Copyright 2009 Applied Research Associates, Inc. All Rights Reserved.

Detailed Puncture Analyses of

Various Tank Car Designs

Final Report – Revision 1

January, 2010

Prepared by:

Steven W Kirkpatrick, Ph.D.

Applied Research Associates, Inc.

2672 Bayshore Parkway, Suite 1035

Mountain View, CA 94043
Applied Research Associates, Inc.

2010

ii

Table of Contents

Section Page

Acknowledgements...................................................................................................................... xiii

Executive Summary..................................................................................................................... xiv

1 Introduction............................................................................................................................. 1

1.1 Problem Definition and Approach .................................................................................. 1

1.2 Tank Car Design Constraints.......................................................................................... 4

2 Model Development and Preliminary Impact Analyses......................................................... 7

2.1 Introduction..................................................................................................................... 7

2.2 Tank Car Model Development........................................................................................ 7

2.3 Side Impact Analyses and Model Validation.................................................................. 9

2.3.1 Test 1 - 14 mph Side Impact (17x23 impactor)........................................................ 10

2.3.2 Test 2 - 15 mph Side Impact (6x6 Impactor)............................................................ 21

3 Material Constitutive and Damage Models .......................................................................... 29

3.1 Introduction................................................................................................................... 29

3.2 Material Stress-Strain Behavior.................................................................................... 29

3.2.1 TC128B Material Properties..................................................................................... 29

3.2.2 Strain Rate Effects .................................................................................................... 34

3.2.3 Properties of Other Tank Car Materials.................................................................... 39

3.3 Bao-Wierzbicki Failure Surface ................................................................................... 39

3.4 Material Characterization Testing and Analyses.......................................................... 42

3.4.1 Tensile Tests and Analyses....................................................................................... 43

3.4.2 Notched Round Bar Tensile Tests and Analyses...................................................... 46

3.4.3 Tensile-Shear Tests and Analyses ............................................................................ 48

3.4.4 Punch Tests and Analyses......................................................................................... 51

3.4.5 Bulk Fracture Charpy Machine (BFCM) Tests and Analyses.................................. 58

3.4.6 Characterization of Other Structural and Energy Absorbing Materials.................... 67

4 Detailed Tank Puncture Analyses......................................................................................... 70

4.1 Introduction................................................................................................................... 70

4.2 Tank Car Puncture Model............................................................................................. 70

4.3 Assessment of Baseline Tank Puncture Conditions...................................................... 73

4.3.1 Effects of Tank Pressure and Diameter .................................................................... 79

4.3.2 Analyses of Tank Jacket Retrofit Designs................................................................ 81

4.3.3 Effects of Impactor Size............................................................................................ 85
Applied Research Associates, Inc.

2010

iii

4.3.4 Effects of Outage Volume ........................................................................................ 98

4.3.5 Effects of Impact Speed.......................................................................................... 105

4.3.6 Summary of the Side Impact Analyses................................................................... 106

5 Detailed Head Impact Analyses.......................................................................................... 112

5.1 Introduction................................................................................................................. 112

5.2 Head Impact Analyses and Model Validation ............................................................ 112

5.2.1 Head Test 1 – Baseline Head Impact Behavior ...................................................... 112

5.2.2 Head Puncture Model ............................................................................................. 115

5.2.3 Comparison of Head Test 1 with Analyses............................................................. 115

5.3 Assessment of Head Impact Puncture Conditions...................................................... 116

5.3.1 Baseline Impact Analyses on Unpressurized Heads............................................... 119

5.3.2 Baseline Impact Analyses on Pressurized Heads.................................................... 122

5.3.3 Analyses of Head Retrofit Designs......................................................................... 132

5.3.4 Head Impact Analyses on Ethylene Oxide Tank Heads......................................... 135

5.3.5 Effects of Impactor Size.......................................................................................... 138

5.3.6 Effects of Impact Speed.......................................................................................... 142

5.3.7 Summary of the Head Impact Analyses.................................................................. 143

6 Detailed Evaluation of Alternate Protection Concepts....................................................... 150

6.1 Evaluation of Foam Protection Systems..................................................................... 150

6.1.1 Validation of Foam Modeling................................................................................. 151

6.1.2 Head Test 2 Measurements and Analyses (Foam System)..................................... 189

6.1.3 Side Impact Puncture Analyses of Foam Systems.................................................. 195

6.1.4 Head Impact Puncture Analyses of Foam Systems ................................................ 199

6.2 Evaluation of EMS Protection Systems...................................................................... 202

6.2.1 Head Impact Puncture Analyses of a Corrugated EMS Systems ........................... 202

6.2.2 Head Test 3 Measurements and Analyses (EMS System)...................................... 212

6.2.3 Side Impact Puncture Analyses of a Corrugated EMS Systems............................. 219

6.2.4 Side Impact Puncture Analyses of an Egg Crate EMS Systems............................. 227

6.2.5 Summary of the EMS Puncture Analyses............................................................... 230

6.3 Double Tank Head Concept Puncture Analyses......................................................... 231

6.4 Analyses of Composite Tank Protection Concepts..................................................... 240

6.4.1 Validation of Composite Modeling ........................................................................ 240

6.4.2 Tank Car Side Impact Analyses with a Composite Jacket ..................................... 250

7 Summary and Conclusions ................................................................................................. 259

8 References........................................................................................................................... 267
Applied Research Associates, Inc.

2010

iv

List of Figures

Figure Page

Figure 1. Updated model of a 105J500W pressure tank car. ...................................................... xix

Figure 2. Calculated Test 1 impact response with cutaway showing lading. .............................. xx

Figure 3. Comparison of the measured and predicted Test 1 force-deflection curves. .............. xxi

Figure 4. Bao-Wierzbicki failure surface and tests used for model calibration......................... xxii

Figure 5. Validation of the notched round bar test behavior for TC128B................................ xxiii

Figure 6. Validation of the combined tensile shear test behavior for TC128B. ....................... xxiv

Figure 7. Simulation of the punch test on the thin TC128B plate material. .............................. xxv

Figure 8. Force-deflection curves for three punch test configurations on TC128B. ................. xxv

Figure 9. Simplified tank model analysis with Bao-Wierzbicki failure assessment................. xxvi

Figure 10. Tank puncture model Test 2 analysis with Bao-Wierzbicki failure assessment. ... xxvii

Figure 11. Calculated puncture behavior of a head and head shield. ..................................... xxviii

Figure 12. Calculated puncture forces as a function of system thickness. ............................... xxix

Figure 13. Loading and failure mechanism for the tank impact and puncture. ......................... xxx

Figure 14. Calculated puncture energies as a function of system thickness............................. xxxi

Figure 15. Relationship between impact speed and kinetic energy for the ram car. ............... xxxii

Figure 16. Puncture prevention thickness requirements for various impact speeds. ............... xxxii

Figure 17. Summary of puncture energies for chlorine tank cars for 

various impactor sizes. ......................................................................................... xxxiii

Figure 18. The effects of jacket thickness on puncture energy for a 500 lb Cl tank. ............. xxxiv

Figure 19. The effects of jacket thickness on puncture energy for a 300 lb EO tank.............. xxxv

Figure 20. The effects of head shield thickness on puncture energy for a 500 lb head........... xxxv

Figure 21. Analyses of advanced protection concepts............................................................ xxxvi

Figure 22. Alternative tank protection concepts.................................................................... xxxvii

Figure 23. Comparison of the force-deflection behaviors for the 

corrugated EMS concepts................................................................................... xxxviii

Figure 24. EMS puncture forces as a function of outer layer thickness. ................................ xxxix

Figure 25. Normalized peak punch test loads versus aerial density for various materials. .......... xl

Figure 26. Summary of side impact puncture performance and goals. ...................................... xlii

Figure 27. Summary of head impact puncture performance and goals. ....................................xliii

Figure 28. Number of Accidents with at Least One Car Releasing Hazardous Materials [1]....... 2
Applied Research Associates, Inc.

2010

v

Figure 29. Calculated derailment behavior for the 36-car train model [9]. ................................... 3

Figure 30. Characteristic force-deflection curve for a tank impact. .............................................. 4

Figure 31. Maximum Dimensions for Plate B [11]. ...................................................................... 5

Figure 32. Diagram of a typical North American general purpose tank car [5]. ........................... 7

Figure 33. Original model generated for an 112A340W pressure tank car. .................................. 8

Figure 34. Updated model generated for a 105J300W pressure tank car. ..................................... 9

Figure 35. Ram car and target tank car staged prior to Test 1..................................................... 11

Figure 36. Shell side impact behavior in Test 1 for Tank Car 3069............................................ 12

Figure 37. Post-test dents in the jacket and shell of Tank Car 3069............................................ 13

Figure 38. Updated model generated for a 105J500W pressure tank car. ................................... 14

Figure 39. Calculated Test 1 impact response with cutaway showing lading. ............................ 15

Figure 40. Comparison of the measured and predicted ram car motions. ................................... 16

Figure 41. Comparison of the measured and predicted Test 1 tank displacements..................... 17

Figure 42. Comparison of the measured and predicted Test 1 force-deflection curves. ............. 18

Figure 43. Measured Test 1 tank pressures at various positions in the lading............................. 19

Figure 44. Comparison of the measured and predicted Test 1 tank pressure. ............................. 20

Figure 45. Test 2 target Tank Car 3074 mounted on support skids............................................. 21

Figure 46. Modified ram car and ram head used in Test 2 .......................................................... 22

Figure 47. Puncture response observed in the full-scale impact Test 2....................................... 23

Figure 48. Measured Test 2 ram car average longitudinal acceleration. ..................................... 24

Figure 49. Measured Test 2 ram car and tank displacements...................................................... 25

Figure 50. Measured Test 2 force-displacement behavior and energy dissipation...................... 25

Figure 51. Measured Test 2 tank car internal pressures............................................................... 26

Figure 52. Updated target tank car and ram car models for Test 2.............................................. 27

Figure 53. Predicted tank damage for the Test 2 impact conditions............................................ 28

Figure 54. Comparison of measured and calculated Test 2 force-displacement response. ......... 28

Figure 55. Material testing data for different TC128B materials. ............................................... 30

Figure 56. Comparison of engineering and true stress-strain data for TC128B. ......................... 32

Figure 57. Tabular true stress curve developed for the TC128B constitutive model. ................. 33

Figure 58. Comparison of the measured and calculated TC128B tensile test. ............................ 33

Figure 59. Example strain rate effects data for structural steels.................................................. 35

Figure 60. Evaluation of TC128B strain rate effects data [37].................................................... 35
Applied Research Associates, Inc.

2010

vi

Figure 61. Yield stress rate effects data and Cowper-Symonds model fit................................... 36

Figure 62. Measured TC128B tensile stress-strain curves at various strain rates. ...................... 38

Figure 63. Tabular true stress curve developed for the various tank car steels. .......................... 39

Figure 64. Local damage criterion for tensile ductile fracture analyses. ..................................... 41

Figure 65. Bao-Wierzbicki failure surface and tests used for model calibration......................... 42

Figure 66. Dimensions of the specimen used in tensile testing. .................................................. 43

Figure 67. Necking behavior observed in the TC128B round bar specimens. ............................ 44

Figure 68. Validation of the tensile test behavior for TC128B.................................................... 45

Figure 69. Validation of the tensile test behavior for A1011. ..................................................... 45

Figure 70. Validation of the tensile test behavior for A516-70. .................................................. 46

Figure 71. Test setup for the TC128B notched round bar specimens.......................................... 47

Figure 72. Validation of the notched round bar test behavior for TC128B................................. 48

Figure 73. Specimen geometries for the combined tensile shear tests. ....................................... 49

Figure 74. Photograph of the combined tensile shear test configuration..................................... 50

Figure 75. Analysis of the specimen behavior in the pure shear orientation............................... 50

Figure 76. Validation of the combined tensile shear test behavior for TC128B. ........................ 51

Figure 77. Simulation of the punch test on the thin TC128B plate material. .............................. 52

Figure 78. Dimensions of the 2-inch-diameter punch. ................................................................ 53

Figure 79. Dimensions of the man hole cover support plates...................................................... 53

Figure 80. Punch test failure mode for the thin TC128B plate material...................................... 54

Figure 81. Simulation of the punch test on the thin TC128B plate material. .............................. 55

Figure 82. Force-deflection curves for three punch test configurations on TC128B. ................. 56

Figure 83. Comparison of the calculated and measured punch test plate profile. ....................... 56

Figure 84. Comparison of the punch loads of the single and layered 90 XF specimens. ............ 57

Figure 85. Comparison of the measured and calculated 90 XF punch behavior. ........................ 58

Figure 86. Photograph of the Bulk Fracture Charpy Machine..................................................... 59

Figure 87. BFCM Test Specimen Dimensions. ........................................................................... 60

Figure 88. Dimensions of the blunt BFCM striker tup. ............................................................... 60

Figure 89. Dimensions of the sharp BFCM striker tup................................................................ 61

Figure 90. Preliminary finite element analysis of a BFCM test. ................................................. 62

Figure 91. Observed BFCM specimen deformations at different impact velocities.................... 62

Figure 92. Calculated BFCM specimen deformations at different impact velocities.................. 63
Applied Research Associates, Inc.

2010

vii

Figure 93. Observed BFCM specimen failure mode for the blunt striker. .................................. 63

Figure 94. Calculated BFCM specimen failure mode for the blunt striker. ................................ 64

Figure 95. Finite element model of the BFCM test. .................................................................... 64

Figure 96. BFCM blunt striker fracture energy for different specimen thicknesses. .................. 65

Figure 97. BFCM sharp striker fracture energy for different specimen thicknesses................... 66

Figure 98. Calculated and measured BFCM blunt striker specimen deformations. .................... 67

Figure 99. Calculated and measured BFCM sharp striker specimen deformations..................... 67

Figure 100. Unconfined compression testing and analysis of foam. ........................................... 68

Figure 101. Tank model and impact zone mesh used for side impact puncture analyses. .......... 70

Figure 102. Calculated tank car impact behavior using two different lading models. ................ 71

Figure 103. Simplified tank model analysis with Bao-Wierzbicki failure assessment................ 72

Figure 104. Tank puncture model Test 2 analysis with Bao-Wierzbicki failure assessment. ..... 73

Figure 105. Detailed impact and puncture sequence for a 300 lb chlorine car............................ 75

Figure 106. Calculated puncture initiation and fracture progression........................................... 76

Figure 107. Side impact force-deflection curves and puncture energies for 

chlorine tank cars....................................................................................................... 76

Figure 108. Calculated energy balance for the 600 lb chlorine tank car impact (R1F)............... 77

Figure 109. Side impact force-deflection curves and puncture energies for the 

EO Tank Cars. ........................................................................................................... 78

Figure 110. Side impact comparison for tank cars at different pressures and diameters............. 79

Figure 111. Evaluation of the effects of pressure on a 500 lb Cl tank......................................... 80

Figure 112. Evaluation of the effects of diameter and thickness at 50 psi. ................................. 81

Figure 113. Evaluation of a 500 lb chlorine tank cars retrofit with thicker jackets..................... 82

Figure 114. The effects of jacket thickness on puncture energy for a 500 lb Cl tank. ................ 83

Figure 115. Analyses of 300 lb EO tanks with different thickness jackets. ................................ 84

Figure 116. The effects of jacket thickness on puncture energy for a 300 lb EO tank................ 84

Figure 117. Tank model and impact zone mesh used for the 12x12 inch impactor. ................... 86

Figure 118. Evaluation of the effects of the ram size on a 500 lb chlorine tank. ........................ 86

Figure 119. Calculated puncture initiation and fracture progression (12 inch x12 inch ram)..... 87

Figure 120. Evaluation of the effects of the ram size on a 600 lb chlorine tank. ........................ 87

Figure 121. Effect of impact speed on the puncture behavior of the 600 lb Cl tank. .................. 88

Figure 122. Evaluation of the effects of the reinforced jacket with the large ram size. .............. 89
Applied Research Associates, Inc.

2010

viii

Figure 123. Summary of puncture forces for chlorine tank cars for various impactor sizes....... 90

Figure 124. Summary of puncture energies for chlorine tank cars for various impactor sizes. .. 90

Figure 125. Calculated energy balance for the 600 lb Cl tank car 12x12 impact (R1J).............. 92

Figure 126. Pressure-volume relationship used for the tank control volume (10.6% outage)..... 92

Figure 127. Effect of variable internal pressure on 500 lb tank impact response........................ 93

Figure 128. Effect of variable internal pressure on 600 lb tank impact response........................ 94

Figure 129. Calculated internal pressure variations in the 500 lb tank impact analyses. ............ 94

Figure 130. Calculated internal pressure variations in the 600 lb tank impact analyses. ............ 95

Figure 131. Updated 500 lb tank impact analysis with different size impactors......................... 96

Figure 132. Updated 600 lb tank impact analysis with different size impactors......................... 97

Figure 133. Summary of puncture forces for chlorine tank cars for 

various impactor sizes. .............................................................................................. 97

Figure 134. Summary of puncture energies for chlorine tank cars for 

various impactor sizes. .............................................................................................. 98

Figure 135. Model of a 23,000 gallon DOT-111A100W tank for analysis of 

outage volume effects................................................................................................ 99

Figure 136. Control volume pressure curves for various outages between 1 and 18 percent.... 100

Figure 137. Calculated impact and puncture behaviors for different outage volumes. ............. 101

Figure 138. Force-deflection curves and puncture energies for different outage volumes. ...... 102

Figure 139. Control volume pressures for impacts with different outage volumes................... 103

Figure 140. Effect of the outage volume on the puncture energy in side impacts..................... 104

Figure 141. Force-deflection curve and impact energy dissipation for an empty tank.............. 104

Figure 142. Side impact damage distribution for 1% and 18% outage volumes....................... 105

Figure 143. Effect of impactor speed on the puncture of a 500 lb Cl tank................................ 106

Figure 144. Side impact puncture force as a function of system thickness. .............................. 109

Figure 145. Loading and failure mechanism for the tank impact and puncture. ....................... 110

Figure 146. Side impact puncture energy as a function of system thickness. ........................... 111

Figure 147. Puncture energy of retrofit tank cars as a function of jacket thickness.................. 111

Figure 148. Pretest conditions of the test article and ram car for Head Test 1.......................... 113

Figure 149. Post-test damage of the test article for Head Test 1. .............................................. 114

Figure 150. Impact damage on the tank head in Head Test 1.................................................... 114

Figure 151. Calculated impact response in Head Test 1............................................................ 115

Figure 152. Calculated and measured force-deflection behaviors for Head Test 1................... 116
Applied Research Associates, Inc.

2010

ix

Figure 153. Tank head model and impact zone mesh used for the head puncture analyses...... 117

Figure 154. Calculated puncture behavior of a head and head shield. ...................................... 118

Figure 155. Calculated energy balance for the 500 lb Cl tank head and Head Shield (R1C).... 119

Figure 156. Calculated force-deflection behaviors for baseline head geometries..................... 120

Figure 157. Buckling behavior for the 11 gauge jacket in the offset head impact. ................... 121

Figure 158. Effect of jacket stiffness on the calculated head impact response.......................... 122

Figure 159. Calculated force-deflection behaviors for pressurized heads................................. 124

Figure 160. Force-deflection comparison for pressurized and unpressurized heads. ................ 125

Figure 161. Force-deflection comparison for pressurized heads and various head shields 

(0.375-inch retrofit jacket thicknesses). .................................................................. 125

Figure 162. Force-deflection comparison for pressurized heads and various head shields 

(11 gauge jackets).................................................................................................... 126

Figure 163. Force-deflection comparisons with different jacket thicknesses............................ 126

Figure 164. Calculated force-deflection behaviors at different impact speeds (300 lb head). .. 127

Figure 165. Calculated force-deflection behaviors at different impact speeds (600 lb head). .. 128

Figure 166. Models developed to assess the effects of the head shield profile. ........................ 129

Figure 167. Calculated Puncture response for the two head shield profiles.............................. 130

Figure 168. Calculated effects of the head shield profile (0.500-inch-thick shield).................. 130

Figure 169. Calculated effects of the head shield profile (0.680-inch-thick shield).................. 131

Figure 170. Calculated effects of the head shield profile (0.8281-inch-thick shield)................ 131

Figure 171. Calculated force-deflection behaviors for various unpressurized head designs..... 132

Figure 172. Comparison of unpressurized 600 lb and retrofit 300 lb tank head impacts.......... 133

Figure 173. Calculated force-deflection behaviors for various pressurized head designs......... 134

Figure 174. Comparison of pressurized 600 lb and retrofit 300 lb tank head impacts.............. 134

Figure 175. The effects of head shield thickness on puncture energy for a 500 lb head........... 135

Figure 176. Calculated force-deflection behaviors for EO Tank head geometries.................... 136

Figure 177. Calculated force-deflection behaviors for 500 lb EO and Cl Tank heads.............. 137

Figure 178. Calculated force-deflection behaviors for 300 lb EO and Cl Tank heads.............. 138

Figure 179. Impactor size effects on the 500 lb Cl tank head.................................................... 139

Figure 180. Impactor size effects on the 500 lb Cl tank head.................................................... 140

Figure 181. Summary of impactor size effects on the puncture forces for Cl tank heads. ........ 141

Figure 182. Summary of impactor size effects on the puncture energies for Cl tank heads. .... 141
Applied Research Associates, Inc.

2010

x

Figure 183. Energy balance for the 12-inch impactor and 600 lb Cl tank head. ....................... 142

Figure 184. Effect of impactor speed on the puncture of a 500 lb Cl tank head. ...................... 143

Figure 185. Calculated puncture forces as a function of system thickness. .............................. 144

Figure 186. Calculated puncture energies as a function of system thickness............................ 145

Figure 187. Extrapolated requirements for a 3 million ft-lb puncture energy. .......................... 146

Figure 188. Calculated 18 mph impact response for the thick pressurized head system........... 147

Figure 189. Calculated Bao-Wierzbicki damage levels in the thick head impact zone............. 147

Figure 190. Calculated 18 mph impact response for the thick unpressurized head system....... 148

Figure 191. Calculated puncture energies as a function of system thickness............................ 148

Figure 192. Calculated puncture forces as a function of system thickness. .............................. 149

Figure 193. Dual platen setup for unconfined compression testing of foam [70]. .................... 152

Figure 194. Compression test of a short (1-inch height), 5 pcf foam specimen [70]. ............... 153

Figure 195. Compression test of a tall (3-inch height), 7.5 pcf foam specimen [70]. ............... 154

Figure 196. Unconfined compression curves for different density foams................................. 155

Figure 197. Unconfined compression curves for different density foams................................. 156

Figure 198. Simulation of an unconfined compression test (5.4 pcf foam)............................... 156

Figure 199. Constitutive model fit and compression test simulation (5.4 pcf foam)................. 157

Figure 200. Simulation of the specimen buckling in unconfined compression tests................. 157

Figure 201. New unconfined compression testing on the 5 pcf foam. ...................................... 158

Figure 202. New unconfined compression testing on the 18 pcf foam. .................................... 159

Figure 203. Comparison of unconfined compression test series on 18 pcf foams. ................... 159

Figure 204. Uniaxial tension test results for the 18 pcf foam.................................................... 160

Figure 205. Comparison of the measured uniaxial stress behavior in 

tension and compression. ........................................................................................ 161

Figure 206. Schematic and photograph of the triaxial compression test apparatus................... 161

Figure 207. Triaxial test results for the 18 pcf foam.................................................................. 162

Figure 208. Comparison of unconfined and low confinement triaxial tests for 18 pcf foam. ... 163

Figure 209. Triaxial test results for the 18 pcf foam.................................................................. 163

Figure 210. Hydrostatic compression test results for the 18 pcf foam. ..................................... 164

Figure 211. Uniaxial strain test results for the 18 pcf foam....................................................... 165

Figure 212. Triaxial test results for the 5 pcf foam (stress difference)...................................... 166

Figure 213. Triaxial test results for the 5 pcf foam (axial Stress).............................................. 166
Applied Research Associates, Inc.

2010

xi

Figure 214. Triaxial test results for the 5 pcf foam (axial Stress).............................................. 167

Figure 215. Hydrostatic compression test results for the 5 pcf foam. ....................................... 167

Figure 216. Uniaxial strain test results for the 5 pcf foam......................................................... 168

Figure 217. Calculated and measured unconfined compression behavior 

(18 pcf foam Mat_63).............................................................................................. 169

Figure 218. Calculated cyclic compression-tension behavior 

(18 pcf foam Mat_63).............................................................................................. 170

Figure 219. Calculated and measured unconfined compression behavior 

(18 pcf foam Mat_63).............................................................................................. 171

Figure 220. True stress correction of the crush curve (18 pcf foam)......................................... 172

Figure 221. True stress correction of the crush curve (18 pcf foam Mat_75). .......................... 172

Figure 222. Comparison of measured foam crush curves from the various tests (18 pcf). ....... 173

Figure 223. Calculated and measured unconfined compression behavior (18 pcf foam Mat_75).

................................................................................................................................. 174

Figure 224. Calculated tensile behavior of a cylindrical foam specimen 

(18 pcf foam Mat_75).............................................................................................. 174

Figure 225. Comparison of measured and calculated tensile behavior 

(18 pcf foam Mat_75).............................................................................................. 175

Figure 226. Comparison of measured and calculated triaxial test behaviors 

(Mat_75 - 120 percent)............................................................................................ 176

Figure 227. Comparison of measured and calculated triaxial test behaviors 

(Mat_75 - 100 percent)............................................................................................ 176

Figure 228. Calculated and measured unconfined compression behavior 

(5 pcf foam Mat_75)................................................................................................ 177

Figure 229. Photographs of the foam sandwich panel puncture test. ........................................ 178

Figure 230. Measured and calculated foam sandwich panel punch test behavior (Mat_63)..... 179

Figure 231. Comparison of measured and calculated foam sandwich panel 

punch forces (Mat_63). ........................................................................................... 180

Figure 232. Measured and calculated foam sandwich panel punch test behavior (Mat_75)..... 181

Figure 233. Comparison of measured and calculated foam sandwich panel 

punch forces (Mat_75). ........................................................................................... 181

Figure 234. Comparison of bonded and unbonded foam sandwich panel bend tests................ 183

Figure 235. Comparison of a foam sandwich panel bend test and analysis (Mat_63). ............. 184

Figure 236. Measured and calculated response for the bonded foam 

sandwich panel (Mat_63). ....................................................................................... 184
Applied Research Associates, Inc.

2010

xii

Figure 237. Measured and calculated response for the unbonded foam 

sandwich panel (Mat_63). ....................................................................................... 185

Figure 238. Measured behaviors for the new 18 pcf foam sandwich panel tests. ..................... 186

Figure 239. Measured behaviors for the new 5 pcf foam sandwich panel tests. ....................... 187

Figure 240. Comparison of the new and original foam sandwich panel tests. .......................... 187

Figure 241. Measured and calculated behaviors for the new 18 pcf foam 

sandwich panel tests (Mat_75)................................................................................ 188

Figure 242. Measured and calculated behaviors for the new 5 pcf foam 

sandwich panel tests (Mat_75)................................................................................ 188

Figure 243. Pretest configuration of the test head and ram car. ................................................ 189

Figure 244. Puncture of the external protective layers in Head Test 2...................................... 190

Figure 245. Internal puncture of the tank head in Head Test 2.................................................. 191

Figure 246. Comparison of predicted and measured behavior for Head Test 2. ....................... 192

Figure 247. Comparison of predicted and measured behavior for Head Test 2. ....................... 193

Figure 248. Comparison of calculated and measured behavior for Head Test 2 (Mat_75)....... 194

Figure 249. Comparison of calculated and measured behavior for Head Test 2 (Mat_75)....... 194

Figure 250. Comparison of calculated head impact response with different foam models....... 195

Figure 251. Evaluation of a retrofit EO tank car with a 0.375-inch jacket and foam................ 196

Figure 252. Impact response of a foam retrofit tank with the large ram (12” by 12”). ............. 198

Figure 253. Evaluation of the effects of the ram size on a 500 lb chlorine tank. ...................... 199

Figure 254. Impact behavior of a multi-layer foam head protection system (R2S). ................. 200

Figure 255. Impact behavior of multi-layer foam head protection systems. ............................. 201

Figure 256. Comparison of foam protection puncture forces to baseline systems. ................... 201

Figure 257. Nomenclature for corrugated EMS structures........................................................ 203

Figure 258. Detail of the corrugated EMS corrugation geometry. ............................................ 203

Figure 259. Drawing of the proposed corrugated EMS structure for Head Test 3.................... 204

Figure 260. Calculated response of the baseline corrugated EMS system (R3A)..................... 205

Figure 261. Calculated response of the corrugated EMS with an 8 inch standoff (R3D). ........ 206

Figure 262. Calculated response of the corrugated EMS and thick cover plate (R3E). ............ 206

Figure 263. Head impact analyses on corrugated EMS end structures...................................... 207

Figure 264. Calculated response of the corrugated EMS with a 12 inch standoff (R3F).......... 209

Figure 265. Performance of corrugated EMS end structures with various standoffs. ............... 210

Figure 266. Performance of corrugated EMS end structures with various standoffs. ............... 210
Applied Research Associates, Inc.

2010

xiii

Figure 267. Predicted Head Test 3 impact response at 15 MPH. .............................................. 211

Figure 268. Predicted Head Test 3 impact response at 18 MPH. .............................................. 211

Figure 269. Performance of corrugated EMS end structures with various standoffs. ............... 212

Figure 270. Pretest configuration of the test head and corrugated EMS system. ...................... 213

Figure 271. Pretest configuration of the test head and corrugated EMS system. ...................... 214

Figure 272. Impact damage to the external EMS protection system in Head Test 3................. 214

Figure 273. Internal puncture of the tank head in Head Test 3.................................................. 215

Figure 274. Comparison of predicted and measured behavior for Head Test 3. ....................... 216

Figure 275. Separation of the outer face sheet on the end EMS structure................................. 217

Figure 276. Buckling of the end EMS corrugation in Head Test 3. .......................................... 217

Figure 277. Fractured EMS end cover plate in Head Test 3...................................................... 218

Figure 278. Comparison of Post-test analysis and measured behavior for Head Test 3. .......... 218

Figure 279. Model for the 6-inch corrugated EMS side impact analysis. ................................. 220

Figure 280. Calculated 6-inch EMS puncture behavior............................................................. 221

Figure 281. Detail of the modified EMS corrugation geometry................................................ 222

Figure 282. Model for the 3-inch corrugated EMS side impact analysis. ................................. 222

Figure 283. Calculated 3-inch EMS puncture behavior............................................................. 224

Figure 284. Comparison of the force-deflection behaviors for the 

corrugated EMS concepts........................................................................................ 225

Figure 285. Extrapolation of the corrugated EMS concept. ...................................................... 226

Figure 286. Comparison of the force-deflection behaviors for the 

corrugated EMS concepts........................................................................................ 226

Figure 287. Side impact model for a 4-inch egg crate EMS system.......................................... 228

Figure 288. Calculated impact forces for a 4-inch egg crate EMS system................................ 228

Figure 289. Calculated side impact damage for a 4-inch egg crate EMS system...................... 229

Figure 290. Puncture force versus face sheet thickness for the egg crate EMS. ....................... 230

Figure 291. EMS puncture forces as a function of outer layer thickness. ................................. 231

Figure 292. Model for the double head puncture protection concept........................................ 232

Figure 293. Summary of the double head concept puncture forces........................................... 233

Figure 294. Summary of the double head concept puncture energies. ...................................... 234

Figure 295. Double head models with different standoff distances 

(2, 6, 12, and 18 inches). ......................................................................................... 234

Figure 296. Effect of the standoff distance on the double head impact response...................... 235
Applied Research Associates, Inc.

2010

xiv

Figure 297. Summary of the standoff effects on the double head impact response. ................. 236

Figure 298. Calculated force-deflection curve for a single 0.8-inch-thick head. ...................... 236

Figure 299. Effect of the head thickness on the double head impact response 

(12-inch standoff). ................................................................................................... 237

Figure 300. Effect of the head thickness on the double head impact response 

(2-inch standoff). ..................................................................................................... 238

Figure 301. Effect of the pressure ratio on the double head impact response 

(12-inch standoff). ................................................................................................... 239

Figure 302. Effect of the pressure ratio on the double head impact response 

(2-inch standoff). ..................................................................................................... 240

Figure 303. Experimental set-up for the quasi-static punch-shear experiment [86].................. 242

Figure 304. Effect of the span ratio on the composite punch force [87]. .................................. 243

Figure 305. Effect of the span ratio on the composite punch test damage mode [87]............... 244

Figure 306. Effect of the specimen thickness on the composite punch force (SPR = 8) [86]... 245

Figure 307. Comparison of the composite damage model with punch test data [86]................ 245

Figure 308. LS-DYNA model of an 11-layer composite punch experiment............................. 246

Figure 309. Comparison of initial punch simulations using ls971s_R2_7600. ......................... 247

Figure 310. Single element response under through-the-thickness compression-shear 

for various SMP versions of LS-DYNA. ................................................................ 248

Figure 311. Comparison of 4-layer punch simulations using ls970 and ls971.......................... 249

Figure 312. Comparison of 11-layer punch simulations using ls970 and ls971........................ 250

Figure 313. Side impact model with a composite jacket. .......................................................... 252

Figure 314. Composite tank model and impact zone mesh used for side impact 

puncture analyses. ................................................................................................... 253

Figure 315. Punch-shear mode in composite jacket at 18 inches of displacement.................... 254

Figure 316. Side impact force-deflection curves for a 500 lb tank car 

with a composite jacket. .......................................................................................... 255

Figure 317. Composite jacket blunting effect at 50 ms (17-inch displacement). ...................... 255

Figure 318. Effective strains in the 500 lb commodity tank at 50 ms 

(17-inch displacement). ........................................................................................... 256

Figure 319. Normalized peak punch test loads versus aerial density for various materials. ..... 257

Figure 320. Calculated puncture forces as a function of system thickness. .............................. 260

Figure 321. Calculated puncture energies as a function of system thickness............................ 261

Figure 322. Relationship between impact speed and kinetic energy for the ram car. ............... 262
Applied Research Associates, Inc.

2010

xv

Figure 323. Puncture prevention thickness requirements for various impact speeds. ............... 262

Figure 324. EMS puncture forces as a function of outer layer thickness. ................................. 264

Figure 325. Summary of side impact puncture performance and goals. ................................... 265

Figure 326. Summary of head impact puncture performance and goals. .................................. 266
Applied Research Associates, Inc.

2010

xvi

List of Tables

Table Page

Table 1. Weight Breakdown of Tank Car Components [11]......................................................... 6

Table 2. Tabular TC128B stress-strain curve values................................................................... 34

Table 3. Parameters used for the tank impact analyses ............................................................... 74

Table 4. Summary of impact analyses to assess outage volume effects.................................... 102

Table 5. Summary baseline side impact analyses...................................................................... 108

Table 6. Summary of unpressurized head impact analyses ....................................................... 120

Table 7. Summary of pressurized head impact analyses. (1)....................................................... 123

Table 8. Summary of pressurized EO head impact analyses..................................................... 136

Table 9. Summary head impact analyses with various impactor sizes. (1) ................................. 139

Table 10. Multilayer foam protection system analyses ............................................................. 199

Table 11. Summary corrugated EMS head impact analyses...................................................... 208

Table 12. Summary of the double head impact analyses........................................................... 232

Table 13. Constitutive model parameters for woven S-2 Glass/SC15 resin composite ............ 246
Applied Research Associates, Inc.

2010

xvii

Acknowledgements

This report is based in part on work funded by The Dow Chemical Company, on behalf of the 

Next Generation Rail Tank Car (NGRTC) Project. The development of the puncture modeling 

methodologies and the testing used to validate the models and develop the constitutive and 

failure models were all performed under the NGRTC project. The analyses of the Ethylene 

Oxide (EO) tank cars were funded by the American Chemistry Council, Inc., on behalf of its 

Ethylene Oxide Panel. 
Applied Research Associates, Inc.

2010

xviii

Executive Summary

Over the last few years, there has been significant attention on the potential for release of 

hazardous materials from railroad tank cars. This attention is primarily a result of a series of 

three accidents or derailments between 2002 and 2005 involving the release of hazardous 

material. In response, a research program was initiated to develop strategies for improving 

railroad tank cars so they can maintain tank integrity for more severe accident conditions than 

current equipment. The research was initiated by The Dow Chemical Company (Dow), Union 

Pacific Railroad, and Union Tank Car Company, working under a Memorandum of Cooperation 

with the Federal Railroad Administration, and Transport Canada, and the U.S. Transportation 

Security Administration. 

The Next Generation Rail Tank Car (NGRTC) Project was organized to include a Core Team 

(consisting of representatives from the five signatories to the Memorandum of Cooperation) and 

a group of Lead Contractors. The Core Team and Lead Contractors worked together to: 1) 

evaluate and select candidate materials, components, subsystems and systems with the potential 

to provide large performance improvements in the safety and security of rail tank cars; 2) select 

conceptual tank car designs incorporating appropriate materials, components and systems for 

improved safety and security; and 3) develop and use appropriate models, analytical techniques 

and testing protocols to demonstrate the efficacy of the tank car concepts. The goal of the 

NGRTC program was to develop a conceptual tank car that had a five to ten times improvement 

in the impact energy required to puncture the tank car. 

A key effort in this program is the development and application of detailed finite element models 

of tank car equipment which can accurately predict the puncture resistance under different 

impact conditions. These analysis tools were developed and validated for the puncture of the 

baseline tank cars for both side and head impact conditions (described in Sections 4 and 5 of this 

report). The models were subsequently applied to assess the puncture resistance of various tank 

car designs. 

This report describes the results of the NGRTC project to develop strategies for improving 

railroad tank cars so they can maintain tank integrity for more severe accident conditions than 

current equipment. Primary emphasis is on the finite element analyses of the impact and 

puncture behaviors of both current and future tank car designs. The scope of this effort includes 

the development of detailed finite element models for tank cars and the use of those models for 

various impact scenarios to assess puncture energies. A summary of the testing performed under 

the NGRTC program is also provided in this report, since it was critical for the development and 

validation of the puncture modeling capability. 
Applied Research Associates, Inc.

2010

xix

Baseline Tank Car Impact Model

The first task in this research was to develop and validate a modeling capability that can be used 

to analyze the impact response of a tank car. The model developed for the 105J500 chlorine tank 

car is shown in Figure 1. The model includes all of the primary tank car structures including the 

jacket and jacket standoffs, commodity tank, manway, bolsters, stub sills, and the addition of the 

outriggers attached to the draft gear to prevent a post-test rollover of the target tank caused by 

the rebound off the reaction wall in the impact testing. 

Figure 1. Updated model of a 105J500W pressure tank car.

Another feature of the tank impact model was the addition of an explicit model of the lading. 

The lading model consists of a low strength viscoelastic material that fills the same volume as 

the slurry lading in the test tank cars. The sloshing of the lading model can be seen in the 

cutaway view of the predicted impact response shown in Figure 2. This lading modeling 

approach was established to capture the momentum transfer of the coupled fluid-structure 

response but minimize effects such as sloshing at the fluid free surface that can cause numerical 

stability problems. 
Applied Research Associates, Inc.

2010

xx

Figure 2. Calculated Test 1 impact response with cutaway showing lading.

Full-scale impact tests were performed on tank cars and the results were used to validate the 

models. The tests were fully instrumented with accelerometers, string potentiometers, pressure 

gauges, and strain gauges. These measurements were compared to the model predictions to 

validate the model. Overall, very good agreement was obtained between the prediction and test 

for the various measurements made. An example of the agreement is the comparison of the 

predicted and measured force-deflection curves, shown in Figure 3. This force-deflection curve 

is an important characteristic of the tank car for a given impact scenario. The area under the 

force-deflection curve is the impact energy that is dissipated (the primary measure used to assess 

the puncture protection levels). The comparison of the force-deflection behaviors shows good 

agreement between the calculation and the test. 
Applied Research Associates, Inc.

2010

xxi

Figure 3. Comparison of the measured and predicted Test 1 force-deflection curves.

Material Damage and Failure Behaviors

The other necessary component of a predictive tank car puncture modeling capability is a 

detailed model that can be used to determine the impact damage and failure of the tank and 

protective system materials. An extensive program of laboratory materials testing was 

performed to characterize the tank car materials of interest. The tests included various material 

characterization tests, such as notched tensile tests and combined tension/shear tests, used to 

calibrate the material constitutive and failure behavior. Strain rate effects on the tank car 

materials were investigated and found to not have a significant effect on the tank puncture 

behavior. Additional component tests, such as punch tests and bend tests, were performed to 

validate the constitutive models. 

The material damage and failure model applied is the Bao-Wierzbicki (BW) model that defines 

the material damage development based on the current stress state in the material and the plastic 

strain increments. The critical strain function is that proposed in the BW criterion and contains 

multiple branches depending on the range of stress state, as shown in Figure 4. The critical 

strain in each branch are governed by the equation
Applied Research Associates, Inc.

2010

xxii

( )

( )

( ) ( )

[( )] ( )

( ) ( ) 





















≤

− + ≤ ≤

− ≤ ≤

+

∞ ≤ −

=

mean eq

mean eq

mean eq mean eq

mean eq

mean eq

mean eq

c mean eq

B

B A A

A

σ σ

σ σ

σ σ σ σ

σ σ

σ σ

σ σ

ε σ σ

3

1

3

3

1 9( ) 0

0

3

1

1 3

3

1

2

And the parameters A and B can be determined by a series of tests under different stress 

conditions including notched tensile tests, with specimens of varying notch radii, and tensileshear tests with different ratios of tension to shear stress. 

Figure 4. Bao-Wierzbicki failure surface and tests used for model calibration.

Although these material tests were used to develop the constitutive and failure model parameters, 

the resulting constitutive models were used to simulate the tests. This provides a validation that 

the material parameters were properly incorporated into the analyses and that the constitutive and 

failure model is capable of simulating the range of material behaviors under different loading 

conditions. The first of these material test series was the notched round bar (NRB) test (shown 

here on TC128B). A comparison of the calculated and measured stress-strain behavior across 

the notch for the three different radii specimens is shown in Figure 5. The comparison shows 
Applied Research Associates, Inc.

2010

xxiii

that the constitutive and damage model were capable of reproducing both the increase in stress 

level and reduction in ductility that occur as the notch radius is reduced. The BW failure 

parameters used provide a good correlation to the observed failures of the specimens. 

Figure 5. Validation of the notched round bar test behavior for TC128B.

A similar comparison of the calculated and measured load-displacement behaviors for the 

combined tension/shear tests are shown in Figure 6. The comparison shows again that the 

constitutive and damage model were capable of reproducing both the decrease in load level and 

increase in displacement that occur as the orientation was rotated from pure tension to pure 

shear. The BW failure parameters used provide a good correlation to the observed failures of the 

specimens. 
Applied Research Associates, Inc.

2010

xxiv

Figure 6. Validation of the combined tensile shear test behavior for TC128B.

For the various steels evaluated in this program, the component punch test was the primary 

laboratory material test used to validate the constitutive and failure models. An example of a 

model and simulation of a punch test on a 0.488-inch-thick TC128B plate is shown in Figure 7. 

The corresponding comparison of measured and calculated punch force-displacement curves for 

a series of three different tests on the TC128B plate is provided in Figure 8. In addition to the 

force-deflection curve, the final profile of the plate specimens after the punch tests were digitized 

and compared to the analyses. Again, the overall agreement between the testing and analyses 

was good. This agreement provides a further validation that the BW failure model is appropriate 

for predicting puncture of the tank cars. 
Applied Research Associates, Inc.

2010

xxv

Figure 7. Simulation of the punch test on the thin TC128B plate material.

Figure 8. Force-deflection curves for three punch test configurations on TC128B.

Tank Car Puncture Modeling

The BW failure modeling capability was combined with the tank car model to complete the tank 

car puncture prediction capability. The tank model was also simplified in these analyses to have 

half-symmetry by removing the manway, bolsters and lading. The simulation of the full scale 

tank car impact test (Test 2 - above the puncture threshold) was performed using the tank 

puncture model, as shown in Figure 9. The model shown was reflected vertically about the 

symmetry plane (seen as a line in the figure) for improved visualization of the impact behavior. 

The impactor in this analysis was a rigid 6x6 inch ram with a 0.50-inch radius around the edges 
Applied Research Associates, Inc.

2010

xxvi

and a total weight of 286,000 lbs. The small rectangular patch of elements under the impactor 

(already punctured in Figure 9) is the fracture zone where the BW failure model was applied. 

The remainder of the tank structure was again modeled with 4-node shell elements and a tied 

shell-to-solid constraint was used at the interface of the two model regions. 

The comparison of the measured and calculated force-deflection behavior for Test 2 with the 

tank puncture model is provided in Figure 10. The comparison shows overall good agreement 

between the calculation and test. The peak load at which the tank was punctured was very 

accurately captured by the model. The measured puncture force was 940,000 lbs and the 

calculated puncture force was 915,000 lbs. The primary discrepancy of the test and model was a 

slightly more compliant behavior in the model seen at the large displacements. This difference 

in compliance could primarily be attributed to the removal of the manway from the tank model in 

this analysis. As a result of the larger displacements in the analysis, the calculated puncture 

energy of 1.26 million ft-lbs is higher than the measured puncture energy of 0.87 million ft-lbs. 

Figure 9. Simplified tank model analysis with Bao-Wierzbicki failure assessment.
Applied Research Associates, Inc.

2010

xxvii

Figure 10. Tank puncture model Test 2 analysis with Bao-Wierzbicki failure assessment.

This combined tank car impact and puncture modeling capability was applied to evaluate a wide 

range of tank/jacket and head/head shield geometries. The side impact condition was a normal 

impact centered on the belt line of the tank. The head impact condition was an offset impact 

point approximately 29 inches vertically downward from the center of the head. 

An example head impact and puncture analysis is shown in Figure 11. The head impact analyses 

included the head, head shield, and a sufficient length of the side shell and jacket to allow for 

buckling to initiate in the jacket support from the loads transmitted by the head shield, as seen in 

Figure 11. 
Applied Research Associates, Inc.

2010

xxviii

(a) Initial impact conditions (b) Puncture response of the head

Figure 11. Calculated puncture behavior of a head and head shield.

The calculated head and shell puncture forces as a function of the combined head and head shield 

thickness are shown in Figure 12. All of the analyses included in the figure are performed with 

the 6x6 inch impactor. The figure shows the analyses are mostly consistent with a linear 

relationship between puncture force and total thickness of the protective layers. 

The linear relationship between the puncture force and total tank system thickness provides an 

indication of the primary failure mechanism initiating the tank puncture. The geometry of the 

ram impacting and indenting a pressurized tank shell is shown in Figure 13(a). A force balance 

analysis in the direction of the impact on a patch of tank shell material is shown in Figure 13(b). 

The forces resisting the impact loads are the pressure on the inside surface of the contact patch 

and the shear stress around the perimeter of the contact patch. For a 100 psi tank pressure and a 

6x6 inch impactor, the pressure load is less than 4 kips on the contact patch. Thus, the average 

shear stress is approximately equal to the impact force divided by the product of the impactor 

face perimeter and tank thickness. 
Applied Research Associates, Inc.

2010

xxix

Figure 12. Calculated puncture forces as a function of system thickness.

The slope of the linear fit in Figure 12 corresponds to an average shear stress in the tank layers 

around the perimeter of the impactor of 39 ksi. By comparison, the yield and ultimate stress 

levels of the TC128B in pure shear are 33 ksi and 49 ksi, respectively (approximately 58% of the 

stress values in pure tension using a Von Mises yield criterion). Thus, the failure mode is 

primarily exceeding the shear capacity around the perimeter of the impact patch. 

The calculated puncture forces for pressurized heads and the thicker head systems tended to fall 

slightly below the linear fit in Figure 12. The proposed mechanism for these lower forces is that, 

for the stiffer head systems, the offset impact creates a larger stress concentration along the upper 

edge of the impactor face and the failure initiates at that location at a lower total force. The more 

compliant head systems allow for a larger dent to form and the impactor develops a more 

uniform stress distribution in the impact patch around the ram face perimeter. 
Applied Research Associates, Inc.

2010

xxx

(a) Geometry of the tank indentation

(b) Free body diagram for the tank contact patch

Figure 13. Loading and failure mechanism for the tank impact and puncture.

The calculated puncture energies for all of the side and head impact analyses are plotted against 

the combined thickness of the system in Figure 14 (6x6 inch impactor only). When comparing 

all of the puncture energies the data falls into four separate groups that are distinguished by the 

impact type (side or head) and the tank pressure. Again, the comparison indicates that the total 

thickness is the relative parameter that determines puncture energy for a given impact condition 

and pressure level, indicating that a retrofit design with an increased jacket thickness should 

provide equivalent protection to a thicker commodity tank systems with equivalent combined 

shell and jacket thicknesses. 
Applied Research Associates, Inc.

2010

xxxi

Figure 14. Calculated puncture energies as a function of system thickness.

The above correlation of the system thickness and puncture energy can be converted to assess 

system requirements for a given impact scenario. The relationship between the impact speed and 

impact energy for the 295,000 lb ram car is provided in Figure 15. This impact energy is 

proportional to the mass of the object so a 286,000 ram car would have energies that are three 

percent lower than those in Figure 15. This relationship between speed and impact energy can 

be combined with the linear fits between system thickness and puncture energy (shown in Figure 

14) to determine the required system thickness to resist puncture for a specified impact speed. 

Examples of the pressurized (100 psi) Cl tank car thickness requirements to resist a puncture of 

the 6x6 inch impactor at various impact speeds is shown in Figure 16 for both side and head 

impacts. The figure shows that as the impact speed increases, the system thickness required 

begins to increase rapidly. A 25 mph side impact would require approximately three inches of 

steel to prevent puncture and a 30 mph head impact would require more than six inches of steel. 

Obviously, these protection levels are not achievable with a traditional tank car design approach 

while maintaining a tank car that is economically viable as a result of both the initial tank car 

cost and the drastically reduced lading capacity. 
Applied Research Associates, Inc.

2010

xxxii

Figure 15. Relationship between impact speed and kinetic energy for the ram car.

Figure 16. Puncture prevention thickness requirements for various impact speeds.
Applied Research Associates, Inc.

2010

xxxiii

The analyses demonstrate that the 25 and 30 mph impact protection levels are not feasible for the 

6x6 inch impactor. However, the analyses demonstrate that they may be achievable with a larger 

ram size. The puncture energies from analyses of side impacts on 500 lb and 600 lb tanks using 

6-inch, 9-inch, and 12-inch square impactors are shown in Figure 17. The figure shows that the

25 mph ram car impact (6 million ft-lb impact energy from Figure 15) is approximately 

equivalent to the puncture energy of the 600 lb tank car impacted by the 12x12 inch ram. 

Figure 17. Summary of puncture energies for chlorine tank cars for various impactor sizes.

Retrofit Tank Cars

A topic of interest for increasing the puncture resistance of tank cars is the relative efficiency of 

adding material to the jacket as opposed to adding material to the commodity tank. This is 

critical for assessing the capability to retrofit existing tank cars to meet equivalent puncture 

resistance levels of a thicker tank car with a higher tank test pressure. 

The assessment of retrofit options was performed using the shell and head puncture analyses. A 

summary of the analyses of the 500 lb Cl tank with various jacket thicknesses is provided in

Figure 18. The graph provides a fit to the analyses to illustrate the relationship between the 

retrofit jacket thickness and the puncture energy. The comparison shows that a 0.375-inch thick 

retrofit jacket is required for the 500 lb Cl tank to provide a similar level of puncture resistance 
Applied Research Associates, Inc.

2010

xxxiv

as a 600 lb tank car with the 11 gauge jacket (as confirmed by the analysis with the 0.375-inch 

jacket). 

Figure 18. The effects of jacket thickness on puncture energy for a 500 lb Cl tank.

Similarly, a summary of the analyses of the 300 lb EO tank with various jacket thicknesses is 

provided in Figure 19. The comparison illustrates that an approximately 0.4375-inch thick 

retrofit jacket would be required for a 300 lb EO tank to provide a similar level of puncture 

resistance as the 500 lb EO car. However, the puncture energy levels calculated for the 500 lb 

and 600 lb Cl tank cars are also indicated in Figure 19. We see from the comparison that the 300 

lb EO tank car retrofit with a 0.375-inch jacket has a puncture energy that is higher than that of 

the 600 lb chlorine car as a result of the lower pressures in EO tanks. Thus, the 0.375-inch 

retrofit jacket in EO cars will provide a protection level equivalent to or better than that of 

retrofit designs of other commodities. 

A summary of the analyses of the pressurized 500 lb tank head with the various head shield 

thicknesses is provided in Figure 20. The puncture energy levels calculated for the 600 lb head 

with the baseline jacket and head shield are also indicated in the figure. The comparison shows

that an approximately 0.60-inch retrofit head shield thickness would be required for a 500 lb 

head to provide a similar level of puncture resistance as the 600 lb head with a 0.50-inch A516-

70 head shield. 
Applied Research Associates, Inc.

2010

xxxv

Figure 19. The effects of jacket thickness on puncture energy for a 300 lb EO tank.

Figure 20. The effects of head shield thickness on puncture energy for a 500 lb head.
Applied Research Associates, Inc.

2010

xxxvi

Analyses of Advanced Protection Systems

The modeling capabilities developed in this program were applied to analyze a series of 

advanced tank car protective structure concepts. Included in these were multi-layered foam 

systems and engineered metal structures (EMS) systems. Example head impact puncture 

analyses on both a layered foam system and a corrugated EMS system are shown in Figure 21. 

A variety of designs were investigated both with analyses and with a full scale head test on each 

of the concepts shown in Figure 21. However, neither of the foam or EMS systems was 

optimized.

 

(a) Layered foam protection system (b) Corrugated EMS protection system

Figure 21. Analyses of advanced protection concepts.

The failure mode of the advanced systems was similar to that of the baseline tank systems when 

impacted by the 6x6 inch impactor. When the load acting on a layer of the advanced protective 

system exceeded the shear capacity of the layer, the layer would be punctured. For many of the 

advanced concepts, the system design would increase the stiffness of the support for the outer 

layers. The increased support structure stiffness would result in outer layer penetration at 

reduced ram displacements. As a result, the protection system and commodity tank would be 

penetrated independently at lower peak forces and lower puncture energies than could be 

obtained in an optimized system. 
Applied Research Associates, Inc.

2010

xxxvii

The performance of advanced protective systems was typically improved when the thickness of 

the outer layer was increased to resist puncture and the stiffness of the supporting structure was 

reduced to increase the displacements of the system before the puncture level was exceeded. 

These effects are seen by considering the three protections concepts shown in Figure 22. The 

concepts include two different corrugated EMS designs and a single monolithic plate, all of 

which have approximately equivalent weight. The different corrugated EMS designs illustrate 

the approach of moving weight from the inner face sheet and core structure to lower the system 

stiffness and increasing the outer face sheet thickness to improve the penetration resistance. The 

maximum extrapolation of this concept is to move all of the EMS weight into a single thicker 

outer jacket.

Figure 22. Alternative tank protection concepts.

The calculated puncture behavior of the three protection options are compared in Figure 23. The 

ram punctures the baseline 6-inch corrugated EMS structure at a load of approximately 200,000 

lbs and a ram displacement of approximately 6 inches. Thus, the energy dissipation of this EMS 

structure is quite small. The ram punctures the modified 3-inch corrugated EMS structure at a 

load of approximately 600,000 lbs and a ram displacement of approximately 12 inches. Thus, 

the energy dissipation of this EMS structure is approximately 5 times that of the original 6-inch 

EMS concept. Alternatively, the 0.5625-inch-thick jacket is not penetrated until a force of 

approximately 1.1 million lbs and an impactor displacement of approximately 30 inches. Thus,

the jacket absorbs significantly more impact energy than either of the corrugated EMS structures 

for the 6-inch impactor. 
Applied Research Associates, Inc.

2010

xxxviii

Figure 23. Comparison of the force-deflection behaviors for the corrugated EMS concepts.

The calculated forces to puncture the outer layer of foam and EMS systems are added to the plot 

of the puncture force versus system thickness in Figure 24. The difference in the data for the 

advanced systems is that the thickness used is the layer thickness only (e.g. outer head shield or 

outer face sheet thickness) rather than the sum of the total tank and jacket layer thicknesses. The 

comparison shows that the foam and EMS structures had the effect of making each layer act 

independently rather than coming in contact and working together to resist the high shear stresses 

around the edge of the impactor. As a result, the maximum achievable puncture forces were 

reduced and the systems do not reach the desired puncture energies. 
Applied Research Associates, Inc.

2010

xxxix

Figure 24. EMS puncture forces as a function of outer layer thickness.

Although the systems with a monolithic jacket performed better than the corrugated EMS 

structures in this example, the conclusion can not be generalized to all EMS designs and impact 

scenarios. The EMS structures would be expected to perform significantly better under a more 

distributed loading. Other benefits of the EMS concepts are that the structural stiffness can be 

used to efficiently carry the train service loads in a “tank within a tank” design. Further analysis 

of other EMS concepts (e.g. core geometries, materials) and more general impact conditions 

would be required to obtain an optimized design. 

A similar set of observations was obtained from the analysis of layered foam systems. In 

general, the systems analyzed did not outperform a single monolithic plate of equivalent 

protection system weight for the 6x6 inch impactor. However, the performance for larger 

impactors and the optimum foam and interface properties to maximize performance were not 

fully explored. In addition, low density foams can contribute to improved thermal insulation. If 

a foam system can maintain a lower average shipping temperature (and thus a lower average tank 

pressure), the overall tank puncture energy could potentially be increased. 

Composite Material Protection Systems

The above analyses indicate that a tank protection system developed using only typical tank car 

materials (steels) will have difficulty reaching the five to ten times protection goals of the 
Applied Research Associates, Inc.

2010

xl

NGRTC program. To achieve a five times increase in the puncture energy of tank cars while 

maintaining a comparable level of utility will require non traditional tank car design approaches 

and advanced materials. These potential materials for advanced protection concepts include 

polymeric composite materials. 

Tank car concepts with composite protection systems were not evaluated as thoroughly as the 

layered foam and EMS concepts. However, some preliminary evaluations were performed to 

assess the potential of composites for puncture resistance. 

A summary of the normalized peak punch test loads for different materials plotted against aerial 

density is shown in Figure 25. Effectively, the normalized punch loads are the shear stress 

around the punch perimeter scaled by the material density. As a result, the punch shear 

resistance is compared on an equivalent weight basis rather than an equivalent thickness basis. 

The data in the figure were taken from the punch tests performed on steels in the NGRTC 

program and composite punch tests reported in open literature sources for a woven S-2 glass 

fiber SC-15 polymeric matrix composite system. 

Figure 25. Normalized peak punch test loads versus aerial density for various materials.

The comparisons in Figure 25 are not perfect due to some differences in the punch test protocols 

between steels and composites. However, the comparison indicates that there is potential for 

modest puncture resistance improvements using composite materials. At the largest aerial 

density testing for the composite, a direct comparison can be made with 11 gauge A1011 and 
Applied Research Associates, Inc.

2010

xli

HSLA 90XF with a similar aerial density. The composite out performs A1011 by approximately 

70% and HSLA 90x by roughly 10%. Further improvements in the performance of the 

composite system could likely be obtained by optimizing the composite material and/or layup for 

these types of impact conditions. However, we have not yet identified any composite systems 

that could achieve a puncture resistance that approach a two times improvement over a 

comparable weight steel system. 

Impact Performance Summary

A summary of the current state of tank car side impact performance is provided in Figure 26. 

From the full scale side impact Test 2, the 105J500W tank car puncture energy was slightly less 

than one million ft-lbs. This corresponds to a 10 mph impact with a 286,000 lb ram car and the 

6x6 inch impactor. From analyses, the puncture energies of the 105J300W and 105J600W tank 

cars are approximately 30 percent less and 40 percent greater than the 105J500W tank car, 

respectively. 

The goal of the Next Generation Railroad Tank Car Project was to increase the puncture energy 

of the tank car by five to ten times over the baseline 105J500W tank car design. This target is 

outlined by the red box in Figure 26. The proposed TIH tank car standard falls within this range,

requiring a puncture energy that is 6 times that of the baseline tank car. However, the levels that 

were found to be achievable with the first generation concepts in the NGRTC project are more

consistent with two to three times the puncture energy of the baseline design. 
Applied Research Associates, Inc.

2010

xlii

Figure 26. Summary of side impact puncture performance and goals.

A similar summary of the current state of tank car head impact performance is provided in Figure 

27. The calculated puncture energy for a pressurized (100 psi) 105J500W tank head with an 11-

gauge head jacket is 610,000 ft-lbs (analysis described in Section 5.3.2). This corresponds to an 

8 mph impact with the 286,000 lb ram car and the 6x6 inch impactor. For comparison, the 

puncture energy of a 105J600W tank head with a 0.50 inch full height head shield is 1.1 million 

ft-lbs. 

The goal of the Next Generation Railroad Tank Car Project was to increase the puncture energy 

of the tank car by five to ten times over the baseline 105J500W tank head as outlined by the red 

box in Figure 27. The proposed TIH tank car standard is above this range, requiring a puncture 

energy that is approximately 14 times that of the baseline tank head. However, the levels that 

were found to be achievable with the first generation concepts in the NGRTC project are more 

consistent with 2.5 to 4 times the puncture energy of the baseline design. 
Applied Research Associates, Inc.

2010

xliii

Figure 27. Summary of head impact puncture performance and goals.
Applied Research Associates, Inc.

2010

1

Detailed Puncture Analyses of 

Various Tank Car Designs

1 Introduction

There is ongoing research to develop strategies for improving railroad tank cars so they can 

maintain tank integrity for more severe accident conditions than current equipment. Research 

results are being used to develop improved tank car designs. In addition, this research was 

performed to support the rail freight operations to prepare for the proposed rulemaking by the 

Federal Railroad Administration (FRA). 

A key effort in this program is the development and application of detailed finite element models

of tank car equipment which can accurately predict the puncture resistance under different 

impact conditions. To date, these analysis tools have been developed and validated for the 

puncture of the baseline tank cars for both side and head impact conditions (described in 

Sections 4 and 5 of this report). These validated tools were applied to assess the puncture 

resistance of various tank car designs. 

This report describes finite element analyses of the impact and puncture behaviors of both 

current and future tank car designs. The scope of this effort includes the development of detailed 

finite element models for tank cars and the use of those models for various impact scenarios to 

assess puncture energies. 

1.1 Problem Definition and Approach

Accident statistics show that the rail industry’s safety performance has generally improved over 

the last few decades. The FRA’s Railroad Accident and Incident Reporting System (RAIRS) 

show that the number of accidents per year with at least one car releasing hazardous materials 

has decreased significantly over the past 25 years, as shown in Figure 28 [1]. However, a series 

of three recent accidents or derailments involving the release of hazardous material have focused 

attention on the structural integrity of railroad tank cars. These events include (1) Minot, ND, on 

January 18, 2002; (2) Macdona, TX, on June 28, 2004; and (3) Graniteville, SC, on January 6, 

2005 [2-4]. 

To better define the collision threat, studies have been performed to both evaluate the accident 

statistics [e.g. 5-7] and analyze the kinematics of freight trains in derailments and collisions [e.g. 

8, 9]. Evaluation of the derailments and collisions has shown that these are complex events with 

a wide range of collisions between the various cars in the train. A 40 mph derailment of a large 

freight train may involve thirty or more cars and the derailment event would last on the order of a 

minute before the train comes completely to rest. An example of a derailment simulation for a 

36-car train model is shown in Figure 29. Impacts on tank cars will include both head impacts 
Applied Research Associates, Inc.

2010

2

and side impacts from objects as small as a broken rail to very blunt objects, such as another tank 

head. Thus, the objective of the tank car development efforts included increased protection in 

both head and side impacts for a range of impact conditions. 

Figure 28. Number of Accidents with at Least One Car Releasing Hazardous Materials [1]. 

To develop an improved tank car design, the physics of the tank impact response need to be 

understood. For the purposes of the analyses performed here, the impactor is assumed to be a 

rigid object. As the impactor strikes the tank, the tank starts to deform. As it deforms, it will 

develop resisting forces. For a given tank configuration (e.g. tank thickness, internal pressure 

level) and a given impact condition (e.g. head or side impact, centered or offset impact location) 

the tank will have a characteristic force-deflection curve, such as that shown in Figure 30. The 

shape of this force-deflection curve is relatively independent of the size of the impactor. The 

area under the force-deflection curve is equal to the amount of impact energy that has been 

dissipated. If the impact speed is sufficiently low, the total energy will be dissipated and the 

impactor will be safely stopped. If the impact speed is sufficiently high, the forces developed 

will exceed the strength of the tank material and the tank will be punctured. The point at which 

this failure of the tank is initiated will depend strongly on the size of the impactor. 
Applied Research Associates, Inc.

2010

3

(a) 36-car train derailment model

(b) Calculated response 25 seconds after derailment

Figure 29. Calculated derailment behavior for the 36-car train model [9]. 

To absorb additional energy prior to puncture, you can increase the force levels required to 

deform the tank, increase the displacements that can be experienced prior to puncture initiation, 

or a combination of both larger forces and larger displacements. The protection strategies are to 

add external energy absorbing structures and protective layers. These structures may have a 

larger standoff distance to extend the total displacement and energy dissipation prior tank 

rupture. Other protection strategies may be to (1) reinforce the tank system (increasing forces), 

or (2) blunt the impact loads (increasing the allowable tank displacements prior to rupture). 
Applied Research Associates, Inc.

2010

4

Figure 30. Characteristic force-deflection curve for a tank impact. 

1.2 Tank Car Design Constraints

A good summary of the overall tank car design constraints is provided by Tyrell, et al. in 

Reference 11. A summary of the primary constraints is provided here. 

Any new developments for tank car designs will still need to satisfy the basic current tank car 

design requirements for all conventional aspects (service, manufacturing, inspection, test, and 

maintenance). The baseline car chosen for this study is a DOT 105J500W car, the same class of 

tank car used in the full-scale crash tests. The tank cars used in the NGRTC test program were 

chlorine tank cars. This baseline car is constructed to Association of American Railroads (AAR) 

plate B clearance standards [10]. The envelope for equipment built to plate B standards is shown 

in Figure 31. The clearance standard dictates the maximum car height, width, and truck center 

spacing for compliant cars. For plate B, the maximum allowable height of any portion of the car 

is 15 feet-1 inch above top-of-rail. At the maximum allowable width of 10 feet-8 inches, truck 

centers cannot exceed 41 feet-3 inches. 

For a car with the maximum distance between truck centers, the length of structure outboard of 

the trucks must also be considered in the car’s design. The swingout at the ends of the car cannot 

exceed the swingout at the center of the car on a 13 degree curve [10]. A car built to plate B 

with maximum distance between truck centers has a swingout at the center of the car of 5.75
Applied Research Associates, Inc.

2010

5

inches. By extending the length of the car outboard of the trucks until this same swingout is 

achieved at the end of the car, a maximum overall car length of 57 feet-10 inches can be 

achieved [11]. 

The baseline car’s tank has an outer diameter of 8 feet-6.3 inches and a nominal capacity of 

17,300 gallons. This tank has a 12.875-inch clearance on either side before encountering the 

maximum allowable car width. This clearance effectively limits the maximum amount of energy 

absorbing material that can be placed around the tank. 

Figure 31. Maximum Dimensions for Plate B [11]. 

One concern associated with the addition of new structure to the car is the additional weight that 

this will contribute. For the improved crashworthiness tank car, the maximum allowable weight 

is 286,000 lb. The new design includes many of the same features of a conventional tank car, 

including the tank assembly, truck assemblies, safety appliances, and an equal lading capacity of 

90 tons of chlorine. In addition, the new design will include a strengthened tank, energy 

absorbing material, and a structural frame to bear the service loads. 

Table 1 shows the weights of components common to both conventional and improved 

crashworthiness tank cars. These preliminary values are given to estimate the amount of weight 

that is available for new structure. Based on the weights of existing tank car components and the 

286,000 lb limit, approximately 39,000 lb are available for new structure (after removal of the 

existing tank jacket and head shields). This available weight comes from increasing the car 
Applied Research Associates, Inc.

2010

6

weight and from incorporating the functions of the baseline car’s jacket and the stub sill/body 

assemblies into alternative components. 

Table 1. Weight Breakdown of Tank Car Components [11].

Component Weight (lbf)

Bare tank (0.777-inch thick) 38,280

Truck Assemblies (2) 20,580

Coupler/Draft gear Assemblies (2) 2,300

Air Brake Assembly 1,310

Fittings and Housing Assembly 1,300

Manway 1,300

Ladders, Handholds, & Railings 1,100

Handbrake Assembly 420

Paint and Markings 100

Chlorine 180,000

Subtotal 246,690

Available for New Structure 39,310
Applied Research Associates, Inc.

2010

7

2 Model Development and Preliminary Impact Analyses

2.1 Introduction

In this section, we provide a brief history of the tank car model development, preliminary impact 

analyses, and model validation for current tank car designs. 

2.2 Tank Car Model Development

A diagram of a general purpose tank car is shown in Figure 32 showing the terminology applied 

in this report [5]. Missing from the figure is any description of an external jacket or head shield 

that were included in the bulk of the impact analyses.

Figure 32. Diagram of a typical North American general purpose tank car [5].

The tank car model used in these analyses was based on a previous finite element model of a 

DOT 112A340W pressure tank car generated to analyze the loads and stresses in the tank car as 

a result of service and salvage operating conditions [12-14]. The LS-DYNA finite element code 

was used for the tank car analyses described in this report [15]. LS-DYNA, developed and 

maintained by the Livermore Software Technology Corporation (LSTC), is an explicit nonlinear 

three-dimensional finite element code for analyzing the large deformation dynamic response of 

solids and structures. LS-DYNA has many features ideally suited to the analysis of vehicle 

collisions and is well accepted in the crashworthiness community as a state-of-the-art code. 

Equipment manufacturers commonly use LS-DYNA in the design of passenger rail car and 

automotive crash safety. The model for the tank car was developed using the TrueGrid 

preprocessor code [16]. 

Two external views of the original DOT 112A340W tank car model are shown in Figure 33. The 

structure includes the tank with ellipsoidal heads, manway, stub sill, bolster, and truck structures. 

The geometry of the tank car structures in the model were based on a design originally developed 
Applied Research Associates, Inc.

2010

8

by Union Tank Car. Figure 33(b) gives a detailed view of the original car end structures in the 

model and the associated model mesh resolution. 

(a) Tank car model (b) Detail of the tank car mesh

Figure 33. Original model generated for an 112A340W pressure tank car.

Subsequent development and application of the tank car model was performed in a study to 

assess the conditions for puncture and catastrophic rupture in the Minot, North Dakota

Derailment [17-18]. For these impact analyses, the model was modified to be a 105J300W 

primarily by adjusting the tank wall thickness to match the values measured and reported in the 

NTSB Materials Laboratory Factual Report [19]. For this application, the external insulation and 

jacket were not included in the model. The stub sill geometry was also modeled in greater detail 

to allow for the inclusion of a model for the coupler and draft gear. The final modification made 

to the model was to significantly increase the mesh resolution in the impact zone. The model 

was refined to have approximately 2-inch elements in the impact zone to resolve the damage 

distributions caused by impact. The updated tank car model is shown in Figure 34.

The starting point for the model used in this study was that of the tank car geometry shown in

Figure 34. However, additional data on specific tank car designs were used to update the models 

used in this study. Modifications include the addition of a jacket and standoffs and the 

modification of the tank dimensions (diameter, length, head and shell thicknesses) to match the 

specific parameters of the various tank designs of interest. 
Applied Research Associates, Inc.

2010

9

(a) Overall view (b) Truck and draft gear geometry

(c) Tank mesh

Figure 34. Updated model generated for a 105J300W pressure tank car.

2.3 Side Impact Analyses and Model Validation

This section describes the impact test results and corresponding analyses that were used to 

validate the models for the detailed impact dynamics. Comparisons to these tests were also used 

to evaluate modeling simplifications that can be applied in subsequent developments of 

improved tank car designs. All of the analyses performed here (and in the remainder of the 

report) were developed using an inch-lbf-s system of units and corresponding plots will be in 

units consistent with this system (e.g. stress in psi, etc). 
Applied Research Associates, Inc.

2010

10

2.3.1Test 1 - 14 mph Side Impact (17x23 impactor)

The first validation of the tank car model was performed for the impact Test 1 from the Next 

Generation Rail Tank Car project on April 26, 2007, at the Transportation Technology Center, 

Inc. (TTCI) in Pueblo, Colorado [20]. This test involved a side impact with a blunt and rigid 

indenter into the side of a type 105 chlorine tank car design. The test was performed by 

impacting a ram car into the side of a ballasted chlorine tank car that was backed by a rigid 

impact barrier. The desired impact speed was 14 mph and the ram car speed was measured 

optically at 13.9 mph using speed trap reflectors placed within 6 feet of the impact point.

Additional parameters for Test 1 were as follows:

• Tank Car Number 3069

• Tank Car Weight (estimated) 265,000 pounds

• Tank Shell: 0.777-inch-thick normalized TC128B

• Insulation and Thermal Protection: ceramic fiber and fiberglass

• Jacket: 11 gauge A1011 steel at 4-inch standoff distance

• Slurry Density 11.6 lb/gallon

• Outage 10.4 percent

• Capacity 17,391 gallons

• Internal Pressure 100 psi

• Ram Car Weight: 285,600 pounds

• Ram Car Speed: 13.9 miles per hour

• Ram Car Energy (derived): 1.86 million foot-pounds

This test and the subsequent Test 2 (described below) were designed to bound the regime 

between no puncture (but with large deformations of the tank) and puncture of the tank. The

instrumentation in the impact tests were designed to measure the necessary accelerations, strains, 

displacements, and pressure measurements to validate the tank car models. In many areas, 

sensors were placed to assure redundancy in the critical information. 

Figure 35 shows tank car 3069 and the ram car staged in position for the April 26 test. The ram 

was aligned to impact at mid height at the center of the tank. The rigid ram had approximate 

frontal dimensions of 23 by 17 inches, with 1-inch-radius edges and corners. The target tank car 

was supported on trucks resting on short stub sections of track normal to the direction of travel of 

the ram car. The target tank car was also backed by a rigid barrier to react to the impact loads.

Figure 36 shows the ram and target cars during the impact test for tank car 3069. As a result of 

the impact speed of 14 mph, the rigid ram dented both the jacket and the shell of the chlorine 

tank car but the tank was not ruptured. The tank car stayed upright throughout the test, but it did 

derail from the stub rail sections. Following the primary impact there was a significant rebound 

of the ram car from the tank car with a rebound speed of 7.6 mph. The tank car also experienced 
Applied Research Associates, Inc.

2010

11

a rebound from the wall after impact. The rebound of the car was countered by outriggers 

installed in the draft pockets of the car. Structural inspection after the impact showed no visible 

damage to the stub sills or their weld joints to the tank. The ram car brakes triggered and set as 

anticipated, which kept the ram car from rolling forward (downhill) for a second impact. 

Figure 35. Ram car and target tank car staged prior to Test 1.
Applied Research Associates, Inc.

2010

12

Figure 36. Shell side impact behavior in Test 1 for Tank Car 3069

The post-test condition of the tank included large dents in both the jacket and tank shell. Figure 

37 shows the post-test dents in the jacket and commodity tank. The post-test permanent dent in 

the jacket was approximately 14 inches deep and in the tank shell it was approximately 10 inches 

deep. The impact also caused one of the weld joints on the jacket near the impact area to fail. 

The back of the tank car was also deformed by the reaction against the rigid impact barrier. 
Applied Research Associates, Inc.

2010

13

 

Figure 37. Post-test dents in the jacket and shell of Tank Car 3069

Comparison of Test 1 with Pretest Predictions

The tank car model was further modified to match the tank car structures used in the full-scale

impact test program. The updated model of the 105J500 chlorine tank car is shown in Figure 38. 

The modifications include the addition of the jacket and jacket standoffs, correction of the 

commodity tank length, diameter, and thickness, and the addition of the outriggers attached to 

the draft gear to prevent a post-test rollover of the target tank caused by the rebound off the 

reaction wall. The mesh was also refined in the impact zone. 

Another modification was the addition of an explicit model of the lading. The lading model 

consists of a low strength viscoelastic material that fills the same volume as the slurry added to 

the test tank. The sloshing of the lading model can be seen in the cutaway view of the predicted 

impact response shown in Figure 39. This lading modeling approach was established to capture 

the momentum transfer of the coupled fluid-structure response but minimize effects such as 

sloshing at the fluid free surface that can cause numerical stability problems. 
Applied Research Associates, Inc.

2010

14

(a) Complete tank car model

(b) Tank mesh

(c) Jacket mesh

Figure 38. Updated model generated for a 105J500W pressure tank car.
Applied Research Associates, Inc.

2010

15

(a) Before impact (time=0.1 s)

(b) Near maximum displacement (time=0.25 s)

Figure 39. Calculated Test 1 impact response with cutaway showing lading.

A comparison of the predicted and measured ram car acceleration and displacement time 

histories are shown in Figure 40. The comparison shows that both the characteristics of the 

double peak acceleration response and the quantitative comparisons of maximum acceleration 

and displacements agree well between the calculation and the test. 
Applied Research Associates, Inc.

2010

16

(a) Acceleration history

(b) Integrated displacement history

Figure 40. Comparison of the measured and predicted ram car motions.
Applied Research Associates, Inc.

2010

17

A series of string potentiometers were included inside the tank to measure the tank distortions 

during the test. These measurements include a series of 5 string potentiometers spaced 

longitudinally on 18-inch centers across the spring line (horizontally between tank belt lines) and 

centered on the impact location. In addition, a vertical string potentiometer was placed on the 

center line of the tank. All of the string potentiometers successfully collected measurements of 

the tank distortions with the exception of the vertical gauge that failed early in the impact 

response. 

A comparison of the predicted and measured tank deformations are shown in Figure 41. The 

solid lines are the predictions and the dashed lines are the measurements. Overall, very good 

agreement was obtained between the prediction and test. The 35-inch maximum ram 

displacement, shown previously in Figure 40, corresponds to a 26-inch peak spring line 

displacement of the tank. The difference between these two values is primarily a result of the 

compaction of the space between the tank and jacket both on the impact face and on the far side 

of the tank car between the tank and the impact wall. 

Figure 41. Comparison of the measured and predicted Test 1 tank displacements.

The ram car accelerations and displacements, shown in Figure 40, can be combined to calculate 

the force-deflection behavior for the target tank car. The comparison of the predicted and 

measured force-deflection curves is shown in Figure 42. This force-deflection curve is an 

important characteristic of the tank car for a given impact scenario. Although it is a 
Applied Research Associates, Inc.

2010

18

characteristic of the tank car and impact condition, it is measured based on the motions of the 

ram car which acts primarily as a rigid body. The area under the force-deflection curve is the 

impact energy that is dissipated and the primary approach that will be applied to improve the 

puncture resistance of tank cars is to add protective structures that increase the area under the 

curve prior to tank puncture initiation. The comparison of the force-deflection behaviors shows 

good agreement between the calculation and the test. 

Figure 42. Comparison of the measured and predicted Test 1 force-deflection curves.

The above comparison of the pretest prediction and impact test serves as a validation of the 

model for the tank collision dynamics and impact deformation behavior. The responses include 

the initial impact, the reaction of the impact loads against the wall, and the post-impact rebound 

from the wall. Additional details of the impact behavior captured in the model included the 

target tank roll motion and lifting forces on the ram car caused by the lateral support of the target 

tank truck on the stub rail sections and the action of the outriggers resisting a post-test rollover 

behavior. 

Test 1 Internal Pressure

The internal pressure in the pretest analysis was calculated by using the tank as a control volume 

and imposing an ideal gas pressure-volume relationship for the gas in the outage region of the 

tank. As a result, the internal pressure at any given time will be uniform everywhere in the tank 

and determined by the instantaneous tank volume. In the test, the pressure was measured at 
Applied Research Associates, Inc.

2010

19

various locations submerged in the slurry lading. The positions included the front and back belt 

line positions as well as locations at the bottom of the tank. Three pressure gauges were 

distributed longitudinally at each of these circumferential locations (12 gauges total). 

A comparison of example measured pressure histories at three different circumferential locations 

is shown in Figure 43. The measurements indicated that there was both a hydrostatic pressure 

increase in the tank from the indentation as well as hydrodynamic pressures from the sloshing of 

the lading within the tank. 

Figure 43. Measured Test 1 tank pressures at various positions in the lading.

A comparison of the predicted hydrostatic internal pressure and the average of the measured 

pressures is shown in Figure 44. The average of the measured pressures is slightly higher than 

predicted. However this discrepancy is probably a result of the hydrodynamic effects that are 

included in the pressure measurements. Overall, comparison indicates that the Test 1 impact 

response and tank indentation will produce a hydrostatic pressure increase of between 10 and 20 

percent. This magnitude of pressure increase is not expected to have a large effect on the impact 

response and puncture energy. 
Applied Research Associates, Inc.

2010

20

Figure 44. Comparison of the measured and predicted Test 1 tank pressure.
Applied Research Associates, Inc.

2010

21

2.3.2 Test 2 - 15 mph Side Impact (6x6 Impactor)

Description of Test Conditions

The full-scale side impact Test 2 was performed on July 11, 2007 at the Transportation 

Technology Center [21]. The test was a side impact with a tapered and rigid indenter into the 

side of a type 105 chlorine tank car of modern design. The test was performed by sending a ram 

car into the side of a ballasted chlorine tank car that was backed by a rigid impact barrier. An 

impact speed of 15 mph (+0.5/-0 mph) was specified for this test. The actual impact speed was 

15.2 mph. The target tank car was modified for this test by lowering the tank position and 

mounting rigid support skids to the bolster as shown in Figure 45. 

Figure 45. Test 2 target Tank Car 3074 mounted on support skids. 

The ram car was also modified prior to Test 2 by lowering the position of the ram arm and 

adding a new ram head with a much smaller contact area, as shown in Figure 46. The ram arm 

was lowered (as was the target tank car) to more closely align the height of the centers of gravity 

(CG) of the target tank, ram car, and the line of impact. The modified ram face was 6x6 inches 

square and had half-inch radiuses on all edges and corners. 

A primary objective of these modifications to the target tank car and ram car were to create a test 

procedure with much more controlled and repeatable impact dynamics that eliminate the postimpact roll of the target tank. The modified ram would also significantly lower the required 

impact speed to puncture the target tank and would help to establish an upper bound on the 

puncture conditions within safe limits for the test facility, given the uncertainties in required 

puncture energy. The impact wall was originally designed based on static 2,700,000 lb load at 3 
Applied Research Associates, Inc.

2010

22

foot height in combination with 300,000 lb load at a 16 foot height (maximum design moment of 

12,900,000 ft-lbs). Recent analysis [22] assessed that the barrier possesses sufficient strength to 

sustain a 3,000,000-lb impact at 65 in. above top of rail (moment of 16,250,000 ft-lbs). 

Figure 46. Modified ram car and ram head used in Test 2 

The pretest and post-test conditions of the tank car are shown in Figure 47. As expected, the 

rigid ram punctured the tank car and left an opening approximately equal to the cross-section 

dimensions of the ram, about 21 inches wide by 32.5 inches high. High-speed video of the 

impact event showed that the car stayed upright throughout the test and exhibited very little body 

roll. Full instrumentation was installed to measure accelerations, strains, and displacements on

both the ram and the tank cars and high quality data was collected in the test for model 

validation. 

Additional parameters for Test 2 were as follows:

• Tank Car Number 3074

• Tank Car Weight (estimated): 265,000 pounds

• Tank Shell: 0.777-inch-thick normalized TC128B

• Insulation and Thermal Protection: ceramic fiber and fiberglass

• Jacket: 11 gauge A1011 steel at 4-inch standoff distance

• Slurry Density: 11.6 lb/gallon

• Outage: 10.4 percent

• Capacity: 17,401 gallons

• Internal Pressure: 100 psi

• Ram Car Weight: 285,400 pounds

• Ram Car Speed (official): 15.2 mph

• Ram Car Energy (derived): 2.14 million foot-pounds
Applied Research Associates, Inc.

2010

23

(a) Pretest target tank car

(b) Post-test tank damage

Figure 47. Puncture response observed in the full-scale impact Test 2. 

The average of the measured longitudinal displacements for the ram car is shown in Figure 48. 

After impact, the impact decelerations rapidly increase over the initial 0.1 seconds to a peak level 

of approximately 3.3 g. At that time the tank is punctured and the ram car acceleration rapidly 

drops to a very low level. Starting at a time of approximately 0.3 seconds, the aft support 

structures on the ram arm engage the tank and the ram car is brought to rest at a time of slightly 

over 0.6 seconds. The remainder of the acceleration pulse is the rebound of the ram car from the 

target tank car.
Applied Research Associates, Inc.

2010

24

Figure 48. Measured Test 2 ram car average longitudinal acceleration.

The measured displacement histories for the ram car and impacted tank shell are shown in Figure 

49. The ram car displacements were found from integration of the ram car impact acceleration 

history shown in Figure 48. The tank displacements are the measurements from the five string 

potentiometers that span the tank spring line. At the time of the puncture initiation (just under 

0.1 seconds), the ram has 24 inches of displacement and the center tank spring line string 

potentiometer measured 19 inches of deformation. Again, the primary difference in these two 

displacements is the compaction of the space between the jacket and the tank shell on the impact 

side of the tank. 

The ram car acceleration, shown in Figure 48, can be multiplied by the mass of the ram car to 

obtain the impact force history. This force history is plotted against the ram car displacement in 

Figure 50. Similarly, this force displacement history can be integrated to obtain the energy 

dissipation up to the point of puncture as shown in the figure. The Test 2 tank car punctured at a 

peak force of 935,000 lbs and a puncture energy of 870,000 ft-lbs. 

The measured pressure histories at the three different circumferential locations are shown in

Figure 51. The comparison of the measurements at different locations indicated that the 

hydrodynamic pressures from the sloshing of the lading were more significant than the 

hydrostatic pressure increase in the tank. The hydrostatic pressure increase is estimated to be 

less than 10 psi at the time of the tank puncture. 
Applied Research Associates, Inc.

2010

25

Figure 49. Measured Test 2 ram car and tank displacements.

Figure 50. Measured Test 2 force-displacement behavior and energy dissipation.
Applied Research Associates, Inc.

2010

26

Figure 51. Measured Test 2 tank car internal pressures.

Comparison of Test 2 with Pretest Predictions

Additional modifications were made to the tank car and ram car models to match the test 

conditions. These include the attachment of the skids to the target tank as shown in Figure 52(a) 

and the creation of a more detailed ram car model and new ram head as seen in Figure 52(b). 

The new ram car model was important to accurately simulate the collision dynamics that are 

influenced by the relative inertial properties of both the ram car and target tank car (e.g. the 

relative CG heights).

At the time Test 2 was performed, a detailed fracture model was not yet incorporated for 

prediction of the puncture energies. The development of the fracture model and application to 

predict puncture energies are described in Sections 3 and 4 of this report. At the time of the test, 

the predictions were performed where failure of the tank was suppressed and the calculated 

damage levels were used to assess whether the puncture initiation was likely. 

An example of the pretest prediction of the tank damage for Test 2 is shown in Figure 53. The 

maximum plastic strain in this test was predicted under the ram and the strain level at the inside 

surface of the tank for a 14 mph impact was approximately 50%. From this strain level it was 

determined that there was a high confidence that the impact speed selected for Test 2 would 

result in a puncture of the target tank. 
Applied Research Associates, Inc.

2010

27

(a) Updated tank car model

(b) New ram car model

Figure 52. Updated target tank car and ram car models for Test 2.

The pretest prediction of the force-deflection characteristics is compared to the measured Test 2 

force-deflection curve in Figure 54. Again, the agreement is excellent up to a post-impact ram 

displacement of approximately 23 inches, at which time the ram punctures the tank wall and the 

measured force drops dramatically. This comparison, along with the Test 1 comparison and the 

evaluation of many other quantitative measures from both tests, validate the ability of these 

models to accurately capture the impact response up to the point of fracture initiation. The 

development and addition of the fracture modeling capabilities is described in the following 

section. 
Applied Research Associates, Inc.

2010

28

Figure 53. Predicted tank damage for the Test 2 impact conditions.

Figure 54. Comparison of measured and calculated Test 2 force-displacement response.
Applied Research Associates, Inc.

2010

29

3 Material Constitutive and Damage Models

3.1 Introduction

The tank car impact analyses require a model for the material constitutive and damage behaviors 

to accurately predict the puncture threshold under various impact conditions. A piecewise linear 

elastic-plastic constitutive model was modified for this purpose to include a version of the BaoWierzbicki (BW) failure criterion [23-25]. This model is has been applied by other researchers 

to assess tank car puncture conditions [26] and is capable of reproducing both the nonlinear 

stress-strain behavior of the material as it deforms into the plastic regime and the fracture and 

failure behavior that depends on the state of stress and plastic strain history in the material. The 

material parameters used in these constitutive models were developed from the material test data 

on TC128B steel, developed under the NGRTC program [27-34]. 

The following sections describe the development of the constitutive and damage parameters used 

in the subsequent tank car puncture analyses. 

3.2 Material Stress-Strain Behavior

The first step in the development of a constitutive model is the development of the nonlinear 

stress strain behavior. This governs the mechanical response of the material and prescribes the 

internal forces (stress) that are developed as the material is deformed (strained). 

3.2.1 TC128B Material Properties

A tabular stress-strain curve was developed based on testing of different samples of TC128B 

[27]. A series of standard tensile tests were performed on different batches of TC128B, as 

shown in Figure 55. The data is consistent within each batch of material but significant variation 

can be found in tank car material obtained from different sources. The new material that was 

tested was at the upper range of strength for TC128B and the material recovered from the tank 

cars used in the test were more consistent with previous test data [14, 19]. As a result, the 

material recovered from the tank car used as the Test 2 target vehicle was used as the baseline 

material for the analyses in this report. 

The data shown in Figure 55 is the measured engineering stress-strain behavior. Engineering 

stress was obtained by dividing the measured loads by the original cross-sectional area of the 

specimen. Similarly, engineering strain was obtained by dividing the change in the specimen 

gauge-section length by the original length. 

The constitutive model in the finite element analyses requires that the engineering data be 

converted to a true stress and true strain. This conversion accounts for the changing cross 
Applied Research Associates, Inc.

2010

30

section of the specimen as it was deformed. The specimen cross section changes (shrinks) 

significantly during the test, and the engineering stress does not yield the “true” stress in this

cross section. Similarly, the engineering strain is not representative of the material behavior, 

especially when a general three-dimensional state of strain exists. As a result, the engineering 

stress decreases as some materials approach failure, implying a weakening of the material. In 

reality, the stress in the cross section is increasing due to the reduction in the cross-sectional area 

(i.e. necking).

Figure 55. Material testing data for different TC128B materials.

There are several different ways to measure stress and strain based on the coordinate system used

[28]. Some are based on material (Lagrangian) coordinates and some on spatial (Eulerian) 

coordinates. These give rise to terms such as “Green” and “Almansi” strain tensors. These are 

important in writing a computer code to solve large strain problems. An alternate approach is to 

define a “true” or “natural” stress and strain. The true stress is based on the load divided by the 

actual cross-sectional area of the specimen and is equal to the engineering stress multiplied by a 

term to correct for the change in cross section.

(1 e) σ T = σ eng + (1)

where σ T and σ eng are the true and engineering stresses, respectively, and e is the engineering 

strain.
Applied Research Associates, Inc.

2010

31

Prior to the onset of localization (necking), the natural or true strain, T ε , is defined as

ln( ) ln(1 e) l

l

o

ε T = = + (2)

This definition comes about from defining the incremental true or “natural” strain as the current 

“change in length” divided by the current length, or

l

dl dε T = (3)

This is in contrast with the definition of engineering strain that references the change in length, 

∆l , divided by the original length, 0l , or

0l

l

e

∆ = (4)

After the onset of localization, the determination of the true strain in the necked region becomes 

more complex and requires measurement of the local neck geometry.

The TC128B engineering test results are compared to the converted true stress and true strain 

data in Figure 56. The true stress curves from the tests do not include a correction for the 

necking behavior. As a result, they are only valid up to the onset of necking at a true strain of 

approximately 15%. The actual true-stress and true strain curves for the material continue to 

strain harden throughout the loading if the effects of necking were corrected. An extrapolated 

true stress curve that corrects for the effects of the necking behavior is added to Figure 56 (solid 

black line). It is this extrapolated curve that is used in the material constitutive model. 
Applied Research Associates, Inc.

2010

32

Figure 56. Comparison of engineering and true stress-strain data for TC128B.

The final step in obtaining the tabular stress-strain parameters for the TC128B constitutive model

was to fit a smooth set of points to the extrapolated true-stress data. This final tabular fit for the 

TC128B is shown in the true stress versus plastic strain curve in Figure 57. The specific values 

for the tabular stress-strain curve are also listed in Table 2. As a validation that this curve 

accurately captures the true stress-strain behavior of the material, a tensile specimen model was

generated and the constitutive parameters were applied to simulate the tensile test response. The 

calculation was analyzed to determine the engineering stress-strain behavior consistent to the 

tests (e.g. using equivalent gauge section length). A plot of the calculated engineering behavior 

compared to the test data is shown in Figure 58. The data shows that the constitutive parameters 

accurately reproduce the material behaviors including the onset and development of necking in 

the specimen. 
Applied Research Associates, Inc.

2010

33

Figure 57. Tabular true stress curve developed for the TC128B constitutive model.

Figure 58. Comparison of the measured and calculated TC128B tensile test.
Applied Research Associates, Inc.

2010

34

Table 2. Tabular TC128B stress-strain curve values

Point No. Plastic Strain (in/in) True Stress (ksi)

1 0.00e+00 58.0

2 8.22e-04 54.6

3 1.30e-02 54.8

4 2.76e-02 66.5

5 5.41e-02 79.5

6 9.87e-02 90.2

7 1.49e-01 96.0

8 1.15e+00 165.0

3.2.2 Strain Rate Effects

Elevated strain rates can influence the strength and ductility of structural materials. For most rail 

crashworthiness applications, these strain rate effects are small compared to the effects of the 

baseline (static) strength and failure modeling. In addition, strain rate effects can have very 

different trends on different material characteristics such as yield stress and failure strain. This 

makes inclusion of rate effects more difficult for problems including material failure, such as a 

tank puncture analysis. The rate effects for TC128B were investigated to assess the potential 

magnitude of the effects on tank car impact and puncture behaviors. The final result was that 

strain rate effects on tank car puncture conditions were small and not included in the analyses 

described in this report. However, the assessment of strain rate effects is provided here for 

completeness. 

An example of high rate testing to establish the rate effects on yield stress for structural steels is 

shown in Figure 59 [35]. The effects were relatively small for strain rates from below 10-2 to 

approximately 103 per second. For the tank car impact analyses the strain rates were expected to 

be on the order of 10 per second and the associated rate effects were expected to be small. 

A series of tests was performed on TC128B as a part of a separate research program [36] and the 

data was subsequently reevaluated for the NGRTC program [37]. The rate effects on the 

TC128B yield and ultimate stress are shown in Figure 60. The data spans stain rates from below 

10-2 to approximately 10 per second. The data shows a much stronger effect of strain rates than 

observed for other structural steels. 
Applied Research Associates, Inc.

2010

35

Figure 59. Example strain rate effects data for structural steels.

Figure 60. Evaluation of TC128B strain rate effects data [37].
Applied Research Associates, Inc.

2010

36

The strain-rate effects on the yield strength were included in the constitutive model for the 

preliminary tank car analyses with the Cowper-Symonds rate effect model [38]. The functional 

form for the rate effects on strength is governed by the equation:













































= +

p

C y y

1/

( ) 1 ε

σ ε σ   (5)

where σ y is the yield strength, ε is the strain rate, and C and p are the Cowper-Symonds 

parameters.

The Cowper-Symonds rate effects parameters, from Equation (5), were fit to the TC128B data 

for rates between 10-1 to approximately 10 per second. The value fit to the parameters C and p 

were 4000 and 5, respectively. The comparison of the model fit to the rate effects test data is 

shown in Figure 61. In the figure, the black line is the fit using the Cowper-Symonds model, 

while the corresponding red line is a fit to the experimental data. The comparison shows that the 

Cowper-Symonds model is capable of modeling the rate effects for only a limited range of data 

and would over predict the strength of the material at both higher and lower rates. Preliminary 

analyses using this model calculates tank and head impact behaviors that were overly stiff and as 

a result, this model was abandoned. 

Figure 61. Yield stress rate effects data and Cowper-Symonds model fit.
Applied Research Associates, Inc.

2010

37

The data at various strain rates on specimens cut from two different TC128B plates are shown in

Figure 62. The comparison of the stress-strain curves show that the data is more complex than 

would be captured by most simple rate effects models. The rate effects are observable in the 

initial yield or ultimate stress but are much smaller in different sections of the curves. In 

addition, the high rate tests have a lower ductility that would not be captured in the models. 

Overall, the strain energy dissipation capacity (area under the curves) appears to be nearly 

identical at all of the measured rates. 

As a result of the limitations of the Cowper-Symonds model for this application, a new rateeffects model was added to the user-defined constitutive model. This new model scales the yield 

stress by the ratio of the strain rate to a reference rate to a power specified by an input parameter 

according to the equation:

C

ref

y y 













 = ε

ε

σ ε σ 

 ( ) (6)

This approach will allow for a simple fit to the data over the full range of rates as shown in

Figure 60. However, it cannot reproduce the effects such as the different slopes of the rate 

effects in the yield and ultimate behaviors or the loss in ductility at high rates. As a result, the 

rate effects were not included in most of the subsequent tank car impact analyses. 
Applied Research Associates, Inc.

2010

38

Figure 62. Measured TC128B tensile stress-strain curves at various strain rates.
Applied Research Associates, Inc.

2010

39

3.2.3 Properties of Other Tank Car Materials

In addition to TC128B, a range of other structural steels are currently used in tank car designs or 

proposed for use in future designs. These range from low strength A1011 steel, currently used as 

a tank car jacket material, to HSLA 90XF steel, proposed as a candidate material for protection 

structures in future tank car designs. The true stress-strain curves used in the constitutive models 

for these materials are shown in Figure 63. The approach used to develop constitutive models 

for these materials was identical to that described for the TC128B above. However, the failure 

behaviors of these alternative materials were not all characterized as thoroughly as for TC128B. 

This detailed characterization of the failure behavior for TC128B is described in the following 

sections of this report. 

Figure 63. Tabular true stress curve developed for the various tank car steels.

3.3 Bao-Wierzbicki Failure Surface

Accurate prediction of the puncture energies of tank cars for various impact conditions requires 

the addition of a detailed damage and failure assessment capability to the material model. These 

damage mechanics, or so-called local fracture mechanics approaches (LFM), provide enhanced 

capabilities for tank car design and puncture assessment. LFM models the microstructural 

deformation and failure processes leading to fracture in terms of continuum parameters averaged 

over a small volume of material [39-46]. In contrast to classical linear elastic and elastic-plastic 
Applied Research Associates, Inc.

2010

40

fracture mechanics (LEFM and EPFM, respectively), which characterize fracture in terms of the 

boundary conditions of the fracture process zone while ignoring the details of the processes 

occurring in that zone, LFM focuses on the evolution of the process zone itself. Although LFM 

may initially seem more complex to formulate and more difficult to apply than LEFM/EPFM, it 

is more versatile and more general than the latter approaches. LFM is also ideally suited to 

implementation into finite element analyses where damage can be evaluated at the local element 

level. 

The key mechanism that needs to be included in the ductile local fracture model for tank car 

puncture analyses is the influence of the stress state on the rate of damage development as the 

material is undergoing plastic deformation. The primary stress state factor that controls the rate 

of damage development is the stress triaxiality, defined as the ratio of the mean stress to the 

equivalent stress (σmean/σeq). The mean stress (or hydrostatic stress) is the average of the three 

principal stresses (stresses on 3 orthogonal axes perpendicular to the principal planes upon which 

no shear stress exists). The equivalent stress, also referred to as the effective stress or the Von 

Mises stress is defined as

[( ) ( ) ( ) ]

1/ 2 2

3 1

2

2 3

2

1 2

2

1

σ eq = σ −σ + σ −σ + σ −σ (7)

where σ1, σ2, and σ3 are the three principal stresses. 

There are many models that include the effects of stress triaxiality on damage development and 

ductility. Several of these have previously been applied within LS-DYNA to analyze various 

ductile fracture problems [e.g. 47-49] including the use of the Gurson-Tvergaard model [50-52] 

for the puncture assessment of pressure tank cars [53]. These models have the ability to include 

the stress triaxiality effects on ductility for tensile loading as illustrated in Figure 64. The 

deficiency with many of these previous local damage models is that they do not include the 

changes in damage development and failure for low triaxiality where the tensile damage and 

failure behavior transitions into a shear dominated fracture behavior. The concern that shear 

fracture behavior is important for tank car puncture assessment led to the selection of the BaoWierzbicki (BW) model in this effort.
Applied Research Associates, Inc.

2010

41

Figure 64. Local damage criterion for tensile ductile fracture analyses.

As implemented, the BW model is a basic form of a ductile fracture criterion [43]. It assumes 

that failure at a material location occurs when the damage within a surrounding characteristic 

volume (VMIC) exceeds a critical value. The damage development and failure criterion can be 

written in the form

( ) ∫ = = 1

c mean eq

p d eq D

ε σ σ

ε

over VMIC (8)

where D is the normalized damage parameter; p

eq dε is an increment in equivalent plastic strain; 

and ( ) c σ mean σ eq ε is the critical failure strain as a function of the stress triaxiality. The 

characteristic volume (VMIC) in this application is the element size which was maintained with a 

characteristic element length of approximately 0.040 inch (1 mm) in the fracture zone. Damage 

accumulation occurs with plastic deformations and the damage is tracked locally in each element 

in the model. When the damage level in any element exceeds the failure criterion (D=1), the 

local failure is propagated in the model by element erosion.

The critical strain function is that proposed in the BW criterion and contains multiple branches 

depending on the range of stress state as shown in Figure 65. The critical strain in each branch 

are governed by the equation
Applied Research Associates, Inc.

2010

42

( )

( )

( ) ( )

[( )] ( )

( ) ( ) 





















≤

− + ≤ ≤

− ≤ ≤

+

∞ ≤ −

=

mean eq

mean eq

mean eq mean eq

mean eq

mean eq

mean eq

c mean eq

B

B A A

A

σ σ

σ σ

σ σ σ σ

σ σ

σ σ

σ σ

ε σ σ

3

1

3

3

1 9( ) 0

0

3

1

1 3

3

1

2

(9)

And the parameters A and B can be determined by a series of tests under different stress 

conditions including notched tensile tests with specimens of varying notch radii [54] and tensileshear tests with different ratios of tension to shear stress. 

Figure 65. Bao-Wierzbicki failure surface and tests used for model calibration.

3.4 Material Characterization Testing and Analyses

A wide variety of material characterization tests were performed to calibrate and validate the 

material constitutive and failure models. These include material tests such as standard tensile 

testing to assess the material stress-strain behavior and testing under various stress states 

(notched tensile or tensile-shear tests) to obtain the characteristics of the failure surface. 

Subsequent tests under more general loading conditions, such as the punch test configuration, are 
Applied Research Associates, Inc.

2010

43

used to validate the models. The approach applied here is to perform detailed analyses all of the 

material testing conditions to assess the accuracy and validity of the models. The comparisons

for the different testing conditions are provided in the following sections of this report. 

3.4.1 Tensile Tests and Analyses

Standard tensile test methods were employed in accordance with ASTM E8 [27]. The thick

TC128B allowed fabrication of a round tensile specimen (diameter of 0.505-inch). For all other 

materials a flat specimen, as shown in Figure 66, was utilized. The specimen had a gage length 

of 2-inch. No machining was performed in the thickness direction and it was tested as received. 

Photographs of the TC128B round bar specimens that were tested are provided in Figure 67. An 

extensometer with a 2-inch gage length was mounted on the specimen during testing. In one 

case, the thickest TC128B condition, strain gages were also mounted (2 gages, oriented opposite 

of each other) on the specimen gage length. Due to the excellent agreement between the strain 

gage and extensometer results, strain gages were omitted in all subsequent testing.

Figure 66. Dimensions of the specimen used in tensile testing.
Applied Research Associates, Inc.

2010

44

Figure 67. Necking behavior observed in the TC128B round bar specimens.

All tests were performed in constant-rate displacement control. Two rates were used, either 0.035 

or 0.050 inch/minute. Due to the large elongation observed in all of the steel materials, typical 

test times ranged from 15-30 minutes under the constant displacement testing. The data 

acquisition system was configured to record data at 3 Hz.

To confirm that the tensile test data reduction and material constitutive parameter extraction 

process was performed correctly, a model of the tensile specimen was generated and the tensile 

test was simulated using the constitutive model. The load and displacements were extracted 

from the analysis using consistent methodologies to the test. A comparison of the measured and 

calculated engineering stress-strain behaviors for TC128B is shown in Figure 68. The 

comparison shows that the model accurately reproduces the stress-strain behavior and captures 

the initialization of the necking response. 

Similar comparisons were made for all of the materials characterized in this program. Examples 

of the A1011 jacket material and A516-70 are shown in Figure 69 and Figure 70, respectively. 
Applied Research Associates, Inc.

2010

45

Figure 68. Validation of the tensile test behavior for TC128B.

Figure 69. Validation of the tensile test behavior for A1011.
Applied Research Associates, Inc.

2010

46

Figure 70. Validation of the tensile test behavior for A516-70.

3.4.2 Notched Round Bar Tensile Tests and Analyses

The notched round bar tensile tests were used to assess the performance of the BW failure model 

in the high stress triaxiality regime. The tests were performed using round bars with various 

notch radii [29] to achieve different levels of confinement at the notch root and thus different 

stress triaxiality levels. Models were created for the different notched round bar specimens and 

the tests were simulated. 

Material was provided from the two full-scale tank shell test articles (Tests 1 and 2) in the form 

of 12x12-inch plates. These plates were taken from near the impact site and were fabricated 

from normalized TC128B with a thickness of 0.777-inch. Specimens were extracted from these 

plates for mechanical testing. All specimens were oriented in the transverse orientation relative 

to the original plate rolling direction. The transverse orientation of the plate used to fabricate a 

tank car ring segment corresponds to the axial direction of the tank.

The notched tensile testing was performed similar to the procedures of the conventional tensile 

test standard, ASTM E8. The inner net diameter of the notched specimen is 0.25-inch and the 

gross outer diameter is 0.5-inch. The three different notch geometries, with radii of 0.25, 0.10

and 0.05-inches, were all gripped in smooth clamping friction grips. A photograph of the test 

setup is shown in Figure 71. 
Applied Research Associates, Inc.

2010

47

Figure 71. Test setup for the TC128B notched round bar specimens.

A comparison of the calculated and measured stress-strain behavior across the notch for the three 

different radii specimens is shown in Figure 72. The comparison shows that the constitutive and 

damage model were capable of reproducing both the increase in stress level and reduction in 

ductility that occur as the notch radius is reduced. The BW failure parameters used provide a 

good correlation to the observed failures of the specimens. 
Applied Research Associates, Inc.

2010

48

Figure 72. Validation of the notched round bar test behavior for TC128B.

3.4.3 Tensile-Shear Tests and Analyses

The combined tensile-shear tests were used to assess the performance of the BW failure model in 

the low stress triaxiality regime (stress triaxiality between 0 and 1/3). The tests were performed 

using a modified Arcan specimen [55] to achieve different ratios of tensile and shear by rotating 

the orientation of the gauge section relative to the loading axis, as shown in Figure 73. A 

photograph of the test setup is shown in Figure 74. Instrumentation includes clip gauges across 

both of the slots separating the upper and lower specimen sections, a string potentiometer to 

measure the displacement along the load path, the load ram LVDT, and the load cell. The tests 

were performed in a displacement control mode at a rate of 0.002 inch per second and data was 

collected at 5 Hz. 

Simulations of the specimen geometry were initially performed to assess that a relatively 

uniform stress state could be achieved in the specimen gauge section, as shown in Figure 75. 

Models were then created for each of the specimens with different gauge section orientations and 

the tests were simulated. A comparison of the calculated and measured load-displacement

behaviors are shown in Figure 76. The comparison shows again that the constitutive and damage 

model were capable of reproducing both the decrease in load level and increase in displacement
Applied Research Associates, Inc.

2010

49

that occur as the orientation was rotated from pure tension to pure shear. The BW failure 

parameters used provide a good correlation to the observed failures of the specimens. 

Figure 73. Specimen geometries for the combined tensile shear tests.
Applied Research Associates, Inc.

2010

50

Figure 74. Photograph of the combined tensile shear test configuration.

Figure 75. Analysis of the specimen behavior in the pure shear orientation.
Applied Research Associates, Inc.

2010

51

Figure 76. Validation of the combined tensile shear test behavior for TC128B.

3.4.4 Punch Tests and Analyses

The above notched round bar and tensile shear tests provide a good set of data to validate that the 

BW failure model can predict the damage development and failure of the TC128B material for a 

wide range of stress states. However, these tests were used to assess the failure parameters A and 

B in Equation 9. Therefore, an independent punch test was developed and applied to validate 

that the model could predict the failure of the material under a more general loading condition.

The specialized puncture fixture was developed for the NGRTC program at the SwRI Materials 

Test Lab and installed in a 220-kip test machine as shown in Figure 77. The punch test fixtures 

are shown with the 1.5-inch diameter punch and the 3-inch diameter receiving hole bore (socalled man hole cover). The fixture was designed so that different punch sizes and different man 

hole cover diameters can be used. 

The punches were fabricated from hardened bar stock material. The punch contact face was flat 

with a 0.30-inch radius around the perimeter. Punches were fabricated with diameters of 1.0, 

1.5, 2.0 and 3.0 inches. The dimensions of the 2.0-inch-diameter punch are shown in Figure 78. 

Man hole covers were fabricated with hole diameters of 2.0, 2.5, 3.0, 3.5, 4.5, and 5.5 inches 
Applied Research Associates, Inc.

2010

52

according to the dimensions shown in Figure 79. As a result, the punch test fixture was 

applicable for evaluating the behavior of multiple materials and structural components [32-34]. 

Figure 77. Simulation of the punch test on the thin TC128B plate material.
Applied Research Associates, Inc.

2010

53

Figure 78. Dimensions of the 2-inch-diameter punch.

Figure 79. Dimensions of the man hole cover support plates.
Applied Research Associates, Inc.

2010

54

During testing, the actuator (actually, the punch frame) displacement was measured with a 

remote LVDT mounted on the actuator. Three string pots were attached to the backside (opposite 

the punch) surface of the test specimen. These three string pots were all in a line with the middle 

location in the center of the punch and the two outer locations at the circumferential periphery of 

the punch. This provided additional measurements of the backside deformation opposite the 

punch on the front of the panel. Load was also measured from the load cell on the 

servohydraulic test frame. Tests were performed in displacement control at an applied rate of 1 

inch/minute. The multiple data channels were recorded with a custom data acquisition system 

operating at 10 Hz. This typically provided total data files of anywhere from 500-2000 points.

The test series used for the validation of the TC128B constitutive and failure model was 

performed on 0.488-inch-thick plate specimens with various combinations of ram and man hole

diameters [32]. The three tests performed used the 1.0, 1.5, and 2.0-inch diameter punches in 

combination with the 2.5, 3.0, and 3.5-inch diameter man hole covers, respectively. Photographs 

of the specimen failure behavior in these punch tests are shown in Figure 80. In all cases, the 

punch test sheared out a plug of material with a diameter approximately equal to the punch 

diameter. 

Figure 80. Punch test failure mode for the thin TC128B plate material.

An example of a model and simulation of a punch test on a 0.488-inch-thick TC128B plate is 

shown in Figure 81. The corresponding comparison of measured and calculated punch forcedisplacement curves for a series of three different tests on the TC128B plate is provided in 

Figure 82. In addition to the force-deflection curve the final profile of the plate specimens after 
Applied Research Associates, Inc.

2010

55

the punch tests were digitized and compared to the analyses. A representative profile 

comparison is shown in Figure 83. 

(a) Cross section of punch test geometry

(b) Punch test response

Figure 81. Simulation of the punch test on the thin TC128B plate material.

The agreement of the measured and calculated behaviors for this punch test provides an 

additional validation that the BW failure model is appropriate for predicting puncture of the tank 

cars. The application of the failure model for assessing puncture energies for different tank 

geometries and impact conditions is provided in the following section of this report. 
Applied Research Associates, Inc.

2010

56

Figure 82. Force-deflection curves for three punch test configurations on TC128B.

Figure 83. Comparison of the calculated and measured punch test plate profile.
Applied Research Associates, Inc.

2010

57

Punch test on 90 XF plate material

In addition to the punch test on the TC128B plate material, the punch test fixture was used to 

evaluate other material and component configurations. One of these components was a foam 

sandwich structure proposed as part of a layered foam protection system. The testing and 

analyses of this foam core panel is described in Section 6.1.1 of this report. An additional punch 

test series of interest was the comparison of the punch behavior of a single plate and multiple 

plates of equivalent total thickness. 

The two punch tests were performed using the HSLA 90 XF material. The first test was 

performed on a single 0.37-inch-thick plate and the second test was performed using three 

stacked 11 gauge sheets (nominally 0.12-inch-thick each) [34]. The comparison of the forcedeflection behaviors for the two tests is shown in Figure 84. The initial portion of the loading on 

the monolithic plate is steeper due to the higher local bending stiffness. However, after 

approximately 0.1-inch displacement, the remainder of the force-deflection characteristics in the 

two tests were very similar. The peak puncture load in the monolithic plate was approximately 

20 percent higher than that of the three layered 11 gauge sheets. 

Figure 84. Comparison of the punch loads of the single and layered 90 XF specimens.
Applied Research Associates, Inc.

2010

58

Analyses were performed for the two test configurations on the HSLA 90XF material. The 

material constitutive model included the Bao-Wierzbicki failure criterion applied in the same 

approach as for the TC128B material. The comparison of the measured and calculated punch 

test behaviors is shown in Figure 85. Again, the model was capable of predicting the overall 

characteristic and strength levels measured in the tests. 

Figure 85. Comparison of the measured and calculated 90 XF punch behavior.

3.4.5 Bulk Fracture Charpy Machine (BFCM) Tests and Analyses

The majority of laboratory material tests used to validate the constitutive and failure models were 

quasistatic. One set of dynamic tests were performed for a previous research program using the 

Bulk Fracture Charpy Machine (BFCM) [56-59]. The BFCM facility, shown in Figure 86,

consists of a large pendulum impactor retrofit with a heavy striker mass to increase the impact 

energy capacity needed to rupture the specimens. The BFCM test facility was designed to 

provide a comparative measure of energy to fracture in an unnotched specimen section. This test 

is a structural test and as such the results are dependent upon specimen and test apparatus

dimensions. 

The nominal span of the dog bone type specimens was 6 inches and the test section was full 

thickness of the plate material being tested by 1.0-inch wide. The dimensions of the specimens 
Applied Research Associates, Inc.

2010

59

used in the test series is shown in Figure 87. During pendulum impact, the specimen was 

subjected to a mixed axial and bending stress as well as the localized stress distributions around 

the contact face of the pendulum striker against the specimen. 

Figure 86. Photograph of the Bulk Fracture Charpy Machine.

A variety of test conditions were evaluated that included different materials [56], different 

impact velocities [58], different specimen thicknesses [57], and different impactor geometries. 

The impactor geometries included both blunt and sharp striker tups with the dimensions provided 

in Figure 88 and Figure 89, respectively [56]. 
Applied Research Associates, Inc.

2010

60

Figure 87. BFCM Test Specimen Dimensions.

Figure 88. Dimensions of the blunt BFCM striker tup.
Applied Research Associates, Inc.

2010

61

Figure 89. Dimensions of the sharp BFCM striker tup.

A preliminary model was generated to analyze the BFCM tests and applied to assess a set of 

impact analyses at various impact speeds that bounded the fracture energy for the specimen [58]. 

The model consisted of the test specimen, rigid grip blocks that are tightly fit to the specimen 

tapered tab geometry, and a rigid tup model with the total striker mass and initial striker velocity 

that is constrained to move in a pure translational motion. The preliminary model and impact 

response for an impact energy below the fracture energy is shown in Figure 90. 

In the BFCM test series, the fracture of the specimen was bounded by striker impacts at 8.4 mph 

and 9.4 mph. The specimens for these two impact conditions are shown in Figure 91. The

corresponding analyses of these two impact conditions with the preliminary BFCM model 

produced the deformed specimen geometries shown in Figure 92. Overall the comparison is 

favorable. The model was able to reproduce the bounds on the specimen fracture conditions 

between 8.4 and 9.4 mph. In addition, the deformation and failure modes appear to be 

consistent. 

A picture of the observed failure mode captured in a high speed video of a BFCM test is shown 

in Figure 93. A region of material ahead of the blunt tup face is mostly undeformed and a pair of 

converging shear planes develop at the impact face near the edges of the tup face that converge 

as they propagate across the specimen thickness. This general deformation and failure behavior 

was captured in the calculations, as shown in Figure 94.
Applied Research Associates, Inc.

2010

62

(a) Initial geometry

(b) Calculated impact response

Figure 90. Preliminary finite element analysis of a BFCM test.

Figure 91. Observed BFCM specimen deformations at different impact velocities.

One significant difference was observed between the measured and calculated behaviors in the 

preliminary BFCM analyses. The test specimen final deformations, shown in Figure 91, have 

tab sections that are no longer aligned with the original specimen axis. This deformation 

indicates that the tabs were allowed to rotate within the tapered grip. The tabs in the preliminary 

model were very tightly constrained and this rotation within the grip was not allowed. 
Applied Research Associates, Inc.

2010

63

Figure 92. Calculated BFCM specimen deformations at different impact velocities.

Figure 93. Observed BFCM specimen failure mode for the blunt striker.

An updated model was developed of the BFCM test configuration, as shown in Figure 95. This 

model contains a much more accurate representation of the specimen constraint conditions. The 

specimen is gripped by a 1-inch deep tapered slot machined into the grip block. The specimen is 

pushed forward in the slot and is initially in contact with the resisting face inside the grip block. 

A retaining plate is then bolted across the open side of the grip block. However, there is a gap 

between the specimen and the retaining plate equal to the difference between the 1-inch depth of 

the grip slot and the thickness of the plate material used to fabricate the specimen. In addition, 

the specimen taper does not have a tight tolerance relative to the grip taper and the specimen fits 

loosely in the grip fixture. As a result, some slip of the specimen relative to the grip block can 

occur in the tests. 
Applied Research Associates, Inc.

2010

64

Figure 94. Calculated BFCM specimen failure mode for the blunt striker.

Figure 95. Finite element model of the BFCM test.

The series of BFCM tests on various thickness TC128B plate specimens was used to validate the 

model for a dynamic impact test condition [57]. A separate set of analyses on these tests were 

previously performed by the Volpe National Transportation Systems Center [60]. The model 
Applied Research Associates, Inc.

2010

65

was used to vary the test conditions and found that the fracture energies and impact response was 

sensitive to many parameters including:

• Gap tolerances between the specimen and grip.

• Model of the specimen retaining plate (rigid, deformable, or not included,).

• Friction coefficient between the specimen and the grip block. 

The comparison of the measured and calculated fracture energies as a function of the specimen 

thickness with the blunt and sharp impactor tups are shown in Figure 96 and Figure 97, 

respectively. The analyses in these comparisons used the baseline TC128B material properties 

(shown in Figure 57) without any rate effects. All of the analyses using the rate effects models 

tended to over predict the fracture energies in the BFCM tests. In addition, these analyses used a 

0.06-inch gap between the specimen tab and grip tapers and a rigid retaining plate on the back of 

the grip block. Overall the comparison of the calculated and measured fracture energies are 

reasonably good. 

Figure 96. BFCM blunt striker fracture energy for different specimen thicknesses.
Applied Research Associates, Inc.

2010

66

Figure 97. BFCM sharp striker fracture energy for different specimen thicknesses.

In addition to the fracture energies, the measured and calculated profiles of the broken specimens 

were compared. Representative examples for tests with the blunt and sharp impactors are shown 

in Figure 98 and Figure 99, respectively. The measured and calculated deformed shapes are very 

consistent. 

The one trend observed in the BFCM testing that was not reproduced in the analyses is the 

increasing fracture energies for the thicker specimens above a linear fit (e.g. the data conforms to 

a quadratic fit). None of the parameter variations evaluated in the analyses were capable of 

reproducing this trend to the extent seen in the test data. Determination of the physical 

characteristics that are responsible for the effect is difficult since there are no active 

measurements (e.g. striker acceleration) that can be used to evaluate discrepancies in the testing 

and analyses. An attempt was made to assess the impact force history from the high speed video. 

However, there was not sufficient resolution to accurately assess the striker acceleration history. 

It is suspected that the test fixture compliance may have contributed to the fracture energy

nonlinearity. In high speed videos of the thicker specimen impact response, a significant amount 

of flexing is observable in the grip blocks. This would potentially allow for the test fixture to 

dissipate some of the impact energy and the amount would be greater for the higher loads 

associated with the thicker specimen. No attempt was made to introduce a similar level of test 

fixture compliance in the analyses to evaluate this effect. 
Applied Research Associates, Inc.

2010

67

Figure 98. Calculated and measured BFCM blunt striker specimen deformations.

Figure 99. Calculated and measured BFCM sharp striker specimen deformations.

3.4.6 Characterization of Other Structural and Energy Absorbing Materials

The above sections describe the characterization of the mechanical and failure behaviors of 

traditional tank car materials (steels). These are the materials used in the baseline tank puncture 

analyses and for the majority of the retrofit protective designs. However, additional energy 
Applied Research Associates, Inc.

2010

68

absorbing materials (structural foams) and advanced structural materials (e.g. composite 

materials) were considered for protective system designs. Analyses of these concepts required 

similar efforts to characterize the material behaviors and develop associated material models. 

The structural foams are primarily characterized by the foam density and associated unconfined 

compressive crush curve. An example of a measured unconfined compressive crush curve and 

associated analysis for a 5.4 pound per cubic foot (pcf) foam is shown in Figure 100. However, 

additional properties such as the tensile strength, interface behaviors and response under more 

general triaxial loading can also be important. As a result, additional tests such as foam 

sandwich panel bend and punch tests were also performed and analyzed. The characterization 

testing performed on foams and associated model development and validation are described in 

Section 6.1.1 of this report. 

Figure 100. Unconfined compression testing and analysis of foam.

Composite materials were thought to be a potential candidate material for a tank protective 

system since composites have many strength characteristics similar to steels but with densities 

that are approximately 1⁄4 of the steel density. Thus a significantly thicker composite protective 

structure could be applied on a tank car without exceeding a comparable weight. Composites 

have also been applied in armor systems for military vehicles achieving significant levels of 

impact and penetration protection [e.g. 61-64]. This also suggests the potential of a composite 

tank car system for significantly increasing the puncture resistance over current tank cars. 

The mechanical and failure behaviors of composites are more complex than many engineering 

materials as a result of the material anisotropy and inhomogeneity. Composites are typically 

fabricated of very strong fibers (e.g. glass, carbon, kevlar) embedded in a polymer matrix 

material (e.g. epoxy). The fibers can typically be either continuous straight fibers arranged in 
Applied Research Associates, Inc.

2010

69

layers (plies) with various orientations or with woven fabric layers. As a result many different 

composite systems exist and these systems cover a wide range of material behaviors. 

The composite system considered for this application uses woven glass fibers in a toughened 

matrix. This material system has been analyzed as part of a program to develop lightweight 

armor systems. The composite damage model development and validation against composite 

characterization tests is described in Section 6.4.1 of this report. 
Applied Research Associates, Inc.

2010

70

4 Detailed Tank Puncture Analyses

4.1 Introduction

In this section, the detailed BW failure model described in Section 3 is applied to assess various 

tank puncture conditions

4.2 Tank Car Puncture Model

As described in Section 3, the detailed failure model uses a fine mesh of brick elements in the 

impact zone with an element dimension of approximately 0.040 inch (1 mm). The model of the 

commodity tank and BW impact zone mesh used in these puncture analyses is shown in Figure 

101. An algorithm in LS-DYNA is used to tie the edge of the shell elements to the solid 

elements around the edge of the impact zone. 

Figure 101. Tank model and impact zone mesh used for side impact puncture analyses.
Applied Research Associates, Inc.

2010

71

As a result of the very fine mesh in the impact zone, the puncture models were significantly 

larger and have correspondingly longer run times. To allow for the evaluation of a wide range of 

impact conditions and tank geometries, some simplifications in the tank model were 

implemented. 

The first simplification was the use of symmetry planes. For the analyses described here, two 

symmetry planes were used to reduce the model to one-quarter of the full tank (a half model was

used for the offset head impacts). This had a small effect since some tank car structural details 

could not be included in the quarter model (e.g. manway and bolsters). The primary effect of 

this approximation was that a side impact centered on the tank may have a slightly reduced 

stiffness for large dent sizes since the manway and surrounding structures are stiffer than the 

bare commodity tank. 

Another simplification in the tank puncture model was in the modeling approach used for the 

fluid lading. In the analyses presented in Section 2 an explicit model was used for the lading. 

However, this explicit model required a significant increase in the model size and also was not 

ideally compatible with a quarter symmetry model. An alternate modeling approach was to 

smear the weight of the lading uniformly into the commodity tank wall. A comparison of the full 

tank car impact response using the two different lading modeling methodologies is shown in 

Figure 102. The comparison shows that the simplification of the smeared lading approach did 

not have a large influence on the impact behavior. 

Figure 102. Calculated tank car impact behavior using two different lading models.
Applied Research Associates, Inc.

2010

72

The final simplification of the impact modeling was that the ram car was not used in the 

analyses. Instead, a rigid model of the ram head was used with the mass of the entire ram car. 

This was a reasonable simplification since the measurements of ram car acceleration at 

distributed measurement locations were all in good agreement. 

The simulation of the Test 2 impact conditions using the tank puncture model is shown in Figure 

103. The model shown was reflected vertically about the symmetry plane (seen as a line in the 

figure) for improved visualization of the impact behavior. The impactor in this analysis was a 

rigid 6x6 inch ram with a 0.5-inch radius around the edges and a total weight of 286,000 lbs. 

The small rectangular patch of elements under the impactor (already punctured in Figure 103) is 

the fracture zone where the BW failure model was applied. The remainder of the tank structure 

was again modeled with 4-node shell elements and a tied shell-to-solid constraint was used at the 

interface of the two model regions. 

The comparison of the measured and calculated force-deflection behavior for Test 2 with the 

tank puncture model is provided in Figure 104. The comparison shows overall good agreement 

between the calculation and test. The peak load at which the tank was punctured was very 

accurately captured by the model. The primary discrepancy of the test and model was a slightly 

more compliant behavior in the model at large displacements. This difference in compliance 

could primarily be attributed to the removal of the manway from the tank model. 

Figure 103. Simplified tank model analysis with Bao-Wierzbicki failure assessment.
Applied Research Associates, Inc.

2010

73

Figure 104. Tank puncture model Test 2 analysis with Bao-Wierzbicki failure assessment.

4.3 Assessment of Baseline Tank Puncture Conditions

The tank puncture model was used to assess the puncture energies for a wide range of tank car 

design variations and impact conditions. A variety of tank cars were evaluated including 

chlorine (Cl) and Ethylene Oxide (EO) tank car designs. As a result, the tank parameters 

investigated include tank diameter (100 and 119 inches), tank thickness (between 0.5625 and 

1.102 inches), internal pressure (50 psi and 100 psi), jacket thickness, foam insulation, and 

impactor size. A summary of the tank geometric parameters used in the various analyses is 

provided in Table 3. The dimensions listed for the tank in the second column in Table 3

correspond to the diameter and length of the cylindrical portion of the tank. In addition, 2:1 

ellipsoidal heads were used to complete the tank structure. The material used for the commodity 

tank shell and/or heads in all of the analyses was TC128B if the thickness was less than one inch 

and A516-70 if greater than one inch. 

An example puncture response is shown in Figure 105. The analysis shown corresponds to the 

300 lb Cl tank geometry (note: the 300 lb tank designation is used in the industry and refers to a 

tank with a 300 psi test pressure). The ram head model was updated in these analyses to include 

the tapered geometry used in the full-scale testing and an updated weight of 295,000 lbs. The 

tapered geometry included a 6x6 inch contact face but flared out to more closely duplicate the 

initial puncture response seen in the impact tests of the NGRTC program. Although the ram 
Applied Research Associates, Inc.

2010

74

head was tapered in the analyses, the contact patch remained the 6x6 inch face until the 

protective layers and/or tank were punctured. As a result, the tapered geometry did not play a 

significant role in the prediction of the puncture energies reported in this report. 

Table 3. Parameters used for the tank impact analyses

Tank Type Tank Shell 

Dimensions (D x L)

Shell/Head 

Thickness (Inches)

Shell/Head 

Material

300 lb Cl 100 x 472 inches 0.5625 / 0.6030 TC128B

500 lb Cl 100 x 472 inches 0.7770 / 0.8281 TC128B

600 lb Cl 100 x 472 inches 0.9810 / 1.1360 TC128B / A516-70

300 lb EO 119 x 505 inches 0.5625 / 0.6030 TC128B

423 lb EO 119 x 505 inches 0.7770 / 0.8281 TC128B

500 lb EO 119 x 505 inches 0.9180 / 0.9784 TC128B

534 lb EO 119 x 505 inches 0.9810 / 1.1360 TC128B / A516-70

600 lb EO 119 x 505 inches 1.1020 / 1.2760 A516-70

The progression of the fracture behavior in the BW failure patch is shown in Figure 106. The 

fracture initiated near the corners of the impactor and propagated initially along the side of the 

impactor face. The crack subsequently ran along the top and bottom of the ram face. Finally, 

the edges of the puncture site are petaled inward as the tapered ram was forced through the tank 

wall. It should be noted that the use of two symmetry planes would potentially influence the 

fracture behavior since asymmetric crack propagation, such as was observed in Test 2, can not 

develop in the model.

The calculated force-deflection behaviors and puncture energies for the 300, 500, and 600 lb 

chlorine tanks are shown in Figure 107. These analyses used a tank diameter of 100 inches and 

an internal pressure of 100 psi. An 11 gauge jacket of A1011 steel was included with a 4-inch

standoff. The shell thicknesses for the 300, 500, and 600 lb tanks were 0.5625, 0.777, and 0.981 

inch, respectively. TC128B Material was used for all of the commodity tanks. The impact speed 

for all three analyses was 20 mph. The calculated puncture loads for the chlorine cars were 

approximately 0.7, 0.9, and 1.1 million lbs and the puncture energies were 0.67, 1.26, and 1.74 

million ft-lbs for the 300, 500, and 600 lb tanks, respectively. 
Applied Research Associates, Inc.

2010

75

Figure 105. Detailed impact and puncture sequence for a 300 lb chlorine car.
Applied Research Associates, Inc.

2010

76

Figure 106. Calculated puncture initiation and fracture progression.

Figure 107. Side impact force-deflection curves and puncture energies for chlorine tank cars.

The energy balance for the 600 lb chlorine car impact is shown in Figure 108. The initial impact 

energy is approximately 4 million ft-lbs (all in the kinetic energy). As the impact progresses, the 

ram is decelerated and the kinetic energy drops off. The energy transfer is from the kinetic 

energy of the ram to the internal energy of the tank (plastic deformations of the tank material) 

and the pressure-volume work caused by the indentation reducing the total tank volume. The 

internal energy of the tank at rupture is approximately one million ft-lbs and the pressure volume 

work is between 500,000 and 600,000 ft-lbs. The hourglass energy and sliding energy in the 

calculation are both negligible indicating that the calculation is stable and does not have any 

numerical energy losses. 
Applied Research Associates, Inc.

2010

77

Figure 108. Calculated energy balance for the 600 lb chlorine tank car impact (R1F).

The design selected for the EO tank car has a 119-inch commodity tank diameter. From the 

Code of Federal Regulations (CFR §179.100-6), the wall thickness after forming of the tank shell 

and heads must not be less than that specified in §179.101, nor that calculated by the following 

formula:

t = Pd / 2SE (10)

Where:

d = Inside diameter in inches

E = 1.0 welded joint efficiency; except for heads with seams = 0.9

P = Minimum required bursting pressure in psi

S = Minimum tensile strength of plate material in psi, as prescribed in §179.100–7

t = Minimum thickness of plate in inches after forming

Equation 10 was used to calculate the thickness of the 500 and 600 lb EO tanks to obtain the 

values provided in Table 3. The 300 lb EO tank thickness is from a required minimum thickness 

of 0.5625 inch and is the same as for the smaller diameter 300 lb Cl tank. In addition, EO tanks 

were analyzed with the 500 and 600 lb Cl tank wall thicknesses of 0.777 and 0.981 inches. The 

corresponding test pressures for these shell thicknesses when applied on the 119-inch diameter 
Applied Research Associates, Inc.

2010

78

EO tank were 423 and 534 lbs, respectively, from Equation 10. These analyses for additional 

intermediate EO tank shell thicknesses provides additional data since other EO tank car designs 

were of smaller diameter (e.g. 106 and 110 inches) and would therefore have correspondingly 

thinner tank shells for the equivalent test pressure. 

The calculated force-deflection behaviors and puncture energies for the 300, 423, 500, 534, and 

600 lb Ethylene Oxide (EO) tanks are shown in Figure 109. The significant differences between 

these analyses and the previous set for chlorine cars was that the tank diameter was increased to 

119 inches and the internal pressures were reduced to 50 psi (there was also a small difference in 

the length of the tank shell). The calculated puncture loads were approximately 0.6, 0.9, 1.06, 

1.1, and 1.17 million lbs and the puncture energies were 1.14, 1.66, 2.13, 2.7, and 2.87 million ftlbs for the 300, 423, 500, 534, and 600 lb tanks, respectively. 

The comparison shows that there was only a small difference in the puncture loads between the 

EO and chlorine cars, but there was a significant difference in the puncture energies. The 

comparison shows that the EO cars had larger displacements before they reached the puncture 

load and, as a result, had the higher energy dissipation levels. The puncture energies in the EO 

cars were approximately 50% higher than those in the chlorine cars. 

Figure 109. Side impact force-deflection curves and puncture energies for the EO Tank Cars.
Applied Research Associates, Inc.

2010

79

4.3.1 Effects of Tank Pressure and Diameter

The primary parameters that can create the difference in the puncture energies seen between the 

EO and Cl tank cars are the tank diameter and the internal pressure. To resolve which of these 

two parameters was primarily responsible, an additional analysis was performed using the 300 lb 

EO car with a 100 psi internal pressure. The comparison of this analysis with the two previous 

analyses on the 300 lb EO car (50 psi) and the 300 lb chlorine car (100 psi) are shown in Figure 

110. The comparison shows that the internal pressure was clearly the dominant parameter that 

contributes to the differences in the previous two sets of analyses. The force-deflection behavior 

and puncture energies of the 300 lb EO and chlorine cars with 100 psi internal pressure were

virtually identical. Thus, the larger tank diameter of the EO car (119 inches versus 100 inches)

had significantly less influence on the response for this tank geometry. 

Figure 110. Side impact comparison for tank cars at different pressures and diameters.

Similar analyses of the effects of internal pressure for the 500 lb chlorine tank were performed at 

50 and 100 psi. The calculated force-deflection curves are shown in Figure 111. Again the 

internal pressure had little effect on the side impact puncture force but the puncture energy 

increased by 60% from 1.26 to 2.01 million ft-lbs. A comparison of the force-deflection 

characteristics and puncture energies for the 500 lb chlorine tank car and the 423 lb EO tank car 

(both with a 0.777 in shell thickness) as well as the 500 lb EO car, all at 50 psi, are shown in 

Figure 112. The comparison shows some effect from changing the tank diameter (both at a 
Applied Research Associates, Inc.

2010

80

constant thickness or a scaled thickness), but the effects were much smaller than those created by 

the change in pressure levels. 

Figure 111. Evaluation of the effects of pressure on a 500 lb Cl tank.
Applied Research Associates, Inc.

2010

81

Figure 112. Evaluation of the effects of diameter and thickness at 50 psi.

4.3.2 Analyses of Tank Jacket Retrofit Designs

A proposed approach to increasing the puncture resistance of tank cars is to require commodities 

currently shipped in 300 lb tanks to be moved to 500 lb tank cars and commodities currently in 

500 lb cars be shipped in 600 lb tank cars. A simple alternative to achieve the same puncture 

protection levels that may allow for retrofit solutions is to maintain the current thickness of the 

commodity tank and to add the additional thickness to the jacket to increase the puncture 

resistance. A series of side impact analyses was performed to evaluate this approach. 

Chlorine Tank Cars

A comparison of a 500 lb chlorine tank cars retrofit with a 0.375-inch and 0.5625-inch jackets is 

shown in Figure 113. The comparison includes the baseline 500 and 600 lb chlorine cars for 

reference. The 500 lb chlorine tank car retrofit with the 0.375-inch jacket has a puncture energy 

of 1.73 million ft-lbs compared to 1.74 million ft-lbs for the 600 lb chlorine car. Therefore, the 

0.375-inch jacket retrofit provides an equivalent level of puncture resistance at roughly 

equivalent weight for the chlorine tank cars. When retrofit with a 0.5625-inch jacket, the 

puncture energy increases to 2.17 million ft-lbs which is significantly greater than the baseline 

600 lb chlorine tank car. 
Applied Research Associates, Inc.

2010

82

Figure 113. Evaluation of a 500 lb chlorine tank cars retrofit with thicker jackets.

A topic of interest for increasing the puncture resistance of tank cars is the relative efficiency of 

adding material to the jacket as opposed to adding material to the commodity tank. An 

indication of this can be obtained from these side impact analyses. The 500 lb Cl tank shell is 

0.204 inch thinner than the 600 lb tank shell. The 0.375-inch retrofit jacket was modified by 

adding 0.256 inch to the original 11 gauge jacket thickness (0.119 inch, nominally) and 

switching the jacket material to something similar to TC128B. As a result, the total thickness of 

the tank shell and 0.375-inch jacket were approximately 0.050 inches thicker than the baseline 

600 lb Cl tank with the 11 gauge jacket. The two analyses had approximately equivalent 

puncture energies. Thus, these analyses indicate that there is not a large difference in efficiency 

of using the additional material to reinforce either the jacket or the commodity tank. 

A summary of the analyses of the 500 lb Cl tank with various jacket thicknesses is provided in

Figure 114. The graph provides a fit to the analyses to illustrate the relationship between the 

retrofit jacket thickness and the puncture energy. The comparison shows that a retrofit jacket 

thickness of 0.375-inch is required for the 500 lb Cl tank to provide a similar level of puncture 

resistance as a 600 lb tank car with the 11 gauge jacket (as confirmed by the analysis with the 

0.375-inch jacket). 
Applied Research Associates, Inc.

2010

83

Figure 114. The effects of jacket thickness on puncture energy for a 500 lb Cl tank.

Ethylene Oxide Tank Cars

Similarly, analyses of a 300 lb EO tank car retrofit with a 0.375-inch jacket and with a 0.5625-

inch jacket (i.e. double tank configuration) are compared to the baseline 300 lb EO car in Figure 

115. The 300 lb EO tank car retrofit with the 0.375-inch and 0.5625-inch jackets have puncture 

energies of 1.85 million and 2.86 million ft-lbs, respectively. These compare to the 2.13 million 

ft-lbs for the 500 lb EO car. 

A summary of the analyses of the 300 lb EO tank with various jacket thicknesses is provided in 

Figure 116. The comparison illustrates that a retrofit jacket thickness of approximately 7/16 inch

would be required for a 300 lb EO tank to provide a similar level of puncture resistance as the 

500 lb EO car. However, the puncture energy levels calculated for the 500 lb and 600 lb Cl tank 

cars are also indicated in Figure 116. We see from the comparison that the 300 lb EO tank car 

retrofit with a 0.375-inch jacket has a puncture energy that is higher than that of the 600 lb 

chlorine car as a result of the lower pressures in EO tanks. Thus, the 0.375-inch retrofit jacket on 

EO cars will provide a protection level equivalent to or better than that of retrofit designs of 

other commodities. 
Applied Research Associates, Inc.

2010

84

Figure 115. Analyses of 300 lb EO tanks with different thickness jackets.

Figure 116. The effects of jacket thickness on puncture energy for a 300 lb EO tank.
Applied Research Associates, Inc.

2010

85

Based on these results, the relative efficiency of adding material to the jacket as opposed to 

adding material to the commodity tank was assessed for the EO tank cars. The 300 lb EO tank 

shell is 0.3555-inch thinner than the 500 lb EO tank. The retrofit requirement of a 0.4375-inch

jacket for the 300 lb EO tank car to achieve equivalent puncture resistance of the 500 lb EO tank 

adds 0.3185 inches when compared to the original 11 gauge jacket. As a result, the total 

thickness of the tank shell and 0.4375-inch jacket were approximately 0.037 inches thinner than 

the baseline 500 lb EO tank with the 11 gauge jacket. As was the case for the Cl tank, these 

analyses indicate that there is not a large difference in efficiency when using additional material 

to reinforce either the jacket or the commodity EO tank. 

4.3.3 Effects of Impactor Size

A potential concern is that the selection of the 6x6 inch ram, as the primary threat used to assess 

puncture resistance, could prevent the adaptation of a concept that could provide significant 

benefit for impacts with larger contact areas. To evaluate the effects of the ram impactor size, 

analyses were performed on the 500 lb chlorine tank car being impacted by both 9x9 inch and

12x12 inch ram impactors with a 0.50-inch radius around the edges. 

The model of the commodity tank and BW impact zone mesh used in the 12x12 inch impactor 

analyses is shown in Figure 117. Again, the refined zone is maintained along the perimeter of 

the impactor face with a characteristic element dimension of approximately 0.040 inch (1 mm). 

A comparison of the force-deflection behaviors and puncture energies with the 6 inch, 9 inch,

and 12 inch square impactors is shown in Figure 117. The higher impact speeds noted in the 

calculations with the 9 inch and 12 inch impactors were required to achieve the higher puncture 

energies associated with the larger impactors. Since the impact behavior is insensitive to impact 

speed (within a limited range), the impact velocities were not considered to be a relevant factor 

in the comparison. The increase to a 9x9 inch impactor increases the puncture load from 0.9 to 

nearly 1.3 million lbs and the corresponding puncture energy from 1.26 to 2.86 million ft-lbs. 

Similarly, the 12x12 inch impactor increases the puncture load from 0.9 to over 1.5 million lbs 

and the corresponding puncture energy from 1.26 to 4.95 million ft-lbs. The puncture initiation 

behavior at the interface of the ram and tank shell for the 12x12 ram are shown in Figure 119. 
Applied Research Associates, Inc.

2010

86

Figure 117. Tank model and impact zone mesh used for the 12x12 inch impactor.

Figure 118. Evaluation of the effects of the ram size on a 500 lb chlorine tank.
Applied Research Associates, Inc.

2010

87

Figure 119. Calculated puncture initiation and fracture progression (12 inch x12 inch ram).

A similar set of analyses were performed on the 600 lb chlorine tank car being impacted by both 

the 9 inch and 12 inch impactors. A comparison of the force-deflection behaviors and puncture 

energies with the 6 inch, 9 inch, and 12 inch square impactors are shown in Figure 120. The 9 

inch impactor increases the puncture load from 1.15 to 1.7 million lbs and the corresponding 

puncture energy from 1.74 to 3.94 million ft-lbs. The 12 inch impactor increases the puncture 

load from 1.15 to over 2.0 million lbs and the corresponding puncture energy from 1.74 to over 

8.5 million ft-lbs. 

Figure 120. Evaluation of the effects of the ram size on a 600 lb chlorine tank.
Applied Research Associates, Inc.

2010

88

One parameter in these analyses that can potentially influence the response is the impact 

velocity. In general we expect that the impact velocity will not have a large influence on the 

puncture behavior. However, the impact dynamics of the dent formation and the relative timing 

of the forces between the ram and the tank and between the tank and the reaction wall will be 

influenced by the impact velocity. An example comparison of the impact behavior for the 600 lb 

tank with the 9x9 inch impactor at 25 and 30 mph is shown in Figure 121. The impact speed has 

little effect on the puncture force. However, in this example, the impact speed has a significant 

effect on the puncture energy since the 30 mph impact fails right before the first peak of the 

force-deflection curve and the 25 mph impact fails in the second peak.

Figure 121. Effect of impact speed on the puncture behavior of the 600 lb Cl tank.

The larger displacements and force levels associated with the larger ram open new opportunities 

to incorporate alternative protection systems to increase puncture resistance and energy 

dissipation. The simple retrofit design with the 0.375-inch jacket impacted by the 12x12 inch 

ram is shown in Figure 122. The thicker retrofit jacket increases the puncture energy to over 8 

million ft lbs with the 12 inch impactor (a 77 percent increase over the 11 gauge jacket). This 

large increase results from the flattening of the force-deflection profile for ram displacements 

greater than 25 inches. Larger impactors, such as a coupler head, are often seen as a hazard in 

severe collisions or derailments. The effective dimensions of the coupler head are difficult to 

define due to the complex geometry. The external dimensions of the coupler head are 

approximately 13x22 inches. The contact area of the knuckle that would be the initial contact is 
Applied Research Associates, Inc.

2010

89

approximately 7x11 inches. By considering a wider range of impactors and collision conditions, 

a more optimized protection design may result that can further reduce the probability of a 

hazardous material release. 

Figure 122. Evaluation of the effects of the reinforced jacket with the large ram size.

A summary of the effect of the impactor size on the puncture force and puncture energy for the 

500 lb and 600 lb Cl tank cars are shown in Figure 123 and Figure 124, respectively. The two

points plotted for the 600 lb tank car and 9-inch ram size represent the results for the analyses at 

25 and 30 mph impact speeds, compared previously in Figure 121. The puncture force for the 

various impactors is compared to linear extrapolations through the forces produced by the 6 inch 

impactor. If the failure mode was dominated by the shear stress around the perimeter of the 

impact face, we would expect that the analyses would approximately follow these linear fits (this 

failure mode is discussed for the 6-inch impactor analyses below in Section 0). The drop off in 

the puncture force at larger impactor sizes (compared to the linear fits) indicates that different 

failure mechanisms are being introduced. The simplest explanation is that the load around the 

perimeter of the impactor was not uniform and that there was a less uniform distribution with 

increasing impactor size (e.g. a greater concentration of the load at the corners of the impactor). 
Applied Research Associates, Inc.

2010

90

Figure 123. Summary of puncture forces for chlorine tank cars for various impactor sizes.

Figure 124. Summary of puncture energies for chlorine tank cars for various impactor sizes.
Applied Research Associates, Inc.

2010

91

The comparison of puncture energies in Figure 124 shows that the increasing impactor size has a 

strong influence on the puncture energy. Note that the linear fit to the puncture energies shown 

in do not pass through the origin of the graph. Higher energies result from the significant 

increases in the displacement prior to puncture as the puncture force was increased (as seen in 

Figure 117 and Figure 120). This shows that any protective system that could blunt the loads 

and make the ram act like a larger impactor against the commodity tank would be an effective 

protection system. 

One potential modeling approximation that could influence the above assessment of impactor 

size was the simplified modeling of the internal pressure. In these analyses the internal pressure 

was a constant 100 psi. However, with the formation of the dent in the tank during impact there 

will be an increasing hydrostatic internal pressure level. For the 6 inch impactor this increase in 

pressure was relatively small (10%-15%). However, the larger dents produced by the larger 

impactor sizes would result in a larger internal pressure increase. 

The energy balance for the 600 lb chlorine car impact and the 12x12 inch impactor is shown in 

Figure 125. In this analysis, the tank was not fully punctured. However, threshold fracture 

initiation at the inside surface of the tank shows that puncture was very near. The initial impact 

energy is nearly 9 million ft-lbs (initial kinetic energy). As the impact progresses the ram is 

decelerated and the kinetic energy drops off nearly to zero. The energy transfer is from the 

kinetic energy of the ram to the internal energy of the tank (plastic deformations of the tank 

material) and the pressure-volume work caused by the indentation reducing the total tank 

volume. The internal energy of the tank at the end of the calculation is approximately four 

million ft-lbs and the peak pressure volume work is approximately 4.7 million ft-lbs. The 

hourglass energy and sliding energy in the calculation are both small. 

Based on the energy transfer to the pressure-volume (PV) work, the effects of the tank volume 

change can be assessed. The 4.7 million ft-lbs PV work corresponds to a change in volume of

326 cubic feet (2438 gallons) with a constant 100 psi (14400 psf) pressure assumption. Using a 

tank nominal capacity of 17,300 gallons, the change in volume is approximately 14%. 

Therefore, the change in volume (∆V) is more than the original outage volume (approximately 

10%) and the pressure would increase significantly during the response.

To determine the influence of the variable internal pressure during the impact an additional series 

of analyses was performed. In these analyses, a control volume was established for the tank and 

the change in volume was used to calculate a corresponding change in pressure. The assumption 

in these analyses is that the tank has a 10.6% outage and the gas in the outage follows an ideal 

constant temperature compression behavior during impact. The resulting relationship between 

the relative volume and pressure in the tank is shown in Figure 126. 
Applied Research Associates, Inc.

2010

92

Figure 125. Calculated energy balance for the 600 lb Cl tank car 12x12 impact (R1J).

Figure 126. Pressure-volume relationship used for the tank control volume (10.6% outage).
Applied Research Associates, Inc.

2010

93

The force-deflection curves with both constant and variable internal pressure for the 500 lb and 

600 lb Cl tanks and the three different impactor sizes are shown in Figure 127 and Figure 128, 

respectively. For the updated analyses with the variable pressure, the impact velocities were also 

reduced by 5 mph to be closer to the expected puncture threshold velocity. Adding variable 

pressure to the model results in a stiffening of the later portions of the force-deflection curve and 

a reduction in the ram displacements prior to the tank puncture. The reductions are more 

significant for the larger ram sizes. This is expected since the larger rams have larger 

displacements prior to the puncture with a larger associated pressure change inside the tank. 

Figure 127. Effect of variable internal pressure on 500 lb tank impact response.

The control volume pressure histories for the three 500 lb Cl tank impact analyses are plotted as 

a function of the ram displacement in Figure 129. In the 6x6 inch impactor analysis the pressure 

change is less than 10% and the previous approximation of a constant 100 psi internal pressure is 

a reasonable simplification for the analyses. For the 9x9 and 12x12 inch impactors, the internal 

pressures increase to approximately 120 psi and 140 psi, respectively, at the point of the puncture 

initiation. These pressure increases are becoming sufficiently large to influence the tank 

effective stiffness during the impact. 
Applied Research Associates, Inc.

2010

94

Figure 128. Effect of variable internal pressure on 600 lb tank impact response.

Figure 129. Calculated internal pressure variations in the 500 lb tank impact analyses.
Applied Research Associates, Inc.

2010

95

The control volume pressure histories for the three 600 lb Cl tank impact analyses are plotted as 

a function of the ram displacement in Figure 130. The pressure histories for the 500 lb tank 

impact analyses are also included as the dashed lines in the figure for comparison. In the 6x6

inch impactor analysis, the pressure change is again less than 10% and the approximation of a 

constant pressure remains reasonable simplification. However, for the 9x9 and 12x12 inch 

impactors, the internal pressures increase to approximately 160 psi and 215 psi, respectively, at 

the point of the puncture initiation. These pressure increases are sufficiently large to 

significantly influence the tank effective stiffness during the impact. 

Figure 130. Calculated internal pressure variations in the 600 lb tank impact analyses.

The rate of pressure rise dramatically increases after about 40 inches of ram displacement, as 

seen in Figure 130. This change in behavior corresponds to the time when the commodity tank 

begins to significantly engage the reaction wall and the back side of the tank (opposite the ram 

impact point) begins to deform. This corresponds to the development of the second peak in the 

force-deflection curve. 

The updated force-deflection curves and integrated puncture energies with the different size 

impactors are summarized in Figure 131 and Figure 132. The addition of the control volume and 

variable internal pressure produces a more linear relationship between the force and deflection 

than obtained with the constant internal pressure (shown previously in Figure 118 and Figure 

120). The corresponding updated summaries of the peak puncture forces and puncture energies 
Applied Research Associates, Inc.

2010

96

for all of the analyses at different ram sizes are provided in Figure 133 and Figure 134, 

respectively. The comparisons show that the variable internal pressure does not have a 

significant influence on the puncture forces but does reduce the puncture energies by 

approximately 25-30% in the analyses with the 9x9 and 12x12 inch impactors. 

Figure 131. Updated 500 lb tank impact analysis with different size impactors.
Applied Research Associates, Inc.

2010

97

Figure 132. Updated 600 lb tank impact analysis with different size impactors.

Figure 133. Summary of puncture forces for chlorine tank cars for various impactor sizes.
Applied Research Associates, Inc.

2010

98

Figure 134. Summary of puncture energies for chlorine tank cars for various impactor sizes.

4.3.4 Effects of Outage Volume

An additional factor that can influence the puncture energy is the percentage of the tank volume 

that is reserved as the outage volume. To evaluate the effects of the outage volume a series of 

impact calculations were performed on an unpressurized tank with a geometry corresponding to 

a DOT-111A100W tank car (no jacket). The specific geometry modeled is a 23,000 gallon tank 

shown in Figure 135. The tank is 110-inches in diameter with a 44-foot 7-inch long cylindrical 

shell and 2:1-ellypsoidal heads. The tank is constructed with a 0.4375-inch-thick A516-70 steel 

tank shell. The tanks are impacted by the standard 6x6 inch impactor (286,000 lbs) at a speed of 

16.2 mph corresponding to an initial 2.5 MJ impact energy. 

The tanks in the analyses were modeled as initially unpressurized with a control volume 

algorithm used to include the pressure-volume effects. Control volume curves were generated

representative of a tank filled with an incompressible liquid to various levels up to as much as 99 

percent of the tank capacity (1 percent outage). The control volume pressure curves (gauge 

pressure) for outage volumes between 1 and 18 percent are shown in Figure 136 and are 

generated assuming the outage volume contains an ideal gas initially at one atmosphere (absolute 

pressure). 
Applied Research Associates, Inc.

2010

99

(a) 1⁄4-symmetry model of tank, ram, and reaction wall

(b) Mesh resolution of the tank and impact patch

Figure 135. Model of a 23,000 gallon DOT-111A100W tank for analysis of outage volume effects.
Applied Research Associates, Inc.

2010

100

Figure 136. Control volume pressure curves for various outages between 1 and 18 percent.

A comparison of the impact deformations at the point of the tank shell penetration (puncture) for 

the 1, 3, and 9 percent outage volumes are shown in Figure 137. The comparison clearly shows 

that the larger outage volumes allow for a larger ram displacement before the tank is punctured. 

The comparison of the corresponding force-deflection behaviors and puncture energies with the 

different outage volumes is shown in Figure 138. A summary of the results from all of the 

calculations is provided in Table 4. The comparison shows that tanks with outage volumes 

between 1 and 18 percent are punctured at similar force levels (392,000 to 467,000 lbs) but at 

significantly different displacements (16 to 65 inches) resulting in puncture energies between 

256,000 and 1,370,000 ft-lbs. 

The calculated control volume pressures in the various analyses are compared in Figure 139 and 

the corresponding pressures at the time of tank puncture are included in Table 4. As expected, 

the displacement required to develop an internal pressure increase is significantly larger for the 

larger outage volumes. Another interesting find is that the calculated pressure levels at the point 

of the tank puncture is significantly reduced for the larger outage volumes. Thus, the fraction of 

the impact force resulting from the impact deformations is larger as the puncture deformation is 

increased and the effects of increasing the outage volume would be expected to have a 

diminishing return. 
Applied Research Associates, Inc.

2010

101

 

(a) Deformation at puncture for 1% outage

 

(b) Deformation at puncture for 3% outage

 

(c) Deformation at puncture for 9% outage

Figure 137. Calculated impact and puncture behaviors for different outage volumes.
Applied Research Associates, Inc.

2010

102

Figure 138. Force-deflection curves and puncture energies for different outage volumes.

Table 4. Summary of impact analyses to assess outage volume effects

Outage 

Volume Tank Shell Impact 

Conditions

Ram Puncture 

Displacement

Puncture 

Force (lbs)

Puncture 

Energy

(ft-lbs)

Pressure at 

Puncture

1 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 16 inches 467,000 256,000 178 psi

2 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 25 inches 464,000 474,000 89 psi

3 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 28 inches 452,000 537,000 53 psi

6 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 36 inches 439,000 715,000 29 psi

9 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 42 inches 428,000 825,000 22 psi

12 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 47 inches 415,000 935,000 18 psi

15 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 53 inches 410,000 1,075,000 17 psi

18 % 0.4375 in

A516-70

16.2 mph

6”x6” ram 65 inches 392,000 1,370,000 21 psi

100 % 0.4375 in

A516-70

16.2 mph

6”x6” ram >110 inches n/a >1,800,000 n/a
Applied Research Associates, Inc.

2010

103

Figure 139. Control volume pressures for impacts with different outage volumes.

The calculated puncture energies are plotted against the outage volume in Figure 140. Also 

included in the figure is an approximate fit to the calculated energies. The functional form of the 

fit is a puncture energy that is proportional to the square root of the outage volume. Thus, the 

effect of an incremental increase of the outage volume results in a corresponding smaller 

increase in the puncture energy as the outage volume grows larger. 

The calculated puncture energies have some natural variability about the functional fit. This 

variability can be seen by considering the impact behavior of an empty tank as shown in Figure 

141. In the empty tank analysis the impactor dents the side of the tank and the dent continues to 

grow until the impactor eventually impacts the far side of the tank and impact wall (greater than 

110 inch dent depth. As the side of the tank collapses the resistance of the tank dent oscillates 

between 150,000 and 300,000 lbs as the dent grows, never reaching a level that is sufficient to 

puncture the tank wall. The timing of the puncture for the different outage levels would be 

influenced by these natural variations in the impact resistance and as a result the calculated 

puncture energies do not fall directly on the smooth functional fit. 
Applied Research Associates, Inc.

2010

104

Figure 140. Effect of the outage volume on the puncture energy in side impacts.

Figure 141. Force-deflection curve and impact energy dissipation for an empty tank.
Applied Research Associates, Inc.

2010

105

An inspection of the puncture results in Table 4 indicates a trend of modest reduction in the 

puncture force as the outage volume is increased. The mechanisms believed responsible for this 

is the uniformity of the stresses around the edge of the impact face. The profile of the impact 

damage during the indentation process for both the 1 percent and 18 percent outage calculations 

is shown in Figure 142. When the tank has a 1 percent outage volume the damage development 

is relatively uniform around the perimeter of the contact patch other than the slight 

concentrations of damage at the corners of the impact face. However, for the 18 percent outage 

the damage along the top and bottom of the impact face is greater than that along the sides of the 

impactor. In addition, the larger dent depths would allow for more plastic bending around the 

edges and corners of the impact face which could result in a less uniform stress state and damage 

development through the thickness of the tank wall. 

(a) Damage Profile for 1% outage (b) Damage Profile for 18% Outage

Figure 142. Side impact damage distribution for 1% and 18% outage volumes.

4.3.5 Effects of Impact Speed

A potential factor for assessing the impact conditions that will puncture a tank is the speed of the 

impacting object. As a reference, the lower bound of impact speed was assessed by performing a 

quasistatic loading of the tank car using the 6x6 inch impactor as a punch. The comparison of 

the dynamic impact response with the quasistatic behavior is shown in Figure 143. 
Applied Research Associates, Inc.

2010

106

Figure 143. Effect of impactor speed on the puncture of a 500 lb Cl tank.

The comparison shows that there are significant differences in the force deflection behavior in 

the dynamic and quasi-static loading. The quasistatic loading results in a lower force level for a 

given indentation depth and a much more linear force deflection characteristic where the reaction 

force is nearly proportional to the indentation depth. The higher force level in the dynamic 

impact is due to the inertial force contribution prodeuced by the need to rapidly accelerate 

portions of the tank to develop the indentation at 20 mph. 

The reduced slope at late time in the dynamic analysis is produced by the drop in the relative 

speeds between the impactor and the tank. This contains contributions from both the impactor 

decelerating and the acceleration of the impacted tank. 

4.3.6 Summary of the Side Impact Analyses

A summary of the NGRTC side impact analyses including the impact conditions and calculated 

puncture forces and energies is provided in Table 5 (excluding the outage effects analyses 

summarized in Table 4). The table includes the tank geometry, impact conditions, and the 

calculated puncture force and energy for each tank impact simulation.

The calculated puncture forces for all of the tank impact analyses using the 6x6 inch impactor are 

plotted against the combined thickness of the tank shell and jacket in Figure 144. The figure 

indicates a linear relationship between puncture force and total thickness of the protective layers. 
Applied Research Associates, Inc.

2010

107

The linear relationship between the puncture force and total tank system thickness provides an 

indication of the primary failure mechanism initiating the tank puncture. The geometry of the 

ram impacting and indenting a pressurized tank shell is shown in Figure 145(a). A force-balance 

analysis in the direction of the impact on a patch of tank shell material is shown in Figure 145(b). 

The forces resisting the impact loads are the pressure on the inside surface of the contact patch 

and the shear stress around the perimeter of the contact patch. For a 100 psi tank pressure and a 

6x6 inch impactor, the total pressure load is less than 4 kips on the contact patch. Thus, the 

average shear stress is approximately equal to the impact force divided by the area of the 

perimeter of the impactor face divided by the tank thickness. 
Applied Research Associates, Inc.

2010

108

Table 5. Summary baseline side impact analyses

Calculation Tank 

Type

Tank 

Shell Shell Jacket Impact 

Conditions

Internal 

Pressure

(psi)

Puncture 

Force (lbs)

Puncture 

Energy

(ft-lbs)

R1A 300 lb EO 0.5625 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 50 psi 616,000 1,145,000

R1B 423 lb EO 0.777 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 50 psi 883,000 1,663,000

R1C 534 lb EO 0.981 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 50 psi 1,141,000 2,709,000

R1D 300 lb Cl 0.5625 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 100 psi 676,000 673,000

R1E 500 lb Cl 0.777 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 100 psi 915,000 1,261,000

R1F 600 lb Cl 0.981 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 100 psi 1,152,000 1,742,000

R1G 300 lb EO 0.5625 in

TC128B

0.375 in

TC128B

20 mph

6”x6” ram 50 psi 776000 1,849,000

R1H 500 lb Cl 0.777 in

TC128B

0.375 in

TC128B

20 mph

6”x6” ram 100 psi 1,010,000 1,732,000

R1I 500 lb Cl 0.777 in

TC128B

0.119 in

A1011

30 mph

12”x12” ram 100 psi 1,512,000 4,885,000

R1J 600 lb Cl 0.981 in

TC128B

0.119 in

A1011

30 mph

12”x12” ram 100 psi (2,085,000)

1 (8,536,000)1

R1K 500 lb Cl 0.777 in

TC128B

0.375 in

TC128B

30 mph

12”x12” ram 100 psi 1,948,000 8,660,000

R1L 300 lb EO 0.5625 in

TC128B

0.5625 in

TC128B

20 mph

6”x6” ram 50 psi 879,000 2,859,000

R1M 300 lb EO 0.5625 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 100 psi 686,000 675,000

R1N 500 lb Cl 0.777 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 50 psi 909,000 2,011,000

R1O 500 lb EO 0.918 in

TC128B

0.119 in

A1011

20 mph

6”x6” ram 50 psi 1,066,000 2,131,000

R1P 600 lb EO 1.102 in

A516-70

0.119 in

A1011

20 mph

6”x6” ram 50 psi 1,171,000 2,874,000

R1Q 500 lb Cl 0.777 in

TC128B

0.5625 in

TC128B

20 mph

6”x6” ram 100 psi 1,090,000 2,175,000

R1R 500 lb Cl 0.777 in

TC128B

0.119 in

A1011

25 mph

9”x9” ram 100 psi 1,286,000 2,857,000

R1S 600 lb Cl 0.981 in

TC128B

0.119 in

A1011

25 mph

9”x9” ram 100 psi 1,701,000 3,941,000

R1T 500 lb Cl 0.777 in

TC128B

0.375 in

TC128B

25 mph

9”x9” ram 100 psi 1,476,000 3,878,000

Note (1) Tank was not fully punctured at this impact velocity.
Applied Research Associates, Inc.

2010

109

Figure 144. Side impact puncture force as a function of system thickness.

The slope of the linear fit in Figure 144 corresponds to an average shear stress in the tank layers 

around the perimeter of the impactor of 39 ksi. By comparison, the yield and ultimate stress 

levels of the TC128B in pure shear are 33 ksi and 49 ksi, respectively (approximately 58% of the 

stress values in pure tension using a Von Mises yield criterion). Thus, the failure mode is 

primarily exceeding the shear capacity around the perimeter of the impact patch. 

The calculated puncture energies for all of the analyses using the 6x6 inch impactor are plotted 

against the combined thickness of the tank shell and jacket in Figure 146. When comparing 

puncture energy the data falls into two separate groups that are distinguished by the tank 

pressure. Again, the comparison indicates that the total thickness was the dominant parameter 

that determines puncture energy for a given pressure level and indicates that a retrofit design 

with an increased jacket thickness should provide equivalent protection. 

A similar comparison of the puncture energy of the retrofit 300 lb EO cars and 500 lb Cl cars as 

a function of thickness is shown in Figure 147. The comparison illustrates that the addition of 

material to the jacket was much more efficient at increasing puncture energy for the more 

compliant EO tank cars (resulting from the thinner baseline tank shell thickness and lower 

pressure). 
Applied Research Associates, Inc.

2010

110

(a) Geometry of the tank indentation

(b) Free body diagram for the tank contact patch

Figure 145. Loading and failure mechanism for the tank impact and puncture.

More advanced solutions to increasing the puncture resistance of tank cars includes the use of 

multi-layer steel and foam systems or engineered metal structures (EMS) such as a corrugated 

steel system being evaluated in the NGRTC program or the egg crate EMS concept proposed by 

the Volpe Transportation Systems Center and FRA [11, 65]. Several of these advanced concepts 

have been analyzed and results are provided in Section 6 of this report. However, of the systems 

analyzed to date, none are providing side impact protection levels that would survive a 25 mph 

impact with a 6x6 inch impactor (approximately 6 million ft-lbs). The principal challenge is that 

the small impactor face concentrates too large a load around the small perimeter of the impactor 

face and as a result, punctures through the protection system. 
Applied Research Associates, Inc.

2010

111

Figure 146. Side impact puncture energy as a function of system thickness.

Figure 147. Puncture energy of retrofit tank cars as a function of jacket thickness.
Applied Research Associates, Inc.

2010

112

5 Detailed Head Impact Analyses

5.1 Introduction

The tank head is subjected to a significant percentage of the overall impacts in collisions and 

derailments. The significant number of head impacts justifies additional protection compared to 

the sides of the tank. For this reason, head shields have been used for many years to provide an 

increased protection against head impacts. Thus, greater head impact protection offers the 

opportunity for substantial improvements in tank car safety. Additionally, the tank head 

comprises a small percentage of the overall tank surface area, so the weight penalty for 

additional protection is not as severe. 

5.2 Head Impact Analyses and Model Validation

In addition to the full-scale side impact tests, three full-scale head impact component tests were 

performed under the NGRTC project. In these tests, only the tank head and external protective 

structures were used to simplify the test from a full tank car impact test. Tank heads were 

unpressurized in these tests and the test head and support structure were ventilated to 

atmospheric pressure. The tests did, however, use an off-center impact point which adds some 

complexity to the impact response. Baseline head impact testing and analyses are described 

here. 

5.2.1 Head Test 1 – Baseline Head Impact Behavior

The first component test of a baseline tank car head test article was performed as part of the Next 

Generation Rail Tank Car project on October 30, 2007, at the Transportation Technology Center, 

Inc. (TTCI) in Pueblo, Colorado [66]. This test involved a 289,550-pound ram car impacting the 

baseline tank head at 8.66 mph. The target impact point was offset approximately 29 inches 

vertically downward from the center of the head. The ram car was equipped with a 6-inch square 

indenter with 1⁄2-inch radius around the edges. The tank head was affixed to the Transportation 

Technology Center impact barrier with a rigid fixture. 

The test article and the ram car staged in position prior to the test are shown in Figure 148. The 

test article was bolted to the wall-offset fixture. The wall-offset fixture was bolted to the impact 

wall and an energy absorbing arrestor system was added to dissipate residual energy if the head 

was punctured. The arrestor system consists of four hydraulic cushion units, two on each side, 

placed to engage the ram car shortly after the head penetration. To withstand any potential loads 

from the arrester system, additional reinforcement was added to the underframe structures on the 

impacting end of the ram car. Some of this reinforcement on the side sills is seen in Figure 148.
Applied Research Associates, Inc.

2010

113

Figure 148. Pretest conditions of the test article and ram car for Head Test 1.

Prior to the baseline Head Test 1, the ram car was weighed to determine the increase from the 

reinforcements. The A-end weight (ram end) was 163,300 lbs and the B-end weight was

126,250 lbs for a total weight of 289,550 lbs. The impact speed, measured by the reflector speed 

trap, was 8.66 mph. The ram punctured the jacket. The commodity tank was significantly 

deformed, but it did not puncture. 

The damage to the jacket and tank head after the impact is shown in Figure 149. A close up 

view of the impact damage to the tank head is shown in Figure 150. The impact behavior shows 

that the ram initially contacts a point approximately 6 inches higher than the contact position at 

the end of the impact behavior. The head indentation in the offset impact geometry results in a 

vertical local displacement and sliding of the head material relative to the inward motion of the 

ram. Ultrasonic thickness testing showed that there was thinning of the tank head material in the 

region of the impact. The peak longitudinal acceleration was about 2.8 g after processing through 

a 60 Hz channel frequency class filter (CFC 60). 
Applied Research Associates, Inc.

2010

114

Figure 149. Post-test damage of the test article for Head Test 1.

Figure 150. Impact damage on the tank head in Head Test 1.
Applied Research Associates, Inc.

2010

115

5.2.2 Head Puncture Model

The head puncture modeling approach was identical to that of the tank side impact puncture 

analyses with a fine mesh of brick elements in the impact zone (both in the head and head 

shields) with an element dimension of approximately 0.040 inch (1 mm). The head impacts 

analyzed matched the Test 1 impact condition (impact point offset below the tank head center). 

As a result, the problem had only 1 symmetry plane running vertically through the test specimen. 

This symmetry plane was used in most of the head impact simulations to reduce the model size 

by one half compared to the full head model. 

5.2.3 Comparison of Head Test 1 with Analyses

The damage to the head was calculated for the Head Test 1 impact conditions, as shown in 

Figure 151. The calculated impact behavior included the puncture of the jacket, shown in Figure 

151 (a), and impact damage that did not fully puncture the head, seen in Figure 151 (b). The 

head damage included the sliding of the head material relative to the impactor as seen in the test. 

The predicted damage shows that this impact velocity was very nearly sufficient to puncture the 

tank head. All of these calculated impact responses are in good agreement with the observed 

Head Test 1 response. 

 

(a) Puncture of the jacket (b) Impact damage on the head

Figure 151. Calculated impact response in Head Test 1.
Applied Research Associates, Inc.

2010

116

The measured and calculated force-deflection characteristics are shown in Figure 152. The 

response shows a very linear behavior for the head impact and the peak load was approximately 

800,000 lbs at a ram displacement of approximately 24 inches. The agreement between the 

force-deflection behaviors for the test and analysis is excellent. This comparison contributes to 

the further validation of the modeling procedures used to assess the puncture energies of tank 

cars under various impact conditions. 

Figure 152. Calculated and measured force-deflection behaviors for Head Test 1.

5.3 Assessment of Head Impact Puncture Conditions

The head puncture model was used to assess the puncture energies for a wide range of tank car 

design variations and impact conditions. The Cl tank head parameters investigated included

head thickness (300, 500, and 600 lb tank heads), internal pressure (unpressurized and 100 psi), 

jacket thickness, foam insulation, and impactor size. The head geometry was a 2:1 ellipsoid with 

a major diameter of 100 inches, consistent with the chlorine car dimensions. The jacket was a 

flanged and dished geometry with a 4 inch standoff at the head center and a 108 inch diameter at 

the transition to the jacket. 

The model of the commodity tank head and BW impact zone mesh used in the head puncture 

analyses is shown in Figure 153. Due to the offset impact condition, only one vertical plane of 

symmetry exists and the model is for one half of the tank head. Again, the refined zone is 
Applied Research Associates, Inc.

2010

117

maintained along the perimeter of the impactor face with a characteristic element dimension of 

approximately 0.040 inch (1 mm). 

Figure 153. Tank head model and impact zone mesh used for the head puncture analyses.

An example head impact puncture response is shown in Figure 154. The ram head model had 

the same tapered geometry used in the full-scale testing and had an updated weight of 295,000 

lbs. The tapered geometry included a 6x6 inch contact face but flares out to more closely 

duplicate the initial puncture response seen in the impact tests. The tank head in this analysis 

was a 500 lb head (0.8281 inches thick) with a 0.50-inch thick A516-70 full height head shield. 

The fractures of the tank and head shield initiate near the upper corners of the impactor and 

propagate along the sides and top edge of the impactor face forming a flap in the head. 
Applied Research Associates, Inc.

2010

118

(a) Initial impact conditions (b) Puncture response of the head

Figure 154. Calculated puncture behavior of a head and head shield.

The energy balance for the 500 lb head with a 0.50-inch thick A516-70 head shield is shown in

Figure 155. In this analysis, the tank head is unpressurized. The initial impact energy is 2.85 

million ft-lbs (all in the kinetic energy). As the impact progresses the ram is decelerated and the 

kinetic energy drops off to approximately 1.64 million ft-lbs. The energy transfer is from the 

kinetic energy of the ram to the internal energy of the tank deformations. There is no pressurevolume work since the tank head is unpressurized. The internal energy of the tank at the end of 

the calculation is approximately 1.2 million ft-lbs (equal to the drop in kinetic energy). The 

hourglass energy and sliding energy in the calculation are both small. This energy balance is 

typical for these analyses of the unpressurized head impacts. 
Applied Research Associates, Inc.

2010

119

Figure 155. Calculated energy balance for the 500 lb Cl tank head and Head Shield (R1C).

5.3.1 Baseline Impact Analyses on Unpressurized Heads

The calculated impact loads for a set of calculations on unpressurized heads is shown in Figure 

156. The head geometries include the baseline 300, 500, and 600 lb tank heads with 11 gauge 

jackets and additional analyses for the 500, and 600 lb tank heads with 0.50-inch full height head 

shields. In this set of analyses, TC128 was used in modeling the 600 pound commodity tank for 

comparative purposes. It was recognized that TC128 is not available in plates over one inch 

thick. Subsequent head impact analyses on the 600 lb head used A516-70 as the head material. 

A summary of these analyses on unpressurized heads, including the impact conditions and 

calculated puncture forces and energies, is provided in Table 6. The calculations shown in 

Figure 156 are summarized in the first five rows of the table. 
Applied Research Associates, Inc.

2010

120

Figure 156. Calculated force-deflection behaviors for baseline head geometries.

Table 6. Summary of unpressurized head impact analyses

Calculation Head 

Thickness

Jacket or 

Head Shield Shell Jacket Impact 

Velocity

Puncture 

Force 

(lbs)

Puncture 

Energy

(ft-lbs)

R1F 0.6030 11 gauge A1011 11 gauge A1011 10 mph 636,000 783,000

R1E 0.8281 11 gauge A1011 11 gauge A1011 12 mph 898,000 903,000

R1C 0.8281 0.500” A516-70 0.500” A516-70 17 mph 981,000 1,196,000

R1D 1.1360 11 gauge A1011 11 gauge A1011 15 mph 1,305,000 1,274,000

R1G 1.1360 0.500” A516-70 0.500” A516-70 18 mph 1,380,000 1,455,000

R1H 0.8281 0.8281” TC128B 0.8281” TC128B 17 mph 1,430,000 1,842,000

R1I 0.8281 0.500” A516-70 11 gauge A1011 17 mph 955,000 1,140,000

R1J 0.6030 0.8281” TC128B 0.375” TC128B 17 mph 1,288,000 1,983,000

One feature of the design that can potentially impact the head impact response is the stiffness of 

the head shield support provided by the side jacket shell. In the example shown in Figure 154, 

the 0.50-inch head shield thickness was extended to the reaction frame providing a high stiffness 

support (Calculation R1C). This would be more representative of the support in a retrofit design 

where the jacket is significantly thicker than the baseline 11 gauge A1011 jacket. 
Applied Research Associates, Inc.

2010

121

For comparison, a second head impact analysis was performed under identical impact conditions 

where the only difference was the substitution of an 11 gauge A1011 jacket in the cylindrical 

section behind the head shield (Calculation R1I). The calculated impact behavior for this head 

geometry is shown in Figure 157. The initial impact and indentation of the head was identical. 

However, after approximately one foot of ram displacement, the loads were sufficiently high to 

initiate the buckling response in the jacket seen in Figure 157. The comparison of the forcedeflection curves for the two corresponding analyses are shown in Figure 158. The comparison 

shows that the bucking of the jacket had little effect on the head impact response (either impact 

force or the puncture energy). 

 

Figure 157. Buckling behavior for the 11 gauge jacket in the offset head impact.
Applied Research Associates, Inc.

2010

122

Figure 158. Effect of jacket stiffness on the calculated head impact response.

5.3.2 Baseline Impact Analyses on Pressurized Heads

The above head impact analyses were performed using a configuration similar to the head impact 

component tests in the NGRTC program. Although these analyses are representative of the fullscale tank head impact behavior, a concern was that neglecting the internal pressure in the head 

impacts will influence the response and potentially change some of the resulting conclusions. As 

a result, a series of analyses with internal pressure were also performed. 

A summary of the pressurized head impact analyses is provided in Table 7. In all of these 

analyses the internal pressure was 100 psi. In this series of analyses, the material used for the 

tank head was TC128B when the thickness was less than one inch and A516-70 when the 

thickness was greater than one inch. The baseline 0.5 inch head shield material was also 

switched to A572-50. The relative stress-strain behaviors for these materials were previously 

provided in Section 3.2.3 of this report. The calculated force-deflection behaviors and puncture 

energies for the pressurized 300, 500, and 600 lb heads, all with a 0.5 inch A572-50 head shield, 

are shown in Figure 159. The 300, 500, and 600 lb heads were Calculations R1K, R1L, and 

R1P, respectively, in Table 7. 
Applied Research Associates, Inc.

2010

123

Table 7. Summary of pressurized head impact analyses.

(1)

Calculation Head 

Thickness

Jacket or 

Head Shield Shell Jacket Impact 

Velocity

Puncture

Force 

(lbs)

Puncture 

Energy

(ft-lbs)

R1K 1.1360 0.500” A572-50 11 gauge A1011 14 mph 1,206,000 1,121,000

R1L 0.8281 0.500” A572-50 11 gauge A1011 10 mph 966,000 916,000

R1M 0.8281 0.8281” TC128B 11 gauge A1011 14 mph 1,289,000 1,321,000

R1N 1.1360 0.500” A572-50 11 gauge A1011 11 mph 1,229,000 1,110,000

R1O 0.6030 0.8281” TC128B 0.375” TC128B 11 mph >1,240,000 >1,190,000

R1P 0.6030 0.500” A572-50 11 gauge A1011 10 mph 813,000 782,000

R1Q 0.6030 0.8281” TC128B 0.375” TC128B 14 mph 1,316,000 1,537,000

R1U 0.8281 0.8281” TC128B 11 gauge A1011 14 mph 1,311,000 1,482,000

R1Y 0.8281 0.680” TC128B 0.375” TC128B 14 mph 1,292,000 1,390,000

R1Z 0.8281 11 gauge A1011 11 gauge A1011 10 mph 813,000 610,000

R0A (2) 0.8281 0.8281” TC128B 11 gauge A1011 10 mph 1,218,000 1,494,000

R0B (2) 0.8281 0.680” TC128B 0.375” TC128B 14 mph 1,195,000 1,252,000

R0C (2) 0.8281 0.500” A572-50 11 gauge A1011 14 mph 952,000 1,100,000

R0D 0.8281 0.680” TC128B 11 gauge A1011 14 mph 1,189,000 1,281,000

R0E 0.8281 0.8281” TC128B 0.375” TC128B 14 mph 1,466,000 1,661,000

Notes: (1) All analyses performed on chlorine tank heads using the 6x6 inch impactor

(2) Analyses with the shallow head shield profile.
Applied Research Associates, Inc.

2010

124

Figure 159. Calculated force-deflection behaviors for pressurized heads.

The comparison of the calculated force-deflection curves for the unpressurized and pressurized 

500 lb head configurations is shown in Figure 160. The solid red lines are for the analysis of the 

unpressurized head and the blue lines are the pressurized head. The differences in the two 

analyses were the addition of 100 psi internal pressure and the head shield material was switched 

from A516-70 to A572-50. The comparison of these two analyses show nearly identical 

puncture forces but the puncture energy was reduced from 1.20 to 0.92 million ft-lbs in the 

pressurized head. The higher stiffness of the head resulting from the internal pressure resulted in 

the puncture occurring at approximately 5 inches less ram displacement.

The calculated force-deflection characteristics for a pressurized 500 lb Cl head (100 psi) with 

various thickness head shields is shown in Figure 161. In the figure, the thicker retrofit head 

shields (0.68 and 0.8281 inch thick) also have a thicker 0.375 inch retrofit jacket supporting the 

head shields. A similar comparison where the 11 gauge jacket was used to support all of the 

head shields is shown in Figure 162. As expected, the increasing thickness of the head shield 

increases the overall stiffness of the system, the peak puncture forces, and the puncture energies. 

The magnitudes of the puncture forces and puncture energies are summarized in Table 7. The 

side jacket thickness was also seen to have a significant contribution, as shown in Figure 163. 
Applied Research Associates, Inc.

2010

125

Figure 160. Force-deflection comparison for pressurized and unpressurized heads.

Figure 161. Force-deflection comparison for pressurized heads and various head shields

(0.375-inch retrofit jacket thicknesses).
Applied Research Associates, Inc.

2010

126

Figure 162. Force-deflection comparison for pressurized heads and various head shields

(11 gauge jackets).

Figure 163. Force-deflection comparisons with different jacket thicknesses.
Applied Research Associates, Inc.

2010

127

Ram Impact Speed Effects

An assumption of the analyses presented in this report is that the impact response is relatively 

insensitive to the impact speed within a limited range. As a result, each system is evaluated by 

selecting a collision speed that is between 25% and 50% higher than the estimated threshold 

puncture speed. The impact energy is then obtained from the resulting force-deflection 

characteristics prior to the point of puncture initiation. The magnitude of the residual kinetic 

energy in the ram is not considered to be significant. In this section, comparisons are made of 

analyses at different impact speeds to assess the validity of this assumption. 

An analysis of a 300 lb tank head with the 0.8281 inch TC128B head shield found that an 11 

mph impact velocity was too low to rupture the head (Calculation R1O). A repeat analysis was

performed at a 14 mph impact velocity. A comparison of the force-deflection characteristics for 

the analyses at the two different impact speeds are shown in Figure 164. The comparison shows 

that the analyses follow a nearly identical force-deflection curve. Thus, the deflection behavior 

was independent of the collision speed within the range of impact speeds investigated. 

Figure 164. Calculated force-deflection behaviors at different impact speeds (300 lb head).

A similar set of analyses for the baseline 600 lb head geometry were performed at 11 and 14 mph 

impact speeds. A comparison of the corresponding force-deflection behaviors and impact 

energies is provided in Figure 165. The comparison again illustrates that there was very little 

difference in the impact response or puncture energies within this limited range of impact 
Applied Research Associates, Inc.

2010

128

velocities. Thus, the methodology was considered to be valid and additional analyses at impact 

speeds closer to the threshold puncture speed were not considered to be necessary. 

Figure 165. Calculated force-deflection behaviors at different impact speeds (600 lb head).

Head Shield Profile Effects

Another tank car design parameter that can potentially influence the head impact behavior and 

puncture resistance is the shape of the head shield. Head shields traditionally have a flanged and 

dished construction. However, the parameters of this head shield profile are not as uniform or as 

tightly constrained as the tank head profiles. An example of two different head shield profiles 

are shown in Figure 166. The first profile has a deeper dish profile that more closely conforms 

to the head geometry. This deep dish profile has been used for the majority of the head impact 

analyses in this report. The second profile has a shallower dish and results in a larger standoff 

distance from the head at locations other than at the center of the head. 
Applied Research Associates, Inc.

2010

129

 

(a) Deep head dish profile (b) Shallow head dish profile

Figure 166. Models developed to assess the effects of the head shield profile.

Three different head shield thicknesses were evaluated with the two different dish profiles. 

These were a 0.50-inch-thick A572-50 head shield and 0.680-inch-thick and 0.8281-inch-thick 

TC128B head shields. An example of the calculated deformation response is shown in Figure 

167. The comparisons of the calculated force-deflection characteristics and impact energy 

dissipation for the 0.500, 0.680, and 0.8281-inch-thick head shields are provided in Figure 168, 

Figure 169, and Figure 170 respectively. There are significant effects of the head shield profile 

on the details of the impact response in each comparison. However, no significant consistent 

trends were observed on the relative impact protection performance between the two designs. 
Applied Research Associates, Inc.

2010

130

 

(a) Deep head dish profile (b) Shallow head dish profile

Figure 167. Calculated Puncture response for the two head shield profiles.

Figure 168. Calculated effects of the head shield profile (0.500-inch-thick shield).
Applied Research Associates, Inc.

2010

131

Figure 169. Calculated effects of the head shield profile (0.680-inch-thick shield).

Figure 170. Calculated effects of the head shield profile (0.8281-inch-thick shield).
Applied Research Associates, Inc.

2010

132

5.3.3 Analyses of Head Retrofit Designs

In addition to the baseline analyses with either an 11 gauge jacket or 0.5 inch head shield, 

analyses of retrofit head systems were performed with 0.680 and 0.8281 inch thick TC128B head 

shields. Force-deflection characteristics for unpressurized 500 lb tank heads with various head 

shields are compared to the baseline 600 lb head with 0.5 inch A516-70 head shield in Figure 

171. The comparison shows that the 0.8281 inch thick TC128B head shield significantly 

increases the puncture resistance of the head. For the 500 lb tank head, the increase from the 

0.50-inch head shield to the 0.8281 inch head shield produced greater than a 50% increase in 

puncture energy (1.14 to 1.84 million ft-lbs). The retrofit 500 lb head also has a higher puncture 

energy than the baseline 600 lb head with a 0.5 inch head shield. 

Figure 171. Calculated force-deflection behaviors for various unpressurized head designs.

The analyses of the unpressurized retrofit 300 lb head with a 0.8281 inch thick TC128B head 

shield is compared to the baseline 600 lb heads with 0.5 inch A516-70 head shields in Figure 

172. The comparison shows that the 300 lb tank head with the thick head shield is punctured 

with a lower peak impact force but sustains larger overall displacements before the head is 

punctured. As a result, the retrofit 300 lb head has a higher puncture energy than the baseline 

600 lb head with a 0.5 inch head shield (1.45 versus 1.98 million ft-lbs). 
Applied Research Associates, Inc.

2010

133

Figure 172. Comparison of unpressurized 600 lb and retrofit 300 lb tank head impacts.

Force-deflection characteristics for the pressurized 500 lb tank heads with various head shields 

are compared to the pressurized 600 lb head with 0.5 inch A516-70 head shield in Figure 173. 

The comparison shows that both the 0.680 inch and the 0.8281 inch TC128B head shields

significantly increases the puncture resistance of the pressurized head. 

For the 500 lb tank head, the increase from the 0.50-inch head shield to the 0.8281 inch head 

shield produced greater than a 50% increase in puncture energy (0.92 to 1.48 million ft-lbs). 

Similarly, the increase from the 0.50-inch head shield to the 0.680 inch head shield produced 

greater than a 30% increase in puncture energy (0.92 to 1.28 million ft-lbs). Both of the retrofit 

500 lb heads have higher puncture energies than the baseline 600 lb head with a 0.5-inch head 

shield (1.48 and 1.28 versus 1.11 million ft-lbs). 

The analyses of the pressurized retrofit 300 lb head with a 0.8281-inch-thick TC128B head 

shield are compared to the baseline 600 lb heads with 0.5-inch A516-70 head shields in Figure 

172. When the 100 psi internal pressure is added to the head systems the overall stiffness of the 

two designs is much closer. The pressurized head comparison shows that the 300 lb tank head 

with the thick head shield is punctured with both higher peak impact force and larger overall 

displacements before the head is punctured. As a result, the retrofit 300 lb head has a higher 

puncture energy than the baseline 600 lb head with a 0.5 inch head shield (1.54 versus 1.12 

million ft-lbs). 
Applied Research Associates, Inc.

2010

134

Figure 173. Calculated force-deflection behaviors for various pressurized head designs.

Figure 174. Comparison of pressurized 600 lb and retrofit 300 lb tank head impacts.
Applied Research Associates, Inc.

2010

135

A summary of the analyses of the unpressurized 500 lb tank head with the various head shield 

thicknesses is provided in Figure 175. The puncture energy levels calculated for the 600 lb head 

with the baseline jacket and head shield are also indicated in the figure. The comparison shows

that a retrofit head shield thickness of approximately 0.65 inch would be required for a 500 lb 

head to provide a similar level of puncture resistance as the 600 lb head with a 0.5 inch A516-70 

head shield. The conclusions are similar for both the unpressurized and pressurized heads as 

shown in Figure 175. 

Figure 175. The effects of head shield thickness on puncture energy for a 500 lb head.

A secondary effect, not shown in the figure, is the effect of the head shield material. The 11 

gauge (0.119 inch) head jacket was made of A1011 Steel, the 0.50 inch head shield was made of 

A516-70, and the 0.675 and 0.8281 inch head shield was made of TC128B. However, these 

material effects are believed to be small compared to the effect of the head shield thickness. 

Another effect illustrated in Figure 175 is that when a thicker retrofit head shield is supported by 

a thicker retrofit jacket over the tank shell (0.375-inch TC128B in this example), the head shield 

provides a greater level of puncture protection (calculations indicated by the green dots). 

5.3.4 Head Impact Analyses on Ethylene Oxide Tank Heads

All of the above head impact analyses were performed using the chlorine tank head geometry 

(2:1 ellipsoid with a 100 inch major diameter). A set of three additional head impact analyses 
Applied Research Associates, Inc.

2010

136

were performed on EO tank configurations to confirm that the response is consistent with those 

from the above analyses on Cl tank heads. The dimensions of the Cl and EO tank heads were 

previously provided in Table 3. A summary of these EO tank head impact analyses is provided 

in Table 8. The analyses include the baseline 300 and 500 lb EO heads, each with a 0.5 inch 

head shields and a 300 lb head retrofit with a 0.8281 TC128B Head Shield. The calculated 

force-deflection characteristics for the EO heads are shown in Figure 176. 

Table 8. Summary of pressurized EO head impact analyses

Calculation Head 

Thickness

Jacket or 

Head Shield Shell Jacket Impact 

Velocity

Puncture 

Force 

(lbs)

Puncture 

Energy

(ft-lbs)

R1R 0.6030 0.500” A572-50 11 gauge A1011 14 mph 929,000 1,116,000

R1S 0.6030 0.8281” TC128B 0.375” TC128B 14 mph 1,347,000 1,931,000

R1T 0.9784 0.500” A572-50 11 gauge A1011 14 mph 1,218,000 1,419,000

Figure 176. Calculated force-deflection behaviors for EO Tank head geometries.

The calculated puncture energy of the retrofit 300 lb EO tank head with the 0.8281 inch head 

shield was over 1.9 million ft-lbs and was approximately 35% higher than the puncture energy 

for the 500 lb EO head with the 0.5 inch head shield. Once again, a simple retrofit of adding 

additional thickness to the head shield was capable of providing a higher protection level than 

the tank with the higher test pressure head thickness. 
Applied Research Associates, Inc.

2010

137

The comparison of the relative impact responses of EO and Cl heads are provided in Figure 177

and Figure 178 for 500 lb and 300 lb configurations, respectively. The 500 lb comparison is 

difficult since a number of parameters were different. The EO tank head has a diameter of 119 

inches compared to 100 inches for the chlorine head. The 500 lb EO tank head has a thickness of 

0.9784 inch compared to 0.8281 inch for the chlorine head. Finally, the EO was pressurized to 

50 psi where analyses were performed for unpressurized and 100 psi pressure levels for the 

chlorine head. However, it was clear that the 500 lb EO tank head had a higher puncture force 

and puncture energy than the 500 lb chlorine head. 

Figure 177. Calculated force-deflection behaviors for 500 lb EO and Cl Tank heads.

The comparison of the 300 lb EO and Cl tank heads is shown in Figure 178. In the 300 lb head 

configurations, both tank heads have the same thickness (0.5625-inch). Thus, the only 

differences were the head diameters (119 vs. 100 inches) and the internal pressure (50 vs. 100 

psi). Again, for the 300 lb configuration, the EO tank head had a higher puncture energy than 

the chlorine head. However, in this case the head systems had equivalent thicknesses and as a 

result had nearly equivalent puncture forces. The additional energy resulted from the lower 

compliance of the EO head and correspondingly larger indentations before the puncture 

initiation. 
Applied Research Associates, Inc.

2010

138

Figure 178. Calculated force-deflection behaviors for 300 lb EO and Cl Tank heads.

5.3.5 Effects of Impactor Size

An important factor for assessing the impact conditions that will puncture a tank head is the size 

of the impacting object. The baseline impactor used in the analyses for this report is the square 

6x6 inch impactor. However, we want to assess other impactors to make sure that any protective 

concepts that are developed will be appropriate for general impact conditions. 

To evaluate the effects of the ram impactor size, analyses were performed on head geometries 

being impacted by both 9x9 inch and 12x12 inch ram impactors with a 0.50-inch radius around 

the edges. A summary of the impact analyses are shown in Table 9. The analyses include both 

the 500 lb and 600 lb chlorine tank head configurations with the 0.5 inch thick A572-50 full 

height head shield. A summary of the head impact analyses with the different impactor sizes is 

provided in Table 9. 

The force-deflection characteristics for the various 500 lb and 600 lb tank head impact analyses 

are provided in Figure 179 and Figure 180, respectively. All of these head impact analyses 

included a 0.50-inch-thick A572-50 head shield. The comparisons show that the force-deflection 

characteristics are mostly independent of the impactor size, up to the point of head shield 

puncture initiation. However, the load level at which the penetration is initiated is strongly 

influenced by the impactor size. 
Applied Research Associates, Inc.

2010

139

Table 9. Summary head impact analyses with various impactor sizes. (1)

Calculation Head 

Thickness

Jacket or Head 

Shield 

Impactor 

Size

Impact 

Velocity

Puncture 

Force 

(lbs)

Puncture 

Energy

(ft-lbs)

R1K 1.1360 0.50-inch A572-50 6 x 6 inch 14 mph 1,206,000 1,121,000

R1L 0.8281 0.50-inch A572-50 6 x 6 inch 10 mph 966,000 916,000

R1V 1.1360 0.50-inch A572-50 9 x 9 inch 25 mph 1,868,000 2,450,000

R1W 0.8281 0.50-inch A572-50 9 x 9 inch 25 mph 1,325,000 2,010,000

R0F 1.1360 0.50-inch A572-50 12 x 12 inch 25 mph 2,334,000 3,672,000

R0G 0.8281 0.50-inch A572-50 12 x 12 inch 25 mph 1,746,000 3,515,000

Notes: (1) All analyses performed on chlorine tank heads pressurized to 100 psi.

Figure 179. Impactor size effects on the 500 lb Cl tank head.
Applied Research Associates, Inc.

2010

140

Figure 180. Impactor size effects on the 500 lb Cl tank head.

The calculated puncture forces and puncture energies for both the 500 lb and 600 lb tank heads 

as a function of the ram size are summarized in Figure 181 and Figure 182, respectively. The 

puncture forces correlate closely to a linear function of the ram size. This again is consistent 

with the puncture initiation resulting from the local shear stresses at the perimeter of the ram 

contact face. The puncture energies also are close to a linear relationship to ram size over the 

range analyzed but the correlation does not go through the axis of the plot. As a result, the 

puncture energies are significantly increased when a larger impactor size is used. 

The calculated energy balance for the 12-inch impactor and 600 lb tank head is shown in Figure 

183. The initial impact energy is approximately 6.2 million ft-lbs (all in the kinetic energy). As 

the impact progresses, the ram is decelerated and the kinetic energy drops off. The energy 

transfer is from the kinetic energy of the ram to the internal energy of the tank head (plastic 

deformations of the tank material) and the pressure-volume work caused by the indentation 

reducing the total tank volume. The internal energy of the tank at rupture is approximately 3.8 

million ft-lbs and the pressure volume work is approximately 400,000 ft-lbs. The hourglass 

energy and sliding energy in the calculation are both negligible indicating that the calculation is 

stable and does not have any numerical energy losses.
Applied Research Associates, Inc.

2010

141

Figure 181. Summary of impactor size effects on the puncture forces for Cl tank heads.

Figure 182. Summary of impactor size effects on the puncture energies for Cl tank heads.
Applied Research Associates, Inc.

2010

142

Figure 183. Energy balance for the 12-inch impactor and 600 lb Cl tank head.

Based on the energy transfer to the pressure-volume (PV) work, the effects of the tank volume 

change can be assessed. The 400,000 ft-lbs PV work corresponds to a change in volume of 28

cubic feet (207 gallons) with a constant 100 psi (14400 psf) pressure assumption. Using a tank 

nominal capacity of 17,300 gallons, the change in volume is approximately 1.2%. Therefore, the 

change in volume (∆V) is much smaller than the original outage volume (approximately 10%)

and the constant pressure assumption of these analyses is a reasonable approximation. The 

hydrodynamic effects of the longitudinal sloshing of the lading in a head impact was not 

evaluated. 

5.3.6 Effects of Impact Speed

A potential factor for assessing the impact conditions that will puncture a tank head is the speed 

of the impacting object. As a reference, the lower bound of impact speed was assessed by 

performing a quasistatic loading of the tank car using the 6x6 inch impactor as a punch. The 

comparison of the dynamic impact response with the quasistatic behavior is shown in Figure 

184. The comparison shows that for this tightly constrained test condition there is little effect of 

the impact speed on the force deflection behavior or puncture energy. 
Applied Research Associates, Inc.

2010

143

Figure 184. Effect of impactor speed on the puncture of a 500 lb Cl tank head.

5.3.7 Summary of the Head Impact Analyses

The calculated head impact forces as a function of the combined head and head shield thickness 

are shown in Figure 185. The head impact analyses are added to the previous calculated forces 

for the side impacts, shown previously in Figure 144. The figure shows the analyses are mostly 

consistent with a linear relationship between puncture force and total thickness of the protective 

layers. However, the pressurized heads and the thicker head systems failed at forces that were on 

average a little lower than the linear fit to the side impact analyses. 

The proposed failure mechanism for the lower puncture forces is that, for the stiffer head 

systems, the offset impact creates a larger stress concentration along the upper edge of the 

impactor face and the failure initiates at that location at a lower total force. The more compliant

head systems allow for a larger dent to form and the impactor develops a more uniform stress 

distribution around the perimeter for the impact patch at the bottom of the indentation. 
Applied Research Associates, Inc.

2010

144

Figure 185. Calculated puncture forces as a function of system thickness.

The calculated puncture energies for all of the side and head impact analyses are plotted against 

the combined thickness of the system in Figure 186 (6x6 inch impactor only). When comparing 

all of the puncture energies, the data falls into four separate groups that are distinguished by the 

impact type (side or head) and the tank pressure. Again, the comparison indicates that the total 

thickness is the dominant parameter that determines puncture energy for a given impact 

condition and pressure level. This shows that a retrofit design with an increased jacket thickness 

should provide equivalent protection to thicker commodity tank systems with equivalent 

combined shell and jacket thicknesses. 
Applied Research Associates, Inc.

2010

145

Figure 186. Calculated puncture energies as a function of system thickness.

The above summary of the puncture energies from the various analyses can be applied to assess 

the system requirements for more severe impact scenarios. For example, we can assess the head 

or side protection requirements to resist a puncture by the ram car with the 6x6 inch ram head at 

an impact speed of 18 mph. The impact energy for the 295,000 lb ram car at 18 mph is slightly 

over 3 million ft-lbs. If we extrapolate the impact energy summary, as shown in Figure 187, we 

would predict that the 3 million ft-lb puncture energy requires approximately 2.4 inches of steel 

for an unpressurized head and 3.1 inches of steel for a pressurized head (100 psi). 

To check the validity of these estimates, a pair of analysis was performed for thicker head 

systems. The first analysis was on a pressurized head system that had a 1.6 inch head and 1.6 

inch head shield for a total thickness of 3.2 inches. The second analysis was on an unpressurized 

head system that had a 1.2 inch head and 1.2 inch head shield for a total thickness of 2.4 inches. 

For both analyses, the head and shield were supported by a side cylindrical shell sections that 

were one inch thick. The heads and shields were assumed to have the baseline TC128B 

constitutive and failure properties used for the majority of tank shells and head analyses in this 

report. 
Applied Research Associates, Inc.

2010

146

Figure 187. Extrapolated requirements for a 3 million ft-lb puncture energy.

The model and calculated impact behavior for the 18 mph impact on the pressurized head is 

shown in Figure 188. The head shield was punctured and fractures were initiated in the head, but 

a full penetration did not occur. The calculated damage in the impact zone is shown in Figure 

189. The calculated level of damage indicates that the 18 mph impact is very close to the 

puncture threshold for this response (material is failed at a damage magnitude of 1.0). 

The model and calculated impact behavior for the 18 mph impact on the unpressurized head is 

shown in Figure 190. The head and head shield were both punctured in this impact with only a 

very small fraction of residual kinetic energy at the time the head is punctured. The ram forward 

motion has been completely stopped in the deformed state shown in Figure 190. 
Applied Research Associates, Inc.

2010

147

 

Figure 188. Calculated 18 mph impact response for the thick pressurized head system.

Figure 189. Calculated Bao-Wierzbicki damage levels in the thick head impact zone.

The updated puncture energy summary including the results of these thick head system analyses 

is shown in Figure 191. These analyses agree well with the expected values based on the 

extrapolation of previous results on thinner systems. Similarly, the updated punch force 

summary including these analyses is shown in Figure 192. Again, the forces agree very well 

with the linear fit to the total system thickness. 
Applied Research Associates, Inc.

2010

148

 

Figure 190. Calculated 18 mph impact response for the thick unpressurized head system.

Figure 191. Calculated puncture energies as a function of system thickness.
Applied Research Associates, Inc.

2010

149

Figure 192. Calculated puncture forces as a function of system thickness.
Applied Research Associates, Inc.

2010

150

6 Detailed Evaluation of Alternate Protection Concepts

In the previous sections of this report, the tank geometries and protection systems evaluated were 

either designs currently in operation or simple variations on those designs where the thickness 

and or materials used for the tank, jacket, or head shields was varied. This section provides 

information on the development of advanced tank car protection strategies. The primary 

approach was to adapt crash, energy absorption, and impact protection technologies to develop 

tank car puncture protection systems that could be adapted on to an existing tank car. 

Technologies investigated included energy absorbing materials (e.g. crushable foams) and 

advanced structural designs (e.g. sandwich structures with various core designs). 

6.1 Evaluation of Foam Protection Systems

Structural foams have been applied to structures for crash safety and energy dissipation 

applications. They can easily be added to complex enclosed spaces and typically perform well in 

crushing when the foam remains confined [e.g. 67, 68]. 

This use of foams for the advanced tank car protection concept has several advantages. The tank 

car and confining jacket layers can be fabricated and assembled using the same facilities and 

equipment used for current tank car designs. The foams can be added to the tank cars relatively 

easily using available foaming equipment. Finally, the foam has beneficial properties for 

insulation and thermal protection of the tank, in addition to the energy absorbing characteristics. 

The primary foam used in this effort was a Dow BETAFOAMTM. This foam is classified as a 

rigid, all-water blown molded foam. BETAFOAMTM is designed to produce large, highly 

contoured, thin parts. In tank cars, the proposed concept was to add structural foams between the 

jacket and commodity tank. The foam would dissipate energy through crushing when impacted 

and could potentially act as a core material to couple the tank and jacket layers as a sandwich 

structure for improved performance. In some designs, multiple layers were used with two 

jackets applied to the tank and different density foams applied between the jackets and between 

the inner jacket and the tank. Variations in the thicknesses of the various layers and foam 

densities could be used to optimize the system. 

In the following sections, the testing and analysis on structural foams and foam protection 

concepts are described. These efforts include a preliminary set of characterization tests on a 

range of foam densities and strengths and associated preliminary analyses to develop foam 

protection system concepts. Subsequently, a more detailed set of material characterization tests 

and component tests were performed using both a low density foam and a high density foam. 

These foams had nominal densities of 5 and 18 pounds per cubic foot (pcf) and corresponding 

crush strengths of approximately 100 psi and 1000 psi, respectively. A more accurate foam 
Applied Research Associates, Inc.

2010

151

constitutive model was developed and validated against the updated characterization testing and 

used in subsequent analyses of the foam system performance. 

6.1.1 Validation of Foam Modeling

There are several constitutive models available in LS-DYNA that are appropriate for different 

types of foam materials. Many of these models are developed to include a strong dependence of 

the strength on the level of hydrostatic confining stress. This behavior is seen in some types of 

foam and is common to soils and geologic materials. Determination of the constitutive 

parameters for this type of model requires a series of triaxial material characterization tests as 

described in Reference 69. 

For much of this effort, the primary information available on the foam materials was a 

preliminary set of unconfined compression tests to measure the crush curves. As a result, a 

simple crushable foam constitutive model was used where the behavior was fit to the unconfined 

compression test only. The specific model was the Material Type 63 constitutive model

(*MAT_CRUSHABLE_FOAM or *MAT_63 in Reference 15) and the crush behaviors were 

obtained from the unconfined compression tests described below. 

As the evaluation of foam systems progressed, a more comprehensive series of foam 

characterization tests were performed to allow the constitutive parameters to be determined for 

more complex foam models. These tests include unconfined compression tests, uniaxial 

compression tests, and triaxial compression tests as well as tensile testing. These tests were 

performed on both low density (nominally 5 pcf) and high density (nominally 18 pcf) 

BETAFOAMTM late in this development effort. Results of these characterization tests are also 

presented in this section. 

Following the more comprehensive material characterization, some deficiencies were identified 

in the constitutive behaviors in the Material Type 63 model (referred to as Mat_63 in the 

remainder of the report). As a result, the Material Type 75 constitutive model 

(*MAT_BILKHU/DUBOIS_FOAM or *MAT_75 in Reference 15) was fit to the data and used 

in subsequent foam impact simulations. This model allows for a more accurate representation of 

the tensile behavior compared to the Mat_63 model and was considered to be appropriate since 

the characterization testing found that the BETAFOAMTM crush strength has a small dependence 

on the level of triaxial confinement. Results with both models are presented in the following 

sections. 

Preliminary Foam Compression Testing

The preliminary characterization of the foam materials was performed using a set of unconfined 

compression tests. These tests were performed at SwRI for a range of foam densities [70, 71]. 
Applied Research Associates, Inc.

2010

152

The specimen geometry utilized during testing was a solid cylinder approximately 1.3-1.4 inch in 

diameter and of varying heights depending upon the thickness of the original foam block 

(nominally 1- or 3-inch for the Dow BETAFOAMTM). A photograph of the unconfined 

compression test conditions for a tall (3-inch thick) specimen is shown in Figure 193. Tests were 

performed under laboratory air conditions with no environmental preconditioning. All of the 

tests were performed in a displacement control mode. 

Figure 193. Dual platen setup for unconfined compression testing of foam [70].

The local measurement of displacement was made by string pots on the loading platens. The 

displacement rate applied during testing was on the order of 0.5-inch/minute. Data was recorded

at a 2 Hz sampling rate for the load cell, LVDT on the actuator, and the string pots. A dualcondition criterion was utilized to terminate the test. The two criteria included a 90-95% total 

strain limit and a 5000 psi stress limit.

An example test on a short (1-inch thick) foam specimen is shown in Figure 194 [70]. The 

specimen is compacted to a large extent to measure the full compaction curve into the region 

where the foam approaches a fully compacted state. After the load is removed the specimen 

remains in a compacted state with little recovery as seen in Figure 194(d). 
Applied Research Associates, Inc.

2010

153

(a) 10 percent Compression

(b) 50 percent compression

(c) 90 percent compression

(d) Post-test configuration

Figure 194. Compression test of a short (1-inch height), 5 pcf foam specimen [70].

A similar test on the tall (3 inch thick) foam specimen is shown in Figure 195. Many of these 

specimens developed a buckling instability as a result of the more slender specimen geometry. 

As a result, the corresponding crush curves for these tests will exhibit structural effects instead of 

a uniform material constitutive response. As much as possible, the constitutive model 

development efforts attempted to use the data from tests where buckling of the specimen was not 

significant. 
Applied Research Associates, Inc.

2010

154

(a) 8 percent Compression

(b) 33 percent compression

(c) 50 percent compression

Figure 195. Compression test of a tall (3-inch height), 7.5 pcf foam specimen [70].

The crush curves measured in the unconfined compression tests on the different foam densities 

are shown in Figure 196. These crush curves are the primary data requirement needed for the 

Mat_63 model. The objective of the testing was to obtain a family of crush curves for various 

strength foams that could be applied in the design study for the foam protection system. Two 

primary modifications were required for use in the constitutive models. First, the behavior under 

high levels of compaction needed to be extrapolated. In the severe impact events, the foam will 
Applied Research Associates, Inc.

2010

155

be compressed to load levels significantly greater than those in the unconfined compression tests. 

Information about the foam solid density and the analyses of the foam sandwich panel punch test 

were used to develop and assess the extrapolated behavior. The extrapolated behavior is 

indicated by the dotted line extension sections of the curves in Figure 197. 

Figure 196. Unconfined compression curves for different density foams.

The second modification to the foam data required was to interpolate between some of the 

measured foam densities to obtain different desired crush strengths. These interpolated curves 

are indicated by the dashed lines in Figure 197. The interpolated data curves correspond to 

approximately 4.75 pcf and 22 pcf foam densities. The addition of these two interpolated foam 

densities results in a set of crush strengths that cover the range of interest and can be applied in 

different combinations to improve the puncture resistance. 

The Mat_63 constitutive models developed using the unconfined compression crush curves were 

then used to simulate the compression tests to validate that the model parameters were properly 

defined. An example simulation of an unconfined compression test for the 5.4 pcf foam is 

shown in Figure 198. The corresponding comparison of the fit to the test data and the simulation 

of the test are shown in Figure 199. The comparison shows that the model is capable of 

accurately reproducing the measured unconfined compression behavior for the foam. In 

addition, when the specimens were given a small initial imperfection, the buckling behaviors 

could also be reproduced as shown in Figure 200. 
Applied Research Associates, Inc.

2010

156

Figure 197. Unconfined compression curves for different density foams.

(a) Initial state (b) Fully compacted

Figure 198. Simulation of an unconfined compression test (5.4 pcf foam).
Applied Research Associates, Inc.

2010

157

Figure 199. Constitutive model fit and compression test simulation (5.4 pcf foam).

(a) Initial state (b) Buckled specimen

Figure 200. Simulation of the specimen buckling in unconfined compression tests.

Updated Foam Characterization Testing

To allow for a more comprehensive modeling and validation of the foam response, a detailed set 

of foam characterization tests were performed. These tests include unconfined compression 

tests, uniaxial compression tests, and triaxial compression tests as well as tensile testing. These 

tests were performed on both low density (nominally 5 pcf) and high density (nominally 18 pcf) 

BETAFOAMTM. Results of these characterization tests are presented in this section. 
Applied Research Associates, Inc.

2010

158

The new set of unconfined compression tests and uniaxial tension tests were performed by The 

Dow Chemical Company. The compaction curves from the unconfined compression tests on the 

5 pcf and 18 pcf foams are shown in Figure 201 and Figure 202, respectively. The test 

specimens were either 1.5 or 2.0 inches in diameter and 2.0 inches long. Both specimen 

geometries resulted in a uniform compression behavior (without significant lateral buckling) and 

the compaction curves all fall within a narrow band. From the series of 10 tests on the 5 pcf 

foam, the average density was 4.97 pcf and the crush strength is approximately 100 psi. From 

the series of 10 tests on the 18 pcf foam, the average density was 19.1 pcf and the crush strength 

is approximately 1000 psi. 

A comparison of the crush strength for the 18.3 pcf foam in the initial test series (SwRI data of 

Reference 70) is compared to the new Dow data in Figure 203. The crush curve in the initial test 

series has a significantly lower steady state crush strength. However, much of this discrepancy 

may be the result of the lateral buckling of the specimen in this test. 

Figure 201. New unconfined compression testing on the 5 pcf foam.
Applied Research Associates, Inc.

2010

159

Figure 202. New unconfined compression testing on the 18 pcf foam.

Figure 203. Comparison of unconfined compression test series on 18 pcf foams.
Applied Research Associates, Inc.

2010

160

The tensile tests were performed on dog bone specimens with gauge section cross sectional 

dimensions of approximately 0.36 by 0.12 inch (9.2 mm W X 3.0 mm) and tested at a rate of 

approximately 2 inches per minute). The 5 pcf foam did not have sufficient strength to support 

the attachment of the grips and loads imparted in the setup of the testing. The tensile tests on the

18 pcf foam are shown in Figure 204. The high density foam has a tensile strength of 

approximately 1200 psi and fail at a tensile strain of approximately 7-8%. A comparison of the 

uniaxial tension and unconfined compression tests for this foam is provided in Figure 205. The 

comparison shows that the tensile stresses are approximately 20% higher than the compressive 

stresses up to the point of tensile failure. 

The triaxial compression tests are commonly used to determine the strength envelope, or yield 

surface as it’s referred to in LS-DYNA, of materials like soil and foams. The testing system 

used for the triaxial foam characterization testing is shown in Figure 206 [69]. For each test, a 

cylindrical specimen of foam is first prepared inside a fluid-tight membrane to prevent 

infiltration of the confining fluid (air). In the triaxial apparatus, it is possible to apply two 

independently controlled components of load to the test specimen, as appropriate to each 

individual test. Pressurized fluid (air) in the vessel is used to impose a hydrostatic stress. The 

other component of load is derived from a piston, which extends through a seal in the top of the 

pressure vessel, loading the cylindrical specimen in the axial direction. Electronic 

instrumentation is used to measure both the applied loads and the resulting deformations of the 

foam specimens. 

Figure 204. Uniaxial tension test results for the 18 pcf foam.
Applied Research Associates, Inc.

2010

161

Figure 205. Comparison of the measured uniaxial stress behavior in tension and compression.

 

Figure 206. Schematic and photograph of the triaxial compression test apparatus.

The first series of tests performed are the basic triaxial tests on the foam specimens. In these 

tests, different levels of confining hydrostatic pressures are applied and then the specimen is 

subsequently compressed axially while the constant confining pressure is maintained. For the 
Applied Research Associates, Inc.

2010

162

high density foam (18 pcf), a series of hydrostatic tests was performed with confining pressures 

of 5, 10, 20, 50, 400, and 800 psi. 

The true stress difference in the triaxial tests on the 18 pcf foam specimens are plotted against 

the measured axial and radial strains in Figure 207. In these tests, the stress difference is the 

additional axial compressive load applied in addition to the confining hydrostatic pressure. From 

the figure, we see that the low levels of confinement (below 50 psi) have a negligible effect on 

the response and would be expected to be equivalent to the unconfined compression test. 

However, when we plot these low confinement tests against the results of the unconfined 

compression tests we see that the two test series have a discrepancy in the crush strength of 

approximately 20% as shown in Figure 208. The source of this discrepancy is not known.

An alternative approach to present the results of the triaxial tests is to add the hydrostatic stress 

to the stress difference and plot the total axial stress against the axial compressive strain as 

shown in Figure 209. The graphs of the triaxial testing show that the foam does exhibit some 

stiffening under an increasing confining pressure. However the sensitivity of the behavior to the 

triaxiality of the load is smaller than seen in many other types of foams. 

Figure 207. Triaxial test results for the 18 pcf foam.
Applied Research Associates, Inc.

2010

163

Figure 208. Comparison of unconfined and low confinement triaxial tests for 18 pcf foam.

Figure 209. Triaxial test results for the 18 pcf foam.
Applied Research Associates, Inc.

2010

164

The other foam characterization tests performed were a hydrostatic compression test and a 

uniaxial strain compression test. The hydrostatic and uniaxial strain compression tests are also 

conducted using the triaxial device. In the hydrostatic compression test, the cylindrical foam 

specimens are loaded only by fluid (air) pressure, without any piston loading. The stresses on 

the specimen are the same in all directions and there is no shear stress on any plane. This is 

referred to as the hydrostatic state of compression. The results of this test are used to define the 

volumetric deformation behavior of the material. When unloaded, the foam expands on a 

different path (slope). This unloading path yields the bulk unload modulus required for some 

constitutive models. LVDT measurements are used to record the axial and radial deformations 

(defining the volume strain). The data from the hydrostatic compression test on the 18 pcf foam

is shown in Figure 18. 

In the uniaxial strain test, the axial stress and confining pressure are applied in such a way that 

the specimen undergoes compressive axial strain with no strain in the radial direction. The 

uniaxial strain loading is accomplished with an automated loading control system using the 

radial deformation measurement as feedback in the control loop. If the radial strain increases, 

the confining pressure is increased to return the radial strain to zero. Because no radial strain is 

allowed in a uniaxial strain test, the axial strain is equal to the volumetric strain in the specimen. 

There is a difference between axial and radial stress, and hence shear stresses exist in the 

specimen. The results of the uniaxial compression tests on the 18 pcf foam is shown in Figure 

211. 

Figure 210. Hydrostatic compression test results for the 18 pcf foam.
Applied Research Associates, Inc.

2010

165

A similar set of material characterization tests were performed on the low density (5 pcf) foam. 

The triaxial compression tests were performed at 2, 5, 10, 20, and 50 psi confining pressures. 

The results of those triaxial tests in terms of the true stress difference and the true axial stress are 

shown in Figure 212 and Figure 213, respectively. Again, the triaxial tests on the low density 

foam shows that the foam crush behavior is relatively insensitive to a confining pressure. In 

addition, the comparison of the triaxial compression tests with the unconfined compression tests 

show that the two test series are in reasonable agreement on the measured crush strength within 

the scatter of behaviors seen between tests, as seen in Figure 214. Potentially the steady state 

crush strength is lower in the unconfined compression test, which is opposite of the variation 

observed in the test series on the high density foam (Figure 208). 

The final characterization tests performed on the low density BETAFOAMTM were the 

hydrostatic compression test and a uniaxial strain compression test. The results from the 

hydrostatic and uniaxial compression test on the 5 pcf foam are shown in Figure 215 and Figure 

216, respectively. 

Figure 211. Uniaxial strain test results for the 18 pcf foam. 
Applied Research Associates, Inc.

2010

166

Figure 212. Triaxial test results for the 5 pcf foam (stress difference).

Figure 213. Triaxial test results for the 5 pcf foam (axial Stress).
Applied Research Associates, Inc.

2010

167

Figure 214. Triaxial test results for the 5 pcf foam (axial Stress).

Figure 215. Hydrostatic compression test results for the 5 pcf foam.
Applied Research Associates, Inc.

2010

168

Figure 216. Uniaxial strain test results for the 5 pcf foam.

Foam Model Development and Validation

The updated foam characterization test data was initially applied to further evaluate and validate 

the Mat_63 constitutive model. The model constitutive parameters were fit to the new 

unconfined compression crush curves and the compression test was simulated. The comparison 

of the measured and calculated unconfined crush behaviors is provided in Figure 217. The 

comparison shows very good agreement showing that the crush curves were properly 

implemented into the constitutive model. 
Applied Research Associates, Inc.

2010

169

Figure 217. Calculated and measured unconfined compression behavior (18 pcf foam Mat_63).

The tensile tests on the 18 pcf foam were subsequently analyzed and a problem with the Mat_63 

model was immediately identified. When the specimen was pulled in tension, the model did not 

develop any stresses, even though a nonzero tensile cutoff strength was specified. This was not 

the expected behavior from the model. 

To further investigate the behavior of the Mat_63 model in tension, a series of analyses were 

performed where the model was first loaded into compression then the load direction was 

reversed and loaded into tension. The load deflection histories for two of these analyses are 

provided in Figure 218. The figure shows a behavior where the foam can support a tensile load 

equal to the magnitude of the maximum compressive load developed in the initial compressive 

phase of the cyclic loading. A physical mechanism that would produce this behavior was not 

clear and we believed this may be the result of a bug in the implementation of the Mat_63 model. 

However, LSTC confirmed that this is the behavior of this model and that there were not plans to 

modify the model behavior. They recommended that we try using the Material Type 75 model 

(referred to as Mat_75 in the remainder of the report) for our application. 
Applied Research Associates, Inc.

2010

170

Figure 218. Calculated cyclic compression-tension behavior (18 pcf foam Mat_63).

The Mat_75 model requires the specification of both a pressure yield stress versus volumetric 

strain curve and a uniaxial yield stress versus volumetric strain curve (e.g. crush curve). The 

uniaxial yield curve can be extended in both the tensile and compressive directions to fit 

different measured behaviors. This capability is a significant improvement over the capabilities 

of the Mat_63 model.

The Mat_75 constitutive parameters were fit to the updated foam data and the unconfined 

compression test was simulated. Initially, the uniaxial crush curve applied previously in the 

Mat_63 model was used for the compression uniaxial yield stress curve in the Mat_75 model. 

The calculated response for this model is compared to the test and Mat_63 calculated behavior in 

Figure 219. The comparison shows poor agreement of the Mat_75 model in the large 

deformation compaction response where the calculated stiffening of the foam material occurs 

prematurely. This occurs because the Mat_63 model is designed to use an engineering stressstrain crush curve as an input and the Mat_75 model is designed to use the true stress-strain 

crush curve as an input. 
Applied Research Associates, Inc.

2010

171

Figure 219. Calculated and measured unconfined compression behavior (18 pcf foam Mat_63).

The Mat_75 uniaxial yield curves were subsequently corrected to the true volumetric strain using 

the approach described in Section 3.2.1. The comparison of the engineering and true strain crush 

curves for the 18 pcf foam is provided in Figure 220. The true strain crush curve is then adapted 

by adding on the tensile test phase to create the uniaxial yield curve required for the Mat_75 

model, shown in Figure 221. The tensile behavior is that defined in the negative volume strain 

section of the yield curve. 

Also shown in Figure 221 is a pressure yield versus true volumetric strain curve required as a 

material model input. The comparison of the measured hydrostatic compression stress-strain 

curve with the compression curves obtained in the other characterization tests is shown in Figure 

222. In general, all of the curves are similar in magnitude. Potentially, the hydrostatic curve can 

be as much as 20 percent above the other curves in different regions of the crush response. To 

investigate the effect of this variability, a sensitivity study was performed using pressure yield 

curves that were scaled from the uniaxial yield curve by factors of 1.0 and 1.2. There was little 

difference in the calculated results for most of the analyses. However, in the triaxial 

compression tests (described below) the pressure curve equal to the uniaxial yield curve (100 

percent pressure yield model) provided a better correlation to the measured response than the 

analyses using the 120 percent pressure yield model. As a result, the 100 percent pressure yield 

model was used in the subsequent Mat_75 analyses of the component and large scale impact 

tests. 
Applied Research Associates, Inc.

2010

172

Figure 220. True stress correction of the crush curve (18 pcf foam).

Figure 221. True stress correction of the crush curve (18 pcf foam Mat_75).
Applied Research Associates, Inc.

2010

173

Figure 222. Comparison of measured foam crush curves from the various tests (18 pcf).

The first analysis performed with the corrected Mat_75 model was the unconfined compression 

test. The comparison of the measured and calculated crush curves is shown in Figure 223. After 

the correction of the Mat_75 material input parameters to the true volumetric strain, the 

constitutive model is able to accurately reproduce the unconfined compression response. 

Included in the figure are the analyses using both the 100 percent and 120 percent pressure yield 

curve models. The variation in the pressure yield curve strength has a small effect on the 

calculated unconfined compression response. 

The calculated response of a cylindrical foam specimen in tension using the Mat_75 model is 

shown in Figure 224. In the initial phase of the loading the specimen stretches uniformly. 

However, at approximately 7 percent strain, the damage localizes to a plane of elements and the 

load drops to a very low level. The corresponding calculated tensile stress-strain behavior is 

compared to the measured tensile behavior in Figure 225. The comparison shows good 

agreement in the measured and calculated tensile response. 
Applied Research Associates, Inc.

2010

174

Figure 223. Calculated and measured unconfined compression behavior (18 pcf foam Mat_75).

Figure 224. Calculated tensile behavior of a cylindrical foam specimen (18 pcf foam Mat_75).
Applied Research Associates, Inc.

2010

175

Figure 225. Comparison of measured and calculated tensile behavior (18 pcf foam Mat_75).

The triaxial compression tests were analyzed using the Mat_75 model with both the 100 percent 

and 120 percent pressure yield curve models. The comparison of the calculated and measure 

response for the 120 percent pressure yield curve is shown in Figure 226. The agreement 

between the measured and calculated response for the low and confinement tests (5 psi and 400 

psi) is pretty good. However, the calculated response for the high confinement (800 psi) test has 

almost no increase in the strength during the crush behavior (hardening in the crush curve) as 

seen in the test. The calculated response for the triaxial compression tests with the 120 percent 

pressure yield curve also had an unstable behavior and filtering of the calculated crush forces 

was required to produce the smooth crush curves seen in Figure 226. 

The comparison of the calculated and measure response for the 100 percent pressure yield curve 

is shown in Figure 227. The agreement between the measured and calculated response for the 

low and confinement tests is similar to that of the 120 percent pressure curve analyses. 

However, the calculated behavior for the high confinement (800 psi) test has improved and 

includes a hardening behavior in the crush curve as seen in the test. In addition, the calculated 

response for the triaxial compression tests with the 100 percent pressure yield curve were more 

stable and resulted in smoother calculated crush forces compared to the 120 percent pressure 

yield analyses. As a result of these differences, the 100 percent pressure curve was used for the 

remainder of the Mat_75 analyses. 
Applied Research Associates, Inc.

2010

176

Figure 226. Comparison of measured and calculated triaxial test behaviors (Mat_75 - 120 percent).

Figure 227. Comparison of measured and calculated triaxial test behaviors (Mat_75 - 100 percent).
Applied Research Associates, Inc.

2010

177

A similar set of analyses were performed for the 5 pcf foam. The analyses were able to achieve a 

similar level of correlation between the calculations and the tests. A representative comparison 

of the predicted and measured unconfined compression behaviors is shown in Figure 228. This 

model was applied for the analyses of the low density foams in the component tests described 

below. 

Figure 228. Calculated and measured unconfined compression behavior (5 pcf foam Mat_75).

Foam Laboratory Component Testing

In addition to the unconfined compression tests to characterize the material behavior, laboratory 

component tests were performed on foam sandwich panels to validate the foam models. These 

tests included a panel bend test and punch testing on the sandwich panel [34, 72-74]. 

Foam Sandwich Panel Punch Test 

The configuration for the foam sandwich panel punch test is shown in Figure 229 [34]. Test 

samples were placed in the specialized puncture fixture available in the SwRI Materials Test Lab 

and used previously for validation of the steel constitutive and damage models as described in 

Section 3.4.4 of this report. The punch was 1.5-inch diameter and the receiving hole bore (man 

hole cover) was 3-inch diameter. During testing, the actuator displacement was measured with a 

remote LVDT mounted on the actuator. Load was measured from the load cell on the 

servohydraulic test frame. Tests were performed in displacement control at an applied rate of 1 
Applied Research Associates, Inc.

2010

178

inch/minute. Data was recorded with a custom data acquisition system operating at 10 Hz. This 

typically provided total data files of anywhere from 500-2000 points.

Figure 229. Photographs of the foam sandwich panel puncture test.

The analyses of the foam panel punch test was performed using both the Mat_63 and Mat_75

models. A comparison of the measured and calculated (Mat_63) punch behaviors are shown in 

Figure 230. The comparison shows very similar deflections and deformation modes. The only 

discrepancy was the visible cracking on the foam seen in the test. This type of discrete tensile 

crack formation was not possible with the simple foam model applied in the analyses. 
Applied Research Associates, Inc.

2010

179

Figure 230. Measured and calculated foam sandwich panel punch test behavior (Mat_63).

The comparison of the measured and calculated force-deflection characteristics for the punch test

is shown in Figure 231. The load-displacement behavior exhibits two peaks, corresponding to 

penetration of the two steel layers. The peak load for this test can be compared to the load when 

a single layer of the HSLA 90 XF face sheet was tested. The panel results exhibit a 20% 

increase in peak load suggesting that the interaction of the foam and face sheets provide a 

combined response to resist puncture. However, this overall increase in puncture force levels 

was relatively modest. 

Two different Mat_63 analyses were performed with different tensile cutoff strengths. The 

comparison shows that a high tensile cutoff stress (1000 psi) results in an over prediction of the 

forces after the upper face sheet was penetrated. When the tensile cutoff stress was reduced to 

100 psi a much better agreement was obtained for the full force-deflection behavior in the punch 

test. The characterization testing performed subsequent to these analyses found that the 1000 psi 

tensile cutoff is much closer to the actual tensile strength but the improved correlation with the 

lower tensile cutoff may be an indication of the stress reduction after the formation of the 

discrete cracks in the test. In addition, the Mat_63 model was shown to have deficiencies in the 

modeling of the tensile behavior as described above. 
Applied Research Associates, Inc.

2010

180

Figure 231. Comparison of measured and calculated foam sandwich panel punch forces (Mat_63).

The comparison of the measured and calculated (Mat_75) punch behaviors are shown in Figure 

232. The comparison shows very similar deflections and deformation modes. Again the model 

does not show discrete tensile crack formation. However, the mat_75 model has the capability to 

reduce the tensile strength in regions of the foam where the tensile strain is exceeded. Thus the 

local reduction in tensile stresses around the punch can be modeled.

The comparison of the measured and calculated force-deflection characteristics for the punch test

is shown in Figure 233. The overall agreement for the analysis with the Mat_75 model is pretty 

good indicating that the improved modeling of the tensile behavior is helpful for simulating the 

punch test response. The slight discrepancy in the character of the second peak in the force 

deflection is probably a result of the symmetry conditions used in the model. By applying a 

quarter symmetry model in the analysis, the response forms a plug type failure of the second face 

sheet rather than the asymmetric flap failure observed in the test. 
Applied Research Associates, Inc.

2010

181

Figure 232. Measured and calculated foam sandwich panel punch test behavior (Mat_75).

Figure 233. Comparison of measured and calculated foam sandwich panel punch forces (Mat_75).
Applied Research Associates, Inc.

2010

182

Foam Sandwich Panel Flexure Testing 

Two sets of flexure tests were performed on sandwich panels. In the preliminary set of flexure 

tests, the specimens were fabricated by bonding face sheets to existing foam blocks [72, 73]. 

These preliminary tests were analyzed with the Mat63 model. The subsequent sandwich panel 

specimens were fabricated by expanding foam between face sheets with different surface 

treatment conditions [74]. These tests were analyzed using the Mat_75 model. 

In the preliminary test series, two foam sandwich panels were tested in a bend test fixture to 

evaluate the performance. The steel face sheets for the panels were nominally cut 17-inch long 

by 15-inch wide. The foam core was constructed from a foam block that was split on the 

thickness centerline in a band saw and after trimming it was essentially 15x15 inch (the steel 

pieces overhung on each end by 1-inch nominally). In one case, the foam was bonded on both 

surfaces and in the second case no adhesive was used. The adhesive used was 3M scotch-weld 

2216 B/A gray. The cure cycle that was applied was 120 minutes at 150° F. Foam surfaces that 

were sprayed with release agent during original foam fabrication (several months prior to testing) 

were sanded and cleaned with isopropyl alcohol prior to adhesive bonding. 

A loading approximating a three-point flexure was applied, with the two lower flexure supports 

being cylinders 3.5-inch in diameter. The upper loading point, or pusher, was a 5x1.5 inch 

pusher with a 0.5-inch radius around the contact face. The nominal span (outer-to-outer loading 

line distance) was 12 inches. The longitudinal axis of the test panel was located along the 12-

inch span as was the major axis of the 5x1.5 pusher. 

A 200-kip servohydraulic test machine was used for the flexure testing. The loading was applied 

with a 2-inch per minute displacement rate to the test specimen. A custom data acquisition 

system was connect to record data from the LVDT, load cell (200 kip capacity) and four other 

string pots also included in the setup. 

The measured load-displacement behaviors for both the bonded and unbonded sandwich panels 

are shown in Figure 237. Initially the bonded panel was much stiffer in flexure than the 

unbonded panel. However, within approximately the first one-half inch of the loading, the bond 

between the face sheets and foam core fails and the behavior approaches that of the unbonded 

panel. In addition, the load-displacement behaviors for both panels have numerous load dropoffs. These load drops likely correspond to where the foam cracked. 
Applied Research Associates, Inc.

2010

183

Figure 234. Comparison of bonded and unbonded foam sandwich panel bend tests.

A series of calculations were performed with the Mat_63 model to investigate the sandwich 

panel performance in the bend test. A comparison of the measured and calculated bend test 

response is shown in Figure 235. The variations evaluated in the analyses included the effect of 

a bonded or unbonded interface, the friction between foam and steel when the unbonded 

condition was used, the tensile cutoff strength for the foam, and modeling characteristics such as 

the contact algorithm between the various panel and test fixture components. Several of these 

parameters were found to have little effect on the behavior. The contact algorithms (e.g. 

automatic, surface to surface) had no effect on the behavior. The variation of the tensile cutoff 

strength had a small effect, even for very low values of tensile strength. However, the interface 

between the foam and the steel face sheets was found to be very important. 

The comparisons of the measured and calculated (Mat_63) force-deflection behaviors for the 

bonded sandwich panel flexure tests are shown Figure 236. The analyses show a similar 

behavior in the initial portion of the curves. However, the model does not include the failure of 

the bonded interface that occurs at approximately 0.4 inch. As a result the calculation was much 

stiffer than the test after the face sheets debond from the foam. 

A comparison of the calculated and measured force-deflection behaviors for the unbonded 

sandwich panel is shown in Figure 237. The calculated response was found to be sensitive to the 

friction coefficient at the foam-metal interface and the analysis shown was for a high friction 

coefficient of 0.8 which provided the best correlation to the measured response. These tests and 
Applied Research Associates, Inc.

2010

184

analyses show that the behavior of the foam-steel interface (both strength and interaction after 

debonding) will be important to optimize the performance of the foam based tank protection 

systems. 

Figure 235. Comparison of a foam sandwich panel bend test and analysis (Mat_63).

Figure 236. Measured and calculated response for the bonded foam sandwich panel (Mat_63).
Applied Research Associates, Inc.

2010

185

Figure 237. Measured and calculated response for the unbonded foam sandwich panel (Mat_63).

In the second set of panel flexure tests, the foams evaluated included both 5 pcf and 18 pcf 

BETAFOAMTM. The steel face sheet surface preparations investigated included untreated, sandblasted, coated, and sand-blasted and coated. Each foam and surface preparation combination 

was repeated three times to establish the variability in the performance, with a total of 24 tests 

conducted. The repeatability within test groups was good both in terms of the load displacement 

behavior and the magnitude of the maximum loads. 

Panels were fabricated by The Dow Chemical Company and were supplied in the ready-to-test 

condition. The test conditions were the same as those used in the preliminary panel flexure test 

series. The specimens were loaded with the constant displacement rate of 2 inches per minute.

The measured force-deflection curves from representative tests on the 18 pcf and 5 pcf panels 

with the various surface treatments are shown in Figure 238 and Figure 239, respectively. As 

expected, the 18 pcf foam panels all sustained higher loads than the 5 pcf panels. Within the 18

pcf group, the non-treated and sand-blasted conditions displayed significantly higher loads than 

the other two coated surface preparation conditions. The peak load in these tests is two to three 

times higher than that of the coated surface specimens and the load does not drop off until a ram 

displacement of approximately one inch. This failure mode was significantly different from 

other conditions and was marked by a loud audible ‘crack’ as the foam layer failed. These tests 

show that the foam-steel interface conditions can have a large effect on the panel (or tank car 

system) response. In post-test specimen inspections, a notable separation between the foam and 
Applied Research Associates, Inc.

2010

186

steel layers was found for the coated and sand-blasted with coating conditions for both the 5-pcf 

and 18-pcf foams, indicating a lack of adhesion.

Figure 238. Measured behaviors for the new 18 pcf foam sandwich panel tests.

The results from the preliminary panel flexure tests are compared to the second test series on the 

18 pcf foam specimens in Figure 240. The preliminary tests agree reasonably well wit the tests 

using the coated face sheet surface treatments. 

The comparisons of the Mat_75 calculated and measured panel flexure behaviors for the 18 pcf 

and 5 pcf foam panel specimens are shown in Figure 241 and Figure 242, respectively. In the 

analyses, the foam is perfectly bonded to the face sheets and any delamination failure would be 

modeled as a failure through the foam elements adjacent to the face sheets. The comparisons 

show that the correlation is best in the early portion of the response before the delamination is 

initiated in the specimens with good bond characteristics. Using this modeling approach with a 

perfect bond would provide an upper bound to the strength and energy dissipation that could 

potentially be achieved with further development of the interface bonding methodologies. 
Applied Research Associates, Inc.

2010

187

Figure 239. Measured behaviors for the new 5 pcf foam sandwich panel tests.

Figure 240. Comparison of the new and original foam sandwich panel tests.
Applied Research Associates, Inc.

2010

188

Figure 241. Measured and calculated behaviors for the new 18 pcf foam sandwich panel tests 

(Mat_75).

Figure 242. Measured and calculated behaviors for the new 5 pcf foam sandwich panel tests 

(Mat_75).
Applied Research Associates, Inc.

2010

189

6.1.2 Head Test 2 Measurements and Analyses (Foam System)

The second component test of a tank car head test article was performed as part of the Next 

Generation Rail Tank Car project on April 30, 2008, at the Transportation Technology Center, 

Inc. (TTCI) in Pueblo, Colorado [75]. This test involved a 295,200-pound ram car striking the 

Option 2 prototype design tank head at 16.9 mph. The ram car was equipped with a 6-inch 

square indenter with 0.50-inch radius around the edges. The tank head was affixed to the 

Transportation Technology Center impact barrier with a rigid fixture. The indenter punctured 

the outer head shield, inner jacket, and the tank head. The ram car was then slowed to a stop by 

the arrestor system mounted on the barrier. The peak longitudinal acceleration before the 

arrestor system engaged was about 3.9 g (CFC60). 

The test article and the ram car staged in position prior to the test are shown in Figure 243. The 

test article was bolted to the wall-offset fixture. The wall-offset fixture was bolted to the impact 

wall and an energy absorbing arrestor system was added to dissipate residual energy if the head 

is punctured. The arrestor system consists of four hydraulic cushion units, two on each side, 

placed to engage the ram car shortly after the indenter penetrates the head. 

Figure 243. Pretest configuration of the test head and ram car.

The Option 2 prototype consisted of three steel layers enclosing two foam layers. The outer 

layer was a 0.50-inch A516-70 head shield that was 127 11/16 inches in diameter. The inner 

jacket was an 11-gage A1011 jacket head that was 121 5/8 inches in diameter. The inner layer 
Applied Research Associates, Inc.

2010

190

was a 0.777 inch TC128 commodity head that was 102 inches in diameter. The inner foam layer 

was about 10-inches thick at the apex of the head and consisted of 4 lb/ft3 foam. The outer foam 

layer was about 2-inches thick at the apex of the head and consisted of 18 lb/ft3 foam. 

The ram car was modified before this test in anticipation of higher peak loads expected for the 

new prototype head designs. Support was added behind the ram to resist bending, the ram was 

strengthened, and a new hardened indenter was installed.

Prior to head test 2, the ram car was weighed to determine the increase from modifications. The 

A-end weight (ram end) was 168,050 lbs and the B-end weight was 127,150 lbs for a total 

weight of 295,200 lbs. The impact speed, measured by the reflector speed trap, was 16.9 mph. 

The ram punctured the outer jacket, inner jacket, and tank head. The damage to the outer jacket 

after the impact is shown in Figure 244. An interior view of the impact damage to the tank head 

is shown in Figure 245. 

Figure 244. Puncture of the external protective layers in Head Test 2.
Applied Research Associates, Inc.

2010

191

Figure 245. Internal puncture of the tank head in Head Test 2.

Head Test 2 was performed prior to the detailed characterization testing of the foams and the 

pretest analyses were performed using the Mat_63 foam models. A comparison of the predicted 

and measured force-deflection behaviors for Head Test 2 is shown in Figure 246. The pretest 

prediction did a very good job of predicting the initial behavior and the puncture of the outer 

head shield which occurs at a force level of approximately 500,000 lbs and a ram displacement 

of approximately 10 inches. However, in the subsequent loading, the pretest prediction over 

predicts the force levels by 200,000 to 300,000 lbs. 
Applied Research Associates, Inc.

2010

192

Figure 246. Comparison of predicted and measured behavior for Head Test 2.

The behavior measured in Head Test 2 can be better understood by comparing to the impact 

behavior for Head Test 1, as shown in Figure 247. We see that the later portion of the curve was 

nearly identical to that of the baseline head. As a result, we know that the foam had either 

significantly debonded from the head or had developed tensile cracks and was not resisting any 

loads. In contrast, the pretest prediction was performed using a perfect bond between the foam 

and the metal layers. In addition, the prediction was performed with a high estimate of the 

tensile cutoff stress. As a result, the model of the layered foam protective system holds together 

and continues to work as a system after the outer head shield layer was penetrated in the pretest 

analysis. 
Applied Research Associates, Inc.

2010

193

Figure 247. Comparison of predicted and measured behavior for Head Test 2.

The results of this test produced the following observations:

1. The models of the foam and detailed puncture models of the steel can accurately predict 

the penetration of the outer layers.

2. With the current formulation of the foam and foam-metal interface bonding, the system 

contributes very little to the puncture resistance after the outer layers were penetrated. 

Additional development of the foam and interface is needed to improve the performance.

Subsequent to the detailed foam characterization testing, a post test analysis was performed using 

the Mat_75 foam models. The other significant change to the post test analysis was that a BaoWierzbicki failure patch was added to the head to calculate the point at which the head is 

punctured. The comparison of the calculated and measured force-deflection behaviors is shown 

in Figure 248. Again there is overall good agreement in the puncture of the outer foam system 

protective layers and the forces at which the head is punctured. The primary difference is that 

the calculated forces begin to increase more quickly after the puncture of the outer layers. 

Again, this potentially is a result of large scale debonding at the foam-steel interfaces. 

The calculated and measured force-deflection behaviors are integrated to obtain the puncture 

energies shown in Figure 249. The comparison shows that the calculated puncture energy is 

approximately 12 percent higher than measured in Head Test 2. 
Applied Research Associates, Inc.

2010

194

Figure 248. Comparison of calculated and measured behavior for Head Test 2 (Mat_75).

Figure 249. Comparison of calculated and measured behavior for Head Test 2 (Mat_75).
Applied Research Associates, Inc.

2010

195

For this head impact test geometry, the influence of the foam modeling is seen primarily in the 

early time response. The later time forces are dominated by the internal tank head response. A 

comparison of the analyses with the Mat_63 and Mat_75 models is shown in Figure 250. The 

calculated response with the Mat_63 model is stiffer in the early indentation of the head shield. 

This probably results from the fact that tensile failures are not properly modeled. In addition, the 

behavior of the Mat_63 model after the head shield transitions directly to the loading of the 

internal head with no indications of the contribution of the inner head shield visible in the 

calculated force-deflection response. The calculated response with the Mat_75 model more 

closely matches the characteristics of the measured early time behavior. 

Figure 250. Comparison of calculated head impact response with different foam models.

6.1.3 Side Impact Puncture Analyses of Foam Systems

A series of additional analyses were performed to investigate the influence of different foam 

system design parameters on the protection performance. These analyses were performed prior 

to the detailed foam characterization and as a result were performed with the Mat_63 foam 

models. 

An example of the difficulties of developing structural systems to protect against the 6x6 inch 

impactor can be demonstrated by adding low density structural foam between the jacket and tank 

in the 300 lb EO car with the thicker outer jacket. A 3.3 pound per cubic foot (pcf) structural 
Applied Research Associates, Inc.

2010

196

foam was added to the 4 inch space between the tank and jacket. The system concept was to 

couple the shell and jacket layers as a sandwich structure and to dissipate some of the impact

energy with the crushing of the structural foam. The crush strength for the 3.3 pcf foam was 

roughly 50 psi [70]. 

The force-deflection curve and puncture energy for this concept using structural foam is shown 

in Figure 251. Without the foam the puncture force was 780,000 lbs and the puncture energy 

was 1.85 million ft lbs. The addition of the foam causes the punch to penetrate the outer jacket 

at approximately 15 inches displacement and the commodity tank was punctured at 

approximately 33 inches. Without the foam the 0.375-inch jacket was not punctured until 36 

inches of ram displacement and the tank ruptures at 43 inches. As a result, the overall puncture 

energy in the commodity tank drops from 1.85 million ft lbs to 1.27 million ft-lbs with the 

addition of the foam. The early failure of the outer jacket appears to result from the foam 

increasing the overall stiffness of the outer jacket behavior and concentrating the localized loads 

around the perimeter of the ram head. 

Figure 251. Evaluation of a retrofit EO tank car with a 0.375-inch jacket and foam.

Another example of a foam system side impact analysis was performed with the 12x12 inch 

impactor to see if the larger contact area improves the performance of foam systems. An 

analysis was performed for a 500 lb Cl tank with a 0.375 inch jacket and 4 inches of 3.3 pcf foam 

impacted by the 12 inch impactor. The calculated impact response is shown in Figure 252. The 
Applied Research Associates, Inc.

2010

197

corresponding force-deflection characteristics for this analysis are compared to other tank 

systems in Figure 253. The comparison shows that the system performs better than the baseline 

tank with an 11 gauge jacket. However, again the foam results in an earlier jacket penetration 

and lower puncture energies than the system retrofit with the 0.375 inch jacket but without any 

foam. 
Applied Research Associates, Inc.

2010

198

Figure 252. Impact response of a foam retrofit tank with the large ram (12” by 12”).
Applied Research Associates, Inc.

2010

199

Figure 253. Evaluation of the effects of the ram size on a 500 lb chlorine tank.

6.1.4 Head Impact Puncture Analyses of Foam Systems

A series of additional layered foam head protection combinations were analyzed. The 

combinations of jacket thicknesses, materials, and foam densities analyzed are summarized in 

Table 10. These analyses were primarily performed using foam models with high estimates of 

the tensile cutoff stress and good bonding between the layers. As a result, the primary 

information developed from these analyses was the response up to the point of the outer head 

shield penetration. 

Table 10. Multilayer foam protection system analyses

Calculation Head 

Thickness Inner Foam Inner Head 

Jacket/Shield Outer Foam Outer Head 

Shield 

R2K 0.8281 10” 4.75 pcf 11 gauge A1011 2” 18.3 pcf 0.50” A516-70

R2N 0.8281 10” 4.75 pcf 11 gauge A1011 2” 18.3 pcf 0.50” A516-70

R2P 0.8281 10” 4.75 pcf 0.25” 90XF 2” 3.30 pcf 0.50” 90XF

R2Q 0.8281 10” 3.30 pcf 0.25” 90XF 2” 7.50 pcf 0.75” 90XF

R2S 0.8281 10” 3.30 pcf 0.25” TC128B 2” 7.50 pcf 0.75” TC128B
Applied Research Associates, Inc.

2010

200

An example of the calculated impact behavior of a foam protection system is shown in Figure 

254. The initial response was as desired for the system where the outer high density foam layer 

and the inner jacket and head shield act together as a sandwich structure. As the outer layers 

were driven inward by the ram, a larger volume of the low density inner foam was crushed. 

However, at relatively small displacements the outer layer was penetrated and the protection 

system was defeated. 

Figure 254. Impact behavior of a multi-layer foam head protection system (R2S).

A series of force-deflection characteristics for various foam head protection systems is shown in 

Figure 255. Note that the detailed puncture modeling was only included in the outer layer so the 

large forces developed at late times by the resistance of the head were not realistic. However, 

the initial puncture of the outer head shield as indicated in the figure should be accurately 

captured. 

The calculated impact forces to puncture the outer layer were compared to those of previous 

analyses in the plot of the puncture force versus system thickness in Figure 256. The difference 

in the data for the foam systems was plotted against the outer head shield thickness only rather 

than the sum of the head and jacket thicknesses. This shows that the foam had the effect of 

making each layer act independently rather than coming in contact and working together to resist 

the high shear stresses around the edge of the impactor. As a result, the maximum achievable 

puncture forces were reduced and the system does not reach the desired puncture energies. 
Applied Research Associates, Inc.

2010

201

Figure 255. Impact behavior of multi-layer foam head protection systems.

Figure 256. Comparison of foam protection puncture forces to baseline systems.
Applied Research Associates, Inc.

2010

202

6.2 Evaluation of EMS Protection Systems

Another proposed protection concept is to use external engineered metal structures (EMS) to 

protect the tank structure and dissipate the impact energy. These EMS concepts have a wide 

range of potential geometries and fabrication techniques. However, the primary concepts 

analyzed here were structures with inner and outer face sheets and a metal core structure that 

creates a metal sandwich structure. The core should maintain the spacing between the face 

sheets to improve the efficiency and stiffness of the EMS system and at sufficiently high loads, 

deform in a mode to dissipate energy. 

There are a large variety of core structures including truss core geometries [76], hexagonal, 

rectangular, and triangular cellular cores [77-79], corrugated geometries, and metallic foam cores 

[80]. These structures were subjected to a range of loading conditions from static loading to 

impact and blast loading. In general, the testing and analyses concentrated on loads distributed 

over an area that is large compared to the characteristic length of the cellular core or core 

thickness. One study investigated cellular sandwich panels subjected to both distributed and

localized blast loads. The study found that a stiff core panel is superior to a solid plate when 

subjected to a planar impulse, but inferior when the load was localized [81]. The performance 

under a localized loading is very important for the puncture performance for a tank subjected to a 

6-inch impactor. 

The two primary core structures investigated here were a corrugated core sheet and an egg-crate 

design where an orthogonal grid of webs, normal to the face sheets, creates square cell geometry. 

6.2.1 Head Impact Puncture Analyses of a Corrugated EMS Systems

A series of analyses were performed on a corrugated head protection EMS system. The 

nomenclature for the corrugated EMS structures is shown in Figure 257. The primary EMS 

structure is a 6 inch deep corrugated structure with the face sheets and corrugated core sheet all 

made of 11 gauge 90 XF material. A detail of the end EMS corrugation geometry is shown in 

Figure 258. A drawing of the initial proposed EMS structure for Head Test 3 is shown in Figure 

259. An objective of these analyses was to further develop this head protection configuration 

prior to Head Test 3 using the corrugated EMS structures as the structural foundation. 
Applied Research Associates, Inc.

2010

203

Figure 257. Nomenclature for corrugated EMS structures.

Figure 258. Detail of the corrugated EMS corrugation geometry.
Applied Research Associates, Inc.

2010

204

Figure 259. Drawing of the proposed corrugated EMS structure for Head Test 3.

The baseline concept was to create a cylindrical side EMS structure and flat end EMS with a 

thicker cover plate to resist puncture. An example of the calculated behavior for this concept 

with a 0.375-inch thick 90 XF end cover plate is shown in Figure 260. The system was

penetrated with approximately 5 inches of indentation caused by the ram with an impact force of 
Applied Research Associates, Inc.

2010

205

586,000 lbs. To improve the performance of this system, the end plate needs to be significantly 

thicker and/or the overall compliance of the system needs to be increased. 

The proposed modification to the system is to extend the outer face sheet on the side EMS 

structure and separate the end EMS structures from the stiff side EMS. The calculated 

performance of this concept with an 8 inch extension of the side face sheet is shown in Figure 

261. This system still uses a 0.375-inch 90 XF end cover plate. The added extension reduces the 

system compliance and some crushing of the side extension was observed before the end EMS

was penetrated. 

Figure 260. Calculated response of the baseline corrugated EMS system (R3A).
Applied Research Associates, Inc.

2010

206

Figure 261. Calculated response of the corrugated EMS with an 8 inch standoff (R3D).

The second proposed modification to the system was to increase the thickness of the end cover 

plate to allow for larger impact forces and greater energy dissipation prior to the puncture of the 

protective structures. The calculated performance of this concept, with the 8 inch extension of 

the side face sheet and 0.777 inch thick TC128B cover plate, is shown in Figure 262. The 

addition of the end cover plate significantly improves the performance and allows for 

significantly more deflection and impact energy dissipation before the system was punctured. 

Figure 262. Calculated response of the corrugated EMS and thick cover plate (R3E).
Applied Research Associates, Inc.

2010

207

The calculated force-deflection behaviors of these three EMS systems are shown in Figure 263. 

The initial system without an end extension was punctured at a displacement of approximately 5 

inches and a force of 587,000 lbs. Adding the end extension of 8 inches does not significantly 

change the peak force the system resists, however the collapse of the extension allows for 

approximately 12 inches of ram travel before the end EMS was penetrated. The thicker end plate 

(0.777 TC128B) response produces approximately two feet of deflection and a force of 864,000 

lbs before the EMS system was penetrated. Thus, the combined effects of increasing the system 

compliance and increasing the load capacity of the outer layers results in a significant increase in 

the energy dissipation capability. 

Figure 263. Head impact analyses on corrugated EMS end structures.

A summary of the corrugated EMS head protection design analyses is given in Table 11. All of 

the corrugated EMS structures were fabricated from the 90XF steel that was 11 gauge thick in 

both the face and core sheets. The variations in the design were the end plates added at the outer 

surface of the system to resist the initial system puncture and the standoff distance added 

between the end EMS section and the side cylindrical curved EMS structures. In all cases, the 

extended end EMS was supported by an extension of the side EMS outer 11 gauge face sheet. 
Applied Research Associates, Inc.

2010

208

Table 11. Summary corrugated EMS head impact analyses

Calculation End Plate Thickness 

& Material

End EMS 

Standoff

EMS Puncture 

Force (lbs)

EMS Puncture 

Energy (ft-lbs)

R3A 3/8 in 90 XF 0.0 Inch 587,000 130,000

R3B 15/32 in TC128B 0.0 Inch 625,000 191,000

R3C 3/8 in 90 XF 8 Inches 515,000 433000

R3D 15/32 in TC128B 8 Inches 610,000 495,000

R3E 0.777 in TC128B 8 Inches 861,000 1,068,000

R3F 0.777 in TC128B 12 Inches 953,000 1,425,000

R3G 0.777 in TC128B 12 Inches 1,059,000 1,313,000

R3H 2X 3/8 in 90 XF 12 Inches 941,000 1,410,000

R3I 2X 3/8 in 90 XF 12 Inches 944,000 1,395,000

R3J 0.777 in TC128B 12 Inches 964,000 1,410,000

The system can be improved by further extending the end EMS to 12 inches of separation. The 

calculated response for this system is shown in Figure 264. A comparison of the force-deflection

characteristics for the 8 inch and 12 inch extensions with the 0.777 inch end plate are shown in 

Figure 265. By increasing the standoff the end structure undergoes a larger deflection before 

penetration and dissipates correspondingly larger impact energy. This 12-inch extension is the 

configuration selected for the EMS structure to be used in Head Test 3. 

A final comparison of the original and modified EMS system performance is shown in Figure 

266. The original EMS design was punctured at a displacement of approximately 5 inches and 

force of 600,000 lbs. By reinforcing the end plate and extending the end EMS structure by 12 

inches, the EMS system now has a displacement of approximately 30 inches and a force of 

approximately 1 million pounds before the EMS structure is penetrated. As a result, the EMS 

puncture energy is increased from 130,000 ft-lbs to over 1.4 million ft-lbs. 

After selecting the EMS configuration for Head Test 3, additional analyses were performed as 

pretest predictions. The primary difference in these analyses is that a patch of refined solid 

elements was added to the head impact zone (BW failure patch) to allow for the detailed 

assessment of both the EMS and head puncture levels. The calculated impact behaviors at both 

15 and 18 mph impact speeds are shown in Figure 267 and Figure 268, respectively. At 15 mph 

the impactor is arrested before the head puncture is initiated. At 18 mph, the head is also 

punctured with negligible residual kinetic energy. 
Applied Research Associates, Inc.

2010

209

 

 

Figure 264. Calculated response of the corrugated EMS with a 12 inch standoff (R3F).
Applied Research Associates, Inc.

2010

210

Figure 265. Performance of corrugated EMS end structures with various standoffs.

Figure 266. Performance of corrugated EMS end structures with various standoffs.
Applied Research Associates, Inc.

2010

211

 

(a) EMS puncture state (b) Final deformation state

Figure 267. Predicted Head Test 3 impact response at 15 MPH.

 

(a) EMS puncture initiation state (b) Head puncture state

Figure 268. Predicted Head Test 3 impact response at 18 MPH.
Applied Research Associates, Inc.

2010

212

A comparison of the predicted force-deflection characteristics and energy dissipation for the 

Head Test 3 configuration at 15 and 18 mph are shown in Figure 269. The two impact behaviors 

are very similar for displacements up to approximately 40 inches where the 15 mph impact 

response is complete. In both calculations the EMS structure is punctured at a displacement of 

approximately 30 inches and a peak load of approximately one million pounds. This also 

corresponds closely to the time that the inner EMS structure engaged the tank head. In the 18 

mph impact response, the ram undergoes approximately an additional 20 inches (50 inches total) 

before the head is punctured. 

Figure 269. Performance of corrugated EMS end structures with various standoffs.

6.2.2 Head Test 3 Measurements and Analyses (EMS System)

The third component test of a tank car head test article was performed as part of the Next 

Generation Rail Tank Car project on July 30, 2008, at the Transportation Technology Center, 

Inc. (TTCI) in Pueblo, Colorado [75]. This test involved a 296,225-pound ram car striking the 

Option 3 prototype design tank head and EMS protection structures at 18.2 mph. The ram car 

was equipped with a 6-inch square indenter with 0.50-inch radius around the edges. The tank 

head was affixed to the Transportation Technology Center impact barrier with a rigid fixture. 

The test article in position prior to the test is shown in Figure 270. The test article was bolted to 

the wall-offset fixture. The wall-offset fixture was bolted to the impact wall and an energy 
Applied Research Associates, Inc.

2010

213

absorbing arrestor system was added to dissipate residual energy if the head is punctured. The 

arrestor system consists of four hydraulic cushion units, two on each side, placed to engage the 

ram car shortly after the indenter penetrates the head. 

Figure 270. Pretest configuration of the test head and corrugated EMS system.

The Option 3 prototype consisted of a 6-inch corrugated EMS structure with a 0.777 inch thick 

TC128B protective end plate. The EMS structure was fabricated from 11 gauge 90 XF material 

with the geometry shown in Figure 258. A 12-inch extension of the outer jacket was added 

between the side and end EMS structures to incorporate a lower strength crush zone and improve 

the overall energy dissipation. 

The ram car was again modified before this test to allow for the testing with the larger protective 

EMS system. The energy absorbing buffers in the test fixture were repositioned to fit the larger 

external diameter of the EMS and the ram car end structures were reinforced to resist the new 

buffer load application points. The modified ram car prior to the test is shown in Figure 271 and 

the weight of the ram car was 296,225 lbs. 

The target test speed was 18 mph. The impact speed, measured by the reflector speed trap, was 

18.2 mph. The ram punctured the EMS head protective structures and the inner 500 lb tank head. 

The damage to the outer EMS structure after the impact is shown in Figure 272. An interior 

view of the impact damage to the tank head is shown in Figure 273. Both the EMS structures 

and head were fully penetrated by the ram head. 
Applied Research Associates, Inc.

2010

214

Figure 271. Pretest configuration of the test head and corrugated EMS system.

Figure 272. Impact damage to the external EMS protection system in Head Test 3.
Applied Research Associates, Inc.

2010

215

Figure 273. Internal puncture of the tank head in Head Test 3.

A comparison of the predicted and measured force-deflection behaviors for Head Test 3 is shown 

in Figure 274. The pretest prediction did a good job of predicting the characteristics of the forcedeflection curve and the displacement levels at which the EMS structures and the head are 

punctured. However, the measured impact forces are on average approximately 30 percent lower 

than predicted by the pretest analysis. 

The discrepancies between the pretest analysis and the test are indicated by the inspection of the 

post-test damage to the structures. After the test, the end EMS corrugation was fully separated 

from the outer EMS face sheet as seen in Figure 275. Post-test inspection of the corrugation 

welds indicated that the separation initiated near the impact location and the weld fractures 

subsequently propagated outward to the edge of the end EMS structure. The result of this 

debonding would be to lower the forces and energy dissipation from the EMS deformations with 

significant reductions in the extent of core buckling as seen in Figure 276. 

Another damage mechanism not expected in the pretest analysis was a fracture in the outer EMS 

end plate (0.777-inch TC128B) that extends from the impact point downward to the edge of the 

plate. The post-test fracture of the EMS end plate is seen in Figure 277. Post test inspection 

determined that this fracture initiated via a shear crack and propagated as a Mode I tensile 

fracture [82]. Thus, the crack appeared to occur after the 6-inch impactor had pierced the 0.777-

inch end plate and was opened up by the wedging effect of the tapered impactor. Shear lips on 
Applied Research Associates, Inc.

2010

216

the Mode I fracture indicate some ductility. Chevron markings on the fracture surface indicate a 

fracture in the transition region of the Charpy energy curve. This fracture would lower the post 

penetration resistance and energy dissipation predicted for the tapered ram head. It was 

subsequently identified that the TC128B end plate was not normalized. However, the potential 

effect of normalization on the formation of this fracture was not assessed. 

Figure 274. Comparison of predicted and measured behavior for Head Test 3.

A post-test analysis was performed to bound the effects of the EMS corrugation weld failures on 

the impact behavior. In the analysis, the end EMS corrugation was fully separated from the outer 

face sheet prior to the impact. The force-deflection characteristics for this analysis are compared 

to both the test and the pretest analysis in Figure 278. We see that the post-test analysis agrees 

well with the measured forces in the initial portion of the impact response. As the later 

deformations develop the characteristics of the test structures and post-test analyses begin to 

diverge. This indicates that the corrugation probably debonded from the face sheet locally under 

the impact at an early time but the progression to the edge of the end EMS structure probably 

occurred as the EMS deformations grew larger. 
Applied Research Associates, Inc.

2010

217

Figure 275. Separation of the outer face sheet on the end EMS structure.

Figure 276. Buckling of the end EMS corrugation in Head Test 3.
Applied Research Associates, Inc.

2010

218

Figure 277. Fractured EMS end cover plate in Head Test 3.

Figure 278. Comparison of Post-test analysis and measured behavior for Head Test 3.

The results of this test produced the following observations:
Applied Research Associates, Inc.

2010

219

1. Head Test 3 was successful in providing high quality test data for model validation on 

EMS structures and evaluating the impact response of a non-optimized EMS head 

protection concept. 

2. The agreement between the predicted and measured impact response was reasonable and 

the primary discrepancies between the model and the test are known (primarily failures of 

the EMS corrugation welds).

3. With a modified design of the corrugation attachment (e.g. adhesive bonding or new 

welding protocols), the predicted level of impact protection for this system should be 

achievable (approximately three million ft-lbs or three times energy absorption for the

baseline 500lb head impact). 

4. The total EMS structure mass in Head Test 3 is approximately two times that of the 500 

lb head used in the test (thus the total test specimen weight is approximately three times 

that of a bare head). 

5. The weight of the end EMS structures (0.777-inch end plate and corrugation) is 

equivalent to a steel plate with an effective thickness of 1.325 inches. Adding a more 

traditional head shield, or multiple layered head shields, with a combined thickness of 

1.325 inches would result in the head test configuration puncture energy of 

approximately 2.5 million ft-lbs (for the unpressurized 500 lb tank head). 

6. Implementation of this design would be more difficult on a traditional tank car due to the 

end EMS extension added to achieve a lower force level crush zone (12-inch end EMS 

extension). The extension of the end EMS structures would require a redesigned and 

extended draft sill to allow clearance between adjacent cars. 

7. The normalized effectiveness (energy dissipation/unit weight) of this EMS structure is 

similar to approach of adding additional thickness to a traditional head and head shield 

design. The additional complexity of implementing this EMS design with an energy 

absorbing crush zone is not warranted for the 6x6 impactor threat. However, with further 

optimization (other core designs, optimized layer thicknesses, improved crush zone 

performance) and for other impactor threats, there is still potential for significantly 

improved tank car protection performance with EMS systems. 

6.2.3 Side Impact Puncture Analyses of a Corrugated EMS Systems

The design of the EMS structure used in Head Test 3 was based on a non-optimized 6-inch

corrugation design constructed from 11 gauge 90 XF steel. The system was modified to have an 
Applied Research Associates, Inc.

2010

220

increased puncture resistance and higher puncture energy in a head impact. However, no efforts 

were made to change the design for a side impact. 

An analysis was performed to assess the side impact response using this baseline 6-inch 

corrugated EMS design. The geometry of this design was shown previously in Figure 258 and 

Figure 259. The model for the side impact analysis is shown in Figure 279. The system is 

applied over a baseline 500 lb chlorine tank pressurized to 100 psi. The standoff from the tank to 

the inside of the EMS structure is 4 inches. The total weight of the 6-inch EMS structure was

32,000 lbs which is approximately equal to the weight of a 0.59-inch thick tank. 

Figure 279. Model for the 6-inch corrugated EMS side impact analysis.

The calculated puncture behavior of the baseline 6-inch corrugated EMS structure is shown in 

Figure 280. The ram punches through the EMS structure at a load of approximately 200,000 lbs 

and a ram displacement of approximately 6 inches. Thus, the energy dissipation of this EMS 

structure is quite small compared to that of a single thicker jacket of equivalent weight. 

Based on the lessons learned from the previous detailed analyses, we can propose a corrugated 

EMS structure for the tank car side protection. We know we need to increase the thickness of 

the outer face sheet as much as is reasonably practical. To achieve this, the core depth was 

reduced to three inches. The proposed design is shown in Figure 281. Effectively, this design is 

a 1⁄2-scale reproduction of the preliminary corrugated EMS design but with a 0.25-inch-thick 

outer face sheet. The core sheet and inner face sheet have been reduced to 16-gauge (0.060-inch-
Applied Research Associates, Inc.

2010

221

thick). The construction material is not specified but any medium to high strength steel with 

good ductility can be used (e.g. TC128B, A516-70, A572-50). 

The model for the side impact analysis with this modified 3-inch EMS design is shown in Figure 

282. The system is again applied over a baseline 500 lb chlorine tank pressurized to 100 psi. 

The standoff from the tank to the inside of the EMS structure is 3.5 inches. The total weight of 

the 3-inch EMS structure was 28,000 lbs which is approximately equal to the weight of a 0.51

inch thick tank. This new design is a weight reduction of more than 4,000 lbs from the original 

6-inch corrugated EMS concept. 

 

Figure 280. Calculated 6-inch EMS puncture behavior.
Applied Research Associates, Inc.

2010

222

Figure 281. Detail of the modified EMS corrugation geometry.

Figure 282. Model for the 3-inch corrugated EMS side impact analysis.

The calculated puncture behavior of the modified 3-inch corrugated EMS structure is shown in 

Figure 283. In this case, the EMS structure deforms under the impactor and engages the inner 
Applied Research Associates, Inc.

2010

223

tank structure well before the outer EMS face sheet is penetrated. The ram punches through the 

EMS structure at a load of approximately 600,000 lbs and a ram displacement of approximately 

12 inches. Thus, the energy dissipation of this EMS structure is approximately 5 times that of 

the original 6-inch EMS concept. The comparison of the force-deflection characteristics of the 

3-inch and 6-inch corrugated EMS structures for these side impact analyses is shown in Figure 

284. 

This example of the performance of different corrugated EMS designs illustrates that the 

approaches required to optimize this type of system are to remove weight from the inner face 

sheet and core structure to lower the system stiffness and to increase the outer face sheet 

thickness to improve the penetration resistance. The extreme extrapolation of this concept is to 

move all of the EMS weight into a single thicker outer jacket as shown in Figure 285. The 

weights of the corrugated EMS systems are comparable to the 0.5625-inch thick jacket analyzed 

previously and described in Section 4.3.2 (analysis R1Q). The total weight of this jacket with a 

4-inch standoff was 29,500 lbs.
Applied Research Associates, Inc.

2010

224

 

 

Figure 283. Calculated 3-inch EMS puncture behavior.
Applied Research Associates, Inc.

2010

225

Figure 284. Comparison of the force-deflection behaviors for the corrugated EMS concepts.

The calculated force-deflection behavior for the 0.5625-inch-thick jacket is compared to those of 

the corrugated EMS structures in Figure 286. The largest difference with this design is that the 

thick jacket is not penetrated until a force of approximately 1.1 million lbs and an impactor 

displacement of approximately 30 inches. Thus, the jacket absorbs significantly more impact 

energy than the corrugated EMS structures for the 6-inch impactor. 

Although the monolithic jacket performs better than the corrugated EMS structures in this 

example, the conclusion can not be generalized to all EMS designs and impact scenarios. The 

EMS structures would be expected to perform significantly better under a more distributed 

loading. Other benefits of the EMS concepts are that the structural stiffness can be used to 

efficiently carry the train service loads in a “tank within a tank” design. Further analysis of other 

EMS concepts (e.g. core geometries, materials) and more general impact conditions would be 

required to obtain an optimized design. 
Applied Research Associates, Inc.

2010

226

Figure 285. Extrapolation of the corrugated EMS concept.

Figure 286. Comparison of the force-deflection behaviors for the corrugated EMS concepts.
Applied Research Associates, Inc.

2010

227

6.2.4 Side Impact Puncture Analyses of an Egg Crate EMS Systems

An analysis was performed for the side impact of an egg-crate EMS system. The parameters of 

the analyses were:

• 500 lb TC128B tank

• 0.777 inch tank thickness

• 4 inch egg crate EMS

• 8 gauge (0.1644”) face sheets

• 5 gauge (0.2092”) core

• 100 psi internal pressure

• Smeared lading

The model used for the egg crate EMS is shown in Figure 287. The impact condition used was a 

20 mph impact of a 295,000 lb rigid 6x6 inch ram. A 4-inch cell size was selected so that the 6-

inch square ram face could not punch between the ribs. The calculated force-deflection 

characteristics and the impact behavior are shown in Figure 288 and Figure 289, respectively. 

The response was a localized denting and crushing of the EMS structure that continues to a full 

penetration of the EMS structure. The puncture of the outer face sheet occurs at a total 

displacement of 6 inches and a force of 439,000 lbs. 

The puncture force plotted against the face sheet thickness for the egg crate system is compared 

to the other structural concepts in Figure 290. The puncture resistance of the outer face sheet 

was marginally more efficient than other systems. However, the total effective thickness of this 

EMS system was approximately equivalent to a 0.7 inch plate. As a result, this system again 

results in the individual layers resisting the impact forces and penetration independently, which 

significantly reduces the performance of the overall system. 
Applied Research Associates, Inc.

2010

228

Figure 287. Side impact model for a 4-inch egg crate EMS system.

Figure 288. Calculated impact forces for a 4-inch egg crate EMS system.
Applied Research Associates, Inc.

2010

229

(a) Initial impact (b) EMS penetration

(c) Detail of EMS crush

Figure 289. Calculated side impact damage for a 4-inch egg crate EMS system.
Applied Research Associates, Inc.

2010

230

Figure 290. Puncture force versus face sheet thickness for the egg crate EMS.

6.2.5 Summary of the EMS Puncture Analyses

A summary of the puncture forces plotted against the end plate and outer EMS face sheet 

thickness for the corrugated systems are compared to the other structural concepts in Figure 291. 

The puncture resistance of the outer layers (summing the thickness of the flat layers in contact 

outside the core structure) was consistent with that of other systems. However, by extending the 

end EMS structure from the side structure, the required impact force levels to crush the system 

were reduced and significantly more energy was dissipated before the puncture force capacities 

were exceeded. 
Applied Research Associates, Inc.

2010

231

Figure 291. EMS puncture forces as a function of outer layer thickness.

6.3 Double Tank Head Concept Puncture Analyses

One proposed concept for improving the tank head puncture resistance is to use a double head 

system. This idea is similar to a double hull concept for large tanker transport ships to reduce the 

potential for an oil spill in a grounding accident. The Geometry for the double head concept is 

relatively simple with an extension of the side shell of variable length and a second head added 

with an identical 2-to-1 ellipsoidal geometry. An example of a model used for a double head 

analysis is shown in Figure 292. In this model a 12-inch standoff distance between the heads is 

obtained by the extension of the tank shell.

The idea proposed for this double head concept is that the standoff distance, the relative 

thickness of the inner and outer heads, and the pressure between the two heads could all be 

varied to optimize the puncture resistance. A summary of the double head analyses is provided 

in Table 12. The sum of the total thickness of the two heads was maintained at 1.6 inches but the 

thickness of the individual heads was varied between 0.6, 0.8, and 1.0 inch. The inner tank 

pressure was maintained at 100 psi but the pressure between the inner and outer heads was varied 

between 0, 50, and 100 psi. The standoff distance between the inner to outer heads was varied 

between 2, 6, 12, and 18 inches. The calculated peak puncture force and puncture energies are 

also provided in the table. 
Applied Research Associates, Inc.

2010

232

Figure 292. Model for the double head puncture protection concept.

Table 12. Summary of the double head impact analyses

Calculation Inner/Outer 

head Thickness

Inner/Outer 

head pressure

Outer Head 

Standoff

EMS Puncture 

Energy (ft-lbs)

EMS Puncture 

Force (lbs)

R5A 0.80/0.80 inch 100/50 psi 12 Inches 1,775,000 1,243,000

R5B 1.00/0.60 inch 100/50 psi 12 Inches 1,925,000 1,451,000

R5C 0.60/1.00 inch 100/50 psi 12 Inches 1,720,000 1,533,000

R5D 0.80/0.80 inch 100/0 psi 12 Inches 1,625,000 1,319,000

R5E 0.80/0.80 inch 100/100 psi 12 Inches 1,900,000 1,049,000

R5F 0.80/0.80 inch 100/50 psi 6 Inches 1,670,000 1,365,000

R5G 0.80/0.80 inch 100/50 psi 18 Inches 1,920,000 1,087,000

R5H 0.80/0.80 inch 100/50 psi 2 Inches 1,650,000 1,475,000

R5I 0.80/0.80 inch 100/0 psi 2 Inches 1,500,000 1,490,000

R5J 1.00/0.60 inch 100/50 psi 2 Inches 1,550,000 1,406,000

The summaries of the peak puncture forces and puncture energies for the double head concepts 

compared to other puncture analyses are provided in Figure 293 and Figure 294, respectively. In 

general, the puncture forces are consistent with those of other head and tank systems. In two 

cases, the puncture forces fall below the lower bound of the previous analyses. In these two 

analyses, the outer head is punctured before the inner head is significantly engaged due to either 

the largest standoff distance (18 inches in calculation R5G) or the combination of a large 

standoff and high pressure between the heads (12-inch standoff and 100 psi between heads in 
Applied Research Associates, Inc.

2010

233

calculation R5E). Thus, these calculations have a puncture force that is controlled by the 

separate head thicknesses instead of the combined system thickness. 

Figure 293. Summary of the double head concept puncture forces.

The puncture energies for the double head system are higher than that of a traditional pressurized 

head and head shield system and more consistent with the previous analyses on unpressurized 

heads. This would indicate that there is at least some benefit to this type of optimized double 

head system for head impact protection. 

A parameter that was found to have a significant trend with respect to the puncture energy in the 

analyses was the head standoff distance. The double head models with the various standoff 

distances analyzed are shown in Figure 295. The corresponding impact behaviors for these 

models with balanced 0.8-inch-thick heads and a 50 psi pressure between the inner and outer 

heads is shown in Figure 296. The analyses show consistent trends with larger peak force levels 

with the smaller standoff distance and larger total puncture energy with the larger standoff 

distance. The increase in peak force levels with smaller standoffs occurs since the inner head is 

more significantly deformed and develops a higher impact resistance force at the time the outer 

head is punctured. The larger total puncture energy with increased standoff is a result of the 

larger total ram displacement required before the system is penetrated. 
Applied Research Associates, Inc.

2010

234

Figure 294. Summary of the double head concept puncture energies.

Figure 295. Double head models with different standoff distances (2, 6, 12, and 18 inches).
Applied Research Associates, Inc.

2010

235

Figure 296. Effect of the standoff distance on the double head impact response.

A summary of the puncture energies as a function of the head standoff distance for all of the 

double head analyses is shown in Figure 297. A linear fit to the data is provided in the figure 

and the slope of the curve corresponds to a 250,000 lb force. This helps to explain the trend 

observed in the analyses. The force displacement curve for a single 0.80-inch head pressurized 

to 50 psi is shown in Figure 298. The head is punctured at a force of approximately 880,000 lbs 

and the average residual force developed as the tapered head is pushed through the punctured 

head is on the order of 250,000 lbs. As a result, the larger standoff distances would add a larger 

energy dissipation contribution of the punctured outer shell before the inner head is punctured. 

The trend of increasing total puncture energies is therefore primarily an artifact of the impactor 

geometry and would not be as significant for improving the puncture protection for other 

impactor geometries (such as a broken rail). This effect is partially responsible for the larger 

average puncture energy of the double heads compared to the baseline pressurized head analyses, 

illustrated in Figure 294. 
Applied Research Associates, Inc.

2010

236

Figure 297. Summary of the standoff effects on the double head impact response.

Figure 298. Calculated force-deflection curve for a single 0.8-inch-thick head.
Applied Research Associates, Inc.

2010

237

The calculated impact behavior for analyses demonstrating the effects of changing the relative 

thickness of the two heads is shown in Figure 299 (analyses with the 12-inch standoff). The 

characteristics of the impact behavior are consistent with effectively summing the response of

two independent heads. The system is slightly more efficient when the outer head is thicker than 

the inner head. However, again this trend may be influenced by the larger post puncture forces 

that develop in the thicker outer head geometry. A similar trend was observed in a pair of 

analyses with the 2-inch standoff distance as shown in Figure 300. For the smaller standoff 

distance, the difference in the total puncture energy is almost entirely due to the puncture of the 

internal head occurring at a ram displacement that is 1-2 inches larger with the thicker inner head 

configuration. 

Figure 299. Effect of the head thickness on the double head impact response (12-inch standoff).
Applied Research Associates, Inc.

2010

238

Figure 300. Effect of the head thickness on the double head impact response (2-inch standoff).

The calculated impact behavior for analyses demonstrating the effects of changing the pressure 

between the two heads is shown in Figure 301 (analyses with the 12-inch standoff). The higher 

pressures result in an earlier puncture of the outer head but a delayed puncture of the inner head. 

The larger total displacement before the inner head is punctured produces higher total system 

puncture energies when the pressure between the heads is increased. The assumption of these 

analyses is that the pressure between the heads remains constant throughout the impact and head 

puncture process. If the pressure drops rapidly after the outer head is punctured, the overall 

puncture energy of the double head system would be reduced as the pressure between the heads 

is increased. 
Applied Research Associates, Inc.

2010

239

Figure 301. Effect of the pressure ratio on the double head impact response (12-inch standoff).

A similar effect of the pressure between the heads was observed in a pair of analyses with the 2-

inch standoff distance as shown in Figure 302. For the smaller standoff distance, the difference 

in the total puncture energy is again almost entirely due to the puncture of the internal head 

occurring at a ram displacement that is approximately 2 inches larger with the increase pressure 

between the two heads. 
Applied Research Associates, Inc.

2010

240

Figure 302. Effect of the pressure ratio on the double head impact response (2-inch standoff).

6.4 Analyses of Composite Tank Protection Concepts

The above analyses indicate that a tank protection system developed using only typical tank car 

materials (steels) will have difficulty reaching the protection goals of the NGRTC program. To 

achieve a five times increase in the puncture energy of tank cars while maintaining a comparable 

level of utility will require nontraditional tank car design approaches and advanced materials. 

These potential materials for advanced protection concepts include polymeric composite 

materials. The potential of these materials for increasing tank car puncture protection is 

described in this section. 

6.4.1 Validation of Composite Modeling

Before composite material systems can be evaluated, a constitutive model for the composites 

including appropriate damage and failure behaviors needed to be developed and validated. The 

constitutive and damage behaviors of composites are more complex due to the material 

anisotropy, inhomogeneity, and the large number of resulting damage modes that can develop 

(e.g. matrix cracking, fiber breakage, and delamination). 

There are several constitutive models available in LS-DYNA for composite materials. However, 

for the local impact and puncture behavior of the tank car protection requirements, an important 
Applied Research Associates, Inc.

2010

241

damage mechanism will be the delamination of the composite in the impact zone. Just as 

required for metals, the approach to accurately model this damage mode requires a local solid 

element patch with a detailed failure model capable of simulating the range of composite damage 

and failure behaviors. 

After an evaluation of available modeling options, the LS-DYNA material type 162 was selected 

(*MAT_COMPOSITE_MSC_DMG) [83]. This model may be used to simulate progressive 

failure in composite materials consisting of unidirectional and woven fabric layers subjected to 

high strain-rate and high pressure loading conditions. The model is a generalization of the layer 

failure model reported in [84]. The Matzenmiller, Lubliner, and Taylor (MLT) damage 

mechanics approach [85] has been adopted to characterize the softening behavior after damage 

initiation. Failure models based on the 3D strains in a composite layer with improved 

progressive failure modeling capability are established for a unidirectional or a plain weave 

fabric composite layer. They can be used to effectively simulate the fiber failure and 

delamination behavior under impact conditions. This model was developed and is supported by 

the Materials Science Corporation (MSC) and requires an additional license to activate the model 

capabilities in LS-DYNA. 

Composite Punch Testing

A composite punch test series was found in the open literature that was appropriate for the 

validation of the composite constitutive and damage model. The punch tests were performed at 

the University of Delaware Center for Composite Materials (UD-CCM) on a fiberglass/epoxy 

composite [86, 87]. The specific composite system used in the tests is one proposed for use in 

ballistic impact protection and has characteristics that make it a good candidate material for tank 

car impact protection (e.g. toughened matrix material system and relatively inexpensive 

constituent materials). The tests span a sufficient range of loading conditions to be a valuable

data set for model validation. 

The composites used in this study were fabricated from two-dimensional [2-D] woven S-2 Glass 

fabric and SC15 resin. Composite laminates were fabricated using 2, 4, 6, 11, and 22 layers of 

the plain weave fabric. Specimens made from these laminates are tested under punch shear 

loading with support-span to punch ratio (SPR) of 2.0 and 8.0, as shown in Figure 303. The 

quasi-static punch shear tests were conducted in the displacement-control mode at a cross-head 

displacement rate of 0.10 inch/min. 

The punch shear test fixture consists of a square support plate (2 inches thick) with a circular 

hole at the center, a relatively thin cover plate (0.5 inch thick) with a central hole similar to the 

support plate, and a cylindrical punch. Sets of support plates and cover plates were fabricated 

with different support span (SS) diameters (DS) of between 0.55 inch and 4.0 inches. Composite 
Applied Research Associates, Inc.

2010

242

plate specimens were bolted on the punch shear test fixture between the support plate and the 

cover plate. A cylindrical punch was used with a 0.5 inch diameter (DP) and a flat tip. The 

combination of punch and support spans provides spans to punch ratios of between 1.1 and 8.0 

(SPR = DS/DP). In addition, a test was performed against a support plate without a hole (SPR = 

0). The calculated force-deflection behavior for a 22 layer thick composite plate and different 

span ratios is shown in Figure 304. The corresponding damage in these punch test specimens is 

shown in Figure 305. Force-deflection behavior of the tests with SPR=8 for various layer 

thicknesses are shown in Figure 306.

Figure 303. Experimental set-up for the quasi-static punch-shear experiment [86].
Applied Research Associates, Inc.

2010

243

Figure 304. Effect of the span ratio on the composite punch force [87].

Composite Punch Modeling

Models of the punch tests were developed to validate the *MAT_COMPOSITE_MSC_DMG 

material model. Our approach was to model the same experiments performed by Xiao et al [86] 

and confirm the agreement between simulations and experiments reported, as shown in Figure 

307. The LS-DYNA models of the punch simulations were developed based on descriptions 

provided by Xiao et al in Reference 86. 

A resulting model of an 11-layer composite punch experiment is shown in Figure 308. The 

constraining ring was modeled as rigid and the punch as steel. Contact was specified between 

the ring and composite and the punch and composite with no friction. Eroding contact was 

defined between the composite layers. Constitutive properties of the composites used in this 

study are provided in Table 13.
Applied Research Associates, Inc.

2010

244

Figure 305. Effect of the span ratio on the composite punch test damage mode [87].
Applied Research Associates, Inc.

2010

245

Figure 306. Effect of the specimen thickness on the composite punch force (SPR = 8) [86].

Figure 307. Comparison of the composite damage model with punch test data [86]. 
Applied Research Associates, Inc.

2010

246

Figure 308. LS-DYNA model of an 11-layer composite punch experiment.

Table 13. Constitutive model parameters for woven S-2 Glass/SC15 resin composite

MID ρ, kg/m3 EA EB EC PRBA PRCA PRCB

1 1.85E+03 27.5 GPa 27.5 GPa 11.8 GPa 0.11 0.18 0.18

GAB GBC GCA AOPT — — — —

2.9 GPa 2.14 GPa 2.14 GPa 2 — — — —

XP YP ZP A1 A2 A3 — —

0 0 0 1 0 0 — —

V1 V2 V3 D1 D2 D3 beta —

0 0 0 0 1 0 0 —

SXT SXC SYT SYC SZT SFC SFS SXY

604 MPa 291 MPa 604 MPa 291 MPa 58 MPa 850 MPa 300 MPa 75 MPa

SYZ SZX SFFC AMODEL PHIC E_LIMT S_DELM —

58 MPa 58 MPa 0.3 2 10 0.2 1.2 —

OMGMAX ECRSH EEXPN CERATE1 AM1 — — —

0.999 0.1 4 0 2.0 — — —

AM2 AM3 AM4 CERATE2 CERATE3 CERATE4 — —

2.0 0.5 0.2 0 0 0 — —

Note: Parameter definitions can be found in the LS-DYNA User’s Manual [15].
Applied Research Associates, Inc.

2010

247

Initial simulations using various shared-memory parallel processing (SMP) versions of LSDYNA did not yield good results in predicting the experimental punch behavior. A sample of 

the runs performed with one of the LS-DYNA versions (ls971s_R2_7600, a formally released 

version) are shown in Figure 309. Initial loading of the composites agrees well with the 

experiments, but the simulated punch behavior consistently failed at lower load and 

displacements. Various modeling parameters that may affect the simulation results were 

considered, but were never adequate to achieve the load and displacements at failure seen 

experimentally. These included the following:

• Material constitutive parameters. Material 162 parameters for the plain weave S-2 

Glass/SC15 composite vary some in the literature [e.g. 84, 88; 89]. Variations in the 

interlaminar shear stress concentration factor (S_DELIM) were also considered since this 

variable is mesh-dependent.

• Contact type between the punch and composite.

• Loading rate and time stepping.

• Number of elements per ply (1, 2, 4, 6 elements per ply considered).

Figure 309. Comparison of initial punch simulations using ls971s_R2_7600.

A likely reason for the lack of agreement between simulations and tests was the functionality of 

material 162 in the versions of LS-DYNA built to run on ARA computers. Significant effort was 

expended in working with LSTC and MSC to have a version of LS-DYNA functioning correctly 
Applied Research Associates, Inc.

2010

248

with material 162. Material 162 is formally supported by MSC, but LSTC provides versions of 

LS-DYNA with this material model. A request was made to MSC for LS-DYNA models that 

could be used to validate that the material model was functioning correctly in the versions of LSDYNA built for ARA. MSC provided single element models to proof various functions of the 

model. However, these single element tests do not provide a comprehensive way of assessing the 

models functionality as there is damage growth that occurs based on the state of surrounding 

elements. Several requests were made to MSC for additional models to test the functionality of 

material 162. Unfortunately, none were provided that could be validated against other results.

It was found that several SMP versions of LS-DYNA provided by LSTC did not pass all of these 

single element tests. In particular, when an element was loaded with through-the-thickness 

compression and shear, the nonlinear response was different than what is reported by UD-CCM 

[90], as shown in Figure 310. Ultimately, for the version ls971s_R4_47600 (all R4 builds of LSDYNA provided by LSTC allow for user-defined materials to be used, which was need for the 

BW model), the response of all single element tests matched the results from UD-CCM. 

However, punch simulations with these versions still did not match experimental results.

Figure 310. Single element response under through-the-thickness compression-shear for various 

SMP versions of LS-DYNA.

According to Xiao et al [86], successful comparisons between simulations and experiments were 

achieved using an older version of LS-DYNA, ls970. In a final attempt to get better agreement 
Applied Research Associates, Inc.

2010

249

with the punch experiments, 4-layer and 11-layer models were run using ls970s_6763. The 

comparison of the results from the ls970 and ls971 versions of the code for the 4-layer and 11-

layer models are shown in Figure 311 and Figure 312, respectively. The load-displacement for 

the 4-layer simulations were very similar using both versions of the code. However, the 11-layer 

simulation was significantly different. The result using ls970s_6763 more closely matches the 

experimental and published simulation results, but still punches through the composite plate at 

roughly half the displacement. 

Figure 311. Comparison of 4-layer punch simulations using ls970 and ls971.
Applied Research Associates, Inc.

2010

250

Figure 312. Comparison of 11-layer punch simulations using ls970 and ls971.

These results give no indication that the material 162 model implemented in any of the versions 

of LS-DYNA received from LSTC works correctly. It consistently produces lower forcedisplacements at failure than seen in experiments and other simulations. Identifying and 

correcting the error would require further collaboration with LSTC, MSC and possibly UDCCM.

Although the punch test simulations consistently produced lower force-displacements at failure, 

the model for material 162 showed reasonably good agreement up to this point. Simulation of 

the composite jacket on a tank car should therefore provide a lower bound on its performance.

6.4.2 Tank Car Side Impact Analyses with a Composite Jacket

A model was developed to assess the puncture resistance of the 500 lb chlorine tank car with a

composite jacket retrofit. A 1.5-inch thick composite jacket was placed around the commodity 

tank, as shown in Figure 313. A composite with this thickness has approximately the same 

weight as a steel jacket with a thickness of 0.355 inches (note that the same tank car retrofit with 

a 0.375-inch steel jacket was assessed in Section 4.3.2). A (0/90) layup with principal material 

directions coincident with the hoop and axial tank directions and equal numbers of 0 degree and 

90 degree layers was used. The model for the commodity tank is the same as discussed in the 

previous analyses. 
Applied Research Associates, Inc.

2010

251

The model developed for the composite jacket is shown in Figure 314. Outside the impact zone, 

the jacket was modeled with shell elements. These were connected to solid elements in the 

impact zone, much like the approach used for modeling the commodity tank. Here, however, the 

solid mesh is made of two separate material models for the composite, material type 2 (*MAT_

ORTHOTROPIC_ELASTIC) for the coarser transition elements and material type 162 (*MAT_ 

COMPOSITE_MSC_DMG) in the path of the impactor. Material type 54 (*MAT_

ENHANCED_COMPOSITE_DAMAGE) was used to model the shell elements. 

Material type 54 uses the ply-level constitutive properties to predict laminate properties by 

specifying the number of plies and their orientations by means of lamination theory. For these 

preliminary analyses, no damage was allowed in the model. The shell model therefore provides 

the correct elastic stiffness for a 1.5-inch thick composite prior to damage initiation. The 

transition elements, modeled with material type 2, also model the composite as linear elastic, but 

with the same through-the-thickness elastic properties used in material type 162. This transition 

region is necessary to reduce the number of nodes through the thickness of the jacket to be 

compatible with using the algorithm in LS-DYNA for connecting the shell elements to the solid 

elements.

The region in the path of the impactor was modeled with material type 162. To reduce the model 

size, this region was model with 12 thick plies of the composite fabric. In reality, a composite 

this thick would require approximately 64 plies of fabric where delamination could occur 

between each layer. To model each ply, a minimum of 64 elements would be needed through the 

thickness. Instead, the composite was modeled with 12 layers with 3 elements through the 

thickness of each ply, resulting in 36 elements through the thickness. Elements therefore have a 

characteristic length of about .042 inches (1 mm). Delamination can still occur in the model, but 

only between each of the 12 layers. 
Applied Research Associates, Inc.

2010

252

Figure 313. Side impact model with a composite jacket.
Applied Research Associates, Inc.

2010

253

Figure 314. Composite tank model and impact zone mesh used for side impact puncture 

analyses.

Preliminary analyses with the tank model were performed without any internal pressure or the 

BW failure model for the steel implemented. This analysis is still ongoing. Currently, the 

analysis has progressed to a time of 36 ms (12.5 inch displacement), as shown in Figure 315. 

Impactor force-displacement is shown in Figure 316. For comparison, the force-deflection curve 

of a 500 lb chlorine tank car with an 11-gauge A1011 jacket and pressurized to 50 psi was added 

to the figure. This configuration was selected since it has a very similar stiffness to the 

unpressurized tank with the composite jacket. 

The current impact load is in excess of 500,000 lbs and the composite jacket has significant 

damage but is still capable of resisting the impact loads around the perimeter of the contact 

patch. Based on the observed level of damage, it is not expected that the calculation would 

predict a puncture force that is significantly greater than that of an equivalent weight steel jacket. 

However, the composite model is also expected to under predict the composite strength in this 

analysis. 
Applied Research Associates, Inc.

2010

254

Figure 315. Punch-shear mode in composite jacket at 18 inches of displacement.

One effect seen in the analyses is the potential for the composite to provide a blunting of the 

impact loads on the commodity tank. A comparison of the impact behavior of a 1.5-inch thick 

composite jacket with an equivalent weight steel jacket is shown in Figure 317 and Figure 318. 

The comparison clearly shows a blunting of the loads on the commodity tank for the same level 
Applied Research Associates, Inc.

2010

255

of impact displacement and this effect has potential for delaying the development of damage and 

penetration of the tank. The magnitude of this blunting effect would be expected to increase as 

the thickness of the composite jacket becomes greater. 

Figure 316. Side impact force-deflection curves for a 500 lb tank car with a composite jacket.

 

(a) Composite Jacket (b) 0.375” A1011 Steel Jacket (0 psi)

Figure 317. Composite jacket blunting effect at 50 ms (17-inch displacement).
Applied Research Associates, Inc.

2010

256

 

(a) Composite Jacket b) 0.375” A1011 Steel Jacket (0 psi)

Figure 318. Effective strains in the 500 lb commodity tank at 50 ms (17-inch displacement).

Empirical Evaluation of Composite Puncture Protection Potential

The detailed finite element analyses did not produce a reliable assessment of the impact and 

puncture behavior of a tank car with a composite protection system. Alternatively, the data 

available in the open literature sources can be used to estimate the potential of a composite layer 

to enhance the tank car puncture resistance. 

As was discussed in previous sections, the tank impact behavior is dominated by punch-shear 

through the jacket and commodity tank. By comparing normalized punch test peak loads from 

the composites and steel testing, an estimate can be made as to their relative performance. A 

normalized peak punch load was calculated as

D t

F

F

p

p

n ρ = (11)

where Fp is the peak punch load, Dp is the diameter of the punch, ρ is the material density and t is 

the plate thickness. Effectively, this is the shear stress around the punch perimeter scaled by the 

material density. As a result, the punch shear resistance is compared on an equivalent weight 

basis rather than an equivalent thickness basis. 

Data were taken from the punch tests performed on A1011, HSLA 90x and TC128B steels by 

SwRI Materials Test Lab, discussed in Section 3.4.4, and normalized [32-34]. Data from the 

composite punch tests performed by Xiao et al [86] for SPR = 2 and SPR =8, as was shown for 

SPR=8 in Figure 306, were similarly normalized. These are plotted against aerial density in 

Figure 319. Steel data are an average of all span ratios tested, from 1.5 to 2.5. The ranges of 

results from different span ratios are also shown in the figure. The effectiveness of the 
Applied Research Associates, Inc.

2010

257

composite appears to drop as its thickness is increased. Some of this effect is the shift from a 

bending failure mode to a punch-shear mode. At larger thicknesses, and higher loads, this may 

also be due to stress concentrations at the sharp punch edges causing localized failure. Recall 

that the composite punch tests used a square punch where the steel tests used a punch with 

rounded edges. Therefore, in comparing the composite and shear punch data, we believe that 

these values provide a lower bound on the performance of the composite.

Figure 319. Normalized peak punch test loads versus aerial density for various materials.

At the largest aerial density testing for the composite, a direct comparison with A1011 and 

HSLA 90x can be made because the 11 gauge steels used in these tests had a similar aerial 

density. The composite out performs A1011 by approximately 70% and HSLA 90x by roughly 

10%. However, if one assumes that the performance of the composite remains the same with 

larger thicknesses (aerial densities above 0.1), the HSLA 90x and TC128B show comparable 

performance.

The above comparison shows that a composite puncture protection layer could perform as well 

as a steel layer and potentially perform better. Improved performance of the composite system 

could likely be obtained by optimizing the composite material and/or layup for these types of 

impact conditions. For example, the simulated failure begins at the punch corner preferentially 

develops along the hoop direction. The current layups is a 0/90 plain weave, with equal numbers 

of layers in the hoop and axial directions. Layers oriented off the principal tank axes and with a 
Applied Research Associates, Inc.

2010

258

greater number in the axial direction may improve the performance. However, it is not expected 

that an optimized layup alone would result in drastic increases in performance. As a result, we 

have not yet identified any composite systems that could achieve a puncture resistance that 

approach a two times improvement over a comparable weight steel system. 
Applied Research Associates, Inc.

2010

259

7 Summary and Conclusions

This report describes the results of a research program to develop strategies for improving 

railroad tank cars so they can maintain tank integrity for more severe accident conditions than 

current equipment. The approach used in the research and development program was to develop 

a tank impact and puncture prediction capability using detailed finite element analyses. Along 

with the modeling efforts, an extensive test program was performed. The testing included both 

laboratory tests of materials and components and full scale impact testing of tank cars and tank 

protection concepts. The test program was critical for both the development and the validation 

of the puncture modeling capability. 

This tank car impact and puncture modeling capability was applied to evaluate a wide range of 

tank/jacket and head/head shield geometries. The side impact condition was a normal impact 

centered on the belt line of the tank and the head impact was an offset impact point 

approximately 29 inches vertically downward from the center of the head. The calculated head 

and shell impact forces as a function of the combined head and head shield thickness are shown 

in Figure 320 (all analyses with the 6x6 inch impactor). The figure shows the analyses are in 

general consistent with a linear relationship between puncture force and total thickness of the 

protective layers. 

The linear relationship between the puncture force and total tank system thickness provides an 

indication of the primary failure mechanism initiating the tank puncture. The 6x6 inch impactor 

produces a large shear stress in the tank wall around the perimeter of the contact patch. The 

slope of the linear fit in Figure 320 corresponds to an average shear stress in the tank layers 

around the perimeter of the impactor of 39 ksi. By comparison, the yield and ultimate stress 

levels of the TC128B in pure shear are 33 ksi and 49 ksi, respectively (approximately 58% of the 

stress values in pure tension using a Von Mises yield criterion). Thus, the failure mode is 

primarily exceeding the shear capacity around the perimeter of the impact patch. 

The calculated puncture forces for pressurized heads and the thicker head systems tended to fall 

slightly below the linear fit in Figure 320. The proposed mechanism for these lower forces is 

that, for the stiffer head systems, the offset impact creates a larger stress concentration along the 

upper edge of the impactor face and the failure initiates at that location at a lower total force. 

The more compliant head systems allow for a larger dent to form and the impactor develops a 

more uniform stress distribution in the impact patch around the ram face perimeter. 
Applied Research Associates, Inc.

2010

260

Figure 320. Calculated puncture forces as a function of system thickness.

The calculated puncture energies for all of the side and head impact analyses are plotted against 

the combined thickness of the system in Figure 321 (6x6 inch impactor only). When comparing 

all of the puncture energies, the data falls into four separate groups that are distinguished by the 

impact type (side or head) and the tank pressure. Again, the comparison indicates that the total 

thickness is the relative parameter that determines puncture energy for a given impact condition 

and pressure level, indicating that a retrofit design with an increased jacket thickness should 

provide equivalent protection to a thicker commodity tank system with equivalent combined 

shell and jacket thicknesses. 

The puncture energy summary in Figure 321 illustrates the important effect of internal pressure. 

If we examine the results from side impact analyses, we find that the puncture energy is 50% to 

75% higher when the internal pressure is reduced from 100 psi to 50 psi. This is a potential 

opportunity to reduce the probability of tank punctures by controlling the tank pressure (e.g. 

modify operational procedures to ship at a lower average temperature). 

The correlation of the system thickness and puncture energy can also be converted to assess 

system requirements for a given impact scenario. The relationship between the impact speed and 

impact energy for the 295,000 lb ram car is provided in Figure 322. This impact energy is 

proportional to the mass of the object. A 286,000 ram car would have energies that are three 

percent lower than those in Figure 322. This relationship between speed and impact energy can 
Applied Research Associates, Inc.

2010

261

be combined with the linear fits between system thickness and puncture energy, shown in Figure 

321, to determine the required system thickness to resist puncture for a specified impact speed. 

Figure 321. Calculated puncture energies as a function of system thickness.

Examples of the pressurized (100 psi) Cl tank car thickness requirements to resist a puncture of 

the 6x6 inch impactor at various impact speeds is shown in Figure 323 for both side and head 

impacts. The figure shows that as the impact speed increases, the system thickness required 

begins to increase rapidly. A 25 mph side impact would require approximately three inches of 

steel to prevent puncture and a 30 mph head impact would require more than six inches of steel. 

Obviously, these protection levels are not achievable with a traditional tank car design approach 

while maintaining a tank car that is economically viable as a result of both the initial tank car 

cost and the drastically reduced lading capacity. 

The modeling capabilities developed in this program were applied to analyze a series of 

advanced tank car protective structure concepts. Included in these were multi-layered foam 

systems and engineered metal structures (EMS) systems. A variety of designs were investigated 

both with analyses and with full-scale head tests. However, neither of the foam or EMS systems 

was optimized. 
Applied Research Associates, Inc.

2010

262

Figure 322. Relationship between impact speed and kinetic energy for the ram car.

Figure 323. Puncture prevention thickness requirements for various impact speeds.
Applied Research Associates, Inc.

2010

263

The failure mode of the advanced systems was similar to that of the baseline tank systems when 

impacted by the 6x6 inch impactor. When the load acting on a layer of the advanced protective 

system exceeded the shear capacity of the layer, the layer would be punctured. For many of the 

advanced concepts, the system design would increase the stiffness of the support for the outer 

layers. The increased support structure stiffness would result in outer layer penetration at 

reduced ram displacements. As a result, the protection system and commodity tank would be 

penetrated independently at lower peak forces and lower puncture energies than could be 

obtained in an optimized system. In general, the performance of advanced protective systems 

was typically improved when the thickness of the outer layer was increased to resist puncture 

and the stiffness of the supporting structure was reduced to increase the displacements of the 

system before the puncture level was exceeded. 

The calculated forces to puncture the outer layer of foam and EMS systems are added to the plot 

of the puncture force versus system thickness in Figure 324. The difference in the data for the 

advanced systems is that the thickness used is the layer thickness only (e.g. outer head shield or 

outer face sheet thickness) rather than the sum of the total tank and jacket layer thicknesses. The 

comparison shows that the foam and EMS structures had the effect of making each layer act 

independently rather than coming in contact and working together to resist the high shear stresses 

around the edge of the impactor. As a result, the maximum achievable puncture forces were 

reduced and the systems do not reach the desired puncture energies. 

Although the systems with a monolithic jacket performed better than the corrugated EMS 

structures in this example, the conclusion can not be generalized to all EMS designs and impact 

scenarios. The EMS structures would be expected to perform significantly better under a more 

distributed loading. Other benefits of the EMS concepts are that the structural stiffness can be 

used to efficiently carry the train service loads in a “tank within a tank” design. Further analysis 

of other EMS concepts (e.g. core geometries, materials) and more general impact conditions 

would be required to obtain an optimized design. 

A similar set of observations was obtained from the analysis of layered foam systems. In 

general, the systems analyzed did not outperform a single monolithic plate of equivalent 

protection system weight for the 6x6 inch impactor. However, the performance for larger 

impactors and the optimum foam and interface properties to maximize performance were not 

fully explored. In addition, low density foams can contribute to improved thermal insulation. If 

a foam system can maintain a lower average shipping temperature (and thus a lower average tank 

pressure), the overall tank puncture energy could potentially be increased. 
Applied Research Associates, Inc.

2010

264

Figure 324. EMS puncture forces as a function of outer layer thickness.

A preliminary evaluation was also performed to assess the potential of composite materials for 

tank car puncture protection. Composite materials have densities that are typically one quarter to 

one third of the density of steel. Therefore, a much thicker composite system could be used to 

protect the tank car for the same added weight. 

A comparison of punch test data indicates that there is potential for modest puncture resistance 

improvements using composite materials. For specific test conditions, the composite out 

performs A1011 by approximately 70% and HSLA 90x by roughly 10% on an equivalent weight 

basis. Further improvements in the performance of the composite system could likely be 

obtained by optimizing the composite material and/or layup for these types of impact conditions. 

However, we have not yet identified any composite systems that could achieve a puncture 

resistance that approach a two times improvement over a comparable weight steel system. 

Impact Performance Summary

A summary of the current state of tank car side impact performance is provided in Figure 325. 

From the full scale side impact Test 2, described in Section 2.3.2, the 105J500W tank car 

puncture energy slightly less than one million ft-lbs. This corresponds to a 10 mph impact with a 

286,000 lb ram car and the 6x6 inch impactor. From analyses, the puncture energies of the 
Applied Research Associates, Inc.

2010

265

105J300W and 105J600W tank cars are approximately 30 percent less and 40 percent greater 

than the 105J500W tank car, respectively. 

The goal of the Next Generation Railroad Tank Car Project was to increase the puncture energy 

of the tank car by five to ten times over the baseline 105J500W tank car design. This target is 

outlined by the red box in Figure 325. The proposed TIH tank car standard falls within this 

range, requiring a puncture energy that is 6 times that of the baseline tank car. However, the 

levels that were found to be achievable with the first generation concepts in the NGRTC project 

are more consistent with two to three times the puncture energy of the baseline design. 

Figure 325. Summary of side impact puncture performance and goals.

A similar summary of the current state of tank car head impact performance is provided in Figure 

326. The calculated puncture energy for a pressurized (100 psi) 105J500W tank head with an 

11-gauge head jacket is 610,000 ft-lbs (analysis described in Section 5.3.2). This corresponds to 

an 8 mph impact with the 286,000 lb ram car and the 6x6 inch impactor. For comparison, the 
Applied Research Associates, Inc.

2010

266

puncture energy of a 105J600W tank head with a 0.50 inch full height head shield is 1.1 million 

ft-lbs. 

The goal of the Next Generation Railroad Tank Car Project was to increase the puncture energy 

of the tank car by five to ten times over the baseline 105J500W tank head as outlined by the red 

box in Figure 326. The proposed TIH tank car standard is above this range, requiring a puncture 

energy that is approximately 14 times that of the baseline tank head. However, the levels that 

were found to be achievable with the first generation concepts in the NGRTC project are more 

consistent with 2.5 to 4 times the puncture energy of the baseline design. 

Figure 326. Summary of head impact puncture performance and goals.
Applied Research Associates, Inc.

2010

267

8 References

1. Tyrell, D.C., Jeong, D.Y., Jacobsen, K., Martinez, E., "Improved Tank Car Safety Research," 

Proceedings of the 2007 ASME Rail Transportation Division Fall Technical Conference, 

RTDF2007-46013, September 2007.

2. National Transportation Safety Board, 2004: “Derailment of Canadian Pacific Railway Freight 

Train 292-16 and Subsequent Release of Anhydrous Ammonia Near Minot, North Dakota; January 

18, 2002,” Railroad Accident Report NTSB/RAR-04/01.

3. National Transportation Safety Board, 2006: “Collision of Union Pacific Railroad Train MHOTU23 with BNSF Railway Company Train MEAP-TUL-126-D with Subsequent Derailment and 

Hazardous Materials Release, Macdona, Texas, June 28, 2004,” Railroad Accident Report 

NTSB/RAR-06/03.

4. National Transportation Safety Board, 2005: “Collision of Norfolk Southern Freight Train 192 with 

Standing Norfolk Southern Local Train P22 with Subsequent Hazardous Materials Release at 

Graniteville, South Carolina, January 6, 2005,” Railroad Accident Report NTSB/RAR-05/04.

5. Barkan, Christopher P.L., Satis V. Ukkusuri, S. Travis Waller, "Optimizing the design of railway 

tank cars to minimize accident-caused releases", Computers & Operations Research, (2005 Elsevier 

Ltd).

6. Saat, Mohd Rapik, Christopher P.L. Barkan, "Release Risk and Optimization of Railroad Tank Car 

Safety Design", Transportation Research Record: Journal of the Transportation Research Board 

(2005).

7. Saat, Mohd Rapik, Christopher P.L. Barkan, "Release Risk as Metric for Evaluating Tank Car 

Safety Performance", Saat & Barkan 05-1826.

8. Jeong, D.Y., Lyons, M.L., Orringer, O., Perlman, A.B., 2007: “Equations of Motion for Train 

Derailment Dynamics,” Proceedings of 2007 ASME Rail Transportation Division Fall Technical 

Conference, RTDF2007-46009.

9. S.W. Kirkpatrick, B.D. Peterson, and R.A. MacNeill, "Finite Element Analysis of Train 

Derailments," ICrash2006. Proceedings of the International Crashworthiness Conference, July 4-7, 

2006, Athens, Greece.

10. Association of American Railroads, Technical Services Division, Mechanical Section—Manual of 

Standards and Recommended Practices.

11. Tyrell, D.C., Jacobsen, K., Talamini, B., Carolan, M., "Developing Strategies for Maintaining Tank 

Car Integrity During Train Accidents," Proceedings of the 2007 ASME Rail Transportation 

Division Fall Technical Conference, RTDF2007-46015, September 2007. 

12. S.W. Kirkpatrick and R.W. Klopp, "Risk Assessment for Damaged Pressure Tank Cars," 2003 

COMPUTER TECHNOLOGY AND APPLICATIONS, ASME Publications, Proceedings of: 

PVP2003 - The 2003 ASME PVP Conference, July 20-24, 2003, Cleveland, Ohio, W. Reinhardt 

and D.J. Martin, Eds.

13. S.W. Kirkpatrick and R.W. Klopp, "Hazard Assessment for Pressure Tank Cars Involved in 

Accidents," ICrash 2000. Proceedings of the International Crashworthiness Conference, September 

6-8, 2000, London, UK, E. C. Chirwa and D. Otte Eds., pp. 223-235. 
Applied Research Associates, Inc.

2010

268

14. R.W. Klopp, S.W. Kirkpatrick, and D.A. Shockey, “Damage Assessment of Tank Cars Involved in 

Accidents: Phase II - Modeling and Validation,” Final Technical Report to the Federal Railroad 

Administration, FRA Report, June, 1999.

15. “LS-DYNA Keyword User’s Manual," Livermore Software Technology Corporation, Version 970, 

April 2003.

16. "TrueGrid Manual, Version 2.1.0," XYZ Scientific Applications, Inc., September 2001.

17. Derailment of Canadian Pacific Railway Freight Train 292-16 and Subsequent Release of 

Anhydrous Ammonia Near Minot, North Dakota, January 18, 2002, NTSB Railroad Accident 

Report No. NTSB/RAR-04/01, March 2004. 

18. Kirkpatrick, S. W., “Analysis of the Tank Car Response in the Minot ND Derailment,” ARA 

Technical Report, June 17, 2005.

19. “Draft of NTSB Material Laboratory Factual Report [In reference to Canadian Pacific Railway 

Derailment and Anhydrous Ammonia Release, Minot, ND, January 18, 2002, DCA-02-MP-002],” 

Frank Zakar, Metallurgy Group Chairman, Feb. 14, 2003.

20. Matthew Witte and Satima Anankitpaiboon, “Fully Instrumented Side Impact Test of Tank Car 

3069,” Transportation Technology Center, Inc. Report No. P-07-033, October 8, 2007.

21. Matthew Witte and Satima Anankitpaiboon, “Second Fully Instrumented Tank Car Baseline Side 

Impact Test of Tank Car 3074,” Transportation Technology Center, Inc. Report No. P-07-034, 

October 8, 2007.

22. Boscardin, Marco D., “Vehicle Impact Wall Evaluation Report, AAR Test Facility, Pueblo, 

Colorado” Boscardin Consulting Engineers, Inc. Report to TTCI, July 9, 2007. 

23. Bao, Y., and Wierzbicki, T., “On Fracture Locus in the Equivalent Strain and Stress Triaxiality 

Space,” International Journal of Mechanical Sciences 46, 81-98.

24. Bao, Y., and Wierzbicki, T., “A Comparative Study on Various Ductile Crack Formation Criteria,” 

Journal of Engineering Materials and Technology 126, 314-324.

25. Young-Woong Lee and Tomasz Wierzbicki, “Quick Fracture Calibration for Industrial Use,” 

Massachusetts Institute of Technology Impact & Crashworthiness Laboratory, Report No: 115, 

August 2004. 

26. Tang, Y., Yu, H., Gordon, J., Jeong, D., and Perlman, A.B., “Analysis of Railroad Tank Car Shell 

Impacts Using Finite Element Method,” Proceedings of the 2008 ASME/IEEE Joint Rail 

Conference, JRC2008-63014, April 2008.

27. Pete McKeighan, “Tensile Properties and Stress-Strain Behavior of Various Steel Products Used in 

Fabricating Test Specimen for the NGRTC Program,” NGRTC Project Memorandum, Southwest 

Research Institute, 21 November 2007. 

28. L.E. Malvern, Introduction to the Mechanics of a Continuous Medium, Prentice-Hall, 1969.

29. Pete McKeighan, “Notched Tensile Properties of Tank Car Material from Full Scale Test No. 1 

(Car No. 3069) and Test No. 2 (3074) Oriented in the Transverse Direction,” NGRTC Project 

Memorandum, Southwest Research Institute, 20 November 2007. 

30. Pete McKeighan, “Notched Tensile Properties of Thickest TC128B Tank Car Material from 

Material Inventory for Fabricating Test Specimens,” NGRTC Project Memorandum, Southwest 

Research Institute, 6 December 2007. 
Applied Research Associates, Inc.

2010

269

31. Pete McKeighan, “Shear Tests on TC128B Normalized Material of Two Different Thicknesses,” 

NGRTC Project Memorandum, Southwest Research Institute, December 2007. 

32. Pete McKeighan, “Puncture Tests on Jacket Material (A1011, 90X) and Normalized TC128B Shell 

Material (two thicknesses)- Revised,” NGRTC Project Memorandum, Southwest Research Institute, 

17 December 2007. 

33. Pete McKeighan, “Puncture Behavior of 0.48-inch Thick A516 Material,” NGRTC Project 

Memorandum, Southwest Research Institute, 12 April 2008. 

34. Pete McKeighan, “Puncture Properties of Different 90x Stacked Layers and a Composite 

Structure,” NGRTC Project Memorandum, Southwest Research Institute, 23 March 2008. 

35. Kirkpatrick, S.W., Bocchieri, R.T., Sadek, F., Robert A. MacNeill, R.A., Holmes, B.S., Peterson, 

B.D., Cilke, R.W., and Navarro, C., “Analysis of Aircraft Impacts into the World Trade Center 

Towers,” NIST NCSTAR 1-2B, National Institute of Standards and Technology (NIST)

Investigation, October, 2005. 

36. Materials Research & Engineering, Inc., “Room Temperature Tensile Tests at different Crosshead 

Speeds,” Technical Report, Prepared for T.L. Anderson, Quest Reliability, LLC, MRE Report

#R07-02-003, February 19, 2007.

37. Pete McKeighan, NGRTC Project Memorandum, Southwest Research Institute, 5 December 2007. 

38. Cowper, G.R. and Symonds, P.S., “Strain-Hardening and Strain-Rate Effects in the Impact Loading 

of Cantilever Beams,” Brown University, Division of Applied Mathematics Report, P. 28, 

September 1957. 

39. Beremin, F. M., 1981, "Study of Fracture Criteria for Ductile Rupture of A508 Steel," in Advances 

in Fracture Research (ICF5), D. François, Ed., Pergamon Press , pp. 809-816.

40. Chaboche, J. L. et al., Editors, 1986, "Séminaire International sur l'Approche Locale de la Rupture," 

Conference Proceedings, Centre de Recherches EDF, "Les Renardières," Moret-sur-Loing, France.

41. Curran, D. R., Seaman, L. and Shockey, D. A, 1987, "Dynamic Failure of Solids," Physics Reports, 

Vol. 147, pp. 253-388.

42. Lemaitre, J., 1986, "Local Approach of Fracture," Engineering Fracture Mechanics, Vol. 25, 

pp. 523-537.

43. Mudry, F., 1985, "Methodology and Applications of Local Criteria for Prediction of Ductile 

Tearing," Elastic-Plastic Fracture Mechanics, L. H. Larson, ed., ECSC, EEC, EAEC, Brussels and 

Luxembourg, Belgium, pp. 263-283.

44. Hancock, J. W. and Mackenzie, A. C., 1976, "On the Mechanisms of Ductile Failure in HighStrength Steels Subjected to Multiaxial Stress States," Journal of the Mechanics and Physics of 

Solids, 24, pp. 147-169.

45. Rice, J. R., and Tracey, D. M., 1969, "On the Ductile Enlargement of Voids in Triaxial Stress 

Fields," J. Mech. Phys. Solids, 17, 201-217.

46. Giovanola, J. H., and Rosakis, A. J., Editors, 1992, Advances in Local Fracture/Damage Models for 

the Analysis of Engineering Problems, AMD-vol. 137, American Society of Mechanical Engineers, 

New York.

47. Giovanola, J. H., and Kirkpatrick, S. W., 1992, "Applying a Simple Ductile Fracture Model to 

Fracture of Welded T-Joints," in Advances in Local Fracture/Damage Models for the Analysis of 
Applied Research Associates, Inc.

2010

270

Engineering Problems, AMD Vol. 137, J. H. Giovanola and A. J. Rosakis, Eds., American Society 

of Mechanical Engineers, New York, pp. 285-303.

48. Giovanola, J.H., and Kirkpatrick, S.W., 1993, "Methodology for Evaluating Strength and Fracture 

Resistance of Weldments Using a Local Approach to Fracture," in Pressure Vessel Integrity-1993, 

PVP-Vol. 250, W. E. Pennel, S. Bhandari, G. Yagawa, Editors, American Society of Mechanical 

Engineers, New York, pp.157-171.

49. Giovanola, J.H., and Kirkpatrick, S.W., and Crocker, J.E., 1996, "Investigation of Scaling Effects in 

Elastic-Plastic Ductile Fracture Using the Local Approach," Proceedings of the First European 

Mechanics of Materials Conference on Local Approach to Fracture, Euromech-Mechamat '96, 

Fontainebleau-France, Sept. 9-11. 

50. Gurson, A.L., “Continuum Theory of Ductile Rupture by Void Nucleation and Growth: Part 1––

Yield Criteria and Flow Rules for Porous Ductile Media.” Journal of Engineering Materials and 

Technology, Vol. 99, 1977, pp. 2-15.

51. Tvergaard, V., “On Localization in Ductile Materials Containing Spherical Voids.” International 

Journal of Fracture, Vol. 18, 1982, pp. 237-252.

52. Tvergaard, V., “Material Failure by Void Growth to Coalescence.” Advances in Applied 

Mechanics, Vol. 27, 1990, pp. 83-151.

53. T.L. Anderson and S.W. Kirkpatrick, “Quantifying and Enhancing Puncture Resistance in Railroad 

Tank Cars Carrying Hazardous Materials,” Technical Report, Prepared for The Chlorine Institute, 

August 7, 2006.

54. Mackenzie, A. C., Hancock, J. W., and Brown, D. K., 1977, "On the Influence of State of Stress on 

Ductile Failure Initiation in High-Strength Steels," Eng. Fracture Mechanics, 9, 167-188.

55. M. Arcan, Z. Hashin and A. Voloshin, “A Method to Produce Uniform Plane-stress States with 

Applications to Fiber-reinforced Materials,” Experimental Mechanics, pp. 141-146, April 1978. 

56. Pete McKeighan, “BFCM Testing on Five Current/Candidate Tank Car Materials,” Project 

Memorandum, Southwest Research Institute, Prepared for Quest Reliability, 10 may 2007. 

57. Pete McKeighan, “Revised Role of Specimen Thickness on BFCM Energy – Blunt Tup,” Project 

Memorandum, Southwest Research Institute, Prepared for Quest Reliability, 17 May 2007. 

58. Pete McKeighan, “Influence of Initial Energy Level on BFCM Energy,” Project Memorandum, 

Southwest Research Institute, Prepared for Quest Reliability, 24 May 2007. 

59. Pete McKeighan, “Initial BFCM Results and the Role of Specimen Dimensions,” Project 

Memorandum, Southwest Research Institute, Prepared for Quest Reliability, 3 April 2007. 

60. Yu, H., Jeong, D.Y., Gordon, J.E., Tang, Y.H., "Analysis of Impact Energy to Fracture Un-Notched 

Charpy Specimens Made from Railroad Tank Car Steel," Proceedings of the 2007 ASME Rail 

Transportation Division Fall Technical Conference, RTDF2007-46038, September 2007. 

61. Hogg, P.J., “Composites for Ballistic Applications,” Composites Processing 2003, CPA, 

Bromsgrove, U.K. 21st March 2003. 

62. AMPTIAC Quarterly, Volume 8 No. 4, Army Materials Research: Transforming Land Combat 

through New Technologies, 2004. 

63. AMPTIAC Quarterly, Volume 9 No. 2, High Performance Fibers for Lightweight Armors, 2005. 
Applied Research Associates, Inc.

2010

271

64. Skaggs, R., Cheeseman, B., Yen, C.F., Dooley, R., Boyd, K., Walsh, S., Campbell, J., Hoppel, C., 

and Goetz. R., “Development of Mine Blast Protected Vehicle Structures for Future Combat 

Systems,” Proceedings for the Army Science Conference (24th), Orlando, Florida, 2005.

65. Carolan, M., Talamini, B., and Tyrell, D.C., "Update on Ongoing Tank Car Crashworthiness 

Research: Predicted Performance and Fabrication Approach," Proceedings of the 2008 IEEE/ASME 

Joint Rail Conference, Paper No. JRC2008-63052, Wilmington, Delaware, USA, April 2008.

66. Russell Walker, Greg Garcia, Satima Anankitpaiboon, and Kari Gonzales, “Tank Car Head 

Baseline Component Test,” Transportation Technology Center, Inc. Report No. P-08-009, April 

2008.

67. B.-J. Kim And S.-J. Heo, "Collapse Characteristics of Aluminum Extrusions Filled With Structural 

Foam for Space Frame Vehicles," International Journal of Automotive Technology, Vol. 4, No. 3, 

pp. 141−147 (2003). 

68. B.S. Holmes and J.K. Gran, "Developing a Research Safety Vehicle Structure with Scale Modeling 

Techniques," Stanford Research Institute Final Technical Report to Minicars, Inc., U.S. DOT, 

National Highway Traffic Safety Administration, Subcontract No. DOT-HS-5-01215, 1976. 

69. Thomas, M.A., Chitty, D.E., Gildea, M.L., and T’Kindt, C.M., “Constitutive Soil Properties for 

Cuddeback Lake, California and Carson Sink, Nevada,” Draft NASA Report No. NASA/CR—

2008–215345, Prepared by Applied Research Associates, Inc., February 2008. 

70. Pete McKeighan, “Foam Material Properties - Free Compression Stress-Strain Behavior,” NGRTC 

Project Memorandum, Southwest Research Institute, 19 October 2007. 

71. Pete McKeighan, “Foam Material Properties - Effect of Humidity and Different Specimen Sizes on 

the Free Compression Stress-Strain Behavior,” NGRTC Project Memorandum, Southwest Research 

Institute, 26 November 2007. 

72. Pete McKeighan, “Three Preliminary Experiments Examining the Flexure Properties of Composite 

and Solid Steel Sub-size Panels,” NGRTC Project Memorandum, Southwest Research Institute, 21 

December 2007. 

73. Pete McKeighan, “Fourth Small Panel Experiment - Flexure Properties of a Steel-Foam 

Composite,” NGRTC Project Memorandum, Southwest Research Institute, 12 February 2008. 

74. Pete McKeighan, “Experiments Examining the Effect of the Steel Surface Preparation on the

Flexural Properties of Bonded Composite Sub-size Panels,” NGRTC Project Memorandum, 

Southwest Research Institute, 27 August 2008. 

75. Fries, R., Walker, R., Anankitpaiboon, S., and Garcia, G., " Tank Car Head Impact Tests: Option 2 -

Multilayer Foam Head Protection; Option 3 - Nonoptimized EMS Head Protection," Transportation 

Technology Center, Inc. Report No. P-08-Draft, Sept. 2008. 

76. Queheillalt, D.T., Murty, Y., and Wadley. H.N.G., “Mechanical Properties of an Extruded 

Pyramidal Lattice Truss Sandwich Structure,” Scripta Materialia 58, pp. 76–79, (2008).

77. Wadley, H. N. G., Fleck, N. A., Evans, A. G., “Fabrication and structural performance of periodic 

cellular metal sandwich structures,” Comp. Sci. and Tech., 63, 2331-2343, 2003.
Applied Research Associates, Inc.

2010

272

78. Fleck N.A., Deshpande V.S., “The resistance of clamped sandwich beams to shock loading,”

Journal of Applied Mechanics, ASME; 71: 1-16, (2004).

79. H.J. Rathbun, D.D. Radford, Z. Xue, M.Y. He, J. Yang, V. Deshpande, N.A. Fleck, J.W. 

Hutchinson, F.W. Zok, A.G. Evans, “Performance of Metallic Honeycomb-Core Sandwich Beams 

Under Shock Loading,” International Journal of Solids and Structures 43, 1746–1763, (2006).

80. Tagarielli, V.L., Fleck N.A., Deshpande V.S., “The Collapse Response of Sandwich Beams with 

Aluminum Face Sheets and a Metal Foam Core,” Advanced Engineering Materials, Vol. 6, No. 6, 

pp. 440-443 (2004).

81. Z. Wei, V.S. Deshpande, A.G. Evans, K.P. Dharmasena, D.T. Queheillalt, H.N.G. Wadley, Y.V. 

Murty, R.K. Elzey, P. Dudt, Y. Chen, D. Knight, K. Kiddy, “The Resistance of Metallic Plates to 

Localized Impulse,” Journal of the Mechanics and Physics of Solids 56, 2074–2091, (2008).

82. Fries, R., Hybinette, C., Douglas, P., and Toms, R.., "Inspection of NGRTC Option 3," NGRTC 

Project Memorandum, Sept. 10, 2008. 

83. “User’s Manual for LS-DYNA MAT162 Unidirectional and Plain Weave Composite Progressive 

Failure Models,” Univ. of Delaware Center for Composite Materials, Materials Sciences 

Corporation, November, 2005. 

84. Yen, C.F., “Ballistic impact Modeling of Composite Materials,” Proceedings of the 7th International 

LS-DYNA Users Conference, Dearborn Michigan, May 19-21, 2002. 

85. Matzenmiller, A., Lubliner, J., and Taylor, R.L. (1995). “A Constitutive Model for Anisotropic

Damage in Fiber-Composites,” Mechanics of Materials, 20, pp. 125-152.

86. Xiao, J.R., Gama, B.A., and Gillespie, J.W. Jr., “Progressive Damage and Delamination in Plain 

Weave S-2 Glass/SC-15 Composites Under Quasi-Static Punch-Shear Loading,” Composite 

Structures Vol. 78, pp. 182–196, 2007. 

87. Waliul Islam, S.M., “A Quasi-Static Penetration Model for Ballistic Energy Absorption of Thick 

Section Composite,” Univ. of Delaware Center for Composite Materials, presented at the UD-CCM 

2005 research Reviews, March 2, 2005. 

88. Deka, L.J., S.D. Bartus, U.K. Vaidya, “Numerical Modeling of Simultaneous and Sequential MultiSite Impact Response of S-2 Glass/Epoxy Composite Laminates”, Composites and Polycon 2007, 

Oct. 17-19, 2007.

89. Gama, B.A., Xiao, J.R., and Haque, Md. J., “Experimental and Numerical Investigations on 

Damage and Delamination in Thick Plain Weave S-2 Glass Composites Under Quasi-Static Punch 

Shear Loading,” Army Research Laboratory Report No. ARL-CR-534, Prepared by the Univ. of 

Delaware Center for Composite Materials, February 2004. 

90. http://www.ccm.udel.edu/Tech/MAT162/SingleElement.htm