical Duplex pump, as made by the Worth-
ington Pump Co. There are two steam cylinders of equal size, each
one driving a pump plunger ; the movement of the pistons corresponds
to that in an engine with two cranks at 90°, and two eccentrics at an
angle of 90* with the cranks. Each piston rod works the slide valve
of the other steam cylinder. One of the levers transmitting the valve
motion must be single armed, the other double-armed, as can easily be
demonstrated if imaginary cranks and eccentrics are drawn for both
gears. The valve of each cylinder is an ordinary D slide or piston
valve, but the cylinder has five ports. The two outermost lead to the
two ends of the cylinder, and are for admission only ; the two inner
open into the bore of the cylinder, and take the exhaust ; the central
port is also for the exhaust. Outside and inside lap, ^V to |\ inch ;
utmost travel of the valve = 2 x (lap + height or breadth of a port) ;
mean speed of the steam through the ports, 4,000 to 6,000 feet per
minute. Maximum number of double strokes per minute = 40 to 50,
but more generally 20 to 30. The motion of the valves is often regu-
lated by adjustable nuts on the valve rods, so that the steam ports may
the longer remain fully open, and the stroke of the pump be varied.

As the motion of the valves is alternating, it follows that both
cylinders are interdependent, and must work at the same time. To
prevent the steam piston from striking against the cylinder covers, the
exhaust ports opening into the bore of the cylinder are closed by the
piston before it reaches the end of its stroke, and the steam remaining
in the clearance space thus forms a cushion. In the new Duplex pumps
made by the Blake & Knowles Steam Pump Co., the openings of the
outer ports in the valve face of the steam chest are not outside, but at the
side of the openings of the inner ports. The former serve for admission
only, the latter for admission and exhaust. The length of the valve is thus
diminished, but at the cost of the breadth. There is also an arrange-
ment for regulating the compression by means of four small compen-
sating valves between the outer and inner ports. A somewhat large
Duplex pump for a duty of 120 tons per hour, which may be used as
a donkey pump or for the ash-ejector, is shown at Figs. 308, 309, Plate
XIV. It has a balanced slide valve, and a pump plunger with a packing
ring of white metal. Each of the steam cylinders has a diameter of 14
inches, each of the pump cylinders of 9 inches. Stroke 12 inches. In

316 MARINE ENGINES AND BOILERS.

the newer kinds of pumps, which work with clean cold water, leather
rings are often used for packing the pump plungers.

§ 1S4, Simplex Pumps.— These have only one steam and one
pump cylinder. Their gearing and method of working are so com-
plicated thai they can hardly be explained without a model. Two such
pumps, each independent of the other, are frequently combined into a
twin Simplex pump. (See Fig. 310.)

Fig. .we.

Fig. 30;

All Simplex pumps arc worked by what may be known as indirect valvt
gears. A small auxiliary steam valve, driven from the piston or piston rod,
works the main distributing valve controlling the steam supply to the
cylinder. If there is only one direct driven valve, dead points cannot be
avoided, further it is impossible to start the pump in any position, and
when working it is very liable to come suddenly and unexpectedly to a
standstill.

PUMPS. 317

§ 185. Weir Pumps. — {See Fig. 310, showing a twin Simplex
pump.) For steam distributing valves of Weir pumps see Fig. 311.
The semicircular main slide valve d has rounded ends m, and m.^ at
either side, which work to and fro in the horizontal auxiliary steam
cylinders h^ and h^ of the valve chest. In the vertical recess at the
back of the main valve, an auxiliary slide valve c, driven from the
piston rod through a single-armed lever, works parallel to the cylinder
axis. Of the two steam passages, ai leads to the bottom end, «„ to the

Fi^. 310.

top end, of the steam cylinder. In the position shown at Fig. 311 the
lower part of the valve is receiving live steam, while the upper part
communicates with the exhaust b, in this position therefore the piston
ascends. It is only after it has returned through about 0-75 of the
stroke, that it Ukes the auxiliarj- valve c with it. The latter gradually
shuts off «), so that the steam in the bottom end of the steam cylinder
expands. At the same time the passage / leading to the auxiliary

318 MARINE ENGINES AND BOILERS.

cylinder A, is opened to exhaust by the auxiliary valve c, and the passage
fa leading to the auxihary cylinder h,„ is opened to admit live steam,
so that as the pressure in h^ is above that of the atmosphere, the main
valve d is forced to the other side of the valve chest. In this way, a^
being open to live steam and a, open to exhaust, steam is admitted to
the piston at the lop end, and drives it down. To prevent the main
valve rf striking the cover of the auxiliary steam cylinder k in its travel,
it is so arranged that it closes the port/j before it has reached its extreme
position, and the steam in A' is compressed, forming a steam cushion.

Weir pumps, similar to Blake pumps, can be r^ulated while run-
ning, by adjusting the lock-nuts on the valve rod. The steam and pump
cylinders are braced together by strong steel columns. The pump
cylinders or barrels are connected by suction or delivery valve boxts,
or by means of both.

Fig. 311.
In Weir's twin air pumps (see Fig. 293) the two steam cylinders are
worked by one main valve, common to both. The photograph shows
clearly the action of the lever working the corresponding auxiliary valve ;
and also the valve box in front, with the two jointed covers at the side,
and the two expansion valves.

S 1S6. Blake Pumps. — For steam distributing valves see Fig. 313.
In these pumps the steam is distributed through a slide valve d, having
a hollow circular back, upon which works the auxiliary valve h, which

ruMPS.

319

Fig. 3] 2.

320

MARINK ENGINES AND BOILlikS.

can be rotated. The ends are shaped like pistons, and work in the
horizontal passages m m of the valve chest. Towards the end of the
stroke of the steam piston, the valve rod o turns the auxiliary valve A
once, through lever it ; and thus the recesses (t e in the auxiliary valve A
form a connection on one side of the valve between the admission steam
passage /, and on the other side between the exhaust passage / and the
auxiliary steam cylinders m m. The auxiliary valve h is thus forced to
the other side, and takes with it the main valve d. Towards the end of
the stroke the main sleam ports a are covered by the piston rings, and
the steam remaining in the cylinder is compressed as in the Duplex
pump. At the beginning of a stroke the steam is admitted through the

KvxiuAirt tju.vE

I
Fig. 313.

auxiliary ports *, which are connected to the main steam ports through
the compression valve c.

To Adjust the Pump. — The pump is set to work dead slow against
a pressure, and the lock nuts on the valve rod adjusted so that the
steam piston knocks against the top and bottom cylinder covers.
These positions are marked on the columns of the pump, and the nuts
brought closer together, until the pointer on the piston rod is from 0'2
to 04 inch short of the marks.

The steam and pump cylinders are connected by three strong
wrought-iron columns. These pumps are set up either singly, or,
like Weir pumps, in sets of two side by side, and connected by the
suction and delivery valve boxes.

Auxiliary Pumps.

§ 187. Bilge Pumps Driven by the Main Engine— These

are as a rule made similar to the feed pumps, and the diameter of the
pump plungers is the same. Ships with engines of less than 100 i.h.p.
have generally only one bilge pump. Ships with larger engines have at
least two, one of which draws water direct from the engine-room bilge,

Fig. 314.

and the other is arranged to draw from all the compartments of the
ship. The pump barrel and valve boxes are generally of cast iron, but
occasionally of brass. The valves are usually rubber flap valves, or in
small pumps solid gunmetal mushroom valves. The suction and de-
livery valves, instead of being placed one above the other, are set side
by side, and can be reached through separate covers. The advantage of
this arrangement is that the suction valve can be easily examined. The

X

MARINE ENGINES AND BOILKRS.

a. Bilge pump.

t.

c. Keed pump.

e. Air pump.

g. Feed pump.

PUMPS. 323

gunmetal plunger is hollow, and is fixed to the steel pump rod by means
of a coned end, and a cap-nut underneath. The pump rod is con-
nected to the crosshead common to all the pumps by means of a cone
and nut, in the same way as the feed pump rod. The air vessel is
of the same capacity as that of the feed pump. In large ships the
delivery pipe leading from the air vessel to the discharge valve in the
side of the ship is often so arranged that it can be shut off at the air
vessel by an automatic valve; by this means, when the pump valve
box is opened for repairs, the water in the discharge pipe does not
run back. This valve must of course open freely when the pump is
working.

§ 188. Sanitary Pumps. — These are generally of the same
dimensions, and arranged in the same way, as the feed and bilge pumps,
unless sp>ecial conditions render necessary a larger quantity of water
than usual. In small ships carrying two bilge pumps, one of them
is generally arranged to act either as a bilge or as a sanitary pump.

§ 189. Arrangement of Pumps.— Figs. 3U, 315, 316 show the
method of arranging the pumps in an ordinary merchant vessel, when
they are driven from the main engine through levers. The two levers
drive a crosshead through two connecting links. The air pump rod e
is connected to the middle of the crosshead ; immediately to the right
and left of it are the two feed pumps c and g (with steel plungers), and
outside these again the bilge pump a on the left, the sanitary pump /
on the right, both with gunmetal plungers.

§ 190. Separate Steam-driven Pumps. — Where these are used
the bilge pump, sanitary pump, fire pump, and circulating pump may
be either of the Duplex or Simplex type. As a rule, in smaller ships, the
donkey feed pump is arranged to perform these various functions. In
large ships, on the other hand, special pumps for the bilges, for delivering
sea water on deck, &c., are provided, so that the various demands may
be satisfied at the same time, and the donkey feed pump not soiled
with the bilge water. The latter is generally used to work the ash
ejector, and is arranged in the same way as the steam feed pump.
The principal difference between them is in the heavier or lighter con-
struction of the various parts, according to the purposes for which either
pump is required. To pump out the ballast tanks, if there are any,
special steam pumps are generally fitted. These are usually Duplex,
but sometimes centrifugal pumps or pulsometers are used. See Lloyd's
Rules concerning pumps, pipes, &c., § 193, page 329.

324 MARINE ENGINES AND BOILERS.

Pump Rods.

§ 191. General Remarks.— The pressure on the plunger of the
air pump is taken as a basis for determining the dimensions of the
rod driving the pump, when worked from the main engine. This
pressure, including the friction of the stuffing boxes and plunger, is
assumed to be 28*5 lb. per square inch of piston area. It also serves
as a basis for designing the circulating pump rods. The allowable
stress in the bearing and connecting-rod bolts of the pump-rod gear
is taken at 3,500 to 6,500 lb. per square inch. The stress allowed
in the rocking beams, which have to withstand a bending strain, can
be taken as 4,000 to 5,000 lb. per square inch.

The sizes of the various pins and journals are calculated by allowing
for a pressure on their working surfaces of from 425 to 700 lb. per
square inch. Their length is generally slightly greater than their
diameter.

The crosshead or beam to which the air pump, feed water, and
bilge pumps are attached, is chiefly subjected to bending strains, and
its sectional area must be determined in accordance with these strains
in each particular case. The allowable stress in it, due to bending,
may be about 3,000 to 5,000 lb. per square inch.

Feed pumps are generally fitted towards the forward end, and bilge
pumps towards the after end of the pump crosshead, but to ensure a
more uniform strain upon the rods, one of the feed and one of the
bilge pumps are frequently placed side by side.

§ 192. Different Pump Arrangements taken from actual

practice. — In the following tables the pump systems of different types
of ships are classified, to show from whence the pumps of each system
draw, and where they deliver to : —

{a) Table No. 30. — Pump arrangement of a river tug.

(d) Table No. 31. — Pump arrangement of a small cargo boat.

(c) Table No. 32. — Pump arrangement of a large cargo and

passenger steamer.

(d) Table No. 33. — Pump arrangement of a large cruiser of

the Imperial German Navy.

The different uses to which the pumps can be put are plotted on what
is known as a " pump diagram."

PUMPS.

325

Table No. 30.
Steam Tug — /et Condensing.

Pump.

Draws from

Delivers to

Air pump (one)

Condenser
Bilge

The sea.

Engine feed pumps
(one or two)

Hot well

Main feed pipe (pos-
sibly through a sur-
face feed heater).

One independent
steam pump

The sea
Hot well
Bilge
Ballast tanks

To deck.

Auxiliary feed pipe.
'I'he sea.
Ballast tanks.

Injector

The sea

Auxiliary feed pipe.

Engine bilge pump
(one or two)

The sea
Bilge
Ballast tanks

The sea.
Ballast tanks.

Bilge ejector

Bilge

The sea.

Table No. 31.
Small Cargo Boat — Surface Condensing,

Pump.

Circulating pump
(double - acting,
driven by main
engine)

Air pump (driven by
main engine, single-
acting)

Two engine - driven
feed pumps

Draws from

The sea
Engine-room bilge

Condenser

Hot well
Condenser

Delivers to

Through the con-
denser to the sea.

Hot well.

Main feed pipe (pos-
sibly through sur-
face feed heater).

326

MARINE ENGINES AND BOILERS.

Pump.

Table No. 31 — continued.

Draws from

One engine - driven
bilge pump

One engine - driven
sanitary pump

Independent steam
pump

Ballast pump

Injector

Hand pump (for filling
the boiler)

Delivers to

Engine room bilge
Bilge piping

The sea

Engine room bilge

Bilge piping

The sea

Bilge piping

Condenser

Boiler

Reserve feed tanks

in double bottom
Ballast tanks

The sea
Ballast tanks
Bilge piping

The sea

Reserve feed tanks in
double bottom

The sea

Reserve feed tanks in
double bottom

The sea.

The deck.
Sanitary tank.
The sea.

Auxiliary feed piping.
The deck.

Through the con-
denser to sea.

The sea.
Ballast tanks.
Through the
denser.

con-

Auxiliary feed piping.

Auxiliary feed piping.

Table No. 32.
Large Cargo and Passenger Steamer — Two Engines^ Surface Condensing,

Pump.

! Circulating pump (one
centrifugal pump to
each engine)

Air pump (one steam-
driven air pump to
each condenser)

Draws from

Both condensers

Delivers to

The sea

Engine room bilge

Through the con-
densers of both en-
gines to the sea.

Auxiliary condenser.

Hot well.

PUMPS.

327

Table No. 32 — continued.

Pump.

Draws from

Delivers to

Two steam-driven feed
pumps (delivering
to feed heater)

Hot well

Reserve feed tanks in

double bottom
Condenser

The sea.
Feed heater.
Main and auxiliary
feed piping.

Two steam - driven
main feed pumps

Feed heater
Hot well
Boiler
Condenser

Reserve feed tanks in
double bottom

The sea.
Feed heater.
Main and auxiliary
feed piping.

Two steam donkey
pumps

The sea
Boiler
Condenser
Main bilge piping
Auxiliary bilge piping
Ballast tank suction
Cooling water from
refrigerating engine

The sea.
Ash ejectors.
Deck and fire hose.
Sanitary tank.

One ballast pump

The sea

Main ballast piping
Main and auxiliary
bilge piping

The sea.

Main condenser.

Auxiliary condenser.

Auxiliary condenser
circulating pump

The sea

Through auxiliary con-
denser to the sea.

Engine bilge pumps
(two per engine)

Engine-room bilge
Main bilge piping

The sea.

Sanitary pumps (one
per engine)

The sea

Deck and fire hose.
Sanitary tank.
Cooling water to bear-
ings.

Circulating pump for
refrigerating engine

The sea

Condenser of refrige-
rating engine.

Drinking-water pump

Drinking-water tank

Galley.

Drinking water filter.

328

MARINE ENGINES AND BOILERS.

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Lloyd's Rules for Pumps and Pumping Arrangements.

§ 193. Rule No. 20. The engines are to be fitted with two feed
pumps, each capable of supplying the boilers ; the pumps, &c., to be
so arranged that either can be overhauled whilst the other is at work.

21. The engines are to be fitted with two bilge pumps, which are to
be so arranged that either can be overhauled whilst the other is at work.

22. In engines of 70 h.p. and under, one feed pump and one bilge
pump will be deemed sufficient, provided they are of adequate capacity.
The main feed pumps may be worked by independent engines, pro-
vided they are fitted with automatic regulators for controlling their
speed. If only one such pump is fitted for the main feed, the auxiliary
feed pump required by paragraph 25 should also be fitted with an
automatic speed regulator.

23. A bilge injection, or a bilge suction to the circulating pump, is
to be fitted.

24. The engine bilge pumps are to be fitted capable of pumping
from each compartment of the vessel. The mud boxes and roses in
engine-room are to be placed where they are easily accessible, and to
the satisfaction of the Surveyor.

25. A steam pump is to be provided, capable of supplying the
boilers with water, this pump to be provided with suctions to the hot
well, and also to the sea. A steam pump is to be so fitted as to pump
from each compartment, to deliver water on deck, aiTd if no hand pump
is fitted in engine-room, it must be fitted to be worked by hand. In
small vessels in which only one steam pump is fitted, it must comply
with all these requirements.

26. In all steam pipes provision is to be made for expansion and
contraction to take place without unduly straining the pipes, and all
main steam pipes are to be tested by hydraulic pressure to twice the
working pressure, in the presence of the engineer surveyor.

27. All discharge pipes to be, if possible, carried above the deep
load line, and to have discharge valves fitted on the plating of the vessel
in an accessible position.

28. No pipes to be carried through the bunkers without being
properly protected.

29. Bilge suction pipes to be arranged to pump direct from each

330 MARINE ENGINES AND BOILERS.

compartment, the roses to be fixed in places where they can be easily
accessible.

A suction pipe from the bottom of the boiler to the steam pump
must be separated from all the other suction pipes belonging to the
pump, that the boiler pressure cannot, through carelessness or ignorance,
find its way into any of the piping.

39. Cocks and valves connecting all suction pipes to be fixed above
the stokehold and engine-room platforms.

40. The arrangement of pumps, bilge injections, suction and de-
livery pipes is to be such as will not permit of water being run from
the sea into the vessel, by an act of carelessness or neglect. Any
defective arrangement to be reported to the Committee.

41. In steam vessels the pumping arrangements, according to the
division of holds, &c., to be as follows : —

42. Holds with Double Bottoms, — In the double bottom of each
compartment of the hold and of engine and boiler space, a steam pump
suction is to be fitted at the middle line, and one on each side, to clear
the tanks of water when the vessel has a heavy list. Where there is
considerable rise of floor towards the ends of vessels, the middle line
suction only will be required. A steam pump suction and a hand pump
are also to be fitted to each bilge, in each hold where there is no well.
When there is a well, one or three steam pump suctions are to be fitted
in the same, according to the rise of floor, whether considerable or little,
and hand pumps are fitted at the bilges.

43. Holds without Double Bottoms. — Where there is considerable
rise of floor, one steam pump suction and one hand pump are to be
fitted in each hold. In vessels with little rise of floor, two or three steam
pump suctions and at least one hand pump to be fitted to each hold.

44. Engine and Boiler Space. — Where a double bottom extends the
whole length of engine and boiler space, two steam pump suctions are
to be fitted to the bilge on each side. Where there is a well, one steam
pump suction should be fitted in each bilge, and one in the well. Where
there is no double bottom in the machinery space, centre and wing
steam pump suctions should be fitted. The rose box of the bilge
injection is to be fitted where easily accessible, and is to be used for
bilge water only. The main and donkey pumps to draw from all com-
partments, and the donkey to have also a separate bilge suction in the
engine-room.

45. Fore and After Peaks, — If the peaks are fitted as water ballast
tanks, a separate steam pump suction is to be led to each. If not used
for water ballast, an efficient pump is to be fitted in the fore peak. If

PUMPS.

331

the after peak is used as a ballast tank, no sluice valve or cock is to be
fitted to the after bulkhead ; but if it is not so used, and if no pump is
fitted in it, a sluice valve or cock is to be fitted to the after bulkhead, to
allow water to reach the pumps when required.

46. Tunnel, — The tunnel well is to be cleared by a steam pump
suction.

47. All Hand Pumps to be capable of being worked from the upper
or main decks above the deep load water line ; the bottoms of the pump
chambers are not to be more than 24 feet above the suction rose, and
the pumps are to be tested by the surveyors, to ensure that water can be
pumped from the limbers. ^ The sizes of the hand pumps to be not less
than those given in the following table : —

I land Pumps in Holds.

Tonnage under Upper Deck.

Diameter of
Barrel.

Diameter of
Tail Pipe.

In vessels under 500 tons -

In vessels of 500 tons but under 1000

tons

In vessels of 1000 tons but under 2000

tons

I In vessels of 2000 tons and above

Inches.
4

Inches.

5

2i

In lieu of hand pumps in each compartment an approved fly wheel
pump may be fitted, if it is connected to the steam pump bilge suction
pipes of these compartments.

48. No Sluice Valve or Cock is to be fitted to the collision bulk-
head.

49. No Sluice Valves or Cocks are to be fitted to the engine-room or
other watertight bulkheads, unless they are arranged so as to be at all
times accessible.

50. When Sluice Valves are fitted, they must be so arranged as to
be controlled above the load water line, and the rods are to be boxed-in
to prevent injury.

51. Sounding Pipes to be fitted on each side of holds and ballast
tanks, and a doubling plate is to be fitted under each.

52. Air Pipes to be fitted to each ballast tank as required.

53. All Cocks and Valves in connection with bilge and ballast
suction pipes are to be fitted in places where they are at all times
accessible.

332

MARINE ENGINES AND BOILERS.

54. The Filling Pipes for deep tanks, which can be used for either
cargo or ballast, must be controlled by valves placed in an accessible
position, and so arranged that when the tank is being used for cargo it
will be impossible to fill it with water. This result is to be obtained by
taking out a short bend or wedge piece, and fitting blank flanges in its
place, or in some other way to be submitted to and approved by the
Committee.

55. The Pipes for bilge or ballast suctions are to be fitted with
flanged joints in convenient lengths, so that they may be easily discon-
nected for clearing. In the case of cast-iron suction pipes, which are
not also used as tank-filling pipes, or which cannot be subjected to sea
pressure, spigot and faucet joints made with indiarubber rings fitted
over the spigots might be adopted, except in the case of bilge suction
pipes passing through ballast tanks, which should be fitted with flanged
joints.

56. The Suction Pipes to fore and aft peaks and to the tunnel well
should not be less than 2^ inches inside diameter, except in vessels
under 500 tons under deck, in which case they may be made 2 inches.

57. The Bilge Injection should not be less than two-thirds of the
diameter of the sea inlet to the circulating pump.

58. The inside diameter of other bilge suction pipes should not be
less than that given in the following table : —

Tonnage under Upper Deck.

In vessels under 5(X) ions -

In vessels of 5(M) tons but under

l(K)Otons . - - .
In vessels of KKKI tons but under

150()tons . - . .
In vessels of 15<NI tons but under

2(XK)tons - - . -
In vessels of 20(10 tons but under

3000 tons . . . .
In vessels of 3000 tons and alx>ve

Engine-room

Centre Suction,

Separate Donkey

Suction, and Hold

Centre Suctions.

Inches.
2

2i

3

3i
3i

Wing Suctions in

Holds where no

Centre Suctions

are fitted, and

Wing Suctions in

Engine-room.

Wing Suctions in

Holds where

Centre Suctions

are also fitted.

Inches.
2

Inches.
2

2

2

21

2

2i

2i

3

2i
25f

In cases where more than one suction to any one compartment is
connected to the pumps by a single pipe, this pipe should be not less
than the size required for the centre suction.

PART III.

SHAFTING, RESISTANCE OF SHIPS,

PROPELLERS.

SECTION I.

SHAFTING.

Thrust Shaft and Thrust Block.

§ 194. Axial Thrust. — In the engines of all screw steamers the
thrust of the propeller is taken up by what is known as a thrust shaft.
It consists of collars forged on to the shaft which press against a thrust
bearing.

The thrust shaft and thrust block are calculated to withstand the
"indicated thrust," that is the axial thrust which would be produced by
the propeller, if all the power generated by the engine in i.h.p. were
utilised, without loss, in- driving the ship forward. If i.h.p. is the indi-
cated horse-power of an engine, n the number of revolutions per minute,
H the pitch of the screw in feet, ? the indicated thrust in pounds, then
the equation will be —

L2LI- X 33,000 = p X H.
n

That is, the work dotie during one revolution of the engine must be
equal to the work due to the thrust during one revolution of the screw,
acting through a distance equal to the pitch of the screw. From the
above equation we get —

p_ I.H.P. X 33,000
nH

If, for instance, the efficiency of the screw is about 65 %, and the
mechanical efficiency of the engine about 8.5 7o> ^^^^ t^^ *' effective
thrust " will be —

0*85 X 0-65 X p = about 0-55p.

As already stated, however, the dimensions of the thrust shaft and
bearing are calculated from the indicated thrust p.

S 195. Thrust Shaft. — In all large ships, with engines above
100 I.H.P., the thrust collars are on a separate length of shaft, which is
made as short as possible, the object being to simplify the construction
and erection of the thrust shaft, and also the fitting and stowing of
the spare thrust shaft. The thrust shaft is therefore only made long
enough to accommodate sufficient thrust collars, two bearings and the

3;tfi MARINE ENOINKS AND BOILERS.

nuts of the coupling bolts for the flange couplings. The diameter d of
the thrust shaft is generally equal to that of the tunnel shaft (see page
340), but it is better to make it thicker, and equal to that of the crank
shaft, as the thrust shaft is more likely to be subjected to bending
strain from the crank shaft than the tunnel shaft.

Fig. 317.

Diameter of the collars D=l-6 to l-9rf(Fig, 317).
Width of collars ^ = 0-13 „ d-lM in lightly built engines, or in

the case of strong shafts.
,, ^ = 01.5 „ 02(/ in heavily built or small engines.

For space between the collars s see § 196.

The number of rings is such that the pressure p exerted by the
" indicated thrust " upon their " effective area" is as follows : —
p= 40 to 55 lb. per square inch for cargo steamers.
p = 5.5 „ 80 „ „ passenger steamers.

/= 70 „ 85 „ „ heavy warships.

/ = 100 „ 130 „ „ light warships.

The coupling flanges of the thrust shaft are the same as those for
the crank shaft, and the material used in the construction of both shafts
is the same.

g 196. Tlirust Block.— In large ships the block is generally a
square trough of cast iron or cast steel, with bearings at each end
(see Figs. 321 to 324). Two heavy screwed rods are fixed one on each
side of the shaft, and to these the horse-shoe thrust collars are attached.
They are secured and adjusted by means of nuts on the two side rods.
The nuts transmit the thrust to the two rods, and through them to the
luain thrust-block casting, and thence to the ship. In large ships the
horse-shoe collars between the thrust collars (Figs. 318 to 320) consist
generally of hollow cast-iron or steel castings (the latter usually only in

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS.

337

warships). Their bearing surfaces are faced writh white metal fitted or
cast on in the ordinary way, and cooling water circulates through
the hollow interior. (See also S 267, page 450.) At the top of each
horse-shoe an oil cup is fitted, which supplies oil between the thrust
collar and the cap. The collars must be strong enough not to
bend with the thrust to which each is subjected. When cralculating
their dimensions it is well to allow very small stresses, as the thrust
is often distributed only over a portion of the horse-shoe. The lugs
supporting the horse-shoe caps on the side rods should he so arranged,
that the line connecting the centres of the rods passes through the

Fig. 320.

centre of gravity of the thrust surface. The latter are usually made
horse-shoe shaped in larger engines, as shown in Figs. 318 to 320.
In the lai^er collars there are generally two eye-bolts, to lift them in
and out more easily.

Material of the horse-shoe collars — Cast steel or cast iron, less

frequently bronze.
Thickness of the horse-shoe collars if solid j = 2 to 2-5i (see Fig. 3 1 7).
„ „ if hollow a = 2-5 to 3* „

Thickness of the white-metal liner

5-1- -08 inch.

MARIXK ENGINKS AXD BOILKkS.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 339

The side thrust rods are made of steel, and screwed with a fine
thread. The nuts between the horse-shoe caps are made of gunmetal, so
that they may not rust on to either the collars or rods. The two side
rods are fitted into strong eyes at either end with nuts on each side.

The allowable stress in the side thrust rods at the bottom of the
thread, due to the indicated thrust, is s = 5,500 to 8,500 lb. per square
inch. It is desirable to allow the horseshoe collars to fit against and
into the sides of the trough casting, either at the ears, or at the tower
part of each shoe. (See Fig. 322.)

Thrust Mock end bearittgs:—\jen%\)\ I =(i-» to l'2rf(Fig. 317). Brasses
are generally of cast iron or gunmetal, and are lined with white metal.

Fig. 325.

The thrust block is secured to the ship as firmly as possible. It is
best to connect it to the engine bed-plate, so that pari of the thrust may
be taken by the foundation bolts. In the engines of large merchant ships
the thrust block is frequently placed upon a separate bed-plate, which
is connected to the engine bed-plate, and to it the thrust block is
secured and fitted with adjusting screws. If the thrust block has no
separate independent bed-plate, it should be firmly wedged between
cleats strongly riveted to the body of the ship.

§ 197. Thrust Blocks in Small Ships.— These have no sepa-
rate end bearings, and are generally not connected to the engine bed-
plate. The thrust is ofien taken by gunmetal rings, made in two halves,
which fit into grooves in the body and cap of the thrust block. The rings
must be prevented from moving round. The cap should be stepped

340

MARINE ENGINES AND BOILERS.

or keyed into the lower part or body of the block. The whole of the
latter may be made either of gunmetal or cast iron.

In very small engines the thrust blocks are made of cast iron or gun-
metal, and the cap and body only are lined with white metal, in which
grooves are turned to receive the thrust collars. In order to transfer the
axial thrust upon the cap to the foundation, the cover must, as already
mentioned above, be stepped or carefully keyed into the lower part of
the block, or the bolts securing it must be an accurate driving fit
(Fig. 325).

Tunnel Shafts and Plummer Blocks.

§ 198. Intermediate or Tunnel Shafts. — These join on to the
after end of the thrust shaft, and transmit the turning moment produced
by the engine to the tail shaft. They are therefore exposed to bending
and twisting strains, but as they have to take fewer shocks and bending
stresses than the crank or propeller shafts, they are generally made
smaller than either of these. If ^ is the diameter of the crank shaft, the
diameter of the tunnel shaft will be ^j = 0-85 to Id.

For German Lloyd's Joules for Shafts see pages 345, 346.

The various lengths of the tunnel shaft are as far as possible made
the same, and their length is determined by the dimensions and type
of the ship. Care must be exercised that the shafts can be easily taken

K ^ i-

^

S

-n

Fig. 326.

in and out and lifted, this being often necessary when overhauling and
renewing the propeller shaft In large ships, where the diameter of the
shaft is from 10 to 24 inches (Fig. 326), /= 16 to 24 feet. The separate
lengths of shaft are joined up, like those of the crank shaft, by flanges.
The tunnel shafts are often made hollow, the proportions of the bore
being the same as for crank shafts. Close to each coupling is a bearing,
each journal {d^ being made slightly larger in diameter than the body
of the shaft to allow for wear.

A flange of one of the intermediate shafts is left very thick, so that,
when everything else is in place in the ship (Fig. 326), any differences
between the distance from the tail shaft to the thrust block, as originally
designed, and as actually built, can be allowed for. The aftermost

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 341

intermediate shaft is sometimes provided with a shoulder or collar on
each side of the last plummer block, to prevent the tail-end shaft from
falling out, in case of an accident to the tunnel shaft.

§ 199. Plummer Blocks or Bearings.— Length /=0-8 to l"2d^.

(See Figs. 326 and 327.) Distance of the bearings apart varies according
to diameter of the shaft ; in large merchant ships and fast steamers there
is a plummer block on each side of the shaft couplings. For the length
of the shaft see S 198.

Fig. 327.

Material used for the block — Cast iron with white-metal lining in mer-
chant vessels.
M „ Cast iron, brass, or cast steel, less fre-

quently gunmetal, with white-metal
lining, in warships.

As the plummer block does not have to take any great upward
thrust, the lower part only need be strongly built and lined with white
metal, the upper part being treated as a protecting cap, and kept quite
thin. On the top is a large oil and grease box, to lubricate the bear-
ings with oil and solid grease. In very large engines the plummer
block is sometimes made a hollow casting, and cooling water is cir-
culated through it.

§ 200. Bulkhead Stuffing Boxes.— At the places where the

342

MARINE ENGINES AND BOILERS.

shafting passes through the watertight bulkheads, a stuffing box made
in two halves has to be provided. It is advisable to cast it in two hah*es
in the form of a flanged plate, so that it can be taken apart, if the shaft

Fig. 328.

has to be lifted and taken out (Fig. 328). For smaller ships the plate on
the bulkhead is made of sheet metal, and the stuffing box fixed to it.

§ 201. Shaft Brake.— At a suitable place in the tunnel a strong
band brake is often placed to lock the shaft. This is used if the latter,
for any reason, has to be uncoupled from the engine. It might other-
wise be set in motion by the screw, as the ship proceeds.

Shaft Couplings.

§ 202. Detachable Shaft Couplings.— These are used to couple
together either the tail shaft and the aftermost tunnel shaft, or any two
lengths of the aftermost shafting. They are provided in case it should
be necessary to draw out the propeller shaft from the stern. As it
would be extremely difficult to carry the propeller shaft through the
ship when built, and fit it into the stem tube from within, one of the
following arrangements is usually fitted in large vessels.

1. In the coupling shown in Fig. 329, the flange, which fits the shaft
exactly, is drawn over the end of the shaft by' hydraulic pressure, and
a strong feather prevents it from turning.
To prevent the propeller shaft working
forward, should the flange get loose, the
diameter of the part of the shaft in the
flange is reduced to d^. To avoid the
danger of its working back, the shaft is
protected by a ring in two parts, which
is fitted into the groove at the forward
end, and held fast in position by
screwing up the flange against it. To
draw off the flange, when the propeller
shaft is taken out of the ship in dock,
hydraulic pressure may have to be used,
propeller shaft ; then

Diameter of the boss -

flange

))

»

Fig. 329.
Let d be the diameter of the

- d^ = l'od.

- D = 2to2'5d.

bore of the movable flange d^ = d-'-—rtod--^^

•30

60

Length of the movable flange
Thickness of collar -
Recess for the collar -
Thickness of flange -

- / =0-8 to ^.

- a =012 to 0-15^.
d^ = d—a,

. / = 0-25 to 0-3^.

2. The removable coupling or flange is often secured to the end of
the shaft by a cone and nut, and a strong feather or key prevents its
tuming. Material of the flange, forged or cast steel. Principal dimen-
sions, as above. Diameter of the screw, about 0*5 to 0*6^. Depth of
nut, about 0*4//; taper about 1 in 6 to 1 in 10.

344

MARINE EXCIXKS AND BOILERS.

S 203. Muff Coni^illgs.— The coupling described nnder § 3(rJ
has the disadvantage that it is very difBcult to take to pieces, and
if the shaft becomes worn by being often remored and replaced,
tt does not 6t tightly enough. A sooUed " muff coupling '' (Fig.
330) is therefore frequently used. It consists of two half sleeves
of wrought or cast steel, which grip the ends of the two shafts by means
of several strong bolts. Each end of tbe shaft is prevented from turn-
ing in tbe sleeve by a strong key. (For dimensions of the latter see
page 401). The two half sleeves are fitted with a tittle clearance be-
tween them, and it is only when the bolts are drawn up tight that they
grip the shaft.

I

Dimensions (see Fig. 330).
Diameter of the propeller shaft
Length of coupling -
Bore „ . -

I =3to3-etf.

Recess for the collars - - d^ = d-a.

External diameter of coupling - - D=r8to2^.

Width of collar left at end of shaft and of recess a =0'12 to 0'15rf.

Thickness of flange / = 05 to 0-6rf.

Total area of all bolts on both sides at the bottom of the thread —

A==0-35toO-5/(^,^?)

The distance c of the bolts from the centre of the shaft is as small
as possible ; they are frequently allowed to cut slightly into the sides of
the shafi. The strength of the bolts should be such that by tightening
them up to a given allowable stress, the pressure of the sleeves is suffi-
cient to grip by friction alone.

g 204. Disconnecting Couplings.— These are sometimes used
in warships having several engines, and they are placed in the shafting
immediately behind the engine, so that if the ship is steaming very

SHAFTING, RESISTANXE OF SHIPS, PROPELLERS. 345

slowly, one of the shafts can be allowed to turn freely without turning
the engine. In triple-screw steamers it is sometimes desirable to be able
to uncouple either the two outermost or the central shaft. These
couplings are made —

I. As claw or clutch couplings.
II. As flange couplings, with circular bolts which can be easily
inserted and withdrawn.

As a rule it is desirable, if possible, to avoid such couplings, as even
when most carefully made they form a weak point in the shafting, and
often begin to knock ; this is due to the twisting strain on the shaft,
when a certain engine speed is reached. Loose couplings have there-
fore been much less used of late years in warships.

§ 205. Tail or Propeller Shaft— In small ships this is almost
always connected to the tunnel shaft by a solid flange coupling, in
larger ships by a detachable coupling (see page 343), so that the tail
shaft may be fitted into the ship from the stern. For mode of attach-

Fig. 331.

ment of the propeller to its shaft see page 401. The length of the
propeller shaft is determined by the length of the stern tube, which again
depends on the build of the ship. It is advisable not to make the
propeller shafts too long, in order that they may the more easily be
put in, examined, and if necessary renewed. In ships wiih fine lines,
long tail shafts cannot always be avoided. If, for constructive reasons,
the shaft must project considerably beyond the body of the ship, a muff
coupling is often placed between the stern tube and the stern bracket.

Diameter of tfu Tail Shaft, — The dimensions of this are generally
greater than those of the crank shaft. Under certain conditions, as when
the propeller strikes against a solid object, this shaft is exposed to enor-
mous strains. In any case, it has to withstand great bending strains
due to the weight of the overhanging screw, and these are much in-
creased if one of the blades be lost, or the bearings of the stem tube
become worn. \{ d^ is the diameter of the crank shaft, then as a rule
^=1 to IWk.

Extract from the German Lloyd's Rules. — If d is the diameter of the
propeller, d must be = 0*6^^ -h 0-03d, or at least = 1 '02^/^.

346 MARINE EXGIXKS AND BOILERS.

Larger shafts are often made hollow ; proportions of the bore are
the same as in crank shafts. In all large, and in weD-built small,
cargo vessels, as well as in all warships, the propeller shaft is gene-
rally cased in brass where it runs in lignum vitae bearings. The brass
sleeve is shrunk on while hot, having been previously bored to a
diameter slightly less than that of the shaft. Thickness of the brass
sleeves for shafts above 10 inches in diameter —

d inches

6 = 0-60 inch +

OO

If the bearings are lined with white metal, the shaft where it runs
in them either has no sleeve, or has a nickel-steel sleeve. Between
the bearings, the shaft, in high-class work, is protected from corrosion
either by a thin brass sleeve shrunk on hot and sweated to the
bearing sleeves, or cast in one w^ith them, or by a rubber coating. The
part of the shaft in the water, between the stem tube and the pro-
peller, is also frequently coated with rubber in warships, and the
rubber protected from injury by a further binding, or by a sheet-metal
casing or by both.

Material used for the Propeller Shaft. — Wrought iron is now
seldom employed ; it is not suitable for this part of the shaft because it
rusts easily in sea-water and is not strong enough. Siemens- Martin
steel is used under all ordinary circumstances. Tensile strength, 28
to 32 tons per square inch; extension, 20 to 25 ""/^ on 8 inches.
Crucible steel is very good for large ships (fast and mail steamers), and
is frequently used for warships. Tensile strength is the same as for
Siemens-Martin steel. Nickel steel \% used for the same typ)es of ships
as crucible steel. Tensile strength up to 38 tons per square inch ;
extension, 25 "/^ on 8 inches.

Stem Tube.

§ 206. General Remarks. — The screw shaft, where it comes
through the skin of the ship, is enclosed in the stern tube, which forms
not only a bearing for the shaft, but a watertight joint round it. The
stern tube is a tube fixed at its forward end to one of the aftermost
bulkheads, and at its after end to the outer plating. If it is very
long, it is often made in two lengths, and supported in the centre. At
the inboard and outboard ends, and sometimes also in the middle,
long bushes are provided in which the shaft runs. The water is
allowed to reach these bushes, and serves to lubricate and keep them
cool ; in fact an artificial circulation is sometimes produced in the
tube, cooling water being either forced into the stem tube or sucked
from it by the engine. (See page 450.) Occasionally the stern tube

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 347

is^closed at its outer end by a stuffing box, and instead of water, oil is
forced into the tube, to lubricate the bearings and preserve the shaft.

At the forward or inboard end of the stem tube is a stuffing box,
which prevents the water in the tube from finding its way into the
tunneL This stuffing box, which is attached to the cross bulkhead at
the after end of the tunnel, must be accessible from the tunnel.

Length of Stern-tube Bearings —

Forward /^ = 3 to id
Aft /o = 4to5-5^

In ships where the screw shaft also runs in a stern bracket be-
yond the stern tube, the bearings of the latter are generally somewhat
shorter.

/j = about 2 to Zd,

A = „ 3 to 4^.
/g = „ 4-5 to 5-5^/.

In single-screw ships the propeller shaft is sometimes carried through
to a bearing in the stern post.

Fig. 332.

Length of Stern Tube, — This depends on the length of the bearings
given above, and on the distances between them ; the latter depend
largely on the form of the ship, and are determined by considerations
governing the accessibility of the stem-tube stuffing box at the end of
the tunnel. Ships with fine stern lines therefore usually have longer
stern tubes than ships with full stern lines. Total length of the stern
tube= 15 to 25^, according to the form of the ship.

Diameter of the Stem-tube Bearings, — If, as is the usual practice,
the shaft is fitted in from the stem, it is better to make the diameter
of the after bearing somewhat larger than that of the forward one.
The reverse holds if the shaft is to be put in from the forward end.

§ 207. Construction of Stern Tubes for Cargo Boats.—

The stem tube is almost invariably made of cast iron.

MAklXK ENGINES AND BOILERS.

If d= diameter of the propeller shaft.

Thickness c=^ + 0-8 inch.

(1 = 1-5 to l-8f.
i=l-2 tol-5c.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 349

Removable gunmetal bushes are fitted into the cast-iron tube, which
when lined form a bearing for the shaft. (See Fig. 332.) These bushes

-f

=^

■^^=

'^—

J

are lined with lignum vitas (see Fig. 337), and for the under side of the
bearing it is best to cut the strips across the grain. The lignum vitce is
held in the bushes by longitudinal gunmetal strips, but these should not

350

MARINE ENGINES AND BOILERS.

be fitted quite at the bottom, so that the shaft may not bear on them.
Lignum vitae is undoubtedly the best material for lining these bushes,
and is generally used ; it keeps in perfect preservation under water,
makes corrosion of the shaft impossible, wears slowly, and is easily
renewed. Renewal must, however, take place as soon as the play on
the after bearing gets too great ; the gunmetal bush can then be drawn
out, when the boat is in dock, by means of screws, for which purpose it
must be provided on the outside with tapped holes ; it is then relined and
driven in again by means of a collar and screws. To allow for their
expansion longitudinally when wet, the strips of lignum vitae must not fit
tightly at the ends of the bearing.

Thickness of the bronze bushes, ^ = f to 1 inch, for shafts from 8 up

to 24 inches in diameter.
Thickness of the lignum vitae strips, /=f to lyV inches, for shafts from

8 up to 24 inches in diameter.

-ff

3-

■■■/■■ ' ^/''i'iW^

Fig. 338.

In ships working in fresh water, the bearings for the shaft are made
in the form of simple cast-iron bushes fitted into the stern tube, and
having longitudinal grooves through which the water can circulate.
White metal is often used instead of lignum vitae strips to form the
bearing, and occasionally the bush is made entirely of white metal, and
drawn into the tube. The composition Of the alloy must be determined
by its durability in sea water. The thickness of the white-metal bushes,
if fitted, is from f to lyV inch.

At the forward or inboard end, the stern tube is bolted to the cross
bulkhead at the after end of the screw tunnel by a strong Range, which is
strengthened by a stout wrought-iron or steel ring ; the after end of the
stem tube must be a good tight fit into the stern post, if so fitted. To
ensure the hole in the stern post coming in the right place, the stern
post must be temporarily bored out, then lined off, and the hole bored
afterwards in correct line and position. A thin wrought-iron ring nut,
screwed on the end of the stem tube with a fine thread, prevents it from

SHAFTING, RESISTANCK OF SHIPS, PROPKLLERS. 351

352 MARINE ENGINES AND BOILERS.

being drawn out. Sometimes the stern tube is fitted in from the stern ;
the flange must then be outside against the stern post, and the nut
in front against the bulkhead. This method of construction has the
advantage of allowing the stern tube, when the ship is in dock, to be
drawn out with ease.

Stern-tube Stuffing Boxes (see Fig. 338). — Width of packing = J to
1 J inches, the larger sizes being employed for larger shafts. Depth of
packing space = 0*8 to 1*5^/, the lower values being employed for larger
shafts. In the larger sizes there should be some arrangement for screw-
ing the nuts on evenly. (See Fig. 338.) Cooling water from the engine
is sometimes admitted at the after end of the stuffing box.

§ 208. Method of Construction for Light Warships.—

The stern tube is generally made of brass, and strips of lignum vitae, or
less frequently of white metal, are fitted direct into it. To economise
weight, the centre piece is often made as light as possible, and not infre-
quently consists only of a light metal tube. In torpedo-boats a steel
tube is sometimes used for the stern tube, with bronze bushes fitted into
it. Brass stern tubes are usually made about half the thickness of cast-
iron stem tubes. (See page 348.) For lignum vitae linings, see page 349.
Fig. 339 shows the stern tube of an armoured cruiser constructed
wholly of gunmetal. The centre is composed of plates riveted together,
and screwed to the forward and after ends, which are cast in gunmetal,
and lined with lignum vitae.

§ 209. General Remarks on Shafts.— The position of the after

end of the shafting is determined by the immersion of the propeller in
the water, and in twin-screw vessels by this circumstance and the
distance of the screws apart. That of the forward end is determined
by the position of the engine. To allow for all these conditions the
shafting generally slopes downwards towards the stem, and in the case
of twin-screw steamers the shafts generally diverge from the centre
line as they go towards the stern. In torpedo-boats, where the centre
of the screw is deeply submerged in the water, this inclination of the
shafting is very marked.

Along the shafting, from the engine-room to the stern-tube stuffing
box, there is a passage called tke shaft tunnel^ separated from the adjoin-
ing watertight hold of the ship by a roof and watertight bulkheads.
It is entered from the engine-room through a watertight door, and has
a platform running down its entire length, to make it more accessible.
Along the roof of this tunnel, lifting gear is often arranged for the
intermediate shafts and their bearings : this gear is used when, in order
to examine the propeller shaft and the after stern tube bearing, the tail
shaft has to be drawn back into the ship.

SECTION II.

RESISTANCE OF SHIPS.

§ 210. Froude*S Method.— This is the most accurate method,
and is based upon experiments of towing models of ships in a tank, and
on what is known as the " Law of Comparison."

The resistance of a ship consists of —

1. Frictional or skin resistance.

2. Eddy resistance. 1 r. ..

3. Wave-making resistance. ) Residuary resistance.

Only the two last-named resistances follow the " Law of Compari-
son," which may be stated as follows : —

Laiu of Comparison, — Let l be the length, b the breadth, t the

draught, v the speed of the ship : further, let /= -, ^= -, /= -, and v be

n n n

the length, breadth, draught, and speed of the model of the ship. Then,

if v2 = ;if!2^ the resistances of the ship and of the model (w and «/) will

be to each other as their displacement s( d and d) —

WD

w d

-=- = «3*

Application of the " Law of Com-
parison.^''— Plot the results of towing
experiments on the model, as in-
dicated by the curve aaa, Fig. 340.
This gives the resistance of the
model at various speeds. Next,
calculate the frictional resistance of
the model, and deduct this from the
total resistance aaa; the curve bbb
will be produced, representing the

t

&p<«<t

Fig. 340.

residuary resistance of the model. If^ the scale be changed, in accord-
ance with the law of comparison, the curve bbb gives the residuary
resistance of the ship. The frictional resistance of the ship must now

* For deductions from this law see Taylor, "Resistance of Ships and Screw
Propulsion."

Z

354

MARINE ENGINES AND BOILERS.

be calculated, and added to the curve bbb, and the curve ccc is thus
obtained, which shows the resistance of the ship, if read to the correct
or ship scale (Fig. 340).

The calculation of the frictional resistance of the model and of the
ship is made as follows : — Let Wg be the frictional resistance in pounds,
F * the wetted surface of the ship or model in square feet, Va, the speed
of the same in feet per second, y density of water, / frictional (or skin)
resistance at unit speed, in pounds per unit of surface immersed in
fresh water, n an index or power, then

Wg =/x y X F X V

n
in

The values of/ and n can be inserted in the above equation for
determining the frictional resistance of the ship, by using the value
given below in Table No. 34. For calculating the friction of the
paraffin wax model in the water of the experimental tank, somewhat
higher values for n and lower for/ than those given in the table should
be used.

Table No. 34.

Constants for Frictional or Skin Resistance of Ships.
(Compare Johow, Hilfsbuchfiir den Schiffbau,)

Length of

the Ship

on the

Water-line.

Ship's Bottom of Iron,
well Painted.

1

Ship's Bottom of

Copper or Zinc

Sheathing.

Old Foul Ship's

Bottom of Copper or

Zinc Sheathing.

Feet.

/

n

/

n

/

n

16-4

32-8

65-6

98-4

131-2

164-0

196-8

230

262-5

295-3

3281

361-0

393-7

0-00026
0-00024
0-00023
0-00023
0-00023
000022
0-00022
0-00022
0-00022
0-00022
000021
0-00021
0-00021

1 -8507
1-8427
1 -8290
1-8290
1-8290
1-8290
1-8-290
1-8290
1-8290
1-8290
1 -8290
1-8290
1-8290

000024
0-00023
0-00023
0-00023
0-00022
0-00022
0-00022
000022
0-00022
000022
0-00022
0-00022
000022

1-9015
1-8525
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270
1-8270

0-00034
0-00031
0-00030
0-00029
000029
0-00028
000028
0-00028
000028
0-00027
000027
0-00027
0-00027

1-8660
1-8525
1-8430
1-8430
1 -8430
1-8430
1-8430
1-8430
1-8430
1-8430
1-8430
1-8430
1-8430

For calculation of P see page 356.

SHAFTING, RESISTANCE OF SHIPS. PROPELLERS. 355

The values of/ do not decrease appreciably —

In Column I. for ships of 550 feet and upwards,
n. » 425

„ III. „ 550 „ „

§ 211. Calculation of the Resistance of Ships, and Power
required for the Engines of Screw Steamers (from Midden-

dorf*).— Let

L denote length of the ship on the water-line in feet.

B „ greatest breadth of immersed midship section in feet.

T „ draught of the ship, excluding keel, in feet.

A „ area of immersed midship section in square feet.

F „ immersed surface of the ship in square feet.

/ „ area of propeller disc in square feet.

Vk 9, speed' of ship in knots.

Vn, „ speed of ship in feet per second,

w „ total resistance of the ship in pounds.

Wj „ frictional resistance of the ship in pounds.

Wj „ residuary resistance w^ = w - Wg.

B.H.P. „ effective or brake horse-power of the engine.

LH.p. „ indicated horse-power of the engine.

f „ a coefficient (see Table No. 35).

„ a coefficient (see Table No. 36).

€

rf ., efficiency (see Table No. 1). Then

1.

w,=

A X B X V *■*

Vb^ + ^l^

2.

W2 =

•00364 X F X v„i«

3. W = Wi+W2.

The effective horse-power of the engine is the product of the resist-

• Compare Middendorf, SchiffFwiderstand ufid Maschinenkistung \Jahrbuch
der Schiffbautechnischen Geselischa/t^ vol. i., 1900).

356

MARINE ENGINES AND BOILERS.

ance of the ship and speed of the screw. As v. is the actual speed of
the ship, Middendorf, to estimate the effective horse-power, has in-
creased this factor, so that

B.H.P. = w{ ^^ + TTTsr-^ * /s^ ) and i.H.p. = - RH.p. (sce Table I.)
^050 167-5 V 3^^/ w ^

Ik —

The values of ( and c are given in the following tables : —

Table No. 35.
Values of the Coefficient (.

B

C

L
B

C

L
B

1
1

1-41

below 8*5

2-Of)

9*3 and below 9*4

1-79

10-2 and below 10*3

8-5 and below 8-6

1-99'

9-4

9-5

1-75

10-3

10-4

1-38

8-6 „ 8-7

1-98

9-5

9-6

1-71

10-4

10-5

l-ST)

8-7

8-8

1-97

9-6

9-7

1-67'

10-5

10-6

1-32

8-8

8-9

1-95

9-7

9-8

102

10-6

10-7

1-29

8-9

90

1-92

9-8

9-9

1-58

10-7

10-8

1*27

9-0

9-1

1-89

9-9

10-0

1-54

10-8

10-9 1-25,

91

9-2

1-86

lOH)

101

1-50

10-9

ll-«

1-24

9-2

9-3

1-83'

1

101

10-2

1-45

1

11-0 and over

1-23

The value of rj (see Table I.) should always be taken as representing
the horse-power of the engine corresponding to the greatest speed.
If the speed for a given horse-power of engine has to be calculated
before the lines of the ship are got out, r may be approximately deter-
mined from the following equation —

Where u = wetted perimeter of midship section in feet.
L = length of the ship in feet,
f = a coefficient (see Table No. 37).

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS.

Table No. 36.
Values of the Coefficient c.

357

L
V 2

c

L
V 2

€

L
V 2

€

L
V 2

€

-0*10

•255

0-28

•214

0-52

•168

0-88

•122

oil

•252

0^29

•211

054

•165

0^90

•ISO

012

•250

0-30.

•209

0-56

•162

0^92

•118

013

•247

0^31

•207

0-58

•158

0-94

•117

014

•245

0-32

■205

^60.1^- •

•155

0-96

•115

015

•242

033

•203

0-62

152

0-98

•114

016

•240

0-34

•201

0-64

•150

. 1-00

•113

017

•237

0-35

•198

0-66

•147

1^02

•112

0-18V

•235

0-36

•196

0^68

•144

ro4

•111

019 -

•232

0-37

•194

0^70

•142

1-06

•110

0-20.

•230

0-38

•193

0^72

•139

1-08

•109

0-21

•228

0-39

191

0-74

•136

MO

•108

0-22

•226

0-40 fvL -

•189

0-76

•134

1-12

•108

0-23

•224

042

185

0-78

131

114

•107

0-24

•222

0-44

182

-0-80, ^^i '

129

116

•107

0-25

•220

0-46

178

0-82'

•127

J-18

•106

0-26

•218

0-48

174

0-84

•126

.1-20

•105

0-27

•216

0^50

171

0-86

124

and over

Table No. 37.
Values of^.

V Speed in Knots.

f

V Speed in Knots.

^

8 and under 9

0^90

19 and under 20

0^79

9 ,

1 10

0-89

20 ,

, 21

0-78

10 ,

, 11

0-88

21 ,

, 22

0-77

n ,

, 12

0-87

22 ,

, 23

0^76

12 ,

, 13

0^86

23 ,

, 24

0-75

13 ,

, 14

0^85

24 ,

, 25

0-74

14 ,

, 15

0-84

25 ,

, 26

0-73

15 ,

, 16

0-83

26 ,

a

, 27

0-72

16 ,

, 17

0-82

27 ,

, 28

0^71

17 ,

, 18

0^81

28 ,

, 29

0-70

18 ,

, 19

0-80

29 ,

1

, 30

0-69

358

MARINE ENGINES AND BOILERS.

Tables

Dimensions^ Resistance, and

(Compiled from the Complete Tables of Middendorf, Jahrbuck

•1

1
Type of Ship.

1

Fast Steamers.

Slower Running Large Cargo
and Passenger Stomers.

1

1
Name of Ship. i

Furst
Bismarck.

Kaiserin

Maria

Theresa.

Kaiser

Wilhelm

d. Grosse.

Lahn.

Aachen,
Halle, &c.

Barbairofisa,

&C.

Pensyl-

Prrtnria.

1

Length L in feet

502o

527-45

625

450.1

3551

533

561

Breadth B in feet

1

57-5

51-83

66

48-82

43 5

60

62

Area of midship section A,

sq. ft. - - - - 1

1226-6

1143-7

16^

nil

843 5

1477

I5CW

Immersed surface F, sq. ft.

3o,830

39,490

54,876

30,235

24,940

46,440

49,722 :

Area of propeller disc /,
sq. ft. -

564*68

5300

7141

388-4

228 1

460

427-7

Number of propellers

2

2

2

1

1

2

2

Speed of ship Vk in knots -

20-7

20-5

22

18

12

14-4

14

Speed of ship Vm in feet per
second ... -

35

34*62

3715

■
30-33

20-26

34-31

23-67

r f - -

1-97

1-45

1-75

1-83

2-0

2-0

1-89

Coefficients - c -

I i - -

•105
110

•105
110

•105
110

•105
110

•105
1-21

•105
114

-ia5

1-15

/Wj in lbs.

75,665

68,965

118,*240

47,529

14,129

36,626

1

Resistance;,,, • ii_
of ship < Wain lbs, -

93,504

101,205

160,270

60,920

22,784

61,983

62,930 1

\ w = w, + Wg lbs.

169,169

170,170

278,510

108,449

36,913

98,609

97,25:2

Indicated horse - power =
l.H.P, calculated -

15,189

15,287

27,005

8,672

2,238

6,688

6,574

Indicated horse - power =
i.H.P. on trial

1
1^

15,725

■0

17,260

26,630

8,465

1,775

6,871

■J -^

5,232

at about

13-2

knotsJ

^ •■

- 1 ■>

1 L

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 359

Nos. 38 and 39.

Horse-power of various Ships.

der Schiffbautechnischen Gesellschaft^ vol. i., 1900.)

Large Warships.

Torpedo'boats.

1

Steam
Launches

and
Pinnaces.

Steam
Trawlers.

1

Prinz
Hrnrich,

Prinz

Regent

Luitpold.

Imp.

Yacht
Hohen-
zollem.

Minne-
apolis.

Powerful and !
Terrible.

1

Gushing.

1

138

Rodgers.

Turbinia.

—

Dora.

4.>0"0

1
382-5

412

ms

160

108-24

52-5

105

51

46

58

71

14-24

1

16

9

10-23

21

11U8

736

1124-4

1646-2

1 47-7

62

23-2

23-45

105-4

33,786

21,272

27,976

40,403

1804-4

2446-8

968-4

592

2651-2

357-6

1
342-25

518-2

596-7

13-88

102-2

16

8-44

48-67

2

2

3

2

1

2

9

1

1

13*6

21-5

231

24 0 21-9

22-5

24-9

32-5

12-5

10-5

22-97

36-31

39-01

40-52 36-98

38 0

4205

1
54-89

21 10

17-73

1-92

20

20

2-0 ' 20

1-67

l-.>4

1-23

20

2

•105

•105

-105

•113 105

-207

•21

•252

•192

•105

1-22

1-12

110

110

110

1-29

1-35

1-2

1-65

1-61

23,701

51,957

111,222

195,173 j 143,740

7,(X)4

12,005

9,768

1

1,2-27

2,100

40,529

59,538

89,420

138,613

117,006

! 5,492

9,029

5,814

608

1,970

64,230

111,495

200,642

333,786

260,746

1 12,496

21,034

15,582

1,835

4,070

4,368

10,595

20,184

36,115

25,542

1

1,618

2,597

2,426

162-7

278

4,-280

9,502

20,088

2,5,930

1,730

2,379

2,071
at about 1
31-7 !
knots.

177

295

«

" 1

1

360

MARINE ENGINES AND BOILERS.

§ 212. Approximate Method for Determining the Horse-
power of an Engine. — To calculate the work of the engine, ap-
proximately and rapidly, for a ship of given deplacement d and
speed V, the following formula may be used —

I.H.P. =

v^x dI

c is here a constant, namely —

For large and fast ships and steamers c = O'0O275.
For large cargo steamers with full lines c -^ 0-00)\90.
For medium-sized warships with fine lines c = 0*004 to '005.
For small ships with fine lines c = 0*005 t(X:006.

SECTION III.
THE SCREW PROPELLER,

§ 213. Introduction. — With our present experience it is not yet
p>ossible to draw up simple formulae which can be universally applied to
determine the leading proportions of screw propellers, or from which,
if given values are inserted in them, the required proportions can be
deduced without difficulty. It must be stated at the outset that no
method of calculation can be accepted, unless the results are compared
with similar data from actual practice, or checked at least by a second
method. The following are the symbols most frequently used : —

I.H.P., indicated horse-power.

B.H.P., brake or effective horse-power.

S.H.P., effective or useful horse-power delivered at the screw propeller.

«, revolutions per minute.

V, speed of the ship in knots.

c, theoretical speed of the screw in knots = ^^^ " ^, ,

6,086-44

u, „ „ " wake " in knots.

St, real slip in knots.

s^ apparent slip in knots.

w, total resistance of the ship.

D, diameter of the screw in feet.

H, pitch „ „

d^y diameter of the boss in feet.

d^ diameter of any given element of the blade in feet, and some-
times also

dy diameter of propeller shaft in feet.

z, number of blades.

A, developed area of one of the blades in square feet.

/, actual breadth of a blade at any given point in feet (a function of
the developed area).

/„, mean width of blade = ^ - g .

D — dn "^ *cet.

*^62 MARINE ENGINES AND BOILERS.

by ratio of mean width of blade to diameter of screw —

D D(D-^„)

2

Z . A

ky ratio of total developed area of screw to the disc area =

\^

hy thickness of blade (i>., greatest thickness at any given cross sec-
tion of blade).
h^ thickness of blade at tip.
^n» n n root.

K^ and Kg, coefficients of the screw.
k^ „ k,yy „ for calculating the stresses in the blades.

S 214. General Remarks. — Every screw consists of a boss and
blades. The after surface of the blade, by which the water is forced
astern as the ship is driven ahead, is called its "/tfr<?." It almost
always forms a part of the helical surface of an ordinary screw. (Compare
page 396.) This surface is produced by a straight line rotating at
uniform speed round an axis which it intersects, while the point of
intersection also moves along the axis at uniform speed, the angle
between this line and the axis remaining constant. That part of the
surface of the screw which is required for the blade is bounded by
the shape or contour of the blade. The surface of each blade belongs
to its own particular |)art of the screw, and the surfaces of all the
blades are parallel to each other. A two-bladed screw is thus two-
threaded, a three-bladed screw three-threaded, and so forth.

A distinction is made between right-handed and left-handed screws.
To an observer looking at a screw propeller from aft, a right-handed
screw, when working ahead, would appear to revolve in the same
direction as the hands of a clock. This motion is said to be " clock-
wise," and the opposite motion to be " counter-clockwise."

The material required to give the necessary strength to the blade is
placed on the back of the driving surface (/>., the surface next the hull
of the ship). (Compare page 396.) The forward surface of the blade
does not conform truly in practice to the surface of a true screw. This
is more marked nearer the boss, where the blade is thickest.

The edge of the blade which, as the ship moves forward, first cuts
the water, is called the " leading edge^^ and the opposite edge is called
the ^^fallaiving edge^

The diameter of the screw D is the diameter of the circle described
by the tips of the blades.

The pitch of the screio h is the distance through which each point
on the surface of the blade travels in the direction of the shaft during

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 363

one revolution, if the screw were considered as rotating in a fixed solid
body.

The ratio — is the ratio of the diameter of the circle described by

H

any point on the surface of the blade about the axis, to the pitch.
This varies, therefore, for the points of the blade situated at varying
distances from the centre, while all the points which lie along a line ab
(Fig. 341) of the developed area corresponding to a fixed distance from

the axis have the same ratio — . The maximum ratio

H

contrary a definite value for each blade.

The width of the blade I (Fig.
341) means the width at any given
point on the developed blade, /.<r.,
the actual developed length of the
face on the cylindrical section con-
centric to the shaft. (For length
of the cross-sectional template see
page 396.)

The mean width /„» is the
quotient —

• _ developed area_ a
length ot blade d - ^n

Mean tvidth ratio h is the
quotient — %

, _ mean width _ /„
diameter d

H

has on the

Fig. 341.

By the term projected area of
a blade is generally understood
the area which is obtained when

the blade is projected on a plane at right angles to the axis of the shaft.
(Compare Figs. 342 and 346.)

The developed area a of the blade (that is the area of its face)
can only be approximately determined, because the area of the screw
cannot be accurately developed.

To Develop the Area of a Screw: First Method, — Given the projected
area, the face of the blade is divided up by any number of concentric
circles, the centre of which is the centre of the shaft. These circles
cut the projected area (Fig 342) in curves which are arcs of circles,
e,g.^ DAE, and the actual face of the blade in helical lines. The
inclination of the latter at a, i.^., the angle which the tangents to

364

MARINE ENGINES AND BOILERS.

the helical lines at a form with the plane vertical to the centre line of
the shaft (the centre of the circles), is obtained from Fig. 344. Ifr
is the radius of any circle cutting the area of the face, h the pitch
of the screw which is constant, the angle of inclination of the helical
lines with the plane at right angles to the axis is, at every point

(and therefore at a), tan o= - — . The tangent to the helical lines is

2/Tr

thus inclined to the centre of the shaft as the line bjojc^ is to the line

'O^O^O

Bo OoCo .

Fig. 343.

CltieaJL J

Hzn

Fij,'. 344.

Now let BoBo' = length of the arc ad, and CoCo' = length of the arc At.
Then the true length of the part of the helical line ad = OoB„ and the
true length of the part of the helical line ae = OoCo.

If we make ab at right angles to ao = OoB^, and ac at right angles
to AG = OoCo, then the points b and c will be the points of the
developed area corresponding to points d and e of the projected area.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 365

In the same way, as many points on the developed area as are de-
sired may be obtained by determining the true length of the lines
of intersection of other circles, concentric to dae. The innermost
of these concentric circles passes through point f, where the centre
line of the blade touches the boss. Thus the line at right angles to
AG passing through f forms the boundary of the developed area at
the boss.

To obtain a side view of the blade (Fig. 343) the point a^ is first
determined upon the line OiFjAj (which is here drawn sloping back-
wards), the distance b — o^^ is plotted out towards the face, and
r=OoCo' towards the back of the blade, measured from a^. The points
where the horizontal and projected lines through d and e intersect the
vertical lines drawn at distances b and c from a^ give the points d^ and
Ej of the side view of the blade. From the sectional elevation at the
axis and the side elevation of the blade, it is easy to obtain the plan
(looking down on the tip of blade). The line Bc = BqCo is the length of
the so-called "cross-sectional template." (Compare Figs. 361, 363, 369,
and § 231.) In the side elevation (Fig. 343) the line D|A^Ei is thus
projected as a sine curve.

If the developed area be given, and it is desired to construct the
projected area from it, horizontal lines are drawn through the developed
area, the angle of inclination a of the helix through a is determined
with the help of Fig. 344, BoOo = ab, CoOo = ac are obtained, and BqBo'
and CoCo' drawn at right angles to OoBo' and OqCo' respectively. If
the lines BoBo' and CoCo' are measured from the point a along the arc
of a circle having the radius r=OA described about o, they will give
the required points d and e of the projected area.

Second Method (see Figs. 345 and 346). — Given the projected area,
to find the developed area.

I^t D and e be, as before, points on the projected area lying
along the arc dae of a circle with radius r. We will here assume that
the area of the screw in the vicinity of point a is represented by the
surface of its tangential plane at this point, which closely corresponds to
it. The angle of inclination of this tangential plane to the plane of
the drawing is equal to the angle of inclination a of the helix of the
screw * passing through point a, and can be found from a right-angled

triangle with its side and base equal to h and 2irr or to ;r- and r
respectively (see Fig. 346). The intersection of the tangential plane

* Strictly speaking this only applies to blades whose planes arc vertical to
the shaft.

366

MARINE ENGINES AND BOILERS.

with the circle through a is thus an ellipse, which is projected as a
circle dae. Half the minor axis of this ellipse is equal to the radius
r, half the major axis is inclined at the angle a to the plane of the
drawing, and when projected is equal to the length r. The major

Fig. 345.

,-r H:ZJC ^^ ff:2Jt ^

Fig. 346.

axis is thus the hypotenuse of a right^ngled triangle, the sides of
which are

r=0oi, and ^

The focal points m and n of an ellipse may, as we know, be obtained,
if from the end a of the minor axis a circle be^described with half the

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 367

major axis as radius. This has been done in Fig. 345. Taking half the

major axis OqF = am as radius, a circle is drawn from a, which cuts the

line at right angles to ao at m and n, the focal points of the ellipse,

the half axes

OoF = am = an; andr=OA.

As the triangle oam = triangle iOoF, then om = if = — . The focal point

H

of the ellipse is thus at the distance — from o, and is independent of r.

It is therefore the same for all ellipses produced by the intersection of
all circles concentric to dae, with their corresponding tangential planes,
and the helical lines of the screw, at the points where the circles intersect
the area of the blade. If the tangential plane be turned round into
the plane of the drawing, the points in the developed area, corre-
sponding to D and E, namely b and c, are obtained. They are found
either by cutting the ellipse passing through any point a by a straight
line projected through d and e, which is easily done by means of the
focal points m and n, or by the following equation —

OoDo' = LD LB = OoDo

and OoE' = le lc = o„e.

'o*-*© ■"*^ '-'^ ''o^o

If the developed area be given, and the projected area be required,
the ellipse bag is drawn for instance through any point b (with the
aid of points m and n), or point a is determined by calculation as
follows: — ^The sum of mb + bn is determined, and ma is made = J
(mb + bn). Through oa draw a circle having the radius 0A = r, and
find the point where it is intersected by the straight line blc at right
angles to oa. The points of intersection d and e are the required
points of the projected area.

§ 215. Number of Blades. — Thvo blades are now scarcely ever
used, as they strain the shaft, and are particularly prone to give rise to
vibrations in the stem of the ship, especially if either of the blades gets
damaged. Three blades are the almost invariable rule in small and
medium sized warships and generally for all small fast vessels. Four
blades are always employed in ordinary cargo boats, steam tugs, ice-
breakers, &c. All other ships have three ox four blades; the larger
number is now usually preferred in large fast steamers and passenger
vessels. In ships with three screws, the middle screw is sometimes four-
bladed and the two outer screws three-bladed.

§ 216. Different Forms of Blade. — Blades are very often

made which are set over towards the stern (see Fig. 355). This shape

368 MARINE ENGINES AND BOILERS.

is usually selected, in order that the tips of the blades may (with a single
screw) lie well away from the stem-post. These blades are designed to
keep the stream of water together, as it is forced astern. For this reason
the blades of screws which run at very high speeds are frequently set
backwards at a considerable angle, and often have to take a severe
stress, as explained in § 226. Blades with variable pitch are seldom
used, and have as a rule no advantage over those with uniform pitch.
A distinction must be made between blades the pitch of which varies
axially, and those in which the pitch varies radially. Torpedo-boats
are fitted with screws having an axially, and at the same time a radially
varying pitch.

The Hirsch, Mangin, and other similar types of propeller are now
seldom fitted.

§ 217. Speed of the Screw, Stream-line Wake, and

Slip. — If the water were a solid body, the screw would move forward
the length of the pitch h at each revolution. The speed of the ship

would then be c= feet per second = knots, say =

^9 ^^ XX ^^ nil

6 080~ '^^® speed c is often called the "speed of the screw."

In reality the screw does not, during one revolution, move forward
through the distance of the pitch h, but through a smaller distance.
The real speed of the ship is therefore less than c. The ratio

_ speed of the screw - speed of the ship ^ ^ __ c- v
speed of the screw * ~' c

affords a means of estimating the retardation of the screw in the water.
The value s^ is called the apparent slip. The water in the rear of the
ship in which the screw works is not at rest, but follows to some
extent the forward motion of the ship. The speed with which the
water streams after the body of the ship is called the speed of the
stream-line wa^e = \j. This is greatest in the part of the water
immediately under the stern-post, and round the wetted skin adjoin-
ing, and is also greater near the surface of the water than deeper
down.*

The value of u, as used in calculations, is the mean speed of
the stream-line wake in the region where the propeller works. The
" fuller " the stern of a ship, the larger will be the stream-line wake.
The effect of the wake upon the screw is the greater, when the latter
works in the strongest portion of the wake current. It is therefore

* Compare Calvert *' On the Measurements of Wake Currents" {/ns/, of Naval
Architecfs, 1893).

SHAFTING, RESISTANCK OF SHIPS, PROPELLERS. 369

more marked in single-screw than in twin-screw steamers. In default
of other basis of calculation, the diagram given in Fig. 347 may be
used to determine u.*

By yii/ness in this diagram is meant displacement -r (length x breadth
X depth) of the immersed portion of the ship. It is best to calculate
u from the results of actual trial trips, and to use the data thus
obtained as the basis of new designs, t (Compare page 381.)

A current or wake of a certain velocity is always present. The
screw therefore has to impart to the ship only the difference between
the velocity of the ship v and the velocity of the wake u. The actual
speed of the ship, relatively to the water at the stern of the vessel, is
thus V - u. As this value is always less than v, the screw must at

t

49S 4S0 ^ ii^ ^ ip^ f/s

Fullness

Fig. 347.

every revolution be retarded by a still greater amount than is shown

p y

in the formula The real slip is therefore —

c

c-0[j-u)

in contradistinction to the apparent slip, which is always less, and is —

c-v u

c c

In ships with full stern lines the stream-line wake velocity u is large.

* Compare Taylor, ** Resistance of Ships and Screw Propulsion."
+ The formula given by Riehn {Z^itschrift des Vereines DetUscher Inghiieure^
18d4, p. 469) cannot be used for the direct calculation of the '*wake" as here
defined, because he there attaches a somewhat different meaning to the term " wake "
( Vorstrom),

2a

370 MARINE ENGINES AND BOILERS.

If at the same time the real slip, as the result of a badly constructed
screw (compare page 373), is very small, it may happen that there is no
apparent slip, or that its value becomes negative. Real slip of course
must always exist ; it can never become zero or a negative quantity.

Example, — Suppose the speed of a ship, during its trial trip, to have
been 10 knots, and the speed of the screw, calculated, from the pitch
and the number of revolutions, to be

The apparent slip is thus negative, and works out at

c-v 9-5-10 p,«oi

K — = — TT^ — = -5*3 /.

c 9-5 '"

The customary explanation is as follows : — The lines of the ship's stem
are very full, there being but one screw, which is placed close to the
stern-post. The effect of the wake current is therefore strongly marked.
If the real slip be taken at only 10 "/^ (compare page 373), the wake
current of the ship is —

u = c(jt - s^ = 9-6(01 + 0-053) = 1-45 knots.

Negative slip is always a sign that the propulsion of the ship is
inefficient, for it shows firstly that the real slip is small, which, according
to § 218 (4), is unfavourable ; secondly, such slip only occurs when the

wake current is large. When such is the
case, more energy is absorbed by the wake
than is recovered from it.*

Angle of Slip (see Fig. 348). — During ^w
revolution, corresponding to the path Dr
of the tip of the blade, the screw moves
forward in an axial direction over a distance
H, if there is no slip ; if there is slip, it moves
only through the distance h - ab. The angle
AOB = ^ is called the slip angle.
Fig. 348.

§ 218. Propeller EiBciency.— If ih.p.

is the indicated horsepower of an engine, then b.h.p. =77 . i.h.p. (see
page 4) is the useful power expended in rotating the screw. A part
only of this power is converted by the screw into axial or forward thrust,

* This customary explanation of ** negative slip" must be used with caution, as
it would be very difficult to demonstrate numerically that the wake can cause the slip
to be less than 0.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 371

the remainder is lost in overcoming frictional and other blade resist-
ances. The efficiency of the scrtiv is therefore —

effective axial thrust of screw s.h.p.

^7s =

B.H.P.

brake horse-power

For calculating the efficiency* from the equations given for s.h.p.
and B.H,p., see § 222.

In well-designed screws the efficiency varies from 40 to 70 "/^ ; the
lower values (40 to 50 ""l^ are for small screws in small ships, with
engines running at a high speed ; the higher values (55 to 70 "*/ J being
for ships with large screws and running at moderately high speeds, t

The efficiency depends —

(1.) Upon the ratio between the pitch and the diameter, />., upon
the inclination of the elements of the blade to the plane perpendicular

(i6
0,i

I
I — I 1 1 I I I I ' I I

r

Y,0»

O q2 ii&^ ^S 0,8

Ratio,
Fig. 349.

to the shaft. Fig. 349 (from Taylor) shows the efficiency of elements of

the blade having varying ratios - •

n

d .

The efficiency of those elements in which - is less than about 0*15

H

is here seen to be very low, therefore the extreme ratio — must be such

H

that, for as large a number as possible of the blade elements, the

* For calculating the efficiency, see Riehn, Zeit. des Ver, Deulschtr Ing,y 1884.

t A high speed within certain limits docs not of itself involve a bad efficiency ;
bat in those t}rpes of ships in which the engines ran at high speeds, the conditions
are unfavoonible to the efficient working of the screw.

372 MARINE ENGINES AND BOILERS.

ratio - conduces to a high efficiency, and lies between the limits
- = 0-2 to 0-9. In accordance with this rule, the extreme ratio of -

H H

is generally found to be from 0*5 to 1*2.

Instances of how largely — varies in practice, and between what

n

wide limits it may lie, are shown in the following table, based upon
results obtained from a large number of actual screws. Thus in

Torpedo-boats having one screw - = |

0-8 to M.

Destroyers

two screws

• • '

}

Small cruisers

two

}}

=

0-7 „ 1-0.

T.arge cruisers and ironclads

two

»

i> ^^

0-7 „ 11.

Fast steamers

two

»>

=:

0-6 „ 0-8.

Freight and passenger steamers

two

n

=5

0-7 „ 0-95.

Cargo boats

one screw

» =

0-7 „ 1-0.

In choosing the ratio - the only thing to be considered is that the

n

value - is not too low for the greater part of the blade.

H

(2.) From the above remarks it follows that the efficiency also
depends upon the shape of the developed area of the blade. If the
greater part of this area lay close to the root of the blade, it would

contain many elements in which the ratio - would be unfavourable to the

H

best efficiency. Therefore the blade must be small at the root. For
constructive reasons we cannot go too far in this direction, neither must
it be too broad at the tip, otherwise it might be easily broken. If, on
the other hand, it is made too thick, the resistance of the leading edge
prejudicially affects the efficiency. (Compare the formula for ly^ § 222.)
It is thus self-evident that the best blades are of the form most generally
met with in practice, the developed area of which varies little from that of
the normal or standard blade (Fig. 3-51). Small deviations from the
normal shape of the developed area * only slightly affect the efficiency.

developed area zk

(3.) The ratio r ^ o*" '^ = i~ affects the efficienc}^

-.D2

* By *' shape of the developed area," the shape of the boundary' lines of the
developed area is always to be understood.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 373

If ^ is very large, it naturally follows that more of the area lies
close to the boss than if, with the same total area of the blade, k were
smaller ; thus, if the value

of k be high, the efficiency
is low. If k be large, the
blades must lie close to-
gether, and disturb each
other by the eddies they
create (compare the smaller
values of a with many- ?>

bladed screws in the for- -sj '

.s»

98

9^

mute § 222).

For the foregoing
reasons the diameter of
the screw is made as large
as possible ; the difficulty of
doing this with the bulging
full stems of torpedo-boats
should be noted. Only
where it is impossible to
increase the diameter,*
and where the use of two

fij

— — -^^^ — ^.^-^ —.^^^^ ^.^i^_ __^_^^

^ il(0S f</ ^S ^ f^s ip§ f JX

Slip,

Fig. 350.

>^ = 0-30 to 0-33

»

19

»

»

»

>^ = 0-33
>^ = 0-32
>& = 0-33
>^ = 0-30
>^ = 0-40

19

)l

)}

))

»>

0-39
0-38
0-39
0-45
0-80

screws is not desirable, do we find high values oik (or of ^, see page 380).
The usual values of k are —
For torpedo-boats and small fast steamers
small warships
ironclads

fast steamers, mail steamers
cargo steamers
flat-bottomed river steamers and tugs

Warships which are to be navigated in shallow water have, under
certain circumstances, higher values of k than those given above.

(4.) The real slip s^ also affects the efficiency, as may be deduced
from the formula for efficiency ly, (see § 222). The most favourable
real slip is about 15 to 20 %.

The connection between the efficiency and the slip for a blade-
element having the ratio - =0*6 is shown in Fig. 350. (Compare

Taylor, see note, page 369).

(5.) In ships with very small screws and running at very high

* As is the case in flat-bottomed river steamers and ferry boats. In these the
blades are very wide, and their edges overlap and almost cover each other.

• »/4

MARINE ENGINES AND BOILERS.

speeds (such as those driven by steam turbines), the efficiency may be
greatly affected by what is known as " cavitation." * This term denotes
the formation of hollow spaces or cavities on the forward side of the
blades, due to the cutting by the propeller of the column of water set
in motion.

§ 219. Other Formulx for calculating the dimensions of ship's
screws are generally based on the equation —

I.H.P. =f . d2(«. h)'.

The constant c, which varies for different types of ships, depends on
the shape of the screw and of the ship (wake). The following formulae
are those generally used t —

/ I.H.P. , /l.H.P.

Vioo/

Here d and h must be stated in feet, a in square feet.

The constants k^ and Ko are obtained from the following table :—

Table No. 40.
Coefficients K^ and Vi^^for Calculating Screivs.
(From Seaton and Rounthwaite's ** Pocket- Book of Marine Engineering.'*)

Type of Ship.

Speed
ofShip

V in
Knots.

Number of

Ki.

Material of

Screws.

Blades.

Blades.

1. Cargo steamer with

full lines •

2. Cargo steamer with

moderately full
lines -

3. Mail and pa.ssenger

steamers, fine lines

4. Do. do.

5. Do., very fine lines

6. Do. do.

7. Warships with very

fine lines -

8. Do. do. •

9. Tori)edo-l)oats, do.

8 to 10

10 „ 13

13„17

13„17
17„22
17 .,22

16„22
16„22
20„2e

1

1

1

2

1
2

2
2

1

4

4

4

4
4
3

4
3
3

17 to 17-5

IS „ 19
19-5 „ 20-5

20-5 ,,21-5

21 „22

22 „23

21 „ 22-5

22 „ 23-5
1 25

1

19 to 17-5

17 „ 15-5

15 „ 13

14-5 „ 12-5
12-5 „ 11
10-5,, 9

11-5,, 10-5

8-5 „ 7
7 „ 6

Cast iron.

\ Cast iron,

bronze, or
j cast steel

Gunmetal

f ^'
bronze.

Bronze or
forged steel.

* For further details see Barnaby, ** Marine Propellers.''

t For similar formulx see Fliege, Zeitschrift des Vereines Deutscher Inghtieure.
1893, p. 1552.

SirAFTIN(;, RESISTANXE OF SHIPS, PROPKLLERS. :\iii

The values for fast steamers with two screws of four blades each
come between lines 6 and 7 of Table No. 40.

Example, — Twin-screw fast steamer " Kaiser Wilhelm der Grosse."
Let us assume the i.h.p. (calculated from the ship^s resistance) =

2 X 14,000, and further ? = 0-636 (see § 218), and « = 78. Assuming

that each screw has four blades, then ? = 4. The diameter, pitch,
and blade area are to be determined. From the above formula
for D —

2_ ir 2

I.H.P.

D' = K

Viooy

3

Taking the value for - as 0*636 —

H

(0-636)^h2 = k,2^"-^-^3

whence 5 ^ k,^ . i.h.p. . 100^

(0-636)2 . n^

Solving the last equation, and taking the value of Kj = 22*75 (mean
value from Table No. 40), h = about 32*8 feet.

In practice h = 33*5 feet (10*2 metres).

Further d = 0*636h = 20*8 feet (6*36 metres).

In practice 0 = 21*32 feet (6*5 metres).

The developed area is obtained from the equation zK^Yi^ /^'^'^\.
According to Table No. 40, the mean of K2= 10*5, w^hence

2rA= 10*5^^ ~'^>j ~ ~ ^^^ square feet (13*4 square metres).

In practice the total developed area of the blades is —

ZK— 136*4 square feet (12*68 square metres).

Calculation of the Area of the Screw from the pressure on the screw.
The developed area of the blade may also be determined by taking as
the basis of the calculation a pressure per square foot of blade area
deduced from actual experiments. A large number of tests have yielded
the following values for the ratio —

indicated thrust

^3-=

developed area of the screw

9i

376 MARINE ENGINES AND BOILERS.

In torpedo-boats it is = about 13 lb. per square inch.

„ fast steamers „ = „ 10

„ cargo and passenger steamers „ = „ 7 to 8*5 „
„ cargo boats „ = „ 5-5 to 7 „

§ 220. Remarks. — It is always better to make the pitch too small
than too large. If the value h be smaller than the corresponding values
given in the formula in § 219, those ratios which prejudicially affect the
efficiency are avoided. On the other hand, by assuming a relatively smaller
pitch, the only difficulty is, that the number of revolutions is increased
beyond that assumed in the calculations, but the engines will then be
better able to utilise to the full the power of the boilers. If h be larger
than its calculated value, it may happen that the cylinders are too small
to produce the turning moment necessary to give the required speed.
Hence the engine will not be able to run at the required number of
revolutions, and cannot impart the desired speed to the ship, although
the boiler may easily be able to supply sufficient power for this speed.

§ 221. Taylor's Method for Calculating a Ship's Screw.—

Although the formulae given in § 219 are partly empirical, those given
below for calculating the screw, taken from Taylor's book, are strictly
theoretical. They confirm the empirical formulae for i.h.p., from which
they differ only because, instead of the constant Kj, expressions occur
based on the diameter ratio, slip, coefficient of friction, &c. In Taylor's
formulae the use of indeterminate coefficients (such as those in which
the physical significance is not clear) is unnecessary, as every coefficient
can be deduced for each separate case from a physical basis. Taylor's
method of calculation necessitates the use of experimental values for
the stream-line wake (compare page 368) ; and as only constants having
a definite meaning are used, this method, if followed, shows where the
weak points of the calculation lie, and thus any great errors may be
avoided.

§222. Taylor's Theoretical Formulae.— Taylor * deduces the
following formulae theoretically : —

B.H. p. = effective work done in turning the screw

s.H.p. = effective or useful work due to the thrust of the
screw (along the ship^s course)

= 3x5: ( "q^^q) X D.^{aj,(l - s,)x, -/[! - s,)y,}.

* The following deductions are taken from the excellent work of D. W. Taylor,
** Resistance of Ships and Screw Propulsion."

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 377

In both formulas d and h must be stated in feet, b.h.p. and s.h.p. in
horse-power, s^ as a decimal fraction ; whilst s, «, <?,/, x^, y^, and z^ are
numerical values. The efficiency of the screw t]^ is obtained by dividing
one formula by the other —

If Tf is the efficiency of the engine (see page 4) then —

B.H.P. = »; X I.H.P., therefore s.h.p. = ^ x rj^ x i.h.p.

In the above formulae, the following symbols are used in addition to
those given on page 361 : —

a = the so-called " coefficient of thrust " (thrust constant), and
^ = 84 - rO- in four-bladed screws.

H

^ = 9*4 - 1*2- in three-bladed screws.

H

a= 10*4 - 1-4- in two-bladed screws

H

y is the coefficient of friction and of resistance due to the shape
of the blade, the mean value of which may be taken at/= 0-045.

Xc, Yc, and Zc are values which depend upon the ratio between the dia-
meter and the pitch ( - J of the separate elements of the blade (see page

363) and on the shape and area of the blade. They may be expressed
by the following equations —

'^J^d' d '' ^'^~y ^ d d' ^^ V ^d' d

dr
d

Here h denotes the width of the developed blade, with radius - = r,
and X, Y, and z are functions of the ratio — for this radius. The

H

values for x, y, and z for various ratios — are given in the following

H

table : —

OI

U

8 MARINE ENGINKS AND BOILERS.

Table No. 41. — Values of x, y, z.
(See Taylor, " Resistance of Ships and Screw Propulsion.")

Diameter Ratio —

11

X.

Y.

1-048

z.

0-1

0-077

0-10

0-2

0-288

1-181

0-47

0-3

0-582

1-374

1-22

0-4

0-912

1-606

2-54

0-5

1-254

1-862

4-60

0-6

1-598

2-134

7-58

0-7

1-939

2-416

11-68

0-8

2-277

2-705

17-09

0-9

2-612

2-999

23-98

1-0

2-944

3-297

32-54

Table No. 42.

Values of

Xf, Yft

Zrfor the Standard Blade (see Fig

. 351)

•

D

H

Diameter

Ratio.

0-00

0-01

0 02

0*03

0-04

o-a>

0 06

!
0-07 0*08

1

0-09

\

K

1

L

i

Values of 3

i 1 1

I

0-4

0170

0-176

0-183 0189

0195

0-202 0-208

0*215

0-2221

0-229

0-5

0-236

0-243

0-250 0-257

0-264

0-272 0-279

0-286 0*294

0-301 ,

0-6

0-309

0-316

0-323 0-331

0-338

0*345 0-353

0*361 0-368:0-376

0-7

0-383

0-391

0-398 0-406

0-414

0-422 ! 0-430

0-438 ■ 0-445 0-453

0-8

0-461

0-469

0-471 0-479

0-492

0-500 0-.'508

0-516 0-523 0-531

0-9

0-539

0-546

0-555 0-562

0-570

0-578 0-586

0-593:0*601 0*609;

Values of Yf.

0-4

0-524 0-527

0-530 1 0-533 ' 0-537 i 0-r>40 0543

0-547

0-550 ; 0-553

0-5

0-557 0-560

0-564 0-568 0*572 0576 0580

0-585

0-589 0-593

J

0-6

0-598 0-653

0-608 0-613 0-618 0623 0628

o-6:u

0-640 0*64,')

0-7

0-6,52 0-658

0-664 0-670 0-676 0-682 0689 0-695

0*702 o-:<»8

0-8

0-714 0-721

0-727 0 733 0740 ; 0746 0752 . 0*759

0*765 0-772

1

0-9

0-778

0-784

0-791 0-797 0-804

0-810 0-817

0-823

0-83CJ 0-8:17

L
1

Values of Z(.

0-4

0-48

0-50 0-51

0-52 0-54

0-55

0-57

0-59 0*61 ' 0-63

0-5

0-65

0-68 0-70

0-73 , t)-75

0-78

0-81

0*84. 0*87 0*91 ;

0-6

0-94

0-98 1 -01

1 -05

1-10

114

119 ' 1-24 , 1*29

1-34

0-7

1-40

1 -45 1 -51

1 -57

1-63

1-69

1*76 , 1-83

1*90

1-97

0-8

2-04

2-11 2-19

2*27

2-34

2-43

2-51

2-60

2*68

2-77

0-9

2-87

2-97 3-06 315

3-25

3-36

3-46

3-57

3*68

3-78

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. ^79

For ordinary work the calculation of the values Xc y^, z^. (" charac-
teristics " of the propeller blade) by means of the above formulae and
Table No. 41 is generally too complicated. In almost all cases that
occur in practice, it is sufficient to take an ordinary blade, that is
a blade with a standard developed shape (Fig. 351), as a starting-point,
and to calculate from a table the values of Xc Vc, z^. In order.

Fig. 351.

however, to make this standard type of blade independent of the area
of the blade, we may write —

Xc = ^Xf, Yc = ^Vf, Z^ = bZf.

b being called the mean width ratio, a value which determines the ratio
of width to length of the blade. (Compare page 363.)

The values Xf, Vf, Zf for the standard blade can be taken from
Table No. 42. (The blades shown in Figs. 359 to 370 agree suffi-

380 MARINE ENGINES AND BOILERS.

ciently with general practice to form a basis for practical calculations.)
It should be noted that the standard blade can only be used to
determine the ratio of the width of each section of the blade to its
maximum width. If b is known, Xc, Yc, and z^ can be at once calculated
from Table No. 42.

By introducing the values x© Yu Zf we get the following formulae for
calculating the dimensions of the screw and its efficiency —

^H.p. = 3 X s(i?^^y D^^KXf +A).

\ 1,000/
S.H.P. = 3 X z[^^\H[asi\ - x,)x,-y(l - :r,)Yj.

For calculating the dimensions of the screw only the formula for
B.H.p. is required.

Method of Calculation, — b.h.p. is calculated from the formulas for
the resistance of the ship. (See § 211.) By a similar method the wake
is then determined. (See § 217.)

The speed due to the screiv is therefore v - u. A real slip is next
assumed. See §218 (4). Then—

and from this the value of nn can be obtained. A number of blades
of a given shape are next assumed (the shape should conform as
nearly as possible to that of the standard blade). This will determine
Xf and Zft /being always = 0045. As a rule the number of revolutions
is also assumed. This gives the pitch —

_ (v - u)6,086-44
«.(l-Jt)-60

- is so chosen, in accordance with the data given in § 218 (1), that a

n

suitable diameter is obtained. Thus —

_ D

D = - X H.

H

These are all the assumptions which may be made with safety, b can then
be obtained from the formula for b.h.p., which determines the area of the

blade. If b had been assumed, instead of the ratio - , the diameter d

H

might have been determined from the formula for b.h.p. If both -
® H

and b are assumed, the number of revolutions n can be calculated from

the same formula.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 381
The mean width ratio b may be selected from the table on page 373 for

Z , A

k =

According to the definition there given —

A

b= -

<'^)

Now as dn = CD, and

f= 01 4 to 0*17 for bosses with blades cast with them.
^=0-24 „ 0*26 „ „ screwed on.

Therefore

(4^)

d2

k Z , A *lz(\-C)

With the help of this formula and the data given for k on page 373,
b may be determined.

§ 223. Example of Taylor's Method of Calculating the
Dimensions and Shape of the Screw.— Let this be done, for

instance, for the S.S. " Deutschland." A speed of 23 knots is
required. From the formulae for ship's resistance, the indicated horse-
power (i.H.p.) for this speed is given as 36,000. Thus for one engine,
B.H.P. = 0*9 X 18,000 = 16,200. From actual practice, with ships of

similar build, a wake current of = 12 7o ™^y ^ assumed for this

class of vessel.

The speed, relative to the surrounding water, which the screw has to
impart to the ship is thus —

v-u = 23(l- 0-1 2) = 20-24 knots.

A slip corresponding to an efficient screw must be selected (see
page 373).

Assuming a somewhat high value of s^ say = 25 '/^, then —

0-75«H = 20-24 M?^:*l

60

Whence «h = 2,733 feet, and c = ^ ' " ' f ? = 27 knots.

o,0o6*44

382 MAtilNE ENGINES AND BOILERS.

This real slip of 25 7o> ^i^^ ^ wake current of 12 °/^, would correspond
to an apparent slip of

., = x,-J=25 7,-10-2 7, = U-8 7^.

This slip corresponds approximately to the apparent slip given by the
results obtained on the trial trip.

The diameter ratio may now be assumed. In determining this, the
shape of the stern post and the draught play an important part. Further,
the curve giving the best diameter ratio must be taken into account (see-
page 371). Allowing for all these, and having regard also to circum-
stances limiting the number of revolutions, a diameter ratio of — =0*65

H

may be chosen.

For further calculations we return to the original formula —

B. H. p. = 3 X i: (^ ^^^^ . D^(as,^t +y%,).

Having regard to the solidity and safe working of the screw, we will
take the number of blades as 2 = 4. The coefficient of thrust (" thrust
constant ") is thus —

<i = 8-4 - 0-65 = 7-75 ; and also/= 0-045 (see page 377).

The shape of the blade is assumed to be of the usual pattern (similar to
the standard type of blade shown on page 379). Therefore the values
Xft Yf, Z{ may be taken direct from Table No. 42, page 378, and with

2 = 0-65, we get Xf= 0-345, Vf=0-623, Zf=M4.

The number of revolutions n may be calculated from the formula
for B.H.P., and the mean width ratio ^ assumed, or if the number of
revolutions be assumed, a definite value will then be obtained for ^,
which determines the area of the blade. Assuming n — 79*5, the pitch
must now be calculated. As « = 79*5, « x h = 2733 feet, then h will
be = 34-35 feet (pitch of actual screw h = 34-44 feet). Hence the
diameter is obtained —

«

D = £ . H = 0-65 X 34-35 = 22-32 feet

H

which corresponds exactly to the diameter of the screw as fitted in the
ship. The mean width ratio b has next to be determined, and is
obtained by working out the equation for b.h.p. and b.

Shafting, resistance of ships, propellers. 383

The values within brackets become

(ijrXjXf+/Zf) = (7-75 X 0-25 X 0-345) + (0-045 X 1-U) = 0-719.

Therefore b = \^^^ = 0-1 85.

4x3x0-204x498-1x0-719

This value almost exactly corresponds to actual practice, in which the
developed area of a blade is in fact 35 square feet. Its length from

boss to tip --^—" = 8-28 feet. The mean width is /„, = 4-22 feet, and
thus the mean width ratio works out at

^ = ^• = 0-189.

The efficiency of this screw can now be calculated. It is as follows
(see page 377) —

s.H.p. asl\ - a)Xc-/(1 -^t)Yc

« B.H.P. aJtXc+^c

In this formula replace Xc by Xf and b will then be eliminated —

(Xc = ^Xf, &c.).

Hence —

as^iy - Jt)Xf-/l - Jt)Yf- 7-75 X 0-25 x 075 x 0-345

-0-045x0-75x0-623 = 0-48.

The denominator of the fraction has already been determined above ;
the efficiency is therefore —

0-48 ^ ^^

4

•

Strength of Propeller Blades
§ 224. Stress in the Propeller Blade due to Thrust and

Tangential Forces. — The greatest stress is at the root of the blade,
and for this reason it is always worked out for a point as near the boss
as possible.

Two forces, a turning force at right angles to the direction of the
shaft (tangential force t), and a force parallel to the shaft (thrust p), act
upon each element of the blade. These forces can be summed up as
a resultant thrust acting on a given point, which will produce the same
bending moment on the root of the blade, as the sum of all the thrust
forces on all the elements of the blade. The distance of this point
from the centre of the shaft is —

, D _ total thrust moment
^2~ total thrust

In the same way the distance of the point from the centre at which the
sum of all the tangential forces is supposed to be concentrated will be —

, D _ total moment of tangential forces
' 2 total tangential force

The constants k^ and k,, are found for the standard propeller blade
from Table No. 43. (For this and the following passages compare
Taylor.)

Table No. 43.
Consfan/s,

H

k,.

/V.J.

;

• D
H

K

k, '

0-4
0-5
0-6
0-7

0-706
0-710
0-692
0-684

0-646
0-658
0-644 ,
0-625

0-8
0-9
1-0

0-688
0-695
0-696

0-614
0-606
0-600

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 385

The thrust p is obtained from the following equation : —
Work due to the thrust per minute = s.h.p.= about 0-7 b.h.p. ;
• for greater safety the maximum efficiency of 70 "/^ is here taken.
Therefore —

\^ ^^Ji = 0*7 B.H.P. X - ft. lb. per mmute.

33,000 z

Taking s^ = 20 */^ = 0*2 for the normal slip j„ then —

28,875 . B.H.P.

p= — z

0X H X »

H being stated in feet, p in pounds, b.h.p. in horse-power.

The tangential force t for one blade is obtained from the well-known
equation —

Moment of t = 5JLZ: x 5,252 x 1

n z

Therefore t x >&. x ? = ?:iH- x 5,252 x 1

2 n z

whence- ^^B.jrP.^ 10.504

n yfcg X D X 5

D being in feet, t in pounds, and b.h.p. in brake horse-power.

If the section of the blade in question is at the distance ^ from-the

axis, then the —

Bending moment due toP = p(>^i^-^)

^"^V^^2~2/

To calculate the dimensions of the cross section of the blade, it is
necessaiy to know the bending moments about the longitudinal axis xx,
and about the transverse axis yy (Fig. 352). These are —

Bending moment about xx, m^ = Mt sin a -j- Mp cos a

„ „ YY, M.2 = Mp sin a - Mt COS a.

The curve at the root of the blade is generally taken as a parabola

2
having an area= ^ Ih (Fig. 353).

2

Distance of the centre of gravity from ab, d^ - A. Moment of

o

inertia about the axis through the centre of gravity, and parallel to

8 • . -1

ab = -=— /A*, Moment of inertia about the axis cd = — l^h,

17o o\J

1 u

386

MARINE ENGINES AND BOILERS.

From these values we get —

(a.) Moment of resistance for axis parallel to ab —

For AB, w. = ^M2. for c, w,= Am2

Fig. 352.

^^

Fig. 363.

(^) Moment of resistance for axis parallel to cd —

For A and B = -L /U,

15

Tension at a and b due to m, = — ^

^ 4 M2

Compression at c due to Mj = ^

Thence

8 '7^2
15mc

Tension at a due to Mo = *-^^^ll
Compression at b due to Mo = ^

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 387

35 M

The maximum tension at a = s^^ = ^ . y l> +

The maximum compression at c = Sji =

'^ 4 ' /£^ I'^h
_105 Mj^

8 " m

Assuming a value for /, k can be calculated from the formula for
s, and Sdi, the lower values of each being taken as a basis. The follow-
ing table gives the allowable stresses for different materials : —

Table No. 44.

Allowable Stresses in Propeller Blades^ due to Thrust and

Turning Moment,

1

1

1
1

Material of the Blades.

1

Working Stress in pounds per sq. in.

S^. Tension.

Sjj. Compression.

L

1

Cast iron

Cast steel

Bronze

Manganese or phosphor bronze -

1

2,000

5,000

3,000

5,000 to 8,500

6,000

10,000

4,000

6,000 to 12,000

The higher values given on the last line are only allowable in light
warships, in which, on account of the weight being so much reduced,
such high stresses are unavoidable.

§ 225. For Working Calculations, the above formulae may be
thus simplified : —

If B be taken as = 0-6, k^ = 0644, then t = 094?. To calculate the

dimensions of the cross section of the blade, for which a = about 45°
(which generally holds for the root of the blade) ; assuming t and p to

be exerted respectively at the distances k^- and ky^ from the centre of
the shaft, then —

M = (o*66? - ^ 1-94P cos 45"

= (o-66--01-4P

= /^0-66? - ^ 40,425 x b.h.p.
V 2 2/ TTWTn

where d, dy and h are in feet, b.h.p. in brake horse-power, and m in ft. lb.

388 MARINE ENGINES AND BOILERS.

If the ratio of ~ is greater than 0*6, the value m must be multiplied

rl

by a factor less than unity. In such cases, instead of m we should
have —

0-97Mif? = 0-65

H

0-94M „ =0-7

0-89m
0-85m
0-82m
0-80m

= 0-8
= 0-9
= 10
= M

The stress in the given cross section of the blade will thus be —

35 M , „ 105 M

The bending moment about the transverse axis may be neglected.
(Compare Example II., page 391.)

Example, — Twin-screw fast steamer " Deutschland." What is the
maximum stress in the blade at a distance of say 4*1 feet (50 inches)
from the centre of the shaft ?

Let D = 22-3 feet, h = 34-4 feet, « = 79-5, ^=41 feet, /= 4-25 feet,
A = 0-86 feet, b. h. p. = 16,200, « = 4, and H=0-65.

n

Therefore M = 0-97(0-66 x 1M5-4 1)^?1?^^1^^

^ ' 4 X 34-4 X 79-5

^ o^ 3-26 X 40,425 x 16,200 , * i oi aaa r. lu

«0-97 X — -V-: — wrri = about 191,000 ft. lb.

4 X 34-4 X 79-5 *

Hence the —

Maximum tension s,, = -— x -7—-^ — ^ ■ = 533,000 lb. per sq. ft

^ 4 4-25 X -74 f n

= 3,700 lb. per sq. in.

Maximum compression Sdx = —^ x ' = 800,000 lb. per sq. ft.

= 5,560 lb. per sq. in.
= about l'5s,i.
The material used is manganese bronze.

§ 226. Stresses in the Blades due to Centrifugal Force.— To

the stresses s^^ and s^i must be added the stress, due to the centrifugal
force, produced by that part of the blade which lies outside the section
now under consideration. Let G be the weight of this part (the shaded

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 389

area in Figs. 354 and 355), s its centre of gravity, w its circumferential
speed, r the radius of the circle it describes. Then the centrifugal force
acting on section xx (Fig. 354) is —

w* G

— X-

This produces in the blade the tensile stress Sb.^ = - lb. per square
inch ; f being the area of the cross section in square inches. The

Fig. 354.

Fig. 355.

maximum tensile stress in the blade (Fig. 354) at section xx is at the
points A and b (Fig. 353), and is —

S, = S,j + s^.
The maximum compression is at c and is —

Sd — Sdi — s

«2-

If the blade is inclined to the rear (Fig. 355), the centrifugal force
also produces a bending moment in the longitudinal plane of the blade.
The amount of this force on the section at the root is about Mc = 0-7^.
If we take the section as forming a segment of a parabola (Fig. 353),
then from this moment we get a tension at points a and b of —

o,« — — —

35

and a compression at c of —

M,

^^^'~^r 8

= l-5s

105

M2

«3'

390 MARINE ENGINES AND BOILERS.

In a blade placed at such an angle, the greatest total tension at a
and B, using the shortened formula, will be (for exact calculations the
tension at a should be more than that at b, see page 386, § 224) —

Maximum total tension s, = s,i + s^ + s^.

Maximum total compression Sd = s^i - s^ + s^g.

The centre of gravity of that part of the blade lying outside the
section here considered may approximately be taken as being at the

distance ^=o+ (o~9)^"^ ^^^^ '^^ *^^^*

Those sections of the blade which are calculated only for bending
stresses from the thrust and tangential force, must be tested, to see
whether they are able to withstand stresses due to centrifugal force, and
if necessary strengthened. For this reason the working stresses in
Table No. 44 will be exceeded by 20 to 30 per cent.

i^ 227. Example I. Effect of Centrifugal Force on the stress

in the blades of the screw of the " Deutschland." The stress in a
section of the blade is taken, as before, at a distance of 50 inches from
the centre of the shaft.

Weight of blade outside this section is 0 = 4,1 80 lb.

Distance of the centre of gravity of this part from the shaft is
r = 6-23 feet, and the area of the section is f = 357*5 square inches. If
« = 79-5, the circumferential speed will be —

w = 51 -8 feet per second.
Therefore c = - x — = 56,000 lb.

Hence s^,., = - = '^ ' ^ = about 1 56 lb. per square inch.
- F 357-5 '^ ^

The blade is inclined to the rear, as shown in Fig. 355 ; the centri-
fugal force therefore exerts not only the tension c, but also a bending
moment upon the section under consideration. ^=11 inches; the
moment of resistance of the section about its axis parallel to ab
(Fig. 353) is—

Tension (in ab) = /A^ = about 645 inches = w,, = 1 'bw^.

Compression (in c) = r-— ^ lA^ = about 430 inches = w^.
^ ^ 105

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 391

T^ . Mc 0-7 X ^ X c 0-7 X 11 X 56,000
Therefore s., = — ^ = = — — — J —

^ w, w, 645

= about 670 lb. per square inch.

, Mc 0-7 X tf X c 0-7 X 11 X 56,000

and Sh« = — ^ = - — = —

'' Wd w, 430

= about 1,000 lb. per square inch.
The total stresses upon the section are therefore —

Tension (in ab) s^ = s^j + 3,2 + ^^z = 3,700 + 156 + 670

= 4,526 lb. per square inch.

Compression (at c) Sj = s^^ - s^^ + Sd3 = 5560 - 156 + 1,000

= 6,404 lb. per square inch.

It is thus shown that the stress in the blade is greatly increased by
the centrifugal force, and this justifies the high factor of safety used in
the figures given in Table No. 44.

§ 228. Hxample II. — The centrifugal force has much more effect on
the blades of screws running at a high speed. In a recently constructed
destroyer, in which n = 325 ; i.h.p. = 2,760 ; h = 7-85 feet = 94-4 inches ;
/-1-64 feet =19-6 inches; r=2-18 feet = 2618 inches; <r = 0-26 feet =

3-14 inches; ? = l-08 = say M; 2 = 3; d = 8-54 feet=102-3 inches;

H

h = 0-295 feet = 3-54 inches ; d= 23 feet = 27-54 inches.
I A^ ~ n.« I ^sumed values. (Compare Table No. 43.)
Hence b.h.p. = i.h.p. x 0*85 = 2,345.

According to the formula in § 224 —

n , k^.iy . z
Bending moment due to t = t(^2 r, ~ r> ) =" 6>9S0 ft. lb.

P = ^-"-P-^^g!gI^ = 8,800 lb.

2. H . «

Bending moment due to p = p (i&j ^ - f ^ = 1 6,200 ft. lb.

From the graphic analysis. Fig. 356, we get —

Moment about the axis xx = Mj == 15,900 ft. lb.

YV = M2 = 7,230 „

392

MARINE ENGINES AND BOILERS.

These moments produce the following stresses —
Tension in ab due to Mj = — x tt-J ^ ttt = 6,750 lb. per sq. in.

Compression at c
Tension at a
Compression at b

ij

ij

ti

_ 105 Ml 1
M2 = 15x^x-L = 950

10,200 „

)>

it

»

>i

II

/2A 144

Hence the greatest tension at a = s^j = 6,750 + 950

= 7,700 lb. per square inch.
„ „ compression at c — Sdi = 10,200 „ „

^

Direction of
rotation'

M . 15900 Ft. lbs

/Y, . 7230 Ft lbs

Fig. 356.

The weight of the part of the blade outside the section, at a distance
~ = ri5 feet from the boss = o = 352 lb. Circumferential speed of the

centre of gravity s of the part of the blade having weight

Ivrn 2irx 2-18x325

w =

60

Therefore

60
352x74^
■j^ r 32-2 X 2-18

w

2

= 74 feet per second.
= 27,500 lb.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 393

Hence

c 27,500 1

S,2 = -

•325 144

= 590 lb. per square inch nearly.

/ r .• 2 ,, 1-64 X -295x2 «.^. - ^v

(f = area of section == - M = = -325 square foot.)

The moment of resistance of the section is —

For tension (in ab)

w,= ~ M2 = 28inche'53.

8

For compression (at c) Wj = ^yp ^^^ =18-6

>>

The moment of the centrifugal force is very considerable here,
because of the slanting position of the elements of the blade, and
the great centrifugal force exerted (^=3*14 inches) —

Me 0-7 X 314 X 27,500 « , ^^ lu • - ♦

S-- = -^= — — =2,160 lb. per square mca.*

"^ w, 28 ' f H

._ M, ^ 0-7 X 314 X 27,500 _^.^^^
'^«"^ 18^6 ^'-^^

»

>i

The total stress on this section of the blade is thus —

s, = S.1 + s^ + s^ = 7,705 + 590 + 2,160 = 10,450 lb. per sq. inch.
Sd = Sdi-s^ + Sd3= 10,200 -590 + 3,250 = 12,860 „

The tensile stress due to centrifugal force is thus shown to be con-
siderable in this case.

§ 229. Thickness of Tip of Blade.— For
solid bronze this thickness ^o is (Fig. 357) —

^o = TV ^o i i'^ch if D = from 6 to 10 feet.
^o= i to| „ if D= „ 10 to 12
^o= i toi „ if D= „ 12 to 19
1 ,, ifD= > 19

>»

>>

/4o =

))

»

For cast iron the thickness of blade at the tip
is about one and a half times the above. The
crowns of the sections at the tip and root are
joined by a straight line, and the thicknesses h for
each section are thus determined. The back of
each section is formed by the segment of a circle ; Fig. 357.

it is also usual to round away the edges of the
sections nearest the boss, on the working or thrust face, (See Figs.
361, 363.)

* This value, 2,160 lb. per square inch, is slightly raised by the components of
the centrifugal force exerted along xx, but the increase is so small that it may be
neglected.

394 MARINE ENGINES AND BOILERS.

The lowest part of the blade is rounded off into the flange with a
very ample curve. The edges of the blade are made as sharp as possible,
and in warships the blades are often finished bright all over.

§ 230. Materials used for Blades. — For all warships, and in
general for first-class steamers, manganese or phosphor bronze is used.
For cargo and passenger vessels of medium size, ice-breakers and large
tugs, cast steel is employed; and cast iron for ordinary cargo boats, and
for small and medium-sized lower class vessels. If cast iron be used, the
boss and blade are cast in one. Cast steel has the disadvantage that it
rusts easily, and also that it is liable to shrink during casting, and cause
inaccuracies in the surface of the screw. In small steam tugs the blades
are often made of steel plates, and riveted to the cast steel boss.

Construction of the Screw.

§ 231. Moulding and Casting the Screw (Fig. 358).— A vertical
spindle, such as is used for loam moulding, is set up on a cast-iron base
plate. On this spindle a striking board or arm is fitted in such a
manner that it is free to rotate, and to slide up and down. This board,
the striking edge of which sweeps out the surfaces of the blades, is
placed with the edge at right angles or inclined to the spindle, according
to the design of the screw. Round the spindle are fixed concen-

Fig. 358.

trically as many special templates as there are blades ; each template
being a triangular metal blade, bent to a radius somewhat larger

than . The angle of inclination a of their upper edges gives the

required pitch of the screw. The space between the spindle and the
ring of templates is filled in with loam. The striking board is then
turned, sliding on the templates which remain fixed, and as it revolves
its combined motion causes the loam to take the required helical

396 MARINE ENGINES AND BOILERS.

form for the face of the screw blades. The pattern thus obtained for
the back or thrust surfaces of the blades is baked; the centre lines
of the blade, and circles concentric to the striking spindle, are then
drawn on its surface. Upon these circles, and symmetrically about the
centre lines of the blades, thin metal templates are erected at right
angles to the moulded surface, representing the cross section of the
blade at the various distances from the centre. For the shape of the
latter see Figs. 361, 363. The spaces between these are filled with
loam, and the forward surface of the screw blades thus obtained is
smoothed off, dried, and painted, and used as the pattern for the
upper part of the mould. When this part is complete, tlie sectional
templates and dried loam are removed from the lower surface, and a
finished mould of the blade remains, which only requires the addition
of the boss or the flange to complete it.

§ 232. Explanation of the Drawings of Screws.

1. Figs. 359 to 362 show the screw of a small tug with one engine ;
the material used for the screw being cast steel.

i.H.p. = 200; v = 7 knots; «=125; D = 6-56feet;
H = 8*53 feet ; « = 4 ; sa = 21 '5 square feet.

The area of the blade in Fig. 359 is developed according to Method
I., § 214. The blades are slightly inclined to the rear. (See Fig. 362.)

2. Figs. 363 to 366 show the screw of a large twin-screw mail
steamer (passenger and cargo boat). The blades are of manganese
bronze ; the boss of cast iron.

i.H.p. = 2x 4,500; v = 14-5* knots; « = 80; D = 18-7 feet;
H = 21*3 feet ; 2 = 4; 2a = 99 square feet.

The area of the blade is developed according to Method I., § 214.
The studs for securing the blade are of Siemens-Martin steel ; the corre-
sponding cap nuts of forged bronze. The boss is strengthened at the
front and back by strong wrought-iron rings shrunk on. The holes in the
flange of the blade are oval, so that the pitch of the screw can be varied
from 20 to 22-7 feet. Here, again, the blades are slightly inclined to
the rear.

3. Figs. 367 to 370 show the screw of a twin-screw armoured cruiser.
The blades and boss are of high tension bronze.

i.H.p. = 2x 8,000; «=140; v = 21 knots; j?=3;
D = 17-4 feet ; h = 16*8 feet.

The blades are vertical to the axis of the screw.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 397

Fig. .359.

• I

I

'/ // ///////A ft/t/i/ft/fttf »-:

2|

Lzi^////////////////////////^

-J-

I

Fig. 36L

Fig. 362.

MARINE ENGINES AND BOILERS.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 399

Fig. 365.

400 MARINE ENGINES AND BOILERS.

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 401

^ 233. Propeller Boss. — 1. In smaller propellers the boss and
blades are cast in one (see Fig. 362). Length of boss l=2-3 to 2-6d.
Maximum diameter of boss <fn = 20 to 2'Sd, Slope of cone of the
propeller shaft 1 in 10 to 1 in 16.

As a rule the centre part of the boss is cut away, firstly to effect a
saving of weight, secondly to facilitate the fitting of the propeller on to
the conical end of the shaft. In order that the turning moment of the

— --t—

Fig. 367.

Fig. 368.

shaft may be transmitted to the boss at its thickest part, the latter must,
especially at the thicker end of the cone, fit accurately on to the shaft.
The boss is prevented from turning on the shaft by one or two strong
keys (two are generally used in warships, although one is really all that
is necessary, even for the largest screws).

Breadth of the key ^ = - + | inch. Thickness of key 05 to 0-6^,

d being the diameter of the propeller shaft. If there are two keys, only

2 c

402 MARINE ENGINES AND BOILERa

0-8^ instead of b is required. The keys must fit both boss and ctmed
shaft accurately at the sides, but a little clearance may be alloired in tbe
boss at the top. The boss is first fitted on to the shaft without the

:-■->'-

'^^'A'

Fig. 370.

keys, then removed, and the keys fitted to the shaft in countersunk
keyways. The boss is replaced, and it should be possible to push il «
far up the cone as before the keys were fitted.

The keys almost always extend the whole length of the boss ; but

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 403

sometimes, if the propeller is small, they occupy only the front half.
The propeller nut has a fine thread, and may be made with either
indentations or projections. (See Figs. 371, 372.)

Fig. 371.

Fig. 372.

Diameter of the nut d^^ = 1 '4 to 1 '5d^. Thickness of nut A^ = 0*75 to
0-S5d^ d^ being the diameter over the thread. The smaller values may
be used for larger nuts. These values hold for nuts where the shaft has
a diameter measured outside the thread of over 5| inches, otherwise d-^
is taken from the table of dimensions of bolts and nuts, and equals the
width across the flats of a hexagonal nut. (See Table No. 70, page 607.)
To prevent the nut slacking back, it should have a left-handed thread
for a right-handed screw propeller, and a right-handed thread for a
left-handed screw propeller. Some method of locking the nut is also
usually provided. To screw on the nut easily, the shaft is continued
for a short distance beyond the nut, and given a diameter slightly less
than that at the bottom of the thread.

2. Bosses with Blades bolted on, — In merchant vessels with propellers
over 10 to 13 feet diameter, and in warships with propellers over 6 feet
6 inches to 8 feet 6 inches diameter, the blades are bolted on to
the boss.

In the best practice, the flanges of the blade are very carefully fitted
to the surfaces on the boss, to prevent the water getting underneath
them, and sometimes a rubber ring is inserted, and screwed up against
the boss.

Thickness of flange of blade /i = 0*18 to 0*22^ for bronze or cast
steel.

Diameter of flange of blade Dj = 1*9 to 2 3^.

Corresponding to this diameter of flange, the external diameter of
the boss is —

<f„ = 2*6 to 3^ for large screws.
^„ = 3*0 to Z'bd for small screws.

404

MARINE ENGINES AND BOILERS.

Length of boss with blades bolted on, l = 21 to 2*6^ (higher values
are for smaller bosses).

Thickness of boss round the cone —

^1 = 0-19 to 0-22^ for bronze,
ri = 0-18 to 0-21^ for cast steel.
^1 = 0-22 to 0'2^d for cast iron.

Thickness of metal at front and back ends of boss —

a/, =

7V^ =

0'22d for bronze,
0-20^ for cast steel.
0*24^ for cast iron.

In all these formulae d is the diameter of the propeller shaft.

The blades are secured by studs fitting very tightly into the boss.

Total stud area at bottom of thread = 065 to 0*85 x area of the shaft.

The number and size of the studs may be more accurately calculated
by taking as a basis the stress upon them due to axial thrust, turning
moment, and the centrifugal force. (See " Strength of Propeller Blades.")

The number of studs should be from 6 to 10 for each blade.

MateriaL — The studs are best made of good steel, especially nickel
steel, on account of its great strength and toughness, even when there

are surface cracks. Bronze is sometimes used,
but is not so reliable, even if of the strongest
and best quality.

The holes in the flanges of the blade are
made oval, so that the pitch can be varied ; a
variation of from yy to -^ above and below
the mean pitch should be provided for. The
nuts used are generally cap nuts made of
bronze, and locked in a suitable way. (See
Fig. 373.)

Propeller Cap. — To protect the nut, and
Fig. 373. to reduce eddy currents, the boss of large

and fast ships is often fitted with a cover of
cast iron or bronze, which is screwed on to the back of the boss
(Figs. 366, 368).

General Remarks. — On the back face of the boss two very large
screwed holes are provided, into which bolts can be inserted, to draw
off the propeller. The arrangement is shown at Fig. 374. For large
propellers the hydraulic press used for dismantling the shaft couplings
is also employed to draw the propeller boss. A spare boss and one or
more spare blades are almost always carried, or a complete spare screw,

/»!»«'»>»■.'.

*^:^:^

"••'.<

W'

SHAFTING, RESISTANCE OF SHIPS, PROPELLERS. 405

when the boss and blades are cast in one. The conical bore of the
spare boss should be fitted as carefully as possible to the actual stern
shaft, and also to the spare shaft, before it leaves the shops. In any
case, for hurried repairs, it is advisable to have a template of the cone,
and of the nut and key of the propeller.

To prevent sea-water penetrating between the boss and the shaft,
the front end of the cone is made watertight by means of a small

Fig. 374.

stuffing box and rubber ring, or some similar arrangement. The whole
boss, where hollow, is filled in with tallow, and tallow is also forced into
the clearances in the conical cap on the boss through a hole provided
for that purpose. The nuts are frequently set round with cement, so
that the surface of the boss may be even and continuous. In war-
ships the nuts are generally covered in with a metal case, thus giving
the boss the shape of a smooth ball (Fig. 368). In order to obviate
any tendency to knock or shock in the screw, it is sometimes carefully

406 MARINE ENGINES AND BOILERS.

balanced so that its centre of gravity coincides exactly with the centre of
the shaft. This is effected by fitting leaden weights under the blades,
if they are screwed on, or at some other convenient place on the boss.

A hole is generally bored in each blade, at about one-third its
length from the tip, into which an eye-bolt can be inserted, for fitting or
removing the blade.

§ 234. Machining the Surface of the Blades.*— In the latest

modern practice the propeller blades of fast steamers are worked up
mechanically by means of a special machine in such a way, that the
surface is trued into the required mathematical shape of a helical screw.
The machine consists of a horizontal arm, which swings to and fro over
the blade ; the centre of oscillation coincides with the propeller axis. A
slide block works along this arm, on which a steel projection moves
vertically up and down in a socket, so that by the combined oscillating
movement of the arm, and vertical movement of the steel projection,
helical lines are described.

* German Patent 145007, Engineering, 1903.

PART IV.

PIPES AND CONNECTIONS.

SECTION I.

FLANGES, VALVES, ETC,

§ 235. General Remarks. — The pipes and connections should
be kept as simple and straight as possible, and the number of flanges,
hangers, valves, &c., as low as possible.

For expansion due to heat, drainage, &c., see Section III., page 423.
All the piping and connections are generally tested by water pressure to
double the working pressure.

§ 236. Pipe Connections. — For thickness of the metal of T pieces
and bends see "Thickness of Valve Bodies," page 415. No pipes
over \ inch diameter carrying high-pressure steam should have branches
brazed on. Screwed joints with unions or sockets (see Fig. 375) should
only be allowed with pipes under \ inch diameter.

Fig. 375.

Fig. 376.

§ 237. Flanges.— (TVawj/o/^r^^ iV^^Af)— [Tables Nos. 45 to 47 in
the German Edition, giving the proportions of gunmetal flanges either
riveted or brazed to copper pipes, have been omitted, because if
converted the equivalent English sizes would all be odd dimensions,
and also because the Engineering Standards Committee are very shortly
bringing out a set of Standard sizes for pipe flanges, which it is hoped
will be universally adopted. These Standards will be procurable from
the Secretary of the Engineering Standards Committee at 28 Victoria
Street, Westminster, S.W., or from Messrs Crosby Lockwood & Son at
7 Stationers' Hall Court, London, E.C. — L. S. R.]

410 MARINE ENGINES AND BOILERS.

Gunmetal fianges are generally brazed to copper pipes, and if of

lai^e diameter may be riveted in addition. (See Figs. 377 and 380.)

Wrought-iron flanges may be attached

to wrought-iron pipes by brazing,

screwing with a fine thread (see

Fig. 381), or by welding, which in

recent practice has been found to

answer well. Joints in iron pipes

■-1 may also be made by means of loose

p flanges. (See Fig. 382.) Gunmetal

-■* or bronze flanges must not be brazed

to iron pipes, on account of their

— 1-- different coefllcients of expansion. For

Fig. 377. lead-pipe connections see Fig. 383.

Figs. 376 and 377 are high pres-
sure, and 378, 379, and 380 low-pressure flanges.

g 238, Jointing.— Low-pressure flanges are jointed with mbber
insertion, &c. ; high-pressure water pipes with rubber insertion coptain-

Fig. 379. Fifi. 380.

ing fine wire gauze ; steam piping with asbestos, rubber asbestos,
corrugated copper rings, wire gauze and red lead, &c. Thick, flimsy
packings should be avoided, as they are apt to blow out It is
important that the flanges should fit the corresponding surfaces evenly
and accurately, and not be turned hollow, as otherwise leakage may
take place through the bolt holes. For steam piping lenticular packing
(metallic discs) is sometimes used ; with this kind of packing the
flanges must be strong and rigid. Pipes and connections which are
exposed to great variations of temperature, t.g., steam piping used
alternately for saturated and superheated steam, are very difficult to
keep perfectly tight. For this reason the practice of late has been lo
use no packing for the flanges of the u.p. steam pipes, but to make them
steam tight by carefully scraping upand polishing the surfaces. Flanges
where the packing is recessed (see Fig. 382) have the advantage tha
it cannot blow out, but on the other hand the pipes have to be forced
apart when being erected, and the flanges are thus subjected to con-
siderable strain.

PIPES AND CONNECTIONS. 411

§ 239. Bulkhead Fitting — Where pipes have to pass through
watertight bulkheads, so-called bulkhead flanges (Fig. 384) or bulkhead
stuffir^ boxes (Fig. 385) must be fitted.

Fig. 361.

Fig.;

Fig. 38».

§ 240. Extract from Reflations of German Lloyd's re-
specting Valves, Cocks, Pipe Connections, and Pumps.—
With the exception of the bilge and the water-ballast tank suctions, and
the steam winch and safety valve exhausts, all the piping must be

Fig. 384.
of copper, except where other material i

Fig. 38fl.

i expressly sanctioned by the
Surveyor of the German Lloyd's. All valves fitted to the ship's side,
and as far as possible all other valves and cocks, must be easily
accessible, placed as far as possible above the flooring of the boiler
and engine-room, and so arranged that no doubt can possibly exist
whether they are open or closed. The heads of all bolts securing the
valves and cocks to the ship's side must be countersunk.

All sea-water inlets must be provided with gratings, and also with a
small steam pipe for keeping the grating clear. Wherever there is any
possibility of water flooding the ship, the pipes must he fitted with two
entirely independent valves, so that, even if carelessly handled, flooding

412 MARINE ENGINES AND BOILERS.

of the ship cannot talce place. For the small watertight spaces in
the run of the ship independent suctions are not required. The
ends of the galvanised-iron suction pipes must be fitted with movable,
easily accessible, galvanised mud-boxes. If the ballast-tank suction or
the bilge-suction pipes are carried through the hold, they must be well
protected, to prevent their being damaged when loading and unloading
the cai^o. Neither steam pipes nor delivery pipes should be led throu^
the hold.

For seagoing ships, the valve seats of all pumps must be of bronze ;
the plungers and pump rods must have bronze rubbing surfaces, and

r«. 386.

pump cylinders must have bronze liners. Feed-pump plungers or
pistons may be of steel. In general all feed pumps (whether driven
direct or independent) must be fitted with a safety valve, which cannot
be shut off from the pump. With automatic feed pumps a safety valve
is not required-

§ 241. Valves. — A distinction is made between ^' slraight-lhrongh"

or globe z-aives (see Fig. 386) and rigkl-angled valves. (See Figs. 39S and

399.) As a rule the former offer more resistance to the passage of the

steam or fluid than the latter, and therefore cause more loss of pressure.

Material. — For v'alves exposed to pressures below 180 lb. per

PIPES AND CONNECTIONS. 413

square inch, cast iron may be used, but it should only be employed
above 130 lb. per square inch when economy is the very first con-
sideration. In merchant ships the exhaust steam valves are always,
and the bilge and ballast water valves are generally, made of cast iron.
These cast-iron valves have gunmetal valves, seats, spindles, stuffing
boxes, glands, and spindle nuts. Cast-steel valves are only used for
steam pipes when exposed to high pressure, and they have the same gun-
metal fittings as cast-iron valves. In warships cast steel is also often
used for those fittings of the bilge pipes which are carried through the
double bottom of the ship, or are situated immediately over it. Gun-
metal valves are largely used, especially for feed-water and small steam
pipes, and for the sea<water piping most liable to corrosion.

Vaive Seats and Cones. — The seats are generally fitted separately,
even with gunmetal valves, in order that they may be renewable. They
must be conical in shape, very tightly and accurately fitted, and held
in place by strong set screws. (See Fig. 387.) In spite of this, in iron
or steel valves, especially where superheated steam is used, the seats,
owing to unequal expansion, are apt to wear loose, and therefore of late
in cast-steel steam valves, the seats are also often made of steel. They
may be either conical or flat (Fig. 387). For steam they are made very
narrow, for water conical and somewhat wider. Angle of cone " = 30*
to 45° {Fig. 398). Width of seat for steam (see Fig. 387) about

b =" 0'08 + — ^ , d being the internal diameter of the pipe.

The Vahe is always of bronze, generally with three or four feather
guides. The Spindle is almost always fitted into the valve in such a

414 MARINE ENGINES AND BOILERS.

way that the latter can move freely. (See Fig. 387.) Care must be taken
that the guides are of such a length that, when the valve is open, it
is guided far enough to prevent it canting, as it closes. In ver)- large
horizontal pump valves it is desirable to have a wide lip to the x'alve
besides the ribs to guide it into the valve chest.

The Spindles are almost always of wrought bronze, but sometimes of
steel, with a square or V thread. In bronze valves the thread is turaed
in the cover, and in cast-iron valves a bronze nut is fitted into tbe
cover, (See Fig. 388.) The valve is opened by turning the spindle, or
the thread on the spindle may work in a nut fitted into a support or
bearing outside the valve, which is turned by a hand wheel or crank.

In the latter case the spindle must be prevented from turning round.
With this arrangement the frictional resistance to the motion of the
valve is considerably greater than where the spindle simply rotates.

The permissible stress in valve spindles of wrought bronze, either
tensile or compressive, at the bottom of the thread, is as follows:—

s = 2,000 to 3,000 lb. per square inch for small, up to 6,500 lb, per
square inch for large valves.
The total load must here be taken as the clear sectional area of the
seat X maximum pressure in the pipe. The thread of the spindle irnisl
be so cut that the valve is closed by turning the band wheel to the
right ((>., clockwise). This must be borne in mind when arranging to

PIPES AND CONNECTIONS. 415

work the valves at a distance, say, from the deck, and the hand wheels,
levers, &c., must be so set that they do not alter the direction of the
motion. Sometimes, in very large valves, where the steam pressure
falls upon the valve, the head of the spindle is made in the form of
a small valve, which opens a very little as the spindle lifts, and before
it carries the valve itself with it. In this way the pressure is equalised
above and below the valve, and the opening of the valve is easily
effected. (See Fig. 401.)

The Lift or amount the valve opens is generally somewhat more
than a quarter the internal diameter of the pipe.

(For valve flanges^ see Translator's Note, page 409.) In cast-iron
valves less than 4 inches diameter the flanges are about 30 '/^, and in
larger valves about 25 ""/^ thicker than gunmetal flanges. Care must be
taken that the bolts can be easily put in place, and especially that
with right-angled valves the flanges are not too close together. The
stress on the caver bolts or studs must be taken as the same as that on
the flange bolts ; and the stud holes must in no case be drilled through
into the steam or water space.

The Vaive Body must allow as free a passage as possible to the
fluid (water or steam), and must therefore be of ample size round the
seating.

Thickness of Vcdve Bodies (^= internal diameter of the pipe in
inches).

Bronze high-pressure valves \g_ d,p j^ ^ ,
for feed and steam pipes / 7,110 ^^

Bronze low-pressure valves - 8 = — — + ^ „

(5 = ^^^ + ^ „ for merchant ships.
d
S = 1=-^ + A " for warships.

Cast-iron high-pressure valves. See Table No. 51, page 433.

(Note. — Very large valve chests must be considerably thicker than
the values given in Table No. 51.)

8 = — -H ^ inch for exhaust steam pipes.

Cast-iron low-pressure valves -j .

S = ~ -H ^ „ for sea-water pipes.

The above thicknesses should only be used in low-pressure cast-iron
valves for pressures up to 85 lb. per square inch, and in high-pressure

416 MARINE ENGINES AND BOILERS.

valves for pressures up to 230 lb. per square inch. In light warships
the walls of the valve bodies are made considerably thinner. For valve
chests, see page 434, For double-seated valves, see page 141. Sea-
suction and discharge valves, see Section II., page 418. Nonreturn
and spring-loaded valves, see page 422. Safety and feed-water valves,
see Boiler Fittings, page 569.

Arrangement of Valves. — All valves inust be so arranged that they
are easily accessible, i.e., that packing can be easily inserted between the

Fig. 389.

flanges, and the bolts put in place, and that the hand wheel or lever
can be easily worked.

The arrangement of piping should also be as simple as possible, and
on this account, for the reasons stated above, " right-angled " valves are
to be preferred to " straight- through " valves. The steam should, as far
as possible, be on the under side of the valve when it is shut down,
in order that the stuffing box may be repacked when the valve is
closed, and the valve easily opened; and lastly, that in main engim-
stop valves, the valve can be adjusted with ease and accuraq' in
spite of backlash. The valves should be grouped as far as possible, to

PIPES ANH CONNECTIONS. 417

reduce the number of joints. Valves connected at ihe bottom end of a
long length of vertical piping must have dtain-cocks fitted above them.

S 242. Sluice Valves afe chiefly used for low-pressure pipes.

ITg. 390.

The valve bodies are in one or two
parts, either of bronze or of cast iron,
wit h gun metal fittings ; the valves them-
selves are nearly always of bronze.
The valve may have faces on one or
on both sides. Single-faced sluices
have lugs or projections on the back,
arranged so as to fit against corre- I'ig- 302.

spending lugs on the valve body, and

thus press the face firmly against the seat. The spindles are of bronze,
and the nuts are generally in the valve, some kind of index being pro-
vided to show how far the valve is open. The valves have guide blocks
at the sides, which run in grooves in the body, so that they do not twist
as they work.

S 243, Plug Cocks are used almost exclusively for water pipes.
For steam they are only employed as indicator cocks, blow-off cocks,
auxiliary steam-cocks, and sometimes instead of reversing valves, &c.
Steam-cocks of large diameter are difficult to move and seize easily,
and should therefore, if possible, be avoided. Cocks are generally used
for bilge pipes, lavatory pipes, and cooling water circulation, &c. They
are then usually made of bronze, with stuffing boxes, and are either right-
angled {Fig. 390) or " straight " or " throughway " (Figs. 391 and 392),

SECTION II.
UNDER'WATER FITTINGS,

J$ 244. Under-water Fittings. — Under this head are included in
general all those fittings which are connected to the outer skin of the
ship. They lie below the water-line, and form the connection from
the sea to the pipes inside the ship.

In cargo vessels the castings are mostly of cast iron, with gun-
metal fittings, whereas in the navy they are always of bronze throughout
Most of the valves in merchant ships consist of ordinary valves opening
inwards, but on warships the old so-called " Kingston " valves are still
frequently met with. These open from within outwards, and are there-
fore less easily accessible for examination and repairs (Figs. 393 and
394). For a smaller Kingston valve see Fig. 395. For smaller pipe
connections ordinary cocks are often used as sea-water cocks.

All openings in the outer skin of the ship, through which water
enters, should be provided with gratings of bronze, wrought iron, or
cast steel, the openings of which should not be more than f inch wide,
and their total sectional area at least 50 7o more than the area of the
valve itself (Fig. 393). Where single separate sea-valves are employed,
as in the stokehold, they are fitted direct to the outer skin, while as
a rule all those in the engine-room are connected to a cast-iron or
wrought-iron chest attached to the skin of the ship, and fitted with
one grating. If of cast iron, the chest is fixed to the skin by gunmetal
studs, if of wrought iron it is riveted on. In the first case the hole in
the skin is strengthened by means of a stiffening ring riveted to it, and
the studs are only fixed into the stiffening ring. If the gunmetal sea-
cocks and valves are fitted direct on to the skin of the ship, it is desir-
able to fit a light zinc collar at the inlet, which can be easily replaced,
and which prevents the outer skin from being eaten away. (See Fig.
396.) In fitting sea-cocks care must be taken that they do not get
stopped up with sand or mud. It is best therefore to have both valves
and valve chests fitted above the floor line, and if possible so arranged
that all the valves and cocks are visible, and can be worked without the
necessity of lifting the floor plates. In any case, all under-water fittings
must be so made that it can be seen at a glance whether they are open
or closed.

PIPES AND CONNECTIONS.

420

MAKINK ENGINES AND BOILERS.

It is best Co fit the gunmetal blow-oflT cocks for the boilers, evapo-
rators, &c., direct to the skin of the ship, and secure thcro with a
specially strong flange and stiffening ring. They are sometimes made
with a projection which passes through the side of the ship and has a
gunmetal flange fitted over it on the outside, in accordance with Lloyd's
Rules (see Fig. 397). It is a better plan, however, to leave the outer
skin quite smooth, and fit a zinc guard ring, which can be easily renewed,
round the orifice (Fig. 396). It is not good practice to have the blow-
olT cocks in a cast-iron chest attached to the skin, as they are liable to

Fig. 394.

Fig. 395.

crack, owing to the sudden and unequal heating caused by the rushes
of hot water. Boiler blow-off cocks must be so arranged that the kej-s
used to work them can only be removed when the cocks are closed.

To free the gratings from ice, &c., sea-cocks are of^en fitted with
a small steam pipe, the diameter of which should be from f to U
inch. It must be fitted directly to the valve chest or the neck
of the valve by means of a copper pipe, through which the steam
plays direct on to the gratings.

It is important that all sea-cocks be strongly constructed, especially if

PIPES AND CONNECTIONS.

421

of cast iron, as they are exposed to corrosion. Valves and seats should
be of bronze, valve spindles of Delta metal, and the thread of the screw
outside the valve box. The collar for carrying the
nut of the spindle is either wholly of bronze, or of
wrought iron with a gunmetal
nut. It is best to make the
glands and studs of the stuiT-
ing box of bronze. Sea-cocks
should always have some ar-
rangement toprevent theplug
being forced out (see Figs. 390
to 392). Bolts and studs,
the latter especially, for gun-
metal cocks or valves should
be of gunmetal, as iron bolls rust very quickly.

;; 245. The Discharge Valves or over

delivery valves of the different pumps are gen

above the water-line in merchant ships, and

of cast iron with bronze fittings, while ir

vessels they are made

>ronze, and ger

elow the wate

cc the numb

in the skin c

le of the sma

Ives are often

to open into the valve chest of the larger valves, the dimensions of
which are correspondingly increased. AH discharge valves for plunger
pumps should be arranged to open outwards automatically with the

422 MARINE ENGINES AND BOILERS.

pressure of the water, and be held in position when open, but not
when closed. (See Fig. 398.) This is of special importance with the
discharge valves of the air, circulating, and bilge pumps attached to
the main engine. Discharge valves serving these pumps are frequently
loaded with a spring (see Fig. 399), but this arrangement is not really
necessary for the discharge valves of centrifugal pumps.

To avoid difficulties in working the air-pump discharge valves of
engines, which, although generally worked surface condensing, can also
be worked jet condensing, these valves may be held closed by a care-
fully adjusted spring, allowing the valve to open automatically in case
of need.

Discharge valves, like sea-cocks, are secured to a stiffening ring
riveted to the skin of the ship, by studs, or bolts with coned heads. As
a rule these valves have a spigot passing through the outer skin, and
ending flush with it on the outside. If they are above the water-line,
the studs securing them may be of wrought iron ; if they are below the
water-line, or are made of bronze, gun metal studs must be used. All
discharge valves must be so arranged as to be accessible at all
times. If they cannot be fitted direct to the outer skin, on account of
the coal bunkers at the side or other impediments, a strong wrought-
iron pipe may be fixed between the skin and the valve. Where it passes
through the coal bunkers, &c., this pipe must be protected from injury
by a stout wood or metal covering.

SECTION III.

MAIN STEAM, AUXILIARY STEAM, AND EXHAUST

FIFING.

1. Main Steam Piping.

S 246. Main Steam Piping^.—If there is only one boiler, the
main steam pipe leads from the main stop valve on the boiler direct
to that on the h.p. valve chest of the engine. If there are several
boilers, the pipes from each boiler are generally connected, and only
one main steam pipe takes the steam to the engine. In the case of
engines of large power (especially on warships) working at high pressures,
instead of one large pipe there are sometimes several smaller ones leading
from the boilers to the engines, which are only connected up when they
reach the engine-room bulkhead. Smaller diameters, reduced thickness
of the piping, and greater safety in working are thus obtained. If there
are two engines, it is better that each should have its own separate
main steam pipe. Generally each of the above pipes can be shut off,
if necessary, at the forward engine-room bulkhead, and the port and
starboard pipes are connected at that point by a junction pipe which
can be shut off at will. If this cannot be done, any break which may
occur must be isolated by means of a blank flange.

^ 247. Draining of Steam Pipes. — To avoid what is known as
" water-hammer," the steam pipe leading from the boiler stop valve to
the engine is frequently given either a slope downwards or upwards.
With the latter arrangement any condensed water can, if the engine is
stopped, flow back to the boiler ; with the second it collects in front of
the stop valve at the engine, or in the separator. Bends in pipes where
any water can collect, or what are known as " water-pockets," must be
strenuously avoided, or if this is impossible, they should be fitted with
drain-cocks. In short steam pipes a separator (Figs. 400, 401) is not
absolutely necessary, but there should be one in every longer steam
pipe as near the engine as possible, to separate the condensed water
and also the priming water from the steam. If there is no separator, it
should be replaced by a drain-cock at the lowest part of the pipe, and
as close as possible to the engine. The condensed water can be led
either into the condenser or into the hot well. The latter is the better
plan, because it helps to heat the feed-water. To ensure continuous
drainage, automatic traps are sometimes fitted, as well as a separator.

424

MARINE ENGINES AND BOILERS.

but they do not as a rule work satisfactorily. It is generally sufficient
to draw off the condensed water by hand at intervals, or so to adjust the
separator drain-cock that the level of water in the separator is always
uniform. Sometimes a special vessel fitted with a gauge glass is pro-
vided below the separator, from which the water is discharged.

Fig. 400.

Fig. 401.

§ 248. Diameter of the Steam Piping.— The internal diameter
is calculated to give a mean speed of the steam through the piping of
from 6,000 to 8,000 feet per minute. (Compare page 128.) The internal
area of main steam pipes is generally from 8 to 10 "/^ less than the
sum of the areas of the branch or contributory pipes from each of the
boilers, as the frictional resistances are proportionately less in the larger
pipes. In the case of long steam pipes, slightly larger diameters should
be allowed than for shorter pipes. Loss of pressure from the boilers to
the valve chest of the h.p. cylinder should not exceed from 4 to 8 lb.

PIPKS AND CONNECTIONS.

425

S 249. Expansion due to Heat. — The steam pipes must be so
arranged that they can expand and contract freely. To effect this the
piping must have large expansion bends or expansion joints. In steam
pipes up to 5 inches internal diameter in which sufficient bends already
exist, special arrangements to allow of expansion are superfluous. Should,
however, the pipes be straight^ expansion joints must be fitted. Care
must also be taken that the expansion of the pipes is not interfered
with by the hangers and supports securing the pipes to the bulkheads
and decks, as any obstruction to free expansion may set up dangerous
strains in the pipes and flanges ; for this reason the pipes are only fixed
in a few places, and allowed to slide elsewhere loosely in the hangers.

Fig. 402.

If the pipes are coiled to allow for expansion due to heat, it is
advisable to calculate the vertical height of the coil according to the
following formula —

A = 2 sjd X L.

Where a = vertical height of the coil in inches.

L = distance between the two fixed ends of the coil in inches.
^= external diameter of the pipe in inches.

For iron pipes it is best to have the short bends of steel. These
should not, however, be used to take up the expansion of the pipes due
to heat, but provision should be made for this by means of a stuffing box
or expansion joint.

If expansion joints (Fig. 402) are used, care must be taken that the
pil^es can only move in the desired direction.

Those parts of the piping ivhich are assumed to be fixed must be suffi-
ciently strong to withstand the strains thrown upon them^ and special care
must be taken to prevent any movement in them. Bulkheads, deck
beams, &c., utilised to support the pipes, must be strengthened (Fig.

426

MARINE ENGINES AND BOILERS.

403). Stop valves, junction pieces, &c., on the boilers when fitted
vrith expansion joints must be strengthened on the boiler side, in order
that hendifig strains on the neck of the valves may be avoided. The gland
of the stuffing box is also often secured with stays to the other end of
the pipe, so that the thrust may be taken up by these stays (Figs.
404-406). In pipes where bends are employed, the strains upon the
supports can only be approximately determined.

In expansion joints the power acting upon the ends of the
pipes tending to make them move in an axial direction, that is, to

D^ir

force the ends of the pipes out of the joint, is p = -— - ./. p being

4

the thrust in pounds, d the ^jc/^r«fl/ diameter in inches of the pipe in the

V.

^

5:

«>

?»• -

«o

^

38

%*

<:

•*

•«*.

1

<§

§

51

SQ

..._._._. ^_ j

J

Fig. 403.

Stuffing box, and / the steam pressure in the pipe in pounds per square
inch above atmosphere. The friction of the stuffing box is here
neglected ; this friction may, if the vibrations of the ship are consider-
able, produce an extra tensile or compressive stress. The use of the
above formula gives quite sufficiently accurate results. The end of the
pipe which slides in the stuffing box is generally of bronze, and is joined
to the pipe itself by a flange (Figs. 402, 403).

To prevent it from being blown out of the stuffing box when the
pipe yields to the thrust p, ties or safety bolts are often fitted, which
connect the stuffing box to a fixed flange on the pipe. Sufficient play
must be allowed, when hot, between the movable flange and nuts of
these tie bolts, that when the pipe cools it does not throw a tensile
strain upon the bolts. The latter are fitted with lock nuts. These

PIPES AND CONNECTIONS.

427

Fig. 404.

Fig. 405.

safety bolts frequently form at the same time the studs for the stuffing
box (Fig. 403) ; it is best to use bronze nuts, as they do not rust, and
are easily screwed up even when hot.

428

MARINE ENGINES AND BOILERS.

The expansion of pipes per foot of length from 32" F. up to the
temperature of the steam is as follows : —

Material.

1

Steam Pressure — Pounds per sq. in. above Atmosphere.

1001b.
per sq. in.

130 lb.
per sq. in.

1601b.
per sq. in.

2001b.
per sq. in.

1 '

•2901b. 1
per sq. in.

Copper -

Wrought iron, steel,
or cast iron

Inch.

0-03
002

•

Inch.
0033

0023

Inch.

0-035
0-024

Inch.

0037
0-025

Inch.
0-038

1

0-026 ,

!

i§ 250. For Arrangement of the Main Steam Pipes in a large

passenger steamer see Figs. 404 to 406. The pipes marked 1, 3, and 4
convey the steam from the forward group and the two central groups of
boilers, while pipes marked 2 take the steam from the aftermost group of
boilers, consisting of three double-ended and one single-ended cylindrical
boiler. The pipes 1 and 2 on the port, and 3 and 4 on the starboard, are
all connected to a cast steel valve chest at the forward engine-room bulk-
head. From here the steam passes to the steam separator below, and
thence through a main stop valve to the main steam pipe to the engine.
The two port and starboard steam separators are connected by a pipe.
To allow for free expansion there are three stuffing boxes in the pipe,
between the steam separators and the engine. To take up the horizontal
thrust along the pipe, the bend in it is stayed, not only to the stop valve
on the steam separator, but also to the main stop valve on the h.p. valve
chest. The stop valves are fixed to the engine-room bulkheads, which
are specially stiffened ; the boiler stop valves are stayed against the
boiler shells ; and the tee piece into which the bend from the single-
ended boiler leads is similarly secured to the bulkhead and deck.

s<

§251. Thickness of Steam Pipes.

1. Thickness of Copper Pipes.
1. Brazed copper pipes

8=-^-*- -06 inch.*
5,688

2. Seamless drawn copper pipes 8 = '^ -f -04 inch.t

0,OoO

If very solidly made S = i^t + -08 in. t If very solidly made 6 = p^ + "08

5,688 a, boo

m.

PIPKS AND CONNKCTIONS.

429

For drawn copper pipes (bound with steel wire) from 4 to 16 inches
the following formula is frequently used : —

\5,688/

i/+0-08

For copper pipes (bound with steel wire) above 16 inches internal
diameter —

■"5;688

d being the inside diameter of the pipe in inches, / the steam pressure
in pounds per square inch above atmosphere, and 3 the thickness of the
pi[>e in inches.

The thickness of copper pipes which have to be bent should be from-
0-02 to 0-04 inch more than the above, but the smallest mean radius of

'^^^toKirr- X

Fig. 406.

the curve must not be less than id. If the diameter of these pipes is
more than 4 inches, they are frequently strengthened by winding thin
steel wire (about No. 11 S.W.G.) round them. This must be so
arranged that if the wire is broken anywhere only a few coils can get
loose. On account of their greater cost and lower tensile strength
copper pipes are now being more and more superseded by iron and steel
pipes. Brazed copper pipes are now seldom used for main steam pipes.
Wrought iron pipes are welded, steel pipes are seamless drawn or
rolled (" Mannesman " tubes).

d "b

3. Thickness of welded iron pipes 8 = ^-^tt; + '12 inch.

^^ 7,110

4. Thickness of seamless steel pipes 8= --^ + "10 inch.

^ ^ 9,954

430

MARINE ENGINES AND BOILERS.

Table No. 48.
Thicknesses of Solid Drawn and Brazed Copper Steam Pipes.
(Calculated from Formulae 1 and 2.)

•

V

s

s
1

c

130 lb. Pressure

above

Atmosphere.

1601b. Pressure

above

Atmosphere.

185 lb. Pressure

above

Atmosphere.

•2301b. Pressure

above

Atmosphere.

Solid
Drawn.

Brazed.

Solid
Drawn.

Brazed.

1

S*^"'' Brnred.
Drawn.

1

Solid
Drawn.

Brar^.

Inches.
1

It

l|

Inch.

0-068

0-068

0-068

0-078

Inch.
0-088
0-088
0-088
0-098

Inch.
0-068
0-078
0-078
0-088

Inch.
0-088
0-098
0098
0-108

Inch.

0-078
0-078
0-088
0-098

Inch.

0-098
0-098
0-108
0-118

Inch.

0-078
0-088
0-098
0-108

1
Inch.

0-098 :

0108
0-118

0-127

1

2
2|

0-078
0-088
0-088
0-098

0-098
0-108
0-108
0-118

0-098
0-098
0-108
0-108

0-118
0-118
0-127
0127

0-108
0-108
0-118
0-127

0-127
0127
0137
0-147

0-118
0127
0-137
0147

0137
0-147
0157
0167

3
3i

0-098
0-108
0-108

0-118
0-127
0127

0-118
0-127
0-127

0-137
0-147
0-147

0-137
0-137
0-147

0-157
0-157
0-169

0-157
0-169
0-177

0-177 i
0-187 i
0196

4
4*

0-118
0137

0-137
0-157

0-137
0-157

0-157
0177

0-157
0-187

0-177
0-206

0187
0-216

0-206
0-236 ,

5

51

0-137
0-147

0-157
0-169

0-169
0-177

0-187
0196

0196
0-206

0-216
0-226

0-236
0-246

0-256 '
0-265 ,

6
6J

0-169
0-177

0-187
0-196

0-196
0-206

0-216
0-226

0-226
0-246

0-246
0-265

0-275
0-295

0-295

0-314

1

7
7*

0-187
0-187

0-206
0-206

0-216
0-226

0-236
0-246

0-256
0-265

0-275
0-285

0-305
0-314

0-324
0-334 1

8
81

0-206
0-216

0-226
0-236

0-246
0-256

0-265
0-275

0-285
0-305

0-305
0-324

9
9i

0-226
0-236

0-246
0-256

0-265
0-275

0-285
0-295

0-314
0-324

0-334
0-344

1

10

0-236

0-256

0-285

0-305

0-334

0-354

PIPES AND CONNECTIONS.

431

Table No. 49.

Thicknesses of Lap Welded WraughUiron Steam Pipes.

(Calculated from Formula 3.)

Internal
Diameter.

Pressures above Atmosphere.

130 lb. per
square inch.

160 lb. per
square inch.

Inch.

0-157
0-169
0-169
0177

185 lb. per
square inch.

2.30 lb. per
square inch.

Inches.
2

2'

Inch.
0147
0157
0157
0-157

Inch.
0-169

0-177
0-177
0-187

Inch.

0-187
0-187
0-196
0-196

3

•'4
3i

3|

0169
0-177
0-177
0177

0-177
0-187
0-187
0-196

0-196
0-196
0-206
0-206

0-216
0-216
0-226
0-226

4

0187
0196

0-196
0-216

0-216
0-226

0-236
0-256

5

0196
0-206

0-216
0-226

0-246
0-246

0-275
0-285

6

6i

0-216
0-226

0-246
0-256

0-265
0-275

0-305
0-314

7

7J

0-236
0-236

0-256
0-265

0-285
0-295

0-334
0-344

8

0-246
0-256

0-275
0-295

0-305
0-324

0-354
0-374

9

0-265

0-305

0-344

0-403

10

0-285

0-3-24

0-364

0-423

11

0-295

0-334

0-383

0-452

12

0-324

0-374

0-423

0-492

13

0-334

0-383

0-442

0-521

14

0-344

0-403

0-462

0-541

15

0-354

0-413

0-472

0-570

:]2

MARINE ENGINES AND BOILERS.

Table No. 50.

Thicknesses of Solid Drawn Steel Steam Pipes.

(Calculated from Formula 4.)

Pressures above Atmosphere.

Internal
Diameter.

130 lb. per
square inch.

160 lb. per
square inch.

185 lb. per
square inch.

230 lb. per
square inch.

Inches.
2

H

Ol

Inch.
0127

0127
0-127
0-127

Inch.
0-127
0-137
0137
0-137

Inch.

0-137

0137
0137
0-147

Inch.

0-147
0-147
0-157
0-157

3

H

3|

0137
0137
0137
0-137

0-147
0-147
0-147
0-157

0157
0157
0-157
0-169

0-169
0-169
0-177
0-177

4

*5

0147
0157

0-157
0-169

0-169
0177

0-187
0-196

5

5i

0-157
0157

0-169
0-177

0-187
0-196

0-206
0-216

6
61-

0169
0177

0-187
0-196

0-206
0-216

0-236
0-246

7
7*

0177
0-187

0-196
0-206

0-216
0-226

0-246
0-256

8
8i

0-187
0-196

0-206
0-216

0-236
0-246

0-265
0-285

9

0-206

0-236

0-256

0-305

10

0-216

0-246

0-275

0-314

11

0-226

0-256

0-285

0-334

12

0-246

0-275

0-314

0-364 1

1

13

0-256

0-295

0-324

0-383

14

0-265

0-305

0-334

0-403

15

0-275

0-314

0-354

0-413

PIPES AND CONNECTIONS.

433

The minimum working stress on the material (which for welded
plp)es, according to formula 3, is about 3,500 lb. per square inch)
must be taken, because under certain conditions the pipes are apt to
corrode, and moreover they cannot be examined internally. In large
pipes the welded seam is sometimes strengthened with a butt strap
riveted on. As considerable difficulty is experienced in bending thick
pipes, wrought-iron and steel pipes of over say 3 inches in diameter
are, where possible, only used in straight lengths, separate bends of
cast steel being provided where bends are necessary.

The thicknesses of cast steel or bronze pipe bends and tee-pieces may
be determined from the following formula : —

(1.) Cast steel bends (on account of difficulties in casting) —

7,110

+ '2 inch for warships ; and

8 = --r^ +■ '2 inch for merchant vessels.
5.680

(2.) Bronze bends — S

TjlO

+ -2 inch.

In warships the thickness of the bronze may be even less.

Cast-iron pipe bends are only used for pressures up to about 185 lb.
per square inch above atmosphere. For thicknesses of cast-iron valves
and bends for steam piping see following table : —

Table No. 51.
Thicknesses of Cast-iron Valves and Bends for Steam Pipes.

Internal

100 lb. per

130 lb. per

160 lb. per

190 lb. per

Diameter.

square inch.

square inch.

square inch.

square inch.

Inches.

Inch.

Inch.

Inch.

Inch.

2

0-35

0-39

0-43

0-47

4

0-47

0-51

0-55

0-59

6

0-55

0-59

0-67

0-70

8

063

0-66

0-74

0-78

10

0-66

0-74

0-82

0-90

12

0-70

0-78

0-90

0-98

U

0-74

0-86

0-98

1-06

S 252. Lag^g^ing^. — To avoid as far as possible losses of heat due to
rculiation from the steam piping, these pipes are coated with a non-
conducting material, such as fossil meal, asbestos, cork, felt, &c., with
an outer covering of sail-cloth, sheet iron, or other suitable material.
The bends and tee-pieces are sometimes covered with lead sheeting.
The flanges are generally left bare, or only provided with a light and
easily removable covering of sheet metal or asbestos.

4;U MARINE EXGIXES AND BOILERS.

2. Auxiliary Steam Piping.

§ 253. Auxiliary Steam Piping.— To drive the various pumps,
&c., in the engine and boiler rooms auxiliary steam pipes are used,
which take steam direct from the boiler through a separate valve. If
there are several boilers in the ship, each boiler has generally a suffi-
ciently large stop valve to supply all the auxiliary steam, but where there
are a very large number of boilers the auxiliary steam pipes are generally
connected only to a few of them. It is best to fit the stop valves
supplying these pipes with loose valves, not attached to the spindles,
so that, should the wrong valve be opened by mistake, steam from the
auxiliary steam pipes cannot find its way into any boiler which is laid
off or not working. In general the same rules hold good for ihe
auxiliary as for the main steam pipes.

Branch pipes leading from the main auxih'ary steam pipe to the
different pumps, &c., must be connected to them by cast-iron, bronze, or
cast-steel tee pieces or bends. With the high steam pressures now in
vogue brazed copper bends are not strong enough. Each branch pipe
is generally fitted with a stop valve where it branches off at the main.
Cocks should not be used in these pipes, as they are never perfectly
tight, and are often difficult to work. At suitable places in the auxiliar)'
steam piping it is well to have distributing valve chests, from which the
pipes are led to the various auxiliary engines; by this means the number
of flanges required is considerably reduced. Besides the connection of
the auxiliary steam pipe to the boiler, it is sometimes also connected
in the engine-room to the main steam piping. This arrangement is,
however, only suitable in ships in which the auxiliary engines for the
centrifugal pumps, steam feed-water pumps, evaporators, &c., have to
be kept at work while the main engines are running.

When arranging the piping to the steam whistle, special care must
be taken that the water condensed in the pipes can flow back into the
boiler, and cannot collect anywhere in pockets. Drain-cocks or valves
should be provided at all points in the auxiliary steam pipes where water
can accumulate when the auxiliary engines are not working. This water
may either be discharged into the bilge, into a condensed water tank
fitted in the bilges, or into the auxiliary engine exhaust pipe. The
internal diameter of the auxiliary steam pipe depends upon the size
and number of auxiliary engines at work at the same time. It is only
necessary to make the area of this pipe from 0-66 to 0*75 of the total
area of all the branch pipes supplying steam to the auxiliary engines.

PIPES AND CONNFXTIONS. 435

3. Exhaust Steam Piping.

^ 254. The Main Exhaust of large engines is led direct into the
main condenser. In small engines on river steamers the exhaust is
frequently led direct to the atmosphere, or may be arranged to dis-
charge either to the atmosphere or to a jet condenser at will.

^ 255. The Auxiliary Engine Exhausts are generally led into
one common pipe. This exhaust may be led either into the

Main condenser,

Auxiliary condenser (if there is one),
Feed-water heater (if there is one),
Or into the open air.

It should be possible to shut off each of the branch pipes by a valve or
cock. One of these valves, generally the one exhausting into the open
air, should take the form of a safety valve, to prevent the pressure of
steam in the exhaust becoming too high through careless management.
The valve should be loaded to a pressure of about 30 lb. per square
inch. The exhaust from the steering engines and from the winch for
raising the anchors is often led, not only into the general exhaust pipe,
but also, alternatively, direct into the open air. The exhaust of the
electric light engines should be taken directly into the main or auxiliary
condenser, so that the steady running of these engines may not be
affected by the varying steam pressure in the exhaust pipe. For use in
case of emergency a connection to the auxiliary engine exhaust pipe
may also be provided. The exhaust pipes of the different auxiliary
engines should be so arranged that each engine can be shut off from the
auxiliary exhaust. The drainage from the auxiliary engines is some-
times led into the auxiliary exhaust. For the blow-off pipe to the safety
valve see page 552.

§ 255a. The Diameter of the Exhaust Pipes of the various

engines should be about 1-2 x diameter of the steam pipe; the diameter
of the main auxiliary exhaust pipe may be taken at slightly less than that
corresponding to the total area of the pipes leading into it.

§ 255b. Thickness of Exhaust Pipes.

Up to 2 inches (internal diameter) = 0*06 inch.
From 2 to 4 inches ( „ „ ) =008 „

,, 4tp8 „ ( „ „ ) =0-10 „

Beyond 8 „ ( „ „ ) =0-12 „

The main exhaust piping should, if possible, be arranged with a fall
towards the condenser, and any unavoidable water-pockets must be
provided with drain-cocks.

SECTION IV.
FEED-WATER PIPES.

% 256. Boiler Feed Pipes. — These serve to supply feed-water to
the boilers. As each boiler installation must, according to Lloyds
Rules, be provided with two entirely independent feed-water systems, it
is usual to have also two quite independent feed-water pipes leading to
separate feed-check valves on each boiler. The pipe which, under
ordinary working conditions, conveys the feed- water to the boilers is
generally called the main feed pipe, and that through which the supple-
mentary supply can be fed into the boilers from the donkey feed pump
or an injector is called the atixiliary feed pipe. The two pipes are
often so interconnected by valves that if the supplementary supply is
damaged the main feed pumps can be utilised, and the engine or maiil
feed-water pumps can deliver through the auxiliary feed pipe. Formerly
the feed-water pipes were usually placed below the floor plates of the
engine and boiler rooms, but in modern practice it is customar)* to
fix them at about the level of the upper part of the boiler, so that the
flanges and pipes may be the more easily kept in view.

§257. General Arrang^ement. — 1. In -the oldest and simplest
arrangement of the main feed-water pipes, the feed pumps, driven from
the engine, drew the hot water from the hot well, and delivered it through
the main feed pipe direct to the boiler. In later practice this method is
only usual in small engines up to about 300 i.h.p.

2. In larger engines a feed-water heater (see page 445) and some-
times also a feed-water filter (see page 439) are fitted into the main feed
pipe between the engine feed pump and the boiler. *

3. In merchant ships with engines indicating more than 1,000 h.p.
the arrangement shown in Fig. 407 is frequently met with. The feed
pumps connected to and driven by the main engines deliver the feed-
water drawn from the hot well into a Weir (see Fig. 415), Blake, or other
type of feed heater fitted as high up as possible in the engine-room,
where the feed-water is heated by the auxiliary exhaust steam. From
the heater the feed-water passes to an automatic duplex steam pump

PIPES AND CONNECTIONS.

437

on the floor of the engine-room, which discharges the heated water
through the main feed pipe to the boiler. The speed of the auto-

Fig. 407.

matic feed pump is regulated by a float in the lower part of the feed
heater, which controls the cock supplying steam to the feed pump.

4:38 MARINE ENGINES AND BOILERS.

In Fig. 407—

1 = steam feed pump suction from hot well.

2 = „ „ „ condenser.
o = ,, ,, „ sea.

4 = „ „ „ feed heater.

5 = auxiliary feed pipe.

6 = main feed pipe.

7 = feed pump delivery pipe to auxiliary feed pipe.

8 = „ „ feed heater.

A = steam from the boiler or auxiliary steam pipe to the relating

cock on feed heater.
<7 = steam direct from boiler or auxiliary steam pipe to feed

pumps.
B = steam from the regulating cock on the feed heater to the

feed pumps,
c = auxiliary engine exhaust.
E = „ „ to condenser.

F = exhaust from auxiliary engines to the l.p. valve casing,
t; = auxiliary exhaust to feed heater.
H = air pipe to condenser.
J = steam to feed pumps.
K = feed pumps exhaust.

In the arrangement shown the two feed pumps may be worked
at the same time, or only one may be worked, the other being kept
in reserve, and not used under ordinary circumstances. The feed-
water filter can be fitted eithpr in the feed-pump delivery pipe 8 from
the engine to the heater, or in the main feed pipe 6.

4. In the feed pipe arrangements of many of the most recent vessels,
only the air pump is driven from the main engines, and there are no
main engine-driven feed pumps. The air pump delivers the condensed
water into a so-called "float tank" (see page 441), from which the two
steam-driven feed pumps draw. Under ordinary working conditions,
however, only one of these pumps draws from this tank, its speed being
regulated by the float. The water is forced into a jet feed heater, which,
as mentioned under heading 3, is fixed as high in the engine-room as
possible. The feed-water is heated by means of the auxiliary engine
exhaust. The second pump then draws the heated feed-water from the
heater, and delivers it into the main feed pipe, the speed of the pump
being regulated by the float in the feed heater. With this arrangement
a filter can be fitted either in the air-pump discharge to the hot well, in

PIP1::S AND CONNECTIONS. 4:!9

the feed-pump delivery to the feed heater, or in the main feed pipe.
Instead of driving the air pump from the main engine it may be
separated from it, and worked as an independent pump, on the Blake,
Weir, Worthington, or other systeni.

5. With air pumps separated from the main engine the water may
be sent direct into a Weir's feed heater, fitted high up in the vessel, from
whence the steam-dtiven feed

pump draws it (the speed of the
pump being regulated by a float),
and delivers it through a filter into
the boiler.

6. In warships each air pump
generally discharges into a hot well,
from which the feed pumps in the
boiler-room draw their supply.

This hot well is sometimes also
arranged to act as a feed-water
heater and purifier, the steam from
the auxiliary steam pipe, and from
the exhaust of the evaporaiois and
drain pipes, being led into it.
The water may be passed through
a layer of coke and sponge to
cleanse it (Schultz system).

Remarks. — As it is necessarj-,
for the preservation of the boiler,
to prevent, as far as possible, air
getting into it, care must be taken
when arranging the feed pump
suction pipes, that as little air as
possible mixes with the water. If

a filter is fitted in the feed pipe, 1-ic- ■*<>»■

a working pressure = boiler pres-
sure +• about .10 lb. per square inch must be allowed for. If it is
fitted in the suction pipe or in the delivery pipe to the Weir's feed
heater, no allowance need be made for any special pressure.

S 258. Feed-water Filter.— Compare S 2.'J7 (6). Blake's filter

may be cited as a typical example of this class of filter {Fig. 408). llie
water enters at the bottom of the hollow cylindrical space, and passes
through filtering cloths, which are placed between alternate layers of

440

MARINE KNGINKS AND BOILERS.

Fig. 44)9

w

^=k

mJ

SktAAas&sa^

Ur

' In Silled ^ziwMnj

Jm/yuptu Oit.%anatmtnt

Fig. 410.

^

1/

tin

/£g</ A/iver^

3

Proper Arran^meitf'

Fig. 411.

PIPES AND CONNECTIONS. 441

sieves, thirteen of which are shown one above the other in Fig. 408.
The sieves and plates for directing the water are pressed together by
a spindle. The clean water passes into the outer shell, and from thence
into the pipe. To clean the filter the cloths are taken out, or else steam
is forced through the reverse way, the dirt being washed down on to the
plates, and then blown out through the central passage. A bye-pass
and valves are fitted to the filter, so that the feed-water can be led direct
to the boiler, while the filter is being cleaned.

§ 259. The Float Tank is generally a galvanised sheet-iron box
(see Fig. 409) in which the float is arranged so that it is not -disturbed
by the intermittent action of the air pump-discharge. A rod carried
upwards from the upper lever of the float works the cock supplying
steam to the Weir feed pump. The float tank, besides a movable
cover for admitting the float, and cleaning holes, has the following
connections : —

Main feed pump suction.

Auxiliar>' feed pump suction.

Feed-heater pump suction.

Air-pump discharge.

Overflow from float tank.

Overflow from feed heater.

Auxiliary condenser air-pump discharge.

Jacket drain.

Cylinder drain.

S 260. Diameter of Suction and Delivery Pipes.— Both

the delivery and the long-suction pipes of the feed pumps should
have air vessels, as near the pump as possible ; these should be
arranged so that the water enters from below. (See Figs. 410 and
411.) The suction pipes should be placed so that air pockets can-
not occur, and should if possible be sloped up towards the pump.
The diameters of the steam feed-pump, suction and delivery pipes
should be so proportioned that the speed of the water in the former
under normal working conditions is from 230 to 315 feet per minute,
and in the latter from 300 to 500 feet per minute. The smaller values
are for pipes where the ratio of length to diameter is large, the higher
values for pipes where ratio of length to diameter is small. Using as a
basis the quantities of feed-water per i.h.p. per hour given on page 310,
the sectional area in square inches of the feed pipes per i.h.p. can be
obtained from the following table : —

442

MARINE ENGINES AND BOILERS.

Table No. 52.

Sectional Area of Suction and Delivery Pipes of Feed Pumps per i.h.p.

for different Speeds of Water,

Velocity of

Compound

Triple Expansion

Quadruple

Water.

Engines.

Engines.

Expansion Engines.

Feet per minute.

Square inch.

Square inch.

Square inch.

230

•00325

•00263

•00248

270

•00279

•00232

■00217

300

•00263

•00217

•00196

315

•00248

•00196

•00186

350

•00217

•00186

•00170

400

•00196

•00170

•00155

500

r

•00155

•00124

•00124

The total sectional area of the feed pipes in square inches may be
found by multiplying the value in the table by the i.h.p. of the engine.
The areas of the branch pipes leading off to the several boilers or
groups of boilers must be of such dimensions that the velocity of the
water through them is rather less than in the main feed pipe.

J5 261. The Thicknesses of Copper Delivery Pipes are deter-

mined in the same way as those of the steam pipes, but to allow for
unavoidable shocks occurring in these pipes a pressure of 1*3 times the
boiler pressure must be taken as the basis of calculation. Therefore—

For drawn pipes ^- rVztu ■•" '^^ \x\q\\, \

Ojboo y

For brazed pipes 5= n^f ^ '^^ ^"ch. |

See Table No. 53.

In war vessels the thicknesses are the .same as for steam pipes, i.e.,
they are calculated from the formula —

In I he above formula d is the internal diameter of the pipe in inches,
/ boiler pressure in pounds per square inch above atmosphere.

PIPKS AM) CONNECTIONS.

443

Table No. 53.
Thickness of Copper Feed-water Pipes in Merchant Ships,

5 = -- - - .-^ + -04 or 06 inch respectively.
5,6oo

e S

Inches.
1

U
1}.

If

-4

3
3
3i

4

5
^*

6-1

I

■ •«
8

130 lb. per
square inch.

Drawn.

Inch.

0-068

0-078

0-078

0-088

0-098
0-098
0-108
0-108

0-118
0-127
0-127

Brazed.

Inch.
0-088
0-098
0-098
0-108

0-118
0-118
0-127
0-127

0-137
0-147
0-147

0-137 0-157
0-157 ' 0-177

0-177
0-187

0-206
0-216

0-226

0-236 . 0-255

0-246 0-265

0-196
0-206

0-226
0-236

0-246

160 lb. per
square inch.

Drawn.

Inch.

0-068

0-088

0-088

0-098

0-108
0-118
0-118
0-127

0-137
0-147
0-157

0-169
0-196

0-206
0-216

0-246
0-255

0-265
0-285

0-295

Brazed.

Inch.
0-088
0-108
0-108
0-118

0-127
0-137
0-137
0-147

0-157
0-169
0-177

0-265
0-275

0-285
0-305

0-314

185 lb. per
square inch.

Drawn.

Inch.
0-078
0-098
0-098
0-108

0-127
0-127
0-137
0-147

0-169
0-169
0-177

0-187 0196

0-216 0-226

0-2-26 0-236

0-236 0-255

0-285
0-305

0-314
0-334

0-344

Brazed.

Inch.
0-098
0-118
0-118
0-128

0-147
0-147
0-157
0-169

0-187
0-187
0-196

0-216
0-246

0-255
0-275

0-305
0-324

0-3.34
0-354

0-364

230 lb. per
square inch.

Drawn.

Inch.
0-088
0-108
0-118
0-128

0-147
0-157
0-169
0-177

0-196
0-206
0-216

0-236
0-265

0-285
0-305

0-344
0-364

0-383
0-403

0-423

Brazed.

Inch.

0-108

0-127

0-137

0-147

0-169
0-177
0-187
0-196

0-216
0-226
0-236

0-255
0-285

0-305
0-324

0-364
0-383

0-403
0-423

0-442

§ 262. Feed Pipe Bends. — Cast bends, tee-pieces, &c., for the
feed pipes are, where economy has to be considered, made of cast iron ;
and in warships and other vessels, where strength and lightness are
especially necessar}', they are made of bronze. They must be so con-

444 marinp: knginks and boilers.

Fig. 415.

PIPES AND CONNECTIONS. 445

structcd that the current of water in the branch pipes is disturbed as
little as possible, and should be shaped as shown in Fig. 412. Globe
tee-pieces (Fig. 413) are also sometimes used, but when fitted in pipes
running in a straight line they cause a greater loss of pressure than
pipes branching off at less than a right angle. Where there are several
boilers, the pressure of water in the feed pipes in close proximity to
the pump is naturally greater than in those further removed, but the
effect of this can be minimised by the adoption of good arrangements
of pipework, and by proper regulation of the feed valves.

g 263, Feed-water Heaters.— A distinction is made between —

1. Surface feed heaters ; and

2. Jet feed heaters.

1. In surface feed heaters the hot steam is either led through a
system of tubes which are surrounded by the feed-water, or the reverse

arrangement is adopted. The condensed steam is led either into the
hot well or into the feed pump suction. The Lundkvist beater (Fig.
416) is fitted either into the air-pump discharge or the feed-pump
suction. The feed-water enters at a, and is discharged at b; the hot
steam (exhaust from auxiliary engines) enters at d, and the condensed
water passes out at C. <> is the blow-off cock.

The Pape & Henncberg heater (Fig. 417) is generally placed in the
feed-pump delivery. The feed-water enters at c and passes out at i) ; the

446 MARINE ENGINES AND BOILERS.

live steam from the boiler enters at a, passes down through the coil of
pipes, and is discharged as water at b. Air and oil may be blown oflf
at E. The heaters may be cleaned by boiling out with soda.

2. Fig. 415 shows an ordinary Weir jet feed heater. The feed-
water enters through valve d, and mixes with the exhaust from the
auxiliary engines, &c., admitted through valve b. The heated feed-
water collects in the lower part of the heater, and passes from thence
to the steam pump, the speed of which is regulated by tfieans of float e,
which controls the cock f supplying steam to the pump. The libe-
rated air can escape at k into the atmosphere, or be led off into the
condenser.

Remarks, — There is no theoretical advantage in heating the feed-
water, unless the exhaust from the auxiliary engines, evaporators, &c.,
is used for the purpose. If live steam is used to heat it, no economy
from a theoretical point of view can be effected

SECTION V.
BILGE PIPES, BALLAST PIPES, CIRCULATING PIPES,

§ 264. Bilge Pipes are the pipes which serve to draw off the water
collecting in different compartments of the ship, and to discharge it into
the sea. They are generally made of lead, often of cast iron, but the
small pipes connecting them with the suction or mud boxes are usually
of copper. The connections at the ends of the different pipes are as
a rule made with copper rings soldered on to the pipes, with loose
wrought-iron flanges behind them. (See Fig. 383.) The thickness of

the lead pipes (2 to 8 inches in diameter) is about ^ = tt^ ^o tk-

The bilge pipes run under the floor-plates, and those passing throogh
the cargo spaces must be protected from injury by strong wooden sleeves.
They should not be carried through the coal bunkers, if it can pos-
sibly be avoided ; but if no other arrangement is feasible, they should
be strongly protected with wood. The lead or galvanised-iron strainers
or mud boxes must be so placed that they can be easily examined and
cleaned. They are frequently made in two halves, that they may be
more conveniently taken asunder. The short piece of pipe attached to
the suction-rose is made of galvanised iron.

Arrangement of Bilge Piping, — According to the regulations of the
German Lloyd's, ships having engines of less than 100 i.h.p., and
according to Lloyd's less than 70 i.h.p., must be provided with at
least one bilge pump; ships with larger engines must have at least
two such pumps, one of which is to be so arranged that it can
draw from all the watertight compartments. An exception may be
made in the case of the comparatively small watertight compartments
in the extreme bows of the ship, for which it is only necessary to provide
a small hand pump worked from the upper deck. Not only the engine
bilge pump and the steam pump, but also the circulating pump, must be
so arranged, that if necessary it can suck from the engine-room bilge
through a non-return valve. The diameter of this suction pipe should
be about 0*66 to 0*75 that of the circulating pump suction. In ships
without a double bottom one suction amidships is sufficient in each
compartment ; but if the ship's bottom is fairly flat, it is desirable to have

448 MARINE enginp:s and boilers.

a suction on each side of the ship, so that the bilge may still be pumped
out if the ship should have a list to one side. There must of course
be some arrangement for shutting off either of these suctions. Double-
bottomed ships must have a bilge suction in each double bottom.

If there is a so-called "well," one suction amidships will suffice;
but if the bottom of the ship is very flat, there must be one suction
on each side. A separate suction pipe leads from each outlet to a
common suction valve chest in the boiler or engine room, from which
the bilge pumps draw. This valve chest should have non-return valves
for each suction, so that if one watertight compartment is flooded, no
water may find its way to another compartment through the bilge
suction pipe. Each valve must be fitted with a name-plate, showing at
a glance to which watertight compartment its suction pipe leads. The
valves are often fitted so that they may be worked from the upper deck.
The connection of the bilge suction valve chests to the bilge pump is
generally through a three-way cock, the centre opening of which is
connected to the suction pipe of the pump, the opening at one side
to the bilge suction valve chest, and that at the other to a second valve
chest, to which the suction pipes from the sea, ballast tank, boiler,
condenser, &c., are connected. By this arrangement no water can pass
from the sea, &c., to the bilge valve chests and thence to the interior of
the ship. To prevent any dirt getting under the valves, cast-iron mud
boxes or traps are generally fitted in each suction pipe, or between the
suction valve chests and the pumps, in which the dirt is retained.
These mud hdkcs must be easily accessible, and always above the
floor-plates. The internal diameter of each bilge suction pipe should
not be less than 2 inches in small ships, and as much as 4 inches in
large ships.

55 265. Ballast Pipes. — These serve to fill and empty the various
ballast tanks. All the tanks of the double bottom, and also the ex-
treme forward compartments, if they are used as ballast tanks, must be
provided with at least one suction pipe leading to the ballast pump.
For the extreme forward compartments, as well as for the tanks of tbe
double bottom directly connected to each other, one suction pipe in the
centre of each space is sufficient ; but m larger ships, on account of the
relatively flat bottom, it is necessary in the remaining tanks to have one
suction pipe in the middle and one at each side, so that the tanks may
be pumped dry when the ship has a list to either side. With this latter
arrangement the suction pipes drawing from the middle of the ship are
generally made larger than those drawing from the sides ; the first is often
called the mam ballast pipings and the others the auxiliary ballast piping.
The diameter of the latter is about 0*6 to 0-7 the diameter of the former.

PIPES AND CONNECTIONS. 449

These various pipes are united in suction valve chests in the engine
and boiler rooms, which are connected by a single pipe to the adjacent
ballast pump. The pipe from the valve chest to the pump may have a
sectional area from two to three times greater than that of any of the
separate suction pipes. Mud boxes are not necessary here, as in the
case of the bilge suction pipes. The valves must be attached to their
respective spindles, as they have to be kept open while the tanks are
being filled. Each valve should have a name-plate showing where the
suction pipe connected to it leads. The size of each suction pipe is
generally so calculated, that, for a mean water speed of 200 to 350
feet per minute, the time required to empty all the tanks is from 4 to 5
hours, and from 2 to 2^ hours for a single tank. From this the size
and duty of the ballast pump can be determi<ned.

^ 266. Diameter and Thickness of Ballast Pipes.— The

ballast pipes are generally of galvanised wrought iron or of cast iron.
Wrought-iron pipes are made in the following thicknesses : —

For internal diameter —

From 2 to 4 inches they are about 0*15 inch thick.
„ 4 „ 6 „ „ 017 to 0-20 inch thick.

„ 6 „ 8 „ „ 0-20 „ 0-23

The thickness of cast-iron pipes is about 0*4 to 0*5 inch.

The connecting pipes and bends and the bulkhead connections are
generally of copper, about 0-12 to 0*15 inch thick. Care must be taken
when fitting in the ballast pipes that the different branches have as
few bends as possible, and that they are not carried through the coal
bunkers, if it can be avoided. If they must pass through the bunkers
or the hold, they should be protected from injury by strong removable
wooden coverings. Sometimes these pipes are not run above, but
through the double bottoms. This arrangement has the advantage that
useful space in the bunkers and hold is not taken up by pipes and their
coverings. On the other hand the pipes are not easily accessible. The
ballast tanks are filled by means of a separate sea-valve, like that used
for the circulating pump. I'he size of valve is proportional to the
diameter of the ballast pump suction. As these pipes are generally of
considerable length, it is best to have an air vessel near the pump, to
avoid as far as possible shocks in the suction pipes.

The discharge pipe of the ballast pump is generally short, and is
connected to a separate discharge valve on the side of the ship. This
can usually be made to discharge above the water-line ; and where this
is possible, it is so fitted. A centrifugal or a plunger pump may be
used as ballast pump. If the arrangements are such that the ballast
pump is used for filling the tanks completely, a centrifugal pump

2 F

450 MARINE ENGINES AND BOILERS.

should be used, because if a plunger pump be employed, and the
pumping is carelessly done, the ballast tanks might easily be burst or
strained. If the ballast tanks are filled by means of plunger pumps,
they must be provided with stand pipes, reaching to the upper deck,
and open at the top.

§ 267. Circulating Water Pipes. — In engines of any consider-
able power, in order to prevent the bearings becoming overheated,
should they tend to work hot, special pipes are provided for the proper
circulation of cooling water. This cooling is effected by means of sea-
water, applied either directly^ by allowing the water to flow actually over
the part of the engine affected, or indirectly by allowing the water to
circulate through hollow spaces in contact with the surfaces to be cooled.
Cranks, eccentrics, crossheads, plummer blocks, thrust blocks, stern
tubes, &c., are ^''directly cooled^^ ; while the crosshead guides are almost
always, and the thrust collars, tunnel bearings, and shaft bearings are
sometimes " indirectly cooled^ The water for the direct cooling is only
applied when necessary, while that for the indirect cooling is always in
circulation. The latter generally runs directly from the parts that have
been cooled into a collecting pipe leading to the suction chamber of
the circulating pump, sanitary pump, or some other sea-water suction
chamber ; the former passes into the bilge, and is pumped overboard
by the bilge pump. The water is drawn from the delivery side of
the circulating or sanitary pumps, or straight from the sea. The water
circulation is conveyed in pipes of from f to J inch internal diameter.
When the cooling water is directly applied, each pipe is so arranged that
it can be separately shut off. Joints are provided so that water can be
sprayed on to any part of the engine that is working too hot. Cocks and
hose-connections are fixed in suitable positions, so that in case of need
large volumes of water may be directed on to any part which requires
to be rapidly cooled. In very large installations portions of the cooling
water systems are sometimes kept separate, there being different pipes
for instance for the thrust blocks and shaftings ; or the direct and the
indirect cooling water systems may be quite apart from one another.

The diameter of the cooling pipes for both direct and indirect
cooling of the thrust bearings, shafting, and other parts of the main
engine must be —

Internal Diameter.
For engines from about 10,000 to 15,000 i.h.p. 3i to 4| inches.

7,000 „ 10,000 „ 3" „ 3i „

4,000 „ 7,000 „ 21 „ 3
1,000 „ 4,000 „ 2 „ 2\ „

2

,, up to about 1,000 „ 1

PART V.

STEAM BOILERS.

SECTION I.
FIRING AND THE GENERATION OF STEAM.

% 268. General Remarks. — The function of a boiler is the
conversion of the energy of a combustible into available heat. Accord-
ing to the Helmholtz law of the conservation of energy, the chemical
composition of the coal determines the amount of heat which, under
the most favourable theoretical conditions, can be obtained from it.
The whole process of combustion is, however, so complicated, that the
actual result attained varies greatly according to circumstances, and it
is therefore necessary to study how it can be best carried on, the prac-
tical means to be adopted, and the effect which they may be expected
to produce.

The subject divides itself naturally under two heads : —

(a) Generation of heat on the grate. Process of combustion,

(b) Transmission of the heat obtained to the water in the boiler :
Generation of steam,

g 269. Process of Combustion. — In general this may be regarded
as the combination of a substance with the oxygen of the air, producing
the phenomenon of fire. Chemical combinations are formed according
to definite proportions of weight ; therefore the quantity of air required
for the combustion of a given quantity of combustible depends upon the
composition of both (Table No. 54:). A chemical combination is nearly
always accompanied by the evolution of heat. It depends quantitively
upon the character of the process (as shown by the products of
combustion). Again, the quantity of heat obtained from the com-
plete combustion of equal parts by weight of the same substances is
always the same. It is thus possible to draw up tables from which
we can calculate approximately the heat developed by the combustion
of every known combustible (Table No. 54). From this table formulae
can be worked out for the quantity of air theoretically required, and the
heat generated per unit weight of substance burnt (heating value).
If a combustible is composed of certain percentages by weight of carbon
(C), hydrogen (H), and oxygen (O), then the

Heating value Q= 14,500 (C + 4-28H) thermal units per pound.

454

MARINE ENGINES AND BOILERS.

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STEAM BOILERS.

455

and the quantity of air by weight theoretically required to be added to
it for complete combustion would be

L = 0-116 (C + 3H-0-4O) lb. per pound, or by volume,
L= 1-52 (C + 8H- 0-4O) cubic feet per pound at 62' Fahr.
and at 30 inches mercury.

More accurate values of Q can only be obtained from calorimetric
determinations. (See Part VI.)

For summary of the heating value of various kinds of coal (mean of
numerous measurements), see below, Table No. 55.

Table No. 55.
Heating Values of Various Kinds of Coal^ &*c.

Kind of Coal.

>t

}»

»i

>i

»>

»

>»

ii

»>

Welsh

Newcastle -

Scotch

Westphalian hard coal

Silesian

Saxon

Saar

Illinois

Missouri

Pennsylvanian

Tasmanian

Chilian

Japanese

Soft brown coal (lignite)

Hard „ . . . .

Dry wood

Refined petroleum - - . -
Distilled petroleum residuum (Mazout)

Heating Value, /.^., British
Thermal Units generated
per pound when Combustion
IS complete. *

B.T.U.
14,850
14,800
14,100
13,500 to 14,040

12,060
10,980
12,960
9,000
9,900
10,800

12,420
11,340
13,680
10,800
11,700
13,320

about 10,800
11,700
12,600
7,200
9,900
3,600 to 5,400
18,000
18,000

>»

19

19

>1

S^ 270. Incomplete Combustion. — In practice these values are

• By heat unit is meant the quantity of heal required to raise the temperature of
1 lb. of water 1 degree Fahr.

456 MARINE ENGINES AND BOILERS.

never attained, as combustion is always more or less incomplete. The
degree of completeness achieved depends in the main on three factors : —

1. The quantity of oxygen introduced with the air for combustion.

2. The temperature in the combustion chamber.

3. The intimate mixing of the fuel and oxygen at the place where
combustion takes place.

To appreciate more fully the effect of these factors, it will be
useful to illustrate the process of combustion by citing a typical
example. For this purpose let us select the ordinary cylindrical boiler
with a horizontal grate, furnace, fire box, and boiler tubes ; there will be
no difficulty in drawing deductions for other types of boilers.

The process of combustion is generally as follows : — The air fur
combustion, entering below the grate through the ash-pit, meets the
lowest layers of coal, w^hich are at a white heat, and carbonic acid gas is
generated. Complete combustion takes place here, because a large
excess of air is always present. The higher the mixture of nitrogen,
carbonic acid gas, and oxygen, rises through the coal on the grate, the less
becomes the excess of air. Combustion is retarded, in consequence of
the fall in temperature produced as the upper layers of coal are reached.
A mixture of gaseous hydrocarbons given off from the fresh coal (such
as methane, ethylene, see Table No. 54), carbonic oxide, carbonic acid,
oxygen, nitrogen, &c., is therefore formed, and appears as flame during
combustion. We have thus to distinguish between two stages of
combustion, viz., that of the solid carbon in the lower strata, and
that of the gaseous constituents above the layer of fuel. The poorer
or the richer a coal is in hydrocarbons, the more or the less im-
portant will be the second stage, with regard to the total process
of combustion, and especially with regard to the production of
smoke. The opinion is widely entertained that smoke consists of small
particles of unconsumed carbon carried off from the grate. But a close
examination of the nature of the soot in the smoke shows that soot
is a residual product of combustion itself, and is present in largest
quantities where the coal consumed is rich in hydrocarbons.

§ 271. Losses by Excess of Air. — The theoretical quantities
of air given in § 269 as requisite for the complete combustion of the
coal are not sufficient in practice, because of the great difference in
volume between the air and the coal (with the theoretical quantity of
air this ratio is about 12,000 to 1), and also because of the impossibility
of exercising a rigid control over the movement of the air through the
fire. Combustion must therefore be carried out with an excess of air,
but this should be restricted within the narrowest limits, because the

STEAM BOILERS.

457

heat carried off uselessly up the chimney increases with the quantity
of gases of combustion : — these escape at a temperature of about 570°
Fahr., and thus the heat efficiency of the boiler is lowered. (Compare
Table No. 56.) In practice, when calculating the dimensions of the
grate and flues, an excess of air of 100 "/^ is generally allowed for, and
the working of the fire is regulated by dampers and ashpit doors.

Table No. 56.

Heat Losses in the Gases of Combustion.

(The loss is expressed in percentages of the heat generated on the grate.)

3 O g

Percentage of Excess of Air.

Ml

I)cg. Fahr.

0

25

50

75
Per cent.

100

125

150

Per cent.

Per cent.

Per cent.

Per cent.

Per cent.

Per cent.

480

8

10

12

14

16

18

20

570

10

12-5

15

17-5

20

22-5

25

660

11-5

14-4

17-3

20-2

231

26

28-9

750

14-2

17-8

21-3

24-9

28-4

32

35-5

The above figures apply only on the assumption that pure carbon is
burnt on the grate. In practice the losses resulting from too great an
excess of air are higher.

§ 272. Grate Area. — This is the starting point for designing
boilers, as a square foot of grate area is taken as a basis of all calcu-
lations. To determine the actual dimensions of the total grate area
for a given boiler or set of boilers, it is necessary to know from experi-
mental data the horse-power generated per square foot of luating surface^
and the ratio of heating surface to grate area. As the dimensions of
each grate are determined by convenience in stoking, the number of
grates required is easily arrived at. According to the quantity of coal
bumty whether more or less^ per square foot of grate area, combus-
tion is said to be either normal or more or less forced. The values
given in Table No. 57 are the result of extensive practical experience.
For more precise values, for different types of boilers, see Tables Nos.
63 to 67 ; for cylindrical boilers, see Table No. 62.

458

MARINE ENGINES AND BOILERS.

Table No. 57.

Pounds of Coal burnt per square foot of Grate^ Quantity of Air required,
and Pressure of Air or Vacuum for the three Methods of Firing in most
general use.

Draught.

Coal burnt per

square foot of

(irate Area,

pounds per

hour.

Quantity of Air reauired per square foot

Urate per hour, taking 100 % £xcess of

Air as a Bans.

1

PrtssQicor

Vaonnm m

Pounds per hoar

per squxu% foot

Orate.

Cubic feet per sauare

foot Grate per Konr

at 62" F. and at 90

inches Mercury.

inches of
Water.

Natural draught
Induced draught -
Forced / Moderate
draught! High

15 to 25
22„30
22„37
37 „ 50

368 to 543
543,, 717
543,, 880
880 ,,1,230

4,900 to 7,200

7,200,, 9,600

7,200 „ 11,800

11,800 „ 16,400

019 to 0-59
0-39 „ lo8 ,
0-39 „ 0-79 1
0-79 „ 2-75 '

Remarks. — 1. The inches of water give the difference in pressure
between the ash-pit and the external air, if they show pressure, and
between the external air and the combustion chamber, if there is a
vacuum (induced draught).

2. In torpedo-boats and destroyers, at full speed, the consumption
of fuel may reach as much as 80 lb. per square foot of grate per hour.

The choice of the system of firing to be adopted, whether natural
or artificial (induced and forced draught), is determined by the desired
rate of combustion of coal on the grate, or evaporation in the boiler.
For slow-going freight and passenger steamers natural draught is
preferable, because simpler. In fast steamers all three kinds are in
use, but only with moderate rates of forcing. War vessels employ
almost exclusively forced draught, and the coal burnt per square foot
of grate varies according to circumstances.

§ 273. Natural Draught is produced by the difference in specific
weight between the hot gases of combustion passing up the chimney
and the outer air. The amount of excess of pressure of the outer air
below the grate is determined by the height of the funnel above the
grate, mean temperature of the escaping gases of combustion, and their
speed through the flues, and the temperature of the external air. Let

h be the head produced by the chimney draught at the grate,
in inches of water -,

STEAM BOILERS. 459

H be the height of the funnel above the grate in feet ;

/g „ mean temperature of the chimney gases, degrees Fahr. ;

/j „ temperature of the outer air, degrees Fahr. ; then

>l= H (l-(L±gJV.12inch.
66-6 \ /2 + 46I/

ITie degree of pressure required to maintain the speed of the gases
against the resistance in the flues is brought into the above formula by
means of the constant *12 to be subtracted, which corresponds to a
mean value, assuming the sectional area of the flues to agree with those
given in Table No. 58.

Table No. 58.

Afean Proportional Dimensions for Calculating the Sectional Area of
Flues of Ship's Boilers with any kind of Draught.

The following sectional areas per square foot of grate may be
used: —

In the ash-pit - - 15 to '20 square feet.
Over the fire bridge - -15 to '18 „

Over the fire bridge - -15 to '18

In the vertical portion -30 to "40 square feet (as large as construc-

of the fire box tive conditions will allow).

Through the tubes - -16 to -20 square feet.

In the smoke box - '40 to -50 „

In the uptake - - '16 to -25 „

In the funnel - - -16 to '24 „

N,B, — The cylindrical boiler has been taken as the basis for this
table; for boilers of other types the required sectional area can be
determined without difficulty, by substituting corresponding values, the
lower values for natural draught, the higher for forced draught with
heated air. In calculating the sectional area for the passage of the
flue gases through the nest of tubes in water-tube boilers, the area
allowed is about "15 to '2 square foot for every square foot of grate
surface.

Example, — Cargo boat in northern latitudes —

H = 82 feet; /i = 68' Fahr. (max.); /2 = 392' Fahr.

h^^^(\^ ^\ - -12 inch = 0-47 inch - 0-12 inch = 0-35 inch.
66-6 \ 853/

In this set of boilers the maximum amount of coal that could be
burnt was 18-5 lb. per square foot of grate per hour. Assuming that if

460 MARINE ENGINES AND BOILERS.

the combustion were more forced, the gases of combustion would escape
at a higher temperature, say a mean of 572** Fahr., then

^ = 6-1^ 0 -iS) -'■'' = '■*' '"^''
so that the maximum quantity of coal burnt per square foot of grate
area per hour might amount to 22*5 lb. If the same steamer were
running in the Tropics, the temperature of the air might be 104** Fahr.,
and the mean exit temperature of the gases of combustion, 392° or
572° Fahr., then h would be

6-110 -D-^-i^'-'^'^-^^-^' -

The difference is comparatively small.

§ 274. Artificial Draug^ht. — If the quantity of coal burnt per
square foot of grate area per hour exceeds 20 lb., the quantity of air
obtained with natural draught is frequently insufficient, and a suppU-
meniary supply must be artificicdly provided. According to the method
of admitting this air, whether it is drawn into or forced through the
fire, the system is known as induced draught, or forced draught The
principal types of the different systems are : — (1) Induced draught— EWis
and Eaves; (2) Forced draught {a) with closed stokehold, (^)with closed
ashpit — Howden and other systems. For the usual arrangement of
these systems of firing, see Figs. 481 to 489. The dimensions of the
sectional areas of the various flues given in Table No. 58 can be used
provided that the higher values therein stated be taken.

§ 275. Centrifugal Fans. — The most important adjuncts to these
forced draught systems are the centrifugal fans, which are almost
exclusively used in ships to furnish the supply of air required. For
calculating the dimensions of the fans, the quantity of air to be de-
livered is taken as a basis, as also the pressures of air (or vacuum)
required, and the power necessary to produce them. It should be
noted that the pressure produced by the fan must be added to that due
to the natural draught.

§ 276. The Dimensions of the Fans may be determined by
means of the subjoined calculations where tbe following factors are
assumed : — The overall diameter of the vane wheels of a fan is from
one and a half times to twice the internal diameter.* The circum-

* In practice the values for the external diameters are r, = 2 feet 6 inches to
4 feet for forced draught ; ra = 3 to 5 feet for induced draught.

STEAM BOILERS. 461

ferential sectional area at any point of the fan measured parallel to the
axis, and at any distance from it, is made the same as the cross section
of the intake, and kept a constant throughout the fan, and according to
its size it is from 0*6 to 0*75 the sectional area of the flue in which
the fan is to be placed. Therefore in proportion to the grate area,
the radial cross section would be (see Table No. 58) : —

1. For induced draughty from '12 to '185 square feet per square foot
of grate area.

2. Y ox forced draughty from '09 to '12 square foot per square foot
of grate area.

By the use of these figures the velocity of the air through the fan
may be determined for every given case.

§ 277. Example I. — Boiler with closed stokehold. Given a mini-
mum of 22-5 lb. and a maximum of 37 lb. of coal per square foot of grate
area i)er hour. According to Table No. 57 we shall therefore require
per hour from 7,200 to 11,800 cubic feet of air at 62' Fahr. and 30
inches mercury. The cross section of the fan (which in this case is
supposed to deliver direct to the ash-pit) we will take according to the
above data at 0-1 square foot per square foot of grate area. We thus
get a velocity of air through the fan of

CA ITa a 1=2^ ^"^^^ P^^ second, and ' = 32-8 feet per

60x60x0*1 60x60x0*1

second.

If the system of draught under consideration were induced draught,
it would be necessary to note that, on account of the higher temperature
of the gases, the volume of air and also its speed would increase in
direct proportion to the absolute temperatures. (Compare also Table
No. 58.)

§ 278. Example II. — Given a boiler with induced draught burning
28*7 lb. of coal per square foot of grate area per hour. According to
the above rule the cross section of the fan wheels is 0*15 square foot per
square foot of grate area. The quantity of air required at 62' Fahr.
and 30 inches mercury is, according to Table No. 57, 9,000 cubic
feet per square foot of grate area per hour. At 752' Fahr. that is

9,000 X -— — — - = about 21,000 cubic feet per square foot grate.

(The reduction in pressure may be neglected as the calculation is based
on the volume.) Thus the radial velocity of the gases through the fan

will be jT/x' /,/^ X ^TTk = (about) 39 feet per second.
60 X 60 0-15 ^ ^

462

MARINE ENGINES AND BOILERS.

Fig. 418.

Fig. 419.

Fig. 420.

§ 279. The Form of the

Vanes (or blades) to ensure the
delivery of a given quantity of air is
of great importance, because it de-
termines the number of revolutions,
the air pressure obtainable, the
maximum quantity of air that can
be delivered with a given breadth
of vane, and the work expended to
produce it. As the fans in every
system are called upon to run at
very varying speeds, an entrance of
the air free from churning at all
speeds cannot be ensured. From
practical considerations the angle
/?! (Figs. 418 to 420) should not
be taken too small, /.^., not less
than 30^ The external angle of
the blade may vary from 60" to

The actual shape of the blade is
preferably designed with arcs of
circles or straight lines, care being
taken that no sudden changes of
direction occur. The shape having
been determined, the pressure to
be attained with a given number
of revolutions, and the expenditure
of work required to produce it, are
also determined, the breadth of the
blade being ascertained by assuming
a given velocity of air. By pressure
is to be understood the difference
in inches of water between the
delivery and suction sides of the
fan, it being a matter of indiffer-
ence whether the latter forces the
air through or sucks it in. The
difference in the weight of the air
in either case is too small to'ap-
preciably affect the results. Let
r, denote the radius of the fan
disc in feet.

}}

9i

STEAM BOILERS. 463

n denote the number of revolutions of the fan per minute.

difference in pressure between the suction and de-
livery sides in inches of water,
radial velocity of the air through the fan (a constant
varying with the type of construction), in feet per
second.

V „ velocity of the fan disc at its outer periphery in feet

per second.

)Sa M external blade angle. (See Figs. 418 to 420.)

T „ absolute temperature in degrees Fahr. of the air

passing through the fan.

/ „ circumferential sectional area in square feet of the

fan disc per square foot of grate area measured
concentrically to the axis, and at right angles to the
radius (see § 276), constant throughout the disc.

A „ work in h.p. expended in driving the fan per square

foot of grate area.

Then A^ = — (v - u cot y3,)2.
A = '23^ X [•093(z; - u cot /3,Y + (•305//)2].

Example /. — Boiler with induced draught but without air being pre-
viously heated. Total grate area 484 square feet ; 25*6 lb. of coal to
be burnt per square foot of grate area per hour. From Table No. 57
there are required per square foot of grate area per hour, 8,500 cubic
feet of air at 62** Fahr. As the fans are placed in the uptake, the
quantity of air at 752' Fahr. to be delivered per hour is —

8,500-— — __ = about 20,000 cubic feet per square foot grate.

The cross section (/) according to the data on page 461 is 0*15
square foot per square foot of grate. Then —

« = ^R — ^ — TT^^ = 37 feet per second.
60 X 60 X 0-15 ^

To keep the engines and fans as light as possible, a high value
for V about = 82 feet per second is taken, always keeping in view the
strength of the material.*

* For parts of machines rotating at high speeds, such as, for instance, the cir-
cumference of fly-wheels, a simple calculation will prove that the stress upon the
material due to centrifugal force depends entirely on the circumferential velocity, and
is not affected by the radius or the num1)er of revolutions.

464 MARINE ENGINES AND BOILERS.

Then according to the above formula, with /J^sSO'

h ^ -_Ji_-- X (82)2 = about 055 inch.
* 7o2 + 461 ^ '

This is the difference in pressure between the suction and delivery
of the fan. To this must be added the vacuum produced by the
natural draught (see page 458), so that the difference in pressure between
the suction side of the fan and the outer air is about 0*55 inch + 0*35
inch = 0*9 inch. Above the fires the vacuum will naturally be some-
what less, owing to the resistance offered by the flues between the
combustion chamber and the fan to the passage of the air through them.
The work required will be —

A = '2|A:1^-^??[093(82)2 + (-305 x 37)2] = -095 h.p.

per square foot of grate area. Total expenditure of work for 484 square
feet of grate area = 46*2 h.p. This is the work required at the shaft of
the fan. Suppose the diameter of the fan be taken (according to the
space available) at 8 feet. Then —

60x82 ,^. , . . ,

n = — ; = 19o revolutions per mmute.

8x?r ^

The number of fans is generally determined by the grouping of the
boilers and arrangement of the uptakes. Assuming that 4 were to be
used, the external width per fan would be —

b= ^-^ =0-73 foot = about 8 J inches.

4x8xir *

With these measurements the fan must be constructed so that it is
not contracted in any part, or if so it must be made correspondingly
wider. If fans of these dimensions were used with a boiler fitted with
the Ellis and Eaves system of forced draught, in which the combustion
gases are cooled by passing through an air-heater before they enter the
fan, T would be 853" instead of 1,213' Fahr. absolute, and

//, = ^ X 822 ^ 0-78 inch.
^ 853

Total vacuum = 0-78 inch + 0*35 inch = M3 inch.

Then « = 25*7 feet per second, and

A = :2?Ji±^Ji^[.093(82)2 + ('305 x 25*7)2]= -075 h.p.
Total work done on the fan shaft, 484 x -075 = 36*2 h.p.

STEAM BOILERS. 465

Example II. — Boiler with closed stokehold; total grate area, 970
square feet ; maximum quantity of coal to be burnt per square foot of
grate area per hour, 35-8 lb. According to Table No. 57, the quantity
of air required per square foot of grate area per hour is 11,500 cubic feet.
Cross section, from page 461, 0*1 square foot per square foot of grate
area.

^ ^ fTi — ar\ — c7\ = ^^ ^^^^ P^r second.
0*1 X 60 X 60

Assuming v, as before, at 82 feet per second, I3^ = 90\ and t at
(es** Fahr. + 461). Then—

/t,= 9-^x82^= 1-26 inch.
* 529

Total excess of pressure between the stokehold and combustion
chamber =1-26 inch + 0-35 inch = 1 61 inch.

A = '^ -^ 9^'^-iil-[-093(82)2 + (-305 x 32)2]

= about '113 H.p. per square foot grate.

Total work expended on the fans is therefore about 110 h.p.
There are assumed to be five stokeholds, each having four fans.
Taking the diameter of the fans as 5-9 feet (see page 460), then

60x82 „^. , .

« = . ^ = 2do revolutions per mmute.

0-9 X TT ^

970 X O'l
The circumferential breadth should be ^= - / - -- — = 0-262 foot,

20 X 5-9 X TT

if there are no contractions or losses from other sources. If any such

have to be provided for, the fans must be made correspondingly wider,

and the power of the fan engine increased. These contractions may

occur either in the fan or in the flues themselves. The most helpful

rule in practice is to calculate the fans and engines for a delivery of

about double the quantity of air actually required, and to adjust the

right number of revolutions by the stop valve, and the cross section of

the air passages by suitable dampers while the fan is at work.

If the fans are driven, not by steam engines, but by electric motors,

which must be run at a given number of revolutions per minute, the

output of the fan must be regulated by the dampers. For this class of

motor, fans with bent-back blades (i8,<90'') are not suitable, because if

the quantity of air increases, the pressure diminishes when the speed

remains constant. Fans should be used with blades bent somewhat

forward (j8,>90°), in which the pressure increases with the quantity

of air admitted, when the revolutions remain constant. With blades

which are radial throughout their length the pressure is constant for

2g

466 MARINE ENGINES AND BOILERS.

any given quantity of air. The most important source of loss of air with
closed stokeholds is the leakage of air through openings and cracks.
This question, however, cannot be treated here; it is best to avoid
these causes of loss by care in details of construction.

S 280. The Number of Blades varies, according to practical con-
ditions, from ten to sixteen, but frequently only half of them are carried
out from rj to r, (Fig. 420), while the other half are fitted only from the
outer periphery down to half the radial depth of the wheel. The outer
angle of the blade P^ has a great effect upon the number of revolutions.

As for the same ^j,

{v-u cos P^ = constant, and must be = about t/j,

then for example —

Withj8= 60' v = v^ + 0'hTiu
„ = 90° v^v^
„ =120° P = ri-0-577«

Therefore in our example where ^i = ^2 and m = 37 (feet per second),
with /3 = 60*' and 120° respectively, the corresponding values would be
61*4 feet and 102-6 feet per second, which would cause a considerable
difference in the number of revolutions.

§ 281. A High Temperature in the Combustion Chamber
is an essential condition for obtaining complete combustion, because the
ignition of the combustible constituents of the coal requires certain de
finite temperatures, and the time available for the process of combustion
is very short. As the air on entering must first be heated before it is in
a position to assist combustion, the intensity of combustion depends on
the excess of air. It is also clear that in any case the temperature of
combustion to be attained depends directly on the excess of air, because
the whole of the heat generated by combustion passes first into the pro-
ducts of combustion, which in their turn convey it by radiation to the sur-
rounding substances. Therefore the admission of only the theoretical
quantity of air required for complete combustion theoretically yields the
highest temperatures of combustion. (Compare Table No. 54.)

In practice, having regard to the material of the furnace, &c.,
excessively high temperatures in the combustion chamber should be
avoided ; nor are they practically attainable on account of the excess of
air required for other purposes (see § 271). Nevertheless, when work-
ing the furnaces, great care must be taken that (by bare patches occurring
in the grate, or by the fire doors being left open too long, &c) the excess
of air does not become so considerable that the flame is extinguished by

STEAM BOILERS. 467

lowering its temperature too far. In furnaces with artificial or forced
draught, in which the speed of the air through the grate is comparatively
high, it is desirable, in order to assist complete combustion, to warm the
air beforehand, or to inject heated air into the space above the grate
(secondary admission of air for combustion, as used by Howden,
Belleville, and others).

§ 282. Mixing of the Gases of Combustion. — The complete

and thorough mixing of the oxygen of the air and the combustible
constituents of the coal is very important in order properly to carry
out combustion. In its first stage (see § 270) this mixing is easily
effected, because the coal remains stationary, and the air passes over
it ; but in the second stage it is much more difficult to effect, because
the constituents are now gaseous, at a relatively high temperature, and
therefore having a high specific volume, move at equal velocity side by
side through the flues. It is only possible to mix them if there are
frequent changes of direction, and if eddies are formed, as for instance
at the fire bridges, or by the injection of secondary air or steam over
the grate.

§ 283. The Useful Heat of Combustion is that heat contained in
the hot gases which can be transferred to the water in the boiler. The
losses here are as follows : — Heat radiated externally from the furnace, heat
remaining in the ashes, incomplete combustion, and heat carried off in the
chimney gases. External radiation from the furnace may be neglected,
as the air surrounding the boiler, which is afterwards drawn into the
furnace, absorbs the greater part of it. Losses in the burnt residuum
(ashes, &c.) seldom amount to more than 3 7o» i^ the grate be in
good condition and properly stoked. Losses arising from incomplete
combustion may, under certain circumstances, be very considerable.
The cause of them in furnaces of ships' boilers is not the want of air for
combustion, for there is always a large excess of air, but from the too great
reduction of the temperature of the flame due to keeping the fire doors
open too long, from putting on too large a quantity of coal at once and in
lumps of too large a size, from allowing the grate to become clogged,
&c.* If the fires be skilfully stoked, these losses may be kept within
comparatively narrow limits, so that the principal loss in the whole pro-
cess consists in the heat carried off by the gases. The amount of this
loss depends wholly on the excess of air, and may under certain con-

* In ships* boilers a third of the fires are cleaned at the beginning of each watch
{i.e., every four hours). Therefore at the beginning of each watch the evaporation
diminishes in a marked degree, and the speed of the engines may be reduced from
3 to 5 %.

468 MARINE ENGINES AND BOILERS.

ditions be considerable. (See Table No. 56.) Experiments to determine
it in any existing boiler can, in ships, only be carried out by chemical
analysis of the flue gases (see Part VI.), that is, by ascertaining the excess
of air from the percentage of oxygen contained in the exhaust gases.

§ 284. Generation of Steam. — The transmission of heat from the
gases of combustion to the contents of the boiler takes place by con-
duction. To effect this transmission economically through the dividing
walls, the latter should be as thin as possible, and their surface as great
as possible (heating surface). Strength and tightness are the first
practical considerations to be attained. The endeavour to have walls
as thin as possible and yet sufficiently solid has led to the construction
of water-tube boilers. The amount of heating surface is determined by
the final permissible temperature of the hot gases, having regard to the
utmost utilisation of the heat in them.

Purely theoretical considerations are of little value here. In the course
of years of practical experience certain data have been obtained for dif-
ferent types of boilers, which are classified in Table >Jo. 59. With the
same kind of flues and the same draught, merchant vessels show higher

, c heating surface h.s. h ^, , • • *i. i ^

values of ^ = = - than warships, as m the latter more

grate surface g.s. r

importance is attached to utilising the weight of metal to the utmost, and

no attempt is made to turn to best account the heat contained in the

gases of combustion when the engines are working at full speed.

g 285. The Efficient Transmission of Heat from the Gases
of Combustion to the Water requires the closest contact between
the gases and the walls of the boiler on the one hand, and between the
water and the boiler wall on the other. To obtain this it is necessary
that the heating surfaces should be kept clean inside and out (from soot,
ashes, salt, or other deposits). The process of transmission of heat is
also greatly assisted by sudden changes in direction and of sectional area
of the flues, and also by suitable arrangement of the fire bridges, &c,

§ 286. The Heat Transmitted to the Contents of the Boiler

has first to heat up and then to evaporate the water. The theoretical
quantity of heat available may be ascertained from Table No. XXII.

The values given in the different columns of Table No. XXII.
(page 694) are obtained from the theory of the mechanical equivalent
of heat, and require some explanation.

If water is heated in a closed space at constant pressure its tempera-
ture rises to boiling point, and the amount of heat required to raise 1 lb.
water through V Fahr, is a unit of heat or thermal unit. The heat

STEAM BOILERS.

469

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470 MARINE ENGINES AND BOILERS.

required per pound to heat the water from 32' Fahr. to boiling point is
sometimes called the water heat of steam (see column 3, Table No.
XXII.). While boiling in an open vessel the temp>erature remains
constant (temperature of boiling water). The steam generated is called
saturated as long as water is present. At the moment when the last
drop of water disappears the steam is dry saturated. According to
Regnault's experiments the temperature of saturated steam is a function
of the pressure only, and therefore in Table No. XXII. a definite
temperature is given for each definite pressure.

The heat required to convert 1 lb. of water into steam at the same
temperature is called the latent heat of evaporation or total latent htai
(column 4 + column 5). This consists of the so-called internal latent
heat and the external latent heat. The internal latent heat is used to
overcome the resistance of the molecular forces which oppose the change
of state of the molecules brought about by the formation of steam, while
the external latent heat represents the amount of heat which has to be
added to overcome the external pressure in pounds per square inch as
the volume (of water or steam) increases. Thus the so-called total heat
of I lb. of dry saturated steam consists of the water heat (column 3)
and the latent heat (column 4 + 5). It signifies the excess of heat
possessed by 1 lb. of dry saturated steam over 1 lb. water at 32' Fahr.
The characteristic of saturated steam is that if heat be withdrawn from
it, it does not produce a fall in temperature, but condensation. If dr)'
saturated steam is still further heated, the pressure remaining constant
it becomes superheated ; z.^., it attains a condition in which it is more
like air, because if heat be withdrawn from it, not condensation, but only
a fall in temperature of the steam, results.

S 287. The Formation of Steam in the Water produces
violent ebullition, due to the sudden increase in volume of the steam
generated. The eddies produced, and the escape of the steam from
the surface of the water in the shape of bubbles, are the reasons for
the formation of wet steam over the surface of the water. The water
mechanically mixed with the steam passes through the engine, but per-
forms hardly any useful work, and may easily give rise to water-hammer;
the generation of dry steam should therefore be aimed at. The means to
ensure this, apart from any fittings especially provided to dry the steam,
are to provide large steam spaces above the water level, to prevent violent
ebullition at the water surface by providing a large surface of water for
disengaging steam, and the use of pure water, the distribution of the
heat as evenly as possible over the total heating surface, together with a
proper circulation of water, and lastly a moderate evaporation per
square foot of heating surface (no undue forcing should be allowed).

STEAM BOILERS. 471

The dimensions of the steam space and the level of the surface of the
water are almost always fixed in ordinary cylindrical boilers by their con-
struction. In water-tube boilers both steam and water spaces are gene-
rally much reduced, because the adoption of these boilers is dictated by
the urgent desire to economise weight, and therefore they yield much
wetter steam than cylindrical boilers. As it is all important to avoid
scum in marine boilers in which the same water is used over and over
again, special fittings are necessary to remove all oil or greasy deposit
from the feed-water. To generate the steam efficiently the heat should
be transmitted as uniformly as possible over the whole heating surface.
It is a well-known fact that uniform transmission of heat forms an
important factor in the efficient generation of steam. (For circulating
apparatus see § 333.) It is necessary that the water evaporated from
the heating surface should be continuously replaced, and the best means
of effecting this is to ensure, by the aid of some auxiliary arrangement,
a regular flow of water to the heating surface. No certain data on this
point can, however, be obtained, as opinions as to the value of the different
methods vary greatly. (See Engineerings 1896, page 583.) In water-tube
boilers, which have headers to convey the water fed into the upper drum
down to the lower, the total sectional area of the headers is calculated
to give '032 square inch per square foot of heating surface. In cylin-
drical boilers the water rises in the neighbourhood of the combustion
chamber, and descends towards the front tube plate.

§ 288. Efficiency of Steam Production. — The measure of the
efficiency of the production of steam is the heat contained in the steam,
per unit of weight, compared to the heat contained in the feed-water
when it enters the boiler. To make the comparison uniform it is usual
to reduce the difference in the amounts of heat in the feed-water and
steam, to evaporation from and at 212" Fahr.

Example. — In a boiler 8 lb. steam are generated per pound of
coal from feed-water of 110** Fahr., and at a pressure of 140 lb. per
square inch absolute. Total heat in the steam at this pressure
from Table No. XXII. = 1,189 thermal units per pound. As the tem-
perature of the feed-water is 110** Fahr. {1,189 -(110-32)} = 1,111
thermal units have been added to each pound of water in passing
through the boiler. From the same table, to generate steam at 212"
from water at 212", 965 thermal units are required; so with 1,108

thermal units ' ^ =1-15 lb. water from and at 212" can be evapo-

rated, and therefore instead of 8 lb. we should have had 9*20 lb. of steam
from and at 212' Fahr. per pound of coal.

The above figures assume that the steam generated is dry satu-

472 MARINE ENGINES AND BOILERS.

rated steam, and it will therefore only hold where the steam is really
dry. Figures giving the results of evaporation (so-called evaporative
values, see Table No. 59*) must therefore be received with caution, as
serious errors may arise (1) because any water carried over mechanically
is credited to the boiler as steam ; and (2) because usually it is not
possible to determine the temperature of the feed-water in the experi-
ment in question, nor the temperature of evaporation to which the
figures were reduced.

§ 289. Transference of Steam from Boiler to Engine— The

steam should be taken from the boiler at a position most likely to give
dry steam. As evaporation is very rapid in marine boilers, and the
level of water is generally much disturbed, due to the motion of the
vessel, special steam domes or drums for taking off the steam are often
provided ; or special steam collectors may be fitted to the boiler, with
the object of preventing the water passing over into the steam piping.
Steam domes are used in cylindrical and Normand water-tube boilers.
In water-tube boilers these arrangements are often insufficient, and an
endeavour is made to obtain re-evaporation by throttling down the
steam, as for instance in the Belleville boiler. The advantages of
throttling the steam are much overestimated.

Example. — Boiler pressure, 200 lb. per square inch absolute ; on the
engine side of the throttle valve, 100 lb. per square inch absolute;
dryness fraction of steam, 0-7, i.^., 70 */^ steam, 30 */^ water.

Heat contained in the steam per pound prior to passing through the
throttle valve (from Table No. XXII.)—

353 -*- 0-7 (844) = 943 thermal units.

Heat contained in the steam per pound after passing through the
throttle valve (from Table No. XXII.)—

298 -I- JC2 (883) = 943 thermal units.

As the heat contained in the steam before and after passing through
the valve is the same, the dryness fraction of the steam will be —

^^ = -8-83- = ^""^

that is, of the water in the steam, before passing through the valve, only

•73 _ '7

— 5 — X 100= 10% has been evaporated. It must here be noted that

* In the data given on Table No. 69, the evaporation is given for pressures of frotn
185 to 230 lb. per square inch alx)ve atmosphere, the feed-water not being previously
heated*

STEAM BOILERS. 473

the expansion of the steam passing through the throttle valve takes place
without work being done, and that while steam at a pressure of 200 lb.
per square inch is generated, only a pressure of 100 lb. per square inch
is available at the engine. By far the best method of drying the steam
is to fit separate steam dryers or superheaters into the uptake. On
account of the high rate of evaporation demanded in a marine boiler,
and the corresponding high temperature of the escaping gases, these
dryers partly utilise the heat still available in the escaping gases, and
the result is a net gain.

Example, — As above, let the absolute pressure be 200 lb. per square

inch, and the dryness fraction of the steam when it leaves the boiler 0*7,

evaporation of water per pound coal (from Table No. 59) 8 lb., excess of air

100 °/^, assuming 13 lb. of air required theoretically per pound of coal, then

13
for every pound of wet steam generated, -_ x 2 = 3-25 lb. of air are re-

o

quired. Temperature of the escaping gases before entering the super-
heater 662'' Fahr., after leaving the heater 392' Fahr. Then per pound
of wet steam 270x0-22x3-25=190 thermal units will be available.

190
With these units we can evaporate o-tr^^"^^ ^' ^^ steam at a pressure

o45

of 200 lb. per square inch absolute. As 0*3 lb. of water is present per
pound of wet steam, there would only be 0*3 - 0*23 = 0-07 lb. of water
in the steam leaving the heater, that \% x — 0*93.

The construction and heating surface of the superheater is chiefly
determined in marine boilers by practical considerations, among which
the chief are space available and ease of access. In general it may be
theoretically of advantage to select a type having as many small tubes
as possible, with very thin walls, because the conditions would then be
the most favourable for the rapid transmission of heat. In any case it
is of course essential to cover the exteriors of the steam pipes with
non-conducting materials, to diminish losses by radiation. It must,
however, be admitted that up to the present it has not been found
practicable to use superheaters and steam dryers in connection with
marine engines, the extra complication to the piping not being counter-
balanced by any corresponding marked advantage.

§290. PercenUge of Water in Steam.— With the great
demands made on modern marine engines, and having regard to the
great extent to which the boilers must under certain conditions be
forced, it is advisable to test the steam for its percentage of water
before it enters the engine. The methods hitherto in use are not very
practical, with the single exception of Peabody's throttling calorimeter,
which is adapted for use with marine boilers (see Part VI.).

SECTION 11.
CYLINDRICAL BOILERS,

§ 291. General Remarks. — Direct tube type cylindrical boilers,
or Admiralty boilers, are only used for small engines, or where the
available head room is much restricted (Figs. 421, 422). Otherwise the
return tube type is always employed. (See Figs. 428 to 435.) They
are made either single-ended (Fig. 434) or double-ended (Figs. 428,
429). Single-ended boilers have a heating surface up to 3,000 square
feet, double-ended up to 6,000 square feet. According to the size of the
boiler, single-ended have from one to four furnaces, double-ended from
four to eight furnaces. The height of the steam space (from the crown
of the boiler to the lowest water level) is about 0*225 to 0*25 of Its
internal diameter. The higher it is, the larger will be the water surface,
other conditions being the same, and the greater the capacity of the
boiler for furnishing dry steam. With comparatively small boilers,
horizontal or vertical steam drums connected to the boiler shell are
often provided, so that the steam may be taken off as far away as
possible from the surface of the water. The drum is generally con-
nected to the boiler through an opening of about 16 inches diameter.

§ 292. Selection of Heating Surface and Grate Area.—

The amount of heating surface is determined by the type of engine
selected, and the quantity of steam required to drive it, the method
of firing the boiler (whether with or without forced draught), and the
class of fuel used. To secure economy of coal consumption natural
draught is best, and it is therefore most generally used in merchant
ships. But where the consumption of coal at maximum power is of less
importance, and the main considerations are economy of space and
weight, as in warships, forced draught of a greater or less degree is
almost always employed.

Table No. 60, page 476, gives the usual power developed per square
foot of grate surface, in cylindrical boilers with natural draught, while
Table No. 61 gives these ratios for various types of marine cylindrical
boilers. In calculating the dimensions of the boiler, the grate area,
Table No. 60, is always taken as the basis, and from the ratio h:R,
Table No. 61, the heating surface is determined.

STEAM BOILERS.

476

MARINE ENGINES AND BOILERS.

Table No. 60.

Maximum Horse-power developed per square foot of Grate Surface

with Natural Drauf^ht*

T^ije of Engine iised.

1

Compound.

Trip

e Expansion.

Quadruple
Expansion.

Steam pressures — pounds per

1
I

square inch

Feet.

130

145

160

175

185

200

215

230

245

Height of Funnel

1 above Grate

26-33

7-4

7-7 8-0

8-4

8-5

8-7

9-0

—

33-50

7-8

8-0

84

8-7

91

9-3

9-4

— ,

50-65

8-0

8-3

8-7

91

9-4

9-7

10-0

10-4

10-7

65-82

8-4

8-8

9-2

9-7

101

10-5

10-8

11-2

11-5

„

82-98

—

10-2

10-6

10*9

11-3

II-8

12-2

98-115

1

10-7

111

11-6 12-0 12-5

1

13-^1

Table No. 61.
Heating Surface and Area of Grate in Cylindrical Boilers.

T>T)eofShip.

Kind of
Draught used.

Type of
Engine.

Consumption
of Coal T per

i.H.p. per
hour, pounds.

Heating Sutface
per i.H.p.in
square feet.

Ratio

H H.S.
R G.S.

Merchant steamers
Do.
Do.

Do.

Do.
Ironclads and
heavy cruisers
Light cruisers

Natural
Do.
Do.

Howden

Moderately

forced

Forced

Compound

Triple
Quadruple

Do.

Do.

1 Do.

Do.

2-2 —2-6
1-62-1 -76
1 -43-1 -65

1 -65-1 -87

1 -65-1 -87

2-0 —2-31

3-6— 4-4
3-2— 3-6
3-2-^ -6

2-2-3-2

2-2 31

1-8 2-7
1 •6—2-2

26-35
30-35
32-^>

36-40

36-4f*

25-30

28-32

§ 293. Furnaces and Grates. — Furnaces may be made smooth,
ribbed, or corrugated. If they are plain and not corrugated, their thick-
ness must be relatively large (especially in the case of large furnaces for
high pressures) ; hence they are generally only used for small boilers,
and if their diameter exceeds 2 feet 8 inches, they are strengthened by
what are known as "Adamson" rings, Fig. 422a. For the high pressures
used in triple and quadruple expansion engines, corrugated furnaces are
almost exclusively employed.

* These powers may, with skilful stoking, be maintained for six hours, and under
favourable conditions of wind, &c., for several days.

t The smaller rates of coal consumption are for larger boilers.

STEAM BOILERS. 477

For Morrison corrugated furnaces see Fig. 423.
„ Fox's „ „ „ 424.

„ Purves' „ „ „ 425.

„ Deighton's „ „ „ 426.

The back of the furnace should be connected to the tube plate by
one or other of the methods shown in Figs. 427, 428, or Figs. 430, 432,
where allowance is made for the unequal expansion of the
furnace. Sometimes these joints, or even the whole of the
lower part of the fire box above the bridge, are fitted with
fire-brick masonry, as shown at Fig. 430.

It is advisable to fit the furnace in in such a way that it can
be taken out when necessary, and a new one put in. For
this purpose the opening in the front end plate must be rather
larger than the external diameter of the furnace. Fig. 427, 422A.
and the back flange must be so shaped that it can be drawn
through this opening. Sometimes the back end of the corrugated fur-
nace is drawn in and flanged, in order to facilitate its removal.

The length of the grate does not as a rule exceed 6 J feet (78 inches),
because a greater length presents difficulties in the way of stoking.

Fig. 423. Fig. 424. Fig. 425. Fig. 426.

The length is about twice to two and a half times the breadth (internal
diameter) of the furnace. Where possible the latter should not be
less than 27 inches, and in very large boilers it should not exceed 50
inches.

Number of furnaces in single-ended boilers —

For diameters up to about 9 feet = 1 furnace

» >i II "*" II — •* II

II II II •»•*' II ^ II

II II above 15 „ --=4 „

Double-ended boilers should have twice as many furnaces.

The grate consists of one or of two sets of bars according to its
length, the ends of which rest on the dead-plate and on the bearers of

478

MARINE ENGINES AND BOILERS.

the fire bridge respectively, and when two sets of bars are fitted the
centre is supported on bearers. (See Figs. 427 to 432.)

length of the cast-iron or wrought-iron fire bars from 20 to 50
inches, sometimes as much as 5 feet 6 inches.

Width of the cast-iron or wrought-iron fire bars at the top about 0-5
to 1 inch.

Space between the bars about f to ^ inch.

The fire bars must have sufficient " play," lengthways, as well as side-
ways, to allow them to expand freely, as they get hot With corrugated

Fig. 427.

furnaces special side bars are fixed at the side of the grate, which should
fit accurately into the corrugations of the furnace. The surface of the
grate is generally somewhat inclined towards the fire bridge, so that, if
the latter is fairly high (about 5 to 8 inches), there may still be a suffi-
ciently large space left above it.

With very large boilers having three or four furnaces, the grates
in the two side or wing furnaces are, for convenience of stoking,
sometimes placed rather lower than the grate of the centre furnace,
(See Fig. 427.) The furnace fronts and doors are generally made
of sheet iron, or of cast iron if induced draught or Howden's
forced draught is used, and fitted on the inside with baffle plates

STEAM BOILERS.

479

Fig. 428.

480 MARINE ENGINES AND BOILERS.

to prevent them from getting burnt The usual size of the opening of
fire doors is from 16 by 12 inches up to 18 by 14 inches. Instead of
fire doors which open sideways, those opening upwards and inwards
(see Figs. 427 and 430) or outwards are often used. They are provided
with counterweights to makt them open and keep open with ease.

The fire bridge is made about 5 to 8 inches above the level of the
grate at the back end, and is either horizontal on the top, or rises a little,

to guide the gases of combustion upwards. The clear height of the
flue above the fire bridge is about 0-id to Q'ihJ, d being the internal
diameter of the furnace.

g 294. Boiler Tubes. — The external diameter of these varies from
2 to 3i inches, their thickness from 1 to 15 inch. Thickness of suy
tubes from 2 to '4 inch, according to size of tube.

STEAM BOILERS.

481

Fig. 430.

1

Fig. 431.

2h

482

MARINE ENGINES AND BOILERS.

Pitch of the tubes : —
With an external diameter of 2 inches, the pitch is 2| to 2J inches.

n

H

^i 11 3|

H

3f „ 3}

3

3J„ 4

H

H „ H

H

i^ „ H

The length of boiler tubes, with natural draught, is equal to about
twenty-three to thirty times the external diameter of the tube ; and with

n

Fig. 432.

artificial draught, about thirty-five to forty times the external diameter
of the tube. Boiler tubes are generally merely expanded into the
smooth-drilled holes in the tube plates, while the stay tubes for stiffen-
ing them are first screwed with a fine thread into both ends, and then
expanded. Sometimes the boiler tubes, as well as the stay tubes, are
turned over and beaded. The holes for the boiler and stay tubes
in the front tube plates are made about ^ inch larger than the external
diameter of the tube, so that the tubes may be easily drawn out from
the front. The front ends of the tubes are correspondingly increased
in size.

STEAM BOILERS.

483

§ 295. Manholes are made in the shell, in the steam drum, or in
the front end of the boiler, and should not be less than about 12 by 16
inches. In exceptional cases they may be 11 by 15 inches. Inspection
holes in the front end and mud holes in the lower part of the boiler are
made smaller. All openings for manholes and mud holes must be
stiffened, either by stiffening rings riveted on, or by flanging the edges
of the openings inwards. The smooth inner edge of these flanges is
generally made to serve as the jointing surface of the corresponding

Fig. 433.

cover. If the manhole is in the shell, its smaller axis must run longi-
tudinally to the boiler, and the opening must be strengthened by a ring
riveted on, either inside or out, to make good the material cut away.

§ 296. Thickness of Material Used.— For marine boilers this
is generally calculated according to the regulations laid down by the
Insurance Companies or Classification Societies. The principal rules
laid down by some of these Companies, and the Hamburg Rules of
1898, are as follows.

MARINE ENGINES AND BOILERS.

STEAM BOILERS.

486

MARINE ENGINES AND BOILERS.

■^ <J

^^1%

%ri

O
ft

c5

4- I ^

01

t2o

•"ti

-A- •-dirti -©2

<5

STEAM BOILERS.

487

488 MARINE ENGINES AND BOILERS.

§ 297. German Lloyd's Rules.

{The metric system has been retained for these rules in order
to prevent complication of the formula.)

(a.) Boiler Shells and Steam Domes tested for Internal Pressure. —
The thickness of the plates and the sectional area of the rivets in a
longitudinal joint are calculated according to the following formulae : —

, V ,Ty b e

1. s^— — x-x -

2 B e-d

o d'v p . D n e

2. = — X — X -

4 2 N a

Here s = thickness of the plate in centimetres.

p = working pressure above atmosphere in kilogrammes per square
centimetre.

D = maximum internal diameter in centimetres.

B = tensile strength of the plate in kilogrammes per square centi-
metre.

N = shearing stress of the rivets in kilogrammes per square centi-
metre.

b = working factor of safety for the boiler plate.

n = working factor of safety for the rivets.

e = pitch of the rivets in centimetres.

a = sectional area of all the rivets in a row.

d- diameter of the rivets in centimetres (compare Figs. 436 to 446).

The values allowed for b and n by the constructor must be the
minimum values when tested for strength (see Table XXXIV., Part
VIII.), and must be given on the drawing submitted to the German
Lloyd's for approval.

The shearing stress n of the material of the rivets, if other values
have not been determined by direct tests, is taken at 0*875 of the tensile
strength in wrought-iron rivets, and 0*8 of the tensile strength in mild
steel rivets.

The factor of safety b of the plate is to be taken at 5, but may be
reduced to 4*75 if the rivet holes in the longitudinal seams are drilled
and the riveting done by machinery, and to 4*5 if, in addition to this,
the longitudinal seams are in double shear. If the plate is more than
i inch thick, the circumferential seams must be double riveted ; if it is
1 inch thick or more, the middle circumferential seams must be treble
riveted. The diameter of the rivet must be at least equal to the mean
thickness of the two plates it connects. If stay bolts are attached to the

STEAM BOILERS.

489

shell plates the strength of the rows of stay bolts must not be less than
that of the longitudinal seams of the body.

The factor of safety n for the material of the rivets is to be taken as
equal to b with lap joints ; if the rivets have double butt straps, as equal
to 1'15^. In the latter case it may be taken as equal to b if the rivet
holes are drilled after the plates are fitted together and the joints
riveted by machinery. The thickness of the butt straps must be at least
0*75 that of the plate. In boilers having no middle circular joint, the
outside butt straps must not be thinner than the shell plate, unless the
riveting is done by hydraulic pressure. With thin boiler plates the dia-
meter of the rivets d must not exceed 2j, with thick plates it must not
be taken at less than x. With lap joints and straight-sided butt straps
the pitch of the rivets e must not be more than eight times the thick-
ness of the plate or of the strap. If the butt straps are straight, the
pitch of the rivets must not exceed ^ = c.j-h4*l centimetres, c being
a constant, taken from the following table : —

Tabie of Values ofc.

Number of Rivets
in one Row.

1
2
3
4
5

c for Lap-jointed Longi-
tudinal Seams.

1-31
2-62
3-47
4-U

c for longitudinal Seams
in Dquble Shear.

1-75
3-50
4-63
5-52
6-00

If the pitch is more than 254 millimetres, the butt-strap riveting must
be zigzag. (See Figs. 433, 443, 445, and 446.) With double-riveted joints
the pitch of the rivets must not exceed 3*75//; with zigzag-riveted joints
the diagonal distance between two rivets must not be less than 2*4^.
The strength of welded seams in wrought-iron boilers is taken at 0*7 of
that of the solid plate. Stiffening rings round manholes, &c., must have
a sectional area sufficient to counterbalance the weakening produced in
the plate.

{b.) Flat Plates, — For flat boiler plates stiffened with gusset stays or
stay bolts, the thickness of the plate is determined from the formula —

J = c . / Vp for steel plates.

J = 1 -12 c. / \/p for wrought-iron plates.

490 MARINE ENGINES AND BOILERS.

Where s — thickness of the plate in centimetres.

/= distance between the stays in centimetres.

p = pressure above atmosphere in kilogrammes per square

centimetre,
c = constant, with the following values : —

c = 0-024 if the plates are in contact with the hot gases and the water,

and the stays are screwed in, and riveted over,
c = 0'022 if the plates are in contact with the hot gases and the water,

and the stays are screwed in, and fitted with nuts outside.
c = 0*021 if the plates are not in contact with the hot gases, and the

gusset stays and stay bolts are screwed in, and have riveted heads,
c = 0-020 if the plates are not in contact with the hot gases, and the

gusset stays and stay bolts are screwed in, and fitted with nuts.
c = 0*018 if the plates are not in contact with the hot gases, and the

diameter of the washer is 0*4/, and its thickness 0*667j.
c = 0*017 if the plates are not in contact with the hot gases, and the

diameter of the washer is 0*6/, and its thickness 0*833j.
€ = 0*016 if the plates are not in contact with the hot gases, and the

diameter of the washer is 0*8/, and its thickness equal to s.

The thickness of the flat plates in the vicinity of the nest of boiler
tubes is determined from the same formula, / being the mean distance
between the stay tubes, taken from centre to centre, and c = 0*020.

To determine the thickness of the flat plates between the nests of the
boiler tubes, we must take

/= distance from centre to centre of the bounding rows of tubes in

centimetres,
c = 0*022 if in the bounding row of tubes every third tube is a stay tube,
c = 0*020 „ „ ., every second tube „

c = 0*019 „ „ „ each tube is a stay tube.

If the top of the combustion chamber is not connected by ties, &c.,
to the boiler shell, but supported on girders which project beyond the
edges of the tube plates, the thickness of the latter must not be less
than

_ px wx^

Where j = thickness of tube plate in cm.

w = width of combustion chamber in cm.

b = distance between the boiler tubes from centre to centre.

d= inside diameter of the boiler tubes in cm.

p = allowable boiler pressure in kilogrammes per sq. cm.

STEAM BOILERS. 491

Boiler plates which are in contact on one side with the hot gases,
and on the other with the steam, must be 10 per cent, thicker than
the calculated value, and it is also advisable to protect them by baffle
plates.

(c) Iron or Steel Furnaces, — ^The thickness of furnace plates is de-
termined from the following formula —

X = 0-00385 VpTdTl

Where j = thickness of plate in centimetres.

p = working pressure above atmosphere in kilogrammes per

square centimetre.
D = external diameter of the furnace in centimetres.
L = length of the furnace in centimetres, or if stiffening rings

are used, greatest distance between two rings.

The thickness of the plate s must not be less than that given by the
following formulae. With ribbed furnaces (Purves* patent), d denotes
the external diameter of the flat part between the ribs ; and with corru-
gated furnaces, it denotes the external diameter at the lowest part of
the corrugation.

P D

s = ' + 0*3 : for plain furnaces without Adamson rings.

740

P.D

900

P D

s = -^^^7:^ + 0*3 : for plain furnaces with one Adamson ring. The distance

between the ring and the plate must not exceed 1*22
metres.

P D

s = ' + 0-3 : for plain furnaces with two Adamson rings. Distance

1010

between the rings not to exceed '79 metre.

P D

j= '^ +0-3 : for plain furnaces with three Adamson rings. Distance

between the stiffeners not to exceed '61 metre.

P D

s = ' -H 0-3 : for corrugated furnaces (Fox, Morrison, or Deighton's

patent) having a tensile strength of 35 to 41 kilogrammes
per square millimetre (22 to 26 tons per square inch).
The thickness of corrugated furnaces should be at least
0*8 centimetre (say % inch), height of corrugation at
least 3 8 centimetres (1^ inch), and length of the plain
end not more than 25*4 centimetres (10 inches).

s ^ T^s7\ + ^'^ • ^or plain furnaces with four Adamson rings, in which

1220

the space between the rings does not exceed 1 foot
8 inches, and also for ribbed furnaces (Purves* patent).
The height of the ribs must not be less than 3-4 centi-

492 MARINE ENGINES AND BOILERS.

metres (If inch) above the plain part, depth of the inner
grooves not more than 1'9 centimetre (f inch), length
between the ribs not more than 22 -9 centimetres (9
inches), and length of plain end not more than 15*2
centimetres (6 inches).

{d,) Stays and Stay Bolts or Screiv Stays, — The stress upon welded
iron stays must not be more than 0*1, and with iron or steel non-welded
stays not more than 0*143 of the tensile strength of the material, but
the following limits must not be exceeded : —

For wrought-iron welded stays, 350 kilogrammes per square centi-
metre (5.000 lb. per square inch).

For wrought-iron non- welded stays, 500 kilogrammes per square
centimetre (7,000 lb. per square inch).

For mild steel, 600 kilogrammes per square centimetre (8,500 lb. per
square inch).

{eJ) Girders for supporting Tops of Combustion Chambers, — The
girders for the flat tops of combustion chambers are calculated from
the following formulae —

, p(w-/)^.Lf V ^ . , AnP(w-/)^.L^ ^ ,
b = -^ -4 — for wrought iron. b = 0*9 -> /„ for steel.

c./i' c.h^

Here w = width of combustion chamber in centimetres.

p = working pressure above atmosphere in kilogrammes per square

centimetre.
/= distance of the stays from each other in the girder, or if there

is only one stay, half the length of the girder, in centimetres.
e = spacing between the girders in centimetres.
L = length of the girders in centimetres.
A = height „ „

b = thickness „ „

c = 420, if there is one stay in each girder,
c = 630, if there are two or three stays in each girder,
c = 720, if there are four stays in each girder.

The ends of the girders must fit on to the vertical end walls of the
combustion chamber, and must project about 4 centimetres (H inch)
above the top.

(/) Donkey Boiltrs. — As far as the rules already laid down for the
construction of boilers are applicable, they apply also to donkey boilers.

(^.) Manufacture of the Boiler. — This can only be efficiently carried

STEAM KOILERS. 493

out, and the coefficient of safety for the strength of the shell plates, as
already mentioned, be determined, if the following conditions are com-
plied with : —

The preparing and working of the material, such as bending, dishing,
and flanging the plates, drilling the holes, &c., must be done with the
greatest care, and in a satisfactory manner. If the rivet holes do not
exactly coincide, they must be rimered with the plates in place. Riveting
the joints and caulking the seams must be done as carefully as possible.
If the holes are punched, they must be enlarged sufficiently to make
good the damaged metal round them. If the edges of the plate are
torn, or the rivets defective, they must be rejected, and sound plates
and rivets substituted for them. All seams must, if possible, be caulked
inside and out.

The shell plates of cylindrical boilers must be bent with the grain.
The butt straps must be cut from plates of the same quality as the shell
plates, and the grain should run in the same direction as that of the
plates. If single butt straps are used, they must be ^ inch thicker than
the boiler plates.

All openings for manholes, steam domes, &c., must be strengthened
with riveted wrought iron or steel rings of flat, angle, or T section, or
better by dishing the plates, so that the weakening of the plate produced
by the hole is completely made good. If there are mud holes in the
bottom end plates, .and these are for other reasons made stronger than
the rest of the boiler, it is not necessary to compensate for the weakening
of the plate. All the larger fittings must be secured by means of studs
or set screws to thick faced seatings or flanges riveted to the sides of the
boiler. These studs must not penetrate the boiler plate. Steel stays
should not be welded. If rivet holes are drilled in a boiler with the
plates roughly erected together, it is advisable, after the holes have been
made, to take the plates apart, and remove the burr. Angle joints
should as far as possible be made by dishing the plates. To rivet by
hydraulic pressure is generally better than to rivet by hand. ^

It depends chiefly on the material used, and its treatment, whether
the steel plates should be annealed, and this is a question which must
be left to the judgment of the boilermaker ; but intense heating of the
plate locally should in any case be avoided. The material to be worked
up must comply with the conditions laid down in the rules for testing
wrought iron and steel to be used in boilers. (Compare Table XXXIV.,
Part VIII.)

§ 298. Hamburg Standard, 1898.— Extract from the rules for

494

MARINE ENGINES AND BOILERS.

calculating the strength of material for new steam boilers (here also
the metric system has been retained) : —

1. Boiler Shells.

The thickness of the plate of the boiler shell is j = d\'^

^ 200k..,

Here s — thickness of plate in millimetres.

d~ inside diameter in millimetres.

p — maximum working pressure above atmosphere, kilogrammes
per square centimetre.

X = factor of safety.

K = tensile strength of the material, kilogrammes per square
millimetre.

;? = strength of riveted joint, as compared with strength of
whole plate.

The thickness of the plate must not be less than 7 millimetres.

The factor of safety x of the plate must not be less than 4*5 at the
weakest part.

If the joints are in double shear, the factor of safety may be reduced
to 4, but this assumes the boiler to have been most carefully constructed
in every way. It should be considered whether, to conform to local
working conditions, additional thickness is required, and this is necessary
if the calculation gives a thickness of plate of less than 10 millimetres.

The strength of material to be used is determined by the Wiirzburg
Standard of 1895.

I. Wrouf;ht Iron {Tensile Strength),

1

Furnace Plates.

1

1 Dished End Plates.

SheU Plates.

1 With the

1
Acrote the

1

, With the

Across the

, 1

With the

Across the

Grain.

Grain.

1 Grain.

Grain.

Grain.

Grain.

Number denoting

quality uf metal

or quality-factor -

, 56

49

50

45

43

3S

Tensile strength in

kilogrammes per

sq. mm.

36

34

35

33

33

30

tons per sq. in. -

22-8

21-5

22-2

21

21

19

Elongation percent.

on 200 mm. (say

8 inches) -

20

15

15

12

10

8

STEAM BOILERS.

495

II. MUd Steel {Tensile Strength).

Furnace Plates.

Disbed End Plates.

Shell Plates.

With the

Across the < With the

Acrosfi the

Acrofts the

With the

Grain.

Grain. Grain.

Grain.

Grain.

Grain.

Number denoting |

quality of metal '

1

or quality-factor - |

62

62 61

61

60 m

: Tensile strength in <

kilogrammes per

sq. mm.

34-40

34-40 36-42

36-42

39-45 39-45

tons per sq. in. -

21 0-25 -3

21 -5-25 -3 '22-8-26 -6

22 -8-26 -6 24 -7-28 -5 24 -7-28 -5

Elongation percent. I

on 200 mm. (say

8 inches) -

25

25 -22 22 1

20

20

If the thickness of the plates is calculated on the basis of a greater
tensile strength than the minimum values given in the Wiirzburg
Standard, proof may be required that the plates really are of the strength
which forms the basis of calculation.

The seams must be riveted in such a way that they will withstand
all tendency to slip, and the resistance of the rivets to shearing stress
must not be less than the strength of the plates allowed for in the riveted
joint The maximum stress in a rivet in kilogrammes per square milli-
metre and in tons per square inch must not exceed —

For a single riveted lap joint (!"^,''"°«'*'"'"^ P*^"" '^""^ m»l™etre.

U'44 tons per square inch.

f 6*5 kilogrammes per square millimetre.

1 4*1 2 tons per square inch.

/6-0 kilogrammes per square millimetre.

1 3-8 tons per square inch.

single riveted double butt strap joint (J^J^ ^^' P^'^ sq. mm.

17*6 tons per sq. m.

/11*5 kg. per sq. mm.

" " 17-3 tons per sq. in.

>»

>»

»>

9)

double

treble

»»

»

»

>>

double

treble

rllO kg. per sq. mm.
\7-0

tons per sq. in.

If butt straps are used for the joints, they must be cut from plates of
at least the same quality as the shell plates.

2. Furnaces.
The thickness of these should be —
(a.) With plain furnace plates —

5 =

2,000

(

1 + - X - -

/ /+

\-dl

+ r.

496 MARINE ENGINES AND BOILERS.

Here s = thickness of plate in millimetres.

/ = maximum working pressure above atmosphere in kilo-
grammes per square centimetre.

i/= internal diameter of furnace in millimetres.

/= length of furnace, namely, greatest distance of the stiffeners
from each other, in millimetres.

a =100 for horizontal furnaces with lap-jointed longitudinal
seams.

a = 80 for horizontal furnaces with butt straps or welded longi-
tudinal seams.

^=1*5 millimetre if/ is as much as 70 lb. per square inch.

^=1*0 „ „ „ 85 „ „

^ = 0-5 „ „ „ 100

^=zero if/ is above 100

(d.) With corrugated and ribbed furnaces, if in the latter the distance
of the ribs is 9 inches apart, then —

/•^ .
1,000

taking r at 3 millimetres.

The thickness of plate must not be taken at less than 7 millimetres.

3. Flat Surfaces.
(a,) Hat Plates, — The thickness of these should be —

= 1-5 + 0-1^^^^

Where j = thickness of plate in millimetres.

/-maximum working pressure above atmosphere in atmo-
spheres.

e — distance of the stays or screw stays from each other in
millimetres.

K = tensile strength of the material in kilogrammes per square
millimetre.

r= 1'323, if the stays or screw stays are screwed and riveted
into the plates.

c= 10314, if they are screwed into the plates and fitted with
a nut on the outside.

^ = 0*9774, if they are screwed into the plates, and fitted inside
and out with nuts and washers, the diameter of which
must be at least equal to four-tenths the distance between
the stays or rows of stays. Thickness of washers at least
\Sy and it must be increased if the diameter of the

STEAM BOILERS. 497

washers is more than 1*5 times the diameter of the
nuts, measured across the corners.
r= 0*8658, if the stays or screw stays on each side of the
plate are fitted with nuts and washers, and the outside
washer is riveted to the plate, and has a thickness of at
least 0-75j and a diameter of at least 0*6^.

{b.) Dog Stays on Flat-bottom Plates,— Wxih these —

VI^J-K'40]

Here 5, /, and k have the same meaning as before.

r= inner radius of rounding of dog stay in millimetres.
^= inner diameter of the bottom plate in millimetres.

4. Rounded Thick Bottom Plates without Staving.
The thickness of these (to resist internal pressure) is —

Here s and k denote the same as before.

r = radius of the arc of the circle in millimetres, assuming that

it is about equal to the diameter of the corresponding

boiler shell.
k ^ allowable stress on the material in kilogrammes per square

millimetre, namely —

For wrought iron up to 4-5 kilogrammes.
For mild steel up to 6*0 kilogrammes.
For copper up to 2*5 kilogrammes.

5. Stays and Screwed Stays.

The stress in these should not exceed —

5 kilogrammes per square millimetre for unwelded iron stays.

6 „ „ „ „ mild steel stays.
3 „ „ „ copper stays.

If the pressure is high (10 atmospheres and above) it is advisable
to screw those longitudinal stays which have nuts, as well as the stay
tubes, into the plates they support, and the former should also be fitted
inside and out with nuts, a suitable washer being fitted under each outer
nut; the stay tubes should, however, be expanded and beaded over.
The end stays should be as long as possible.

2 I

498 MARINE ENGINES AND BOILERS.

6. Stays for Fire-box Tops.
The projecting unattached girders must be calculated as follows : —

h _/(«/ - d)e I

Here p — maximum working pressure in atmospheres.
w = width of fire box in millimetres.
^= distance of stay bolts apart in millimetres.
e = distance of girders apart in millimetres.
/= length of girder in millimetres.
A = depth of girder in millimetres.
d = width of girder in millimetres (or total thickness of girder

plates).
r=420 if there is one stay to each girder.
c= 630 if there are two or three stays to each girder.
^= 720 if there are more than three stays to each girder.

If the girders supporting the top are suspended, they must be cal-
culated according to the altered proportions of the load upon them.

§ 299. Extract from Rules of the " Bureau Veritas."

(a.) The thickness of the shell plates and rivets is calculated from
the following formulae —

Here /= thickness of shell plate in millimetres.

P = working pressure above atmosphere in atmospheres.

D = greatest inside diameter of boiler shell in centimetres.

R ^ allowable tensile stress, kilogrammes per square millimetre.

The latter is—

For iron or steel r^ "minimum tensile strength of material

4
For iron r = 7*9 kilogrammes per square millimetre, if

the minimum tensile strength of material

is not known.

Further a = ^~'

P

p = pitch of rivets in outer row in millimetres.
d= diameter of rivet holes in millimetres

(2.)

STEAM BOILERS. 499

P. D ./

'2s

Here p and d are the same as before.

/= pitch of rivets in the outer rows in centimetres,
s = maximum stress in pounds per square inch which will be
allowed on the rivets.

For steel rivets, one-fifth part of the ultimate tensile
strength can be taken.

For iron rivets, 6*3 kilogrammes per square millimetre,
corresponding to an ultimate tensile strength of about 31*5
kilogrammes per square millimetre.
A = total shearing surface in square inches of the rivets (/.^.,
twice the area of the rivet hole when the rivet is in double
shear), when machine riveted.

Only \^ of the full area must be taken when the
riveting is done by hand.

■

/rt \ P . D . / C. S

(3.) B = — -

^ ' 2r R

Here p, d, /, r, and s have the same equivalents as before.

B = sectional area in square millimetres of the plate, on portion

/ of the joint, along the line of its supposed rupture,

assuming thickness of plate to be reduced by 1 millimetre,

due to corrosion,
c = total sectional area of rivets which are exposed to shear

in the length /.

(4.) /=#-+!

^ ^ 20aR

a = I

Here /= thickness in millimetres of single butt strap, or sum of
thicknesses, if there are two straps.
^ = pitch of rivets in inner row in millimetres,
^"s diameter in millimetres of rivet holes in inner row.

/Remarks. — If the above formulae are used, it is assumed that all the
rivet holes are drilled after the plates are bent. These holes must be
at the distance of the diameter of one rivet from the edge of the plate.
In zigzag riveting, the distance between the rows is to be such that there
is no fear of a rupture through plate or butt strap along the zigzag
line. When stays are bolted through the shell, they should be so
arranged that they do not weaken the shell plates more than the

500 MARINE ENGINES AND BOILERS.

riveted joints. For circumferential seams, double riveting will be
required if the thickness of the plates exceeds 12^ millimetres (| inch).
In double-ended boilers with six furnaces, treble riveting will be
required for the circumferential seams connecting the shell rings with
each other ; it is not required for the end seams.

(d,) Flat Plates,— Fox these /» 1*5 + /(«« + b^)^^±

Here p = working pressure above atmosphere in kilogrammes per

square centimetre.
/= thickness of plate in millimetres.
a = pitch of stays in one row in centimetres.
b = distance in centimetres between two rows of stays. In case

of irregular staying, the mean distance between the stays is

to be substituted for Va^T^.
T = tensile strength of material of plates in kilogrammes per

square millimetre.
K = 0*735 when the stays are screwed into the plates and

riveted over.
K = 0-578 when the stays are screwed into the plates, and fitted

with outside nuts at either end.
K = 0*542 when the stays are screwed into the plates, and fitted

with nuts and washers inside and out, and the outer

washer has a diameter of at least 0*4a, and a thickness of

2
at least ^ /.
o

K = 0*481 when the stays are fitted with inside and outside

nuts and washers, the outside washer being riveted to

the plate, and at least 0*75/ thick, and its diameter at

least 0*6^.

In flat plates which are in contact on one side with the steam, on
the other with the hot gases —

'=3+y?^^

9t

When the front plates are in two parts, the lap joints must be double
riveted if the plate is 13 millimetres {\ inch) or above.

{c) Stays and Screw Stays. — For these —

d= 3 millimetres + Vl?^

STEAM BOILERS. 501

Where //= inner diameter of the stay in millimetres.
Q = total load upon stay in kilogrammes.
T = tensile strength of the material in kilogrammes per square
millimetre, namely —

For steel, the lower limit assumed for tensile strength
(tensile strength, 35 to 47 kilogrammes per square milli-
metre).

For iron, 35 kilogrammes per square millimetre.

The stay tubes must be screwed into the plates they support.
(d.) Cylindrical Furnaces.

(1.) The thickness of plain cylindrical furnaces should be —

V'-^

- for iron.

- for mild steel.

2c

/ = required thickness of plate in millimetres.

D = outside diameter of furnace in centimetres.

p = working pressure in kilogrammes per square centimetre above

atmosphere.
L = length of furnace in centimetres, or if made with efficient rings

the length between the rings.
c = 588 when the furnace is truly circular, and the longitudinal

seams are welded and butt-jointed, or lapped, bevelled, and

double riveted.

The thickness of tube must not, however, be less than —

/=- ' for iron plates.
l^-wo- ^^^ steel plates.

DO

(2.) With corrugalcd (urnsices —

p and / having the same meaning as before.
Ds= outside diameter in centimetres measured across the top
of the corrugations.

It is here assumed that the

Depth of the corrugations is at least 4 centimetres (1 J inches).
Length „ „ 15 „ (6 „ ).

502

MARINE ENGINES AND BOILERS.

(3.) With ribbed furnaces —

D denoting the greatest outside diameter between the ribs in centimetres.

Here it is assumed that the distance between the ribs = 23 cm.

Height of the ribs = 35 mm.

It is assumed that the material used has a tensile strength of 41 to
47 kilogrammes per square millimetre (26 to 30 tons per square inch).

§ 300. Extract from Lloyd s '' Reg:ulations for British

and Foreign Shipping.

1)

1. Boiler Shells,

/=— ^ — for iron boilers.

T =

C. B
A . D
C . B

+ 2 for steel boilers.

Where /= thickness of plate in inches.

T == thickness of plate in sixteenths of an inch.

A — working pressure in pounds per square inch.

D = mean diameter of shell in inches.

B = percentage of strength of joint.

c = coefficient according to following table.

Values of c for Iron Boiler Shells

(Lloyd's Rules).

For Plates

For Plates

For Plates

Description of Joint.

\ inch thick
and under.

{inch thick and
above ^ inch.

above f inch
thick.

T-ap joint —

Punched holes -

155

165

170

Drilled holes

170

180

190

Double butt strap joint —
Punched holes -

170

180

190

Drilled holes

180

190

200

Values of c for Steel Boilers,

c = 21 with double butt straps of equal width.

c = 20 25 with double butt straps, the outside strap having one row of

rivets more than the inside.
c= 19*5 with lap joints.

STEAM BOILERS.

503

If the tensile strength of the boiler shell plate is more than 27 tons
per square inch, c may be correspondingly increased.

The inside butt strap must be at least | the strength of the longi-
tudinal joint.

B=:the minimum percentage of strength of the longitudinal joint for
plate at joint, found as follows : —

H = ^^^

xlOO

and for the rivets at joint

. 90 for iron plates and iron rivets, drilled holes.

B = ^^^ . 85 „ steel

B =

n,a
JTt
n.a

B = ^.70 „ steel

)»

}>

steel

iron „

Where / = pitch of the rivets in inches.
/= thickness of plate in inches.
^= diameter of rivet holes in inches.
n = number of rivets used per pitch in a longitudinal joint.
a = sectional area of rivet in square inches.

In case of rivets in double shear 1*75^ is to be used instead of a.

Proper deductions are to be made for openings in shell.

All manholes in circular shells to be stiffened with compensating
rings.

Shell plates under domes in boilers so fitted, to be stayed from the
top of the dome, or otherwise stiffened.

Note, — For the shell plates of superheaters or steam chests enclosed
in the uptakes or exposed to the direct action of the flame, the coefficients
in the above tables should be | of those given.

Allowable Strains upon Stays, Stay Bolts, and Stay Tubes in pounds

per square inch (Lloyd's Rules).

Diameter at Bottom of
Description of Stay. , Thread less than

1^ inch.

1

Diameter at Bottom of

Thread more than

\\ inch.

Iron stays -

Steel stays (screw stays)*
Steel stays (other stays)*

6,000 lb. per sq. in.

(welded).
8,000 lb. per sq. in.
9,000 „

7,500 lb. per sq. in.
(unwelded).
9,000 lb. per sq. in.
10,000 „ „

* No steel stays are to be welded.

504 MARINE ENGINES AND BOILERS.

The maximum stress in stay tubes must ilot exceed 7,500 lb. per
square inch.

2. Flat Plates or Stayed Surf aces, — For these t = p /'-

c X T-
- orA= — 5-

p-

Where r = pitch of stays in inches. If the pitch in the rows is

not equal to that between the rows, the mean of the
squares of the two pitches is to be taken.

T and A as before.

c = constant, the values of which are given in the table on
next page.

Where doubling plates are employed and securely riveted to the flat
plates, their thickness / in sixteenths of an inch is /=2 (p^/- - t)

/ /\2

or A = \ ^/ where / is not to be less than two-thirds of r.

p2

In the case of front plates of boilers in the steam space, c is to be
taken at 20 7^ less, except where the plates are shielded from the direct
action of the heat.

For the wide water spaces between the nests of tubes t = p I-

or

C X T^

A = — — . Here p = horizontal distance from centre to centre of the

bounding rows of tubes, and c as follows : —

c= 120, if every third tube is a stay tube, and not fitted with nuts

outside the plates.
c= 130, if every third tube is a stay tube, but with nuts outside the

plates,
c = 140, if every second tube is a stay tube, and not fitted with nuts

outside the plates.
c= 150, if every second tube is a stay tube, with nuts outside the

plates.
c = 160, if every tube is a stay tube, and not fitted with nuts outside

the plates.

For steel tube plates in the nest of tubes the strength to be taken
from

U0xt2 ^ ^ „ / A

= AOrT = P./_

= P l±-

V uc

where t and a are as before, and p = the mean pitch of stay tubes from
centre to centre.

STEAM BOILERS.

505

For the steel tubes in the vicinity of the nest of boiler tubes c must
be taken as = 140 ; p is the mean pitch of the stay tubes.

Values of c for Flat Plates (Lloyd's Rules).

Kind of Plate.

Kind of Stays or Stay Bolts.

Thickness of Plates.

c.
90

Iron or steel

Screw stays with riveted

yV in- and under.

heads.

Do.

Do.

Above jV inch.

100

Do.

Screw stays with nuts.

tV in. and under.

110

Iron -

Do.

Above yV ^nch.

120

Steel

Do.

Above yV and
under ^^ inch.

120

Do. -

Do.

yV in. and above.

135

Iron -

Screw stays with double
nuts.

Do.

140

Steel not exposed

to the fire

Do.

Do.

175

/

Stays with double nuts,
and washers outside
the plates, having a

\

Iron - - i

diameter = ? and thick-

> Do.

150

\

T

ness = - .
2

J

Steel not exposed

to the fire

Do.

Do.

185

r

Stays with double nuts,
and washers riveted to
the outside of the

\

Iron - - <

plates, having a dia-
meter = 0-4p and thick-

Do.

160

\

T

ness = -.

/

Steel not exposed

md

to the fire

Do.

Do.

200

1 Iron -

1

Stays as above, but dia-

Do.

meter of washers = |p

175*

1

and thickness = t. j

Steel not exposed

to the fire

Do.

Do.

220

The thickness of the tube plates of combustion chambers, in cases

* Or 190 if P be taken as the pitch of the stays in the row.

506 MARINE ENGINES AND BOILERS.

where the pressure on the top of the chamber is borne by these plates,
is not to be less than that given by the following rule —

A. w. D
T =

1,750 (D-^

Where a = working pressure in pounds per square inch.

w = width of combustion chamber over plates in inches.
D = horizontal pitch of tubes in inches.
T = thickness of tube plates in sixteenths of an inch.
i/= inside diameter of plain tubes in inches.

3. The thickness of the girders supporting the top of the combustion
chamber is taken at —

A (l - P) D . L C X ^/2 X T

T = — i \. or A =

Q.d' (L - P) X D X L

Where

L = width of combustion chamber between the plates \

T = thickness of girders at centre I

i/= depth of girder at centre ^ .

D = distance from centre to centre of girders

p = pitch of stays in girder

A = working pressure in pounds per square inch

c = 6,600 for steel girders with one stay to each girder.

0 = 9,900 „ ., two or three stays to each girder.

0=11,000 „ „ four or five „ „

0 = 11,550 „ „ six or seven

c= 11,880 „ „ eight or more

9) »

4. Fiirnaces, — The strength of plain furnaces, where l>120t, is

"V 1,075,20

1,075,200x12
or A = ' '

200 L X D

Here a = working pressure in pounds per square inch,

T = thickness of plates in inches.

L - length of the plain cylindrical part in inches, measured from
the centre of the rivets connecting the furnaces to the
flanges of the end and tube plates, or from the banning
of the curvature of the flanges of the furnace, where it is
flanged or fitted with Adamson rings.

D = outside diameter of furnace in inches.

STEAM BOILERS.

507

If the length of the plain part of the furnace is less than 120 times
the thickness of the plate, then —

a^d + 50l _ 50 X (300t - l)

^ IpOO" ^''^ 5

If the above formulae are employed, it is assumed that the steel will
have a tensile strength of not less than 26, and not more than 30 tons
per square inch. If the material of furnaces has a less strength than
26 tons per square inch, then for each ton per square inch below the
minimum tensile strength of 26, the coefficient is to be correspondingly
decreased by oVth part.

With steel corrugated tubes t = ^i^ + 2, or a = ^iiilzA)

T be

c ,
c .

ng = thickness of plate in sixteenths of an inch.

= outside diameter of corrugated furnaces, or smallest out-
side diameter of ribbed furnaces in inches.

= 1259 with Fox, Morrison, Deighton, or Beardmore's cor-
rugated furnaces.

= 1160 with ribbed furnaces (spacing between the ribs 9
inches).

= 912 with spirally corrugated furnaces.

= 945 with Holmes' patent furnaces, corrugation not more
than 16 inches apart centre to centre, and not less than
2 inches high (t for plain portions, and D for plain
parts).

Table
Cylindrical

Type of Ship - - - - |

Fast
steamer

Fast
steamer

Cargo
steamer

Caigo
steamer

Kind of Draught used

Natural

Howden

Natural

Natural

Type of Boiler

Double
ended

Double
ended

Double
ended

Single
ended

Heating Surface - sq. ft.

6,466

6,090

«5,>ToU

2,000

Grate Area - - - sq. ft.

200

155

110

55

Heating Surface : Grate Area ratio

321

39-3

36-3

36-4

Press, above Atmosphere lbs. per sq. in.

178

214

214

1

214

Internal Diameter - tfttkgs

200

196

162

162

C/5

Length .... inrAes

244

244

238

132

Thickness of Shell Plate - inches

1/ir

m

ly.

\h

i

Tensile Strength of material \
^ tons per sq. in. ]

28—30

33—36

28 32

•28-32

Type of turnace ....

Purves

Morison

Morison

Morison

Number of Furnaces - - - -

8

8

6

3

s

Internal Diameter - - inches

47i

43-5

39-5

39-5

u

g

Thickness of Furnace Plate inehes

\\

M

H

\\

fa

Number of Combustion Chambers perl
boiler /

3

3

3

3

0

Minimum clear Space between Boiler)
V Shell and furnace - - inehes]

4-84

5

5-86

4-44

/Outside Diameter of Boiler Tubes and\
/ Stay Tubes - - - incAesJ

3i-3}

2i

3

3

Thickness of Boiler Tubes (welded iron\
tubes) - - - - inches]

016

013

015

015

1

Thickness of Stay Tubes (welded iron\
tubes) -. - • - ituhes]

0-31

0-19-0-31

0-39

0-39

Number of Boiler Tubes •

674

592

352

171

5 I Number of SUy Tubes - . -

194

364

222

111

c

Length between the Tube Plates inches

92^

94

94

92

Vertical Pitch - - - inches

416

3-85

4-09

4^

«

Horizontal Pitch - - ittches

413

3-85

4-09

409

Thickness of Tube Plates - inches

0-94

0-94

0-98

0-98

Lowest Water-I^vel above middle of \
Boiler .... inches)

5216

52-08

42-44

42-44

\lleight of Combustion Chamber inc/ics

49-36

53*38

47-2

'26-77

No. 62.
Boilers.

Scout

Qosed
stokehold

Double
ended

4,800
172
27-9
170
180
200

lA

39-5

2x2
3-58

Cruiser

Closed
stokehold

Double
ended

4,067
122
33-2
185
150
214

H

Purves Morison

6
37-5

013

3-74

2i

0-11

Cargo
steamer

Natural

Single
end^

2,582
74
351
206
183
140

28—32

Morison
3

m

5

Cargo
steamer

Natural

Single
ended

1,400
43
321
170
130
133

28—32

Morison

39-5

3-46

015

015

Cargo
steamer

Natural

Double
ended

3,850
122
31-4
185
163
224

lA
28—32

Morison
2x3
39-5

4-72

3 ♦

015

Cargo
steamer

Natural

Single
ended

1,130
37
31 1
114
126
124

i
28-32

Plain

2

37

i

5-70

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Cargo
steamer

Natural

Single
ended

970

32

30

170

126

126

26—29

Plain
2
36

31

Cargo
steamer

Natural

Single
ended

753

26

29

150

96

106

S

26—29

Plain

2
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Natural

Single
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430
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140
83
95

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0-23

580
136
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3-85
0-86

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0-23

618
190
83.
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3-38
0-90

40-63

0-39

251
125
90
3-93
3-93
0-96

46-18

0-39

144

62

88i

3-93

3 89

0-82

35-43

0-31

430
164
86i

4

4
0-86

40-78

0-39

132

40

86

3-97

3-93

0-70

32

84

54

83

4-48

4-48

0-92

80

40

79

3-85

3-85

0-70

62

18

67
3-9
3-9
0-70

2204

45-43

26-77

24-8

49-60

20-47

SECTION III.
LOCOMOTIVE BOILERS.

S 301. Dimensions of Locomotive Boilers. — Locomotive
boilers were Tornierly in use for all light war vessels. Since the beginning
of the last decade they have been almost entirely superseded by water-
tube boilers, and are now scarcely ever constructed, except to replace
old boilers of this type. For method of construction see Figs. 447, 4-l!<.
In newer boilers of this type the lower edge of the fire box, which is
attached to the boiler shell, forms the lower part of the space dividing
the fire box from the shell (in Figs. 447, 448 the division is made by
a rectangular wrought-iron
ring).

Dimensions, &c., of a
modern marine locomotive
boiler. Power developed,
900 i.H.P. with forced, 600
i.H.p. with natural draught.
Pressure above atmosphere,
about 1601b. per square inch.
Heating surface, 1,700 square
feet. Grate area, 470 square
feet. Total length, 19 feet.
Diameter of the shell (cylin-
drical part containing the
tubes) = 1 feet 9 inches.
Length of the tubes, 7 fed
*■ ■ \\ inch. External diameter

of the tubes, 2 inches.
Number of tubes, 420, of which 90 are stay tubes. Width of the furnace
(rectangular part containing the grate), ^ feet 6 inches at bottom, T
feet 9 inches at top. Height, 8 feet 2 inches, of which 4 feet 2 inches
are below the middle of the shell. Width of grate = 2 feet x 3 feet
6 inches. Length of grate = 6 feet 11 inches.

When used in light war vessels, these locomotive boilers are generally
highly forced (pressure of air in the combustion chamber or stolcebold

STEAM BOILERS.

511

being as much as 4^ to 5 inches, and the fuel burnt per square foot of
grate surface per hour as much as 82 lbs.

The ratio of heating surface to grate area varies considerably, from

H

" = 35 to 55. See also § 321, and Table No, 68.

to

SECTIOX IV.
WATER-TUBE BOILERS.

i 302. General Remarks. — Water-tube boQers aie now almost
always fitted in light war vessels, torpedo-boats and destroyers, and
cruisers. In the larger class of warships they are frequently fitted
conjointly with cylindrical boilers. Hitherto they have been but little
adopted in the mercantile marine.

The advantages of water-tube boilers may be stated in general
tobe*—

1. Capacity to raise steam rapidly.

2. Facility in responding to a change of speed in the engines.

3. Facility with which they can be fitted on board ship.

4. Less danger in the event of an explosion.

5. Capacity for being greatly forced.

6. Reduced weight.

The disadvantages are —

1. Increased susceptibility to irregular feeding.

2. Increased susceptibility to irr^ular stoking.

3. Increased susceptibility to rapid fouling.

4. Difficulty of internal cleaning.

A distinction is made between water-tube boilers with large and with
small tubes. The Belleville, Diirr, and Niclausse boilers belong to
the first class. To the second class belong the Yarrow, Normand,
Thornycroft, Schulz, and other similar boilers, none of which (except
the Yarrow) differ from the general type by any very marked charac-
teristics. A description is given below of some typical examples
of water-tube boilers. A few general remarks, which apply to the

♦ Compare the remarkable essay in Marine Rtutdschau^ 1901, page 524, **The
Water-tube Boiler Question in the German Navy," by Marine Oberbaurat Kohn von
Joski; also Schiffbau^ 1901, iii., page 129 e( seq,^ ''Further Contributions to the
Water-tube Boiler Question in the German Navy," by ZUblin.

STEAM BOILERS. 513

construction of them all, will be found in the description of the different
systems.

For dimensions of grate surface, see Table No. 59, page 469.

For calculation of the heating surface, see Tables of Details of
Boilers in the description of the different systems.

For flue area in the nests of boiler tubes, see page 539.

For circulation of water and sectional area of the headers, see
page 548.

For securing the brickwork for the furnaces, see page 539.

For fitting the tubes into the upper and lower drums, see pages 526
and 539.

For regulating the feed- water, see page 516.

For zinc plates, see page 526.

Material for Water-tube Boilers. — The boiler plates are made of
the same material as those of cylindrical boilers, viz., Siemens-Martin
steel of about 27 to 32 tons per square inch tensile strength for plates
which are not welded or in contact with the fire, and 23 to 27 tons
for those exposed to the fire. Elongation of the former, at least 20 per
cent. ; of the latter, at least 25 per cent. Stress upon the plates about
8,500 pounds per square inch.

Tubes. — These are always seamless or solid drawn, and made of
soft iron or Siemens-Martin steel. They are either drawn without a
seam, or, in the latest practice, are rolled while hot, without a seam,
according to the Mannesmann process, and without being further drawn.
The latter class of tubes are as durable as the seamless drawn tubes.

§ 303. Belleville Boiler. — This boiler consists in the main of
vertical parallel sets of tubes known as " elements," arranged side by
side, which are connected at their lower end with a feed collector, and
at their upper end to a steam drum, both being placed at right angles
to the direction of the tubes (see Figs. 449 to 454). The elements
consist of straight solid drawn steel tubes (from about 6 feet 6 inches
to 7 feet long, and 4 to 4| inches diameter, *2 inch thick) arranged
by means of so-called " junction boxes " to form zigzags. Each tube is
inclined about 4** to the horizontal. The ends of the tubes are fitted into
junction boxes of cast steel or malleable cast iron, about *3 inch thick.
They are screwed into the back junction boxes with a fine thread, and
connected to the front junction boxes by a short pipe and socket, forming
a sleeve or muff coupling. The joint is made tight by means of screwed
rings, which act as lock nuts (see Fig. 449). In each front junction box
there are two small oval holes, closed by a door opening inwards, for

2k

514

MARINE ENC;iNES AND HOILERS.

cleaning the tubes. The lowest junction box of each element has a
conical opening, which is fitted with a nickel ring which forms a joint
with a conical nipple secured to the feed collector. £ach of the upper-
most junction boxes is secured by means of a flange to a short length of
vertical pipe screwed into the steam drum.

Fig. 449.

The feed collector is a horizontal rectangular steel tube about
4 X 4 X ^\ inches thick. The pipes forming the connection to the
lower row of junction boxes are screwed into the top of the collector
side by side. The steam drum l lies above and parallel to the feed
collector. It is a riveted, or welded, mild-steel or wrought-iron tube,
about 20 inches internal diameter, and has cast-steel ends. Pipes

STEAM BOILERS.

515

with flanges, to form the joints with the upper row of junction boxes,
are screwed into the bottom of this drum. To separate the water
carried over with the steam, a complicated system of baffles is pro-
vided inside the steam drum. The feed collector and steam drum lie
in the same vertical plane as the front junction boxes. Between
them, and at right angles to them, are fitted seven or eight vertical

J ,

\\ i|i III ^l^:

Fig. 450.

elements, each consisting of about twenty tubes if there is ' not,
and about fourteen tubes if there is, an " tconomiser." The junc-
tion boxes are supported one on the other (see Fig. 450) by
small lugs. The feed-water is sprayed into the steam drum (through
a valve) and then flows down into cast-steel mud boxes, through
two vertical pipes c, which are fitted to each steam drum. Here

516 MARINE ENGINES AND BOILERS.

it deposits any dirt, which can then be easily drawn off. The water
passes thence to the feed collector, and from there into the tubes forming
the elements. The lowest water level is situated about halfway up the
elements. The lowest rows of tubes sometimes consist of "Serve'"
tubes, f.^., tubes with internal ribs, a cross section of which is given at
Fig. 451.

As the level of water in the elements is subject to great fluctuations,
an automatic feed-water regulator (see Fig. 452) is fitted to all Belle\'ille

boilers. It consists of a tank 1, which is connected
to one of the lower junction boxes through pipe 2,
and to one of the upper junction boxes through pipe
3. A float 5 moves up and down in the tank. If the
level of water falls, the float draws down lever 6, forces
up lever 7 through rod 8, and o[>ens valve 4 in the
feed-water pipe by means of connecting rod 9. The
Fig. 451. arrangement is adjusted by weights 10 and spring 11,

A being the flxed, and b the adjustable weights.
Valves 12 and 13 serve to open connection with the feed- water pipes
14 and 15 respectively. On the outside of tank 1 are fixed the water
gauge 17, pet-cocks 18, and bracket 19 for holding the regulator.

To utilise the heat in the escaping gases of combustion, which are
at a fairly high temperature, and to further cool them, and prevent
overheating of the uptake and funnel, a feed-water heater or econommr
is placed above the boiler in the newer types (see Figs. 449 and 450,
and the photograph Fig. 454). It is constructed in the same way
as the boiler; internal diameter of the tubes = 2| inches. The
junction boxes are malleable castings or cast steel. The rise of tem-
perature of the feed-water in the economiser is 120" to 180^ The
water passes from the automatic feed regulator a'b' (Fig. 450) through
pipe c to the feed-water collector of the economiser AjBiCi, flows
through the elements of the latter, and collects in the cross pipe h,
which unites the upper ends of all the elements. Thence it passes
through a pipe h and the feed valve k to the steam drum l. Between
the boiler and economiser there is a combustion chamber db\ into
which a current of air b^ can be forced, so that the gases may be
thoroughly mixed, and complete combustion ensured. Air is also some-
times forced through nozzles into the furnace immediately above the
layer of coal on the grate, so that the smoke may as far as possible be
consumed. The pressure of this air, which is compressed in separate
air-compressors, is about 20 lb. per square inch.

In the Belleville boiler the steam is almost always generated at a
pressure of about 40 lb. per square inch above the working pressure at
the engines, and is reduced to the latter pressure immediately before

STEAM BOILERS. 517

it enters the engine by means of a rtducing valve (see Fig. 453).
The steam enters the reducing valve from the boiler through the
opening on the left, and, after passing through the slots, passes out
through the opening on the right to the steam pipe and thence to
the engine. At the same time it finds its way into the spaee above the
valve through holes pierced in it. This exerts a pressure on the small
piston, working in the stuffing box, and tends to close the valve.
The upward pressure on this piston acts in a contrary direction to the

B Fig. 452.

springs, which press down the piston through a single-armed lever, and
tend to hold the valve open. If the required tension be given to the
springs by means of the hand wheel and spindle shown on the left in
Fig. 45.3, the throttling of the steam can be regulated at will. By using
a higher pressure of steam in the boilers, a certain small reser\-e of
energy is provided, which is an advantage, considering the small quantity
of water and heat contained in the Belleville boiler. For the effect of
throttling upon the dryness of the steam see page 472 ; for the dimen-

518 MARINE KNIilNKS AXD liOILERS.

sions of boikT casing, grale, &c., sec Figs. 449 and 454. The dimensions
and efficiency of the Belleville boiler are given in Table No. 6.t, Thesr
boilers can be placed either athwart ship or fore and aft.

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520 MARINE ENGINE:S AND BOILERS.

§ 304. Diirr Boiler (see Figs. 455, 456).— This boiler consists
of an upper drum, below which, and riveted to it, is a flat water
space or header chamber. Straight parallel boiler tubes, closed at their
further end, are fitted in the back plate of the water space. The water
space is separated into two i>arts throughout its whole length by a
dividing wall parallel to the tube plate. Into this wall small tubes
are fixed, which are passed into the boiler tubes, and extend almost
throughout their whole length.

The water level is rather above the bottom of the upper drum.
The steam generated in the boiler tubes rises up in the back part of
the water space between the dividing wall and the back tube plate;
while in the front part the water descends, and finds its way into the
inner circulating tubes, inside the boiler tubes.

The waUr space is seamless, having no riveted joints, the metal
being merely welded together, and stiffened with stays and stay bolts
between the holes for the tubes and the caps. The water space is
widened at the top like a wedge to facilitate the circulation.

Both the boiler tubes and the back wall of the water space are slightly
inclined, and are fixed into the back tube plate by forcing them into
conical openings, the axes of which are not quite in line with the axes of
the corresponding tubes. The diameter of the tubes is decreased at the
back, where they fit loosely into a wrought-iron grating. The two out-
side vertical rows of tubes at either side of the boiler are bent round to
the right and left, immediately behind the tube plate, so as to form a
tube wall on both sides of the grate, enclosing the nest of tubes. The
inner water-circulating tubes are fitted into the dividing wall of the
water space with bell-mouthed ends, and are made of thin sheet iron.

The caps which serve to close the holes provided for the insertion of
the tubes are introduced from within, and are pressed into the conical
holes by means of a forged screw and collar, no packing being required
(see Fig. 457). The back end of the tube is also made tight either by
a conical cap on the inside, or a bronze lock nut on the outside ; the
latter arrangement makes a better joint, and is more quickly taken to
pieces (Fig. 457). The upper drum is generally at right angles to the
tubes, as shown in Fig. 455. This is made of mild steel plates riveted
together, and stiffened with stays where it is cut away for the water space.

The superheater or steam dryer consists of several horizontal boiler
tubes fixed into the upper drum, and lying in the uptake. The steam
from the drum is caused to circulate through these tubes by means of
the inner circulating pipes. The dried steam passes back through the
outer tubes into a pipe in the steam drum, and thence to the main stop
valve (see Fig. 455). The circulating pipes receive the wet steam from

STEAM HOILEKS.

522 MARINE ENGINES AND HOILKRS.

the upper drum through a pipe having small slots cut in it. The
grate covers the whole area below the nest of tubes. The comhui-
tioti chamber is lined with firebrick. Baffle plates are fixed above some
of the rows of tubes, so that the flames are first forced backwards
towards the back wall, then to the front, and then again backwards as

Stfetf Vtive

ihey rise through the nest of tubes. By means of iron rods passing
through the holes for the stay bolts, these baffle plates can be shaken
by the stoker and freed from the ashes which settle on them.

S 304a. Dimensions of a Diirr boiler having 3,350 square feet

STF.-''^*

ISOII'V-P'

S __"

4

_Jliiii=i=^'*-T*

524 MARINE ENGINES AND BOILERS.

of wetted heating surface. Area of the superheater, 176 square feet ;
grate area, 80 square feet for a working pressure of 190 lb. per square
inch.

Upper Drum, — Diameter, 45 inches; length over all, 13 feet 7
inches ; thickness of plate, f inch.

Water Space. — Width at top, 11 '8 inches; at bottom, 7 '8 inches;
height, 7 feet 6 inches; breadth at top, 11 feet 5 inches; at bottom, U
feet 10 inches; thickness of plate, ^ to 1 inch.

Tubes, — 20 rows : 16 rows at the top = 427 tubes; external diameter,
3^ inches; thickness, \ inch. Next 2 rows = 56 tubes; external diameter,
3 J inches; thickness, u\ inch. Then 1 row = 28 tubes; external dia-
meter, 3^ inches ; thickness of plate, /^ inch. The lowest row = 28
tubes, has an external diameter of 3 J inches; and thickness, '2 inch. All
the tubes are 7 feet 4 inches long. The tubes forming the water wall are
3J inches external diameter, -128 to '2 inch thick, 7 feet long. Tubes of
the superheater =44; external diameter, 2 J finches; thickness of plate,
•138 inch; length, 6 feet.

Centre division plate of the water space, \ inch thick ; distance from
front tube plate, 4| inches.

Length of grate — ^ feet 6 inches: 6 firedoors, each 15 inches wide
X 14 inches high.

Height of boiler over all, 15 feet ; breadth of front, 15 feet 1 inch ;
depth of boiler, 8 feet.

Lowest water levels 13J inches below the centre of the upper drum.

Boiler Casing, — Thickness of inner plate, '125 inch ; of outer plate,
■04 inch ; space between them, 2 to 2 J inches. For dimensions and
results of actual boilers see Table No. 64.

STEAM BOILERS.

525

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526 MARINE ENGINES AND BOILERS.

§ 305. Yarrow Boiler (see Figs. 459 to 461a).— This boiler
consists of a riveted upper drum of Siemens-Martin steel, which is
connected to two lower drums by straight tubes. In the larger boilers
the lower drums are riveted, but in smaller boilers, where they cannot
easily be examined internally, they are made in two halves and bolted
together.

External diameter^In the older and smaller types, 1 to l^ inch,
of tubes ) In the newer and larger types, up to If inch.

Thickness of Tubes : — For an external diameter of 1 inch, about ^jf
inch for the inner rows of tubes, about ^v ^"^^ ^^^ '^^
outer; and ^ inch with an external diameter of If
inch. The tubes are secured to the upper and lower
drums by expanding and coning the ends (see Fig.
458). To stiffen the tube plates of the lower drum a
Fig. 458. ^^^ tubes in the nest may be expanded into holes, the
sides of which are slightly ribbed (see Fig. 467), ifTlhe

same way as is done with bent tubes. They serve also to a certain

extent as stay tubes.

Dimensions,

Torpedo-boat (see page 41*). — One boiler. Heating surface, 1,420
square feet ; grate area, 24*5 square feet ; length over all, 9 feet 6 inches ;
breadth over all, 7 feet 9 inches ; height over all, 5 feet ; length of grate,
6 feet 3 inches ; width, 4 feet ; internal diameter of upper drum, 3 feet ;
thickness of metal, "4 inch ; external diameter of tubes, 1 inch \ number
of tubes, about 1,600; lower drum, 9 by 16 inches; pressure above
atmosphere, 185 lb. per square inch. The tubes, which are seamless
solid drawn steel, are expanded into the upper and lower drums. The
tube plates where the tubes enter are made \\ inch thick for I inch
tubes, and from If to 2 inches thick for If inch tubes. As is usual in
most boilers of this class, the upper and lower drums are fitted with zinc
plates. These must only be used in the upper drum with very great care,
and in such a manner that, if eaten away, pieces of zinc cannot possibly
fall off. If attention is not paid to this, and especially if, instead of
rolled zinc plates, cast zinc plates are used, which fall to pieces verj-
easily, one or more of the tubes may get stopped up, become red hot,
and burst. There is usually no automatic feed-regulating valve, the feed-
water being admitted to the lower part of the upper drum by means of
an internal feed pipe provided with slots. In modern practice feed-water
regulators, as described at page 542, are occasionally fitted.

* Compare SchiffbaUy iii., No. 5, ** Further Contributions to the Question of
Water-tube Boilers," by Von ZQblin.

STEAM BOILERS.

MAKINK KNGINES AND ROILKRS.

STKA.M UOILKK

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STEAM BOILERS. 531

The circulation of water is produced by the steam and water ascend-
ing through the tubes exposed to the greatest heat, while the water
descends from the upper to the lower drum through those tubes which are
least in contact jwith the flame. The grate is placed between the two
nests of tubes, and the flames And their way up sideways through them.

Consumption of coal per 'i^ In larger ships, from 1*8 to 2*3 lb.

i.H.P. per hour j In torpedo-boats, about 3*15 lb.

i.H.p. per square foot of|In larger ships, 17 to 21 i.h.p.

grate /in torpedo-boats, 23 to 28 i.h.p.

Ratio of heating surface to grate surface = 45 to 58.
For details of these boilers see Table No. 65.
Recently some Yarrow boilers have been fitted with Howden's forced
draught and air heaters. In these the ratio h : r is taken at about = 50.

Casing of the Yarrow Boiler, — Plate XI Va. shows a set of Yarrow
boilers for a Chilian warship, with casing and uptake. The arrangement
of the boiler fittings can also be clearly seen. On the front of the boiler
casing there are light and easily removable dampers, for withdrawing the
ashes. The smoke box has double walls, the outer shell being isolated
with non-conducting material (asbestos), which is protected by a metal
covering.

Ji 306. Normand Boiler (see Figs. 462 to 466).— This boiler con-
sists of an upper drum, and two lower drums parallel to it, and connected
with it by two nests of solid drawn steel tubes. The grate is placed
between the two lower drums. The tubes are slightly bent, and can
thus expand freely when heated, but the curvature is too slight to per-
mit of the steam collecting in pockets. The two lower drums are con-
nected at one end to the upper by downcast pipes of large diameter,
which are outside the casing, and are intended to promote the circula-
tion of water from the upper drum (into which the feed-water first passes)
into the lower drums. At the other end the two lower drums are
connected by stays or a downcast pipe with branches to each drum.
If the stays happen to come within the casing, water circulates through
them to keep them cool.

The tubes of the outermost and innermost rows are placed so close
together that they form a compact screen or "tube wall." The inner-
most tubes at the front end of the boiler have a space left between them,
from top to bottom. The flames are thus forced to find their way forward
from the furnace through these openings, and backwards through the
nest of tubes to the back of the boiler, where they strike against a vertical
baffle plate, under which they are compelled to pass. At the end of the
nest of tubes they pass out either (a) forwards, in the same direction, or
{b) sideways, or {c) into the uptake through openings left at the top of

MARINE KNCINES AND BOILERa

STEAM BOILERS.

534 MARINE ENGINES AND BOILERS.

the outermost rows of tubes. Various modifications of this arrangement
have been tried.

In Fig. 462 the downcast pipes at the further end of the boiler
are not seen ; but the stays in the front end of the boiler, and the open-
ing in the outer tube wall below the funnel, are clearly shown. Here the
flames enter the tubes at the back end of the grate.

Dimensions of a medium-sized boiler : Upper drum, about 3 feet
3 inches in diameter; lower drum, about 1 foot 8 inches; length of
upper drum, about 10 feet; number of tubes, 1,000, having 1^ inch
external diameter and \ inch thickness (corresponding to about 3,000
square feet heating surface, and 70 square feet grate area). The mean
level of water reaches to about one-quarter of the way up the upper drum.
Sometimes the length of the grate is as much as 8 feet 3 inches and
more; the amount of space above the grate is so large that, in spite
of its considerable length, it is more easily stoked than any other type
of boiler.

For details of these boilers see Table No. 66.

Normand Boilers of the Russian Cruiser ^^Bogatyr^^ (compare Table
No. 66). — This ship is fitted with sixteen Normand boilers, having a
total heating surface of 50,700 square feet and 985 square feet grate
area. Boiler pressure, 256 lb. per square inch. Of these sixteen
boilers, four have each 70 square feet grate area, and twelve have
each 58 square feet grate area. Each of the boilers has a heating
surface of 3,170 square feet. There are 972 tubes to each boiler, each
tube having an external diameter of 1 J inches, and thickness ~ inch.
Overall length of each boiler, 1 1 feet 9 inches. Overall width of the
larger boilers, 14 feet 3 inches; of the smaller, 12 feet 8 inches. Total
height, including the steam dome, 14 feet 9 inches. Water space per
boiler, 140 cubic feet. Steam space, 70 cubic feet. Each boiler has two
furnace doors. Length of grate, 7 feet 4 inches. Width of grate in the
larger boilers, 9 feet 6 inches ; in the smaller, 7 feet 9 inches. Internal
diameter of the upper drum, 3 feet 3 inches ; thickness, 0*70 inch.
Internal diameter of each of the lower drums, 18 inches ; thickness, 0*94
inch. Internal diameter of each of the two downcomers, 10 inches.
At the back end of the boiler a hollow tie connects the upper to each of
the lower drums. A number of dampers are placed in the sides of the
ash pit, which admit a plentiful supply of air, and open inwards, so that
they are opened automatically by the pressure of air in the stokehold, and
close in the same way if any steam escapes from the tubes into the furnace.
The steam is taken off through an internal steam pipe op>ening into the
dome, which prevents any priming water being carried off w^ith it.

The flame first enters the nest of tubes near the fire door, passes

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536

MARINE ENGINES AND BOILERS.

through over the tubes in a horizontal direction, and on reaching the
end is directed downwards by a baffle plate; finally it passes out,
partly through an opening in the tubes, at the upper end of the outer-
most row, and partly horizontally at the end of the boiler throughout the

Fig. 464.

whole length of the tubes. The gases of combustion are led through
the uptakes into three funnels, the front one being 8 feet 2 inches, and
the two other funnels 9 feet 2 inches in diameter. Height of the
funnels above the grate, 72 feet.

The construction of the boiler is shown in Figs. 462-466 ; the con-

STEAM BOILERS.

537

struction of the casing and lagging (double metal walls with asbestos
cloth laid over them, and secured with iron bands) and arrangement of
the boiler mountings is seen in Fig. 463.

Fixing the tubes into the plates of the upper and lower drums is

^aaSma^Simmmiimmmiimmmmimmmmi^

Fig. 465.

effected, as in all boilers with bent tubes, by expanding them into holes,
and bell-mouthing the end which projects over. In many cases the
holes have spiral grooves, about -^^ inch deep. The edge of the hole
must be slightly rounded off, where the bell-mouthing takes place, so
that there may be no sharp edges.

MARINE ENGINES AND BOILERS.

STEAM BOILERS.

539

The brickwork is secured to the boiler casing by iron bolts, with T-
shaped heads, which are secured at their other end to the casing by
nuts or wedges (Fig. 468).

Unrestricted Area of Flues, — The unrestricted opening left to admit
the gases of combustion into the nest of tubes is about ^ to | of the

Fig. 467.

Fig. 468.

grate surface. The unrestricted area for the passage of gases through
the nest of tubes, and their exit from the nest, should be not less than
1 and not more than \ the area of the grate ; while the area of the
uptake and the funnel is generally about \ that .of the grate.

§ 307. Small Tube Water-Tube Boilers.— The various types

belonging to this series of water-tube boilers do not markedly differ from
each other, and may therefore be treated together.

§ 308. "Daring" Type Thomycrofk Boiler.— One of the oldest

of the better known types is the "Daring" type Thornycroft boiler
(see Fig. 469). A grate is situated on either side of the central lower
drum. The flames enter the inner nest of tubes at the bottom along
the whole length of the grate, and pass out at the top along the whole
length of the nest of tubes. On the outer side of each grate the com-
bustion chamber is bounded by a so-called "water wall," ^^^ by a
double row of tubes, which are bent together so as to form a single con-
tinuous wall ; these tubes open at their lower end into a lower drum of
very small diameter, which is connected through a pipe with the central
lower drum. The water, which is fed into the upper drum, passes
down into the central lower drum, through a row of tubes in the middle,
between the two inner nests of small tubes. Except for these last-
named tubes, all the water tubes discharge into the steam space above
the water level of the upper drum.

§ 309. " Speedy " Tjrpe Thornycroft Boiler.— Another older

type still is the "Speedy" Thornycroft boiler (see Fig. 470). On
each side of the grate is placed a nest of tubes, all the tubes of

540 MARINE ENGINES AND HOILERS.

which discharge into the steam space of the upper drum. The
inner and outer tows of each series fonn a water wall, except thai
in order to allow free ingress and ^;ress to the gases into and froin
the nest of tubes, the inner wall is left open along the whole length
of the grate at the bottom, and the outer row along the whole
length at the top. The circulation of water from the upper dniin—
into which the feed-water is fed — to the lower, is by means of down-
comers at either end of the boiler. This method of constiuction
has now been abandoned, because the steam collecting above the
openings of the small tubes into the upper drum, renders them

I

liable to overheating, and also when the boiler is laid by, the uppc
curves of the tubes form "air-pockets," and injury to the boiler is
the result.

5 310. Thoniycroft Boiler.— For other types of boilers bebnginj;
to the systems mentioned in g 307, see Figs. 471, 472, and also Figs-
473, 474. In the larger sizes, Figs. 471, 473, these boilers have three
lower drums; the middle one is connected to the upper dram by
two nests of tubes, and the two outer drums each by a single nest of
tubes. At either end of the upper drum a lai^e downcomer pipe leads
to each of the two outer lower drums, while the centre one is connected

STEAM BOILERS.

541

to the upper drum from end to end by a row of slightly curved down-
comer pipes. As none of these pipes are exposed to the heat of the
fire, they are able to carry down the water from the upper drum to the
lower part of the boiler. Between the lower drums are the two grates.
The two innermost and the two outermost rows of each set of tubes are
bent so as to form tube-wall rows, with openings at the bottom on the
fire side rows, and at the upper ends on the rows nearest the uptake.
To make these tubes touch each other, their diameter is increased to

Fig. 470.

about 1 J inch, whereas the diameter of the other tubes is only 1^ inch.
The flames, therefore, can only penetrate the nest of tubes at the bottom,
where they do not touch each other, and pass out at the top of the
** tube wall," where the tubes enter the upper drum. The dischsitge
opening for the hot gases should be as small as possible, and placed
immediately below the funnel in the middle of the boiler. The boiler
shown at Figs. 471 and 472 is designed to give 1,750 i.h.p.; heating
surface of tubes, 3,600 square feet ; grate area, 63 square feet ; pressure
above atmosphere, 230 lb. per square inch.

542

MARINE ENGINES AND BOILERSl

In the smaller sizes (see Figs. 473 and 474) tfacre are only tro
lover drams, from each of which a nest of tubes leads to the upper dnuc
The grate is situated between the two nests of tubes. The inner and
outer rows of both nests of tubes fonn, as before, continuous tube vaCs.
which only afford a passage to the hot gases through the inner rows a: |
the bottom, and through the outer rows at the top. The tubes, as

Fig. 471.

before, are solid or seamless drawn steel, galvanised on the outside,
and expanded into the upper and lower drums ; the tube ends are bell-
mouthed on the inside of the drum.

The boiler for a torpedo-boat, shown in Figs. 473, 474, has a
heating surface of 1,150 square feet, grate area of 21 square feet, and a
working pressure of 146 lb. per square inch. The water level is about
one-quarter of the way up the upper drum.

The feed is regulated by a solid drawn W.I. float inside the upper

STEAM BOILERS.

543

drum, which rises and falls with the water level, and thus actuates a
horizontal rod through a bell crank lever. The rod passes through the
front of the boiler, and actuates a slide or lift valve, throttling the
admission of feed-water if the water level is high, and admitting it freely
if the water level is low. The casing is composed of double plates (see

(33) r/fn

sff ctrxumfi

Centrt Drums S' 7
..Wing Drums II' l'

Fig. 472.

Figs. 473, 474), between which asbestos is packed, the outside being
lagged with non-conducting material.

S 311. Recent Thornycroft Boilers. — In recent boilers of this
type, in the construction of which Schulz has rendered considerable
service, the flue gases, in order to more fully utilise the heat contained
in them, are led through specially constructed passages. Such a boiler
is shown at Fig. 475. The flames enter the central nest of tubes

544

MARINE ENGINES AND BOILERS.

.9 .P

STEAM HOILERS.

o4o

to

2 M

546

MARINE ENGINES AND BOILERS.

Table No. 67.—

Name of Shifx

I.H.P.*

Maker's
Figures.

Type of
Engine.

Steam Pressure,
pounds per sq. in.

At

Engine.

In
Boiler.

230

C 3

H

—

7,000

4

sq. ft '
14,40f)

H.M.S. "Proserpine,"
third-class cruiser

7,000

2 en^nes,

3 cylinder
tripleexpansion

8

250

300

19,475

"Missouri"and"Ohio,»
American battleships,
1900

16,000

2 engines,

4 cylinder

tripleexpansion

12

— ^

250

53,2a)

" Novik," Russian cruiser,
1901

17,000

3 engines,

4 cylinder
tripleexpansion

12

^■^

256

49,5(JII

**^gir," Geiman coast
defence vessel

4,980*
3,680*

2 engines,

3 cylinder
tripleexpansion

16,140
16,140

••Niobe," small German
cruiser

8, 110*
3,890*

2 engines,
tripleexpansion

—

—

—

21,730
21,730 ,

"litis"

1,378*
1,293*

Triple
expansion

—

—

4,000

4,00(1

1

H.M.S. "Handy,"
torpedo-boat destroyer.

4,600

2 engines,

3 cylinder
tripleexpansion

3

^-

185

(max.

210)

9,501)

"Eber," gunboat, Ger-
man Navy ; built 1903
at the Vulcan Works,
Stettin

1,300

2 engines,

3 cylinder
tripleexpansion

4

about

180

lb. per

sq. m.

185

3,»30

1

•* M," small cruiser, Ger-
man Navy ; built 1903

10,000

2 engines,

3 cylinder
tripleexpansion

10

200

215

30,iJ00

" Preussen," battleship,
German Navy; under
construction (1903) at
the Vulcan Works,
Stettin

16,000

3 engines,

3 cylinder

tripleexpansion

8
water
tube

6
cylin-
drical

185
185

192
192

34,400,

1

lo,0C«">

"Mecklenburg," battle-
ship, German Navy ;
delivered (1903) by
Vulcan Works, Stettin

13,6a)
14,350*

3 engines,

3 cylinder

tripleexpansion

6

watei
tube

6
cylin-
drical

185
185

185
185

29,0fP0

i5,a()0

1

* The figures marked with an asterisk are the results given on trial trips.

STEAM BOILERS.

547

Tkornycroft^ Schulz^ and similar Boilers.

1^

■
X

• •

s
,2

57-2
55-7

0-48
0-36

27-8
20

Consumption
1 5: . of Cool per
' ■ i.H.p. per
hour.

Consumption
^ ' of Coal per sq.
ft. of Grate
per hour.

1 0

inchei>.

3-2

.\uthority.

aq. ft.
250

350

Communicated by the

builders.
Engineering, 1899, iL,

pp. 216-218.

960

55-4

0-30

16-6

—

Maritu Rundschau^
1901, p. 449.

860

57*5

0-34

19-8

—

—

Ibid,^ p. 459.

273-7
-273-7

59
59

0-30
0-22

18
13

•156

22-5

—

Ibid,, pp. 556-559.

415-8
415-8

52-3
52-3

0-37
0-18

20
0-93

•173

17-4

—

Ibid.

81-7
81-7

TjOI
501

0-33
0-31

16-8
15-8

•146

25

—

Ibid,

188

50-4

0-48

24-3

—

30

Engitueringy 1896, i.,
p. 246.

75

51-4

•33

17-2

—

max.

allowable

2-56

580

50-8

-33

17

max. all.
256

650

52-7

—

—

—

max. all.
2-56

450

33-4

—

—

—

—

max. all.
0-47

670

—

-16

at

13 t

Schiffbau, 1903, No. 2.

300

1

9610

I.H.p.

0-35 t

t When developing 14,350 i.H.r.

548 MARINE ENGINES AND BOILERS.

at the bottom along its whole length, and turn in an upward direction.
After passing the centre of the furnace, they pass in a downward
direction under a solid wall of tubes in the centre of the outer nest
(extending the whole length of the grate, and from the central upper
drum almost to the wing drum in the vertical direction), and thence
upwards to the funnel, as indicated by the arrows. A row of down-
comers is placed in the middle between the two walls of tubes, inside
the central nest. Sectional area of these downcomers about -03 square
inch per square foot of heating surface. The lower wing drums are
either provided with special downcomers, or connected to the central
lower drum, by horizontal tubes. For particulars of boilers of this type
see Table No. 67. The boiler casings are made of thin plate, and
stiffened on the front and back of the boiler with angle irons. They
are provided with circular openings for cleaning the tubes. Wherever
the flames can strike against the casing, it is protected by a fire-brick
lining. At a greater distance from the grate asbestos millboard is used.
Reckoning from within outwards these parts are covered first with
asbestos, then the plating, then an air space, then plating again, and
lastly a covering of asbestos cloth, held in place by iron straps. Where
the casing is protected by a wall of tubes, the inner asbestos millboard
is omitted.

SECTION V.
SMOKE BOX, FUNNEL, AND BOILER LAGGING,

(For arrangements of induced and forced draught see Figs. 481 to
483 and 486 to 489.)

§ 312. The Smoke Box is generally made with double walls.

Thickness of inner plate, in small light vessels jj\ to \ inch, in
larger vessels ^ to J inch, in exceptionally large heavy steamers up
to \ inch. Thickness of outer plate about /^ to \ inch ; in lighter
ships not more than yV inch. Air space between the plates, 2 to 4
inches. The plates are connected to each other by bolts and ferrules.

For cross-sectional area of smoke box see page 459.

Doors are fitted in the smoke box for cleaning the tubes ; there is
generally one for each nest of boiler tubes. To prevent the cold air
streaming in through the open 6re doors, when cleaning the fires, the
uptakes from the different furnaces, or in cylindrical boilers from the
different groups of tubes, are frequently divided in the smoke box, and the
draught in them controlled by separate dampers. The doors are hung
on stout hinges and closed with strong catches ; they are not only made
double, but have baffle plates on the inside, against which the hot gases
impinge on issuing from the tubes. Thickness of the centre plate, from
about ^ to ^ inch ; of the outer plate, tht ^^ ^ inch ; and of the inner,
if\ '^ 1 inch. The smoke-box doors open either sideways or upwards.
In the case of large funnels, iron ladders are provided for the inspection
of the uptake and funnel, access being obtained through carefully closed
manholes. Compare construction of uptake and funnel for a large war-
ship, at the end of this section, § 316.

§ 313. Funnel. — In merchant ships the cross section of the funnel
is almost always circular, in war vessels it is often elliptical. (For
sectional area see page 459.) The height depends on the construction of
the ship. (In the largest fast steamers the height of the upper edge of
the funnel above the grate is about 100 feet.) The funnel proper is
either connected direct to the uptake, or to the funnel seating and
through it to the uptake. The cross section of the funnel seating is

550

MARINE ENGINES AND BOILERS.

generally square at the bottom, and octagonal or round at the top (see
Figs. 476 and 479). Thickness of the funnel plate at the bottom
(below the upper deck), /^r to } inch ; further up (above the deck), I to
f\ inch; at the top, uV to ^\ inch. From the uptake to the deck
coaming the funnel is invariably protected by a casing (thickness, ^% to
\ inch ; distance between it and the funnel, from 2 to 4 inches), and
also above the deck coaming one is usually fitted (of the same thickness ;

Fig. 4

1 1'

Fig. 476.

Fig. 478.

distance from the funnel, 4 to 6 inches). The casing is generally smooth-
riveted on the outside, with inside butt straps, and the funnel is similarly
riveted when there is no casing. Between the casing and the funnel a
wide space is often purposely left to make the funnel look bigger. In
warships this space is often closed at the top by a metal hood.

§ 314. Fixing of Funnel.— The expansion of the metal due to
heat must here be allowed for. The uptake is always supported on the
boiler, and in large ships the funnel is supported from the funnel

H

STEAM BOILERS.

551

Fig. 480.

552 MARINE ENGINES AND BOILERS.

casing, being secured above the uptake by means of angle-iron brackets.
(In large installations it is supported on angle-iron brackets resting on
similar brackets bolted to the funnel casing, see Fig. 476). To allow
for the expansion of the uptake between the boiler and the base of the
funnel, the uptake below the latter is sometimes cut through, and a
double butt-jointed ring is fitted round the part where it is cut, in which
the funnel and the casing can expand (Fig. 480). The upper part of
the funnel either slides in the casing itself, if the latter is bolted to
the deck coaming (Fig. 478), or it expands together with the casing
and apron (see below) out of the ring or casing, which is bolted to
the deck coaming (Fig. 476). Where the funnel has no casing (Fig.
477), it expands with the apron through the deck coaming.

To prevent water finding its way down the funnel into the boiler-
room during bad weather, the hole in the deck casing into which the
funnel fits is often protected by a metal apron secured to the funnel or the
casing, or occasionally the casing itself may be enlarged at the bottom
to fill up the hole completely (Fig. 478). A short distance below the top
of the funnel, one, or sometimes two, strong rings are riveted on, to
which shackles for the guy ropes are fitted. If there is a casing, the
outer ring is firmly bolted to the inner casing by means of bolts and
ferrules. The stays, which consist of chains or wire rope, are held with
lanyards to the body of the ship, and must be loosened as soon as the
funnel gets hot.

§ 315. Funnel Dampers. — At the bottom of the funnel a damper
is provided for regulating the draught. If there are several boilers
discharging into a common funnel, the latter is frequently divided up
by longitudinal partitions from top to bottom, so that the smoke from
each boiler passes up through a separate channel. In this way the
draught is not affected if any one boiler is not working, flach of these
divisions has a damper of its own. The smoke from the donkey boiler
is also led into the main funnel, but in a separate pipe, and carried
up to the top of the main funnel so as to improve the draught when
in port, and allow the main funnel to be overhauled. The blow-off
pipe from the safety valves is generally carried some way up the funnel,
either in front or behind. To facilitate inspection of the funnel, an
iron ladder is generally fitted inside it. In small river steamers some
arrangement is usually provided for lowering the funnel, corresponding
counterweights being provided.

§ 316. The uptake and funnel for a war vessel of the large

cruiser or battleship type are constructed more or less on the following
lines : —

STEAM BOILERS. 553

Uptake, — Each boiler has a separate uptake carried up to between
decks, and fitted with a separate damper, so that, if desired, any boiler
may be isolated. There is a separate damper for each combustion
chamber, but these dampers must not impair the free access to the
tubes. In the bent parts the thickness of the plate is about /^ inch,
elsewhere about /^ inch. Where the uptake passes over those parts of
the boilers which are in contact with the steam, an air space of at least
2^ inches is provided. The lower portion of the uptake is bolted to
the boiler, and can expand into the lower part of the casing between the
armoured decks. The upper part of the uptake is supported between
decks by brackets or slings, and can expand between the funnel and
the armoured upper-deck. When the uptakes are inclined or at an
angle it is sometimes necessary to support or suspend them. The
hatchways for the uptake and funnel in the armoured deck are pro-
tected with armoured gratings. The deck coamings for the uptake and
funnel, which consist of deck beams, or vertical plating, are fitted with
brackets to take the armoured gratings. The uptakes are accessible
through doors opening inwards.

Funnels, — The thickness of the funnel plates is about /„ inch at
the bottom, and ^ inch at the top. They are provided inside with iron
stiffening rings. The funnel stays are provided with adjustable shackles,
fitted with bronze nuts, to secure the funnel, which is generally attached
to the armoured upper deck.

Uptake and Funnel Casings. — Both uptakes and funnels are fitted
throughout their length with a casing about | inch thick. The casing
plates are jointed by means of inside butt straps. The casings are
either fitted with removable doors, or are partly bolted together so that
the outside of the uptakes may be accessible for cleaning. The spacing
between the uptake or funnel and the corresponding casing is generally
from 3 to 4 inches. From the deck upwards the spacing between the
funnel and casing is generally about 12 inches. The annular passage
thus formed is utilised to carry away the hot air, and allows it to pass
out below the apron. With water-lube boilers it is best to have
dampers below the armoured deck, by means of which, with forced
draught, the exit of air from the boiler-room up the annular space round
the funnel can be regulated. These dampers are so arranged that
they can be worked by hand from some convenient [)osition.

The casings are made air-tight at their lower end, where they join
the uptake. To regulate the pressure of air in the boiler-room, and
the ventilation of the space between the uptake and casing, slides or
dampers are often placed at suitable points in the lower part of the
casing. When blowing steam through water-tube boilers, to clean the

554 MARINE ENGINES AND BOILERS.

external surface of the tubes, these dampers must be carefully closed.
The casing of the uptake is lagged with asbestos mats secured in position
by outside plating, or by means of iron bands. When the bulkheads
round the uptake form at the same time a part of the ventilating shaft,
they must be lagged with non-conducting material. The upper part of
the funnel casing is often fitted with a ring of round iron about 6 inche.
from the plating, which is intended to facilitate painting the funnels,
putting on the covers, &c. Iron rungs are frequently fitted on the
outside of the funnel casing, to form a ladder from the deck to the top
of the funnel.

Funnel Covers. — Each funnel is generally provided with a galvanised
iron cover made in sections.

J$ 317. Boiler Lagging. — In cylindrical boilers about two-thirds
of the upper circumference is covered with a non-conducting material
(fossil meal, preparations of cork, &c.) and cased in plating jV ^i^ch thick.
The lower portion of the boiler shell is usually covered with an easily
removable covering or lagging (such as asbestos fibre, felt, strips of
cork, &c.). Felt should only be used if the pressure in the boiler does
not exceed 110 to 140 lb. per square inch, otherwise it is liable to get
charred. In water-tube boilers the boiler casing is made double, the
inner plate being from /^ to \ inch thick (thinner for torpedo-boats) ;
next this is a sheet of asbestos or an air space % to 2 inches thick:
then another plate from /^ to ^\ inch thick ; and lastly a layer of
non-conducting material or asbestos mats. (Compare water-tube boilers.)
Where the flame impinges directly on the boiler casing, the inner plate
is further protected with asbestos millboard.

SECTION VI.
FORCED DRAUGHT.

§ 318. General Remarks. — Forced draught may be divided
into —

1. Draught produced by sucking out the gases of combustion,

or "induced draught."

2. Draught produced by creating an air pressure in the ash-pit,

or " the glosed ash-pit system," as in Howden's system.

3. Draught produced by creating an air pressure in the stoke-

hold, or "the closed stokehold system."

S 319. Induced Draught (Ellis' and Eaves' system, see Figs. 481
to 483). — This system is generally so arranged that natural draught can
also be used. The gases of combustion are sucked out of the smoke box
through a special tube ss, and then forced through the tube d into the
funnel by means of a fan. A vacuum is thus formed below the fan, and
the air for combustion is caused to flow rapidly through the fire. Before
being led into the ash-pit, the air is heated in an air-heater placed
in the smoke box. This apparatus, which is now usually arranged
vertically, consists of a series of parallel pipes (about 2| inches internal
diameter) through which the hot gases from the combustion chamber
pass on their way to the smoke box. The air enters the heater v at the
top, through dampers kk, circulates round the pipes, and passes out
through passages w to the fire f. The main body of the air enters the
grate (see Fig. 484) through the horizontal dampers, and some through
the damper and slots above the grate. The ash-pit doors are perforated
with holes, so that a little cold air may be admitted, to keep down the
temperature of the grate. To prevent too much cold air reaching
the smoke box while cleaning the fires, it is often partitioned off, and
the gases from each fire ascend through a separate passage, which can
be separately controlled by a damper. (In Fig. 481 these dampers,
which lie immediately above the air-heater, are not shown.) In double-
ended boilers the common combustion chamber is divided in two by a
firebrick wall.

556

MARINE ENGINES AND BOILERS.

Fig. 481.

STEAM BOILERS.

007

Fig. 482.

558

MARINE ENGINES AND BOILERS.

"Serve" tubes (see page 516), which allow of a considerably
smaller ratio of heating to grate surface, are sometimes used ; if the
tubes are of the ordinary kind, " retarders," />., spiral coils composed of
metal strips, are often fitted into the tubes. The construction of the
boilers is the same for induced as for natural draught.

For plain tubes with retarders, the ratio of

heating surface ^f^^^^O^^^g
grate surface

For " Ser\'e " tubes, the ratio of

heating surface ^ f^^^ 23 j^ 35
grate surface

Fig. 483.

Serve tubes may thus be made much shorter than " the ordinan' "
plain tubes, but they have the disadvantage of being very difficult to
clean, and for this reason they have not been fitted so frequently of
late years.

Combustion per square foot of grate area per hour (with moderate
rates of forcing) —

Fast steamers and mail steamers, 23^ to 25^ lb.
Cargo and passenger boats, 20^ lb.
Cargo steamers, 17 J to 18 J lb.

STEAM KOILERS. 559

These are the figures recommended by the patentees, but ihey
are generally exceeded in actual practice.

Temperature of the gases in the fan, 430^ to 530* Fahr.
Temperature of the air after passing through the heater, about

300° Fahr.
Vacuum in the fan (with high rates of forcing), from 2 to 2 J inches

of water.
Vacuum above the fires, from 0*5 to 0*8 inch of water.
Vacuum in the ash-pit, from 0*25 to 0*4 inch of water.

Diameter of the fan up to 8 feet 6 inches ; maximum number of
revolutions, with high rates of forcing, from 250 for very large fans, up
to 40Cf per minute for smaller fans. The fan casings are supported by
the boilers, and connected to the uptake (for the supports see s/, Fig.
482); the engines m for driving them are situated between decks
(Fig. 483). As the fans tend to rise with the casing as it becomes
heated, a universal joint must be provided in the shafting. Com-
pound or two-cylinder engines are used, and one often suffices to
drive several fans. The engines are generally completely enclosed
to protect them from dust, and to keep them efficiently lubricated.
Air is usually admitted on one side of the fan only : the fan centre is
generally made of cast steel, the thickness of the side plates is from v^.r
to -./j inch, and that of the vanes from ./^ to ^ inch. Number of
vanes, 10 to 16, but half of them are frequently short, and fixed to the
periphery of the fan. Fig. 485 shows a fan 7 feet in diameter, driven
by a compound single-acting engine. Diameter of h.p. cylinder, 6^
inches; of l.p. cylinder, 11^ inches ; stroke, 7 inches. For further details
of these engines see § 275 ^/ se^,

§ 320. Howden's System of Forced Draught (compare S 274).

— I. H.p. generated per square foot of grate surface : for fast steamers,
working for short periods, 16 to 17; for longer voyages, with inferior
kinds of coal, 14 to 15. Ratio of heating surface to grate surface, from
38 to 42 to 1. The higher the rate of forcing, the shorter the life
of the plant. For arrangement see Figs. 486 to 489. The fan v is
driven by the steam engine m, attached to the fonvard bulkhead of
the engine-room. The air is led into the stokehold through a passage
K, in which a damper is fixed at the place where it passes through the
engine-room bulkhead. The passage opens into a cross channel or
collector Q, which serves to supply the after furnaces with air ; from this
another passage k^ passes to the front, and also opens into another
cross channel Q supplying the forward furnaces. From these passages
the air passes through the dampers e into the air-heater r, and circulates

MARINE ENGINES AND BOILERS.

STEAM BOILERS.

561

2 N

562

MARINE ENGINES AND BOILERS.

Fig. 486.

Looking on to the Forward
Bulkhead of Engine-room.

Fig. 487.

STEAM BOILERS.

563

Fig. 488.

After Stokehold looking
Forward,

Fig. 489,

564 MARINE ENGINES AND BOILERS.

round the vertical tubes com[)osing the heater, which are heated by the
hot gases. After leaving the heater, the heated air is led through
passages at the side of the smoke box to the air passage surrounding
the furnaces, and enters the latter through adjustable dampers (in the
same way as shown in Fig. 484 for induced draught). These dampers
are so arranged that they close automatically when the fire doors are
opened, in order to prevent the flames from entering the stokehold, and
injuring the firemen.

Sometimes the fans are fitted above the boilers, and mounted on
girders riveted to the ship ; the engines are then placed between decks ;
a universal joint is here unnecessary, as both engines and fans are
secured to the ship. Diameter of the fans, up to 8 feet 6 inches.
Number of revolutions, about 250 per minute. Pressure of air in the
air passages, about 2 inches of water. Temperature of the hot air, from
about 212** to 300'' Fahr. The engines are double-acting, single-cylinder
or compound.

§ 321. Closed Stokehold System. — In warships with cylindrical,
locomotive, or water-tube boilers, forced draught is almost always*
obtained by creating a pressure of air in the stokehold. The latter is
made as air-tight as possible, and, for purposes of ingress and ^ess,
fitted with an air-lock, />., double doors, one of which must not be
opened until the other has been closed. The fans are sometimes placed
above the armoured deck, and force the air through passages, or direct
from the casing, into the stokehold. They may be fitted in the stoke-
hold itself, or hung from the deck, in which case the air is delivered from
the fan direct into the stokehold. In both cases the fans draw the air
from ventilators fitted on the deck. The speed of the air in the suction
trunk about 20 feet, in the delivery pipe about 30 feet per second. A
damper must be fitted in the suction or delivery pipe of each fan, to
prevent the escape of air through any fan which is not working. The
air must be equally distributed throughout the stokehold, which may be
done, if necessary, by special baffle plates, and the personnel of the
stokehold must not be inconvenienced by the fans.

For the necessary quantity of air, shapes of the vanes, power of fan
engines, &c., see § 275 et seq. External diameter of fan wheel, 4 to 6
feet. Number of revolutions, up to 600 per minute. The air may be
sucked into the fan on one or on both sides. Diameter of the intake,
about 0*7 X external diameter of fan wheel with single suction, and about
0*5 X external diameter of fan wheel with suction on both sides. Width
of fan at periphery, 0-1 3 to 0*1 6 x external diameter of fan wheel. Thick-

* The closed ash-pit system of forced draught, or what is known as " under air
blast/' is now seldom used, apart from Howden's system, on mercantile steamers.

STEAM BOILERS.

565

ness of fan casing plates, jp^ to ^^ inch ; of the vanes, about ^V
The boss may be of bronze or cast steel.

inch.

Pressure of Air, — Even at the highest rates of forcing this is never
above 4 inches (and this only with locomotive and water-tube boilers).
As a rule it does not exceed 2J inches for water tube and locomotive
boilers, and 1 inch for cylindrical boilers.

Fig. 490.

Fan Engines, — These may be single or double acting, one-cylinder,
two-cylinder, or compound engines. " Chandler " engines are largely
used \ they are single-acting, wholly enclosed, so that the cranks work in
an oil bath, and have long bearings lined with white metal. The exhaust
steam is led into the auxiliary exhaust pipe, and the pipe carrying steam
to the fan engines must be fitted with a sufficient number of drain-cocks.
For data of actual vessels see Tables Nos. 68 and 68a.

566 MARINE ENGINES AND BOILERS.

STEAM BOILERS.

567

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SECTION VII.

BOILER FITTINGS AND MOUNTINGS,

§ 322. Boiler Safety Valves. — The dimensions of these valves
must be so calculated that during continuous firing with the stop
valve closed, all the steam generated can escape through them
without the pressure rising more than 10 7o above the working
pressure. The sectional area of the safety valves thus depends upon
the amount of heating surface, the boiler pressure, and the method
of firing (whether with natural or forced draught), and also upon the
construction of the valves themselves. Safety valves must be so
arranged that they lift quickly when the maximum allowable working
pressure has been exceeded, and close rapidly as soon as the pres-
sure has dropped again to the working pressure. In boilers worked
with natural draught, experience has shown that the total internal area
of the safety valves is large enough, if calculated from the data given
in the following table : * —

Table No. 69.

Sectional Area of Boiler Safety Valves,

Pressure above
, Atmosphere, pounds
per square inch.

Unrestricted sec-
tional area in
sq. inches per
1(X) sq. feet of
heating surface

1

70 ' 85

1

100

115
1-24

130

140

160

170
•865

185

•808

200
•78

215
•75

230

1-89 1-62

1

1-46

114

104

•951

•736

* See C. Hartmann, ** Official Regulations respecting the Registration and Testing
of Marine Boilers."

570 MARINK ENGINES AND BOILERS.

If the boiler is worked with forced draught, or the grate axea is
more than j}j that of the heating surface, the total sectioned area of the
safety valves must be correspondingly increased. In ships, the boilers
of which have to be tested according to American regulations, the
total area of the safety valves = ^J^ of the total grate surface. Marine
boilers must, in accordance with regulations, always have at least two
safety valves. The internal diameter of the valves does not as a rule
exceed 4 inches. If the total sectional area required for the safely
valves of one boiler exceeds about 25 square inches, it is divided
between three, four, or more valves of equal size.

S 323. Load on Valves. —The /oad on each valve is now almost
always produced by direct-acting steel springs (see Figs. 492 to 495)
of either square or round section. The dimensions of the springs
must be so selected that with the normal load on the valve they are

compressed to the extent of at least j to — (see below), the maxi-
mum shearing stress upon the material must not meanwhile exceed
about 12*5 tons per square inch.

=v^

Here </= thickness of steel spring of square or round section in inches.

D = mean diameter of coil in inches (centre to centre of coil).

s = load on the spring in pounds.

c = constant, = 8,000 for round steel, 11,000 for square steel, or
for naval work 11,000 and 15,000 respectively.

And also if —

/ = compression of the spring by the load s in inches.

// = number of complete coils.

/= thickness of steel spring of square or round section in inches.

a = constant = 1,400,000 for round steel, 2,000,000 for square
steel.

then/= ^^ULl

STEAM BOILERS.

572 MARINE ENGINES AND BOILERS.

The load s on the spring, that is, on the valve, should be

Where d — internal diameter of the valve in inches.
•08 = width of valve seat (inches).
p = maximum allowable steam pressure above atmosphere in
pounds per square inch.

Only those coils of the spring must be taken into account wiiich are
free and do not touch each other, under normal conditions of loading.
When the spring is loaded, there should be a space of from \ to fJr *"^^
between each of the coils.

§ 324. Safety Valve Casings. — The safety valve casing may be
made of cast iron, bronze, or cast steel. Figs. 492 to 495 show bronze
safety valves as used on warships. The valve seats are always put in
separately, if the casing is of cast iron or cast steel, and secured with
three or four studs. The valve has three or four guiding ribs or feathers.
The seat is generally flat, and not more than '10 inch wide. The spindle
and seat are made as shown in the drawing, so that the escaping steam
may strike against the overhanging rim of the valve, and thus hold
the latter wide open. The bottom end of the valve spindle should
always be situated somewhat below the actual level of the valve seat.
To prevent any jamming of the valve spindle, it should be allowed
ample "play," not only at the bottom of the spring casing, but also
within the spring and the socket at the top.

The cap fitted over the spindle is connected to it by a key or
cotter in such a way that the spindle can move freely, should the ap
become jammed.* It also prevents any interference with the spindle
from the outside. The cotters should be secured by a padlock, or
sealed with wire and lead by the surveyor. In order to twist the
valve easily, the socket should be fitted with a hexagonal nut, or
with two lugs opposite each other. The easing gear, which acts
upon the cap by means of a small lever, is arranged so that it can
be worked from either the engine-room or stokehold. The safety valve
blow-off pipe is carried up alongside the funnel, and has a sectional
area equal to, or rather smaller than (about 0*66) the total area of the
valves opening into it. The condensed water which collects in the
casing above the valve must be carried off by a | to 1 inch drain pipe
into the bilge, or to a tank. This pipe must never be shut off. In
w^arships the safety valve is often united in the same casing with the

* In Fig. 492 the clear space under the cotter has been omitted in error.

/

STEAM BOILERS.

573

main or auxiliary stop valve, so that one hole only in the boiler is
necessary for both (see Figs. 493 to 495). This is especially desirable
in water-tube boilers, where space in the upper drum is much restricted.

Fig. 495.

§ 325. Steam Stop Valve. — The internal diameter of the boiler
stop valve, opening into the main steam pipe, is determined from the
formula —

-v^

074 MARINE ENGINES AND BOILERS.

Where d= diameter of valve in inches.

H = heating surface of the boiler in square feet.
^ = absolute boiler pressure in pmunds per square inch.
e= 126 to 1'35 for boilers with natural draught.
^= 1-4 to 1'62 for boilers with forced draught.

The stop valves for the auxiliary steam pipe are of the same
diameter as the pipes connected to them (see § 253). The valves
should be as strong as possible, and the flanges joining them to
the boiler should be made specially thick and stiffened with ribs.

Fie. 407.

In other respects all that has been
said concerning stop valves in ^ -^I
holds good for these valves also.
F^. 496. If there is no steam dome lo

the boiler, it is advisable to have an
internal steam collecting pipe at the top of the steam space, which
connects with the main stop valve and the safety valve. The lon^
tudinal or cross slits or holes on the upper side of the collecting pipe
have a total sectional area about twice as lai^e as that of the pipes
connected to them. Copper or brass was the material formerly osed
for these pipes. In modem practice, copper, or alloys of copper, arc
no longer admissible for fittings inside the boiler, as they may set up
corrosion, due to galvanic action. The interna! steam pipes are there-
fore made of soft iron.

STEAM BOILERS. 575

§ 326. Feed Check Valves. — In these the valve is not con-
nected to the spindle, and the lift can be frequently adjusted by a
screw spindle and hand wheel. A special stop valve is placed between
the feed check valve and boiler, so that in case of need the latter
may be entirely disconnected from the feed pipes. Both valves are
made of bronze, and either fitted separately, or case in one piece (see
Figs. 496, 497). Where the valve is not connected to the spindle,
separate seats should be fitted. Both the valve and spindle should
be made as strong as possible, and the thread of the valve spindle
should be outside the valve. Sometimes the main and auxiliary
feed check valves are in one casting, which is fastened direct on
to the boiler. In boilers working at moderate steam pressures (up to
about 120 lb. per square inch) a three-way cock is often fitted, leading
to both the feed check valves. To reduce the lift of the valve within
the smallest limits, the diameter of the valve seat is made about *16 to
•31 inch larger than that of the feed pipe connected to it (see § 256).
The feed check valves must be as easy of access as possible. The
pipe inside the boiler connected to the feed check valve is now generally
made of iron, for the reasons given above, preferably of cast iron,
because it is not so liable to rust, and is carried up as far as the lowest
water level, where it is pierced with holes, to distribute the water as
much as possible.

§ 327. Water Gauges. — These are generally tubular and made
of glass, and have an external diameter of from | to f inch ; but
gauges with fiat glasses (Klinger's patent) are also used. The top
and bottom gauge cocks, or sometimes only the bottom one, are often
fitted with self-closing valves (Leser, Dietze, Westphal, Burgemeister's
patents, &c.), to prevent the firemen being scalded, should a glass
break. The cocks are packed with asbestos. It is also necessary to
protect the glass by some arrangement (such as wire netting, strong
plate glass, &c.), so that, if the gauge glass breaks, the splinters may
not injure any one. The length of the glass tubes between the nuts
of the stuffing boxes is from about 10 to 14 inches.

In the Maas water gauge the glass is fixed into a -separate holder
or carrier, and the two can be taken out and fitted in together. The
glass may thus be fixed to the holder at a distance from the boiler.
The holder is secured to the heads of the gauge cocks by means of
cones.

As a rule the gauges are not fitted direct to the boiler, but on to a
gauge column of cast iron or brass, which is connected by copper tubes
with the steam and water spaces respectively. The internal diameter of
these pipes is from | to 2 inches, according to the size of the boiler (the

I

576 MARINE ENGINES AND BOILERS.

Hamburg authorities require a diameter of at least 45 millimetres = say
If inch), and special valves or cocks are often arranged so as to shut
off the connection with the boiler. The height of the water gauge
should be so arranged that the lowest working level of water is not more
than 4 inches above the nut of the lower stuffing box, when screwed
down. To mark it, an index should be fixed on the gauge column
having a plate bearing the words — " Lowest water level." The gauge
column is secured by wrought-iron brackets to the smoke box or
boiler.

§ 328. Pet-cocks or Valves. — These may be fitted either direct
on to the boiler, or if the conditions will not allow of this, they may
be connected to the boiler by straight, thick, wrought-iron pipes. To
admit of readily cleaning the valves or cocks and pipes from scale
and sediment, they must have a straight way through them. The
cocks are often asbestos packed to render them steam tight, and
enable them to be easily worked. The cocks or valves may be
screwed into the boiler, or preferably, attached to it by means of
flanges and f inch studs. If the pipes are of wrought iron, they are
secured to the boiler with a fine thread and lock nut ; a wrought-iron
flange is generally fixed to the other end of the pipe, to which the pet-
cock is fitted. Internal diameter of the pet-cock \ inch to f inch.
The lowest pet-cock is in line with the lowest water level ; the upper
cock is placed about 4 inches above it.

§ 329. Density Cocks or Valves. — A small cock or valve is

fitted to the boiler, by means of which the attendant can draw off
some of the water, and determine its density or percentage of salt.
This valve or cock is so placed that it can be easily reached from
the floor of the stokehold, and is generally made similar to the
pet-cocks or valves on the gauge columns and pipes. Diameter from
J to f inch.

§ 330. Blow-off Cocks or Valves. — In order to keep the pipe
connections as short as possible, this valve or cock is placed close
to the bottom of the boiler, and it and the spindle are usually made
of gunmetal, the thread on the spindle being outside the valve. The
valve must be so fitted that the pressure tends to lift the valve. The
diameter of this valve, and of the pipe connected to it, is generally
made rather less than that of the boiler feed pip)e. The internal
blow-off pipe is made of cast or wrought iron, copper, or brass, and
terminates about | to | inch above the bottom of the boiler.

STEAM BOILERS. 577

§ 331. Scum Cock. — This is also made of brass, and so fitted to
the boiler that the pressure tends to lift the valve, the thread of the
spindle being outside. Diameter of the valve is about 0*65 to 0-8 the
diameter of the boiler feed pipe. The internal scum pipe or trough is
carried along the lowest water level inside the boiler at about the
middle of the surface of the water. The end of the pipe is funnelled
or dish-shaped, to collect the scum.

§ 332. Boiler-emptying Plug. — A plug 1 to IJ inch diameter
is sometimes fitted to the lowest part of the boiler. It consists of a
gunmetal casting, bolted to a stiffening ring, and fitted with a screwed
gunmetal plug.

§ 333. Apparatus for Improving the Circulation of Water

in the Boiler. — In order to avoid unequal heating of the water and
boiler in a cylindrical boiler when raising steam, and to prevent undue
straining, which causes leakage, an apparatus for heating the water
and producing circulation (known as a hydrokineter) is often used. It is
made like an injector, and fitted in a suitable position inside the boiler.
The steam to drive it is obtained from a donkey boiler. The apparatus
sucks up the cold water from the bottom, heats it, and delivers it into
the upper part of the boiler. It is sometimes fitted to the internal feed
pipe of the boiler. The steam pipe to the hydrokineter is generally made
from J to 1 inch in diameter, and to prevent the water in the boiler from
escaping, the pipe should have a non-return valve fitted to the boiler
shell, in which the valve is not fixed to the spindle. The diameter of
the steam nozzle is about {^ to J inch.

§ 334. Summary of Remarks in reference to marine boiler
fittings. All fittings are usually secured with studs to wrought iron or
cast steel stiffening rings riveted to the boiler shell ; the studs must not
penetrate the latter. The flanges to hold the different fittings should
be made as strong as possible and strengthened with ribs, while the
bolts securing them should, if possible, be \ inch thicker than the corre-
sponding ordinary bolts. Bolts less than f inch diameter should never
be used. All fittings must have projections exactly fitting into, and
passing through the boiler plate, to which the internal pipes can be
secured.

§ 335. Regulations affecting Marine Boiler Fittings.

1. German Government Regulations.

Safety Valves, — All ship's boilers must have at least two safety valves.
With the exception of ocean-going vessels, one of these valves must be

2o

578 MARINE ENGINES AND BOILERS.

SO placed that the prescribed load on it can be easily verified from the
deck. The valves must be so arranged that they can be eased when
necessary, and that they blow off the moment the blow-off pressure is
reached in the boiler.

Feed Check Valves, — Every boiler must have a feed check valve,
which is closed by the pressure of the water in the boiler when the
feed is cut off.

Water Gauges. — The level of the top of the combustion chamber
measured athwartships must be indicated on the casing or shell of everf
marine boiler in a clear and permanent manner ; and two water gauges
must be placed on this shell, in the same plane athwartships, sym-
metrical to the centre line of the boiler, and as far apart, right and
left of it, as possible. There must also be a second arrangement for
indicating the water level in the boiler. Each of these fittings must
have a separate connection with the inside of the boiler, or if they have
a common connection the gauge column must have an area of at least
60 square centimetres, say 9 J square inches. If pet-cocks are used, the
lowest of them must be in the same plane as the lowest water level.
All pet-cocks must be so arranged as to have a straight way through,
for cleaning the pipe from scale and deposit.

The lowest permissible water level must be clearly indicated on the
gauge column, and must be at least 4 inches above the top of the
combustion chamber. This minimum distance must be maintained
when the ship is inclined at an angle of 4" to the horizontal, in the
boilers of river and lake steamers, and of 8" in the boilers of ocean-going
vessels. This rule does not apply to water-tube boilers consisting of
tubes over 4 inches diameter, nor to combustion chambers in which
there is no danger of the plate in contact with the steam space becoming
red hot. The danger may, as a rule, be considered as obviated if
the wetted heating surface of the boiler is, with natural draught, at
least twenty times as much, and with forced draught at least forty times
as much, as the grate area.

Pressure Gauges, — Each boiler must have two pressure gauges, on
which the highest permissible steam pressure must be marked. One of
these must be easily visible to the firemen, and the other on deck at a
place convenient for observation, except in ocean-going ships. If there
are several boilers in the ship, connected up, it will be sufficient to
have one pressure gauge on deck, in addition to the one fitted on each
boiler.

Marking. — E^ch boiler must, in accordance with regulations, have
the highest steam pressure, maker's name, factory number, year of com-

STEAM BOILERS. 579

pletion, and the standard lowest water level marked on it in a clear and
permanent manner. These particulars must be shown on a metal plate,
tixed to the boiler with copper rivets in such a way that it is visible
when the boiler is lagged.

Testing. — Every new boiler must, after it is put together, and before
it is lagged, be tested hydraulically with all openings sealed. For
boilers which are not intended for a pressure of more than 5 atmo-
spheres, or 70 lb. per square inch above atmosphere, the test pressure
to be double the working pressure ; for all other boilers, the test pressure
to be 5 atmospheres, or 70 lb. per square inch above the working pressure.
One atmosphere is here equivalent to 1 kilogramme per square centi-
metre, or 14*22 lb. per square inch.

2. German Lloyd's Regulations.

Safety Valves. — Every boiler must have at least two safety valves,
fitted directly to the boiler, with easing gear of sufficient lift. Suitable
methods must be adopted, so that the valves cannot be loaded in excess
of their specified pressure. The safety valve casings must be fitted with
a drain pipe. Superheaters which can be disconnected from the main
boilers must also be fitted with a safety valve.

Main Stop Valve. — Every boiler must be provided with a main stop
valve, to shut off the main steam piping. If there are several boilers,
and only one superheater, the latter must be so arranged that it can be
shiit off from the main steam piping.

Feed Check Valves, — Every boiler must have two feed check valves,
which can be regulated by means of an outside screw ; they must be
fitted direct to the boiler, and arranged to serve the two feed pipes.
The pressure in the boiler should tend to keep the valves shut, and they
must be arranged so that they can be examined without having to empty
the boiler.

BlouH>ff Valves or cocks must be fitted direct to the boiler.

Water Gauges. — Two of these, with an index showing the lowest
water level, must be provided for each boiler. They must be in the
same plane athwartships, at the same height, and symmetrical to the
centre line of the boiler, and be clearly visible. The internal diameter
of the pipes connecting the steam and water spaces must not be less
than If inch in boilers having 320 square feet of heating surface, or
under, and in larger boilers not less than 1| inch. Below the mark
indicating lowest water level there should be at least 4 inches of gauge
glass. With double-ended boilers there should be two gauge glasses

580 MARINE ENGINES AND BOILERS.

at one end, as already mentioned, and one in the middle at the other
end. It must be possible to shut off the gauges easily from the boiler.
Besides these water gauges, there should also be test-cocks or some
other suitable arrangement for ascertaining the water level.

Pressure Gauges. — Every boiler must be provided with two pressure
gauges, upon which the working pressure is marked in red : one
should be in view of the firemen, the other in a prominent and
easily accessible position in the engine-room. Suitable connections
must be provided, to which the test gauges can be fitted.

Marking the Boiler. — Every boiler must have the highest allowable
steam pressure, maker's name, factory number, year of completion, and
standard lowest water level marked on it in a clear and permanent
manner. These particulars must be shown on a metal plate, fixed to
the boiler with copper rivets in such a way that it is visible after the
boiler has been lagged.

Test Pressure. — Boilers which are intended for a working pressure of
not more than 70 lb. per square inch are tested for double the maximum
working pressure (above atmosphere) ; all other boilers are subjected to
a test exceeding the working pressure by 70 lb. per square inch.

3. British Lloyd's.

Safety Valves. — Every boiler is to be provided with two safety valves,
loaded to the working pressure in the presence of the surveyor. If the
working pressure is more than 60 lb. per square inch, the valves must
be loaded to 5 lb. above the normal working pressure. If ordinary
valves are used, their total area is to be at least ^ square inch to each
square foot of grate surface. If any special forms of valve are used,
they must be tested under steam in the presence of the surveyor ; in
no case must the steam pressure rise more than 10 '/^ above the working
pressure. An approved safety valve is also to be fitted to the super-
heater. Donkey boilers may be fitted with one safety valve only, if
the area of the latter is at least \ square inch for every square foot of
grate area. Every safety valve must be so arranged that it cannot be
overloaded when steam is up, and it must be provided with suitable
easing gear which must lift the valve itself. All valve spindles must
extend through the cover of the valve casing, and be fitted with a
socket and cross handles, allowing the valves to be raised, turned
round on their seats, and tested to see that they are in proper working
order, at any time.

Stop Valve. — Each boiler must be fitted with a stop valve, so that it
can be worked independently.

STEAM BOILERS. 581

Blow-off Cocks, — Every boiler must be fitted with a blow-off cock,
which is independent of the blow-off cock fitted to the ship's side.
Every boiler must be provided with an accurate pressure gauge.

Test Pressure, — The boilers must be tested by hydraulic pressure
up to double the working pressure in the presence of Lloyd's surveyors,
and carefully gauged while under test.

4, Bureau Veritas.

Safety Valves, — Each main boiler must have at least two safety
valves of approved design. Their total area must be such that with not
less than twenty minutes' hard firing the pressure on the safety valve does
not exceed the working pressure by more than 10 7o' If the boiler is
worked with forced draught, the sectional area of the safety valves must
be increased to correspond with the higher evaporative power of the
boiler. Arrangements must be made and gear provided to ease the
safety valves from the deck, or from the floor of the stokehole.

Main Stop Valve, — If there are several boilers in the ship, they must
be capable of being worked together or independently ; for this purpose
stop valves must be provided between the boilers and the common
superheater, or between the different superheaters and the main steam
pipe. The steam piping to the auxiliary engines must be independent
of the main steam piping, so that the main engine may not be affected
when the winches and pumps are worked. There should be an arrange-
ment for shutting off the boiler stop valve, either from the floor of the
stokehold or from the deck.

Blow-off Cocks, — The blow-off pipe must be so arranged that one
cock is fitted direct to the outside of the boiler shell, and one to the
ship's plating, and the scum cock must be fitted in a similar manner.

Water Gauges, — Every boiler must be fitted with at least two water
gauges, either with one gauge glass and one set of test-cocks, or with
two gauge glasses. Boilers fired at both ends must have a similar
arrangement at each end.

Every boiler must be fitted with at least one pressure gauge.

*

PART VI.

MEASURING INSTRUMENTS.

MEASURING INSTRUMENTS.

§ 336. Pressure Gauges. — There are two principal types of
gauges used for marine purposes, w\z,, flat-spring gauges (Fig. 498) and
round 'Spring gauges (Fig. 500) (Bourdon gauges). The principle
underlying both types is based upon the phenomenon that the deflection
of a spring subjected to a moderate strain is directly proportional to the
force acting on it. The deflection of these springs is indicated on a
suitable dial, the scale of which is determined experimentally by means
of a mercury column open to the air. As the elasticity of the springs
diminishes with increase of temperature, the pressure gauges must not
be brought in contact with hot walls or hot steam. Therefore the
makers stipulate that the instruments must only be exposed to moderate
heat, and protected from the hot steam by a small column of water
(Figs. 499, 501).

Every pressure gauge must be fitted with a stop-cock (usually a
three-way cock), to shut off* the steam pressure or to throttle it if neces-
sary, should the spring be exposed to violent fluctuations of pressure in
the steam space. For boiler pressure gauges, a flange (Fig. 502) must be
provided, to which the test gauge can be attached. These gauges are
generally made with double pointers and springs, on the Bourdon
principle; they are extremely sensitive and very accurately adjusted.
Wherever an engine plant of any size is at work, these gauges should
always be to hand, for continually checking the working gauges.

§ 337. Thermometers. — For low temperatures, up to about 400"
¥ahr,, ordinary mercury thermometers are used, and should be surrounded
with a casing, to protect them from injury. When it is desired to take
the temperature of steam in piping, &c., under pressure, stuffing boxes
and other protections should be provided, to prevent the thermometers
from being blown out by the pressure. As even with the most carefully
constructed thermometers there is always a risk of their breaking, it is
advisable, in order to obviate this danger, to protect them by a solid
casing or cup screwed into the steam pipe, &c. (see Fig. 503). The
heat is more effectually transferred if this cup is filled with mercury or
oil It may also be filled with fine iron filings, which can be extracted
from the cup by means of a magnet. It should be noted that quick-

!

586

MARINE ENGINES AND BOILERS.

silver is unsuitable for use with brass cups, on account of its action on
the brass. It is often contended that enclosed thermometers are in-
accurate, but this objection does not apply in practice, as for the degree
of accuracy required in these measurements, any slight injurious effect
of the casing may be neglected.

Fig. 498.

Fig. 499.

Fig. 600.

Fig. 501.

If the temperatures to be measured approach the boiling point of
mercury (648* Fahr.), the ordinary mercury thermometers cannot be
employed. In this case, mercury thermometers partly filled with
nitrogen are used ; />., thermometers in which the top of the glass tube
is widened, and filled with compressed nitrogen gas, to prevent the

MEASURING INSTRUMENTS.

Tig.aoi.

0

a

588 MARINE ENGINES AND BOILERS.

mercury boiling. Pressure ihermomeierSy based on the principle that at
certain pressures saturated vapours acquire well-defined temperatures,
may be usefully employed. These thermometers, called *^ mercurial
pyrome/ers" are made in the following way: — At the bottom is a
flat wide vessel (the evaporator), and to this a glass tube, as long as
may be required, is attached, carrying at one end the pressure gauge,
with lever and pointer (Fig. 504). The tube is filled with ether, water,
or quicksilver, according to the height of the temperature to be
measured. As long as these instruments are used to indicate tempera-
tures below red heat, />., up to about 1,300' Fahr., they are perfectly
reliable. Above that, they must be often recalibrated, by means of
standard air thermometers. Graphite pyrometers are much used for
commercial purposes, but are not suitable for marine work, as they are
too inaccurate, if constantly employed.

§ 338. Analysis of the Flue Gases. — The determination of the
excess of air is a valuable aid in judging of the efficiency of combustion,
and this may be best effected on board ship by the use of the Orsai
apparatus (Fig. 505). It is generally arranged to test the gases for their
percentage of carbon dioxide, carbon monoxide, and oxygen, and in-
directly of nitrogen. Three absorption bottles, a, ^, and r, are each
connected by a cock to a glass suction or capillary tube, bottle a being
filled with solution of caustic potash, b with pyrogallic acid, and c with
ammoniated solution of cuprous chloride. At one end of the capillary
tube is a three-way cock, which has a connection on one side to the
smoke box, on the other to a water jet pump. To the other end the
measuring burette d is attached, which has a capacity of 100 cubic
centimetres, and is carefully marked to scale in centimetres. At the
lower end of the burette is a rubber tube ^, which forms the connection
with a levelling bottle // filled with water.

The gases are analysed as follows : — The absorption fluids are first
introduced into that half of one of the double bottles which is nearest
the cock, and the cocks closed. The levelling bottle is raised till the
measuring burette is filled with water up to the top, and the three-way
cock is turned to open communication between the smoke box and the
jet pump. By means of the latter, all the air is sucked out of the con-
necting pipe, and it is filled with flue gases only. The three-way cock
then connects the smoke box with the measuring burette, and the
levelling bottle is lowered, till there are exactly 100 cubic centimetres of
gases in the burette, and the cock shut off. The cock to a is then
opened, and by raising the levelling bottle the gases are passed over
into the bottle containing caustic potash, for absorbing the CO^ The
surface of the reagent is increased by putting little tubes of glass into

MEASURING INSTRUMENTS.

589

the bottle. Absorption being complete, the levelling bottle is lowered,
the rest of the gases passed back into the burette, the cock at a closed,
and the volume of the remaining gases read off. The difference in the
readings gives the percentage by volume of carbonic dioxide in the
flue gases. By opening the other cocks at b and <r, one after the other,
the gases are passed successively into the other bottles, the oxygen and
carbonic monoxide are absorbed, and the final residuum is nitrogen.
Great accuracy is required in making these analyses, and special care
roust be taken that all the cocks and connections are perfectly tight.
The order given above in which the gases are absorbed by the different
reagents must be strictly adhered to.

J5 339. Draught Gauge. — The draught is measured in inches of
water. For this purpose the best
and simplest apparatus is a glass
tube bent in the shape of a U, Fig.
506 (U gauge). In practice the
mistake is often made of not stating
where the draught was measured.
To make the measurements of
greater value, it should be men-
tioned, not where the vacuum was
measured, but between which parts
of the boiler, for instance between
the ash-pit and combustion chamber, or l)etween the smoke box and
the outer air.

:=3^

Fig. o()6.

§ 340. Determination of the Heating Value of Coal.—

The better kinds of coal develop, more or less, the same quantity of
heat during combustion (have the same calorific value), nevertheless it
is sometimes useful to determine experimentally the heat efficiency of a
given fuel. It is not advisable that the engineer should himself under-
take these experiments (by burning the coal with oxygen in a Mahler
bomb), because they require much practice and occupy considerable
time; it is better to have the coal tested in a .chemical laboratory,
but great care must be exercised in selecting the samples. To get a
representative sample of the coal, a shovelful is taken from different
parts of the coal of which it is desired to ascertain the calorific value,
carefully crushed and then well mixed. From different parts of this
heap small samples are again taken, and further mixed, and the process
is continued until a small representative sample is obtained, which is
hermetically sealed, and forwarded for testing. As the percentage of
water contained in the coal materially affects the results, it should be
stated whether the samples are taken from wet or dry coal.

590 MARINE ENGINES AND BOILERS.

§ 341. Determination of the Amount of Moisture in the
Steam : Dryness Fraction. — The best apparatus yet introduct^ for
this purpose is Peabody's throttling calorimeter (see Fig. 507). This in-
strument does not require complicated arrangements, or occupy much
time in manipulation. The apparatus consists of a vessel of about 120
cubic inches capacity, which is connected through a valve with the steam
pipe and the condenser respectively, and is carefully protected from loss
of heat hy radiation, A pressure gauge is fitted in the steam pipe, and a

Fig. 507.

thermometer and pressure gauge are fixed in the calorimeter. T^e valve
fj is so adjusted that the steam in the calorimeter is much throttled, and
is thus at a considerably lower pressure than in the steam pipe. If ibe
absolute pressure in the steam pipe is /„ then the heat of vapoHsatioD ^^
and the latent heat r-^ corresponding to this pressure, can be ascertained
from the Steam Tables, and the values f j and r^ corresponding to the
pressure p^ of the throttled steam can be ascertained in the same way.
As, for every pound of mixture in the steam pipe, there will be j;, lb, of

MEASURING INSTRUMENTS. 591

steam, and (1 -x,) lb. water, the heat contained in 1 lb. of steam
will be Qi = ?i + A:if,. As soon as the steam in the calorimeter comes
to rest, it will contain, except for any minute losses from radiation, the
same amount of heat as when in the steam pipe. This will consist of
^2 = ?i + 'a + ^ Ca - 'z)' *■ being the specific heat of dry saturated steam
expanding at constant pressure (about 048), (^ the temperature shown
in the calorimeter, and i„ the temperature of saturation corresponding

Fig. ;*M.

to the pressure /j. The loss from radiation being negligible, Q|=Q),
that is, y, + a:,r, = ?2 + '■j + Af-i - 'j), from which x^ can be calculated.

These determinations will only yield reliable results if the apparatus
used is very accurate, and the percentage of water in the steam does not
exceed 5 '/__. If ihe steam is supposed to contain more than 5 */__ of
water, a separator must be fixed just in front of the calorimeter, and the
results obtained from it added to those of the calorimeter. To avoid
any radiation of heat into the surrounding air, the jacket round the

592 MARINE KNGINES AND BOILERS.

separator is so heated that the thennometer in it registers exactly the
same tempenttuTe as the one in the calorimeter.

§ 343. Indicators aad their use.— Those most generally em-
ployed in Germany are —

1. Schafffr &• Budenberg.

2. Drtytr, Rosenkrans, i5f Droop.

3. Cresby.

Fig. :m.

Figs. 508 to 510 show the latest types made by these firms. In all
the instruments the compression of a spring adjusted lo scale is utilised
to represent the varying pressure of the steam during one stroke of ihe
engine. The deflections of the spring, proportionally enlarged by the
parallel motion of the pencil, are transferred to the paper, and the
manufacture of the instrument has now been brought lo such perfection,

MEASURING INSTRUMENTS. 593

that the scale of the spring is a constant throughout the height of the
diagram.

g 343. Study of the Indicator and its Accessories : Pre-
parations for indicating.— I. Remove the paper cylinder and
lubricate the axis or pivot of the drum. The tension of the drum
spring must be such, that it is sure to secure the return motion of the

drum. The tension must therefore be greater if the number of revolu-
tions is high. Too great a tension is, however, to be avoided.

2. Remove the pencil and piston attached (according to the instruc-
tions given with each instrument). The inside of the indicator cylinder
must be perfectly free from any flaws, irregularities, or dirt, and must be
cleaned if necessary with sof^ wadding and oil ; any hard or carbonised

594 MARINE ENGINES AND BOILERS.

oil must be removed with petroleum. The use of sharp or pointed
tools must be carefully avoided, otherwise the bore of the cylinder may
be affected, and the piston no longer fit accurately ; if repairs are required
they should only be done by skilled hands.

3. To disconnect the Piston^ Piston-rod and Springs. — The screws
after they have once been loosened with a key must work easily, and
be cleaned if necessary with petroleum, as oil deposit, especially when in-
dicating high pressures, makes the instrument work stiffly. The pencil
and paper should be tested while the engine is running light (this pre-
caution is often neglected). Little can be altered at this stage, but
the result of this test shows whether the instrument can be used or not
for a given experiment.

4. To adjust the Spring. — Before this is done, it should be known
between what limits of pressure the instrument is required to indicate.
If the Spring is too weak, the indicator piston will hit the cylinder top,
and an imperfect diagram will be drawn ; if it is too strong, the top of
the diagram will not be fully developed, and the area will be too small.
The springs are calibrated and marked with a number or fraction
which represents the distance they compress in inches, per pound per
square inch of steam pressure.

5. Pencils. — Metal pencils are generally supplied with the instru-
ments, with prepared paper, on which the diagrams are drawn. It is
often better, however, to use ordinary smooth writing paper and a
common lead pencil, as the diagrams thus drawn are clearer. This is
especially the case when several diagrams have to be taken one over
the other, the various lines being more sharply defined.

6. To fit the Indicator on to the Engine. — ^This is done without putting
in the indicator piston and pencil. During the interval before indicating
the engine, the indicator cocks must be put in and tested, to see if the
threads of the cock and indicator fit each other, if the plugs of the
cocks work easily, and if they are properly adjusted.*

7. An indicator table should be drawn up, showing at a glance what
indicator springs are to be used for the different cylinders, what make
of indicator is used for the top and bottom of each cylinder, and any
errors to be allowed for, as for example : —

* It often happens that the plug is so fitted in that instead of being able to draw
the atmospheric line, steam blows through. If the washer cannot be turned through
90', the plug itself must then be turned through ISO**, and this can only be done
when steam is shut off.

MEASURING INSTRUMENTS.

595

Make and
Number of
' Indicator

Scale of
Spring.

H.p. Cylinder.

Top,

Bottom.

No. I.
M.P. Cylinder.

Top.

Bottom.

Schafier &
Budenberg
No. 648 I No. 649 No. 483 No. 487

Crosby

No. 2.
M.p. Cylinder.

Top.

Bottom.

Crosby
No.486INo.485

L.P. Cylinder.

Top.

Bottom.

Dreyer &

Rosenkranz

No. 590 I No. 591

Remarks

xio inch
per lb.

Steam
jacketed

^^ 5 inch
per lb.

Steam
jacketed

^ inch
per lb.

^ inch
per lb.

Pencil

not
working

Spring
newly
tested

iV inch
per lb.

iVinch
per lb.

^ inch ' ^ inch
per lb. per lb.

Alu-
minium
drum

Alu-
minium
drum

8. Before putting in the pencil and paper, steam should be well
blown through the indicators, to clean and heat the passages. On the
low-pressure cylinder it is advisable to blow through only during the
period of admission, to prevent cold air being drawn in. Immediately
before setting the indicator, the paper and pencil should be tested,
and the piston and spring examined to see if they are properly and
firmly screwed up, and whether the spring corresponds with the scale
given in the table. The piston is then dipped in thick cylinder oil,
well smeared, and fitted in. The precaution of keeping the pencil and
its gear free to move on the indicator cylinder is often overlooked, but
this is important, to avoid any hitch while taking a diagram.

§ 344. The Driving Gear. — I'he principal point to be kept in view
is that the paper and drum must move in proportion to the stroke of
the piston. To effect this a lever (Fig. 511), pivoted at a short distance
from one end, is almost always used. For other methods of driving the
indicator see Figs. 514, 515. It may be worked in one of two ways,
either from the crosshead, or by an eccentric on the shaft. In the
former case the indicator cord must be at right angles to the central
position of the driving lever (see Figs. 512 and 513); in the latter it
must be so arranged, that the eccentric and the engine piston are at
the top of their respective strokes at the same moment.

Fig. 515 shows a method of driving the indicator in high-speed
engines. The link motion must never have any play or backlash, and
the cord must not be too heavy or the bends too sharp, otherwise the
diagram will be distorted. The end of the driving lever, on to which
the cord is hooked, must be free from sharp edges and corners, to
avoid any risk of injury to the operator. After putting on the cord the

596 MARINE ENGINES AND BOILERS.

MEASURING INSTRUMENTS. 597

drum must be examined to see if the diagram can be properly drawn
in its highest and lowest positions of the crank, or whether there is
any defect in guiding the cord.

§ 345. Putting on the Paper. — The paper must lie smoothly all
over the surface of the drum, otherwise errors will occur in drawing the
diagram, and the paper may get torn. If necessary the clips must be
removed and readjusted.

§ 346. Planimeter. — To work out the diagrams the Amsler polar
planimeter is generally used. It consists of a movable arm cd^ and a
fixed arm ab^ Fig. 516. If the instrument is in proper adjustment,
and the pencil d is made to trace exactly the outline of the diagram,
starting from any given point, the reading on the wheel, multiplied by the
constant for the apparatus, gives the area of the diagram in units corre-
sponding to the constant. There is usually a sleeve by which the fixed
arm can be shifted along the movable arm, and the planimeter adjusted
to the scale of the diagram. In this case the reading gives the mean
height of the diagram in a definite unit of measurement. In using the
planimeter it is always advisable first to draw the vertical tangents to the
diagram (at right angles to the atmospheric line), in order to mark off
exactly the length of the diagram upon the atmospheric line ; the pro-
cess of placing the diagram in position is then quite easy. It should
be noted, however, that there are always slight variations in the lengths
of diagrams, and the distance between the extreme points should be
checked for each diagram. It is also necessary to check the constants
of the planimeter, and if they are not known they must be determined,
by accurately drawing a number of rectangles of known size, and then
measuring them by the planimeter. The calculated area of each rect-
angle divided by the planimeter reading gives the constant.

To find the constants forming the basis of a diagram already plotted,
all the rectangles are drawn with the same base line, and the plani-
meter traced over them ; in this case the height measured, divided by
the reading, gives the constant required. Care must be taken when
using the planimeter that the wheel runs easily over smooth and un-
creased paper. To begin measuring the diagram, the movable arm is
placed at a tangent to the diagram, and at the same time at right angles
to the fixed arm. In this position, errors in plotting at the beginning
and end of the measurement have least efi'ect on the result. The edge
of the wheel must be carefully protected from injury.

§ 347. Schlick's Pallogjaph. — To determine the vibrations in
the hull of a ship, the Schlick pallograph has been much used of late.

698

MARINE ENGINES AND BOILERS.

Fig. 516.

Horizontal VibraNons

Vertical Vibrations.

Time Base, 5 Seconds,

Fig. 517.

up.

Fig. 518.

MEASURING INSTRUMENTS. 599

This instrument mainly consists of four parts, a strip of paper driven
by clockwork, and three pencils, one of which records the vertical
vibration, the second the horizontal, and the third marks the time
base. The vibrations are rendered visible by heavy weights suspended
from springs, and connected to the two first-named pencils. The vertical
vibrations are recorded by a weight which can only swing vertically
relative to the point of suspension, and the horizontal by a second
weight which can only swing horizontally. The principle of the appa-
ratus is based on the phenomenon that heavy masses are not susceptible
to sudden changes of speed. Therefore the weights suspended from
springs, and the pencils, do not take part in the comparatively rapid
successive vibrations of the hull of the ship, but they are followed by
the frame of the api)aratus and the strip of paper, because these are
fixed to the ship. The pencil for the time diagram is moved regularly
to and fro by clockwork, and draws a stepped line upon the paper.
Fig. 517 shows a diagram from this apparatus, from which the character
of the vibrations can be clearly followed.

§348. Instrument for Measuring the Uniformity in the
Turning Moment of an Engine. — A useful instrument for testing
whether the turning moment of a marine engine is uniform is shown in
Fig. 518. A collar b with four cotters is placed on the shaft w. Upon
it a metal hoop r, carrying a strip of paper round it, can be fixed by
sliding it along parallel to the axis of the shaft. A tuning fork con-
trolled by an electro-magnet is made to vibrate strongly and uniformly
by means of an interrupter, and the vibrations are transferred by a pencil
to the blackened paper strip. While the engine is running, the tuning
fork is carried along the shaft by means of a screwed spindle, and thus
draws a continuous waved line on the paper. After the experiment the
strip of paper is taken off, and from the variations in the lengths of the
undulations the uniformity of the tummg moment can be determined.*

§ 349. Fottinger*s Torsion Indicator is used to register auto-
matically by means of a curve the amount of the actual turning moment,
f .^., the actual tangential pressures during one revolution. The diagram
of the torsional angles and curves gives respectively, on a different scale,
the actual turning moment, as the twist of the shaft is exactly propor-
tional to the effective turning moment.

The apparatus. Fig. 519, consists of a tube in two parts which is
passed over the shaft, the forward end of the tube being firmly secured
to it. The after free end carries a disc marked i, and opposite to

• QompBLxe Journal of the Sthiffbautecknischen Gestlhchafty vol. i.

600

MARINE ENGINES AND BOILERS.

the latter a similar disc marked ii is fixed on the shaft. Two points
on discs i and ii which, if the shaft is not in tension, are opposite each
other, will, on account of the twisting of the shaft when running, be
shifted along a short distance s, which can easily be calculated from the
turning moment at any given instant, &c., from § 3. This slight move-
ment (about "06 to '08 in.) of disc i in reference to ii, is enlarged from 18
to 28 times, by means of the arrangement of levers (shown at Fig. 519).
The tie rod a ^ is connected with disc i by the screw k, and moves the
shorter arm of the bellcrank lever ^ c d, the axis of which c k fixed to
disc II. To the longer arm c </, the rod ^ ^ is attached at d, and works
the indicator lever /c^. This is capable of angular movement round a
point/ which is fixed to disc ii. The other end g of this lever carries
the pencil for drawing the diagrams, the point of which faces towards

Rtcordinj dnim

the centre of the shaft. When working, g swings from its point of
rest in the direction of the axis of the shaft, in proportion to the
turning moment at that instant, and all the levers, discs, tube, &c.y
rotate with the shaft.

Between the pencil and the shaft, and concentric with the latter, is
a drum which is pushed axially along a slide between them, and upon
this as many torsional diagrams as may be desired can be drawn. Paper
is stretched round the drum, and upon it the pencil draws the curve
proportional to the torsion, /.^., the actual turning moment. Another
pencil, not shown in Fig. 519, draws the datum line. To take off the
diagram, the drum is pushed to the left, and a fresh paper can then be
fitted to it.

Fig. 520 shows the construction of the first type of this instrument
The open end of the measuring tube is centred to the shaft by four

MEASURING INSTRUMENTS.

i

602 MARINE ENGINES AND BOILERS.

large adjustable rollers. The levers and their machinery are made in
duplicate (and fixed to two opposite points) in case one of the pencils
should fail to act. By means of the bellcrank lever m m, and a stop
fixed to the body of the ship, the pencils can be raised from the paper,
and the levers thrown out of action at the same time. Before starting
the instrument, the zero point of the movable pencil is determined by
turning the shaft backwards and forwards, and the stationary pencil is
fixed in the position thus given. The actual work of the engine is found
by measuring the diagrams with a planimeter. The apparatus registers
automatically the variations in turning moment produced in a shaft of
a given length. With a somewhat modified torsional indicator of this
type, the b.h.p. of one of the 20,000 h.p. engines of the fast steamer
" Kaiser Wilhelm II." was determined.

With another arrangement of the apparatus, suited to steam turbines
and electric motors, the mean turning moment can be read off to scale
on a stationary indicator.**^

See page 441,/aArducA der SckiffbauUcknischen GeseUschaft,

PART VII.

VARIOUS DETAILS,

VARIOUS DETAILS.

r

+r

i

c

01

D

C

§ 350. Bolts, Nuts, and Screw Threads, &c.

The allowable stress upon small screws is always taken as much less
in proportion to that upon large screws, as the former are apt to be easily
broken when being screwed in. This must be specially taken into
account with bronze screws, which offer a smaller resistance to twisting.
For dimensions of screw bolts with Whitworth thread, see Table No. 70.

In all screws over f inch diameter, the tensile strength of which has
to be as large as possible, such as bolts for connecting rods, crossheads,
valve rods, &c, a fine thread is used (see Table No. 71). The heads
of bolts that must be kept from
turning, or which cannot be held
fast while the nuts are being screwed
on, must be fitted with a projec-
tion or feather (see Figs. 521, 522).
A good sized fillet must always unite
the head to the turned part of the bolt.
When square heads are used, each side
of the square must be made the same
width as the corresponding hexagonal
nut across the flats. Tap bolts or studs
should be screwed into cast iron to a
depth of one and a half times the
diameter of the bolt ; but if necessary
the depth may be decreased to one
and a quarter times the diameter. If

they are to be screwed into tight-fitting holes, they must have a groove
running lengthwise across the thread, to allow the air to escape while
the bolt is being screwed home. Bolts and pins, the nuts of which
have constantly to be taken off, should have the top rounded off, and
their ends turned down for a short distance to the bottom of the thread
to allow of the nuts being easily screwed on and off. This is
especially necessary in the case of large horizontal screws, such as the
screws for propeller nuts.

Iron or steel studs must not be screwed into bronze, as the iron
rusts rapidly, especially if it is exposed to the action of sea water. The

0

:)

Fig. 621.

Fig. 522.

606

MARINE ENGINES AND BOILERS.

heads of countersunk screws should not have a slit (Fig. 523), as

turning the screws burrs the slits,
and injures the countersinking. It
is better to use a small square-headed
"* screw (Fig. 524). In screws which
have to be constantly taken out, it
is advisable to case-harden the heads.
Steel bolts or pins should have wrought-
iron nuts, which do not wear the thread
of the bolt so quickly as steel nuts.
Nuts which have to be constantly taken
oflf must be case-hardened if they 'are
to retain their hexagonal shape.

Lock Nuts, — For small bolts sepa-
rate lock nuts are generally used, for
the depth of which see Table No. 70, or
for ring or " Penn " nuts see Table No. 72. Split pins must not be
considered sufficient to prevent the nut slacking back, but only to keep
it from falling right off. For particulars and the methods of securing
the nuts, see below.

Table No. 70.

Bolts with Whiiivorth Thread.'*'

Depth of nut = diameter of bolt = a.

Depth of lock nut = depth of bolt head = ^ = ^1a,

Width of head, if square \

Diameter of head, if round l^ = 1 4^ -j- '2 inch.

Width over flats, if hexagonal )

Width measured over angles c— 1*155^.

Fig. 523.

Fig. 525.

• The dimensions of nuts vary greatly in the different tables. The following table
must therefore be taken as affording a basis for calculation only, and their actual
dimensions may vary with the requirements of the different clients. For instance, the
Imperial German Navy specifies particular sizes for nuts.

VARIOUS DETAILS.

607

1

No. of threads to the inch = «. .'. pitch = - inch.

n

1*28 .
Diameter at bottom of thread d= a-l"2Sx pitch — a- — ^ inch.

* ft

Diameter of

Bolt
in inches.

Diameter at

Bottom of

Thread in inches.

a.

if.

i

§
i

H
H
1}
14
It
li
H
2

21
2i
2}
3

H

34
3|

4

*i
44

;>

H

5i

5f

6

0134
0186
0-241
0-295
0-346
0-393
0-456
0-509
0-622
0-733
0-840

0-942
1067
1161
1-286
1-369
1-494
1*590
1-715

1-930
2-180
2*384
2-634

2-855
3-105
3-323
3-573

3-804
4-a>4
4-284
4-534

4-762
5-012
5-239
5-489

w « *

m

ft.

24

20
18
16
14
12
12
11
10

9

8

7
7
6
6
5
5

44
44

4

4

34
34

3*
3i
3
3

2i

2}

2«

28
2ft
24

24

Hexagonal Nats.

Width

over

Flats

in inches.

A.

-448
-525
•6014
•7094
-8204
•9191
1-011
1-101
1-3012
1-4788
1 -6701

1*8605
2-0483
2-2146
2-4134
2-5763
2-7578
3-0183
31491

3 546

3 894
4*181

4 531

4^85
5175
5-55
5-95

6-375
6-825
7-3

7-8

8-35
8-85
9-45
10-0

Width

over

Angles

in inches.

r.

•3«

OjC

a'

1

c.

® e •

/

-5173

-164

•6062

*2187

•6944

-2734

•8191

3281

•9473

-38*28

1-0612

-4375

1 1674

-4921

1-2713

•5468

1-5024

-6562

1 -7075

•7656

1-9284

•875

21483

•9843

2-3651

10937

2-5571

12031

2-7867

1-3125

2-9748

1-4218

3-1844

1 -5312

3-4852

1-6406

3-6362

1-75

4-0945

1-9687

4-4964

21875

4.8278

2-4062

5-2:^19

2-625

5-60(12

2-843

5-9755

3-062

6-4f)85

3-281

6-8704

3-5

7-3612

3-718

7-8819

3-937

8-4293

4-156

9-0066

4-375

9*6417

4-593

10-2190

4-812

10-9119

5031

11-5470

5-25

0-157
0*177
0-177
0196

0196
0-236
0-236
0-275
0-275
0-315
0-315
0-3^

0-a54
0*393
0-393
0-472

0-472
0-551
0-551
0-629

0-629
0-669
0-669
0-708

0-708
0-748
0-748
0-787

s-g-s

^•t

0-236
0-275
0*275
0-314

0-314
0-354
0-354
0-433
0-433
0-472
0-472
0-551

0-551
0-630
0-630

0-708

0-708
0-826
0-826
0-944

0-944
1-062
1*062
1181

1*181
1*259
1*259
1*338

0-014
0-027
0-046
0-068
0-094
0121
0-163
0-203
0-304
0-422
0-554

0-697
0-894
1-058
1-299
1-472
1-753
1-986
2-311

2-925
3-732
4-464
5-450

6-402

7-563

8-673

10027

11-365
12-908
14*404
16*146

17-810
19*72
21-57
23*64

* Calculated from the diameter at the bottom of the thread.

t As a rule, the point starts immediately above the nut, so that there are no threads
outside the nut.

608

MARINE ENGINES AND BOILERS.

Table No. 71.

Bolts with Fine Threads {for Connecting Rods^ CrossheadSy

Main-bearing BoltSy dr'c).

*> a
c

cS

¥

2i
2^

<* S e

0-665
0-756
0-901
1019
1118
1-247
1-374
1-484
1-610
1-732
1-818
2-039
2-287
2-539

o -^

i!

15

13

13

12

10

10

10

9

9

9

7

6

6

6

B

o c g

O ^ u

V

0-347
0-472
0-638
0-816
0-988
1-2-23
1-482
1-730
2-036
2-356
2-598
3-266
4-109
5 064

g.s

S3

PQ

« 5 r-

o

2-787
2-995
3-244
3-496
3-716
3-976
4-212
4-448
4-685
4-921
5-110
5-400
5-629
6-653

u

H

a

6-102

7-052

8-265

9-600

10-850

12-418

13-937

15-544

17-239

19-021

20-573

22-848

24-894

34-769

-V;-L±:

Table No. 72.
jRing Nuts.

Fig. 624a.

Fig. 525a.

/ and g are the same as in Table No. 70 ; * ^ is the same as in
Tables Nos. 70 or 71.

As a rule the turned end begins immediately above the nut.

VARIOUS DETAILS.

609

Table No. 72 — continued.

a.

^.

■

1.

/'.

Inches.

m.
Inches.

/I.

0,

/.

Inches.

Inches.

Inches.

Inches.

Inches.

Inches.

Inches.

f

0-59

019

0-23

0-10

1-22

114

2-00

_

3

0-67

0-19

0-23

0-10

1-37

1-33

2-20

1

0-78

0-23

0-31

0-11

1-57

1-49

2-75

6
16

n

0-86

0-23

0-31

oil

1-69

1-61

2-87

5
TIT

4

0-94

0-23

0-31

on

1-89

1-81

3-07

5
TF

^

102

0-23

0-31

oil

2 04

1-96

3-22

6
IF

110

0-23

0-31

013

2-24

2-16

3-46

6
TF

M8

0-31

0-35

013

2-44

2-36

3-85

.3
IT

^

1-30

0-31

0-35

0-15

2-59

2-52

4-01

.1

H

n

1-41

0-31

0-35

015

2-75

2-67

4-17

7f

2

1-49

0-31

0-35

0-19

2-91

2-83

4-33

F

H

1-69

0-39

0-39

019

3-22

311

5-04

J

2l

1-85

0-39

0-39

0-19

3-58

3-46

5-39

1

2}

2-08

0-39

0-39

0-23

3-93

3-81

5-74

1

3

2;28

0-39

0-39

0-23

4-29

4-17

6-10

i

H

2-44

0-47

0-55

0-27

4-64

4-52

6-85

5

■ff

^m

H

2-60

0-47

0-55

0-27

500

4-88

7-20

6

3|

2-75

0-47

0-55

0-27

5-31

5-19

7-52

R
F

4

2-95

0-47

0-55

0-31

5-67

5-55

7-87

5
F

4i

315

0-59

0-63

0-31

6-02

5-86

8-62

41

3-34

0-59

0-63

0-31

6-37

6-22

8-97

3

T

4|

3p4

0-59

0-63

0-35

6-73

6-58

9-33

T

5

3-74

0-59

0-63

0-35

7-08

6-93

9-68

5i

3*93

0-63

0-78

0-39

7-44

7-28

10-43

7

5J

413

0-63

0-78

0-39

7-79

7-63

10-78

7

5J

4-33

0-63

0-78

0-39

8-07

7-91

11-06

7
IT

6

4-52

0-63

0-78

0-47

8-42

•

8-26

11-41

7
F

^ 351. Screw Spanners. — In the equipment of the ship one or
more sets of single (see Fig. 529) and double ended spanners (see Fig.
526) are provided. For large nuts, which have to be constantly taken
off (connecting-rod bolts, main-bearing bolts, &c.), a light spanner is
generally used for screwing them on, and a heavy spanner, wl\ich can
be hammered, for tightening them up. For large heavy nuts the spanners

* In the ring nuts shown at Fig. 5*2«5a, / may be taken somewhat smaller.

2q

iBlO

MARINE p:N(iINES AND BOILERS.

Fig. 526.

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VARIOUS DETAILS.

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612

MARINE ENGINES AND BOILERS.

are generally made in the form of ring spanners, and if there is not much
room to turn them, they sometimes have twelve angles instead of six.
Very large spanners (for propeller nuts, stern tubes, &c.) frequently have
an eye at the end of the handle, for turning them with a rope and tackle.
Such a spanner for a slotted nut is shown at Fig. 528.

Fig. 529.

Spanners are made of steel rough ground on the grindstone and
the sides in the jaws machined. In the best practice they are specially
finished bright and then hardened.

Table No. 73 gives the dimensions of single-ended spanners.

§ 352. Platforms. — Platforms are fixed where necessary and at
convenient heights. In the engine-room one is usually arranged a
little above the crank shaft, at about the same level as the top of the
bedplate; and in the boiler-room at about '2 feet 6 inches to 4 feet
below the centre of the furnace. The platform consists of chequered
plates of convenient size, so that they can easily be taken up, to
get at the pipes above the double bottom. Thickness of plates /^ to ,f.^
inch, and f inch in the boiler-room measured over the ribs. The
plates are laid on angle-iron frames, which are secured to the engine
and the body of the ship, and supported where necessary on wrought-
iron uprights or supports. Cocks and valves below the platform have
a square end on the valve spindle, to turn them by means of a box
key which is passed down through a hole in the platform, fitted with
a cover.

§ 353. Edging Plates about I inch thick and 6 to 16 inches
high are put round the engine and all the openings in the platform in
the engine-room. The cranks are protected by metal plates, to prevent
the oil and water being splashed over the platform. These plates must
be easily removable, and must obstruct the examination or feeling of
the cranks as little as possible. The quick running wheels of the turning
and reversing engines, &c., should be protected by metal casings.

§ 354. The Gratings in heavily built ships are composed of a
framing of flat iron about 2^ to 3 inches deep and ^^^ to f inch thick,
with f inch bars riveted in. Sometimes bars of wedge or T-shaped
section are used. They are spaced 2 to 2^ inches apart. Only such
sections of bars should be used as are easily cleaned. In gratings
where weight has to be considered, the frames may be made of C-iron
or light flat iron, and the bars made only ^- inch diameter. Gratings in
the boiler-room always have round iron bars about } inch diameter.

Allowable width of gratings —

With round iron bars ^ inch diameter = 18 inches.
With round iron bars J inch diameter = 24 inches.

Broader gratings must be braced across the middle by flat iron bars

614 MARINE ENGINES AND BOILERS.

If wedge-shaped or T-section bars are used, the width of the gratings may
be considerably increased, corresponding to the thickness of the bar
used. If there is plenty of room, the gratings may be from 2 feet
6 inches to 3 feet wide, but where space is restricted they are often only

1 foot wide. The top gratings are made as light as possible, and all
gratings must be provided with handrails. They are supported on
wrought-iron or cast-iron brackets, spaced from 6 to 1 2 feet apart. In
very large engines, with Stevenson's link motion, the lowest grating is
a little below the valve gear, the middle one a little above the lower
edge of the cylinder, and the top grating on a level with the cylinder
covers. Small engines have only one grating, at the same l^eight as
the bottom of the cylinder.

S 355. Ladders. — These have sides of flat iron, and rungs of cast
iron, chequered plate, &c. In the boiler-room, instead of steps, there
are rungs of round iron ^ to f inch diameter. Sometimes the rungs con-
sist of one bar of round and one bar of square section side by side, the
latter being in front. Ladders are placed nearly vertical in the boiler-
room, or if there is sufficient space inclined at 75** to S0\ Width of the
main ladder in the engine-room, 1 foot 8 inches to 2 feet 6 inches ; of
those in the boiler-room, 15 to 20 inches; distance between steps, ^
to 10 inches.

§ 356. Balusters and Handrails are provided wherever neces-
sary. The handrails are made of smooth drawn or polished round iron or
steel tubing from -J to 1 inch diameter. The wrought-iron balusters
carrying the handrail are about 3 feet high, and have a ball head at the
top, through which the handrail passes. They are secured at the
bottom to the frame, or to the bracket carrying the frame, by a flange
with two bolts, or by a collar, thread, and nut. The diameter of the
balusters if 3 feet high, as in merchant vessels, is usually 1 inch at the
top, 1^ inch at the bottom, and diameter of ballhead about 1| to

2 inches. In warships the balusters are made lighter and shorter.

§ 357. Lifting Gear over the Engines. — In medium-sized and
large merchant ships there are travelling cranes above the engines, by
means of which the cylinder covers, pistons, &c., and even the upper
cylinders themselves, when placed tandem one above the other, can be
lifted and moved either lengthwise or athwartships. The cranes run
lengthways of the ship upon I section girders placed at each side of
the engine hatch. The crane girders are C or I section. The lifting
machinery on the crab consists of worm gear, or of a screwed spindle
only, the manoeuvring of the crane being done entirely by hand. In

VARIOUS DETAILS. 615

very large ships two cranes may be used over each engine, running on
the same set of rails, and occasionally they are electrically driven.

§358. Lifting Gear for Engines of Warships.— In warships

it is impossible to have a travelling crane, on account of the armoured
deck. Below the latter, and immediately above the cylinder covers,
there are fitted short CI or I section girders placed side by side, two and
two together, along the lower members of which small travellers work,
with gear to lift the cylinder covers and pistons. The latter can only
be taken out after they have been disconnected from the piston rods.

§ 359. Engine Foundations. — The stresses upon the engine
foundation are produced by —

1. Weight of the engine.

2 Force exerted by the total mass and weight of the engines, due
lo pitching and rolling of the ship.

3. The vertical and horizontal forces and moments due to the

inertia of the reciprocating and rotating parts.

4. The moment on each crank, due to the turning moment of the

engine.

5. Bending moments resulting from the thrusts due to the steam

pipes, in the case of two cylinders placed side by side, but not
stayed. To these must be added the stresses which the foun-
dation has to resist as part of the hull of the ship.

The stresses under 1, 2, and 5 are generally small in comparison with
those under 3 and 4. These latter may, in large engines, become very
considerable, especially if the oscillations of the ship are a multiple
of, or synchronise with the number of revolutions of the engine.

Some idea of the extent and direction of these stresses may be ob-
tained by plotting the moments for each position of the crank, and then
constructing the polygon of forces. In order to determine the stresses
transmitted to the body of the ship, the starting point may be taken at a
place where the foundation of the engine merges into the double bottom
of the ship. In reciprocating engines it is often sufficient to determine
the greatest equivalent twisting moment of an outer crank, in relation to
the plane of the adjacent inner crank. In this case the stresses upon the
hull of the ship can only be small. The influence of the length of the
connecting rod is here neglected, as it is not important. If an engine
has strong diagonal cross braces betw^een the different columns or
entablatures, the foundation may be made much lighter, on the assump-
tion that the longitudinal bracing converts the framing of the engine
into a solid girder.

.^ 360. Construction of the Engine Foundation (see also
Bedplates).

VARIOUS DETAILS.

617

(a.) Merchant Ships with Double Bottom. —The flanges for securing
the bedplate are generally all in one horizontal plane; the bedplate being
simply laid upon the plates forming the top of the double bottom. The
framing of the double bottom under the engine is generally deeper than
elsewhere, and a few horizontal girders, of the same depth as the double
bottom, are fixed into it, over and above those required for stiffening
the ship. The top plates of the double bottom are especially strengthened

Fig. 530.

Fig. 531.

(being made up to 1^ inch thick for fast steamers, in medium-sized
merchant ships \ inch to 1 inch, and from \ inch to | inch in smaller
ships), in order to accommodate the holding-down bolts, &c. (see
Bedplates).

(A) Warships with Double Bottom, — The framing of the engine
seating is, as a rule, deeper in the middle than at the sides, and is braced
fore and aft by longitudinal beams. These longitudinal beams are

618

MARINE ENGINES AND BOILERS.

situated immediately under the longitudinal members of the bedplate,
and rest on, and are riveted to the double bottom. Under the bed-
plate these longitudinal girders are usually braced by cross box girders,
which are cut away to accommodate the engine seating. The space
between the girders and the cross frame of the engine bedplate is
fitted with wooden packing pieces (compare Figs. 530, 531).

(r.) If there is no double bottom^ the longitudinal girders are laid
across the frames, and strongly connected to them (there should be at
least four rivets at each intersection). It is best to fix the longitudinals
to the frames or the body of the ship by intercostals reaching down
between the frames. If the engines are heavy, and the bedplate is level

Fig. 532.

along the bottom, a series of longitudinal girders may be placed above
and across the frames, and plated over, so as to form a surface for the
bedplate to rest on.

§ 361. Boiler Seatings (see Fig. 532). —The bearers, composed of
angle irons and plates, extend over about one-quarter of the circumference,
and should be so arranged that they do not cover the riveted seams,
stay bolts, &c. The number of bearers, composing the seatings on
either side of the boiler, depends on the size, and especially on the
weight of the boiler, and may vary from two to six. The size of the
angle iron used is from 2 J x 2 J x § in. to 4 x 4 x J in. ; thickness of the
iron plates from j\ to \ in. The seating nearest the centre of the ship

VARIOUS DETAILS. 619

is generally made in two parts, that the boiler may be more easily
stowed in the wing of the ship.

Each boiler bearer should, as far as possible, rest on a frame (or on
projecting parts of the ship's bottom); if this cannot be done, they
should be supported on girders carried by the frames, so as to transmit
the strains directly to the latter. Above the double bottom the bearers
should be connected by side plateS, either riveted or bolted on. In
order that the boiler may be got at between the bearers, holes are cut
in the connecting plates and bearers. To make the lower part of the
boiler shell accessible, the lower edge of the boiler should, if possible,
be at least 1 2 inches above the double bottom of the ship.

To keep the boiler in place when the ship rolls, &c., from two to four
eyes are riveted to the upper p)art of the boiler shell, and are secured
with tie bars to the nearest deck beams, " stringers," or to the adjacent
boilers. These eyes may also be used for lifting the boiler. Another
method of securing the boiler is to rivet four eyes to it immediately
above the bearers, and connect them with . the latter by means of
short ties. In this way the boiler is fastened independently of the
deck, which always " gives " more or less, and also independently of the
skin of the ship. To prevent the boiler shifting longitudinally, strong
angle-iron brackets are riveted or bolted to the double bottom of the
ship, and butt against the front and back of the boiler.

§ 362. Lubrication* of the Steam Spaces.-— This is generally

effected with mineral oil, to which graphite is sometimes added. The
oil is forced into the lubricating pipes by a piston working in a cylinder,
and driven by steam or by means of a special lever from one of the
working parts of the engine. The pipe leads either to the main stop
valve only (to lubricate the steam), or in larger engines to each valve
chest as well as the main stop valve. The cylinders are either connected
to the oil pipe referred to above, or they have separate lubricators so
designed that oil finds its way into the cylinders, in spite of the pressure
of steam existing in them. The lubricator generally *^has an arrange-
ment whereby oil can be quickly forced out by hand or by steam
pressure, should a piston or valve require special lubrication, and there
should also be some means of shutting off the pipe supplying any

particular part of the engine. To prevent steam
from say the h.p. or the m.p. cylinders reaching
the L.P. through the branches of the main oil pipe,
there is a small non-return valve at the mouth of
each branch pipe, which is forced open by the
pressure in the lubricator.

The piston rods are swabbed from time to
time with mineral oil, which is retained for a short
time by the wicks attached to the outer end of the
stuffing boxes.

J5 363. Lubrication of otlier Parts.—

Almost all the other parts of the engine are lubri-
cated with vegetable or animal oil, which is applied
either by ?iand^ by siphon lubricators^ or by automatic sight-feed lubri-
cators. For the rotating or non-reciprocating parts, the oil flows from
the oil boxes direct to the bearings. The reciprocating parts are supplied
with oil from a stationary lubricator. A pipe carries the oil from the
latter to an open cup fixed to the reciprocating part, or the oil is wiped
off a small brush or feather at the end of the pipe into the cup. Several
small oil pipes are usually run from one large lubricator, which should
be easy of access. In reciprocating parts (such as the crosshead, &c)
some arrangement should be provided, to prevent the oil being flung

* Compare the arrangements of engines and photographs in former chapters.

Fig. 533.

VARIOUS DETAILS.

621

out of the lubricator. A small piece of perforated metal or wire gauze
is generally placed in the upper part of the cup, to prevent any dirt
finding its way into the bearings (see Fig. 533).

The oil cups on the reciprocating parts must be solidly screwed in,
so that they may not work loose, and fly off when the engine is running ;
the oil pipes must also be secured in a substantial manner. The most
important parts of the engine, such as the crank and crank-shaft
bearings, are fitted with arrangements for oiling by hand, besides the
siphon lubricators. Care must be taken that the oil pipes reach into
the brasses, and fit them tightly.

The oil is distributed throughout the bearings by means of grooves,
which are cut in the working surface of the bearing (the grooves
should be from ^ to f inch wide, according to the size of the bearing,
and half as deep). The bearings must be so arranged that the oil
cannot find its way out sideways, but is forced to distribute itself evenly
over the bearing.

Internal Diameter of Lubricating Pipes,

Engines below 1,000 i.h.p. -

from 1,000 to 3,000 i.h.p.
„ 3,000 to 7,000 „
„ 7,000 I.H.P. and upwards

M

»l

Crank and
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Valve Gear,
&c.

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^ to I
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.

The Plummer blocks are sometimes lubricated with solid grease, as
well as with oil.

Bearings subjected to light strains are often only provided with
small holes fitted with Stauffer lubricators.

Worm gear should, if possible, be run in an oil bath.

Besides the arrangements here described, a large number of special
types of lubricators are also employed.

8 364. Ash Hoists. — In smaller ships the ashes are drawn up to
the deck in buckets with chains by small hand winches, through one of
the ventilators, and then thrown overboard. In larger ships the winches
are worked by steam (diameter of cylinder and stroke about 4 inches).
Number : — Either one for each stokehold or one for every two stoke-
holds, if conveniently situated near each other. The winches are fixed
in the boiler-room casings, at the level of the upper deck. The buckets
are hoisted up in the ventilators by means of wire ropes. The winches
are so arranged that the steam is cut off automatically, when the bucket
reaches the top.

S 365. Ash-ejectors. — In recent large ships the ashes are dis-
charged by what are known as " ash-ejectors " (see Fig. 534). They
are shovelled into a cast-iron hopper t, on the top of which is a strong
cast-iron grating, through which the ashes fall into the hopper, and
on which they can be further broken, if necessary. At the bottom of
the hopper t is a jet d, through which, in the direction of the discharge
pipe R, a stream of water at a pressure of 150 to 200 lb. per square
inch is forced, and carries the ashes with it through the pipe discharging
them overboard at least 3 feet above the water line. The pipe can
be closed by a flap valve k, so that water cannot find its way to the
stokehold in bad weather. The upper part of the pipe in the bend
is formed of a separate piece s, which can be easily renewed when
worn out. The water under pressure is delivered by the auxiliary feed,
or other special pump.

To prevent water finding its way from the ash-ejector into the stoke-
hold, the water pressure must be turned on suddenly. This may be
efiected by means of an ordinary cock, a valve, or an adjustable
differential piston.*

In the latter arrangement (see Fig. 535) the water under pressure
enters from the pump at a. The spaces g and h are placed in communi-
cation through openings e and f in the hollow piston. In the position
shown in the drawing, the water passes through e and f to pipe c, which

• Patentee Howaldtswerke, Kiel.

VARIOUS DKTAILS. 623

leads overboard through a cock and non-return valve. If the regulating
cock a be turned through 90° so that passages/ and rare made to com-
municate, the water under pressure drives ihe piston, with diameter D,
to the right, and closes pipe c. Simultaneously pipe B is opened, and

Fig. 534.

the water led off to the jet in the ash-ejector. Diameter of nozzle of jet
^ to J inch. Diameter of discharge pipe, 4 to 6 inches. Thickness
about ^ to 1 inch. Inclination of the discbai^e pipe to the horizontal,
about 60°. As the bent upper part of this pipe requires frequent

624 MARINE KNGINES AND BOILERS.

renewal, it must be easily accessible. Diameter of hopper at the top,
about 14 lo It* inches. Diameter of delivery pipe Trom the pump,
about 2 to 2} inches. Number of ash-ejectors, the same as of steam
ash hoists.

:; 366. Ventilation of the En^ne and Boiler Rooms. —The
engine and boiler rooms of merchant vessels are almost always venti-
lated by the natural draught produced by the motion of the ship, or

by the force of the wind. The air is led below by means of thin
iron pipes of large diameter, fitted at their upper ends with revohing
cowls.

With a velocity of air of about 10 feet per second, the cress sution
tf the ventilaters should be sufficient, to deliver into the stokehold
about 350 cubic feet of air per pound of coal burnt. Besides the
engine-room hatch, which has to be closed in bad weather, special
ventilators with cowls are fitted.

g 366a. Area of Engine-room Ventilators. —-If g be the floor

VARIOUS DETAILS.

625

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626 MARINE ENC.INES AND BOILERS.

space of the engine-room, a the total area of all the ventilators leading
into it, then —

A = T^ to -% generally = -^ to

160 80' ^^ ' 120 100

These values naturally vary considerably, according to the different sizes
of the engine-room. In the case of very small engines, the skylight
alone is often sufficient for ventilating purposes.

§ 367. Ventilation of the Engine and Boiler Rooms in

Warships. — In warships, if the boilers are worked with the closed
stokehold system, the boiler-rooms are ventilated by the fans delivering
air under pressure into the stokehold (compare page 567, Closed
Stokehold System). The engine-rooms of warships are also ventilated
by special fans, which deliver air, where specially required, by means
of ventilators made of thin plating. Table No. 74, compiled from the
latest arrangements used in the German Imperial Navy, gives details
of the dimensions of these fans.

§ 368. German Lloyd's Rules for Spare Gear for Eng^ines

and Boilers.

Steamers registered at the German Lloyd's must have on board
the following spare gear for engines and boilers : —

(a.) Steamships classed \s and Atl. {Ocean-going and Atlantic Steamers),

1 set of crank-pin brasses.

1 set of crosshead pin brasses.

1 set of bolts and nuts for crank-pin end of connecting rod.

1 „ „ crosshead end „

1 set of main-bearing bolts and nuts.

1 set of bolts for each size of couplings.

1 set of springs and rings for each piston.

1 valve rod for each cylinder. If all the valve rods are of the same size,

one spare rod is suff.cient.
1 air-pump piston rod or connecting rod.
1 circulating-pump piston rod or connecting rod.
1 set of valves for air pump.
1 set of valves for circulating pump. If there is a centrifugal instead of

a plunger circulating pump, brasses and bolts for the latter of every

size, as well as a valve rod, piston, and piston rod.
1 set of feed-pump valves. If the pumps are of the same size, one set

to every two pumps.
1 set of bilge-pump valves. If the pumps arc of the same size, one set

to every two pumps.
1 set of feed-check valves.
1 complete set of straps to each pump and lever.

1 safety-valve spring if all the safety valves are of the same size ; if not,

one spring to each valve.

2 */^ of condenser tubes and ferrules.
2 '/^ of boiler tubes.

6 sets of water-gauge glasses with packing rings, for each boiler.
\ set of fire bars.

10 "*/, each of bolts and studs for the air and circulating pumps with
their nuts ; half a bundle of hoop iron ; one \ inch thick and one

628 MARINE ENGINES AND BOILERS.

I inch thick iron plates ; 2 bars of flat and 3 of round iron : half
bar of steel ; one dozen each of screws, nuts, and washers of i
inch, ij inch, -; inch, and 1 inch diameter respectively.

To every 50 boiler tubes, including stay tubes, one tube stopper.

Of engineer's tools, sufficient to execute small repairs on board.

It is also recommended to have in reserve — 1 crank shaft; 1 tail
shaft; 1 propeller or half set of blades, if the latter are not cast in
one with the boss.

(d,) Steamships Class K (Large Coasting Steaffiers),

1 set of crank-pin bolts and nuts.

1 „ crosshead „

1 „ main-bearing „

1 valve rod for each cylinder. If all the valve rods are of the same

size, one is sufficient.
1 set of air-pump valves.
1 „ circulating-pump valves.

1 „ feed-pump valves.

If the pumps are all of the same size, one set of valves to every

two pumps.
Piston bolts, set screws, bolts, and nuts for the air and circulating pumps :

1^ 7« of each of the number in use.

2 ^1^ of the number of condenser tubes, with packing for the tube ends.
4 sets of water-gauge glasses with packing rings, for each boiler.

\ set of fire bars.

1 boiler-tube stopper to every 50 boiler tubes.

A stock of flat and round iron bars, &c., and the necessary engineers
tools.

(r.) Steamships Class k {Small Coasting Steamers).

Bolts for every size of coupling.
1 crank-pin bolt.
1 crosshead bolt.

1 main- bearing bolt.

Half set of circulating and air pump valves, of bilge and feed pump
valves, to each pump, or if all the pumps are of the same size,
one set to every two pumps.

2 sets of water-gauge glasses with packing rings, for each boiler.
J set of fire bars.

2 boiler-tube stoppers.

I

§ 369. Lloyd's Rules for Spare Gear.

Spare gear is not required in steam yachts. The spare gear men-
tioned in the following list will be required to be carried in all steam
vessels classed in the Society's Register Book, viz. : —

2 connecting-rod or piston-rod top end bolts and nuts.

2 „ ,, bottom end bolts and nuts.

2 main-bearing bolts.

I set of coupling bolts.

1 set of feed and bilge pump valves.

1 set of piston springs (where common springs are used).

A quantity of assorted bolts and nuts.

Iron of various sizes.

In addition to the foregoing, it is recommended that the following
articles be carried, with a view to expedite repairs and lessen delay in
distant ports, viz. : —

Crank shaft.

Propeller shafc.

Propeller, or full set of blades.

Stern bush or lignum vitae lining

for bush.
1 pair of connecting-rod brasses.
1 pair of crosshead brasses.
1 set of link brasses.
1 eccentric strap complete.
Air pump rod.

Circulating pump rod.
H.p. valve spindle.
L.P. valve spindle.

1 set of check valves.
6 cylinder cover bolts.
6 junk ring bolts.

4 valve chest cover bolts.

2 dozen condenser tubes.

1 cylinder escape valve and spring.
1 set of safety valve springs.

PART VIII.

VARIOUS TABLES.

632 MARINE ENGINES AND BOILERS.

LIST OF TABLES IN PART VI H.

TAni.B PACE

I. Squares, Cubes, Square Roots, Cube Roots, Recip-
rocals, Natural Logarithms, Circumferences,

Areas of Circles from i to 1,000 - - - a34

IL Common Logarithms from l to 100 - - - 074

in. Sines a^d Cosines - - - - - 677

IV. Tangents and Cotangents - - - - 679

V. Various Equivalents - - - - - 681

VL Cos ui + A cos 2<o ... - - . 681

VII. Inches and Millimetres - - - - 68f

VIII. Square Metres and Square Feet - - - 686

IX. Square Feet and Square Metres - - - 687

XI. Pounds and Kilogrammes - - - - 688

XII. Kilogrammes and Pounds - - - - 689

XIII. Pounds per Square Inch and Kilogrammes per

Square Centimetre - - - - - 690

XIV. Kilogrammes per Square Centimetre and Pounds

PER Square Inch - - - - - 691

XXI. Comparison of Thermometers - - - 69:^

XXII. Properties of Saturated Steam - - - 694

XXIII. Expansion of Rigid Bodies by Heat - - - 698

XXIV. Melting Points of Various Materials - - 698

XXVa. Specific Gravity of Woods - - - - 699

XXVb. Specific Gravity of Metals - - - - 699

XXVc. Specific Gravity of Various Materials - - 700

XXVd. Relative Weights of Coals - - - - 700

XXVe. Specific Gravity of Fluids - - - - 700

XXVf. Specific Weights of Gases at 30 Inches of Mer-
cury AND 32^ Fahr. - - - - - 701

XXVI. Strength and Elasticity of Various Materials - 702

XXVI I. Strength and Elasticity of Manganese Bronze - 704

XXX. Moments of Inertia " i " and Internal Moments of
Resistance or Moduli of Section "w" for Cir-
cular Sections of Diameter"//" - - - 705

LIST OF TABLES. 638

TABLE PAGE

XXXL Bending Moments - - - - - 707

XXXII. Torsional SxRENCiXH - - - - - 711

XXXIII. Strength of Struts ----- 712

XXXIV. German Lloyd's Rules for Iron and Steel for

Boilers ------- 713

XXXV. German Lloyd's Rules for Steel and Cast Steel

FOR Parts of Engines . _ - - 714

XL. Weight of Machinery ----- 715

XLI. Weight of Boiler Equipments Compiled from the

German Navy ------ 716

XLIl. Weight of Cylindrical Boilers - - - 716

N.B. — The missing Tables refer to data which it has not been thought
necessary to include in the English Edition.

As the German plates have been used for Tables I. to XXL, the decimal
points are indicated by commas.

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Oi Ok Ok Ok Ok Ok

i

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Ok O* ^ W W Ok ^ I

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sis

ill:

674

MARINE ENGINES AND BOILERS.

Table No. II.
Cominon Logaritfams from z to zoo.

Nr.

0

1

2

s

4

6

8

7

8

9

Dlf.
tcnnoa.

10
11
12
13
14

0000
0414
0792
1139
1461

0043
0453
0828
1173
1492

0086
0492
0864
1206
1523

0128
0531
0899
1239
1553

0170
0569
0934
1271
1584

0212
0607
0%9
1303
1614

0253
0645
1004
1335
1644

0294
0682
1088
1367
1673

0334
0719
1072
1399
1703

0374
0756
1106
1430
1732

43—40

39-36
36-34
34-31
31-29

15
16
17
18
19

1761
2041
2304
2553

2788

1790
2068
2330
2577
2810

1818
2095
2355
2601
2833

1847
2122
2880
2625

2856

1875
2148
2405
2648

2878

1903
2175
2430
2672
2900

1931
2201
2455
2695
2923

1959
2227
2480
2718
2945

1987
2253
2504
3742

2967

2014
2279
2529
2765
2989

29—27
27—26
26-25
24-23
28—22

20
21
22
23
24

3010
3222
3424
3617
3802

3032
3243
3444
3636
3820

3054
3263
3464
3655
3838

3075
3284
3483
3674
3856

3096
3304
3502
3692
3874

3118
3824
3522
3711
3892

3139
3345
3541
3729
3909

3160
3365
3560
3747
3927

3181
3385
3579
3766
3945

3201
3404
3598
3784
3962

22—20
21—19
20—19
19-18
ia-17

25

26
27

28
29

3979
4150
4314
4472
4624

3997
4166
4330
4487
4639

4014
4183
4346
4502
4654

4031
4200
4362
4518
4669

4048
4216
4378
4533
4683

4065
4232
4393
4548
4698

4082
4249
4409
4564
4713

4099
4265
4425

4579
4728

4116
4281
4440
4594
4742

4133
4298
4456
4609
4757

18—17
17—16
16—15
16-15
15—14

30
31
32
33
34

4771
4914
5051
5185
5315

4786.
4928
5065
5198
5328

4800
4942
5079
5211
5340

4814
4955
5092
5224
5353

4829
4%9
5105
5237
5366

4843
4983
5119
5250
5378

4857
4997
5132
5263
5391

4871
5011
5145
5276
5403

4886
5024
5159
5289
5416

4900
5038
5172
5302
5428

15—14
14—13
14-13

18
18—12

35
36
37
38
39

5441
5563
5682
5798
5911

5453
5575
5694
5809
5922

5465
5587
5705
5821
5933

5478
5599
5717
5832
5944

5490
5611
5729
5843
5955

5502
5623
5740
5855
5966

5514
5635
5752
5866
5977

5527
5647
5763
5877
5988

5639

5658
5776
5888
5999

5551
5670
5786
5899
6010

13—12
12—11
12—11
12—11
12—11

VARIOUS TABLES.

675

Table Na II.
Common Logarithms from z to zoa

Nr.

0

1

2

8

4

6

«

7

8

9

Dif.
fwvDoe.

40
41
42
43
44

6021
6128
6232
6335
6436

6081
6138
6243
6345
6444

6042
6149
6253
6355
6454

6053
6160
6263
6365
6464

6064
6170
6274
6375
6474

6075
6180
6284
6385
6484

6085
6191
6294
6895
6493

6096
6201
6304
6405
6503

6107
6212
6314
6415
6513

6117
6222
6325
6425
6522

11-10
11—10
11—10

10
10—9

45

46
47
48
49

6532
6628
6721
6812
6902

6542
6637
6730
6821
6911

6551
6646
6739
6830
6920

6561
6656
6749
6839
6928

6571
6665
6758
6848
6937

6580
6675
6767
6857
6946

6590
6684
6776
6866
6955

6599
6693
6785
6875
6964

6609
6702
6794
6884
6972

6618
6712
6803
6893
6981

10—9
10—9
10-9

9
9-8

50
51
52
53
54

6990
7076
7160
7243
7824

6998
7084
7168
7251
7332

7007
7093
7177
7259
7840

7016
7101
7185
7267
7348

7024
7110
7193
7275
7356

7033
7118
7202
7284
7364

(7042
7126
7210
7292
7372

7050
7135
7218
7300
7380

7059
7143
7226
7308
7388

7067
7152
7235
7316
7396

9-8
9-8
9-8
9-8
8

55
56
67
58
59

7404
7482
7559
7634
7709

7412
7490
7566
7642
7716

7419
7497
7574
7649
7723

7427
7505
7582
7657
7781

7435
7513
7589
7664
7738

7443
7520
7597
7672
7745

7451
7528
7604
7679
7752

7459
7536
7612
7686
7760

7466
7543
7619
7694
7767

7474
7551
7627
7701
7774

8—7
8-7
8—7
8-7
8—7

60
61
62
63
64

7782
7853
7924
7993
8062

7789
7860
7931
8000
8069

7796
7868
7938
8007
8075

7808
7875
7945
8014
8082

7810
7882
7952
8021
8089

7818
7889
7959
8028

QAQfi

7825
7896
7966
8035
8102

7832
7903
7978
8041
8109

7839
7910
7980
8048
8116

7846
7917
7987
8055
8122

8-7
8-7
7-6
7-6
7—6

65
66
67
68
69

8129
8195
8261
8325
8388

8136
8202
8267
8331
8395

8142
8209
8274
8338
8401

8149
8215
8280
8344
8407

8156
8222
8287
8351
8414

8162
8228
8293
8357
8420

8169
8235
8299
8363
8426

8176
8241
8306
8370
8432

8182
8248
8812
8376
8439

8189
8254
8319
8882
8445

7-6
7-6
7—6
7-6
7-6

676

MARINE ENGINES AND BOILERS.

Table No. II.
Common Logaritlims from z to xoa

Nr.

0

1

8

8

4

5

6

7

8

9

f«ren.««.

70
71
72
78
74

8451
8518
8573
8633
8692

8457
8519
8579
8639
8698

8463
8525
8585
8645
8704

8470
8531
8591
8651
8710

8476
8537
8597
8657
8716

8482
8543
8603
8663
8722

8488
8549

8669
8727

8494
8555

8615
8675
8733

8500
8561
8621
8681
8739

8506
8567
8627
8686
8745

7-6
7-6

6
6-6
6-^

75
76
77
78
79

8751
8808
8865
8921
8976

8756
8814
8871
8927
8982

8762
8820
8876
8932
8987

8768
8825
8882
8938
8993

8774
8831
8887
8943
8998

8779
8837
8893
8949
9004

8785
8842
8899
8954
9009

8791
8848
8904
8960
9015

8797
8854
8910
8965
9020

8802
8859
8915
8971
9025

6-6
6-5
6-5
6-5
6—6

80
81
82
88
84

9031
9085
9138
9191
9243

9036
9090
9143
9196
9248

9042
9096
9149
9201
9258

9047
9101
9154
9206
9258

9053
9106
9159
9212
9263

9058
9112
9165
9217
9269

9063
9117
9170
9222
9274

9069
9122
9175
9227
9279

9074
9128
9180
9232
9284

9079
9133
9186
9238
9289

6-6
6—6
6-5
6-6
6-6

85
86
87
88
89

9294
9345
9395
9445
9494

9299
9350
9400
9450
9499

9304
9355
9405
9455
9504

9309
9360
9410
9460
9509

9315
9365
9415
9465
9513

9320
9370
9420
9469
9518

9325
9375
9425
9474
9523

9330
9380
9430
9479
9528

9335
9385
9435
9484
9533

9340
9390
9440
9489
9538

6-5

5

5
5-4
5-4

90
91
92
93
94

9542
9590
9638
9685
9731

9547
9595
9643
9689
9786

9552
9600
9647
9694
9741

9557
9605
9652
9699
9745

9562
9609
9657
9703
9750

9567
9614
9661
9708
9754

9571
9619

Ckti0fi
uOOO

9713
9759

9576
9624
9671
9717
9763

9581
9628
9675
9722
9768

9586
9633
9680
9727
9773

5-4
5-4
5-4
5-4
6-4

95
96
97
98
99

9777
9823
9868
9912
9956

9782
9827
9872
9917
9961

9786
9832
9877
9921
9965

9791
9886
9881
9926
9969

9795
9841
9886
9930
9974

9800
•K54o
9890
9934
9978

9805
9850
9894
9939
9983

9809
9854
9899
9943
9987

9814
9859
9903
9948
9991

9818
9863
9908
9952
9996

6-4
6-4
5-4
6-4
6-4

/

VARIOUS TABLES.

677

Table No. III.— Sines and Cosines.

C o « i n e 8

C

wr

2y

w

-w

«r

-w

0

1

2
3

4

5
6
7
8
9

11
12
13
14

1^0000

0,99d4
0i99o6
0.9976

VXSSS
0,9998
0,9993
0,9985
0,9974

0,9962
0,9945
0,9926
0,9903
0,9877

0,9959
0,9942
0,9922
0,9899
0,9872

0,9811
0,9776
0,9737
0,9696

1,0000
0,9997
0,9992
0,9983
0,9971

0,9939
0,9918
0,9894
0,9868

1,0000
0,9907
0,9991
0,9981
0,9969

0,9999
0,9996
0,9989
0,9980
0,9967

0,9999
0,9995
0,9988
0,9978
0,9964

0,9954
0,9936
0,9914
0,9890
0,9863

0,9833
0,9799
0,9763
0,9724
0,9682

0,9951
0,9932
0,9911
0,9886
0,9858

0,9929
0,9907
0,9881
0,9853

0,9999
0,9994
0,9986
0,9976
09962

0,9822
0,9788
0,9750
0,9710
0,9667

0,9926
0,9903
0,9877
0,9848

Tj;5gIS

0,9782
0,9744
0,9703
0,9659

89^

88

87

86

85

83
82
81
80

75"
78
77
76
75

0,9848
0,9816
0,9782
0,9744
0,9703

0,9659
0,9613
0,9563
0,9511
0,9455

0,9652
0,9605
0,9555
0,9502
0,9446

0,9838
0,9805
0,9769
0,9730
0,9689

0,9636
0,9588
0,9537
0,9483
0,9426

0,9827
0,9793
0,9757
0,9717
0,9674

0,9580
0,9528
0,9474
0,9417

73
72
71
70

15
16
17
18
19

20
21
22
23
24

26
27
28
29

0,9397
0,9336
0,9272
0,9205
0,9136

0,9887
0,9325
0,9261
0,9194
09124

0,9051
0,8976
0,8897
0,8816
0,8732

0,9644
0,9596
0,9546
0,9492
0,9436
0,5877
0,9315
0,9250
0,9182
0,9112

0,9038
0,8962
0,8884
0,8802
0,8718

0,9367
0,9304
0,9239
0,9171
0,9100

0,9026
0,8949
0,8870
0,8788
0,8704

0,9357
0,9294
0,9228
0,9169
0,9088

0,9621
0,9572
0,9620
0,9466
0,9407

0,9013
0,8936
0,8857
0,8774
0,8689

0,9346
0.9283
0,9216
0,9147
0,9076

0,9613
0,9663
0,9611
0,9466
0j^9397

0.9001
0,8923
0,8843
0,8760
0.8676
T^555T
0,8496
0,8403
0,8307
0,8208

0,9336
0,9272
0,9206
0,9136
0,9063

0,8988
0,8910
0,8830
0,8746
0,8660

69
68
67
66
65

31
32
33
34

36
37
38
39

0,9063
0,8988
0,8910
0,8830
0,8746

0,8660
0,8572
0,8481
0,8387
0,8290

0,8646
0,8567
0,8466
0,8371
0,8274

0,8631
0,8642
0,8460
0,8366
0,8268

0,8616
0,8526
0,8434
0,8339
0,8241

0,8602
0,8611
0,8418
0,8323
0,8225

T53575
0,8481
0,8387
0,8290
0,8192

^4
68
62
61
60

68
57
66
55

0,8193
03090
0,7986
0,7880
0,7772

0,7660
0,7547
0,7431
0,7314
0,71H3

0,7071

0,8175
0,8073
0,7969
0,7862
0,7763

0,7628
0,7412
0,7294
0,7173

0,8158
0,8056
0,7951
0,7844
0,7735

0,7609
0,7392
0,7274
0,7163

0,8141
0,8089
0,7934
0,7826
0,7716

0.7490
0,7373
0,7264
0,7133

0,8124
0,8021
0,7916
0,7808
0,7698

0,8107
0,8004
0,7898
0.7790
0,7679

0,8090
0,7986
0,7880
0,7772
0,7660

"BT

53

52

51

50

48
47
46
45

40
41
42
43
44

0,7585
0,7470
0,7363
0,7234
0,7112

0,7566
0,7451
0,7333
0,7214
0,7092

0,69flS

0,7431
0,7314
0,7193
0,7071

^b

TSJCBT

0,7009

0,6988

^H

w

«r

Aff

or

'W

IT

Sines

^.B. — Decimal points indicated by commas.

678

MARINE ENGINES AND BOILERS.

Table No. III.— Sines and Costnes.

Sines

ly

Qtr

0,0058
0,0283
0,0407
0,0681
0,0766

w

0,0087
0,0262
0,0486
0,0611
0,0786

nor

0.0116
0,0291
0,0466
0,0640
0,0814

my

w

0
1
2
8

4

T
6
7
8
9

0,0000
0,0176
0,0849
0,0628
0,0698

0,1046
0,1219
0,1892
0.1664

0,0029
0,0204
0,0878
0,0662
0,0727

0,0146
0,0820
0,0494
0,0669
0,0648

TJ75m

0,0349
0,0628
0,0698
0,0672

89^

88

87

86

86

11

12
18
14

IF
16
17

18
19

20
21
22
28
24

26
27

28
29

0,1908
0,2079
0,2260
0,2419

0,0901
0,1074
0.1248
0,1421
0,1698

0,2766
0,2924
0,8090
0,3266

0,1765
0,1937
0,2108
0,2278
0,2447

0,8684
0,8746
0,8907
0,4067

0,2784
0,2962
0,8118
0,8283

0,3448
0,8611
0,8773
0,3934
0,4094

0,4263
0,4410
0,4666
0,4720
0,4874

0,0930
0,1108
0,1276
0,1449
0,1622
THTST
0,1966
0,2136
0,2306
0,2476

0,0959
ail82
0,1806
0,1478
0,1661

0,2812
0,2979
0,8146
0,8311

T5317F
0.8638
0,3800
0,8961
0,4120

0,1822
0,1994
0,2164
0,2836
0,2604

0,1161
0,1834
0,1607
0,1679

0,1016
0,1190
0^363
0,1686
0,1708

JKUm
0,1219
0,1392
0,1664
0,1737

0,1851
0,2022
0,2193
0,2363
0,2632

0,2672
0,2840
0,3007
0,3173
0,3338

0,2700
0,2868
0,3036
0,3201
0,3366

0,3502
0,3666
0,3827
0,3988
0,4147

0,3529
0,3692
0,3864
0,4014
0,4178

0,1880
0,2061
0,2221
0,2391
0,2660

's:ms

0,2896
0,3068
0,8228
0,8393

0,1908
0,2079
0,2260
0,2419
0,2688

0,3557
0,3719
0,3881
0,4041
0,4200

0J3766
0,2924
0,3090
0,3266
0^3420

0,3746
0,3907
0,4067
0,4226

83
82
81
80

IS-
IS
77
76
76

74
73
72
71
70

W

68

67

66

66

0,4226
0,4384
0,4640
0,4696
0,4848

0,4436
0,4692
0,4746
0,4899

T),4305
0,4462
0,4618
0,4772
0,4924

o]5m

0,6226
0,5373
0,6619
0,6664

0,4331
0,4488
0,4643
0,4797
0,4960

0,5100
0,6260
0,5398
0,6644
0,6688

0,4358
0,4614
0,4669
0,4828
0.4976

0,5125
0,6276
0,6422
0,6668
0,6712

0,4384
0,4540
0,4696
0,4848
0,6000

0,5150
0,5299
0,6446
0,6692
0,6736

63
62
61
60

90
81
82
88
84

ofioo(y

0,5150
0,6299
0,6446
0,6692

0,5736
0,6878
0,6018
0,6167
0,6293

0,5025
0,6176
0,6824
0,6471
0,6616

0,6901
0,6041
0,6180
0,6816

0,6050
0,6200
0,6348
0,6496
0,6640

0,6926
0,6066
0,6202
0,6338

0,5948
0,6088
0,6226
0,6861

0,6018
0,6167
0,6298
0,6428

68
67
66
66

^5
86
87
88
89

0,6661
0,6691
0,6820
0,6947

JJWT

0,6688
0,6713
0,6841
0,6968
0,7092

0,6604
0,6734
0,6862
0,6988
0,7112

0,5851
0,6972
0,6111
0,6248
0,6383

0,6648
0,6777
0,6906
0,7030

0,5854
0,6996
0,6134
0,6271
0,6406

0,6539
0,6670
0,6799
0,6926
0,7061

0,6691
0,6820
0,6947
0,7071

0,7193

3r

68
62
61
60

ID
41
42
43
44

45

0,6495
0,6626
0,6766
0,6884
0,7009

W

0,7133

0^7158

wzi

Oj7173

ir

48
47
46
46

4f

Dtgne

mr

4ff

wr

ir

C o • { n e •

Dogiw

N.B. — Decimal points indicated by commas.

VARIOUS TABLES.

679

Table No. IV.— Tangents and Cotangents.

Cotana^nts

0'

IC

49,1089
26,4816
18,0750
13,7267

42,9641

24,5418
17,1698
18,1969

^

38,1885
22,9038121
16,8499
12,7062

iff

85,9898
34,3678
,4704
15,6048
12,2505

w

w

a
1

2
3
4

6

7
8
9

57,9900
28,6862
19,0811
14,3007

9,5144
8,1444
7,1154
6,8138

9,2553
7,9530
6,9682
6,1970

10,7119
9,0098
7,7704
6,8269
6,0844

107J7S5
8,5556
7,4287
6,5606
5,8708

68,7501
31,2416
20,2056
14,9244
11.8262

8,3450
7,2687
6,4348
5,7694

S7;S9QS

28,6362
19,0611
14,3007
11,4300

9,5144
8,1444
7,1154
6,3138
5,6718

88
87
86
85

8r

83

82
81
80

11
12
13
14

5,6713
5,1446
4,7046
4,8315
4,0108

10,3854
8,7769
7,5958
6,6912
5,9758

■75"

78

77

76

75

IS

16
17

18
19

^77551
3,4874
8,2709
3,0777
2,9042

5,5764
5,0658
4,6383
4,2747
3,9617

8,6891
3,4495
3,2371
3,0475
2,8770

5,4845
4,9894
4,5736
4,2193
3,9136

3,4124
3,2041
3,0178
2,8602

5,8955
4,9152
4,5107
4,1653
3,8667

5,3098
4,8430
4,4494
4,1126
3,8208

■g;5S5g

3,3402
3,1397
2,9600
2,7980

5,2257
4,7729
4,3897
4,0611
3,7760

5,1446
4,7046
4,3315
4,0108
3,7321

■p573
3,2709
3,0777
2,9042
2,7475

21
22
23
24

2:7475
2,6051
2,4751
2,3559
2,2460

■5;725S
2,5826
2,4545
2,3369
2,2286

2,6985
2,5605
2,4342
2,3188
2,2113

8,6059
3,3759
3,1716
2,9887
2,8239

8,5261
3,3052
3,1084
2,9319
2,7725

li

73

72

71

70

■^
26
27
28
29

2,1283
2,0853
1,11480
1,8676
1,7917

2,6746
2,5387
2,4142
2,2998
2,1943

2,6511
2,5172
2,3945
2,2817
2,1775

2,6279
2,4960
2,3750
2,2637
2,1609

2,6051
2,4751
2,3559
2,2460
21445

69
68
67
66
65

2,1445
2,0503
1,9626
1,8807
1,8041

1:7205
1,6534
1,5900
1,5301
1,4733

2,1123
2,0204
1,9347
1,8546
1,7796

1,7090
1,6426
1,5798
1,5204
1,4641

2,0965
2,0057
1,9210
1,8418
1,7675

1,6977
1,6319
1,5697
1,5108
1,4550

"2:5555
1,9912
1,9074
1,8291
1,7556

1,9768
1,8940
1,8165
1,7438

2,0503
1,9626
1,8807
1,8041
1,7321

63
62
61
60

m

81
82
83
84

36
87
38
39

1,7321
1,6643
1,6003
1,5399
1,4826

1,6864
1,6213
1,5597
1,5013
1,4460

1:5755
1,6107
1,5497
1,4919
1,4370

1,6003
1,5399
1,4826
1,4282

59
58
57
56
55

15-
41
42
43
44

1,4282
1,3764
1,3270
1,2799
1,2349

1,4193
1,3680
1,3190
1,2723
1,2276

1,4106
1,3597
1,3111
1,2647
1,2203

1,4020
1.3514
1,3032
1,2572
1,2131

1,3934
1,3432
1,2954
1,2497
1,2059

X3848
1,3351
1,2876
1,2423
1,1988

1,3764
1,3270
1,2799
1,2349
1,1918

53
52
51

50

XlSOJ
1,1106
1,0724
1,0355
1,0000

1,1918
1,1504
1,1106
1,0724
1,0355

1,1847
1,1436
1,1041
1,0661
1,0295
0,9942

1,1778
1,1369
1,0977
1.0599
1,0236

1,1709
1,1303
1,0913
1,0538
1,0176

1,1640
1,1237
1,0850
1,0477
1,0117
0.9770

1,1572
1,1171
1,0786
1,0416
1,0068
0,9713

0,9667

49
48
47
46
45
IT

A5

1.0000

0,9884

0,9827

5V

IRT

20^

10^

V

Tangents

N*B, — Decimal points indicated by commas.

680

MARINE ENGINES AND BOILERS.

TaUe No, IV.-

Tmncenbi and Cotangeata.

Dmvm

Tangents |

Dcgrw

C

lar

Qtr

sc

w

60^

60*

0

0,OuOO

0,00s»

0,00&8

0,0087

0,0116

0.014fi

0,017K

M

1

0,0176

0,0204

0,0288

0,0262

0,0291

0,0320

0,0349

88

2

0,0849

0,0878

0,0408

0,0487

0,0466

0,0496

0,0624

87

8

0,0524

0,0668

0,0682

0.0612

0,0641

0,0670

0,0699

86

4

0,0699

0,0729

0,0758

0,0787

0,0816

0,0846

0,0876

85

6

0,0876

0,0904

0.0d34

0,0963

0,0092

0,1022

0,1061

84

6

0,1061

0,1081

0,1110

0,1189

0,1169

0,1198

0,1228

83

7

0,1228

0,1267

0,1287

0,1317

0,1846

0,1876

0,1405

82

8

0,1406

0,1486

0,1466

0,1496

0,1624

0,1554

0,1584

81

9

0,1684

0,1614

0,1644

0,1678

0,1703

0,1733

0,1763

80

Id

OliTftJ

0,17ftS

0,1823

0,1863

0,1884

0,l9l4

0^944

79

11

0,1944

0,1974

0,2004

0,2035

0,2066

0,2095

0,2126

78

12

0,2126

0,2166

0,2186

0,2217

0,2248

0,2278

0,2309

77

18

0,2809

0.2389

0,2870

0,2401

0,2432

0,2462 0,2498

76

U

0,2498

0,2624

0,2556

0,2686

0,2617

0,2648 0,2680

76

Ih

0,3fi80

0,27ll

0,8742

0,2773

0,2806

0,2836

0,2868

*4

16

0,2868

0,2899

0,2981

0,2962

0,2994

0,3026

0.3067

73

17

0,3067

0,3089

0,3121

0,8163

0,3185

0.3217

0,3249

72

18

0«8249

0,8281

0,3314

0,8346

0,3378

0,8411

0,3448

71

19

0,8448

0,3476

0,3609

0,3541

0,3674

0,3607

0.3640

70

^

0,3^

0,3^73

15|570B^

0,3739

0,3772

0,8806

0,8839

69

21

0.8839

0,8872

0,3906

0,3939

0,8978

0,4007

0,4040

68

22

0,4040

0,4074

0,4108

0.4142

0,4176

0,4211

0,4246

67

28

0,4245

0,4279

0,4314

0,4348

0,4388

0,4418

0,4462

66

24

0,4452

0,4487

0,4522

0,4557

0,4592

0,4628

0,4663

65

%

0,46&i

0,4fidd

0.4734

0,4770

0,4806

0,4841

0,4877

64

26

0.4877

0,4913

0,4960

0,4986

0,6022

0,5059

0,5096

63

27

0,6095

0.6182

0,6169

0,5206

0,6243

0,6280

0,5317

62

28

0,6817

0,6366

0,6392

0,6430

0.5467

0,5506

0.6543

61

29

0,6648

0.6681

0,5619

0.5668

0,66%

0,6785

0,5774

60

SO

0,6774

0,68l&

0,6861

0,68dl

0,6930

0,6969

0,6009

59

81

0.6009

0,6048

0,6088

0,6128

0,6168

0,6208

0,6249

58

82

0,6249

0,6289

0,6830

0,6871

0,6412

0,6463

0,6494

57

88

0,6494

0,6586

0,6677

0,6619

0,6661

0,6703

0,6746

56

84

0,6746

0.6788

0,6830

0,6873

0,6916

0,6969

0,7002

66

3&

0,7008

0,7046

0,7089

0,7183

0,7l77

0,7221

0,7266

54

86

0,7266

0,7810

0,7855

0,7400

0.7446

0,7490

0,7536

58

87

0,7686

0,7681

0,7627

0,7673

0,7720

0,7766

0,7818

52

88

0,7818

0,7860

0,7907

0,7964

0,8002

0,8050

0,8098

61

89

0,8098

0,8146

0.8196

0.8243

0,8292

0.8342

0,8391

60

40

0,8391

0,8441

0,8491

0,8641

0,8691

0,8642

0,8693

49

41

0,8698

0,8744

0,8796

0.8847

0,8899

0,8962

0,9004

48

42

0,9004

0,9067

0,9110

0,9163

0,9217

0,9271

0,9326

47

43

0,9826

0,9880

0,9486

0,9490

0,9646

0,9601

0,9657

46

44

0,9667

0,9718

0,9770

0,9827

0,9884

0,9942

1,0000

45

45 1.0000

1.0068

1.0117

1,0176

1.0236

10296

1.0856

44

-Davma

60' 1

W 40- 1

^'

56^ Id' 1 O' 1

TU««M

Cotangents

J\r.B. — Decimal points indicated by commas.

^

CO

X

to
« lo 'I!

00 o» S
_ ©51 « 2

oSg^kS 8

.n B n I

VARIOUS TABLES.

!^

GO 00

C9 o

CO

c*5

8 eo ^ ^

1. »>

S -2

1

m

>

O

s

11 II

CD *^

1-1 r-

* «

«o "^ a

^ C» CO

^ ^ 00

CQ O 0>

II II II

o

8

+

I

>

I

d

I

681

!

II

r.j'S

il

tH|"^

'Hi*©
II

II

II

II

II
II

d

3

-H

3
3 a

3
3.

§8

a

O

o

m

8

04

m

.8

09

d

o

®S8S

0 C» CO t«
O CO CO O

01 1-H d^ e*»

_r _r <-? i-r

iH v^ o o

+++-H

22 ^ *• 5r
M 2 t« O

<N iH 0> t«
vH iH O O

++ + +

u5 oO o> O

8 rS 8 8

o* o- o" ©•

+ + +

Q3 M vH O
$9 O) iH S

0« r-t tH O
+ + +

09 iH

+ + +

+++

++++

®ss«

S c« s

++ 1

0> t« 09

S8i

cT o- ©•

++ 1

*o 09 O
f« ^ 2

« d «

++ 1

8 CO Q

iH vH ^

I I I

iH M o;
o* o* o'

I I I

iH e« «
©*d"d

I I I

I I I

+ +

SiS8

sas

09 Q ^

o* ©- ©•

I I I

lO ^ O
^ CO l--

I I I

^§1

^ CO *•

©•• ©- ©'

I I I

o'o*

I I

tH tH O

I I

I I

CO O

o o

I I

cT© ©

I"
d

I I I

8S8

tH iH ri

S 3 8

«g

CO

t<* C* 00

o*^ d* d*
I I I

IS

^ 00
i> t- r-

I I I

T^ O* G
'^ '^ t&

r- r- t-

<S <:> <S

III ^

8

CO

+++

•1 e« ej

o" ©- o*
++ +

o o o

+++

+++

III

I I I

S 8 8

1

I

I

MARINE ENGINES AND BOILERS.
Table No. VII.

A^.^.— Decimal points indicated by commas.

VARIOUS TABLES.

683

Table No. VII.
Inches and Millimetres.

Indiet

mm

Inotacs

mm

Inohad mm

iBOhM

nun

Inehca

mm

26V,
2BV«
26»/,

26V.

25»/,

26»A
25'/,

688,16
641,34
644,51
647,69
650,86
654,04
657,21

SOVe
30V.

30V,

30*/.
30Vp

765,16
768,34
771,51
774,69
777,86
781,04
784,21

35%
35%
35%
35%
35%

35V4
36%

892,16
895,33
898,51
901,68
904,86
908,03
911,21

40V,
40V.
40'/,

40V.
40*/,
40'/.
40'/,

1019,2
1022,3
1025,5
1028,7
1031,9
1035,0
1038.2

46V,
46V.

45'/,

45V.
46V.
45'/.
46'/,

1146,2
1149,3
1152,5
1156,7
1158,9
1162,0
1165,2

26

660,39

31

787,39

36

914,38

41

1041,4

46

1168,4

26V4

26V,

aev.

26»/,

26V«
26'/,

663,56
666,74
669,91
673,09
676,26
679,44
682.61

31V8

3iy.

31V,

31V,
31Vs
3IV4

31%

790,56
793,74
796,91
800,09
803,26
806,44
809,61

36V4
36%
36%
36%
36%
36%

917,56
920,73
923,91
927,08
930,26
933,43
936,61

41V,
41V.
41'/,

41V.
41V,
41'/.
41'/,

1044,6
1047,7
1050,9
1054,1
1057,3
1060,4
1063,6

46V,
46V.
46'/,

46V.
46'/,
46'/.
46'/,

1171,6
1174,7
1177,9
1181,1
1184,3
1187.4
1190,6

27

685,79

32

812,79

37

939,78

42

1066.8

47

1193,8

27V,

27V4

27»

27V.

27V,

27V«
27'/,

688,96
692,14
695,31
698,49
701,66
704,84
708,01

32V.

32V4

32V,

32V,

^'/»
32V4

32%

815,96
819,14
822,31
825,49
828.66
831,83
835,01

37%
37%
37%

37V,
37%

37%

942,96
946,13
949,31
952,48
955,66
958,83
962,01

42V.
42V,
42»/,
42V,
42%
42'.
42'/,

1070,0
1073,1
1076,3
1079,5
1082,7
1085,8
1089,0

47V,

47V.
47'/,

47V,
47'/,

47'/.
47'/,

1197,0
1200,1
1203,3
1206,5
1209,7
1212,8
1216,0

28

711,19

33

838,18

38

965,18

43

1092,2

48

1219.2

28«/,

28V«
28»/,

28V.
28»/,
28»/«
28'/,

714,36
717,54
720,71
723,89
727,06
730,24
733,41

33V8

33%
38V8

33V,

33V8
38%

33%

841,36
844,53
847,71
850,88
854,06
857,23
860,41

38%

38V4
38%
38%
38%

38V4

38%

968,36
971.53
974,71
977,88
981,06
984,23
987,41

437.
43V.
43'/,

4av,

43V,
43'/.
43'/,

1095,4
1098,5
1101,7
1104,9
1108,1
1111,2
1114,4

48V,

48V.
48'/,

48V.
48'/,
48'/.
48'/,

1222,4
1225,5
1228,7
1231,9
1236,1
1238,2
1241,4

29

736,59

34

863,58

39

990,58

44

1117.6

49

1244,6

2»V,
29V«
29*/,

29V.

29»/.
29'/,

739,76
742,94
746,11
749,29
752,46
755,64
758.81

34Vg
34%
34%
34%
34%
34%
34V5

866,76
869,93
873,11
876,28
879,46
882,63
885,81

39%
39%
39Vg
39V,
39%
39%
39%

993,76
996,93

1000,1

1003.3
1006,5
1009,6
1012,8

44V,
44'/.
44'/.
44V,

44V,
44'/.

44'/,

1120,8
1123,9
1127,1
1130,3
1133,5
1136,6
1139,8

49V.
49V.

49V,
49V.
49V,
49*/.

49'/,

1247,8
1250,9
1254,1
1257.3
1260,5
1263,6
1266,8

30

761,99

35

888,98

40

1016,0

45

1143,0

50

1270,0

N.B.— Decimal points indicated by commas.

MARINE ENGINES AND BOILERS.

Tkbte No. VII.
inclici ■»H Uillinutnft.

\.B. — Decimal points indicated by comoiBS.

VARIOUS TABLES.

685

Table No. VII.
Inches and Millimetres.

Inches

mm iMbM

tn^

IndiM

mm laohM

mm Inohet

mm

76V,
76V«
76»/,

75'/.
76»/,

76»/4
76'/,

1908,1
1911,8
1914,6
1917,7
1920,8
1924,0
1927,2

80V,

8OV4
80»/,
80'/,
SO*/,

80»/.
80'/,

2085^
2038,4
2041,4
2044,6
2047,8
2061,0
2054,2

85V8
85V4
86%
85V,
85»/,

85»/4

85V,

2162,2
2165,4
2168,4
2171,6
2174,8
2178,0
2181,2

90V.

9OV4

90»/.

90V.
90»/.
90»/,
90V,

2289,2
2292,4
2295,4
2298,6
2301,8
2305,0
2308,2

96V.
96V4
95V.

95V,
95V.
95»/,

95V,

2416,2
2419,4
2422,4
2425,6
2428,8
2432,0
l2435^

76

1980.4

81

2057,4

86

2184,4

91

2311,4

96

2438,4

76'/,
76V4
76»/,

76V,
76»/,
76»/,
76V,

1983,6
1936,7
1989,9
1943,1

1946,3
1949,4
1962,6

81'/,

81«/4

81V,
81V«
81V,
8IV4
81'/,

2060,6
2063,8
2066,8
2070.0
2073,2
2076,4
2079,6

86Vg
86V4
86Vg
86V.
86Vg
66V4
86V,

2187,6
2190,8
2193,6
2197,0
2200,2
2203,4
2206,6

91V.
9IV4
91V.
91V,
91V.

9IV4
91V,

2314,6
2317,8
2320,8
2324,0
2327,2
2330,4
2333,6

96V.

96V.
96V,
96V.
96V4
96V,

2441,6
2444,8
2447,8
2451,0
2454,2
2457,4
2460,6

77

1966,8

82

2062,8

87

2209,8

92

2336,8

97

2463,8

77V.

77'/«

77V,
77'/,

77V,
77V«
77V,

1968,9
1962,1
1966,8
1968,6
1971,6
1974,8
1978,0

82'/,

82V4
82»/,
82'/,
82Vt
82V4
82'/,

2086,0
2089,2
2092,2
2095,4
2098,6
2101,8
2105,0

87V8
87V4
87V.
87V,
87V.
87V4
87V,

2213,0
2216,2
2219,2
2222,4
2225,6
2228,8
2232,0

92V.

92V.
92V,

92V.

92»/,

92V,

2340,0
2343,2
2346,2
2349,4
2352,6
2356,8
2359,0

97V.
97V4
97V.
97V,
97V.
97V4
97V,

2467,0
2470,2
2473,2
2476,4
2479,6
2462,8
2486,0

78

1981,2

R8

2108,2

88

2235,2

93

2362,2

96

2489.2

78V,
78'/4
78»/,

78V,
78»/,

78«/4
78V,

198^3
1987,6
1990,7
1998,9
1997,0

2000,2
2008,4

88«/,

88V4
88»/,
88'/.

83»/,

83»/4
88'/,

2111,4
2114,6
2117,6
2120,8
2124,0
2127,2
2130,4

88V.
88V4
88V.
88V.

88»/.
88»/,
88V,

2238,4
2241,6
2244,6
2247,8
2251,0
2264,2
2267,4

93V.
93V,

93V,

93»/,

93V4
93V,

2365,4
2866,6
2371,6
2374,6
2378,0
2381,2
2384,4

98V.

98V4
96»/,

96V,
98»/.
98»/,
98V,

2492,4
2495,6
2498,6
2601,8
2605,0
2508,2
2511,4

79

2006,6

84

2138,6

89

2260,6

94

2387,6

99

2614,6

79V,

79V4
79»/,

79V,
79»/,
79*/.
79*/,

2009,8
2013,0
2016,0
2019,2
2023,4
2086,6
2028,8

84'/,

84V4
84V,
84'/,
84»/,

84V4
84'/,

2186,6
2140,0
2148,0
2146,2
2149,4
2152,6
2156,8

89V.

89V4
89»/.

89V.
89*/.

89^/4
89V,

2263,8
2267,0
2270,0
2273,2
2276,4
2279,6
2282,8

94V.

941/4

94V.
94V,
94V.
94V4
94V,

2390,8
2394,0
2397,0
2400,2
2408,4
2406,6
2409,8

99V,

99V4
99»/.

99V,
99»/.

99»/4

99V,

.2517,8
2621,0
2524,0
2627,2
2630,4
2533,6
2636,8

80

3088,0

86

2159,0

90

2286,0

96

2418,0

100

2M0,0

AT.^.— Decimal poinU indicated by commas.

686

MARINE ENGINES AND BOILERS.

Table Na VIIX.
Square Metrea and Square Feet

Bqoftre
M«tr6t

■»sr

Bq«in
Itotrci

%r

BqiMn
Mrtrw

"3sr

Sqnwe
Itotnt

%r

1

10,764

26

279,872

51

548,979

76

818,087

2

21,529

27

290,636

52

559,744

77

828,851

8

82,298

28

301,400

53

570,508

78

839,616

4

43,057

29

312,165

54

581,272

79

860,880

5

58,822

80

322,929

55

592,036

80

861,144

6

64,586

81

333,693

56

602,801

81

871,908

7

75,850

82

344,458

57

613,565

82

882,673

8

86,114

83

365,222

58

624,329

83

893,437

9

96,879

84

365,986

59

635,094

84

904,801

10

107,643

85

376,750

60

645,858

86

914,966

11

118,407

86

387,515

61

656,622

86

925,730

12

129,172

37

398,279

62

667,387

87

936,494

18

139,936

38

409,043

63

678,151

88

947,268

14 150,700

39

419,808

64

688,915

89

968,023

15

161,464

40

430,572

65

699,680

90

968,787

16

172,229

41

441,336

66

710,444

91

979,561

17

182,993

42

452,101

67

721,208

92

990^6

18

193,757

43

462,865

68

731,972

93

1001,080

19

204,522

44

473,629

69

742,737

94

1011,841

20

215,286

45

484,394

70

753,501

96

1022,608

21

226,060

46

495,158

71

764,265

96

1033,873

22

236,815

47

505,922

72

776,030

97

1044,137

23

247,579

48

516,686

73

786,794

98

1064,901

24

258,343

49

527,451

74

796,558

99

1066,666

25

269,108

50

538,215

75

807,822

100

1076^480

N.B. — Decimal points indicated by commas.

VARIOUS TABLES.

687

Table No. IX.
Feet and Sqnare Metres.

Sqnftn

Bqoan
M«tnt

%r

Sqnftn
M«t(W

BOOATC

BqoAn
M0lrM

"^r

BqiMn
Metnt

1

0,0929

26

2,4154

51

4,7379

76

7,0604

2

0,1858

27

2,5088

52

4,8808

77

7,1588

3

0,2787

28

2,6012

58

4,9237

78

7,2462

4

0,3716

29

2,6941

54

5,0166

79

7,3891

5

0,4645

80

2,7870

55

5,1095

80

7,4820

6

0,5574

81

2,8799

56

5,2024

81

7,5249

7

0,6503

82

2,9728

57

5,2953

82

7,6178

8

0,7432

83

8,0657

58

5,8882

88

7,7107

9

0,8361

84

8,1586

59

5,4811

84

7,8086

10

0,9290

85

8,2515

60

5,5740

85

7,8965

11

1,0219

86

3,3444

61

5,6669

86

7,9894

12

1,1148

37

3,4378

62

5,7598

87

8,0828

18

1,2077

38

3,5802

63

5,8527

88

8,1752

14

1,8006

89

8,6281

64

5,9456

89

8,2681

15

1,3985

40

3,7160

65

6,0385

90

8,8610

16

1,4864

41

o,oUd9

66

6,1314

91

8,4589

17

1,5798

42

3,9018

67

6,2248

92

8,5468

18

1,6722

43

3,9947

68

6,8172

98

8,6897

19

1,7651

44

4.0876

69

6,4101

94

8,7826

•

20

1,8580

45

4,1805

70

6^5030

95

8,8255

21

1.9509

46

4,2784

71

6,5969

96

8,9184

22

2,0438

47

4,8668

72

6,6888

97

9,0118

28

2,1367

48

4,4592

73

6,7817

98

9^1042

24

2,22%

49

4,5521

74

6,8746

99

9,1971

25

2,3225

50

4,6450

75

6,%75

100

9,2900

JV.^.— Decimal points indicated by commas.

688

MARINE ENGINES AND BOILERS.

Table No. XI.
Pounds and Kilognunmes.

Founda

KOoffr.

Pouida

KOofr.

Pounds

KDofr.

Poondi

KOoffr.

1
2
8

4
5
6
7
8
9

0,4586
0,9072
1,3608
1,8144
2,2680
2,7216
3,1752
3,6287
4,0823

41
42
43
44
45
46
47
48
49

18,5973
19,0509
19,5045
19,9581
20,4117
20,8653
21,8189
21,7724
22,2260

81
82
83
84
85
86
87
88
89

36,7410
37,1946
37,6482
38,1018
38,5554
39,0089
39,4625
39,9161
40,3697

121
122
123
124
125
126
127
128
129

54,8847
55,8383
55,7919
56,2455
56,6991
57,1527
57,6063
58,0599
58^135

10
11
12
13
14
15
16
17
18
19

4,5359
4,9895
5,4431
5,8967
6,8503
6,8039
7,2575
7.7111
8,1647
8,6182

50
51
52
53
54
55
56
57
58
59

22,6796
23,1382
23,5868
24,0404
24,4940
24,9476
25,4012
25,8548
26,3084
26,7619

90
91
92
93
94
95
96
97
98
99

40,8233
41,2769
41,7305
42,1841
42,6877
43,0918
43,5449
48,9985
44,4521
44,9057

130
131
132
133
134
135
186
137
138
139

58,9671
59,4207
59,8742
60,3278
60,7814
61,2350
61,6885
62,1421
6^,5958
68,0494

20
21
22
23
24
25
26
27
28
29

9,0718
9,5254
9,9790
10,4826
10,8862
11,3896
11,7984
12,2470
12,7006
13,1542

60
61
62
68
64
65
66
67
68
69

27,2155
27,6691
28,1227
28,5763
29,0299
29,4835
29,9371
30,3907
30,8448
81,2979

100
101
102
103
104
105
106
107
108
109

45,3593
45,8128
46,2664
46,7200
47,1736
47,6272
48,0808
48,5344
48,9880
49,4416

140
141
142
143
144
145
146
147
148
149

63,5030
63,9566
64,4102
64,8638
65,3174
65,7710
66,2246
66,6782
67.1317
67,5853

80
«1
82
83
84
35
36
87
88
89

13,6078
14,0614
14,5149
14,9685
15,4221
15,8757
16,8292
16,7298
17,2865
17,6901

70
71
72
73
74
75
76
77
78
79

31,7515
32,8051
82,6587
83,1123
33,5658
84,0194
34,4780
34,9266
35,8802
35,8838

110
111
112
118
114
115
116
117
118
119

49,8952
50,3488
50,8024
51,2560
51,7096
52^682
52,5168
53,0704
53,5240
53,9775

150
151
152
158
154
155
156
157
158
159

68,0389
68,4925
68,9461
69,3997
69,8533
70,3069
70,7605
71,2141
71,6677
72,1212

40

18,1487

80

36,2874

120

54,4311

160

72,5748

//.B.^Decuoal points indicated by commas.

VARIOUS TABLES.

689

Table No. XII.
KUognumnes and Ponnds.

KUogr.

Pounds

Kilogr.

Pouida

KUogr

KUogr.

Poonda

1

2
8

4
:5
6
7
8
9

2,2046

4,4092

6,6139

8,8185

11,0231

13,2277

15,4324

17,6370

19,8416

41

42
43
44
45
46
47
48
49

90,3895

92,5941

94,7987

97,0034

99,2079

101,4126

103,6172

105,8218

108,0264

81
82
83
84
85
86
87
88
89

178,5743
180,7789
182,9836
185,1882
187,8928
189,5974
191,8020
194,0067
196,2113

121
122
123
124
125
126
127
128
129

266,7591
268,9638
271,1684
273,3780
275,5776
277,7823
279,9869
282,1915
284,3961

10
11
12
13
14
15
16
17
18
19

22,0462
24,2508
26,4554
28,6601
30,8647
33,0693
85,2739
87,4786
39,6882
41,8878

50
51
52
53
54
55
56
57
58
59

110,2311
112,4357
114,6403
116,8499
119,0495
121,2542
123,4588
125,6634
127,8680
130,6727

90
91
92

93
94
95
96
97
98
99

198,4159
200,6205
202,8251
205,0298
207,2344
209,4390
211,6431
213,8482
216,0529
218,2575

130
131
132
133
134
135
136
137
138
139

286,6004
288,8054
291,0100
293,2146
295,4192
297,6238
299,8285
302,0330
304,2337
306,4423

20
21
22
?3
24
25
26
27
28
29

44,0924
46,2970
48,5017
50,7068
52,9109
55,1155
57,3202
59,5248
61,7294
63,9340

60
61
62
63
64
65
66
67
68
69

132,2773
134,4819
136,6865
138,8911
141,0958
143,3004
145,5050
147,7096
149,9142
152,1189

100
101
102
103
104
105
106
107
108
109

220,4621
222,6667
224.8713
227,0760
229,2806
231,4852
233,6898
235,8945
288,0991
240,3087

140
141
142
143
144
145
146
147
148
149

308,6469
310,8516
313,0562
315,2608
317,4655
319,6700
321,8747
824,0793
326,2839
328,4885

SO
31
32
33
34
85
86
87
38
39

66,1386
68,3433
70,5479
72,7525
74,9571
77,1617
79,3664
81,5709
83,7756
85,9802

70

71
72
73
74
75
76
77
78
79

154,3235
156,5281
158,7327
160,9374
163,1419
165,3466
167,5512
169,7559
171,9605
174,1651

110
111
112
113
114
115
116
117
118
119

242,5083
244,7129
246,9175
249,1222
251,3268
253,5314
255,7360
257,9407
260,1453
262,3499

150
151
152
153
154
155
156
157
158
159

330,6932
332,8978
335,1024
337,3120
339,5116
341,7163
343,9209
346,1254
348,3301
351,1348

40

88,1848

80

176,3697

120

264,5545

160

352,7394

N.B, — Decimal points indicated by commas.

2x

690

MARINE ENGINES AND 1K)ILKRS.

Table No. XIII.
Pounds per Square Inch and Kilogrammes per Square Centimetre.

^

It

1^'

II

ii

mndeper
aare Inch

•

' .a

to O

it

ig-

t ■

£?

8.

fiy

8.

£7

&

£8*

&

1

0,0703

36

2,530

71

4,991

106^

7,452

141

9,913

2

0,1406

37

2,601

72

5,061

107

7,522

142

9,983

3

0,2109

38

2,671

73

5,131

108

7,593

143

10,054

4

0,2812

39

2,741

74

5,202

109

7,663

144

10,124

5

0,3515

40

2,812

75

5,272

110

7,733

145

10,194

6

0,4218

41

2,882

76

5,342

111

7,804

146

10,264
10,339

7

0,4921

42

2,952

77

5,413

112

7,874

147

8

0,5624

43

3,022

78

5,483

113

7,944

148

10,405

9

0,6327

44

3,093

79

5,553

114

8,015

149

10,475

10

0,7030

45

3,163

80

5,624

115

8,085

150

10,546

11

0,7733

46

3,233

81

5,694

116

8,155

155

10,897

12

0,8436

47

3,304

82

5.764

117

160

11,249

13

0,9140

48

3,374

83

5,834

118

8,296

165

11,600

14

0,9843

49

3,444

84

5,905

119

8,366

170

11,952

15

1,0546

50

3,515

85

5,975

120

8,436

175

12,303

16

1,1248

51

3,585

86

6,045

121

8,507

180

12,655

17

1,1952

52

3,655

87

6,116

122

8,577

185

13,006

18

1,2655

53

3,725

88

6,186

123

8,647

190

13,358

19

1,3351

54

3,796

89

6,256

124

8,718

195

13,710

20

1,406

55

3,866

90

6,327

125

8,788

200

14^061

21

1,476

56

3.936

91

6,397

126

8,858

210

H76

22

1,546

57

4,007

92

6,467

127

8,929

220

15,46

23

1,616

58

4,077

93

6,537

128

8,999

230

16,16

24

1,687

59

4,147

94

6,608

129

9,069

240

16,87

25

1,757

60

4,218

95

6,678

130

9,140

250

17,57

26

1,827

61

4,288

96

6,748

131

9,210

260

18,27

27

1,898

62

4,358

97

6,819

132

9,280

270

18,98

28

1,968

63

4,428

98

6,889

133

9,350

280

19,68

29

2,038

64

4,499

99

6,959

134 1 9,421

290

20,38

30

2,109

65

4,569

100

7,030

135

9,491

300

21,09

31

2,179

66

4,639

101

7,101

136

9,561

310

21,79

32

2,249

67

4,710

102

7,171

137

9,632

320 : 22.49

33

2,319

68

4,780

103

7,241

138

9,702

330

23,19

34

2,890

69

4,850

104

7,312

139

9,772

340

23,90

35

2,460

70

4,921

105

7,382

140

9,843

350

24,60

A^. .5. —Decimal points indicated by commas.

VARIOUS TABLES.

691

Table No. XIV.
Kilos^nunmes per Square Centimetre and Pounds per Square Inch.

•

is

11

is

II

Kilogr.
persq. em.

Is

II

Kilogr.
persq. om.

Pounds per
square inch

Poundl per
square inch

Kilogr.
per sq. om.

Pounds per
square inch

0^1

1^422

3,6

51,203

7,1

100,984

10,6

150,77

14,1

200,55

0,2

2,844

3,7

52,625

7,2

102,407

10,7

152,19

14,2

201,97

0,8

4,267

8,8

54,048

7,3

103,829

10,8

153,61

14,3

203,39

0.4

5,689

3,9

55,470

7,4

105,251

10,9

155,03

14,4

204,81

0,6

7,111

4,0

56,892

7,5

106,674

11,0

156,46

14,5

206,24

0,6

8^533

4,1

58,315

7,6

108,096

11,1

157,88

14,6

207,65

0,7

9,956

4,2

59,737

7,7

109,518

11,2

159,30

14,7

209,08

0,8

11,378

4,3

61,159

7,8

110,940

11,3

160,72

14,8

210,50

0,9

12,800

4,4

62,582

7.9

112,363

11,4

162,14

14,9

211,92

1.0

14,223

4,5

64,004

8,0

113,785

11,5

163,57

15,0

213,35

1.1

15,645

4,6

65,426

8,1

115,207

11,6

164,99

15,1

214,77

1,2

17,067

4,7

66,849

8.2

116,630

11.7

166,41

15,2

216,19

1.3

18,490

4,8

68,271

8,3

118,052

11,8

167,83

15,8

217,61

1,4

19,912

4.9

69,693 :

8,4

119,474

11,9

169,26

15,4

219,04

1.6

21,334

5,0

71,116

8,5

120,897

12,0

170,68

15,5

220,46

1.6

22,757

5,1

72,538

8,6

122,319

12,1

172,10

15,6

221,88

-1,7

24,179

5,2

73,960

8,7

123,741

12,2

173,52

15,7

223,30

1.8

25,601

5,3

75.382

8.8

125,164

12,3

174,94

15,8

224,73

1.9

27,024

5,4

76,805

8,9

126,586

12.4

176,37

15,9

226,15

2.t)

28,446

5,5

78,227

9,0

128,008

12,5

177,79

16,0

227,57

2.1

29,868

6,6

79.649

9.1

129,431

12,6

179,21

16,5

234,68

2^

31,291

5,7

81.072

9,2

180,853

12,7

180,63

17,0

241,79

2;3

32,713

5,8

82,494

9,3

132,275

12.8

182,06

17,5

248,91

2,4

34,135

5,9

83,916

9,4

133,698

12,9

183,48

18,0

256,02

2,6

35,558

6,0

85,339

9,5

135,120

13,0

184,90

18,5

263,13

2,6

36,980

6,1

86,761

9,6

136,542

13,1

186.32

19,0

270,24

2,7

38,402

6,2

88,183

9,7

137,965

13,2

187,75

19,5

277,35

•2,8

39,824

6,8

89,606

9,8

139,387

13,3

189,17

20,0

284,46

2,9

41,247

6.4

91,028

9,9

140,809

13,4

190,59

20,5

291,58

3,0

42,669

6,5

92,450

10,0 142,230

13,5 192,01

21,0

298,69

3,1

44,091

6,6

93,873

10,1

143,650

13.6 193,43

21,5

305,80

3.2

45,514

6,7

95,295

10,2

145,080

13,7 194,86

22,0

312,91

3,3

46,936

6,8

96,717

10,3

146,500

13,8 196.28

22,5

320.02

8.4

48,358

6,9

98,140

10,4

147,920

13,9 197,70

23,0

327,13

3,6

49,781

7,0

99,562

10,5

149,340

14,0

199,12

23,5

334,25

Al^. ^Decimal points indicated by commas.

692

MARINE ENGINES AND BOILERS.

Table No. XXI.
ComiMuisoii of ThennometerB.

Centigrade

lUanmur

Fahienheit

Centagnuie

B^aumw

Fchxenlwit

20

— 16

- 4

+ 20

+ 16.0

+ 68^0

— 19

— 15,2

- 2,2

h21

.

1-16,8

.

h 69,8

-18

-14,4

- 0.4

-22

-

-17,6

-

- 71,6

-17

— 13,6

h 1,4

-23

-

-18,4

-

- 73,4

— 1«5

— 12,8

-

^ 3,2

-24

-

-19,2

-

- 75,2

— 15

— 12,0

-

- 5,0

-25

-

-20,0

••

- 77,0

— 14

-11,2

-

- 6,8

-26

-

-20,8

-

- 78,8

— 13

-10,4

«

- 8,6

-27

-

-21,6

-

- 80,6

-12

- 9,6

-

-10,4

-28

-

-22.4

-

- 82.4

-11

- 8,8

-

-12,2

-29

-

-23,2

-

- 84,2

— 10

- 8,0

+ 14,0

+ 30

+ 24,0

+ 86.0

— 9

- 7,2

.

hl6,8

[-81

.

1-24,8

.

h 87,8

— 8

- 6.4

-

-17,6

-32

-

-26,6

-

- 89,6

— 7

— 5,6

-

-19,4

-33

-

-26,4

-

- 91.4

— 6

- 4,8

-

-21,2

-34

-

-27,2

-

- 93.2

— 5

- 4,0

-

-23,0

-36

-

h28,0

-

- 95,0

— 4

- 3,2

-

-24,8

-86

-

-28,8

-

- 96.8

^ 3

- 2,4

-

-26,6

-37

-

-29,6

-

- 98,6

— 2

- 1,6

-

-28,4

-38

-

-30,4

-

-100,4

— 1

- 0,8

-

-80,2

-39

-

-31,2

-

h 102,2

0

0

+ 82,0

+ 40

+ 82.0

+ 104,0

.

h 1

h 0,8

—

h33.8

h41

m.

[-32,8

.

1-105^

-

- 2

- 1,6

-

-35,6

-42

-

-33,6

-

-107,6

-

- 8

- 2,4

-

-37,4

-43

-

-34,4

-

-109.4

-

- 4

. 8,2

-

-39,2

-44

-

-36,2

-

-111,2

-

- 5

- 4,0

-

-41,0

-45

-

-36,0

-

- 118,0

-

- 6

- 4,8

-

-42,8

-46

^

-86,8

-

-114.8

-

- 7

- 5,6

-

-44,6

-47

-

-37,6

-

- 116,6

-

- 8

- 6,4

-

-46,4

-48

-

-38,4

-

- 118.4

-

- 9

h 7,2

-

-48,2

-49

-

-39.2

-

-120.2

+ 10

+ 8,0

4-60,0

+ 50

+ 40,0

+ 122.0

-

hll

h 8,8

_

h51,8

1-51

+ 40.8

.

[-128,8

-

-12

- 9,6

-

-63,6

-62

-41,6

-

-125,6

-

-13

-10,4

-

-55,4

-53

-

-42,4

-

-127,4

-

-14

-11,2

-

' 57,2

-64

-

-43,2

-

h 129,2

-

-16

-12,0

-

-69.0

-55

-

-44,0

..

- 131,0

-

-16

-12,8

-

-60.8

-56

-

-44,8

-

-132,8

-

-17

-18,6

-

-62,6

-57

-

-45,6

-134,6

-

-18

-14,4

-

-64,4

-58

-

-46.4 1 -

-136,4

■■

-19

-15,2

—

-66,2

-59

-47.2

-

-138,2

N.B, — Decimal points indicated by comma&i

VARIOUS TABLES.

693

Table No. 'XXI.— continued.
Comparison of Thermometers.

Centigrade

Rdaumur

Fahrenheit

Centigrade

Reaumur

Fahrenheit

+ G0

+ 48,0

+ 140,0

+ 100

+ 80,0

+ 212,0

h61

^

[-48.8

h 141,8

1-101

_

f- 80,8

[-213.8

-62

—

-49,G

- 143,6

-102

-

- 81,6

-216,6

-63

-

-50,4

-145,4

-103

-

- 82,4

-217,4

-64

-

-61,2

- 147,2

-104

-

- 83,2

- 219,2

-65

-

-52,0

- 149,0

-105

-

- 84,0

-221,0

-66

-

-62,8

-150,8

-106

-

- 84,8

-222,8

-67

-

-63,6

- 162,6

-107

-

- 86.6

-224,6

-68

-

-54,4

-164,4

-108

-

- 86,4

-226,4

-69

-

-65,2

- 156,2

-109

-

- 87,2

h 228,2

+ 70

+ 56,0

+ 158,0

+ 110

+ 88,0

+ 280,0

h71

_

h56,8

h 159,8

-HI

—

h 88,8

\- 231,8

-72

-

-57,6

- 161,6

-112

-

- 89,6

-283,6

-73

-

-58,4

-163,4

-113

-

- 90,4

-236,4

-74

-

-59,2

-166,2

-114

-

- 91.2

-287,2

-76

-

-60,0

- 167,0

hll6

-

- 92,0

-239,0

-76

-

-60,8

-168,8

-116

-

- 92,8

-240,8

-77

-

L61,6

- 170,6

-117

-

- 98,6

- 242,6

-78

-

h62,4

- 172,4

-118

-

- 94,4

-244,4

-79

-63,2

- 174,2

-119

-

- 95,2

- 246,2

+ 80

+ 64,0

+ 176,0

+ 120

+ 96.0

+ 248,0

h81

_

h64,8

h 177,8

(-121

h 96,8

[-249,8

h-82

-

-66,6

- 179,6

-123

-

- 97,6

-261,6

h83

-

-66,4

- 181,4

-123

-

- 98,4

-263,4

-84

-

-67,2

-183,2

-124

-

- 99,2

-266,2

-85

-

-68,0

-186/)

-125

-

-100,0

-267,0

-86

-

-68,8

- 186,8

-126

-

-100,8

-268,8

-87

-

-69,6

-188,6

-127

-

- 101,6

-260.6

-88

-

-70,4

-190,4

-128

-

-102,4

-262,4

-89

-

-71,2

- 192,2

-129

-

-103,2

-264,2

+ 90

+ 72,0

+ 194,0

+ 180

+ 104.0

+ 266,0

1-91

«

[-72,8

h 195,8

[-131

n

1-104,8

[-267,8

-92

-

-73,6

- 197,6

^m

-132

-

-105,6

- 269,6

-93

-

-74,4

- 199,4

-133

-

-106.4

-271,4

-94

-

-75,2

-201,2

-134

-

- 107,2

- 273,2

-95

-

-76,0

-203,0

-135

-

-108,0

-275,0

-96

-

-76,8

-204,8

-97

-

-77,6

-206,6

-98

-

-78,4

-

-208,4

-99

-79,2

- 210,2

N.B. — Decimal points indicated by commas.

694

MARINE ENGINKS AND IJOILERS.

g

SI

X

X

B

2L

6

•^ <f>

U3

l-H O

B

O a*

pa o

^ 3

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VARIOUS TABLES.

695

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d98

MARINE ENGINES AND BOILERS.

Table No. XXIII.
Expansion of Rigid Bodies by Heat,

Material.

1

Expansion

between

32* and 212* F.

per Unit of

Length.

0-0029
0-0018
0-0008
0-0011
0-0016

Material.

Expansion

between

32* and 212* F.

1 per Unit of

Length.

Lead -
Bronze
Oak -
Cast iron -
Copper

Wrought fron
Steel -

Pine -
Zinc -

0-0012
0-0011 to
00012
0-0035
0003

Table No. XXIV.
Melting Points of various Materials,

Material.

Antimony

Lead

Bronze, average

Gold

Cast iron, white

Do. grey

Copper - - .-

Platinum

Quicksilver

Wrought iron

Sulphur

Silver

Steel

White metal

Bismuth

Bismuth, 8 parts ; tin, 3 parts ; lead, 5 parts
Zinc . - - - - -
Tin

Degrees Fahr.

800

630
1650
2190
1930
2190
2010
4530
-40
2730 to 3230

230
1830
2430
410 to 500

500

210

800

450

VARIOUS TABLES.

699

Table No. XXVa.
Specific Gravity of Woods,

Material.

Oak -

Ash -
Pine -
Fir -
Cork -
Pitch pine
Guaiacum
Teak -

Specific Gravity.

Green.

0-89 to 1-06
0-78 „ 0-93
0-79 „ 0-99
0-81 „ 1-00

}}

Seasoned.

0-65 to 0-92
0-54 „ 0-85
0-38 „ 0-48
0-46 „ 0-76

0-24

0-66
1-26 to 1*34
0-88 „ 0-98

))

Table No. XXVb.
Specific Gravity of Metals,

MateriaL

Specific

Material.

Specific

Aluminium -

Gravity.

1

2-6 to 2-7 1

Gravity.

Iron —

Antimony

6-7 '

Pig, white steel -

7-6

Lead

11-3 to 11-4

Siemens - Martin

Bronze —

1

mild steel

7-85

Deha metal

8-6 '

Siemens - Martin

Bell metal -

8-8 1

steel

7-85

Manganese

1

Tool steel -

7-86

bronze -

8-5

Copper

^'% to 9

Brass

8-4 to 8-7

German silver

8-4 „ 8-7

Muntz metal

8-5 j

Platinum

21-5

Phosphor bronze

8-8

Quicksilver -

13-6

Gun metal -

8-7

Silver -

10-5

Iron —

White metal -

7-1

Pure -

7-8

Tin - - -

7-3

Cast -

7-25

Zinc

7-1

Pig, grey -

6-8 to 7-5

700

MARINE ENGINES AND BOILERS.

Table No. XXVc.

Specific Gravity of Various Materials,

Material.

Specific
Gravity.

Material.

Specific
Gravity.

1

Asbestos

1-2

Chalk -

1-8 to 2-6

Asphalt

-

M to 1-5

Marble -

2-7

Cement, powder -

115 „ 1-7

Cotton wool, fairly

1 }

„ set -

-

2-7 „ 3

loose

016

Hard fat

-

0 92 „ 0-94

Sulphur

2-0 ,

Rubber, raw -

-

0-92 „ 0-96

Talc or Mica

0-91 :

„ vulcanised

1-45

Bricks, ordinary -

1-94

„ fire -

1-91 '

1

Table No. XXVd.

Relative

Weights of Coals /«

fr Unit Volume (Water = I).

Relative

Relative

Material.

Weight

Material.

Weight

( Water =1).

Coke, screened

(Water=l).

1

Anthracite -

.

1-4 to 1-7

0-4

German brown coal

0-8 „ 1-5

Hard coal

1-2 to 1-5

Coke -

«•

1-4

„ screened

0-75

Table No. XXVe.
Specific Gravity of Fluids,

Material.

Specific
Gravity.

0-79

Material.

Specific
Gravity.

Alcohol, pure

Hydrochloric acid

Turpentine -

0-85

(concentrated) -

M9

Linseed oil -

-

0-94

Sulphuric acid (con-

Mineral oil -

-

0-90 to 0-92

centrated) -

1-84

Rape oil

-

0-91

Sea-water

1025

Petroleum

-

0-80 to 0 90

Distilled water at

Quicksilver -

-

13-6

4* Centigrade or

Nitric acid ((

con-

39" Fahr. -

1-00

centrated) -

1

1-53

VARIOUS TABLES.

701

Table No. XXVf.

Specific Weights of Gases at 30 inches of Mercury and 32' Fahr.

{Water =- 1),

Gas

Specific
Weight.

Gas.

Specific
Weight.

Carbonic oxide
Carbonic acid gas -
Coal gas

Air - - -
Methane, marsh gas

0-001250
0001978
0-000690
0001293
0-000720

Steam (see Table

XXII.).
Oxygen
Nitrogen
Hydrogen

0-001429
0-001256
0-0000896

If G is the specific weight of a gas at 32" Fahr. and 30 inches, and
Gj the specific weight of the same gas at /** Fahr. and / inches of
mercury —

G, =G
^ 30

\l+a/j

in which the coefiicient a is nearly the same for all gases and = -002036
per degree Fahr.

735 mm. of mercury 1 _ ^ -^.^^-.t^Ki^r^ _ / ^ ^^g- P^"" square cm.
28-6 inches of mercur>'/ " ^ ^tmospnere- | ^^,^^ j^ ^^ ^^^^^^ .^^^^

702

MARINE ENGINES AND BOILERS.

Table No. XXVL
Strength and Elasticity of Various Materials.

«

> Material.

Breaking

Strength.

Tons per

square inch.*

Elongation
per cent.

on an 8 inch
.Length.

Elastic

Limit.

Tons per

square inch.

1

Modulus

of

Elasticity.

Tons per

square

inch. 1

1
1

, Common grey cast
iron
Common grey cast
iron under com-
pression

7-6 to 9
44 „ 47

9,530

Extra good cast iron,
for cylinders and
the like

Extra good cast iron,
for cylinders and
the like, under
compression -

10-5 „ 13-75
47 to 51

1

6,650

Good welding iron,
small forgings

22 „ 24

U „ 18

12 to 16

12^700

Siemens-Martin mild
forged steel, accord-
ing to percentage of
carbon present

24 „ 27

20 „ 25

12 „ 19

13,600

Siemens-Mart i n
forged steel, for
shafts, &c.

29 „ 35

20 „ 25

12 „ 22

14,000

Forged crucible steel,
best quality -

29 „ 35

20 „ 25

12 ^'2

14,000

Nickel steel, forged -

35 „ 41

20

24

14,000

Tool steel, unhardened

48 „ 57

35 and over

14,000

Steel castings (Sie-
mens-Martin)

25 „ 32

18 to 20

12 to 19

13,600

Boiler plates of mild
steel

24 „ 27

20 „ 25

16

13,600

Best crucible steel wire
for ropes

95 „ 115

^^__

Rolled brass

9-5

7,000

* For strength in compression or shear, at the most four-fifths of the strength in
tension may be assumed.

VARIOUS TABLES.

703

Table No. XXVL — continued.

Material.

*

Breaking

Strength.

Tons per

square inch.*

12 to 19

Elongation
per cent.

on an S.iuch
Length.

Elastic'

Limit.

.Tons per

square incl).

Modulus

of

Elasticity.

Tons per

square

inch.

Best gunmetal, bronze
for valves, &c.

at least
10 to 20

5,700

Muntz metal, rolled

1

or forged

22

1

»

Delta metal, forged -

22 to 24

i

11-5

6,350

Manganese bronze,!
cast (propeller
blades) -

19 „ 29

15 to 25

— >^

Manganese bronze,
drawn (bolts)

25 „ 32

20 „ 40

Copper plates -

13 „ 15

38

9

7,000

Copper wire

Up to 24

Oak (with the grain) J

7

tension

4-2

compression

7

1
1

760

Pine (with the grain)./

f
tension
2-8
compression

7-6

]
)

}-

760

Ash (with the grain)-/

tension
4-2

630

I

compression

1

Note, — Unless otherwise mentioned, the table refers to the breaking
strength in tension.

The figures for elongation give the percentage increase of the
original length which occurs in a test bar 8 inches long when broken.

* For strength in compression or shear, at the most four-fifths of the strength iq
tension may be assumed.

t Compare Table No. XXVII,

704 MARINE ENGINES AND BOILERS.

Table Na XXVII.

Strength and Elasticity of Manganese Bronze*
{Parsons' Manganese Bronze.)

Quality No. 1. — ^Tbis cannot be cast, but is only obtainable
hammered, rolled, or drawn.

Uses, — Screws, spindles, nuts, piston rods, circulating pump
shafts, &c.

Strength. — 22 to 39 tons per square inch.

Elastic Limit. — 11 to 16 tons per square inch (and up to 34 tons
per square inch).

Extension. — 20 to 45 per cent.

Quality No. 2. — For castings.

Uses. — Propeller blades, sundry engine parts in torpedo-boats, stem
posts for small vessels, &c.

Strength. — 23*5 to 37 tons per square inch.

Elastic Limit. — 15 to 20*5 tons per square inch.

Extension. — 18 to 29 per cent.

Note. — This extension is that obtained on a test piece 2 inches
long. The highest values given for strength and extension are those
of a single test.

* From data given by the firm.

VARIOUS TAHLKS.

705

Table No. XXX.

Moments of Inertia " i " and Internal Moments of Resistance or Moduli
of Section " w ''for Circular Sections of Diameter "^."

d

d^TZ

64

10

490-9

10-5

596-4

11

718-7

11-5

858-5

12

1,018

12-5

1,179

13

1,402

13-5

1,630

14

1,886

14-5

2,178

15

2,485

15-5

2,833

16

3,217

16-5

3,638

17

1

4,100

17-5

1

4,604

' 18

5,153

1 18-5

5,750

1 19

6,397

19-5

7,098

; 20

1

7,854

1

21

9,547

22

11,499

23

13,737

24

16,286

25

19,175

32

d

'=64
22,432

d^T

W = -- .

32

9817

26

1,726

113-6 '

130-7

149-3

27
28
29

26,087
30,172
34,719

1,932
2,165
2,394

169-6

30

39,761

2,651

188-6

215-7

31

45,333

2,925

241-5

32

51,472

3,217

269-4

33

58,214

3,528

! 299-3

34

65,597

3,859

331-3

1

35
36

73,662
82,448

4,209
4,580

; 365-6

37

91,998

4,973

402-1

38

102,354

5,387

441-0

39

113,561

5,824

482-3

1
1

526-2 !

40

125,664

6,-283

572-6
621-6
673-4
728-0

41
42
43
44

138,709
152,745
167,820
183,984

6,766
7,274
7,806
8,363

785-4

45

201,289

8,946

909-2
1,045
1,194
1,357

46
47
48
49

219,787
239,531
260,576
282,979

9,556
10,193
10,857
11,550

1,534

50

306,796

12,272

1

2 Y

700

MAKINK I:N(;INKS AM) lUMLKRS.

Table No. XXX. — continued.

d

d^TT

^=64

d^T
w- - -

32

1

d

64

d^
32

51

332,086

13,023

76

1,637,662

43,096

52

358,908

13,804

77

1,725,571

44,820 I

1

53

387,323

14,616

7X

1,816,972

46,589

54

417,393

15,459 ,

79

1,911,967

48,404

55
56

449,180
482,750

16,334
17,241

80

2,010,619

50,2(55

57

518,166

18,181

81

2,113,051

52,174

58

555,497

19,155

82

2,219,347

54,130

59

594,^10

20,163

83

2,329,605

56,135

60

636,172

21,206

84

85

2,443,920
2,562,392

58,189
60,292

61
62
63
64

679,651
725,332
773,272
823,550

22,284
23,398
24,548
25,736

86

87
"^"^
89

2,685,120
2,812,205
2,943,748
3,079,853

62,445
64,648
66,903
69,210

65

876,240

26,961

90

3,220,623

71,569

66
67
68
69

931,420

989,166

1,049,556

1,112,660

28,225
29,527
30,869
32,251

91
92
93
94

•

3,366,165
3,516,586
3,671,992
3,832,492

73,982
76,448
78,968
81,542

70

1,178,588

33,674

95

3,998,198

84,173

71
72

1,247,393
1,319,167

35,138 '
36,644

96

97
98

4,169,220
4,345,671
4,527,664

86,859
89,601
92,401

73

1,393,995

38,192

99

4,715,315

95,259

74

1,471,963

39,783

75

1,553,156

41,417

1

100

4,908,738

98,175

VARIOUS TABLES.

0'

/u/

Table No. XXXI.
Bending Moments.

Method of Support.
Curve of Moments.

Prfssure on Supports a and B.

Bending Moment m and Maxitnum

Bending Moment Mi,wx.

A= P
M = P . X

Mmax = P . /

Cantilever loaded at
Weakest sectionr at a.

end.

A =

p

M = ..V

^1 tlt3 V

P./

'max

Beam supported at ends.
Weakest section in middle.

A-p/|- B-P.

M = P . j . AT

"'max .

Beam unevenly loaded.
Weakest section at c.

MARINE ENGINES AND BOILERS.

Table No. XXXI.— continued.

Method of Support.

Pressures on Supports a snd n.

Bending Moment M and Maximum

Bendiiig Moment M,c„

A

)

>■

<s

\

P„(/,+/.) + P,/,

A

C--^

Sih

^Tl = A/, M,, = B/j

'

.

/

— T

M = A.J-P,{*-/,).

= B(/-.x)-P.,(/-/3-A-).

1

—

(L

5

(d)

Beam with two loads and
unfixed ends. Weakest section
at p,.

A = 2/ = P=/./

B

p = toad per unit length.

.......

. B.

Cantilever. Evenly distri-
buted load. Weakest section
at A.

A = P, + P = P,+//

M = OT, + OTj=P,.ir + -^"

)

Mm. = M] + M,. = P,/ +-\,-

Cantilever. Combined load,
evenly distributed and at end.
\Veakest section at a.

VARIOUS TABLES.
Table No. XXXI.—ceniinueii.

Method of Support.
Curve of Monienls.

Cun'e oj Moments :
Paraliola with Apex at C.

I'lesiures on Supports a and B
Bending Moment M, and Mtxinii

Bendb(; Momtnt Mnn..

p . I- vl

\Veakest section at c.
Beam with evenly distri-
buted load and unfixed (;nd!>.

-^(f-D

Weakest sections at a, d,
and c. Built-in beam cen-
trally loaded.

■mi^ffi^mmi':

=^G-MO

Weakest sections at a and b.
Built-in beam with evenly
distributed load.

710 MARINE EXGIXES AXD BOILERS.

Table No. XXXI. — continued.

Bending Moments.

IvCt M = bending moment in inch-pounds at any section.

I = moment of inertia of the section taken round that neutral
axis which, lying in the plane of the section, passes
through its centre of gravity, and is normal to the
direction of the bending force.

^ = the distance from the neutral axis of the most highly
stressed fibre.

w = - = intenul moment of resistance (or modulus) of the sec-
e

tion in inches^.
Then the permissible bending or skin stress: —

s = — lb. per square inch,
w

For s there should be inserted the permissible values of stress when
under tension.

VARIOUS TABLES.

711

Table No. XXXIL
Torsional Strength,

M = twisting moment in inch-pounds.

G = torsional modulus of elasticity = about 5,000 tons per square inch
for iron and steel.

/= length in inches of specimen twisted.

s = torsional stress about four-fifths of permissible tensile stress.

1. Circular Section.

IT «
M = -d3.S

2. Ring Section.

M = - . s

16 D

3. Rectangular Section.

M = '^b'^h, s.

•7

{h>b,)

4. If </> is arc through which shaft is twisted, then for round or
circular section —

. 32 M ,

r = — s ' - ^

</» is the length of the arc, measured on a radius of 1 inch, through
which two sections at a distance / from each other are twisted.

712

MARINE ENdlNKS AM) BOILERS.

Table No. XXXIII.
Strength of Struts.

/= length of strut in inches.

I = smallest moment of inertia in inchest

£ == modulus of elasticity pounds per square inch.

p = load in pounds at which bending first occurs.

^-diameter of strut in inches.

Case 1 . One end of strut built in, the other end free.

p =

ir2 EI

4' /-^

Fig. I.

Case 2. Both ends free, but constrained to move in the direction of
load.

7r2 EI

Fig. II.

Case 3. One end built in, the other free but constrained to move in
direction of load.

'It- EI

Case 4. Both ends built in.

o EI

P = 4^^-^

Fig. III.

Fig. IV.

w^m^m

The above formulas apply for struts with the proportion —

Case 1. /= hd for cast iron ; /= 12</ for wrought iron.
„ 2. /-lO// „ J /=24^

„ 3. l^Wd „ ; /=33^

„ 4. /=20^ „ : 7=48^

>j

))

>j

VARIOUS TAIJLES.

h-1 •>

Table No. XXXIV.
German Lloyd^s Rules for Iron and Steel for Boilers,

1. Wrought-iron Plates,

Shell Plates.

Tensile
Strength.

tons pel I kg. per
SK). in. ,sq. mm.

With the grain - 21 | 33

Across the grain

19

30

Per cent.
Elongation
in 8 inches

End Plaiesi.

Tensile
Strength.

tons per, kg. per
sq. in. sq. mm.

•22 35

21

33

Per cent.
Elongation
in 8 inches.

12

8

Furnace Platen

Tensile
Strength.

Per cent.
; Elong.ition
in 8 inches

tons peri kg. per

sq. if. 'sq. mm.

23 36
21-5 34

18
12

For plates more than 1 inch thick the strength may be reduced by
0-32 ton per square inch (0-5 kilogramme per square millimetre) for
every extra 0*08 inch (2 millimetres) of thickness.

2. Steel and Mild Steel Plates,

I Tons per sq. in.
Tensile strength \

\ Kg. per sq. mm.

Per cent, elongation in 8 inches

I Tons per sq. in.
Ten.sile strength -|

\ Kg. jier sq. mm.

Per cent, elongation in 8 inches

30-5 I 30

! 48 I 47
; 20 -20

26
41
23

•29

28-5

28

27-5

27

46

45

44

43

42

20-5

21

21 -5

22

22-5

25-5 25

40 39

23-5 24

'24

23-5

23

22

:»

37

:)6

35

24-5

25

25-5

26

For shell plates 20 per cent, elongation is sufficient. Furnace plates
and end plates should have a tensile strength not exceeding 28 tons per
square inch, with a corresponding elongation.

14

MARINE EN(;iNES AND BOILERS.

Table No. XXXIV. — co7itinued.

3. Stays and Rkfets of Iron or Steel.

1

Wrought Iron.

Steel.

Tensile Strength.

I*er cent.
Elonga-
tion in
8 inches.

Tensile Strength.

Per cent.
Elongation

in 8 inches.

tons per

kg. per

tons per

kg. per

sq. in.

sq. mm.

. sq. inch.

sq. mm.

Stays

at least

at least

at least

22 to 29

35 to 45

—

Rivets

23

36

15

22 „ 29

35 „ 45

at least

at least

at least

Angles

23

36

16

^^^^

Table No. XXXV.
German Lloyd^s Rules for Steel and Cast Steel for Parts of Engines,

Crank Shafts, Propeller Shafts, and Lengths of Shafting

in Wrought Steel.

Strength, 25 to 30 tons per square inch ; elongation, 20 per cent,
on 8 inches.

Special and nickel steel to have at least 20 per cent, elongation.

Stem and Stern Posts, &c., in Cast Steel.

•Strength, 25 to 35 tons per square inch j elongation, 15 per cent, on
an 8 inch length.

Note. — For the cast steel portions of engines and boilers 25 to 28
tons is now usually allowed.

VARIOUS TABLES.

715

Table No. XL.
Weight of Machinery (from various published data).

Under heading " Engines" are included main engines with propeller
and shafting.

Under heading " Boilers " are included boilers with small and large
mountings, fire box, and funnel (boiler empty).

Under heading " Pipework " are included all pumps, water, steam,
and exhaust pipes, gratings, platforms, &c.

Type of Engine.

Torpeclo-l)oals and destroyers.
Triple-expansion engines* —
Waler-tul^ boilers

Light cruisers. Triple-expan-
sion engines * —
With cylindrical boilers
With water-tube lx)ilcrs

Battleships. Triple - expansion
engines * —
Cylindrical boilers

Fast liners. Cylindrical
boilers —
Triple-expansion engines
(Quadruple-expansion engines

Cargo boats. Cylindrical

boilers —

Large, boats with quadruple-
expansion engines

Medium -sized, with triple-
expansion engines

Small, with compound engines

Steam pinnaces. Twin or com-
pound engines —
Locomotive boilers

Weight in pounds per h. i>,

Engines.

Boilers.

18 to 31

4,1 „ 67
45 ,, 67

67 „ 83

113 „ 135
126 „ 157

22to:W

Pipework.

170 „ 214

158 „ 200
145 „ 180

16 „ 30

78 „ 100
45 ,, 67

90 „ 113

124 „ 146
135 „ 170

170 „ 214

170 „ 214
170 „ 190

45 „ 67

22 „ 34

34 „ 45

54 „ 67

54 ,, 67

Total.

9 to 18 I 50 to 83

100 „ 113

90 „ 100
78 „94

145 ,,200
113 „ 124

190 ,,204

290 „ 350
314 „ m)

440 „ ^30

416 „ 510
395 „ 465

9 „18

70 „ 115

• The weights have been calculated on the maximum n.P. with forced draught.

16

MARINE ENGINES AND BOILERS.

Table No. XLI.
Weight of Boiler Equipments Compiled from the German Navy*

(Weights of boilers without water, but with large and small mountings,
fire box, and funnel, as well as all pipework and auxiliary machinery
in the boiler-room.)

Type of Boiler.

Weight per h.p. in pounds (reckonetl on
the maximum continuous h.p.).

Cylindrical boiler

143 to 188, mean 163

Locomotive boiler -

71 „ 157,

»>

108

Diirr boiler - - - -

89 „ 138,

»»

130

Thornycroft boiler -

64 „ 92,

»

82

! Belleville boiler

1

100 „ 108,

»

105

This table does not include weights of boilers for torpedo boats.

Table No. XLIL
Weight of Cylindrical Boilers {unclothed and 7vithout mountings^ <S^'^.).

Cylindrical single-ended boilers, weight per square foot of heating
surface —

For 114 to 142 lb. per square inch, 31 2 to 39-5 lb.
For 170 to 214 lb. per square inch, 35-4 to 49 lb.

Double-ended boilers of the same diameter, and of double the
heating surface of the corresponding single-ended boilers, weigh about
12 to 17 '/„ per square foot of heating surface less than the latter.

* From Kohn v. Jaski, Manm- Rundschau^ 1901, v.

APPENDIX.

FINAL REPORT (JUNE 1904) TO THE
LORDS COMMISSIONERS OF THE
ADMIRALTY OF THE COMMITTEE
ON NAVAL BOILERS.

THE COMMITTEE ON NAVAL BOILERS.

n.M.S. "Bulwark" at Rapallo,
VUhJwic 1904.

Sir, — I have the honour to submit, herewith, to be laid before
the Lords Commissioners of the Admiralty, the final Report of the
Boiler Committee of which I am the President. Although I have not
been present at the experiments carried out during the last two years, I
have received from time to time all the reports, and they show the great
care and pains taken by the Committee to obtain correct results.

2. With reference to our previous Report, I am compelled to say
that my experience with the Belleville boilers on the Mediterranean
Station has been very favourable to them as a steam generator, and it
is clear to mc that the earlier boilers of this description were badly
constructed and badly used. We have had no serious boiler defects in
any of the ships out here, and the fact that two ships are about to be
recommissioned with only the ordinary annual repairs being undertaken
shows that their life is not so short as I originally supposed. However,
the second commission of these ships will be a very good test of the
capabilities of their boilers.

3. In conclusion, I cannot express too highly my opinion of the
work done by my colleagues on the Committee.

I have the honour to be,

Sir,
Your obedient Servant,
(Signed) COMPTON DOMVILE,

Admiral and Commander-in-Chiefs

President of the Boiler Committee.

The Secretary to the Admiralty.

REPORT OF THE COMMITTEE ON

NAVAL BOILERS.

June 1904.

1. The Committee on Naval Boilers appointed by the Lords Com-
missioners of the Admiralty in September 1900, having completed their
investigations and experimental trials, and being in a position to recom-
mend standard types of boiler for use in H.M. Navy, as requested in
their Lordships' letter of 28th February 1901, have the honour to
submit their final Report.

2. A Statement of the work of the Committee up to May 1902 was
given in paragraph 2 of their Report of that date. Since then the
reboilering of H.M.S. " Medea " with Yarrow large-tube boilers, and of
H.M.S. " Medusa " with Diirr boilers, together with the necessary
machinery alterations, have been completed under the supervision of
the Committee, and the boilers of both ships have been thoroughly
tested. The results obtained are recorded in a separate Report. As
requested by their Lordships in their letter S 1 JJSt ^^ ^^^ November
1902, the Committee have also carried out a series of trials of the
Babcock & Wilcox boilers of H.M.S. " Hermes," which extended from
7th October 1903 to 16th May 1904. These trials also form the
subject of a separate Report.

3. The Committee have from time to time reported the results of
their investigations, and they have also answered such questions as
have been put to them by their Lordships. The Reports and other
documents which have already been forwarded include : —

(tf.) The Interim Report forwarded on the 19th February 1901.

ip,) Minutes of the evidence given before the Committee, together
with the Appendix thereto, forwarded 26th April 1901.

(f.) Report on the trials of the "Hyacinth," "Minerva," and
"Saxonia," together with a Summary of Conclusions, forwarded 27th
November 1901.

2z

722 APPENDIX.

(d,) Progress Report for the year 1901, forwarded 31st Decentot,

1901. • iv
(c.) Report on the relative economy and efficiency of Belle j n

and cylindrical boilers in commissioned ships, forwarded 29th AAuy>^

1902. - '
(/) Report of May 1902, together with the Appendix thereto.
(g.) Report on the trials of the " Seagull," " Sheldrake," " Espil

and "Fantdme," forwarded 5th August 1902.

There are now submitted with this Report : —

(A.) Report on the trials of the " Medea" and " Medusa."
(i.) Report on the trials of the " Hermes."

4. The Report of May 1902 was intended to be final as r^ards the]
Belleville boiler, and the Committee have since seen no reason to
modify the opinion expressed in paragraph 6 of that Report, viz., that it
is "undesirable to fit any more of this type in H.M. Navy."

5. In paragraph 5 of their Report of May 1902, the Committee
stated that the experience obtained by them since the date of their
Report of February 1901, had confirmed them in the opinion that the
"advantages of water-tube boilers for Naval purposes are so great,
chiefly from a military point of view, that, provided a satisfactory type
of water-tube boiler be adopted, it would be more suitable for use in
H.M. Navy than the cylindrical type of boiler."

In their Reports of 1901 and 1902, the Committee expressed the
opinion that four different types of water-tube boiler, viz. : —

(a.) Babcock & Wilcox,

(d,) Niclausse,

(c) Diirr, and

(d.) Yarrow large-tube,

were sufficiently promising to justify their use in H.M. Navy in com-
bination with cylindrical boilers. Having concluded their experimenul
investigations, they are now satisfied that two of these four types, viz,, the
Babcock & Wilcox, similar to that tried in the " Hermes," and the
Yarrow large-tube, similiar to that tried in the " Medea," are satisfactory
and are suitable for use in battleships and cruisers without cylindrical
boilers. In the Babcock & Wilcox boiler, the generating tubes are
nearly horizontal ; in the Yarrow boiler, they are nearly vertical. Each
type has its particular advantages, and only long experience on general
service can show which is, on the whole, the better boiler. For the

APPENDIX. 723

^^st ftatnt, the Committee unanimously recommend both types as suitable

ival requirements.
' of Bt- n making these recommendations, the Committee recognise that
id :"^: upkeep of any water-tube boiler is likely to be heavier than that of
cylindrical boiler, but they are of opinion that the two types they
thcu ff recommend will cost less for upkeep than the other types of large
^ "feaight-tube boiler which they have had under trial.

6. The Committee make these recommendations after investigations
id trials carried out under their superintendence extending over a
I- eriod of nearly four years.

The ships in which each type of boiler has been tried by the Com-
mittee are —

v.-

Cylindrical - - H.M.S. " Minerva " and R.M.S. " Saxonia."

Belleville - - H.M.S. " Diadem " and H.M.S. " Hyacinth."

Babcock & Wilcox H.M.S. "Sheldrake," H.M.S. "Espi^gle,"

and H.M.S. " Hermes."

Niclausse - - H.M.S. " Seagull " and H.M.S. " Fantome."

Durr - - - H.M.S. "Medusa."

Yarrow large-tube H.M.S. " Medea."

7. Although the Committee have no knowledge of any type of
water-tube boiler which is likely to prove more suitable for His Majesty's
ships than the two recommended, there are other types which may be
considered worthy of trial later on. If any type of boiler is considered,
in future, to be of sufficient merit to justify its trial in the Navy, it is
recommended that it be fitted in a new vessel not smaller than a
second-class Cruiser.

8. As in their previous reports, the Committee do not offer any
remarks on the most suitable type of boiler for small vessels of high
speed. From the nature of the case, some form of " express " boiler
with small tubes closely pitched is absolutely necessary in order to
obtain such a ratio of output to weight of boiler as is required in
torpedo-boats and destroyers. For small Cruisers, however, which have
to keep the sea and act with the fleet, it is probable that a boiler such as
the Yarrow large-tube would, on the whole, give better results than the
" express " types which have hitherto been fitted.

9. In reference to paragraph 3 (if) of their lordships' letter of 6th
September 1900, and to the Committee's Report of May 1902, they
desire to call attention to the breakdown of the " Hyacinth's " machinery
on 16th February 1903, and to the trouble experienced with the

724 APPENDIX.

bearings in the " Hermes " during the homeward run from Gibraltar,
which strengthen the recommendation of the Committee contained in
paragraph 13 (a) of the Report of May 1902, viz., "They consider it
desirable, where practicable, to increase the length of stroke and reduce
the number of revolutions per minute as compared with the recent
practice in His Majesty's Service."

10. The principal comparative results on which the recommenda-
tions of the Committee are based are set forth in the succeeding
paragraphs. F'ull details are given in the separate Reports of Trials.

11. Thermal Efficiency of Boilers,— The full Tables which are
appended to the Committee's Reports give the efficiency of each type
of boiler under very varied conditions. The results are here sum-
marised : —

The best obtained with the Babcock & Wilcox boilers of the
'' Hermes " were during the trials of furnace gas baffling, the boilers in
the middle boiler-room, with vertical baffles and a forced air supply
over the fires, giving the high efficiency of 81 per cent, on a 30 hours-
trial when 20 lbs. of coal were being burnt per square foot of firegrate
per hour, and an efficiency of 77*8 per cent, on a 29 hours' trial when
burning 27 lbs. per square foot, these rates of combustion corresponding
to the ordinary rate of steaming and to the full power of the boilers
respectively. The boilers of the " Hermes," with the restricted uptake
baffling and without any special air supply over the fires, had a maxi-
mum efficiency of 75 8 per cent, on a 12 hours' trial when burning 20*5
lbs. per square foot per hour. On three trials of over 24 hours' duration
each and when 19 lbs. were being burnt per square foot per hour, the
efficiency was in each case practically 71 per cent. When burning 29
lbs. per square foot per hour for 7 hours, the efficiency of these boilers
was 66-3 per cent. ; but the weather during this test was so bad that
the trial, which was to have been of 8 hours' duration, had to be stopped
after the seventh hour on this account. During the baffling trials,
however, in good weather, an efficiency of 70*3 was obtained on the 30
hours' trial, when burning 27 lbs. per square foot per hour, or practically
the full output.

The maximum efficiency of the Yarrow boilers of the " Medea,"
viz., 75-7 per cent., was obtained on a 26 hours' trial when burning
1 8 lbs. per square foot per hour ; their efficiency, when burning at the
maximum rate of combustion, viz., 40 lbs. per square foot per hour for
8 hours, was 69*5 per cent. On trials of over 24 hours' duration each,
burning from 17 to 21 lbs. per square foot per hour, the efficiency
remained at or over 75 per cent.

APPENDIX. 725

The Belleville boilers of the " Hyacinth " had a maximum efficiency
of 77-2 per cent, recorded on a 24^ hours' trial, when 16 lbs. of coal
were being burnt per square foot of firegrate per hour. When burning
20 lbs. per square foot per hour for 1 1 hours, the efficiency was 73*3 per
cent.; and burning 17*4 lbs. for 24 hours, it was 71*8 percent. The
efficiency of these boilers on an 8 hours' trial in fine weather, when
burning 27 lbs. per square foot per hour, corresponding to the full
output of the boiler, was 65 per cent.

The maximum efficiency of the Durr boilers of the " Medusa " was
64*8 per cent, obtained on an 8 hours' trial, when burning 35 lbs. per
square foot per hour, this being the maximum rate of combustion with
these boilers; the efficiency, when burning 16 lbs. per square foot per
hour for 26 hours, was 63-8 per cent. On trials of over 24 hours'
duration each, and burning 18 lbs. and 21 lbs. per square foot, the
efficiencies were 61*7 per cent, and 60-3 per cent, respectively.

Of the cylindrical boilers tried, those of the " Saxonia," on the only
trial made, which was of 13 hours' duration, and on which 20 lbs. per
square foot per hour was burnt, had the high efficiency of 82-3 per cent.

The maximum efficiency obtained with the cylindrical boilers of the
" Minerva " was 69*7 per cent., which was recorded on a 25 hours' trial
when burning 14 lbs. per square foot per hour ; on a trial of 8^ hours'
duration, with retarders in the plain tubes and burning 29 lbs. per square
foot the efficiency, was 68*4 per cent.

In the smaller ships, the maximum efficiency of the Babcock &
Wilcox boilers tried was 66 per cent, on a 12 hours' trial burning 18 lbs..
per square foot per hour in the " Sheldrake," and 73*2 per cent, on a
9 hours' trial burning 13 lbs. per square foot in the "Espifegle." The
maximum efficiency obtained by the Niclausse boilers of the " Seagull "
was 66-9 per cent, on an 8 hours' trial burning 13 lbs. per square foot,
and by those of the " Fantome," 69-8 per cent, on a 9 hours' trial
burning 14 lbs. per square foot.

12. A noticeable feature in connection with the boiler efficiencies is
the improvement in the results obtained with the later boilers of the
Babcock & Wilcox type. The earliest of these fitted in the
"Sheldrake" in 1898, showed efficiencies ranging from 66-0 percent,
to 59*2 per cent. ; the boilers fitted in the " Espiegle " in 1901 showed
improved efficiencies ranging from 73-2 per cent, to 63*1 per cent.
Those of the "Hermes," fitted in 1903, show a still further gain in
economy, the efficiencies ranging from 75-8 per cent, to 66-3 per cent.,
and the same boilers, after modification, showed, on one occasion, the
high efficiency of 81 per cent. It is noticed in this connection that the
three sets of Babcock & Wilcox boilers tried differ from each other in

726 AppEKnrjf.

the arrangement of their heating surface and furnace gas baffling. The
boilers of the "Sheldrake" were fitted throughout with tubes IJJ
inches in diameter, without any baffles for furnace gases ; the boilers of
the " Espi^gle " were fitted throughout w^ith tubes 3^^ inches in
diameter, vertical baffles being placed among the tubes and causing a
zigzag flow of the gases ; the boilers of the ** Hermes " were fitted with
two rows of 3}- J inches diameter tubes immediately over the fire, the
remainder of the tubes being l|j{ inches diameter, and the baffling of
the furnace gases was effected by a restriction of the space for the
passage of the gases between the top row of tubes. Those boilers of
the ** Hermes " which showed the efficiency of 81 per cent, were
similar in construction to those last mentioned, but the baffling of the
furnace gases was by a vertical system which caused a zigzag flow of
the gases over the heating surface, and, in addition, a forced air supply
was introduced above the fires (see " Report on the Trials of H.M.S.
• Hermes.* ")•

The arrangement of the heating surface in both the earlier and
later boilers of the Niclausse type was the same, and the thermal
efficiencies of the two sets of boilers were very similar.

13. Wetness of Steam. — As explained in other Reports, the wetness
of the steam was taken throughout the Committee's trials by means of
a Carpenter's calorimeter. Experience in the " Medusa " satisfied
the Committee that the results registered by this instrument arc
trustworthy.

As regards the production of dry steam at all rates of combustion
the Yarrow large-tube and the later Babcock &: Wilcox boilers have
given the best results.

14. Loss of Water. — The loss of feed water with each of the four
types of boiler under consideration has been moderate throughout the
Committee's trials. In the runs to Gibraltar and back, carried out with
the " Medea " and " Medusa," the loss of water was small, being at
the rate of 1*6 and 1*8 tons per 1000 horse-power per day respectively.
On the 140 hours' endurance trial of the ** Hermes," the loss was 3'^
tons per 1000 horse-power per day.

The loss of water may be expected to be greater in boilers fitted
with many doors than in those fitted with but few, and to increase as
the doors and joints become worn. In this respect, the Yarrow boiler,
having only three manhole doors, has an advantage.

15. Examination and Cleaning of Interiors of Tubes. — Of the
boilers tried by the Committee, the Yarrow boiler can be internally

APPENDIX. 727

examined and cleaned in the shortest time and with the least amount
of labour — to obtain access for such an examination and cleaning it is
only necessary to remove manhole doors. The Babcock & Wilcox
type is less easily examined and cleaned — two small doors have to be
removed for each tube, and these have to be rejointed after the
examination and cleaning have been completed. In order to carry
out a thorough examination of the tubes of the Diirr boiler, it is neces-
sary to remove a hand-hole door at the front of each tube, the
diaphragm washer of the internal tube, the internal tube itself, and
the cap nut at the back end of the generating tube ; but, in order to
carry out a thorough cleaning, it is also necessary to remove the gene-
rator tubes from the boiler, and after the cleaning is complete, these
have to be replaced, this being a long and tedious process. The work
connected with the examination and cleaning of the tubes of Niclausse
boilers is very similar to that necessary with the Diirr boiler. Further,
the cap nut at the back end of the Diirr boiler tube permits of each
tube being readily emptied, while owing to the back end of Niclausse
boilers being inaccessible, some process is necessary to empty the tubes
when required, such as blowing the water out of the tube by a special
pump and hose.

The necessity for being able to withdraw each of the tubes in a
direct line with its axis, renders the clear space required for the
installation of Diirr and Niclausse boilers considerably more than
would be required for boilers of other types. For warships, where the
stokehold space is very limited, this must necessarily cause considerable
inconvenience in the arrangement of pipes and auxiliary machinery.

16. External Cleaning of Tubes, — In both the "Medea "and the
" Hermes " it is possible to partially clean the tubes externally, when
the fires are alight, by means of air lances.

The tubes in the " Medea " can be thoroughly cleaned externally
when the fires are not alight, as they can be swept in three directions,
viz., from the furnace, from the smoke-box, and from the front of the
boiler. The tubes of the Diirr and Niclausse boilers cannot be so
thoroughly cleaned externally in place as those of the ** Medea," the
number of rows being greater and the overlapping of the baffles pre-
venting portions of certain tubes being touched. In the Babcock &
Wilcox boilers, the tubes can be swept horizontally through side doors
fitted to the casings, but, as the boilers in the " Hermes " were
originally fitted, the sweeping in a vertical direction was difficult.
After the alterations of baffling, the sweeping vertically can be carried
out, but necessitates the removal of portions of the baffles. It is to be
recognised that any system of baffling among the tubes, however it may

728 APPENDIX.

improve the circulation of the gases, renders the cleaning of the tubes
themselves more difficult.

17. Bending of Tubes, — After the "Medusa" had completed her
preliminary runs, it was found that all the tubes of the bottom rows
had curved upwards in the middle, the maximum bending being 1^,^
inches, and these tubes were removed and straightened before starting
on the Committee's trials. These tubes had to be straightened again
in August 1903, and again at the conclusion of the Committee's trials
in February 1904. When the Committee visited H.M.S. "Berwick"
in April 1904, it was noticed that the tubes of the bottom rows of the
boilers (Niclausse type) were bent upwards, and the members were
informed that the maximum bending on the 22nd March 1904 was
^ inch ; the ship was new in 1903, and only commissioned in December
of that year. With the Niclausse, and also with the Diirr boiler, con-
siderable bending of the tubes at the bottom rows must be expected,
and it will be necessary to straighten these tubes when the amount of
bending exceeds J inch. This will entail a considerable amount of
extra work with these types of boiler, which will be off service for
corresponding periods. The upward bend of the generator tube is
often greater than the space between the inner and outer tubes, and as
the inner tube, which is only supported at the two ends, remains
straight, it is liable to touch the outer tube at some point, thus impeding
the circulation of water between them. To prevent this, it may be
necessary to support the inner tube at the middle of its length as well
as at the back end, so that it must bend with the outer tube.

In the case of the Yarrow boilers of the " Medea," the Committee
experimented in six of the boilers with the firerows of tubes pur-
posely bent, as described on page 9 of the " Report on the Trials of the
* Medea ' and * Medusa,'" with the object of overcoming some slight
leakages of tube ends which showed themselves when working under
forced draught. In two boilers, the tubes of the firerows were left
straight. Although these bent slightly in use, no trouble was experi-
enced with them ; and, during the later trials, these boilers proved to
be as satisfactory as regards freedom from leakage as those in which
the firerows had been put in bent. The Committee have suggested in
their letter of the 21st December 1903, concerning the Yarrow boilers
proposed for H.M.S. "Warrior," that the tubes of the firerows be bent
one inch from the straight, and this recommendation they think should
apply to future designs.

In the Babcock & Wilcox boilers of the " Hermes," although some of
the tubes of the bottom rows have bent, no leakage of tube ends has

APPENDIX. 729

resulted, and it has not been necessary to remove any tubes for
straightening or renewal.

18. Corrosion of Tubes and Wear of Casings and Uptakes, — In none
of the four types of water- tube boiler which were recommended for
trial by the Committee has there been any considerable corrosive decay
of tubes, and the ordinary wear has been very slight. On the conclusion
of the Committee's trials, the tubes of the boilers of the " Medea "
and " Hermes " had not deteriorated to any appreciable extent. This
applies also to the " Medusa," except that the internal tubes have shown
signs of roughening.

In the " Medusa " (Diirr boilers), there was some buckling of the
side casings of the boilers, and some of the casing doors at the back of
the boilers became warped <yid burnt.

No trouble was experienced in connection with the casings and
uptakes of the Yarrow boilers of the " Medea," and very little with
those of the Babcock & Wilcox boilers of the " Hermes." From the
experience of the Committee with the boilers of the S.S. "Martello,"
employed on the Atlantic trade for nearly four years, and also from
their experience to date with the " Hermes," it is considered that the
durability of the casings and uptakes of Babcock & Wilcox boilers will
prove to be satisfactory under the ordinary conditions of Naval Service.

In the Yarrow boiler, the temperature of the furnace gases is con-
siderably reduced before they reach any part of the side casings, and,
in consequence of this moderate temperature, the casings and uptakes
of the "Medea's" boilers were uninjured on the conclusion of the
Committee's trials.

In this respect, the Yarrow boiler is superior to the other types of
water-tube boiler which have been tried by the Committee.

19. Liability to Damage from being Forced, — The makers of the Dun-
boilers stated that not more than 35 lbs. of coal should be burnt per
square foot of firegrate per hour in the " Medusa." The Committee
consider that this limitation of the quantity of coal to be burnt was
prudent, as the overheating and bending of tubes in one of the boilers
during the full-power homeward run from Gibraltar were, in the opinion
of the Committee, due to the fact that the safe limit had been exceeded.
It is also considered that the limitation of the amount of coal to be
burnt per square foot of grate applies with even greater force to the
Niclausse boiler, as the supply of water to the tubes is freer in the case
of the Diirr boiler than in that of the Niclausse. As the result of their
trials, the Committee find that the Yarrow boiler can be severely forced

730 APPENDIX.

without danger, and that the Babcock & Wilcox boiler can with safety
be forced to the extent shown in the Reports.

20. Skilled Firing Required, — The satisfactory stoking of water-tube
boilers requires a higher degree of skill than that of cylindrical boilers,
and this is more necessary with the large grates of the Diirr, Niclausse,
and Babcock & Wilcox boilers than with the smaller grates and better
shape of combustion chamber of the Yarrow. The stoking in the
" Medea," " Medusa," and " Hermes " was good throughout the
the trials, and towards the end was excellent. Under ordinar)' service
conditions, such good firing can hardly be expected, at least until a
vessel has been some time in commission. Good results can, however,
be obtained with Yarrow boilers with engine room complements new to
the ship, as shown by the trials to Malta and back, which have been
made by the ** Medea " since the completion of the Committee's trials
with that vessel.

21. Superheated Steam, — The Diirr boiler was the only one tried by
the Committee which had any arrangements for superheating. It was
fitted with complicated directing plates in the steam collector and with
superheater tubes. The fittings in the steam collector are undesirable,
and they and the superheater tubes will probably require frequent
renewal, while the amount of superheat obtained by their use was small,
even when the temperature of the funnel gases was abnormally great
The results obtained were not sufficient to enable the Committee to
express any opinion as to the value of superheating as applied to Naval
boilers.

22. Feeding of the Boilers, — No trouble has been experienced with
the feeding of any of the four types of boilers under consideration. In
the " Medea " and " Medusa," the boilers were fitted with automatic
feed regulators. It was found, however that these were not sufficiently
sensitive in opening and closing (allowing a variation of level of about
6 inches in the gauge glass) ; the feed was therefore regulated throughout
the trials by hand, no trouble being experienced in doing this. For a
similar reason, the feed was regulated by hand in the " Hermes " during
the trials. In the " Medea," the feed regulators were inside the steam
collectors and interfered with the examination and cleaning of the
middle rows of tubes. The Committee consider that the balance of
advantages rests with the omission of automatic feed regulators in
boilers such as the Yarrow large-tube and the Babcock & Wilcox,
where there is a fairly large reserve of water in the boiler.

APPENDIX.

731

23. Saii IVa/er.^The Report on the trials of the " Medea " and
" Medusa " contains a description of experiments made on the Yarrow
and Diirr boilers in regard to their behaviour when working with
brackish water. These experiments, so far as they went, indicated
that neither type of boiler was likely to give trouble from this cause.
In the case of the Yarrow boiler, this result has been corroborated by
the fact that on a recent voyage the " Medea " is reported to have bad
leaky condenser tubes and a corresponding density in the boilers without
any bad effect.

24. Relative Weights, — In the case of the " Hermes," the " Medea,"
and the " Medusa," the new boilers were installed without any altera-
tions being made in the stokehold floor spaces. A comparison of weights
and maximum output of the boilers gives the following results : —

Output of

Weight of

(1 •!

Maximum
Output of

S|eam per
hour per ton

TjTJc of Boiler.

1

Ship.

Boiler*rooin
Installation.

Steam per
hour.

of Boiler-
room

1

Tons.

Weights.

lbs.

lbs.

\

Cylindrical

'*Saxonia"

1,000 (abt.)

132,600

132-6

Ditto

"Minerva"

567

167,100

295

With retarders.

, Ditto

Ditto

5.18

156,20<l

280

As originally fitted.

: Belleville -

** Hyacinth "

454

178,700

394

Yarrow

** Medea"

330

157,800

478

\

DUrr -

** Medusa"

314

158,000

503

■

Babcock &

i Wilcox -

1

"Hermes"

400

2rjo,000

410

With vertical baffles,
and forced air
supply above the
fires.

. Ditto

Ditto

481

! 182,300

380

As originally fitted.

Ditto

•'Sheldrake"

12.5

43,840

351

Ditto

**Espi^le"

95

24,780

261

•

Niclausse -

"Seagull"
"Fantome"

135

48,450

.359

1 Ditto

76-5

, 22,750

1

297

1

1

25. The Committee are under great obligation to Mr C. J. Wilson,
F.C.S., who has, during the four years of their work, given his valuable
personal attention to the analysis of funnel gases and of coal samples
without any remuneration. They are also much indebted to Messrs
Thomas Wilson, Sons, t^ Company, for permission to examine the
boilers of the S.S. " Martello," and to Mr W. S. Hide, the Superintending
Engineer of that Company, for affording the Committee facilities for
carrying out the inspections and giving information concerning the
results obtained in the running of that vessel.

732 APPENDIX.

26. The Conimhtee desire in conclusion, to place on record their
appreciation of the assistance which they have received from their
Secretaries. Captain Browning, R.N., acted as Joint-Secretary until
his appointment to H. M.S. "Ariadne" in 1902. Engineer-Lieutenant
W. H. Wood, R.N., has continued to act as Secretary throughout their
whole work. The diligence and energy which the latter officer has
shown in carrying out his work, his knowledge of the scientific as well
as of the practical side of marine engineering, and his capacity for dealing
both with details and with general organisation, have been invaluable
to the Committee throughout, and especially in connection with the
carrying out of their boiler trials at sea, a work of no little difficulty and
complexity, and they desire to bring his services to the favourable
notice of their Lordships.

(Signed) COMPTON DOM VILE,

Admiral and Chairman.

JAS. BAIN.

JOHN INGLIS.

ALEX. B. W. KENNEDY.

JOHN LIST.

J. T. MILTON.

JOS. A. SMITH.

Wm. H. Wood,

St'creiary,

INDEX.

ACCELERATION, angiilar, 60
— crank-pin, 60

— of the piston, 60

— of the rotating masses, 60

— radial, 62

Actual work exerted by the engine, 6
Adiabatic expansion, 34
Admission, work done during, 17
^\ir pressure in stokehold, 665

— pump, 284

— pump, principal dimensions of, 284

— pump body, 289

— pump bucket, 292

— pump, Edwards', 293

— pump rod, 292

— pump suction and delivery pipes, 286

— pump valves, 286
All-round reversing gear, 261
Amount of eccentricity, 166

— of feed- water required, 310
.\nalysis of coal, chemical, 589

— of flue gases, 588
Angle cock, 417

— of crank, 58

— of lead, 168

— valve, 412
Angular acceleration, 60

— velocity of crank, 58
Apparent slip, 368
Approximate calculations, 19, 360
Area, developed, of a ship's screw, 363

— of blades, 363

— projected, of a ship's screw, 363
Areas of circles, 634
Arrangement of cranks, 106

— of cylinders, 106

— of main engines, 123
Arrangements for applying bjakc, 342

(see Slmft Brake)

Ash ejector, 623

— hoist, 623
Aspinall governor, 141
Atmospheric line, 5
Auxiliary engine exhaust, 435

— pumps, 321

— steam piping, 433

BALANCE cylinders, 223
— cylinder pistons, 223

— weights, 83

Balancing the moving parts of engines, 82

— the moving parts of engines, example

on, 92

— Schlick system of, 85
Ballast piping, 448

— pump, 326
Balusters, 614
Bearing caps, 243
Bearings, cooling water to, 450

— lubrication of, 621

— tunnel shaft, 341
Bed-plates, engine, 234
Beldam valves, 286
Belleville boiler, 513
Bending, resistance to, 707
Bilge piping, 447

— pumps, 321
Blades, area of, 363

— bolted on, 403

— breadth of, 363

— form of, 367

— number of, 367

— strength of, 384
Blake pump, 318
Blocks, thrust, 336
Blow-oflF, boiler, valve, 576, 581
Board of Trade, rules for shafts, 212

734

INDEX.

Boiler, Belleville, 513

— blow-off valve, 576

— Durr, 520 '

— emptying plug, 577

— fittings, 569

— fittings, marine, regulations affecting,

577

— flues, sectional area of, 459, 480

— lagging, 554

— Normand, 531

— rivets, 486

— safety valves, 569, 577, 579, 580, 581

— seating, 618

— stop valve, 573

— Thomycroft, 540

— tul>es, 480

— Yarrow, 526

Boilers, construction of, rules for, 488

— cylindrical, 474

— locomotive, 510

— power of, 457

— test pressures, 579, 580, 581

— water tube, 612

— weight of, 716

Bolted on propeller blades, 403
Bolt heads, 605
Bolts, 605

— holding down, 239

— main bearing, 243

— nuts for, 608

— tables of dimensions of, 607

— with fine thread, 605

— with Whitworth thread, 606
Boss, propeller, 401

— propeller, strengthening the, 401
Bracing, diagonal, 250

Brake, horse -power, 3

— shaft, 342

Brass, specific gravity of, 699

— strength of, 702
Brasses, connecting-rod, 196

— main bearing, 241
Breadth of blades, 363
Breaking strength of iron, 702

— strength of various materials, 702
British thermal units, 32

Bronze, manganese, strength of, 703, 704

— specific gravity of, 699

— strength of, 703
Brown's reversing engine, 256
Buckley piston packing, 186
Built-up crank shafts, 214

Bulkhead fittings, 41 1
— stuffing boxes, 341
Bureau Veritas, 498, 581
Byepass valves, 441

c

ALCULATION from the theoretical
diagram, 25

— of a ship's screw, Taylor, 376

— of covers, 134

— of cj'linder dimensions, 16

— of flat surfaces, 134

— of the resistance of ships, Middendoif,

355

— of triple engine, example of, 20

— of valve chest passages, 134
Calorific values of coals, table of, 455
Calorimeter, throttling, Peabody's, 590
Caps, bearing, 243

— horseshoe, thrust, 337
Cast iron, properties of, 702

— steel, 702, 714
Caxntation, 374
Celsius thermometer, 692
Centrifugal force, 62, 390

— pump casing, 303

— pump spindle, 303

— pumps, 298

— pumps, valves of, 298
Check valve, feed, 575 '
Chemical analysis of coals, 454, 589
Circular sections, moments of inertia of, 705
Circulating pipes, 450

— pumps, 294

— pumps, centrifugal, 298

— pumps, reciprocating, 294

— water, amount of, 294
Circumferences, table of, 634
Circumferential speed of crank, 58
Clearance, 14

— top and bottom, 189

Closed stokehold system of forced draught,

564
Coal, consumption of, 452, 469, 476, 519,

525, 530, 534, 543
Coals, calorific values of, 455

— chemical analysis of, 454, 589

— hard, 455

— specific gravities of, 700
Cock, salinometer, 576

— scum, 577

— through way, 417

INDEX.

735

Cocks, cylinder drain, 139, 143

— density, 576

— pet, 576

Coefficient of friction, 221
Coefficients for ship's screw, 474
Coefficients of expansion, 698

— of performance, 357

— tables of, 357, 374
Collars of thrust shaft, 336
Columns, 246

— cylinder, longitudinal bracing of, 122,

247
Combined diagram, 18
Combustion, 453
Compound engines, 9

— engines, expansion in, 9
Compression, 169

— effects of, 169

— resistance to, 702'
Condenser, 39

— cooling surface of, 273

— fittings and connections, 280

— jet, 280

— shell, 277

— surface, 272

— tube surface of, 273

— lubes and lube plates, 276

— vacuum in, 193
Condensers, 272

Condensing engines, efficiency of, 20
Connecting-rod, 193

— brasses, 196

— calculation of top end, 195

— fork, 195

— influence of length, 103

— shaft, 194
Connections, screw, 409
Constants, table of, 384, 679
Construction of boilers, rules for, 488, 578
Consumption of steam, 310

Cooling surface of condenser, 273

— water to bearings, 450
Copper, melting point of, 698

— specific gravity of, 699
Cobines, table of, 677
Cotangents, table of, 679
Couplings, crank -shaft, 219

— friction, 344

— muff, 344

— propeller shaft, 345

— shaft, 343

Cover, cylinder, studs for, 133

Covers of valves, 134
Crank, 58

— angle of, 58

— circle, angular velocity at, 58

— pin. 214

— pin acceleration, 60

— pin, tangential pressure on, 63

— shaft, 208

— shaft couplings, 219

— shafts, built up, 214

— shafts, forged, 216

— webs, 216

Critical number of revolutions, 78, 104
Crosshead, 199

— and slide, 199

— pin, 196

Cross sectional area of funnel, 459

— section of receiver pipe, 128, 130
Crucible steel, strength of, 702
Cube roots, table of, 634

Cubes, table of, 634
Cur\'es, speed, 353
Cut-off, 21

Cylinder columns, longitudinal bracing of,
122, 247

— cover, studs for, 133

— covers, 152

— drain cocks, 139, 143

— drainage, 143

— fittings, 139

— flanges, 133

— hydraulic tests, 138

— lagging, 139

— liner, 130

— passages, 134

— proportions, 21

— proportions, tables of, 43

— ratios, 23

— relief valves, 144

— stuffing boxes, 156

— lest pressures, 138

— walls, strength of, 130
Cylinders, arrangement of, 106

— draining of, 143

— warming up the, 37, 143
Cylindrical boilers, 474

— lx>ilers, furnaces and grates of, 476

— boilers, grate sur&ce of, 474

— boilers, heating surface of, 474

— boilers, rules for the construction of, 488

— boilers, tubes of, 480

— boilers, weight of, 7}6

1

736

INDEX.

DEGREE of regularity in revolutions
of an engine, 599
Delta metal, specific gravity of, 099

— metal, strength of, 702
Density cocks or valves, 576
Determination of cylinder dimensions, 3
Developed area of a ship's screw, 363
Diagonal bracing, 250

Diagram, MUlIer, 169

— Rankine's, 18

— tangential pressure, 63

— theoretical, 17

— valve, 169

Diameter of steam piping, 424
Diameter ratio, extreme, of screw, 371
Different forms of blade, 367
Dimensions of main bearings, 244

— of wheels for turning gear, 269
Direct loss of work due to clearance, 36
Discharge valves, 421

Double beat valve, 141

— ported slide valve, 166
Drain cocks, cylinder, 139, 143
Drainage of cylinders, 143
Draining of steam pipes, 423
Draught, Ellis and Eaves* induced, 555

— gauge, 589

— Howden's forced, 559

— natural, 458

Dryness fraction of steam, 590
Duplex feed pump, 315
DUrr boiler, 520

ECCENTRICITY, amount of, 166
Eccentric rods, 228

— rods, crossed, 173

— rods, open, 173

— sheaves, 231

— strap, 232
Eccentrics, 229
Edging plates, 613
Edwards' air pump, 293

Effect of length of connecting rod on

balance, 103
Effects of compression, 169
Efficiency of condensing engine, 20

— from steam diagram, 17

— of engine, 4, 17

— of propeller, 370

— total, 18
Ejector, ash, 622

Elasticity, limit of, 702

— modulus of, 702

Ellis and Eaves' induced draught, 555
Empirical formulae for calculation of a

ship's screw, 374
Engine, bed-plate, 234

— columns, 246

— efficiency' of, 4, 17

— foundations, 616

— position in ship, 104

— quadruple expansion, 9

— reversing. Brown's, 256

— reversing the, 256

— seating, 616

— starting the, 125

— stop valve, 141

— stroke of, 40

Engines, arrangement of main, 123

— Imlancing of (see Balancing the Moving

Parts), 82

— compound, 23

— efficiency of, 4, 17

— fan, 463

— for cruisers, 23, 45

— for driving centrifugal pumps, 31)6

— for large merchant ships, 53, 246

— for small merchant ships, 246

— for torpedo-boats, 45

— marine, revolutions per minute of, 41

— multiple expansion, 9

— power of, 3

— single expansion, 9

— twin cylinder, 9

— weight of, 715

Equivalent, mechanical, of heat, 33, 694

E vaporisation, heat of, 694

Example of calculation of triple engine, "IHK

24
Exhaust lead, 168

— piping, 434
Expansion, adiabatic, 34

— coefficients of, 698

— due to heat, 425

— in compound engines, 9

— of steam, 10

— total, 10

— work done during, 1 1

FAHRENHEIT thermometer, 585, 692
Fan engines, 463
Fans, 460

INDEX.

737

Feed check valve, 57«5

— heater, injection, 446

— pump barrels and valve lx)xes, 311

— pump valves, 314

— pumps, 310

— pumps, size of, 31 1
Feed-water filter, 439

— heater, 445

— pipes, 436

— required, amount of, 310
Filter, feed- water, 439
Fine threads, liolts with, 60o
Fire liars, 478

Firing and generation of steam, 4o3
Fittings, boiler, 569

— bulkhead, 411

— cylinder, 139

— under-water, 418

Fitting the slipper block to crosshcad,

418
Fixing the blades of a propeller, 403

— the funnel, 550
Flanges, 409

— cylinder, 133
Flanges, table of, 409
Float tank, 441

Flues, boiler, sectional area of, 459, 480
Forced draught, closed stokehold system,
5&4

— draught, Howden's, 559

— feed lubricator, 620
Forged crank shafts, 216
Form of blades, 367
Fottinger's torsion indicator, 599
Friction, coefficient of, 221

— couplings, 344
-^ of valves, 221
Frictional losses, 3

— resistance of ships, 354

Froude, calculation of ship's resistance,

353
Fuels, 455
F'unnel, 549

— cross-sectional area of, 459
Furnaces, 467

GASES, specific gravity of, 701
Gauge, draught, 589

— vacuum, 585

— water, 575
Gear, lifting, 614

Gears, reversing, 251
Generation of steam, 468
German Lloyd's, rules [for forgings and
castings for engines, 714

— Lloyd's, rules for iron and steel boilers,

488, 579

— Lloyd's, rules for pumps, 329

— Lloyd's, rules for shafting, 210

— Lloyd's, rules for spare gear, 627

— Lloyd's, rules for valves, &c., 411
Governor, Aspinall, 141

Grate surface, 457

— surface of cylindrical boilers, 466, 476

— surface of locomotive and water-tulxi

boilers, 510, 513
Grates, 477
Gratings, 612
Guide, slipper, 2(N)
Guides, 205

Guides, pressure on, 205
**^," values of, 681

HAMBURG rules, 493
Hard coals, calorific values of,
455
Heater, feed, injection, 444

— feed-water, 445
Heating surface, 468

— surface, cylindrical boilers, 474

— surface, locomotive boilers, 511

— surface, water-tube boilers, 512

— the receivers, 39

— value of coal, determination of, 589

— values of hard coals, 455
Heat losses in the engine, 37

— mechanical equivalent of, 32

— mechanical theory of, 32

— of evaporisation, tabic, 694
Heusinger valve gear, 181
Hoist, ash, 623
Holding-down bolts, 239
Hollow shafts, 209
Horse-power, brake, 3

— effective, 3

— indicated, 3
Horseshoe thrust caps, 337
Howden's forced draught, 559
Hull, vibration of the, 104
Hyperbola, rectangular, construction of,

11
Hyperbolic logarithms, 634

3 a

'88

INDEX.

IMMERSED midship section, 355
Inches to millimetres, 682
Indiarubber valves, 286
Indicated horse-power, 3

— horse-power, measurement of, 4
Indicator connections, 593

— diagram, 9

— diagram, construction of, 27

— gear, 595

Indicators and their use, 592

Indirect loss of work, 37

Induced draught, Ellis and Eaves', 555

Influence of multiple expansion, 37

Injection, feed-heater, 444

Instrument for measuring the uniformity

of turning moment, 599
Instruments (see Measuring Instruments),

585
Intermediate shafts, 340
Iron, breaking strength of, 7<)2

— melting point of, 698

— specific gravity of, 699

JACKETS, steam, 37
Jet condenser, 280
Joy's assistant cylinder, 224
— valve gear, 179
Junk rings, 190

K

ATZENSTEIN packing, 158

Kilc^rammes per square centimetre
to pounds per square inch, 691
— to pounds, 689
Kinghorn valves, 286
Kingston valves, 418
Klug valve gear 176

LADDERS, 612
Lagging, boiler, 554

— cylinder, 139

— materials, 139, 554

— steam pipe, 43ii

Latent heat of steam, 468, 694
I^w of comparison, 353
Lead, angle of, 168

— exhaust, 168

— linear, 168

Leading edge of propeller blade, 362
Lead, melting point of, 698

I

Lead piping, 447

— specific gravity, 698
Levers, valve, 226
Lifting gear, 614

Lift of pump valves, 286

— of valves, 140, 415
Lignum vitae strips, 349
Limit of elasticity, 702
Linear lead, 168

Lloyd's rules for pumps and i>umping ar-
rangements, 329

— rules for shafts, 211

— rules for spare gear, 629

— rules for the construction of boilers,

502,580
Load on safety valve, 570

— on slide valves, 221

Loads, breaking, for various materials,

702
Lock nuts, ring, 608
Locomotive boilers, 510

— boilers, grate surface of, 510

— boilers, heating surface of, 5 11
Logarithms, common, 674

— hyperbolic, 634

Longitudinal bracing of cylinder columns,

122
Losses by throttling or wiredrawing during

admission, 35

— due to condensing, 37

— due to friction, 3

— in the engine, 4
Lubrication, 620

— of bearings, 621
Lubricator, forced feed, 620

MACHINING the blades, 406
Main bearing bolts, 198, 243

— bearing brasses, 241

— bearings, 196, 24*)

— engines, arrangement of, 123

— exhaust, 434

— steam piping, 423

— stop valve, 140, 579, 581
Manganese bronze, specific gravity of,

699

— bronze, strength of, 703, 704
Manholes, 483
Manoeuvring valve, 139

Marine boilers, regulations affecting fittings
of, 577

INDEX.

739

Marine engine, revolutions per minute, 41
Marshall valve gear, 179
Masses, reduction of the, 62, 85
Material of blades, 394
Materials for crank shafts, 220

— strength of, 702, 704, 714

— various, breaking loads of, 702
Maximum allowable pressure on brasses,

197,240

— load, 181

Mean piston speed, 41, 59
— ■ pressure, actual, 17

— pressure, theoretical, 12

— turning moment, 65
Measuring instruments, 585
Mechanical equivalent of heat, 32

— theory of heat, 32
Melting point of copper, 698

— point of iron, 698

— pointof lead, 698

— points of various metals, 698
Mercurial pyrometer, 588
Metallic packing, 158

Metals, various, melting points of, 698
Method of handling the reversing gcjir,

256
Middendorf, resistance of ships, 355
Midship section, immersed, 355
Millimetres to inches, 682
Modulus of elasticity, 702
Moment turning, instrument for measuring

uniformity of, 599

— of inertia of circular sections, 705
Moments, bending, 707

Moulding and casting a ship's screw,

395
Moving parts of the steam engine, 60
Moving parts, weight of, 79, 92
Mud boxes, 447
MUller valve diagram, 169
Multiple expansion engines, 37

NATURAL draught, 458
— logarithms, 634
Negative slip, 370
Nodes in dbrations, 104
Normand boiler, 531
Number of blades, 367

— of revolutions, 41
Nuts, dimensions of, 608

— for bolts, 607

OBJECT of the steam jacket, 37
Oil grooves, 621
Openings in the valve face, 165
Orsat apparatus, 588
Oscillations, torsional, of shafts, 78

P ALLOGRAPH, Schlick's, 597
Paraffin model of ships, 354
Packing, Buckley's piston, 186

— Katzenstein, 158

— metallic, 158

— piston, 184

— piston-rod, 156

— rings, piston- valve, 161

— Schelling, 159

— United States, 160
Particulars of surface condensers, 309
Peabody's throttling calorimeter, 590
Pet cocks, 576

— valves, 576

Phosphor bronze, specific gravity of, 699
IT, value of, 681
Pin, crosshead, 196
Pipe connections, 409

— flanges, 409

— joints, 410

— receiver, cross section of, 128, 130

— steam, lagging, 433
Pipes and connections, 407

— feed-water, 438

— suction, 442

Hping, auxiliary steam, 433

— Ijallast, 448

— bilge, 447

— exhaust, 434

— lead, 447

— main steam, 423
Piston clearances, 189

— packing, 182

— packing, Buckley's, 186

— rings, 185

— rod packing, 181

— rods, 191

— speed, 41

— - speed, mean, 41, 59

— stroke of, 41

— valve liners, 136

— valve ports, 136

— valve rings, 161
Pistons, 183

— cast iron, 183

'40

INDEX-

Pistons, cast steel, 1H2

— steel, thicknevs of, 1H3

- weight r»f, 02
I'itch of propeller blades, variable, 268

— of rivets (sec Rules for Construction ;

of Boilers)

— of ship's screvv, 3ft2
I'Unimeter, 597
PUtes, edging, 613
Platforms, gratings, and ladders, 613
Ping cocks, 417
Plummer blocks, 341
Ports of piston \'alves, 136
Position of the engine in the ship,

104
Pounds per square inch to kilos per
square centimetre, 69(i

— to kilogrammes, 688
Power of boilers, 457

— of engines, 3

— of engines, and speed of vessel, 355

— required to drive the valves, 220
Pressure diagram, tangential, 63

— - mean, 12, 17

- (m guides, 205

— or pull on connecting rod, 58

— - tangential, on crank pin, 63
Principal dimensions of reversing engines,

264
Process of combustion, 453
Projected area of a ship's screw, 363
Propeller, 361

— blade, leading edge of, 362

— blades, allowable strain in, 383

— blades, bolted on, 403

— blades, strength of, 384

— blades, variable pitch of, 368

- l)os8, 401

— - lioss, strengthening the, 4fll

— - cap, 404

- diameter ratio of, 378

— efficiency of, 370

— - fixing the blades, 403

— moulding and casting of, 363, 395

— - pitch of, 362

— shaft, 345

— slip, 368
--thrust of, 337
Properties of cast iron, 702

— - of saturated steam, 694
Proportion of cylinders, 43
Pump, air, 284

Pump aurangements taken from actual
practice, 324

— ballast, 326

— Wlge, 321

— Blake, 318

— body, 297

— circulating, 294

— direct driven, 310

— piston and piston rod, 297

— rods, 324

— steward's, 323

— suction and deliver)-, 297

— valves, 297, 314

— valves, lift of, 286

— VVorthington, 315
Pumps, 284

— centrifugal, 298

— feed, 310

— Weir, 317
P>Tometer, mercurial, 588

QUADRANT blocks, 226
Quadruple expansion engine, 9

RADIAL acceleration, 62
Ramsbottom rings, 185
Rankine's diagram, 18
Ratios of heating surface to grate surface,

469
R^umur thermometer, 692
Receiver, 27

— pipe, cross section of, 128, 130
Reciprocals, 634

Reduction of the masses, 62

Regulatioas affecting marine boiler fittings,

577
Relief valves, cylinder, 144
Required amount of feed- water, 310
Resistance of ships, 353

— of ships, Middendorf, 355

— to bending, 707

— to compression, 702

— to torsion, 711

— wave-making, 352
Reversing engine, all round, 261

— engine, Brown's, 256

— engine, direct-acting, 256

— engines, principal dimensions of, 264

— gears, 251

— shaft and lever, 251

INDEX.

741

Reversing the engine, 256

— valve, 261

Revolutions of an engine, degree of regu-
larity of, 599

— per minute of a marine engine, 41
Revolving slide valve, 261

Ring lock nuts, 608
Rings, junk, 190

— of thrust block, 336

— piston, 185

— Ramsbottom, 185
Rivets, boiler, 486

— pitch of (see Rules for Construction of

Boilers)

Rods, valve gear, 220

Rule of mean ordinates, 5

Rules for the construction of boilers —
Bureau Veritas, 498 ; German Lloyd's,
488 ; Hamburg Standard, 493 ; Lloyd's,
502

— for the construction of cylinders, 139

SAFETY" valve, boiler, 569, 577
— valve, load on, 570
Salinometer cock, 576
Saturated steam, properties of, 694
Schelling packing, 159
Schlick's pallograph, 597
Schlick system of balancing, 85
Screw connections, 409

— spanners, 609

— stays, 492, 497, 500, 503
Scum cock, 577

Seating, boiler, 618

— engine, 617

Sectional areas of ship's boiler flues, 459,

480
Securing the blades of a propeller, 403
Separately driven air pumps, 292

— driven steam feed pumps, 314
Separator, steam, 421

Shaft brake, 342

— couplings, 343

— crank, 208

— propeller, 345

— thrust, 335

— transmission, .340

— stuffing box, 341, 352

Shafts, sleeves for, in stern tul)e, 347

— torsional oscillations of, 78

— tunnel or intermediate, 340

Shearing stresses, 702
Ship's hull, vibration of, 104

— resistance, 353

— resistance, calculation of, 355

— screw, 361

— screw, developed area of, 363

— screw, dimensions and shape of, 374

— screw, efliciency of, 370, 376

— screw, empirical formulx for calculation

of, 374

— screw, extreme diameter ratio of, 371

— screw, fixing the blades, 403

— screw, moulding and casting of, 395

— screw, pitch of, 362

— screw, projected area of, 363

— screw, Taylor's method of calculation of,

376

— screws, 361

Ships, frictional resistance of, 355

— paraffin model of, 354
Simplex feed pump, 316
Simpson's formula, 6
Sines, table of, 677

Single eccentric valve gear, 176

— expansion engines, 9

Sleeves for shafts in stem tube, 347
Slide valve, 161

— valve, double ported, 166

— valve, load on, 221

— valve, principal dimensions of, 166

— valve, revolving, 261
Slip, 368

— apparent, 368

— n^ative, 370
Slipper, 200-
Sluice valves, 417

Small water-tube boilers, 539

Smoke box, 549

Spanners, 609

Spare gear, German Lloyd's rules for, 627

— gear, Lloyd's rules for, 629
Specific gravity of brass, 699

— gravity of bronze, 699

— gravity of coals, 700
- gravity of copper, 699

— gravity of Delta metal, 699

— gravity of gases, 701
-- gravity of iron, 699

— gravity of lead, 699

— gravity of manganese bronze, 699

— gravity of phosphor bronze, 699

— gravity of wood, 699

742

INDEX.

Speed, circumferential, of crank, 58

— curves, 353

— of piston, 41

— of the screw, 3d8

— of vessel and power of engine, 355
Square feet to square metres, 687

— metres to square feet, 686
.Squares, 634

Starting tTleengine, 125

— valve, 143

Stays, screwed, 492, 497, 500, 503
Steam consumption, 310

— diagram, efficiency from, 17

— dryness, fraction of, 590

— expansion of, 10

— generation of, 468

— jackets, 37

— pipe lagging, 433

— piping, 423

— piping, auxiliary, 433

— piping, main, 423

— pressures, determination of, 29

— saturated, properties of, 694

— separators, 421

— space, 474

— superheater, 520

— tables, 694

— theoretical work of 1 lb. , 34

— to cylinders, 143

— to jackets, 37

— total heat of, 694

— velocity of, 128
Steel castings, 702

— crucible, strength of, 702
Stephenson's link motion, 174, 227

— valve gear, 174
Stern tube, 346

— tube stuffing boxes, 352
Steward's pumps, 323
Stokehold, closed system, 5<S4

— air pressure in, 565
Stop valve, Imiler, 573

— valve, engine, 140

— valve, main, 140, 579, 581
Stream-line wake, 368
Strength of blades, 384

— of brass, 702

— of bronze, 702

— of cast iron, 702

— of crucible steel, lif2

— of cylinder walls, 130

— of Delta metal, 702

I

Strength of manganese bronze, 703, 704

— of materials, 702

— of nickel steel, 702

— of propeller blades, 384

— of steel castings, 702

— of struts, 712

— of various metals, 702

— of wood, 703

— torsional, 711

Strengthening the propeller boss, 401
Stress in propeller blade due to thrust, &c.,

384
Stresses in blades due to centrifugal force,

388

— in columns and framing, 249

— shearing, 702
Strips, lignum vitse, 349
Stroke of engine, 40

— of valve, 415
Struts, strength of, 712
Studs for cylinder cover, 133
Study of the valve gear, 88
Stuffing boxes, bulkhead, 341

— boxes, cylinder, 156

— boxes, stern tube, 352
Suction air vessel, 441

— pipe, 286, 442
Superheater, steam, 520
Surface condensers, 272

— condensers, particulars of, 309

— tube of condenser, 273

— wetted, 355, 358

TABLE of calorific values, 455
— of circumferences, 634

— of coefficient, 356

- of constants, 6, 354

— of cotans, 384, 679

— of cube roots, 634

— of cubes, 634

— of dimensions of various ships, 42

— of tangents and cotangents, 679

- of various values, 681

Tables of dimensions of bolts, 607

— of flanges, 409

— of sines and cosines, 677

— steam, 694
Tail end shaft, 345

— rods, 192

Tangential pressure diagram, 63

— pressure on crank pin, 63

INDEX.

743

Tangents, table of, 679

Tank, float, 441

Taylor, calculation of a ship^s screw,

376
Test pressures of boilers, i)79, 580,

581

— pressures of cylinders, 138
Theoretical diagram, 17

— efficiency, 17

— work of 1 lb. of steam, 34
Thermal units, British, 32
Thermometer, Celsius, 692

— Fahrenheit, 585, 692

— Reaumur, 692
Thickness of blade at tip, 393

— of cap, 245

— of cylinder jacket, 131

— of cylinder liner, 130

— of cylinder wall, 131
- of steam pipes, 428

Thornycroft boiler, 540
Throttle valve, 141
Through way cock, 417

— valve, 412

Thrust block collars, 336

— shaft rings, 335

— blocks, 336

— caps, horseshoe, 337

— collars, 335

— of the propeller, 337

— shaft, 335

— indicator, 599
Torsion, resistance to, 711 ^
Torsional oscillations of shafts, 78

— strength, 711
Total expansion, 10
Trick valve, 166
Tube, stern, 346

— surface of condenser, 273
Tubes, cylindrical boiler, 480
Tunnel shaft bearings, 341

— shafts, 340
Turning gear, 266

— gear, dimensions of wheels, 269

— moment, 63

— moment, instrument for measuring uni-

formity of, 599

— moment, method of determining the,

63

— moment of the multiple crank engine,

63
Tw^in-cylinder engines, 9

UNDER water fittings, 418
United States packing, 160
Uptake and funnel for war vessel, 552
Useful work, 6
Utilisation of steam in the engine, 32

VACUUM gauge, 585
— in condenser, 39
Values of ir, ^, », e, 681
Valve, boiler blow-off, 576

— boiler stop, 573

— casing, 134

— chest, 134

— covers, 134

— cylinder relief, 144

— diagram, MUller, 169

— diagram, Zeuner, 171

— double beat, 141

— engine stop, 141

— feed check, 575

— gear, 176

— gear, Heusinger*s, 181

— gear, Joy's, 1 79

— gear, Klug, 176

— gear, Marshairs, 179

— gear rods, 220

— gear, single eccentric, 176

— gear, Stephenson's, 174

— gear, study of the, 88

— levers and quadrants, 226

— lift of, 140, 415

— manoeuvring, 139

— pump, lift of, 286

— reversing, 261

— rod guides, 227

— rods, 222

— safety, lx)iler, 569, 571

— seats, 413

— slide, revolving, 261

— starting, 141

— stroke of, 166

— throttle, 141

— through way, 412

— trick, 166
Valves, 412

— air pump, 286

— Beldam, 286

— byepass, 143, 441

— cylinder relief, 144

— density, 576

— discharge, 421

'44

INDEX.

VaWes, friction of, 2*21

— indiarubber, 286

— Kinghorn, 286

— Kingston, 418

— pet, 576

— piston, ports of, 136

— relief, 144

— sluice, 417

Variable pitch of propeller blades, 368
Variation in crank pin velocity, 69
Variations in cut-off, 172

— in turning moment on the shafting, 74
Various details, 605

— values, Uble of, 681
Velocity of steam, 128

Ventilation of engine and boiler rooms, 624
Vibraaonofthehull, 1(>4
Vibrations, nodes in, 104

WAKE, streamline, 368
Warming up the cylinders, 37(
143
Water gauges, 57«>

— tube boilers, 512

— tube boilers, grate surface of, 513

— tube boilers, heating surface of, 513
Wave-making resistance, 352

Weight of boilers, 715, 716

— of cylindrical boilers, 716

— of engines, 715

— of moving parts, 79, 92

— of piston, 92
Weights of valve gear, 93
Weir's pump, 317
Wetted surface, 355, 358

White metal, melting point of, 608

— metal, specific gravity of, 699
Whitworth thread, 606 '

— thread, bolts with, 606
Woods, specific gravity of, 699

— strength of, 703

Work done during admission, 9

— done during expansion, 11

— theoretical, of 1 lb. of steam, 34

— usefiil, 6
Worm wheels, 269
Worthington pump, 315

Y
Z

ARROW lx>iler, 526

KUNKR valve diagram, 171

Printed at The Darien Press. Edinburgh.

1

Fcap. SivOf 750 pages, with ntunerous lUttstrations, boand in Limp
Leather, Gilt Edges, Rounded G>rners, 12s. 66, net.

THE

NAVAL ARCHITECT'S AND SHIPBUILDER'S

POCKET-BOOK

OF

FORMULAE, RULES, AND TABLES

AND

Marine Ens:ineer'5 and Surveyor's Handy Book of Reference.

By CLEMENT MACKROW,

Member ot the iuMtltutlan ot Nmvml Arcbitects,
Lecturer on Nmvml Arcbitecture mi tbe Bow mnd Bromhy Inmiitute,

EIGHTH EDITION, carefully Revised and Enlarged.

SUMMARY OF CONTENTS-

Algebraical Signs and Symbols — Decimal
Fractions — Evolution — Trigonometry —
Practical Geometry- — Mensuration of Super-
ficies— Mensuration of Solids — Mensuration
of the Surfaces of Solids — Centres and
Moments of Figures — Moments of Inertia
and Radii of Gyration— Algebraical Expres-
sions for Simpson *s Rules, &c. — Mechanical
Principles — Centre of Gravity — Laws of
Motion — Displacement, Centre of Buoy-
ancy, &c. — Centre of Gravity of Ship's
Hull, &c — Stability Curves and Meta-
centres — Sea and Shallow-water Waves —
Rolling of Ships — Propulsion and Resist-
ance of Vessels — Speed Trials — Sailing,
Centre of Effort, &c. — Distances down
Rivers — Steering and Rudders of Vessels —
Launching Calculations and Velocities-
Weight of Material and Gear — Gun Par-
ticulars and Weights — Standard Wire and
Plate Gauges — Riveted Joints and Riveting
— Strength and Tests of Materials, &c. —
Notes on Materials — Binding and Shearing
Stresses, &c. — Moments of Inertia —
Strength of Shafting, Pillars, Wheels, cStc.
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Sections, Catenarian Curves, &c.— Mecha-
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Regulations for Boilers and Engines — Board
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of Chains, &c. — Lloyd's Rules for Ships —
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&c. — Ships' Fittings and Tests — Seasoning
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Timber, Plastering, Bricklaying — Expan-
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Rules for Horse-power, Screw Propellers,
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of Vessels — Lloyd's Rules for Yards, Masts,
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and English Terms — English and Foreign
Weights and Measures — Decimal Equiva-
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Reckoners —Tables of Circular Measure —
Tables of Areas of and Circumferences of
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Roots of Numbers — Tables of Logarithms
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&c.

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««-

T. List of British Standard Sections (included in No. 6), Is. net.

This Publication gives the sizes (but not the Properties, which are to be found in
Publication No. 6) of the following sections : — Equal and Unequal Angles, Bulb Angles,
Bulb Tees, Bulb Plates, Zed Bars, Channels, Beams, and Tee Bars.

2. Specification and Sections of British Standard Tramway Rails and

Fishplates, 21s. net.

This Publication contains two Specifications — (i) Tramway Rails, (2) Fish Plates.
The sections for straight track vary from 90 lbs. to no lbs. per yard, each section
increasing in weight by 5 lbs. per yard. Sections are also given of the curved rails
corresponding to each weight of rail for straight track. Fully dimensioned diagrams of
every section are included.

3. Report on the Influence of Gauge Length and Section of Test

Bar on the Percentage of Elongation, by Professor W. C.
Unwin, F.R,S., 2/6 net.

This Publication contains a very full Report on a series of Tensile Tests carried out by
Prof. Unwin in 1903, on Ship and Boiler Plates, and were made for the Committee
during their deliberations in connection with the forms of Standard Tensile Test Pieces.

4. Properties of Standard Beams (included in No. 6). Demy 8vo,

sewed, Is. net.

5. Report on Standard Locomotives for Indian Railwajrs (to be

published at a later date).

This Publication contains the Recommendations of the Committee on Types of
Standard Locomotives for Indian Railways. It includes Instructions to Designers, and
Drawings for three classes of Metre Gauge and two classes of Broad Gauge Locomotives.

6. Properties of British Standard Sections, 6s. net.

This Publication includes all the information contained in Nos. I and 4, together
with Areas, Weights, Moments of Inertia, Moments of Resistance, &c, for all Standard
Sections of Structural Steel. Demy 8vo, cloth.

7. British Standard Tables of Copper Conductors and Thicknesses of

Di-Electric, 2/6 net.

This Publication includes the Resolutions of the Committee with respect to Standards
for Copper Conductors, and contains Tables of Dimensions for Single, Concentric, and
Stranded Conductors, Armour, and Di-£lectric.

8. British Standard Specification for Tubular Tramway Poles, 6s. net.

This Specification contains the Sizes and Dimensions of Standard Tubular Tramway
Poles, together with the Tests recommended by the Committee.

9. British Standard Specification and Sections of Bull-Headed Railway

Rails, 10/6 net.

This Publication includes the Specification and Sections of Bull-headed Railway
Rails. The Sections vary from 60 to xoo lbs. per yard, each Section increasing in
weight by 5 lbs. per yard. Fully dimensioned Diagrams of every Section are included,
together with Tables of Dimensions, and Sketches showing the forms of Standard
Tensile Test Pieces.

1 o. British Standard Tables of Pipe Flanges, 2/6 net.

This Publication contains particulars of Flanges for different ranges of Steam Pressure
up to 325 lbs. per square inch, and Water Pressure up to 200 lbs. per square inch.
Dimensions of the Standard Bends and Tees recommended by the Committee are also
given.

[ Contintud on next page.

Publications of Engineering Standards Committee— contiaued.

1 1. British Standard Specification and Sections of Flat- Bottomed Railway

Rails, 10/6 net.

This Publication will include the Specification and Sections of Flat-Bottomed Rail-
way Rails. The Sections vary from 20 to 100 lbs. per yard, each Section increasing in
weight by 5 lbs. per yard. Fully dimensioned Diagrams of every Section are included,
together with Tables of Dimensions, and Sketches showing the forms of Standard
Tensile Test Pieces.

12. British Standard Specification for Portland Cement, 2/6 net.

This Specification contains the Tests for Portland Cement recommended by the
Committee. Dimensioned Diagrams are given of the Standard Briquette, Jaws for
holding same, Needle for ascertaining setting time, and the Apparatus for conducting
the *' Le CAa/e/ier" Test.

t

1 3. British Standard Specification for Structural Steel for Shipbuilding:,

2/6 net.

This Specification includes the Tests recommended by the Committee for Steel Plates,
Angles, Channels, Bars, and Rivets used for Shipbuilding, together with Sketches
showing the forms of the Standard Tensile Test Pieces.

14. British Standard Specification for Structural Steel for Marine

Boilers, 2/6 net.

JThis Specification includes the Tests recommended by the Committee for Steel Plates,
Angles, Stay Bars, and Rivets used in the Construction of Marine Boilers, together with
Sketches showing the forms of the Standard Tensile Test Pieces.

15. British Standard Specification for Structural Steel for Bridg^es (to

be published shortly), 2/6 net.

' This Specification will include the Tests for Steel used in Bridge Construction,

together with Sketches showing the forms of Standard Tensile Test Pieces.

16. British Standard Specification for Telegraph Material (to be pub-

lished shortly), 6s. net.

This Publication will include the Resoluiions of the Committee with respect to
Standards for Copper Conductors, and will contain Specifications for general Telegraph
Material, together with Tables for Hard and Annealed Copper Wire.

17. Interim Report on Electrical Machinery (C.L. 224).

This Report gives Standard Pressures and Frequency ; Speeds for Direct Current
Generators, D.C. Motors and A.C. Induction Motors. Definitions and Standard Test
Conditions are also included ; the latter are, however, still under consideration, and
must not, therefore, be regarded as final.

18. Forms of Standard Tensile Test Pieces (C.L. 116).

This contains Dimensioned Diagrams, and will l>e inserted in all Specifications in
which reference is made to the Standard Tensile Test Pieces.

The above publications can be obtained from Crosby Lockwood & Son,
Publishers to the Committee, 7 Stationers' Hall Court, Ludgate Hill, London,
E.C., or at the Offices of the Committee, 28 Victoria Street, Westminster,
London, S.W.

NOTE, — With the exception of Nos, 4 and 6, the Volttmes are issued in Foolscap size,

hound in strong Paper Covers.