Flywheel Energy Storage Systems for Rail

Transcription

1 Imerial College London Deartment of Mechanical Engineering Flywheel Energy Storage Systems for ail Matthew ead November 2010 Thesis submitted for the Diloma of the Imerial College (DIC), PhD degree of Imerial College London 1

2 I declare that the research resented in this Thesis is my own work and that the work of others is roerly acknowledged and referenced. Matthew ead 2

3 Abstract In current non-electrified rail systems there is a significant loss of energy during vehicle braking. The aim of this research has been to investigate the otential benefits of introducing onboard regenerative braking systems to rail vehicles. An overview of energy saving measures roosed within the rail industry is resented along with a review of different energy storage devices and systems develoed for both rail and automotive alications. Advanced flywheels have been identified as a candidate energy storage device for rail alications, combining high secific ower and energy. In order to assess the otential benefits of energy storage systems in rail vehicles, a comutational model of a conventional regional diesel train has been develoed. This has been used to define a base level of vehicle erformance, and to investigate the effects of energy efficient control strategies focussing on the alication of coasting rior to braking. The imact of these measures on both the requirements of an energy storage system and the otential benefits of a hybrid train have been assessed. A detailed study of a range of existing and novel mechanical flywheel transmissions has been erformed. The interaction between the flywheel, transmission and vehicle is investigated using a novel alication-indeendent analysis method which has been develoed to characterise and comare the erformance of different systems. The results of this analysis roduce general design tools for each flywheel transmission configuration, allowing aroriate system configurations and arameters to be identified for a articular alication. More detailed comutational models of the best erforming systems have been develoed and integrated with the conventional regional diesel train model. The erformance of roosed flywheel hybrid regional trains has been assessed using realistic comonent losses and journey rofiles, and the fuel saving relative to a conventional train quantified for a range of energy storage caacities and ower-train control strategies. 3

4 Acknowledgments I would like to thank my suervisor, oderick Smith, for his suort and guidance. I am also very grateful to Keith Pullen for his invaluable knowledge, advice and enthusiasm. Many thanks go to my colleague and friend Pablo Martinez. His knowledge and insight has resulted in many thought-rovoking conversations and fruitful ideas. I would also like to thank Catherine Griffiths for her advice during the writing-u rocess and all the members of the Future ailway esearch Centre for roviding such a stimulating research environment. Finally, I owe an enormous debt of gratitude to my arents. This would not have been ossible without their endless suort and encouragement. 4

5 Nomenclature Symbol Unit Meaning E J Energy fc litres/car-km Fuel consumtion F N Force G - Seed ratio of gear air in control gearbox GPE J Gravitational otential energy h m height J kg m 2 Moment of inertia K - Seed ratio of gear air KE J Kinetic energy m kg Mass P W Power - Characteristic gear ratio of lanetary gearset r - Overall transmission seed ratio t s Time T Nm Torque U - Flywheel utilisation factor v m/s Velocity x m Dislacement η - Efficiency φ - Variator seed ratio ω rad/s Angular velocity 5

6 Abbreviations Abreviation ATOC BL CGB CVT DDC DMU DOD ESS FC FDC FESS FHT FMG FWC GA GHG GM HDT IAM ICE IEA IPCC IVT LDV Li-ion LPG ML NiMH OC PGS PST SSB TI SE SOC TC ULEV UMTA VLA Meaning Association of Train Oerating Comanies Branch-line Control gearbox Continuously variable transmission Dual differential couled Diesel multile unit Deth of discharge Energy storage system Fluid couling Final drive couled Flywheel energy storage system Flywheel hybrid regional train Flywheel motor generator Flywheel couled Genetic algorithm Greenhouse gas General Motors Hybrid regional diesel train Indeendent Analysis Method Internal combustion engine International Energy Authority Intergovernmental Panel on Climate Change Infinitely variable transmission Light duty vehicle Lithium ion Liquid etroleum gas Main-line Nickel metal hydride Outut coued Planetary gearset Power-slit transmission ail Safety and Standards Board ailway Technical esearch Institute Secific energy State of charge Torque converter Ultra low emission vehicle Urban Mass Transortation Administration Valve regulated lead acid 6

7 Table of contents Abstract... 3 Acknowledgments... 4 Nomenclature... 5 Abbreviations... 6 List of figures List of tables Introduction Background to energy consumtion in the UK rail network Proosed energy saving otions Objectives of the research Overview of Thesis structure Literature review Hybrid ower-train technology Overview of hybrid vehicle classification Energy storage devices Energy and ower caabilities of storage devices Additional considerations Develoments in hybrid rail vehicles Continuously variable transmissions for flywheel energy storage systems Fixed ratio gearbox with sliing clutch Variator technology Power-slit rinciles for a simle differential Advanced ower slit configurations Mechanical flywheel hybrid vehicles ail vehicle energy modelling Otimised control

8 Hybrid rail vehicle analysis ail vehicle modelling Summary Effects of driving strategy on conventional and hybrid diesel regional trains Aroach to vehicle modelling Vehicle dynamics module Power-train module Vehicle control module Descrition of route data Driving strategies for conventional vehicles Effect of timetabling constraints on fuel consumtion Simle vehicle model of a hybrid regional train Constant force regenerative braking with coasting Constant ower regenerative braking with coasting Effect of timetable restrictions on erformance of hybrid regional train equirements for HDT regenerative braking system Summary Indeendent analysis of clutch and brake controlled flywheel transmissions Brake-controlled flywheel transmissions Connection otions for differential gearing Aroach to indeendent analysis of PGS transmissions Analysis of multi-pgs brake-controlled transmission esults for single PGS brake-controlled flywheel transmission esults for multile PGS brake-controlled transmissions egional rail alication of flywheel system with brake-controlled transmission Indeendent analysis of clutch-controlled flywheel transmissions Clutch-controlled ower-slit transmission modes Multi-PGS clutch/brake controlled PST

9 CGB kinematic requirements Performance of CGB-controlled PST CGB torque and seed requirements Strategy for limiting CGB mass egional rail vehicle alication of CGB-controlled PST Summary Indeendent analysis of variator-controlled transmissions Direct variator-controlled transmission Single regime variator-controlled ower-slit transmissions Ideal variator-controlled ower-slit hase Indeendent analysis of single regime variator-controlled PSTs Comarison of single regime variator-controlled transmissions Single PGS, 2-regime variator-controlled transmissions Combined FDC/FWC 2-regime transmission Synchronous shift 2-regime transmissions Comarison of multile regime variator-controlled transmissions Multile-regime FDC/DDC transmissions egional rail alication of variator-controlled PSTs Summary Effect of vehicle and flywheel losses on FESS erformance Characterisation of vehicle and flywheel losses Flywheel resistance losses Vehicle resistance losses Methodology for assessing accuracy of Indeendent Analysis Method FESS erformance for regional train alication including losses FESS erformance for assenger car alication including losses Summary Investigation of flywheel hybrid regional train

10 7.1. FESS configurations considered for hybrid train Integration of FESS and conventional diesel regional train models Torque and seed deendence of FESS comonent efficiencies FESS control strategy esults of FHT modelling Effect of FESS control strategy on FHT erformance Effect of initial FESS condition on FHT erformance Comarison of mechanical FESSs for regional rail alication Summary Summary and conclusions Summary of research and contributions ecommendations for future work Aendices Aendix A Identifying exression for normalised CGB seed Aendix B Further results from the Deendent Analysis Method Aendix C Consideration of PGS meshing losses Bibliograhy

11 List of figures Figure 1-1 Proosed reductions in energy related CO 2 emissions in order to stabilise atmosheric concentration at 450 m [3] Figure 1-2 (a) World energy-related CO 2 emissions [4] and (b) Global transort energy use by mode in 2000 [2] Figure 1-3 Average energy intensity of assenger and freight transortation in IEA member countries [3] Figure 1-4 Proosed CO 2 reduction measures for transort in order to meet BLUE Ma scenario [3] Figure 1-5 Estimated breakdown of UK electric train energy usage after ower generation losses [11] Figure 1-6 Estimated breakdown of UK diesel train energy usage after engine losses [11] Figure 2-1 Classification of hybrid electric vehicles [16] Figure 2-2 Tyical series electric hybrid architecture Figure 2-3 Parallel hybrid architecture Figure 2-4 Examle of comlex ower-slit hybrid architecture Figure 2-5 Examles of flywheel design for (a) mechanical [33] and (b) electrical [34] FESSs Figure 2-6 Comarison of secific energy and ower for energy storage devices Figure 2-7 Tyical charge and discharge efficiency of a lead-acid battery as a function of SOC [51] Figure 2-8 Effect of SOC on secific ower of a Li-ion cell at 25 c [52] Figure 2-9 Effect of temerature on secific ower of a NiMH battery at 60% SOC [52] Figure 2-10 Imlementation of Li-ion battery modules in a vehicle alication [49] Figure 2-11 Tyical ower-train efficiencies for electrical and mechanical FESSs during discharge Figure 2-12 Overview of the NE Train layout and oerational erformance [67] Figure 2-13 (a) Lithium-ion battery system develoed by TI [70] and (b) Mitrac suercaacitor based energy storage unit [71] Figure 2-14 Illustration of conventional and hybrid shunting locomotives [75] Figure 2-15 (a) Illustration of ACT-1 flywheel-motor-generator unit [78] and (b) ULEV-TAP2 comlete electrical flywheel system [61]

12 Figure 2-16 PPM vehicle and ower-train configuration [79] Figure 2-17 Secific energy, secific ower and energy density for Li-ion suercaacitor and advanced flywheel devices and integrated units (arrows show the effect of control and ackaging requirements) Figure 2-18 Illustration of a full toroidal variator [82] Figure 2-19 (a) Illustration of ush-belt variator oeration showing low and high gear ratios [84] and (b) details of ush-belt construction Figure 2-20 Toroidal variator efficiency for an outut/inut seed ratio of 1 [85] Figure 2-21 Descrition of generic differential gearing unit Figure 2-22 Possible range of values for ractical simle and idler PGSs [95] Figure 2-23 Illustration of the ossible oerating regions for a variator-controlled outut couled ower-slit transmission with arrows showing direction and magnitude of ower flow (adated from [96]) Figure 2-24 Proosed flywheel hybrid ower-train for assenger car [93] Figure 2-25 Examle of the oeration of the Gyreacta in vehicle acceleration mode (simlified diagram does not include PGS-flywheel gearbox) [103] Figure 2-26 Illustration of roosed flywheel hybrid ower-train [104] Figure 2-27 Illustration of flywheel hybrid ower-train roosed by General Motors Figure 2-28 Flybrid mechanical FESS develoed for Formula 1 alication [109] Figure 2-29 Illustration of longitudinal dynamic forces Figure 3-1 Flow chart of calculation rocedure for forward-facing model structure consisting of control, ower-train and vehicle dynamics modules Figure 3-2 Secific fuel consumtion and max torque curve for a 315kW engine Figure 3-3 Power-train outut values of (a) tractive effort, (b) SFC, (c) transmission efficiency and (d) tank-to-wheel efficiency as a function of vehicle seed for ower notches 1, 3, 5 and Figure 3-4 Schematic illustration of vehicle control subsystem oeration Figure 3-5 Data for (a) Main-line and (b) Branch-line routes used in the simulations Figure 3-6 Examles of calculated vehicle seed rofiles for main-line route with no coasting (flat out) and coasting trigger deceleration values (D coast ) of 0.1 and 0.07 m/s Figure 3-7 Comarison of measured fuel consumtion (fc) with results of model using flat-out vehicle oeration matching the time-tabled dearture times for main-line (ML) and branch-line (BL) routes

13 Figure 3-8 Effect of increasing coasting and reducing maximum vehicle seed on relative journey time and fuel consumtion for main-line oeration with station dwell time of 60s Figure 3-9 Effect of increasing coasting and reducing maximum vehicle seed on relative journey time and fuel consumtion for branch-line oeration with station dwell time of 30s Figure 3-10 Illustration of the cause of fuel consumtion increase when timetable-limited coasting is used to achieve overall journey time Figure 3-11 Effect of coasting on main-line hybrid erformance with ESS ower limitations Figure 3-12 Effect of coasting on branch-line hybrid erformance with ESS ower limitations Figure 3-13 Calculated fuel saving and required energy caacity for hybrid train with 1MW ESS using constant ower and constant force braking strategies with various degrees of coasting on mainline (ML) and branch-line (BL) routes Figure 3-14 Braking energy available during aroach to (a) ML and (b) BL stations using uniform and timetable-limited coasting to achieve overall journey time; dotted lines shows minimum ESS caacity required to cature all available energy with uniform coasting (assuming 100% ESS efficiency) Figure 3-15 Hybrid fuel savings relative to conventional vehicle using timetable-limited coasting to achieve overall journey time (assuming 100% ESS efficiency) Figure 3-16 ML and BL fuel consumtion for conventional and hybrid vehicles using uniform and timetable (TT) limited coasting strategies (ESS ower and energy caacities as stated and assuming 100% efficiency) Figure 4-1 Examles of (a) brake and (b) clutch controlled flywheel transmissions Figure 4-2 Definition of relationshi between the flywheel rotor and the outut characteristics of the flywheel unit Figure 4-3 Schematic diagram of brake-controlled flywheel transmission with m differentials (arrows show the direction of ower flow during flywheel discharge using PGS 1) Figure 4-4 Examle of a ossible configuration for 2-PGS ring-brake controlled transmission (note: 0 < 1 < 2 < 0.5 in this configuration) Figure 4-5 Illustration of simle flywheel discharge/charge cycle with a 2-PGS ring-brake transmission, constant T fd and DOD ov = 75% Figure 4-6 Value of as a function of DOD ov for a single PGS brake-controlled gs transmission Figure 4-7 (a) discharge and (b) charging efficiency as a function of η gs and DOD ov for a 1-PGS brake-controlled transmission Figure 4-8 ound-tri efficiency of the 1-PGS brake-controlled transmission as a function of and DOD ov gs 13

14 Figure 4-9 The value of U as a function of DOD gs ov for the 1-PGS brake-controlled transmission Figure 4-10 Normalised energy dissiation at the control brake during discharging and charging of flywheel, as a function of and DOD gs ov for a 1-PGS brake-controlled transmission Figure 4-11 Total energy dissiation in the control brake during a full charge/discharge cycle er unit energy delivered to the vehicle during flywheel discharge as afunction of and DOD gs ov for a 1- PGS brake-controlled transmission Figure 4-12 Values of n x required to achieve maximum discharge as a function of DOD ov for a 5-PGS brake-controlled transmission Figure 4-13 discha max, rge x max U x and round-tri efficiency as functions of DOD ov for a 5-PGS brake-controlled transmission (using the n values secified in Figure 4-14) Figure 4-14 Total energy dissiated at transmission brakes in a full charge/discharge cycle er unit energy delivered to the vehicle during flywheel discharge as afunction of and DOD ov for a 5-PGS brake-controlled transmission Figure 4-15 Limitations on the secific energy caacity (SE) of a 1-PGS brake-controlled flywheel system for a regional rail alication (E avail = 40 MJ, η gs = 100%) Figure 4-16 Limitations on the secific energy caacity of 1-5, 10 and 15 PGS brake-controlled flywheel systems for a regional rail alication (E avail = 40 MJ, η gs = 100%) Figure 4-17 FDC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) Figure 4-18 FWC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) Figure 4-19 Illustration of DDC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) Figure 4-20 Examle of a 3-PGS ower-slit transmission with 4-seed CGB Figure 4-21 Influence of DOD ov on the CGB gearing ratios required for maximum efficiency Figure 4-22 The values of η eq for a 3-PGS, 4-seed CGB controlled transmission during flywheel discharge with gear ratios chosen for maximum efficiency (η A, η B, η C and η cgb are equal to 95%) Figure 4-23 Overall charge, discharge and round-tri efficiencies for the 3-PGS, 4-seed CGBcontrolled transmission (with η gs s and η cgb = 95%) Figure 4-24 Comarison of maximum ossible discharge efficiency as a function of DOD ov between 15-hase brake-controlled and 3-PGS, 4-seed CGB-controlled transmissions Figure 4-25 Maximum values of normalised seed, torque and rated ower at CGB x 14

15 Figure 4-26 Comarison of normalised rated variator ower as a function of U, calculated to achieve maximum discharge efficiency or equal max CGB torque (defined as the equalised modes case) in each ower-slit mode Figure 4-27 Values of normalised PGS ratio required to achieve equalised maximum CGB torques in all modes, and the resulting transmission efficiency Figure 4-28 Mass of system comonents as a function of DOD ov using a 3-PGS, 4-seed CGB transmission roviding a useful energy caacity of 30 MJ Figure 4-29 Maximum secific energy of FESS as a function of available energy using a brake/clutch controlled transmission with 4-seed CGB and 3 PGSs, and a 4-PGS ring-brake transmission (E avail = 40 MJ, η gs s and η cgb = 100%) Figure 5-1 Configuration for a direct (toroidal tye) variator-controlled transmission Figure 5-2 Performance characteristics of the direct variator flywheel transmission as functions of normalised gear ratio, K (η var = 85%) Figure 5-3 Definition of inut and outut couled configurations with arrows showing direction of ower-flow with no ower recirculation Figure 5-4 Characteristics of simle inut and outut couled PSTs oerating with no ower recirculation (φ t = 6.25) Figure 5-5 Schematic diagram of FDC variator-controlled PST configuration (arrows show direction of ower-flow during clutch/variator controlled flywheel discharge) Figure 5-6 Contour mas for a FDC variator-controlled PST showing (a) η discharge, (b) DOD ov, (c) η round-tri, (d) U and (e) P var, rated as functions of and eq (with η var = 0.85, η gs = 0.95) dotted line shows min(p var, rated ) w.r.t. U Figure 5-7 Illustration of the effect of eq values on transmission losses (shown as ercentage of transmission inut energy dissiated in comonents) Figure 5-8 Characteristics of FDC variator-controlled PST when P var, rated is minimised for a given U Figure 5-9 Values of and eq which minimise P var, rated for a given U Figure 5-10 FWC variator-controlled PST (arrows show direction of ower-flow during flywheel discharge) Figure 5-11 Characteristics of FWC variator-controlled PST when P var, rated is minimised for a given U Figure 5-12 and eq of a FWC variator-controlled PST which minimise P var, rated for a given U Figure 5-13 Normalised rated variator ower as a function of U for 1-regime flywheel transmissions Figure 5-14 Transmission efficiencies as a function of U for 1-regime flywheel transmissions

16 Figure 5-15 Variator seed ratio as function of time for the three single regime transmissions during flywheel discharge with otimum conditions for U = 60% and the same maximum flywheel KE Figure 5-16 Normalised rated variator ower and breakdown of transmission losses for single regime variator-controlled transmissions during flywheel discharge with U = 60% (transmission losses exressed as % of total transmission inut energy) Figure 5-17 Schematic diagram of ossible combined FDC/FWC variator-controlled PST configuration Figure 5-18 Schematic diagram of FDC synchronous shift PST configuration (arrows show direction of ower-flow during both regimes of flywheel discharge) Figure 5-19 Schematic diagram of FWC synchronous shift PST configuration (arrows show direction of ower-flow during flywheel discharge) Figure 5-20 Normalised rated variator ower and discharge efficiency as a function of U for 1 and 2 regime flywheel transmissions Figure 5-21 Variator seed ratio as function of time for the three 2-regime PSTs during flywheel discharge with minimised ower conditions for U = 60% using the same maximum flywheel KE 178 Figure 5-22 Breakdown of losses for 2-regime variator-controlled PSTs during flywheel discharge with U = 60% (losses exressed as % of total transmission inut energy) Figure 5-23 Possible configuration for a 3-PGS FDC/DDC variator-controlled PST Figure 5-24 Variator seed ratio as function of time during flywheel discharge with equalised modes for U = 60% Figure 5-25 Efficiencies, normalised rated variator ower and normalised PGS ratios as functions of U for 3-PGS FDC/DDC variator-controlled PST (η var = 0.85, η gs = 0.95) Figure 5-26 Secific energy caacity of variator-controlled PSTs (E avail = 40 MJ, η gs s = 95%, η var = 85%) Figure 5-27 Secific energy caacity of 3-PGS variator and CGB controlled PSTs (E avail = 40 MJ, η gs s and η cgb = 95%, η var = 85%) Figure 6-1 Schematic illustration of time arameters for analysis of FESS erformance including vehicle and flywheel losses Figure 6-2 (a) Flywheel utilisation and (b) transmission discharge (solid) and round-tri (dashed) efficiencies as a function of for 1-regime FDC variator-controlled PST (E useful = 20 MJ using IAM design tool) Figure 6-3 (a) ated variator ower and (b) useful energy delivered to vehicle during flywheel discharge (E useful, solid line) and minimum braking energy required to fully recharge flywheel (E recharge, dashed line) as functions of for 1-regime FDC variator-controlled PST (E useful = 20 MJ using IAM design tool) Figure 6-4 (a) Overall round-tri system efficiency and (b) secific energy of FESS as a function of for 1-regime FDC variator-controlled PST (E useful = 20 MJ using IAM design tool)

17 Figure 6-5 Maximum SE as a function of useful energy caacity for regional train alication using 1-PGS FDC and 3-PGS FDC/DDC variator-controlled FESSs with available braking energy limited to 40 MJ Figure 6-6 Tyical discharge/charge event for assenger car in urban driving (extracted from ECE urban drive cycle) and associated drive cycle characteristics used in the DAM Figure 6-7 Maximum SE as a function of useful energy caacity for 3-PGS variator-controlled PST in a assenger car alication with available braking energy limited to 60 kj Figure 7-1 Variator efficiency ma (in ercent, including hydraulic control system) as a function of seed ratio and normalised inut torque [133] Figure 7-2 PGS efficiency as a function of maximum torque in the three PGS branches divided by the PGS torque rating (exerimental data from Mantriota [139]) Figure 7-3 Overview of the FHT model structure Figure 7-4 Effect of FESS control strategy on fuel saving for 3-PGS variator-controlled FESS relative to conventional vehicle using a range of uniform coasting conditions Figure 7-5 Effect of FESS control strategy 1 on the maximum oerating seed of the variator (relative to the rated value) in a 3-PGS variator-controlled FESS using a range of uniform coasting conditions Figure 7-6 An examle of the effect of FESS control strategy (CS) on FHT erformance (with D coast = 0.1 m/s 2 ) Figure 7-7 Fuel saving relative to conventional vehicle using 3-PGS variator-controlled transmission and the stated FESS useful energy caacities with flywheel initially either fully charged or emty Figure 7-8 Fuel saving relative to conventional vehicle for main-line oeration using 4-PGS brakecontrolled and 3-PGS variator-controlled PSTs and the stated FESS useful energy caacities with equal initial and final flywheel SOC Figure 7-9 Fuel saving relative to conventional vehicle for branch-line oeration using 4-PGS brake-controlled and 3-PGS variator-controlled PSTs and the stated FESS useful energy caacities with equal initial and final flywheel SOC

18 List of tables Table 1-1 The relationshi between emissions and climate change [2] Table 1-2 Traction energy consumtion in the UK railways (rimary fuel energy) [10-11] Table 2-1 Potential fuel savings for diesel light-duty vehicle (LDVs) using range of currently available technologies, and associated cost increase [3] Table 2-2 Feasible energy storage devices and the required ESS transmissions for assenger vehicle alications Table 2-3 Comarison of cell characteristics for the main battery chemistries Table 2-4 elationshi between and simle PGS [95] Table 2-5 elationshi between and PGS with idler lanets [95] Table 3-1 Vehicle arameters for model of 3-car diesel-hydrodynamic vehicle Table 3-2 Effect of coasting strategy and timetable restrictions on fuel consumtion Table 4-1 Direction of ower flow in PGS branches (+ve in, -ve out) and associated torque equations including the effect of constant PGS transmission losses for all ossible cases Table 4-2 Torque and energy dissiation relationshis for a hase of oeration using brakecontrolled transmission Table 4-3 Summary of normalised arameter deendence for 1-PGS brake-controlled transmission Table 4-4 Exressions for eq and η eq for FDC, FWC and DDC clutch-controlled ower-slit modes Table 4-5 Exressions for (ω cgb,o ) and (T cgb,o ) for FDC, FWC and DDC clutch-controlled owerslit modes Table 4-6 The limiting values of eq and η eq for the equivalent PGS brake-controlled transmission Table 4-7 Control sequence for 3-PGS, 4-seed CGB clutch/brake controlled PST during flywheel discharge (X indicates the element is engaged) Table 4-8 Exressions for CGB gear ratios (where eq, j is the equivalent value of the oerating hase using the j th CGB gear in the resective ower-slit mode) Table 4-9 equired sread of gear ratios in the CGB for DOD ov = 80% (assuming a constant actual or equivalent PGS efficiency of η x for all hases) brake-controlled hases are shown in bold Table 4-10 Normalised torque exressions for 3-regime (FDC/DDC/DDC) CGB-controlled PST

19 Table 4-11 Parameters of 3-regime CGB controlled transmission used in equalised modes analysis Table 4-12 Values used in analysis of 3-regime CGB-controlled flywheel transmission for a regional rail vehicle Table 5-1 Characteristic relationshis for simle inut and outut couled PSTs [94] Table 6-1 Vehicle data for Ford Focus assenger car [107] Table 7-1 Secifications for a 3-PGS variator-controlled FESS with a range of useful energy caacities in order to achieve minimum mass (using the design tool data resented in Chater 5 with the stated maximum variator seed and PGS torque) Table 7-2 Secifications for a 4-PGS brake-controlled FESS with a range of useful energy caacities in order to achieve minimum mass (using the design tool data resented in Chater 4 with the stated maximum PGS torque) Table 7-3 Energy breakdown (in MJ s) for conventional and flywheel hybrid regional trains on main-line (ML) and branch-line (BL) routes using uniform coasting driving strategy (FESS useful energy caacity of 20 MJ, equal initial and final flywheel SOC)

20 1. Introduction This Thesis describes an investigation into how novel flywheel energy storage systems may rovide a means of reducing energy consumtion in rail vehicles through the imlementation of regenerative braking. This technical challenge is connected to both the environmental and economic ressure to reduce vehicle emissions and energy consumtion related to transort. In recent years, the issue of global warming has generated great scientific and ublic interest regarding the effects of human activity on the environment. The roduction of greenhouse gasses (GHGs) has been linked with climate change, and in the light of scientific evidence and reorts such as the Stern eview on the economic imacts of climate change [1], a general consensus aears to have been reached on the need to limit such emissions. Table 1-1 shows the relationshi between global temerature rises and the atmosheric concentration of GHGs as redicted by the Intergovernmental Panel on Climate Change (IPCC) [2]. The target levels of CO 2 emissions required by the year 2050 in order to limit these temerature rises are also shown relative to 2000 levels. Figure 1-1 is taken from a 2008 International Energy Agency ublication [3] and resents a roadma for limiting global GHG emissions over this eriod in order to stabilise atmosheric CO 2 concentrations at 450 m, as illustrated by the BLUE Ma scenario. This is shown relative to a Baseline scenario, which corresonds to the redicted global emissions that will occur with the energy and climates olicies that were imlemented by It is clear that in order to limit CO 2 levels, large reductions in emissions are required from both the ower generation sector and end-users. The reduction of transort emissions is a key element in meeting this target, and illustrates the ressing need to develo technologies and strategies to imrove vehicle energy efficiency. Table 1-1 The relationshi between emissions and climate change [2] 20

21 Figure 1-1 Proosed reductions in energy related CO 2 emissions in order to stabilise atmosheric concentration at 450 m [3] Transort reresents a large roortion (23% in 2007) of global energy-related CO 2 emissions, as shown in Figure 1-2. There has also been steady growth in transort emissions since the 1970s, driven rimarily by increasing car ownershi in both develoed and develoing countries. Gt CO2 (a) (b) Figure 1-2 (a) World energy-related CO 2 emissions [4] and (b) Global transort energy use by mode in 2000 [2] Figure 1-2(b) shows the breakdown of global rimary energy use in the transort sector by mode. It is clear that transort energy use is dominated by road vehicles, esecially assenger cars (LDVs). The large market for these vehicles, combined with rising fuel costs and tightening emissions legislation has driven research and develoment within both industry and academia focussed on imroving the efficiency of road vehicles. The small contribution of rail (1.5% of transort energy use) illustrates its limited otential to affect overall GHG emissions. The relatively small market and long service life of rail vehicles also mean that oortunities for imroving energy efficiency are restricted. 21

22 ail does however have an imortant role to lay in current and future transort strategy due to its inherent advantages as a mass transortation system. These include; i. The efficiency of the steel-on-steel wheel-rail contact which avoids the hysteresis losses associated with neumatic tyres, ii. The aerodynamic rofile of rail cars in convoy leading to low vehicle drag er unit mass, iii. The use of an electrical ower delivery system which allows high energy efficiency and the exloitation of low carbon ower generation. These advantages lead to low energy intensity (defined here as the rimary energy use er assenger-km) comared to other modes of transort. This is shown in Figure 1-4 for assenger and freight transortation modes using data for member countries of the International Energy Agency. Figure 1-3 Average energy intensity of assenger and freight transortation in IEA member countries [3] The advantages of rail transort mean that it is set to lay an imortant role in achieving the transort emissions reduction required to meet the BLUE Ma scenario described in Figure 1-1. These reductions can be achieved through imroving engine and ower-train efficiencies, using alternative fuels (i.e. bio and synthetic fuels) and enabling shifts between different transort modes. The required contribution of these measures is shown in Figure

23 Figure 1-4 Proosed CO 2 reduction measures for transort in order to meet BLUE Ma scenario [3] The emission reduction due to modal shifts shown in Figure 1-5 is achieved largely through the greater use of high seed rail for journeys u to 750 km, greater use of rail to transort freight and the greater use of trams, buses and light rail in urban areas. These modal shifts are combined with redicted imrovements in the energy efficiency of rail transortation. In order to achieve these targets growth is required in international railway caacity. The cost of building the required infrastructure is however relatively high, with high seed rail costing in the region of 5 to 50 million US$/km [5-6] deending on toograhy, the cost of land (which can be articularly high around urban areas) and labour costs. The cost of conventional rail systems are significantly lower, but are still affected by these factors. The develoment of existing rail networks is also constrained by the need to maintain rail services. This can be seen in the UK, where the ercentage of electrified track is currently only 40% and lans to increase electrification are limited to a few mainline routes [7-8]. ather than relying on electrification to imrove energy efficiency, an alternative and otentially more cost-effective aroach is to develo clean and fuel efficient self-owered rail vehicles. This requires a thorough understanding of current traction systems in order to identify the main areas of energy loss and assess the benefits of otential future technologies and oerating strategies aimed at imroving fuel efficiency. The analysis resented in this Thesis has been erformed in the context of the UK railways, the key features of which are described in the following sections. 23

24 1.1. Background to energy consumtion in the UK rail network Transort accounted for 21% of total greenhouse gas emissions in the UK [9], and has seen steady growth over the last 20 years. The total energy consumed by rail vehicles in the UK is around 15 million MWh er year, and is divided aroximately equally between electric and diesel traction [10]. Electric vehicles can be subdivided into high voltage (25kV AC) intercity trains, and both high and low voltage (750V DC) multile units which oerate redominantly regional and commuter services. A relatively small amount of energy is used by electric freight locomotives. Diesel assenger vehicles can be divided according to the transmission and service tye, with diesel-electric trains roviding intercity services and diesel-hydrodynamic vehicles roviding the majority of regional services. A considerable amount of energy is also used by diesel freight trains. The Association of Train Oerating Comanies (ATOC) roduce yearly energy consumtion data for rail vehicle in the UK, and the figures for 2004/05 are shown in Table 1-2. This data is for the rimary fuel energy consumed which includes losses associated with the generation and distribution of electricity, but does not include the energy use associated with the roduction and distribution of the diesel fuel or the fuel used at the ower station. Diesel Electric Energy consumtion (million MWh) Intercity DEMU 2.2 egional DHMU 2.5 Total assenger 4.7 Freight 2.7 Intercity 25kV 2.3 EMU 25kV 1.3 EMU 750V 3.9 Total assenger 7.5 Freight 0.3 Total 15.2 Table 1-2 Traction energy consumtion in the UK railways (rimary fuel energy) [10-11] A 2007 reort by the ailways Safety and Standards Board (SSB) [11] rovides detailed estimates of the total energy losses for UK rail vehicles. The estimated ower generation efficiencies for both electric and diesel traction are as follows: Average electric ower generation efficiency = 40% 24

25 Average diesel engine fuel efficiency = 32% These estimates are made on the basis of current technology (and tyical duty-cycle in the case of diesel engines), and will be subject to future develoments in electrical generation (such as the ossible exansion of renewable and nuclear caacity) and IC engine technology resectively. A shift in ower-train technology to hybrid or fuel cell systems also has the otential to raise the average oerating efficiency of self-owered trains, thereby reducing rimary energy consumtion. Once the ower generation losses have been taken into account, a breakdown of where the sulied electrical or mechanical energy is used in the vehicle gives an indication of where efforts to imrove efficiency should be focussed. This is shown according to vehicle and service tyes in Figures 1-6 and 1-7 for electric and diesel trains resectively. Intercity (25kV) egional (25kV) Suburban (750V) Figure 1-5 Estimated breakdown of UK electric train energy usage after ower generation losses [11] Diesel Intercity Diesel egional Figure 1-6 Estimated breakdown of UK diesel train energy usage after engine losses [11] It is clear from Figures 1-6 and 1-7 that after ower generation losses (at either a ower station or the diesel engine) the combined traction requirements of accelerating the vehicle and overcoming running resistance account for the majority of energy used. The balance 25

26 between these inertia and resistance losses is seen to deend on the service tye. For intercity services the train is likely to send long eriods at high seed, leading to high resistance losses. For regional and suburban services the train is likely to sto more frequently, dissiating large amounts of kinetic energy. As of 2009 all the 25kV AC network and a large roortion of the 750V DC network is comatible with regenerative braking [8]. The inertia losses shown in Figure 1-6 are therefore likely to be higher than current levels, although factors such as ower and efficiency limitations in the electrical ower-train limit the roortion of braking energy that can be regenerated. A further issue known as line recetivity also limits regeneration in DC systems, resulting in regenerated electrical energy being dissiated in resistors if there is insufficient ower demand from other vehicles in the same section of track. While this issue can be tackled by using inverting equiment at the DC substation [12-13] this technology is not used in the UK railways. For diesel trains, it is clear from Figure 1-7 that significant fuel savings could be achieved through the efficient cature and reuse of vehicle braking energy. This can be achieved (along with a range of other otential benefits) by imlementing hybrid ower-train systems in self-owered rail vehicles. In fact, the otential of hybrid technology to reduce fuel consumtion and emissions has been given as one of the reasons for not ursuing widesread electrification of the UK network [14]. Hybridisation is therefore included in the range of measures which have been identified as suitable for reducing the energy consumtion of the UK railways, as discussed in the following section Proosed energy saving otions The desire for a sustainable and efficient UK railway network has led to the investigation of measures aimed at reducing the traction energy consumtion of both electric and diesel rail vehicles. A 2007 reort roduced by the ail Safety and Standards Board (SSB) [11] has identified a wide range of strategies and technologies which could be imlemented to reduce overall energy consumtion in the UK railways. The otential of each otion was assessed in terms of estimated network-wide energy savings and timescale for introduction. Each measure considered was then ranked on a cost-benefit basis. The results suggest that the oortunities to reduce energy consumtion in a cost effective way fall into one of two categories short term which involves changes to the way current vehicles are oerated, and long term which relates to new technologies which require more develoment or can 26

27 only be introduced on new rolling-stock. The most romising measures identified in the reort are summarised below; Short term energy saving measures Energy efficient driving techniques and timetabling Matching train caacity to demand (e.g. shorter trains during off-eak eriods) educing diesel engine idling Long term energy saving measures egenerative braking with on-board energy storage for diesel trains Intelligent engine control on vehicles with distributed ower-train educing vehicle weight er seat educing vehicle drag It is imortant to note that some of these measures are inter-related. For examle, the introduction of hybrid systems is likely to be combined with efficient ower-train control strategies, while the additional comonents (articularly the energy storage devices) will tend to increase the mass of the vehicle ower-train. The focus of this Thesis is on assessing the otential of imlementing regenerative braking with on-board energy storage for diesel trains. Analysis has therefore been erformed using detailed comutational modelling of a tyical UK regional diesel train, where the benefits of this technology are exected to be greatest due to the high braking losses that occur in conventional vehicles of this tye. The develoment of systems to cature, store and reuse energy that would otherwise be lost is constrained by a wide range of technological and ractical issues. The tye of energy storage device used and the transmission connecting it to the vehicle drive are critical in determining the hybrid system characteristics and ultimately the fuel consumtion of the vehicle. Many different tyes of energy storage device are available. Selection of the most aroriate device for a given alication is achieved by consideration of the following: The vehicle characteristics and duty cycle requirements The degree of hybridisation and the ower-train control strategy The efficiency of the energy storage system Practical issues relating to safety, reliability, weight and installation requirements The cost, lifesan and maintenance requirements of the system 27

28 The ossibility of using diesel-electric hybrid technologies similar to those develoed in the automotive industry has been investigated by the SSB for a new fleet of UK diesel regional trains [15]. This reort concluded that new hybrid vehicles reresented oor value for money due to the small order size combined with uncertainties over ossible future electrification and otions to extend the life of existing vehicles. Simler hybrid systems involving intelligent control of conventional distributed ower-trains and self contained bolt-on regenerative braking systems should therefore be considered as an alternative means of reducing fuel consumtion. The research described in this Thesis identifies advanced flywheel energy storage technology and diesel hybrid ower-train architectures as suitable for rail vehicles, and the otential alication of aroriate systems to diesel-owered trains oerating UK regional services is investigated at both system and oerational levels. A key relationshi which has not been fully addressed in the literature is the effect of driving strategy on the design requirements and otential benefits of hybrid regional diesel trains (HDTs). As energy efficient driving has been identified as a short term measure which can be alied to existing vehicles through imroved driver training and/or in-cab advisory systems, this should be considered as a baseline against which to assess the fuel savings ossible through hybridisation. This has also been assessed using detailed comutational modelling of a conventional diesel-hydrodynamic regional train oerating a reresentative UK service, allowing the effect of hybridisation to be isolated for a range of driving strategies Objectives of the research The research described in this Thesis focuses on assessing the erformance of hybrid diesel regional train (HDTs) and a range of flywheel energy storage systems using mechanical transmissions. The aims are summarised below; 1. Investigate the effect of driving strategy on the fuel consumtion of conventional and hybrid diesel regional trains in order to identify the realistic fuel savings ossible through hybridisation, and the energy and ower requirements of aroriate regenerative braking systems. 2. Investigate the comromises in the secification of a mechanical flywheel energy storage system (FESS) and identify flywheel and transmission arameters in order to maximise the secific energy caacity of the system. 28

29 3. Develo design tools to allow the comarison of different transmission systems, and secify aroriate FESS arameters for a given alication. 4. Assess the benefits of the best erforming mechanical FESSs through detailed comutational modelling of a flywheel-hybrid regional diesel train. A summary of the Thesis content is resented below, roviding an overview of the methods used to achieve these objectives Overview of Thesis structure Chater 2 A review of hybrid systems (including consideration of energy storage devices and transmission otions), comutational methods for calculating rail vehicle energy consumtion and the otential of efficient driving strategies. Chater 3 An analysis of the effect of driving strategy on vehicle fuel consumtion using detailed comutational models of a conventional diesel-hydrodynamic regional train and a generic hybrid diesel train. Chater 4 An analysis of a range of clutch and brake controlled mechanical transmissions for FESS alications. A novel indeendent analysis method (IAM) has been imlemented to rovide an absolute basis for the comarison of different transmission configurations and to identify aroriate system arameters for a given alication. Chater 5 An analysis of a range of variator-controlled transmissions for mechanical FESS alications using the IAM. Chater 6 Validation of the IAM results through the detailed comutational analysis of the effect of flywheel losses and vehicle resistance on the erformance of mechanical FESSs for secific alications. 29

30 Chater 7 A detailed comutational analysis of roosed mechanical hybrid regional diesel trains, integrating the mechanical FESS models (using arameters identified from the IAM) with the conventional diesel regional train model. The effect of driving strategy and energy storage caacity on fuel consumtion is investigated for realistic main-line and branch line services. Chater 8 Conclusion, contributions of the research and recommendations for future work. 30

31 2. Literature review egenerative braking and energy efficient driving have been identified in Chater 1 as two key measures for reducing rail energy consumtion and harmful emissions. This literature review is therefore divided into two main sections. Section 2.1 comrises an overview of hybrid ower-train systems and comonents, while Section 2.2 considers methods of analysing rail vehicle dynamics, ower-train oeration and driving strategies in order to assess energy consumtion Hybrid ower-train technology As discussed in Chater 1, this Thesis focuses on the fuel consumtion of regional rail vehicles where there is a strong case for imlementing hybrid systems to cature braking energy that would otherwise be dissiated. The recovery, storage and reuse of braking energy can be imlemented using a wide range of technologies and control systems, and has the otential to allow some or all of the following benefits to be achieved; Potential benefits of hybridisation educed fuel use and emissions due to; o The cature and reuse of braking energy o educed engine idling when no traction ower is required o More efficient rovision of auxiliary ower requirement o Efficient control of the engine during owering o Downsizing of the installed engine caacity Imroved vehicle acceleration educed engine noise and emissions around stations educed wear on friction brake comonents The viability of these benefits deends uon the tye of regenerative braking system imlemented. A simle bolt-on system oerating in arallel with a conventional owertrain may allow the recovery of braking energy and achieve a reduction in fuel consumtion, but will be less flexible than a fully integrated hybrid ower-train where the caacity and oerating range of the engine can be otimised. There are also a number of limiting factors when considering regenerative braking systems for regional rail vehicles, as discussed below. 31

32 Constraints on the hybridisation of rail vehicles Increasing vehicle mass (as this affects traction energy consumtion and infrastructure maintenance) Increasing vehicle comlexity (as this affects the vehicle reliability and maintenance costs) Increasing life cycle costs Vehicles are required to oerate on a range of routes and services, which limits the otential otimisation of ower-train comonents and control strategies The choice of energy storage device and the architecture of the regenerative braking system will have a strong influence on these factors, and are therefore extremely imortant considerations. Existing and roosed devices and systems for both rail and automotive vehicles are exlored in the following literature review Overview of hybrid vehicle classification Much of the develoment of hybrid ower-train technology has been focussed on automotive vehicles. It is therefore instructive to consider the otential of hybrid systems for this alication alongside a range of other roosed technologies aimed at imroving fuel efficiency. The data in Table 2-1 is taken from a recent IEA reort [3] and shows the otential fuel savings and the redicted cost of vehicle and diesel ower-train technologies. Energy Saving Measures Diesel hybrid Non-engine imrovements Tyres 0.5-4% Aerodynamics 0.5-4% Lights 0-2% Efficient aux % 25% weight reduction 10-11% Engine imrovements Higher comression ratio, no throttle 2-3% Direct injection 7-8% Imroved combustion 3-4% Hybrid system 15-17% Total imrovement comared to baseline vehicle 40-55% Cost of imrovements ($) Table 2-1 Potential fuel savings for diesel light-duty vehicle (LDVs) using range of currently available technologies, and associated cost increase [3] 32

33 Many of these measures are equally alicable to diesel-owered rail vehicles, although the resulting imrovements in fuel economy will obviously differ. With resect to automotive vehicles Table 2-1 shows that hybridisation can achieve large fuel savings, esecially when combined with other measures. This however, results in a considerable estimated increase in vehicle cost of around 20% for a tyical US vehicle rice of $25,000. It is clear that there is a comromise between imroving vehicle efficiency and increasing the vehicle cost which is highly deendent on the characteristics of the hybrid system used. A range of hybrid-electric systems have been roosed and develoed for automotive alications as shown in Figure 2-1. This rovides a good basis for understanding different hybrid configurations and system requirements. Vehicle classification Proulsion device Energy source IC engine vehicle Micro hybrid Mild hybrid Full hybrid Electric vehicle Fuel cell electric vehicle Figure 2-1 Classification of hybrid electric vehicles [16] While Figure 2-1 is focussed on hybrid-electric systems, the vehicle classifications also aly to more general hybrid configurations using different transmission tyes and energy storage devices, and are discussed in more detail below. Series Hybrid The series hybrid vehicle is ossible when the medium of energy transfer in the vehicle transmission is the same as that of the energy storage device. This can be achieved using either a hydraulic or electrical system, where energy storage is achieved using a hydroneumatic accumulator or batteries/caacitors resectively. An examle of this 33

34 configuration is shown below, with arrows illustrating the ossible direction of ower flows in the system. Electrical transmission Power source G PE M/G Vehicle drive Energy storage device G generator M/G motor/generator PE ower electronics Figure 2-2 Tyical series electric hybrid architecture Automotive vehicles using the electric hybrid architecture shown in Figure 2-2 are often referred to as lug-in hybrids, as they are designed to oerate rimarily as electric vehicles. In this case, the ICE or fuel cell has a relatively low ower rating, and is used to extend the range of the vehicle by sulementing the tractive ower and/or recharging the battery. The series hybrid architecture can also be used for multile energy storage devices with comlementary characteristics [17]. An examle of this aroach is the use of batteries to achieve good energy storage caacity and allow a long vehicle range, combined with suer-caacitors [18-19] or flywheel motor/generator devices [20] to allow high transient ower flows during acceleration and braking. Full Hybrid As illustrated in Figure 2-1, a full hybrid can be defined as a ower-train where the ower source can deliver ower to the vehicle via a searate ath than the energy storage device. In the simlest form, this corresonds to a arallel hybrid configuration where the engine/fuel cell and energy storage devices have searate transmissions. 34

35 Power source Energy storage device Transmission Transmission Vehicle drive Figure 2-3 Parallel hybrid architecture More comlex full hybrid configurations are also ossible through the use of mechanical differentials such as lanetary gear sets (PGSs) to rovide multile aths for ower-flow. An examle of this is the comlex ower-slit hybrid architecture shown in Figure 2-4. IC Engine M/G D Battery PE M/G electric motor/generator PE ower-electronics D mechanical differential M/G Vehicle drive Figure 2-4 Examle of comlex ower-slit hybrid architecture This arrangement rovides a greater degree of flexibility in the oeration of the engine and ower-train than is ossible with a arallel hybrid, allowing fuel consumtion to be otimised through the use of aroriate control strategies [21]. Mild Hybrid Hybrid vehicles of this classification are essentially conventional, ICE owered vehicles with a low owered electrical motor/generator and limited energy storage caacity arranged in a arallel hybrid configuration. The energy storage device can be recharged from the engine or by regenerative braking, and can be used to rovide auxiliary ower load (enabling engine shut down when stationary) and limited tractive ower assistance during acceleration (enabling a small degree of engine down-sizing). 35

36 Micro Hybrid Micro hybrid vehicles are similar to mild hybrids but the energy storage devices have a smaller caacity, can only be recharged from the engine and do not rovide any tractive ower to the vehicle. Fuel savings are achieved through the alication of engine shutdown to minimise fuel consumtion when no tractive ower is required. A key comonent of all these hybrid ower-trains is the energy storage device. This is likely to account for a significant roortion of the additional cost and weight of a hybrid vehicle, and will strongly influence the transmission, configuration and ackaging of the ower-train. The different energy storage otions are discussed in the following section Energy storage devices Develoments in the automotive industry have largely focussed on using electrical systems and battery technology to comliment internal combustion engines, which has become the most established hybrid technology [22]. There are however a range of ractical energy storage otions, each of which have advantages and disadvantages. The four most established tyes of energy storage device are shown in Table 2-2, along with the tye of transmission required for integration in an energy storage system (ESS). Energy storage device Electro-chemical battery Caacitor Flywheel Hydro-neumatic accumulator ESS transmission Electrical Electrical Electrical or Mechanical Hydrostatic Table 2-2 Feasible energy storage devices and the required ESS transmissions for assenger vehicle alications A brief descrition of each of these energy storage technologies is given below. Batteries For ractical hybrid vehicle alications three main tyes of electro-chemical battery have been identified; lead-acid, nickel-metal-hydride (NiMH) and lithium-ion (Li-ion) [23-25]. Lead-acid batteries are a well established technology, and are widely used as an auxiliary storage device in automotive alications. They consist of lead and lead oxide 36

37 electrodes with a liquid electrolyte containing sulhuric acid. Modern valve-regulated lead-acid (VLA) batteries are sealed units which are more robust than a simle lead-acid cell and rovide an economical, low maintenance otion for alications where energy caacity er unit mass is not critical. Nickel-metal-hydride and lithium-ion are both tyes of dry cell batteries where the electrolyte is in the form of a aste. For NiMH the electrodes consist of nickel oxidehydroxide and an intermetallic comound, with an alkaline electrolyte. These batteries are widely used in consumer electronics and electric vehicles due to their reasonable cost and good energy storage caacity. Li-ion batteries consist of carbon and metal-oxide electrodes with a lithium salt electrolyte, and achieve better energy caacity than NiMH batteries but at a higher cost. They are also used in consumer electronics (esecially in roducts where low size and mass are imortant), and are beginning to see use in electric and hybrid vehicles such as the Tesla sorts car [26] and the Nissan Leaf assenger car [27]. Suercaacitors Suercaacitors are high erformance electrolytic caacitor devices consisting of two electrodes of activated carbon (which has a very high surface area) searated by an electrolyte layer. Unlike in an electro-chemical battery, no chemical reactions occur during oeration and the boundary between the electrolyte and electrode acts as an insulator. Alying a voltage across the electrodes therefore causes charge to accumulate, storing electrical energy. The high surface area of the electrode and the thin insulating boundary layer result in a very high caacitance. The energy storage caacity of these devices is however limited by a maximum cell voltage of around 3V [28]. Advanced flywheels Advanced flywheels use rotors constructed from high strength comosite materials which rotate at high seed (tyically 10,000 to 100,000 rm), thereby storing kinetic energy. The rotor is contained within a vacuum environment in order to minimise aerodynamic losses, and the flywheel casing is designed to contain the rotor in the event of a catastrohic failure. This tye of flywheel technology has been develoed and demonstrated in several research rograms [29-32], and advanced flywheels have been shown to achieve much higher energy storage er unit mass than conventional steel flywheels. 37

38 The energy caacity of a flywheel deends on both the rotor inertia (related to the material density and geometric shae) and the maximum safe rotational seed (related to material strength and shae of the rotor, the alied safety margin and the bearing system used). The design of the flywheel will therefore affect the secific energy, but a range of other factors such as controlling failure modes and the costs of comonents, materials and manufacturing must also be considered. The design and erformance of advanced flywheels are also deendent on the choice of transmission system. Practical otions for transferring ower to and from the flywheel involve the use of either; i. An electrical motor/generator integrated with the rotor ii. A mechanical transmission caable of oerating over a continuous ratio range Figure 2-5 shows examles of advanced flywheel devices develoed for electrical and mechanical systems. (a) Containment discs (b) Flywheel rim Vacuum seals Touchdown ring Flywheel hub Flywheel bearing Vacuum ort Seed sensor Containment ring Flywheel housing Figure 2-5 Examles of flywheel design for (a) mechanical [33] and (b) electrical [34] FESSs The electrical flywheel otion can only oerate in a system with a full electrical transmission. The flywheel-motor-generator (FMG) unit is therefore sometimes referred to as a mechanical battery, due to its conversion of electrical energy to kinetic energy (rather the chemical energy as in an electro-chemical battery). With suitable control 38

39 electronics the FMG can therefore be considered as a direct cometitor to battery or suercaacitor units for hybrid electric vehicle alications. The mechanical flywheel otion relies on a direct mechanical link between the high seed flywheel and a continuously variable transmission (CVT) caable of transmitting ower over a continuous range of gear ratios. While the FMG requires an external electrical traction system, the mechanical flywheel can connect directly to the vehicle drive shaft via the CVT, and has the advantage of reduced comlexity of the flywheel rotor. The mechanical flywheel is therefore suitable as a bolt-on FESS to enable regenerative braking on existing vehicle architectures. A general discussion of all asects of the design rocess for advanced flywheels is given by Genta [35], while detailed descritions of the design and testing of mechanical and electrical flywheels for vehicle alications are given by Shah [29] and various ublications by researchers at the Center for Electromechanics, University of Texas at Austin [32, 34, 36] resectively. Hydro-neumatic accumulator Hydraulic energy storage systems can be imlemented in hybrid vehicles through the use of a high ressure accumulator and a low ressure reservoir [37-38]. The accumulator is a ressure vessel artly containing a quantity of gas which is sealed by a membrane. Hydraulic fluid can then be umed from the low ressure reservoir into the accumulator, comressing the gas and thereby storing energy. By allowing the ressurised hydraulic fluid to flow out of the accumulator and through a hydraulic motor the stored energy can be recovered as mechanical work. The energy storage efficiency of accumulator devices can be imroved thorough the use of an elastomeric foam in the gas section of the accumulator to minimise thermal losses [39]. The maximum round-tri efficiency of this tye of energy storage system can be as high as 80% [40-41]. A conventional variabledislacement hydraulic um/motor suffers from relatively low efficiency at art load, although develoments in digital dislacement hydraulic um technology can achieve high efficiency over a wider ower range [42-43]. The energy storage caacity of the system is however fundamentally limited by the maximum and minimum allowable volume and ressure of the gas contained in the accumulator. The characteristics of these devices are described in the following sections and considered in terms of the suitability for alication in rail vehicles. 39

40 Secific energy (J/kg) Energy and ower caabilities of storage devices While many factors are imortant when selecting an energy storage device for a given alication, an initial assessment can be made by considering the energy and ower caacity er unit mass of a device. This can be achieved by erforming agone tests, where the energy that can be extracted from a fully charged device at constant ower demand is measured. This data can then be used to construct a agone lot [44-45]. Figure 2-6 shows energy storage device data from a range of sources for agone lots [46-47] and manufacturer s data [28, 48-50]. A key feature of agone lots is the fact that the ratio of secific energy to secific ower defines a characteristic discharge time for a given device, which can be seen as dashed diagonal lines in Figure 2-6. By identifying the tyical duration of charge and discharge events for an energy storage device in a given alication, the tye of device caable of meeting these requirements with minimum mass can be identified Batteries 1000s 1s ange of tyical rail vehicle braking times High-seed flywheels Hydraulic accumulators 1ms 10 2 Suer-caacitors Electrolytic caacitors Secific ower (W/kg) Figure 2-6 Comarison of secific energy and ower for energy storage devices Electro-chemical batteries, caacitors (electrolytic and develoments in suer-caacitor technology), advanced flywheels and hydro-neumatic accumulators are shown in Figure 2-6. The characteristic discharge times are also shown for the range of tyical rail vehicle braking times (with uer and lower limits reresenting intercity and metro tye services resectively). Currently batteries and suer-caacitor devices can rovide the energy and 40

41 ower caacity required to store the braking energy of the vehicle. Advanced flywheels however can meet these requirements with the highest secific energy and ower, and therefore with the lowest mass. The hydro-neumatic accumulator is seen to have a very low secific mass ( times lower than advanced flywheels) which limits its alication in regenerative braking systems. This tye of energy storage device has therefore not been considered in any further detail Additional considerations There are a range of additional factors relating to the imlementation of an energy storage device in a hybrid ower-train including; i. The safety requirements of the device, ii. The efficiency with which energy can be stored, iii. The cost, lifesan and maintenance requirements, iv. The system integration and control requirements. These issues are discussed below for battery, suercaacitor and advanced flywheel systems. Electro-chemical batteries An imortant feature of all batteries is the effect that state-of-charge (SOC) has on the charge/discharge efficiency, charge accetance and the lifesan of the device. A limitation of batteries is that increasing DOD tends to reduce both the efficiency and life [51], although the magnitude of this effect deends on the battery chemistry. The effect of SOC on charge and discharge efficiency is illustrated in Figure 2-7 for a lead-acid battery. Figure 2-7 Tyical charge and discharge efficiency of a lead-acid battery as a function of SOC [51] 41

42 The ower that can be delivered or acceted by the battery is also strongly deendent on the SOC as illustrated for a Li-ion battery in Figure 2-8. Charge Discharge Figure 2-8 Effect of SOC on secific ower of a Li-ion cell at 25 c [52] A suitable comromise is therefore required to identify aroriate oerating conditions for a articular battery tye in a given alication. An examle of this is the NiMH battery ack used in the Toyota Prius hybrid electric vehicle, where the control strategy maintains a SOC of between 50-70% at all times [51]. The actual secific energy of the battery ack is therefore only a fifth of the maximum energy caacity er unit mass. Another imortant factor affecting the erformance of all batteries is the oerating temerature. This is a articular roblem with NiMH batteries where low temeratures can significantly reduce the ower caacity of the battery, as illustrated in Figure 2-9. NiMH batteries also suffer from low charging efficiency at high temerature (>50 c) [52]. Careful monitoring and control of cell temeratures is therefore a necessary feature of battery acks for vehicle alications. Prismatic module Cylindrical module Figure 2-9 Effect of temerature on secific ower of a NiMH battery at 60% SOC [52] 42

43 Another issue with battery acks is the need to revent significant variation in SOC between cells, which occurs through differences in the self discharge rates of the cells due to manufacturing or temerature variations. A battery ack for vehicle alications therefore needs careful management in order to maintain a consistent SOC in all cells, and revent overcharging or fully discharging cells which cause ermanent damage. In the Li-ion vehicle battery ack described by Saft et al [49] each module allows the charge and discharge of the individual cells to be controlled and monitored by the vehicle control systems. Each module is also designed to rovide over-charge detection and short-circuit rotection. The modules are liquid cooled to enable thermal management during both warm and cold weather oeration. The integration of batteries in a vehicle system is illustrated in Figure These requirements add to the overall weight and cost of battery systems, but are necessary to ensure safe and reliable oeration. BATTEY SYSTEM OVEVIEW HIGH VOLTAGE UNIT VEHICLE MANAGMENT UNIT Figure 2-10 Imlementation of Li-ion battery modules in a vehicle alication [49] A comarison of the key erformance arameters of the different battery chemistries is resented in Table 2-3. This comarison of different battery tyes rovides reresentative data which has been collated from a several sources [24, 51, 53-54]. Lead-acid NiMH Li-ion Secific energy Wh/kg Secific ower W/kg Energy density Wh/litre Normal life years Cycle life (at 80% DOD) cycles Actual cost US$/kWh Table 2-3 Comarison of cell characteristics for the main battery chemistries 43

44 The cost of different battery systems deends on a number of factors including material and manufacturing costs and roduction volumes [55], and are therefore likely to vary with demand and technological develoments [56]. The values of cost/kwh given in Table 2-3 reflect the current cost of the different battery systems, although lower costs (articularly for Li-ion technology) are being targeted. The self-discharge rate is highest for Ni-MH batteries, but in all cases the losses will be insignificant for hybrid vehicles alications as the battery is unlikely to be left fully charged for long eriods, and will be recharged quickly during use. Suercaacitor As no chemical reactions take lace within a suercaacitor device there is little material degradation over time, and a long lifesan (>100,000 cycles) without significant loss of erformance is ossible [57]. The lack of an electrical-chemical energy conversion also means that suercaacitors can achieve high discharge efficiencies of >95% when oerating within the rated ower [58], although this decreases during higher ower oeration. However, as with battery systems there is significant variation in cell roerties which can lead to uneven SOC (and therefore voltage) within a ack of suercaacitor cells. Charge equalisation circuits are therefore used to balance the cells and revent damage [59]. These control circuits are required to manage the erformance and lifetime of the suercaacitor devices, but add to both the cost and size of suercaacitor systems. The cost of suercaacitors in US$/kWh is currently much higher than batteries [57], and they are more suited to higher ower alications where the cost in US$/kW is critical. Costs are however exected to imroved, with a target of < 50/kWh exected to be achieved in the next 5-10 years [47]. Advanced flywheel The energy storage efficiency of a FESS deends largely on the characteristics of the transmission used. A comarison of the two transmission otions in terms of the flywheel-towheel efficiency of the devices is illustrated in Figure 2-11 using reresentative motorgenerator and gearing efficiencies. 44

45 FMG Motor/generator Power electronics, motor and final drive FW-to-wheel efficiency Electrical FESS Flywheel ~95% ~85% ~81% CVT Final drive Mechanical FESS Flywheel ~85% ~95% ~81% Figure 2-11 Tyical ower-train efficiencies for electrical and mechanical FESSs during discharge While the efficiency values stated in Figure 2-11 are aroximate, and will deend on factors such as vehicle seed and alied tractive effort as well as the details of the ower-train comonents and configuration, it is clear that the efficiencies of the two systems are broadly comarable. These efficiency values do not include the effect of flywheel self discharge due to aerodynamic and bearing losses. For a well designed flywheel oerating in a low ressure environment these losses are tyically in the region of 2-4% er minute [33] which is much higher than self-discharge rates for batteries or suercaacitors. While this makes FESSs unattractive as long-term storage devices, these losses are unlikely to have a large imact in hybrid vehicle alications where the flywheel is only fully charged for short eriods for examle, station dwell times for trains are tyically around 1 minute. These losses should however be taken into account when considering the system design and control strategy for hybrid flywheel vehicles. Other factors such as the ease of integration, mass, cost and auxiliary requirements of the flywheel system deend on the transmission system and are discussed below. a) Ease of integration For FMGs the integration of an electrical machine with the flywheel adds further design constraints to the flywheel geometry and increases the cooling requirements due to electrical motor-generator losses occurring within the vacuum containment. The seed range of the motor/generator required to maintain the required ower flow also limits the seed range of the flywheel, leading a tyical DOD of around 60% [60]. The FMG aroach does however allow flexibility in the ositioning of the energy storage device within the vehicle, and allows gimbal mounting in order to reduce both the flywheel bearing loads and the gyroscoic forces 45

46 transmitted to the vehicle [61-62]. The mechanical flywheel system requires a recision high-seed vacuum seal to allow a direct shaft connection between the flywheel and CVT [29], and is more restricted in the ositioning and orientation of the flywheel unit. If necessary the gyroscoic forces acting on the vehicle can be eliminated by using two counterrotating flywheels with arallel axes, although bearing loads must still be considered. The DOD of the flywheel is constrained by the characteristics of the mechanical transmission. The choice of flywheel DOD for a articular mechanical FESS and alication is not clearly defined or justified in the literature. b) Mass The electrical FESS develoed for the ULEV-TAP2 roject [61, 63] (see Section for more details) has an installed mass of around 1 tonne, only 35% of which is due to the FMG including the containment and gimbal mount. The remaining mass is largely due to the ower electronics required to integrate the FESS with the diesel-electrical ower-train. An argument in favour of mechanical flywheel transmissions over electric systems on the ground of reduced mass and cost is resented by Cross and Hilton [64]. Both the secific energy and secific ower of the Flybrid mechanical FESS (see Section for more details) are higher than those of the ULEV-TAP2 electrical flywheel unit, desite the fact that the electrical system requires searate electric traction motors in order to achieve regenerative braking. This illustrates the high secific torque that can be achieved with modern mechanical continuously variable transmissions. c) Cost The cost of FESSs for vehicle alications is difficult to secify as no system is currently in roduction. Whether the high cost of rototye systems can be reduced deends on the size of the market, and whether comonents (articularly the flywheel rotor) can be standardised for a range of alications. This may rove difficult in alications where a uniquely otimised flywheel caacity and geometry is necessary or beneficial. The flywheel device develoed by Flybrid Systems for F1 alications has been designed to minimise the carbonfibre content of the flywheel rotor, as this has been identified as a major comonent of the cost [64]. For electrical FESSs, a recent SSB study into energy storage technologies for rail [47] suggests that the cost er unit energy caacity of FMGs could otentially be considerably lower than suercaacitors due to the relatively low material costs. The overall cost of electrical systems should however also be considered. In a reort to the California 46

47 Energy Commission [65] one of the major manufacturers of stationary electrical flywheel systems for the ower generation sector estimated that 60% of the total system cost is associated with the ower electronics, while the remaining 40% is the FMG unit. These systems oerate with a similar range of charge/discharge times (10-100s) to those required in regenerative braking systems for rail vehicles, and so the ratio of flywheel energy caacity to ower requirement of the electronics will be similar. It is therefore reasonable to exect that the cost of a flywheel device for use with a mechanical transmission will be significantly lower than the electrical otion, due to the low materials cost of the transmission and the simler flywheel rotor design (which does not require an integrated motor and electrical control system). d) Safety and reliability The safety and reliability of the system is essential. This has been achieved through the use of careful design and control of the flywheel rotor in order to ensure that it is not exosed to dangerous stresses. Furthermore, containment systems have been designed and tested in order to ensure that the kinetic energy of the flywheel can be safely dissiated in the case of a catastrohic rotor or bearing failure [36, 53, 64]. As the oerating range of the flywheel is constrained to ensure safety, there is little material degradation during use. This means that there is no decrease in erformance over time, and the lifesan of these devices is exected to be high. egular maintenance of the bearings and seals is however likely to be required [47]. In conclusion, while there are still challenges relating to the mass roduction and long-term reliability of advanced flywheels for vehicle alication, they have been shown to reresent a technically viable, efficient and otentially cost-effective technology for hybrid vehicles Develoments in hybrid rail vehicles The requirements of hybrid systems for rail are summarised by Lu et al. [66] in terms of the ower-train configuration, energy storage device, vehicle control strategy, system otimisation and ractical imlementation. It is clear that there is a great deal of flexibility in the way that hybridisation can be imlemented. This is illustrated by the range of hybrid systems that have been roosed and develoed for rail alications, although none are currently in wide-sread use. Overviews of the different systems are resented below. 47

48 Hitachi hybrid trains The Jaanese manufacturer Hitachi has develoed a single-car diesel-electric hybrid commuter train called the New Energy Train [67]. This vehicle has a maximum seed of 100 km/h and the ower-train consists of a 331 kw diesel engine with an electrical transmission and a 10 kwh Li-ion battery ack, arranged as shown in Figure The ower-train control strategy limits the DOD of the battery ack to around 30% in order to achieve a redicted life of over 5 years. During oeration, regenerated energy has been shown to rovide 14-33% of the total energy demand on the routes shown in Figure Figure 2-12 Overview of the NE Train layout and oerational erformance [67] It is interesting to note that there is no significant downsizing of the diesel engine for comarison the 160 km/h UK Class 170 DMU (diesel multile unit) has a similar installed engine caacity of 315 kw er car. This is robably due to the need to meet the maximum tractive ower requirement on a range of routes with different characteristics in terms of gradient and stoing attern. It does however suggest that the rimary benefit of hybridisation for rail vehicle is the recovery and reuse of braking energy rather than imroving the engine oerating efficiency. Hitachi have also been involved with the develoment of the hybrid diesel-electric ower-train used in the unique Hyabusa track measurement train oerated in the UK by Network ail [68]. This is a diesel-electric locomotive with Li-ion battery storage rated at 48 kwh. The battery ack weighs around 1 tonne and has an exected life of 8-10 years, suggesting that a relatively low DOD is used during oeration. Around 80% of the regenerated braking energy was redicted to be recovered for the next owering cycle, which suggests that the energy storage system has an overall round-tri efficiency of around 70%. The vehicle was oerated on all main-line routes around the UK as art of the routine track insection rogramme between mid 2007 and late eorted fuel savings of 12-48

49 20% were achieved deending on route [69], although no information is available regarding the effect of the drive cycle or the baseline fuel consumtion figures used. TI Li-ion esearch The Jaanese ailways Technical esearch Institute (TI) have conducted research investigating the use of Li-ion batteries in an electric tram [70]. While this is an electric rather than hybrid vehicle, the research rovides useful information about the requirements of relatively large scale Li-ion energy storage systems. The battery system tested includes a battery rotection management system and battery management units for each module of Li-ion cells, and is shown in Figure 2-13(a). The secific energy and ower of the system is therefore significantly lower than for a single Li-ion cell. As the battery system was used as the rimary energy source for the electric tram it was discharged from 100% SOC to around 60% SOC through a series of small discharge/charge cycles with SOC variations of around 2.5%/1.5% resectively, corresonding to vehicle owering and braking. In this alication a life of over 100,000 cycles was exected. For a hybrid diesel alication a change in SOC of around 30% of caacity would be exected during charge and discharge events (as seen with the Hitachi NE Train), reducing the lifesan. The secific energy and ower of this system with a 30% DOD can be estimated from the data given. (a) (b) Figure 2-13 (a) Lithium-ion battery system develoed by TI [70] and (b) Mitrac suercaacitor based energy storage unit [71] Bombardier Mitrac energy storage system Bombardier Transortation have develoed a suercaacitor based energy storage unit (shown in Figure 2-13(b)) which can be integrated with their Mitrac range of electric ower-train comonents [72]. Measurements of the energy storage device in oeration on a 49

50 DC electric light train have been ublished [73], showing a reduction in both the eak ower demand at the line and net energy consumtion. The system is also roosed as a solution to regenerative braking in diesel-electric vehicles. When comared to other energy storage systems, the mass and volume of the suercaacitor unit are however very large for its energy caacity. Green Goat hybrid shunting locomotive The Green Goat hybrid shunting locomotive has been develoed for use in railway switching yards [74-75]. This tye of alication with low mean ower and short eriods of high ower oeration is articularly suited to a hybrid ower-train. A small diesel engine is used at otimum efficiency to charge a large bank of VLA batteries, as shown in Figure Figure 2-14 Illustration of conventional and hybrid shunting locomotives [75] These batteries then rovide auxiliary and traction ower for the vehicle. This achieves a reduction in fuel consumtion and emissions comared to a conventional shunting locomotive which requires a much higher installed engine caacity. The requirements of the ower-train for this tye of alication are articularly aroriate for fuel cells oerating at constant ower, and analysis of a fuel cell-battery hybrid has also been erformed [76]. Both these systems work well in situations with intermittent oeration at low seed, but are of limited otential for main-line diesel vehicles. Aiesearch ACT-1 electric flywheel unit for rail In 1972 the Urban Mass Transortation Administration (UMTA) initiated the ACT-1 rogram to develo an efficient DC electric light rail vehicle with an onboard electric FESS 50

51 [77]. The flywheel energy storage device was develoed by the comany Garrett Aiesearch [78], and is illustrated in Figure 2-15(a). The unit consists of a laminated steel rotor with a design seed of 11,000 rm roviding maximum kinetic energy storage of 4.5 kwh. The rotor is connected to a 600 kw DC motor-generator via a fixed ratio gearbox. The rotor oerates within a vacuum environment to reduce aerodynamic drag, resulting in losses of around 3% er minute at design seed. The FESS unit weighs around 3 tonnes resulting in a relatively low secific energy of 1.45 Wh/kg. In oeration, the unit was mounted with the axis of flywheel rotation arallel to the longitudinal axis of the vehicle. Under exected oerating conditions, the gyroscoic forces acting at the mounting oints of the FESS were found to be less than the weight of the unit, roducing no noticeable effect on the vehicle dynamics or the oerating life of the system. (a) Gearbox (b) Blower Motor/generator Flywheel Containment Figure 2-15 (a) Illustration of ACT-1 flywheel-motor-generator unit [78] and (b) ULEV- TAP2 comlete electrical flywheel system [61] ULEV-TAP2 light rail ower-train During a hybrid diesel-electric ultra-light rail vehicle with electric FESS was develoed through the Euroean Commission sonsored ULEV-TAP 2 roject [61, 63]. The aim of the roject was to develo an efficient alternative to electrified systems, which could also meet strict emissions and noise limits. The electric FESS unit develoed for this alication is shown in Figure 2-15(b). The FMG device uses a comosite rotor with a maximum seed of 22,000 rm to achieve a useful energy caacity of 4 kwh. This results in a secific energy of 10.6 Wh/kg for the FMG. However, it is clear from Figure 2-15(b) that many other comonents (redominantly ower-electronics) are required in order to integrate the FMG with the vehicle ower-train. The FMG accounts for only 38% 51

52 of the total comonent weight, and the secific energy of the FESS unit is therefore considerably lower than that of the FMG, at around 4 Wh/kg. A ower-train using the FESS with an efficient diesel-generator unit was redicted to achieve a 40% fuel saving and 50% reduction in emissions comared to a conventional diesel-electric light rail vehicle, although no details of the drive cycle are given. Parry Peole Mover light rail vehicle A simle flywheel energy storage concet has been develoed for light rail alications by Parry Peole Movers [79]. The vehicle contains a conventional low seed flywheel constructed from laminated steel and weighing around 0.5 tonnes (giving a relatively low secific energy of 2 Wh/kg for the rotor alone). This flywheel can be charged either from an external electric ower source or through regenerative braking. A small LPG engine acts as a range extender, roviding additional tractive ower to the vehicle through an automatic gearbox and shared mechanical driveline. Figure 2-16 PPM vehicle and ower-train configuration [79] The vehicle is currently oerating a single niche alication, roviding a regular service on a 1.3 km stretch of isolated branch line at Stourbridge, UK. The size, weight and relatively low efficiency of the flywheel system however make it unattractive for use on larger and higher seed DMU tye vehicles. Voith hybrid systems for rail The comany Voith Turbo Ltd have develoed a range of energy saving features for their rail ower-train systems [80]. The conventional ower-train consists of a diesel engine with a hydrodynamic transmission. Voith have included Microbrid and Hydrobrid otions which both utilise a degree of onboard energy storage with a hydrostatic um/motor and a hydro-neumatic accumulator. The Microbrid system is a micro hybrid with the ability to 52

53 rovide auxiliary loads, while the Hydrobrid is a mild hybrid caable of roviding a degree of traction boost. Few details are available regarding the caability of the systems, and neither system aears to be in current service. The available data for energy storage systems that have been develoed for rail alications can be used to assess the effect of system requirements on the overall energy and ower caacity of the Li-ion, suercaacitor and electric flywheel storage devices. Figure 2-17 shows the secific energy and energy density lotted against secific ower for both the device and the comlete energy storage unit oerating with ractical DOD. Li-ion FMG Li-ion FMG Suerca Suerca Figure 2-17 Secific energy, secific ower and energy density for Li-ion suercaacitor and advanced flywheel devices and integrated units (arrows show the effect of control and ackaging requirements) For the energy storage devices oerating with ractical DOD limits, the ULEV FMG is seen to combine high secific energy and secific ower. When the mass of the FMG control systems are included, the secific energy is still much higher than the suercaacitor unit but dros to about half that of the Li-ion battery unit. However, the higher secific ower of the ULEV system comared to the Li-ion unit means that it could still be the lighter otion for regenerative braking alications with a characteristic charging time of u to 50 seconds. This comarison of systems shows that advanced flywheels are a viable and attractive energy storage device for rail vehicles. The control requirements for an electrical flywheel device have been seen to reresent a significant roortion of the weight and cost of the system. While mechanical flywheel systems have been identified as otentially low-cost, efficient and easily integrated method of enabling regenerative braking on vehicles with any 53

54 ower-train, no such systems have been develoed for rail alications. The following sections therefore consider relevant transmission technology and mechanical flywheel systems that have been develoed for other alications Continuously variable transmissions for flywheel energy storage systems While the design challenges of the advanced flywheel device have to a large extent been overcome, there are considerable difficulties associated with the design of a transmission to allow efficient transfer of energy between the vehicle and flywheel over a wide range of oerating conditions. A comlete transmission system caable of achieving this continuous variation in seed ratio is referred to in this Thesis as a continuously variable transmission (CVT). Different otions for achieving a mechanical CVT are discussed in the following sections Fixed ratio gearbox with sliing clutch One of the simlest ways of imlementing a CVT is through the use of a fixed ratio gearbox with a clutch designed to oerate with continual sli. This allows the CVT to achieve any seed ratio within the range of the gearbox. As the torque exerted on each side of a sliing clutch is equal and oosite it can be used to transfer energy from a higher seed inut shaft to a lower seed (or stationary) outut shaft. This results in a roortion of the inut energy being dissiated due to friction in the sliing clutch. These losses can be reduced by using multile gear ratios, reducing the average seed difference between the inut and outut shafts during oeration. A CVT consisting of a 12-seed gearbox and a hydraulically actuated wet multi-late clutch has been investigated for use in automotive flywheel energy storage alications by Beachley et al. [81]. An average transmission efficiency of over 90% was redicted for this system, and otential advantages in terms of reliability and ease of develoment (due to the use of conventional automotive ower-train comonents) were identified. The raid gear changes required during acceleration (several times a second during high ower oeration) and during the switch from acceleration to regenerative braking were however felt to require a sohisticated control system in order to achieve good driveability and smooth ower transfer. 54

55 Variator technology A variator is defined here as a device caable of transmitting ower while achieving a continuous variation in seed ratio (outut seed divided by inut seed). The term variator should not be confused with the term CVT which is a comlete transmission system which may or may not contain a variator device. There are two main tyes of mechanical variator that have been develoed for ractical vehicle imlementation. The first of these is the toroidal variator, an examle of which is illustrated in Figure This tye of variator transmits ower via a traction fluid under high ressure at the contact oints between roller elements and toroidal cavities located on an inut and outut disc. Varying the angle of the rollers about a fixed axis (erendicular to the rotational axis of both the roller and the inut/outut discs) changes the radius of the contact oints on the two toroidal surfaces. This allows a continuous variation in the ratio of outut to inut rotational seed between a maximum and minimum ratio defined by the geometry of the roller and toroidal surface. Outer discs (tyically inut) oller Toroidal cavity Inner discs (tyically outut) Figure 2-18 Illustration of a full toroidal variator [82] The torque transmitted by the toroidal variator is deendent on the contact ressure between the rollers and the inner/outer discs. By controlling this ressure, the rollers automatically assume the osition required to achieve the necessary seed ratio between the inut and outut discs the variator is therefore described as torque controlled, and can be imlemented with a simle control system without feedback [83]. 55

56 The second tye of ractical variator is the ush-belt design, and has been develoed by comanies including Bosch/Van Doorne for automotive alications with maximum torques of u to 400 Nm [84]. The comonents of the ush-belt variator are illustrated in Figure (a) (b) Low High Figure 2-19 (a) Illustration of ush-belt variator oeration showing low and high gear ratios [84] and (b) details of ush-belt construction This device uses two v-ulleys with variable sacing mounted on arallel shafts, connected with a v-belt consisting of a large number of metal segments mounted on steel bands (seen in Figure 2-19). Power is transmitted through comression of these elements, allowing much higher torques than can be achieved with a tensioned v-belt drive. The sacing of the two v- ulleys is controlled simultaneously using hydraulic actuators in order to vary the gear ratio between the secondary and rimary ulleys. The variator is therefore ratio-controlled, and requires a feedback control system in order to maintain a demanded outut torque. Both toroidal and ush-belt variators have similar erformance and oerating limitations. The range of ratio coverage that can be achieved is limited by the roller/cavity and ulley geometry resectively. For both tyes, a maximum ractical gear sread of around 6.25 is ossible, centred around a 1:1 gear ratio [64, 84]. Both devices are otentially bi-directional (i.e. ower can flow in either direction) but are not reversible, as the sign of the seed ratio is fixed by the mechanism. The maximum efficiency of variator devices is tyically lower than that of fixed ratio gearing. While the loss mechanisms for toroidal and ush-belt variators are different, the overall effect of losses is similar, with efficiency varying as a function of inut seed, inut torque and seed ratio. In toroidal variators the main losses are due to sliage at the roller 56

57 contact oints and friction in bearings. Exerimental tests show that maximum efficiencies of around 95% can be achieved [85], as shown in Figure Figure 2-20 Toroidal variator efficiency for an outut/inut seed ratio of 1 [85] It is clear from Figure 2-20 that toroidal variator efficiency is only weakly deendent on the actual inut seed, but decreases significantly at low inut torque. The loss mechanisms for ush-belt variators are discussed in some detail by Akehurst et al. [86-88], and are dominated by friction in the ush-belt, at the belt/ulley contact oints and in the bearings. This results in similar maximum efficiency levels as the toroidal variator, and the effect of inut seed on the efficiency is again seen to be small [89]. For the modelling of CVT ower-trains, the efficiency of mechanical variators is therefore often assumed to be a function of only seed ratio and inut torque. This allows a quasi-static aroach to be alied using exerimentally derived efficiency mas. The variable seed ratio of variators means that they can be used as a CVT. However, as the minimum seed ratio ossible is greater than zero, it is aarent that a sliing clutch is required to transfer ower to a low seed outut shaft (e.g. for the initial acceleration of a vehicle) until the minimum seed ratio is reached. The energy dissiated in the clutch during this eriod of oeration will reduce the overall efficiency of the CVT. Another issue relating to this tye of direct variator CVT is that the variator is required to transmit the full ower flowing through the CVT. For high ower alications (such as regenerative braking in rail vehicles) this could result in a heavy and costly system. Finally, the losses in the variator and the ower demand of the variator control system may result in a relatively low CVT efficiency, articularly when oerating at low torque. 57

58 Power-slit rinciles for a simle differential It is clear that there are a number of limitations regarding the use of a variator device as a CVT. Some of these issues can be addressed by considering mechanical ower-slit transmissions (PSTs). These are achieved by using a differential gearing unit with 2 degreesof-freedom to slit the main ower flow into two aths. The kinematic and torque relationshis that exist between the branches of the differential allow the overall ower flow to be controlled by controlling the ower flow in one of the two ower-slit branches. Fundamental analysis of ower-slit transmissions has been erformed for a range of configurations consisting of a variator and PGS [90-93]. The way in which the three branches of a PGS are connected in these transmissions affects the overall oeration, but the general analysis method used by White [94] allows an analysis of the system to be erformed for the generic differential unit shown in Figure Branch: where Figure 2-21 Descrition of generic differential gearing unit The kinematic and ideal torque relationshis for the generic differential illustrated in Figure 2-21 are shown in Equation 2-1 and 2-2 resectively. 3 2 Or alternatively (2-1) T T2 T1 (2-2) 1 These equations aly to any tye of differential in any configuration. It has been shown [95] that the relationshi between the ratio of ring to sun diameter of a lanetary gearset (PGS) tye differential and the value of can be defined for each of the six ossible configurations, as shown in Table 2-4 for a simle PGS and Table 2-5 for a PGS with idler lanet gears. 58

59 Table 2-4 elationshi between and simle PGS [95] Table 2-5 elationshi between and PGS with idler lanets [95] It can be seen that once an aroriate value of has been identified for a given alication, the most aroriate PGS configuration can be deduced for these tyes of differential by referring to Figure (b) Basic ratios for a simle lanetary gearset (a) Basic ratios for a lanetary gearset with idler lanets Figure 2-22 Possible range of values for ractical simle and idler PGSs [95] This allows a ower-slit system to be analysed for a given alication as follows; i. Analyse system using the equations for a generic differential, 59

60 ii. iii. Identify an aroriate characteristic differential gear ratio for the alication, Identify the most ractical differential tye and branch connections in order to achieve the required characteristic gear ratio This is a owerful analysis method for understanding PSTs as it eliminates the need for an exhaustive study of all ossible ermutations of differential tyes and connection otions. Simle variator-controlled PSTs can be classified as either inut or outut couled, deending on whether the variator connects between the control branch and the ower inut or ower outut branch of the PGS. Analysis of these configurations shows that several oerating regions are ossible, as illustrated in Figure 2-23 for the case of an outut couled PST. P d = P variator / P in Diff. Var. Oerating region Figure 2-23 Illustration of the ossible oerating regions for a variator-controlled outut couled ower-slit transmission with arrows showing direction and magnitude of ower flow (adated from [96]) It can be shown (see Chater 5) that in order to achieve genuine ower-slit oeration (with ower flow through each of the two branches being less than the overall ower flow) the ratio sread of the PST is always smaller than the ratio sread of the variator. This sread can only be increased by allowing ower-recirculation to occur. As well as extending the ratio sread, the resence of ower-recirculation also makes it ossible to achieve a geared neutral condition (i.e. zero transmission outut seed with a non-zero inut seed) and reverse drive (i.e. negative outut seed for ositive inut seed) these are often described as infinitely 60

61 variable transmissions (IVTs). This however requires the variator to be able to handle large ower flows, and limits the overall efficiency of the PST due to high ower losses in the variator and gearing. This reduction in efficiency has been demonstrated analytically and exerimentally for the two main tyes of simle (single regime) ower-slit oeration [97-99]. The comromise between ratio coverage and transmission efficiency has been investigated by Beachley et al. [93] for a flywheel hybrid assenger car alication. The general hybrid ower-train configuration used in the analysis is shown in Figure Figure 2-24 Proosed flywheel hybrid ower-train for assenger car [93] A range of simle variator-controlled PSTs were considered for the CVT, with varying degrees of ower-slit and ower-recirculation oeration. The analysis focuses on the efficiency of the CVT rather than quantifying fuel savings. In order to roduce general results that were indeendent of the control olicy for the system, simulations were erformed assuming a range of constant flywheel seeds. The results are therefore not useful in assessing the effect of factors such as energy storage caacity, flywheel self-discharge or the hybrid ower-train control strategy on the overall FESS erformance. They do however confirm that ower-recirculation results in average CVT efficiencies considerably lower than average variator efficiencies. A load factor (calculated assuming no transmission comonent losses and identical variator dimensions in all cases) was used to assess the relative variator size required for each case, which was shown to increase with the roortion of ower flowing through the variator Advanced ower slit configurations The comromise between overall ratio sread and the size and efficiency of the transmission which occurs with a simle variator-controlled PST can be addressed by considering more comlex multile-regime configurations. The simlest form of multile-regime PST consists 61

62 of one regime of ower-slit oeration and a second regime in which the variator rovides a direct connection between the transmission inut and outut. This tye of transmission has been investigated for both engine [100] and flywheel [93] alications. A fundamental analysis of a wide range of other 2-regime transmission configurations has been erformed by White [92, 95], who identifies and characterises 48 different configurations for a 2-regime CVT with synchronous gear-change using only a variator, two coaxial differential gears and two clutches to engage each regime. This illustrates the comlexity of identifying an aroriate configuration for a articular alication, as there are a large number of ways that ower-slit oeration can be achieved with multile PGSs. Secific configurations of multile-regime transmissions have been develoed for use as a conventional IC engine transmission [96, ] using mechanical variators or electrical motor/generator. Clutch and brake comonents are used to switch different differentials in and out of the ower-flow ath enabling several regimes of ower-slit oeration to be achieved. These transmissions achieve high efficiency, and are comact and light enough to relace conventional automotive transmissions, although none are currently in roduction. There are no examles in the literature of alying these multile-regime ower-slit transmissions to FESS alications. In conclusion, mechanical CVTs can achieve high efficiency in a relatively comact and lightweight ackage and are therefore attractive for FESS alications. The use of a mechanical CVT does however imose kinematic restrictions on the oeration of the FESS, which is related to the configuration and secification of the transmission comonents. It is therefore very imortant to be able to assess and comare a range of mechanical CVT systems, and identify the most aroriate system for a given alication. Several transmissions have been develoed secifically for FESSs in alications ranging from assenger buses to motorsorts, and are discussed in the following section Mechanical flywheel hybrid vehicles A range of hybrid vehicles utilising flywheel energy storage devices with mechanical transmissions have been roosed in the literature, and are described below. While some of these systems have been exerimentally roven, none have achieved commercial roduction. 62

63 Gyreacta transmission The Gyreacta transmission concet [103] oerates with two differential gear units in the form of lanetary gear sets (PGSs). These PGSs are used to achieve flywheel assisted acceleration or regenerative braking, with one PGS engaged in each mode (and the other allowed to freewheel). Each PGS has one of its three branches connected to the vehicle final drive and another to the flywheel (via a 4-seed gearbox). A clutch allows the engine to be connected to either of the remaining branches of the two PGSs. Inut Engine connected to PGS sun gear Outut Vehicle connected to PGS carrier branch Flywheel connected to PGS ring gear Figure 2-25 Examle of the oeration of the Gyreacta in vehicle acceleration mode (simlified diagram does not include PGS-flywheel gearbox) [103] These connections are made in such a way that during acceleration ower flows from the engine and flywheel to the vehicle (as shown in Figure 2-25), while during regenerative braking ower flows from the engine and vehicle to the flywheel (which is therefore referred to as ower-assisted braking ). In either mode, the engine torque alied to one branch of the PGS establishes aroriate reaction torques at the vehicle and flywheel branches of the selected differential. The use of an additional 4-seed gearbox between the PGSs and the flywheel allows the engine to oerate at aroriate seeds over the range of flywheel and vehicle oerating seeds. A further mode of oeration allows the flywheel to be charged from the engine while the vehicle is stationary. The main drawback of this tye of system is that the engine is required to oerate during regenerative braking, and sulies a significant roortion of the energy stored in the 63

64 flywheel. This has two results; firstly, the energy caacity of the flywheel needs to be larger than the braking energy that is actually recovered (increasing the system mass and flywheel losses) and secondly, that significant transmission losses occur as energy is transferred between the engine, flywheel and vehicle. University of Wisconsin research A 1970s research rogram at the University of Wisconsin develoed a flywheel hybrid system consisting of a 1-regime hydrostatic PST connected in series with a conventional 4- seed gearbox [81, 104]. This transmission was used to connect either an engine or a laminated steel flywheel to the final drive of the vehicle, as illustrated in Figure Figure 2-26 Illustration of roosed flywheel hybrid ower-train [104] The flywheel had a useful energy caacity of 0.5 kwh and a self discharge rate of around 2% er min at the maximum seed of 10,000 rm. When comared to a conventional vehicle, simulation results for the flywheel hybrid vehicle redicted an increase in fuel economy of 58% for an urban drive cycle. The majority of this fuel saving was achieved through more efficient oeration of the engine. However, the transmission was found to oerate with relatively low efficiency, with the total amount of regenerated energy being similar to the total losses at the flywheel and hydrostatic PST. General Motors flywheel research vehicle In the 1980 s General Motors investigated the otential of FESSs and develoed a hybrid assenger car with a direct-variator transmission [105]. This simle transmission was chosen over more flexible and efficient configurations due to its lower mechanical 64

65 comlexity, size and cost. Figure 2-27 illustrates how the transmission connects either the flywheel or the engine to the final drive of the vehicle. CVT Engine Variator Flywheel Figure 2-27 Illustration of flywheel hybrid ower-train roosed by General Motors Comutational modelling was used to assess the system erformance of the flywheel hybrid ower-train for both urban and highway drive-cycles. Comared to a conventional vehicle, the fuel consumtion was found to be 36% lower for the urban cycle, but 11% higher for the highway cycle where transmission losses were high and little regenerative braking is ossible. During the urban cycle 69% of braking energy was redicted to be available to charge the flywheel, 50% of which could then be used to deliver traction or auxiliary ower. This gives a low overall regeneration efficiency of around 35%, and GM concluded that the imrovements in urban driving efficiency were insufficient to justify further develoment of the concet. Imerial College London Mechanical Hybrid Vehicle esearch at Imerial College London has investigated the otential of a mechanical hybrid vehicle using a FESS [ ]. Two flywheel transmission otions have been considered; a simle friction-brake controlled PGS arrangement and a conventional single-regime variator controlled PST. The oerating range of the PST was extended by considering additional eriods of sliing-clutch and brake controlled oeration. The erformance of the two transmission systems was investigated through comutational modelling and fuel savings were redicted for both assenger car and bus alications. The brake-controlled transmission was found to be very inefficient, with the calculated fuel savings largely due to reduced engine idling. The variator-controlled PST achieved much higher efficiency, resulting in fuel savings of 10-25% over a conventional vehicle. However, neither FESS was otimised in terms of transmission or system efficiency. This is seen in the low DOD of 25% that is achieved using the secified system arameters for the variator-controlled 65

66 PST, leading to high flywheel losses and an oversized energy storage caacity. Also, no information is rovided regarding the design and control requirements of the transmission comonents (articularly the variator), making it difficult to assess the racticality of the system. Flybrid Systems The comany Flybrid Systems have develoed a FESS for use in a arallel hybrid configuration in motorsort alications [109]. The flywheel consists of carbon fibre wound around a steel hub and can oerate at a maximum seed of 60,000 rm. For racing cars, oeration of the flywheel is rarely required to accelerate the vehicle from stationary and so a transmission with a geared neutral is not essential. Therefore, the transmission used is a direct toroidal variator with clutch. As shown in Figure 2-28 the flywheel is connected to one branch of the variator via a fixed gear ratio. The inut drive to the FESS connects to the vehicle final drive. Figure 2-28 Flybrid mechanical FESS develoed for Formula 1 alication [109] This transmission rovides sufficient ratio coverage to meet the alication requirements, and oerates with high efficiency due to the consistently high ower demands. The differing requirements of a FESS for assenger car alications have however been acknowledged, and alternative transmissions have been roosed using a direct variator with a range-extending gearbox or single-regime ower-slit transmissions with a range of different connection oints for the FESS within the vehicle ower-train [110]. esults from this study suggest imrovments in fuel efficiency of between 15-25% are ossible over a conventional car for a range of standard urban and extra-urban drive-cycles, with over half 66

67 of the fuel savings shown to be a result of reduced engine idling when no engine ower is required. No details are however given regarding the selection of gear ratios, comonent sizing or the ower-train control strategy used in these simulations. Conclusions on mechanical transmissions for FESS alications It is very difficult to comare the different mechanical transmission systems in the literature, as they have all been develoed and assessed for a secific alication. When alied to FESSs, transmission efficiency has been identified as a critical factor in achieving effective regenerative braking, suggesting that PSTs should be designed to minimise or avoid ower-recirculation in order to reduce losses and achieve a lightweight and comact transmission. The design of the transmission will however directly affect the oeration range of the flywheel, and this interaction requires careful consideration in order to achieve good overall system erformance in terms of energy caacity, regeneration efficiency, mass and cost. The identification of aroriate system arameters for high-ower mechanical FESSs using a range of single and multile regime transmissions is the focus of Chaters 4 and 5 of this Thesis ail vehicle energy modelling The assessment of the otential benefits of hybrid regional trains requires detailed modelling of the vehicle ower-train and oerating conditions. The erformance of any rail vehicle is highly deendent on both the route considered and the way in which the vehicle is controlled. This is esecially imortant in the case of hybrid vehicles where the requirement of the regenerative braking system and the amount of energy available for recovery deends on the braking strategy used by the driver. The relationshi between control strategy and energy consumtion must therefore be characterised for both conventional and hybrid trains in order to make a comrehensive comarison of vehicle erformance. A range of comutational aroaches are described in the literature to investigate the oeration of electric and diesel owered rail vehicles. These can broadly be categorised as multile vehicle system models which tend to focus on the timetabling of services or the substation ower demand on electric railways [ ], and single vehicle models which focus on the control and energy consumtion of an individual vehicle. In both cases, an accurate model must include details of the vehicle dynamics, the ower-train, the route 67

68 rofile (including station sacing and gradients) and the control strategy alied by the driver. The tye of comutational method alied deends on the aims and scoe of the research. The two main areas of research in the literature are otimised control methods and hybrid rail ower-trains Otimised control A considerable amount of research effort has been focussed on identifying the most energy (or fuel) efficient driving rofile for a single rail vehicle travelling along a defined route with secified timetable restrictions. The nonlinear nature of this roblem means that otimisation methods are required. Both analytical methods (where a solution is calculated) and numerical methods (where a solution is reached iteratively) have been alied to this roblem. An analytical aroach using differential equations to describe the vehicle trajectory and identify an otimum control strategy is described by Golovitcher and Lui [113]. This can be alied to a route with multile stos and variable seed restrictions and gradient, and the roosed alication of this method for use in automated train oeration systems is described in some detail. The otimisation is however only erformed to minimise the traction energy required at the wheel, and therefore assumes constant ower-train efficiency under all oerating conditions. The ower delivered by the ower-train is also assumed to be continuously variable, rather than notch-based as is common in diesel owered trains. The otimised control of notch-controlled diesel trains has been investigated by Howlett and colleagues [ ]. A similar analytical aroach is used to identify otimum trajectories for diesel trains with discrete control of the engine throttle, assuming each throttle setting allows a constant flow rate of fuel to the engine. The ower-train efficiency is again assumed to be constant. Much of this research is embodied in the Metromiser on-board comuter system which can advise train drivers on fuel efficient driving strategies, mainly through the identification of aroriate coasting oints. This system has been shown to achieve energy savings of 10-15% for metro and suburban tye services [11]. Numerical otimisation methods have also been alied. Wong and Ho [116] assessed the feasibility and erformance of classical and heuristic numerical searching methods in locating coasting oints for a train travelling between two stations with a secified journey time. The analysis included the effects of seed restrictions, gradient and ower-train characteristics, and was imlemented using a time-ste based single train model. They found that a heuristic genetic algorithm (GA) search method achieved a robust and comutational efficient 68

69 solution for the otimum coasting oints. The research also showed that in cases with relatively short station sacing (less than 10 km) the otimal solution was achieved with a single coast oint rior to braking on the aroach to the station sto. Only with relatively long station sacing (~30 km) were multile coast oints required to achieve an otimum trajectory. Hwang [117] describes an otimisation method using a combined fuzzy logic and GA hybrid scheme for identifying the otimal comromise between energy consumtion and journey time in a high seed railway. An examle case was resented and the location of aroriate coasting zones were identified. A similar aroach has been alied to DC suburban railways [118] using a detailed time-ste based single-vehicle model. The use of fitness functions allowed weighting to be given to reducing either journey time or energy consumtion, and the effect of regenerative braking on the otimum trajectory was also considered. In both these studies the otimum driving strategy was found to be achieved largely through the use of coasting rior to braking on the aroach to seed reductions and station stos Hybrid rail vehicle analysis Several comutation studies have been erformed to assess the otential benefits of alying hybrid ower-trains to rail vehicles. Hillmansen and oberts [119] modelled a hybrid commuter train by assuming constant ower outut from a diesel engine and modelling the storage device as having constant efficiency and sufficient caacity to rovide the fluctuating ower demand over the course of a secified drive cycle. A constant regenerative braking efficiency was also used to define the fraction of available braking energy actually available at the storage device. Fuel consumtion was calculated over reresentative commuter and high-seed drive cycles using the hybrid model and a standard diesel model to rovide a basis for comarison. With a 50% regeneration efficiency and a storage device efficiency of 80% during both energy inut and outut, fuel savings of 35% and 28% were calculated for the high seed and commuter trains resectively. However, the fuel consumtion was calculated for an otimised combination of engine caacity and energy storage caacity, which is highly deendent of the articular drive cycles used. This is unlikely to be ossible in ractice, as rail vehicles are required to oerate a range of routes and services. 69

70 Jefferson and Marquez [120] described a Matlab/Simulink based generic ultra light rail vehicle (ULV) hybrid model. esults were obtained for a diesel-mechanical vehicle with automotive gearbox and constant transmission efficiency of 85%, and a diesel-electric hybrid vehicle. The diesel engine in the hybrid was assumed to oerate at constant ower outut and otimum efficiency, with energy storage rovided by a FMG unit. Using simulations with the same oerating conditions over a simle drive cycle, the hybrid vehicle was found to consume 36% less fuel than the conventional vehicle. The results are however calculated using a relatively simle ower-train model, and are limited to light rail vehicles oerating at low seed and with frequent stos. Destraz et al. [121] have modelled a suercaacitor based energy storage system for a hybrid diesel-electric rail vehicle. A constant ower outut from the diesel engine was again assumed, and the relationshi between required storage caacity and engine ower outut was investigated for a articular route. As engine size is decreased, both the fuel saving and the required storage caacity were shown to increase. An economic analysis was erformed and the most aroriate configuration was redicted to achieve a 44% reduction in fuel consumtion. This simulation work was however erformed using drive cycle data for trains oerating in a mountainous art of northern Italy, and the severe duty cycle is likely to have an influence on the erformance of the hybrid vehicle. The otimised downsizing of the engine is again deendent on the articular drive cycle studied, and may limit the vehicle erformance on other routes. A major drawback of these three studies is the fact that the drive cycles used are reresentative of an aggressive driving strategy, with no coasting used on the aroach to station stos and seed reduction. This effectively reresents a worst case driving strategy which maximises the fuel consumtion of a conventional vehicle and maximises the roortion of fuel saved by the hybrid vehicle. A more efficient driving strategy aimed at closely matching the timetable for a given route is likely to decrease the amount of braking energy available for regeneration, and so reduce the ercentage fuel saving associated with hybridisation. A more flexible aroach to train control has been alied by Wen et al. [122] for the comutational analysis of a hybrid-electric high seed intercity train. The inclusion of a driver module (discussed in more detail in Section 2.2.3) allows a range of driving strategies to be studied. A range of routes were considered, and both aggressive and efficient driving strategies alied. For the aggressive strategy the hybrid train was found to reduce fuel consumtion by 15-25%, falling to 8-19% when the efficient strategy was alied. 70

71 It is clear that there is significant otential to reduce fuel consumtion through the alication of hybrid technology. However, the interaction between efficient driving and hybridisation needs to be characterised for regional rail vehicles in order to assess the realistic fuel savings ossible and the requirements of the hybrid system. This can be achieved through detailed comutational modelling ail vehicle modelling Comutational modelling can be used assess energy consumtion in rail vehicles by considering the following three areas; vehicle dynamics, ower-trains and duty cycle analysis. a) Vehicle dynamics All rail vehicle energy models are based around the longitudinal motion of the vehicle, as illustrated in Figure olling resistance, F s Gravitational resistance, F g Gradient, α Velocity Vehicle weight, m v g Aerodynamic resistance, F a Tractive force, F t Figure 2-29 Illustration of longitudinal dynamic forces The longitudinal equation of motion can therefore be exressed as; m m x F F F v r (2-3) t r g where m v is vehicle mass; m r is equivalent mass of rotating comonents; x is the longitudinal vehicle dislacement from an absolute reference oint; F r is the total resistance to motion (rolling and aerodynamic); F g is force due to gravity and F t is tractive force roduced by the ower-train. The total resistance to motion is the sum of both rolling friction and aerodynamic drag forces. These resistance forces are functions of a large number of variables including vehicle 71

72 mass, length and rofile; vehicle and wind velocity; condition of axles and bearings; curvature of track. As some of these conditions vary with time, an established method of characterising the vehicle resistance is to take a range of measurements during tyical vehicle oeration and identify the relationshi between resistance force and vehicle seed for average conditions [123]. This average resistance curve can be aroximated as a quadratic function of vehicle seed (the derivative of dislacement) which is known as the Davis Equation, and is defined in Equation 2-2; F r 2 A Bx Cx (2-4) The terms A, B and C are constants known as the Davis coefficients and can be emirically related to the aerodynamic and friction forces as discussed by ochard and Schmid [124]. By assuming a lumed mass model for the train, the force due to gravity, F g, can be aroximated as a function of the vehicle mass, m v, and the instantaneous angle of incline, α(x), which is +ve when travelling uhill. This is defined in Equation 2-3; F g x m sin (2-5) v While this does not take into account the distributed mass of the vehicle along its length and therefore the continuous change in resistance force as the vehicle asses over a oint where the gradient changes, the effect on energy calculations is negligible as the total change in gravitational otential energy over a route is unaffected. b) Power-train modelling This thesis focuses on the energy consumtion of both conventional and roosed hybrid regional trains. The ower-train in these vehicles consists of a number of distributed diesel engines with hydrodynamic transmissions. A suitable modelling aroach is therefore required in order to assess the oeration of the combined engine-transmission unit. In order to calculate fuel consumtion, the dynamic oeration of both the diesel engine and the transmission comonents (torques converter and fluid coulings) can be accurately modelled using emirical steady-state erformance mas [46]. This is due to the fact that the rate of change of engine and transmission oerating oints are relatively low for rail vehicles, and so the ower-train erformance is close to the steady-state values. A standard time-ste based quasi-static analysis method has therefore been used to create a detailed ower-train model. This has been imlemented using Matlab/Simulink software due to its simle grahical interface and rovision of aroriate numerical solvers for time-ste based analysis. 72

73 A steady-state fuel consumtion ma has been obtained for a 330 kw diesel engine from the ADVISO automotive simulation software [125]. This has a similar ower rating to the diesel engines used in DMUs, and can be scaled to achieve the required maximum ower and maximum engine seed. This rovides the fuel consumtion (in g/kwh) as a function of engine torque and rotational seed. An imortant consideration in the modelling of a dieselhydrodynamic ower-train is the interaction between the engine and the hydrodynamic elements of the transmission. A standard aroach to the analysis of torque converters and fluid coulings is described by several authors [46, ]. Emirical data for the torque converter (TC) and fluid couling (FC) comonents can be used to create a steady-state oerating ma defining the engine oerating oint as a function of the transmission outut seed and the engine ower setting. c) Control/drive cycle The calculation of fuel consumtion for road vehicles relies on standard drive cycles (velocity-time rofiles) which are often used to determine the tractive ower required at the transmission outut [46, 125]. This requires a backwards-facing calculation method which uses the required tractive ower (calculated during each time-ste) to deduce the oerating oint of the IC engine, and hence the fuel consumtion. For a journey between two oints, the constraints of rail vehicle oeration mean that in theory the drive cycle can be redicted with more accuracy than is the case with road vehicles, where factors such as route selection, traffic intensity and signalling create considerable uncertainty. The slack in rail journey timetables does however allow variation in driving strategy, which has been shown to have a significant effect on the energy consumtion of conventional diesel and electric trains [see section 2.2.1]. Furthermore, this variation in driving strategy can lead to significant variation in both the energy storage requirements and otential benefits of hybrid trains [see section 2.2.2]. A forwards-facing calculation rocedure allows the vehicle ower to be secified, and the resulting vehicle acceleration to be deduced. This method has been alied in a number of single-train models, where the locations at which braking must be initiated are found by erforming a forward-facing calculation alied backwards in time to calculate the required velocity rofile on the aroach to a known seed limit when a secified braking force is alied [ ]. This aroach rovides a means of obtaining a simle drive cycle, but does not allow the effect of different control strategies (esecially those involving a degree of coasting) to be easily investigated. An alternative aroach is described by Wen et al. [122] for the modelling of conventional and hybrid high seed diesel-electric trains. This uses a 73

74 forward-facing calculation rocedure combined with a vehicle control subsystem. The tractive ower delivered to the vehicle during a time-ste is secified according to the current vehicle seed and osition with reference to overall route rofile (including consideration of gradients, seed limits and station locations), although the details of the control logic are not fully exlained. This enables different driving styles to be investigated by varying a range of control arameters. A similar aroach has therefore been used for analysing the erformance of conventional and hybrid diesel regional trains in this Thesis Summary Advanced flywheel energy storage devices have been identified as a candidate solution for regenerative braking systems due to their energy and ower characteristics, long life and otentially low cost. Mechanical flywheel transmissions offer a means of achieving high regenerative braking efficiency with a lightweight and comact system that can be alied to any tye of vehicle in a arallel-hybrid arrangement. The research resented in this Thesis therefore considers the alication of mechanical FESSs to regional diesel-hydrodynamic trains where braking losses have been identified as a major source of inefficiency. The effect of regenerative braking on fuel consumtion is however deendent on the driving strategy alied to the vehicle. Efficient driving strategies have been shown to significantly reduce braking losses, and have been identified as a short term measure to reduce the energy consumtion of conventional rail vehicles. The interaction between efficient driving and regenerative braking must therefore be studied in a rigorous way in order to identify aroriate solutions and assess the realistic benefits of hybridisation. 74

75 3. Effects of driving strategy on conventional and hybrid diesel regional trains Vehicle braking in non-electrified rail systems wastes energy. This Chater considers two aroaches to reducing braking losses in regional diesel trains; efficient driving strategies and regenerative braking. The interaction of these two aroaches is critical in secifying the requirements of a hybrid train and assessing the relative fuel saving. Comutational models of conventional and hybrid diesel-hydrodynamic regional trains have been develoed using real route data to generate a simle control algorithm and investigate the effect of driving strategy on fuel consumtion and journey time. The current modelling redicts fuel savings of u to 40% for the hybrid train when an aggressive control strategy is used. This fuel saving is halved when an efficient driving strategy is emloyed, which also reduces the required energy storage caacity. The model rovides a tool for identifying effective control strategies which should be imlemented to reduce fuel consumtion for both conventional and hybrid trains. It also rovides a realistic basis for assessing the otential of hybridisation for reducing fuel consumtion, and allows an initial assessment of the energy and ower requirements of the energy storage system Aroach to vehicle modelling The first stage in the develoment of a hybrid regional diesel train (HDT) model has been the construction of a conventional diesel-hydrodynamic (non hybrid) model. This aroach allows a direct comarison of the fuel consumtion for both the conventional and hybrid trains. Matlab/Simulink software was chosen because of its simle grahical environment and the availablity of numerical solvers for discrete time-ste based simulations. A forward facing model structure has been selected to allow a vehicle control module to be imlemented, enabling the effect of different control strategies to be investigated. The modelling uses a standard quasi-static aroach for the vehicle dynamics and ower-train comonents. A control module has been created which secifies the required vehicle ower or braking demand based on the current seed, osition and the secified route data (including seed limits and station locations). Figure 3-1 illustrates the model structure which consists of three modules describing the vehicle dynamics, ower-train and control. 75

76 Figure 3-1 Flow chart of calculation rocedure for forward-facing model structure consisting of control, ower-train and vehicle dynamics modules The following three sections describe the standard aroaches taken to modelling the vehicle and ower-train, and the secific control aroach alied in this study Vehicle dynamics module The vehicle energy consumtion is related to the longitudinal motion of the vehicle, as described by Equations (3-1). m m x F A Bx Cx 2 m sin x v r (3-1) t v This ordinary differential equation can be solved within Matlab/Simulink using numerical integration methods when the tractive force, F t, is secified. The vehicle arameters used in this analysis (including the emirical Davis equation coefficients) were secified using data for a tyical UK 3-car diesel-hydrodynamic train with maximum seed of 160km/h [ ] and are shown in Table 1. An equivalent rotational mass, m r, of 10% of the actual vehicle mass, m v, is used as this reresents a tyical value for mainline DMUs [123]. 76

77 Vehicle Parameter Mass, m v tonnes Equivalent rotational mass, m r 0.1 m v Installed ower 945 kw Auxiliary ower load 67.5 kw Final drive ratio 2.6 Wheel radius 0.4 m Davis equation coefficients: A [kn] B [kn / (m/s)] C [kn / (m/s) 2 ] Table 3-1 Vehicle arameters for model of 3-car diesel-hydrodynamic vehicle Power-train module The ower-train rovides the tractive ower and consists of the diesel engine and hydrodynamic transmission. This module defines the interaction between these comonents, and calculates the fuel consumtion and tractive ower delivered to the vehicle as a function of the vehicle seed and ower setting. The fuel consumed by the engine is calculated using data from a 330 kw diesel engine ma (obtained from the Advisor automotive vehicle simulation software [125]) which was scaled to match the installed ower of a single 315 kw engine as shown in Figure 3-2. This ma shows the oerating efficiency of the diesel engine in terms of mass of fuel consumed er unit of energy generated as a function of the engine torque and seed. Figure 3-2 Secific fuel consumtion and max torque curve for a 315kW engine 77

78 Using a standard aroach [ ] a model of a 2-seed hydrodynamic transmission consisting of one torque converter and one fluid couling has been develoed. The torque converter and fluid couling characteristics have been secified to match the ublished maximum tractive effort curve of the 3-car Class 170 Turbostar vehicle. Transmissions of this tye are widely used in regional diesel vehicles in the UK, but have not been considered in revious analysis of HDTs. The ower-train model allows the oerating oint of the engine and transmission to be defined for a given vehicle seed and ower notch setting. A total of 7 ower notch settings have been assumed for the transmission, with engine torques equal to the maximum torque shown in Figure 3-2 multilied by 0.4 to 1 (in stes of 0.1) reresenting notches 1 to 7. The tractive effort, transmission efficiency, engine SFC (secific fuel consumtion) and tank to wheel efficiency (assuming a secific energy of 12.1 kwh/kg for diesel fuel) are shown in Figure 3-3 as a function of vehicle seed for ower notches 1, 3, 5 and 7. (a) (b) (c) (d) Figure 3-3 Power-train outut values of (a) tractive effort, (b) SFC, (c) transmission efficiency and (d) tank-to-wheel efficiency as a function of vehicle seed for ower notches 1, 3, 5 and 7 78

79 The four diagrams in Figure 3-3 show the oerating characteristics of the ower-train. The fuel efficiency of the engine is relatively high under all oerating conditions with the SFC varying between 190 and 218 g/kwh. This suggests that there will be little benefit from control strategies aimed at otimising the engine oerating oint. Transmission efficiency is found to be relatively high for vehicle seeds above 50km/h, but dros raidly at lower seeds. These low seed losses are therefore significant, and can be reduced by either; Imroving the low seed efficiency of the conventional ower-train, Introducing a regenerative braking system to reduce or eliminate ower delivery via the conventional ower-train during low seed oeration. In terms of fuel consumtion, hybridisation of this tye of rail vehicle therefore has the dual benefits of recovering energy that would otherwise be dissiated during braking and using this energy to dislace low efficiency oeration of the conventional ower-train. The tractive force calculated by the ower-train module is assed to the vehicle dynamics module allowing the vehicle acceleration, seed and osition to be calculated. Introducing the control module then allows the required ower setting to be defined, as described in the following section Vehicle control module The vehicle control module has been constructed to use simle logic for owering, seed holding and seed reductions. Three scenarios are considered in order to define the ower demand. These are: i) Acceleration during vehicle acceleration below the holding seed (where v hold < v line ), the vehicle is commanded to oerate at a secified engine ower. ii) Seed holding as the vehicle reaches the holding seed the ower notch is changed to the minimum setting which will maintain vehicle acceleration. When the line-seed is reached the maximum ower setting to achieve deceleration is selected. If the gradient is sufficient to accelerate the vehicle beyond the line seed with no tractive ower alied, braking is used to limit the vehicle seed. iii) Seed reduction this alies during the aroach to reductions in line seed and station stos. Four trigger seeds are calculated during each time ste based on the distance and height difference between the current vehicle osition and; a) the next line seed reduction (Δx slow and Δh slow ) 79

80 b) the next sto location (Δx sto and Δh sto ). These trigger seeds reresent a hyothetical vehicle seed at which a secified constant deceleration, D, would achieve the required seed reduction. For cases (a) and (b) two trigger seeds are calculated using a lower constant deceleration value for a coasting trigger seed, and a higher value for a braking trigger seed. They are calculated by considering conservation of energy as defined in Equations 4-6. dissiated during constant deceleration KE GPE 0 Energy (3-2) m m Dx m m ( v v ) m gh 0 v r v r target trigger v (3-3) 2 v trigger 2 mv vtarget 2( Dx gh) (3-4) m m v r where v target is the next line seed limit (zero when calculating trigger seeds for station stos). Once the vehicle seed equals the coasting trigger seed, the vehicle is allowed to coast until the braking trigger seed is reached. At this oint a braking force is alied to the vehicle in order to achieve the secified constant deceleration. All results resented for constant force braking use a moderate braking deceleration rate, D brake, of 0.5m/s 2. By varying the coast trigger deceleration value, D coast, different degrees of coasting can be alied for a secific route, and the effect on journey time and fuel consumtion can be assessed. Figure 3-4 illustrates the oeration of the vehicle in each of the three scenarios described. Figure 3-4 Schematic illustration of vehicle control subsystem oeration 80

81 This shows that the oeration of the vehicle over a defined route can now be investigated using a range of simle vehicle control strategies. The effect of these strategies is assessed by comaring the calculated journey time and fuel consumtion in each case Descrition of route data The vehicle analysis is erformed using route data for real sections of main-line and branchline routes in the East Anglia region of the UK. The station locations, line seed limits and variation in height of both routes are shown in Figure 3-5. (a) (b) Figure 3-5 Data for (a) Main-line and (b) Branch-line routes used in the simulations An examle of the outut velocity rofile calculated by the conventional vehicle model using the main-line route data and three different driving strategies is shown in Figure 3-6. This illustrates the effect of varying the D coast arameter in the control module. Figure 3-6 Examles of calculated vehicle seed rofiles for main-line route with no coasting (flat out) and coasting trigger deceleration values (D coast ) of 0.1 and 0.07 m/s 2 81

82 Using this real route data in the vehicle simulation allows the calculated fuel consumtion to be comared to the time-averaged values reorted in a recent industry study [128] for the small fleet of Class 170 diesel hydro-dynamic trains oerating a range of services in this area (including these two routes). The results are shown in Figure 3-7 and are calculated for when the vehicle is driven flat out (i.e. no coasting used during seed reduction) with station dearture times matching the ublished time-table. Figure 3-7 Comarison of measured fuel consumtion (fc) with results of model using flatout vehicle oeration matching the time-tabled dearture times for main-line (ML) and branch-line (BL) routes The calculated fuel consumtion er car-km (where car refers to a single carriage) is similar for both the main-line and branch-line routes, and lies towards the uer range of the time averaged fuel consumtion measured in oerational service. The results are sufficiently accurate for the investigation of relative fuel savings for efficient driving strategies and regenerative braking; however, more data is required for a full validation of the model Driving strategies for conventional vehicles The model resented above can now be used to investigate the effect of driving strategies on conventional and hybrid regional trains. There are two simle driver control strategies considered for the conventional vehicle. These are: i) Limiting the maximum allowable seed of the vehicle (subject to the line seed limit) ii) Alying a degree of coasting rior to braking on the aroach to seed reductions and station stos 82

83 These measures lead to reductions in both the energy dissiated during braking and the work done against resistance at the exense of increased journey time. Figures 3-8 and 3-9 show the calculated fuel consumtion (in litres/car-km) against the calculated total journey time normalised by the scheduled journey time for the route. Constant station dwell times of 60s and 30s have been assumed for the main-line and branch-line routes resectively. Figure 3-8 Effect of increasing coasting and reducing maximum vehicle seed on relative journey time and fuel consumtion for main-line oeration with station dwell time of 60s Figure 3-9 Effect of increasing coasting and reducing maximum vehicle seed on relative journey time and fuel consumtion for branch-line oeration with station dwell time of 30s Figures 3-8 and 3-9 show the effect of alying coasting and seed limits; each data oint is calculated for a articular value of D coast and maximum vehicle seed. The findings fall into the following three categories. 83

84 1. There is considerable scoe to look at energy efficient driving techniques for these routes, as simulation with a flat out driving strategy takes around 85% of the scheduled journey time. 2. For a given journey time the maximum fuel saving is achieved by accelerating the vehicle to the full line seed limits and alying the required amount of coasting during deceleration. 3. There is a different effect on the two routes when alying the strategy of limited maximum vehicle seed. On the main-line route this results in a significant reduction in fuel consumtion regardless of how much coasting is used. On the branch-line route however significant reductions in fuel consumtion are only ossible through coasting. In summary, coasting is the most fuel efficient of the simle driving strategies considered. The vehicle control algorithm used here resents a simle method of identifying aroriate coasting oints for drivers Effect of timetabling constraints on fuel consumtion The analysis resented in Section 3.2 assumes overall journey time is the only time constraint on the train services considered. A constant value of D coast has therefore been used throughout the journey, defined here as a uniform coasting strategy. However, the imact of driving strategies is also affected by timetabling constraints. When using coasting, these restrictions make it necessary to aly aroriate values of D coast to meet the scheduled arrival times at each intermediate station, defined here as timetable-limited coasting strategy. The calculated fuel consumtion for three different driving techniques are shown in Table 3-2; flat out driving with scheduled intermediate dearture times (the worst case driving style as shown in Figure 3-7), uniform coasting to achieve overall journey time (i.e. uniformly alied D coast value) and timetable-limited coasting to achieve overall journey time (i.e. individual D coast values for each inter-station route section). 84

85 Fuel Driving Strategy Main-line Branch-line litres/car-km saving litres/car-km saving Flat out with scheduled intermediate dearture times (worst case) Uniform coasting to achieve overall journey time % % Timetable-limited coasting to achieve overall journey time % % ercentage saving relative to the worst case fuel consumtion Table 3-2 Effect of coasting strategy and timetable restrictions on fuel consumtion These results show that the fuel consumtion is lowest when the uniform coasting strategy is used. Comared to timetable-limited coasting, uniform coasting achieves significant fuel savings of 10% and 8% on main-line and branch-line routes resectively. This can be exlained by considering the general form of the fuel consumtion (fc) vs. journey time curve when coasting is alied for a journey between two stations, as shown in Figure (fc) 1 (fc) 2 fc Ft d fc dt 2 d fc 2 dt is ve is ve fc 1 fc 2 T- T T T+ T Time Figure 3-10 Illustration of the cause of fuel consumtion increase when timetable-limited coasting is used to achieve overall journey time The general shae of fc-time curve means that the increase in fc for a time reduction, -ΔT, is more than the decrease in fc for a time extension, +ΔT. Hence using less coasting in one section of the route and more in another tends to increase the overall fc for the same overall journey time. Comaring the results of the uniform coasting analysis with the route timetable shows that shorter inter-station distances have less margin in the schedule and so less coasting can be used. The extra time available for coasting on longer sections makes little difference to the fuel consumtion as significant coasting is already being used, and (fc)/ T is therefore small. 85

86 In ractice this means that time-table restrictions should be relaxed where ossible. This would be advantageous in reducing fuel use and encouraging drivers to take a consistent aroach to energy efficient vehicle control. The issue of time-tabling also has a direct imact on the erformance of hybrid trains due to the different braking strategies required. This is considered in Section Simle vehicle model of a hybrid regional train The second section of this analysis focuses on hybrid trains. In this model a generic energy storage module calculates the recovery and reuse of vehicle braking energy. This allows the otential benefits and basic requirements of a hybrid train to be quantified without requiring detailed knowledge of a articular energy storage system (ESS) or hybrid ower-train architecture. The erformance of the regenerative braking is investigated by limiting the maximum ower flow that the ESS can accet during braking; the results resented below show ower limits of 0.5 MW, 1 MW and an unlimited case (reresenting maximum ossible energy recovery). Two driving strategies are considered: i) Coasting strategy with constant force braking (as this was seen to be most effective in reducing fuel use for the conventional vehicle) ii) Coasting strategy with constant ower braking and maximum deceleration limit of 0.5 m/s 2 (to maximise braking energy recovery with ower-limited ESS) The energy storage module uses the recovered vehicle braking to relace energy delivered to the vehicle by the conventional ower-train oerating at the average efficiency for the journey. This allows the fuel consumtion of the hybrid train to be calculated Constant force regenerative braking with coasting The coasting strategy (i) results in fuel consumtion values shown in Figures 11a and 12a for the main-line and branch-line routes resectively. These values are calculated for the secified ower limitations and assuming an ESS with 100% efficiency, allowing the maximum otential fuel savings to be identified. The energy storage caacity required in both cases is shown in Figures 11b and 12b. 86

87 (a) (b) Figure 3-11 Effect of coasting on main-line hybrid erformance with ESS ower limitations (a) (b) Figure 3-12 Effect of coasting on branch-line hybrid erformance with ESS ower limitations For both the main-line and branch-line routes the fuel consumtion of the hybrid train with unlimited ESS ower caacity is seen to be aroximately constant across the range of coasting conditions. This illustrates the fact that coasting strategies and regenerative braking are alternative methods of utilising the finite kinetic energy of the vehicle rior to braking. The ower limitations of the ESS result in increased fuel consumtion. This is most significant for the main-line route where the high vehicle seed leads to high initial braking ower, severely limiting the amount of energy recovered. 87

88 Constant ower regenerative braking with coasting The outcome of the constant force braking analysis shows there is scoe for further benefits if a constant ower braking strategy (ii) is used. The trigger seed as a function of distance has been calculated for every station sto and change in line seed by erforming back calculations (including the effects of resistance and local gradient) from these oints, with the braking force limited by ower (1MW has been considered) and a maximum vehicle deceleration of 0.5m/s 2. This braking strategy can again be combined with coasting. An illustration of the effect of regenerative braking control strategies (i) and (ii) is shown in Figure 3-13 for an ESS ower limit of 1MW. The fuel consumtion is resented as a ercentage saving relative to the conventional vehicle using coasting and constant force braking. (a) (b) Figure 3-13 Calculated fuel saving and required energy caacity for hybrid train with 1MW ESS using constant ower and constant force braking strategies with various degrees of coasting on main-line (ML) and branch-line (BL) routes Figure 3-13 shows that the constant ower braking strategy increases the fuel saving for the HDT on both routes. In total the benefit is however much smaller for the branch-line (BL) than the main-line (ML). Although the results have been calculated for an ESS efficiency of 100%, the nature of the analysis method means that the calculated fuel saving relative to the conventional vehicle can be scaled in roortion to the round-tri efficiency of the ESS, which is likely to be in the region of 50-80% for ractical regenerative braking systems. In 88

89 summary the model shows that under the aggressive driving strategy u to 40% fuel savings are ossible for the HDT with an ideal ESS. However, when efficient driving strategies (i.e. coasting) are used the fuel savings are reduced by half. The model rovides a tool for identifying effective control strategies which should be imlemented to maximise the benefits of a HDT. It also allows an assessment of the aroriate ower and energy caacity of an ESS in order to minimise the additional cost and weight of a HDT Effect of timetable restrictions on erformance of hybrid regional train The revious analysis of HDTs assumes that a uniform coasting strategy is alied. It has already been shown in Section that a timetable-limited coasting strategy significantly increases the fuel consumtion for a conventional vehicle due to an increase in braking energy losses. While this means that more energy is available for regeneration, a larger energy storage caacity is required to cature this energy. Furthermore, the uneven nature of the coasting strategy means that available braking energy varies considerably along a route. This is illustrated in Figure 3-14 which shows braking energy available on the aroach to each station sto using either uniform or timetable limited coasting with constant force braking. 89

90 (a) 33.4 MJ (b) 37.1 MJ Figure 3-14 Braking energy available during aroach to (a) ML and (b) BL stations using uniform and timetable-limited coasting to achieve overall journey time; dotted lines shows minimum ESS caacity required to cature all available energy with uniform coasting (assuming 100% ESS efficiency) It is clear from Figure 3-14 that in order to maximise braking energy recovery for the timetable-limited coasting strategy energy storage caacities of 125 and 55 MJ are required for the main-line and branch-line resectively. The effect of timetable-limited coasting on the fuel savings ossible for the hybrid vehicle with a range of ower caacity limitations is shown in Figure The energy caacity is assumed to either be sufficient to store all available braking energy, or limited to the energy storage caacity values identified from the uniform coasting analysis for the main-line and branch-line routes. 90

91 Energy caacity = 34.3 MJ Energy caacity = 37.1 MJ MAIN-LINE BANCH-LINE Figure 3-15 Hybrid fuel savings relative to conventional vehicle using timetable-limited coasting to achieve overall journey time (assuming 100% ESS efficiency) The values of fuel savings using the stated energy caacity limits and a ower limit of 1MW are similar to those obtained for the uniform coasting strategy (see Figure 3-13). However, the timetable-limited coasting strategy causes an increase in the actual fuel consumtion of the HDT as shown in Figure Figure 3-16 ML and BL fuel consumtion for conventional and hybrid vehicles using uniform and timetable (TT) limited coasting strategies (ESS ower and energy caacities as stated and assuming 100% efficiency) These results show that the uniform coasting strategy allows lower fuel consumtion to be achieved with a HDT through better utilisation of the available braking energy. However, significant fuel savings can also be achieved using timetable-limited coasting with ractical ESS energy and ower limitations. 91

92 equirements for HDT regenerative braking system The results resented in this Chater show that significant fuel savings can be achieved through the use of hybrid systems to enable regenerative braking. When combined with efficient driving strategies an ESS with energy caacity of around 40 MJ and ower caacity of 1 MW (i.e. a characteristic discharge time of 40 seconds) has been identified as suitable to achieve high levels of braking energy recovery. As the conventional diesel-hydrodynamic ower-train has been shown to oerate with good efficiency over a wide range of oerating conditions, the hybrid system can be designed as a ure regenerative braking system and used to deliver tractive ower during (otherwise inefficient) low vehicle seed oeration. The energy and ower requirements combined with the short eriods of energy storage and the high secific torque ossible with mechanical transmissions mean that advanced flywheel systems are highly suited to this alication, and are considered in more detail in the following Chaters Summary In conclusion, there are a number of significant findings relating to both conventional and hybrid driving strategies. The most significant finding is that the otential savings for a hybrid regional diesel train (HDT) using efficient control are around half of that found using an aggressive driving strategy. This saving is significantly lower than redicted by other models but is realistic considering there are a range of driving strategies in use. Fuel savings reviously associated with hybrid trains can in fact be achieved through efficient control methods which can be imlemented in the short-term at low cost. This must be taken into account when considering the case for imlementing hybrid trains in the future. Different tyes of oeration will require different amounts of energy storage caacity for the HDT. This caacity should be tailored to secific routes in order to minimise the additional cost and weight of the hybrid train. 92

93 The constant ower braking strategy can increase energy recovery in cases where braking is imlemented at high vehicle seed if sufficient storage caacity is available to store the additional energy. Of the strategies studied for conventional vehicles, the use of maximum line seed limits with coasting rior to braking can achieve the lowest fuel consumtion for a given journey time with fuel savings of u to 36% relative to the worst case driving strategy. Advisory maximum seed limits are shown to be an effective means of regulating journey time and fuel consumtion for the main-line route. Alying coasting to meet the fixed timetable for intermediate stos increases the fuel consumtion of conventional vehicles by around 10% comared to a driving strategy where coasting is uniformly alied to achieve the overall journey time. This increases the available braking energy, but regeneration is limited by the HDT energy and ower caacity. The results of the conventional vehicle model show that under a range of driving conditions significant fuel savings are ossible using hybrid vehicles. Driving strategies governed by a simle control algorithm have been shown to achieve low fuel consumtion when alied to both conventional and hybrid trains. The required energy caacity of a HDT has been calculated as a function of the ESS ower caacity and the driving strategy used. It is imortant to now consider the comonents, configuration and control strategy of the ESS in order to more accurately assess the design requirements and erformance of hybrid trains. For mechanical flywheel systems, a key element of the FESS is the mechanical CVT (continuously variable transmission). It is therefore essential to; a) understand the interaction between the CVT, flywheel and vehicle b) secify aroriate gearing ratios for a given CVT configuration c) identify the ower transmission efficiency of a given CVT configuration d) identify the ower and torque requirements of CVT comonents e) Identify the most aroriate CVT configuration for a given alication These issues are addressed in Chaters 4 and 5 for FESSs containing CVTs controlled by brake, clutch and variator comonents. 93

94 4. Indeendent analysis of clutch and brake controlled flywheel transmissions Advanced flywheel technology has been identified as a candidate for reducing the energy consumtion of regional diesel trains. The challenge is to investigate how to use flywheel devices in a ractical regenerative braking system to achieve energy efficiency. A key element is to exlore the comromise between reducing vehicle fuel consumtion and increasing the weight of the vehicle. This Chater therefore aims to address mechanical flywheel transmission design and erformance using lanetary gearing controlled by brake or clutch comonents. The transmissions considered are similar to ower-shift automatic transmissions widely used in automotive ICE (internal combustion engine) ower-trains, but the requirements are different for flywheel alications. The transmission design is determined by the need to otimise system erformance, as discussed below for the two cases; i) ICE transmissions (current alication) An ICE is able to generate torque over a well defined range of seeds, with the transmission gearing ratios secified on the basis of overall fuel-efficiency (as the gear ratios affect the engine oerating conditions) and driveability. The engine is connected to the vehicle through a fixed gear ratio for a large roortion of the time, while ower-shifting is used intermittently to rovide smooth gear changes. This means that the ICE actively controls the oerating oint of the transmission. ii) Flywheel energy storage (roosed alication) In contrast, flywheels are assive devices which can only be charged and discharged when an aroriate torque is alied by the transmission. This can be achieved using fixed ratio gearing designed to oerate with a continuous state of ower-shift. The gearing ratios for this alication are secified on the basis of maximising the useful energy storage caacity er unit mass of the FESS (flywheel energy storage system). Previous aroaches to analysing FESSs have focussed on detailed modelling of individual vehicles, drive cycles and systems [ , 109] in order to reduce fuel consumtion for a articular case study. This aroach has a number of drawbacks. Firstly, it results in a secific system being defined by a large number of arameters, making the otimisation 94

95 rocess comutationally intensive and alicable only to the articular case considered. This tends to obscure the fundamental relationshis between the key FESS design arameters and the system oerating characteristics, making a general comarison of systems difficult. Secondly, an emhasis on minimising fuel consumtion can fail to consider imortant system design comromises relating to the mass, cost and efficiency of comonents and the comlexity of the transmission. In order to overcome these issues and assess the otential of a number of mechanical transmission configurations oerating over a wide range of conditions a new design tool has been created. This design tool consists of normalised datasets that characterise the erformance and design requirements of each transmission configuration. They are derived from analytical relationshis between the key system arameters. By secifying the vehicle arameters and the required energy storage caacity for a given alication, the design tool allows an initial estimate of the otimum gearing ratios required to achieve maximum secific energy for the FESS. This rovides an absolute basis for the comarison of a range of mechanical transmissions, and a strong starting oint for more detailed comutational analysis of secific FESS alications. The elements of the transmissions considered in this Chater are all widely used in standard vehicle ower-trains, and the systems resented therefore reresent a ractical, lowcost means of imlementing efficient flywheel energy storage. The outcome of this analysis suggests high erformance can be achieved using a simle and comact transmission consisting of a number of PGSs and a small counter-shaft control gearbox (CGB). These systems have the advantage of high secific energy when comared with ublished data for electrical flywheel systems Brake-controlled flywheel transmissions This section analyses simle brake-controlled transmissions consisting of lanetary gearing. In its simlest form, the transmission consists of a single PGS as shown in Figure 4-1(a). This system has been roosed by Diego-Ayala et al. [106] but for a secific automotive alication which has very different energy and torque requirements than a rail vehicle. The system is controlled by alying a braking torque to the control branch of the PGS (in this case the ring), which exerts oosite torques on the vehicle and flywheel. This allows the flywheel to be discharged, with a roortion of the energy being transferred to the vehicle while the remainder is dissiated due to friction at the control brake. This results in a fundamental limit on the overall efficiency of the transmission. 95

96 (a) fd PGS with fw ring 0 (b) Fixed gearing with ratio, K ω fd ω fw ω fd Final drive Flywheel Flywheel Clutch ing brake ω fw Figure 4-1 Examles of (a) brake and (b) clutch controlled flywheel transmissions An analysis of the systems illustrated in Figure 4-1(a) and 4-1(b) shows that they are equivalent when the gear ratios K and are equal. When multile gear ratios are used in either configuration the erformance of the transmissions can be imroved by reducing the total energy dissiated at the brake or clutch. Both tyes of transmission are attractive because of the mechanical simlicity and ease of control. PGSs are however referred for the flywheel transmission, as they have the following advantages; They can achieve high gear ratios which are required when using high seed flywheels, A large range of gear ratios can be achieved using single and multile stage PGSs, They can rovide high torque caacity er unit mass due to load sharing in the multile lanet gears, It is easier to manage the cooling and maintenance of PGS brake comonents than clutches. While the brake-controlled PGS transmission has been studied in its simlest form (using a single PGS), in this investigation a comrehensive analysis of multi-pgs brake-controlled transmissions has been erformed in order to identify the requirements for otimum FESS erformance, and to quantify the benefits of using multile PGSs. It is found that the transmission arameters can be chosen to achieve otimum system erformance within ractical limitations (here considered in terms of system mass and energy available for recovery during braking). The analysis method below has been develoed to rovide these results in a generalised form which is indeendent of the alication itself. The normalised 96

97 transmission arameters and associated erformance rovide a solid basis for comaring a range of mechanical transmission systems and allows the identification of aroriate gearing ratios for any vehicle alication. The analysis of PGS brake-controlled transmissions is described in the following sections Connection otions for differential gearing Using the general aroach to analysing differential gearing described by White [94] eliminates the need to secify the tye of differential gearing used and the way it is connected to the vehicle and flywheel. The differential is simly considered to have three branches and a characteristic gear ratio,, as defined in Chater 2. The kinematic and ideal torque relationshis for the generic differential are restated in Equation 4-1 and 4-2 resectively (4-1) T T2 T1 (4-2) 1 These equations aly to any tye of PGS in any transmission configuration. A general analysis (using Equations 4-1 and 4-2) can be erformed to identify an aroriate value of for a given alication, allowing the most aroriate PGS configuration to be deduced. In the case of a flywheel transmission this will deend uon the characteristics of both the vehicle and the flywheel unit. The results of a general analysis can therefore rovide a design tool that illustrates the effect of these factors on the erformance and the design requirements of various flywheel transmission configurations Aroach to indeendent analysis of PGS transmissions To generate this new design tool, the following analysis method is alied. The oeration of a differential gearing unit is described by the kinematic relationshi (Equation 4-1) and the ideal torque relationshi (Equation 4-2). The effect of PGS gearing losses can be included in the analysis by using a simle assumtion of constant transmission efficiency, η gs. The ideal torque equations can then be modified (deending on the direction of ower flow in the three branches) to take these losses into account. The torque equations can therefore be defined as in Table 4-1 for all 6 ossible cases of PGS ower-flow (a-f). 97

98 Power-flow case P1 P P 2 3 T 1 T 2 1 a + gs 1 b + gs 1 c gs 1 d + + gs 1 e + 1 f T 3 T gs gs gs 1 gs 1 1 Table 4-1 Direction of ower flow in PGS branches (+ve in, -ve out) and associated torque equations including the effect of constant PGS transmission losses for all ossible cases In the ring-brake transmission illustrated in Figure 4-1(a) the sun gear can be defined as corresonding to branch 1 and the carrier to branch 3 of a generic differential. As ower is always flowing out of the PGS at the ring (branch 2) and is dissiated in the brake, flywheel charging and discharging with this tye of transmission therefore corresonds to ower-flow cases (e) and (a) in Table 4-1 resectively. The kinematic and torque equations for the PGS are then identical to the secific case described by Diego-Ayala et al. [106] where the ratio of the PGS is defined in terms of the articular configuration and the number of teeth in the ring and sun gears. The general form of these equations are however used in all subsequent analysis due to the ability to identify both the most aroriate tye of differential and how the branches should be connected to the vehicle and flywheel for a given alication. As the aim of this Chater is to develo a generic design tool which is indeendent of the vehicle characteristics and drive cycle, the effects of dissiative losses at the vehicle (due to rolling and aerodynamic resistance) and flywheel (due to self-discharge) are disregarded. These assumtions can be justified as follows; i. The maximum tractive effort is much greater than the force due to resistance at low vehicle seeds (max(t trac ) >> T veh res ) ii. Advanced flywheel self-discharge losses of 2-4% er minute are reorted in the literature [109]. As tyical acceleration and braking events will be of the order of 1 98

99 minute, this resistance torque is much lower than the maximum torque exerted on the flywheel by the transmission (max(t fw ) >> T fw losses ) The effect of these assumtions will be assessed in Chater 6. The simlification is however extremely useful, as by disregarding the losses at the vehicle and flywheel, the analysis corresonds to the case when vehicle and flywheel acceleration are infinite. The results are therefore indeendent of time, and so do not deend on any articular drive-cycle. With no vehicle or flywheel losses, the following equations aly at the inut and outut of the transmission (where the fd subscrit refers to the conditions on the transmission side of the final drive gearing): T T fd fw d fd J fd (4-3) dt d fw J fw (4-4) dt In Equation 4-4, J fw and ω fw reresent the outut inertia and seed of a flywheel unit, which can include a fixed gear ratio between the high seed rotor and the outut as shown below in Figure 4-2; J Outut shaft fw fw K K 2 fw rotor fw J rotor Flywheel unit Fixed gear ratio, K fw otor Figure 4-2 Definition of relationshi between the flywheel rotor and the outut characteristics of the flywheel unit Defining the flywheel in this way rovides freedom to secify ractical values for the PGS(s) and identify the required value of K fw for a given maximum rotor seed and kinetic energy caacity. Differentiating the PGS kinematic relationshi (Equation 4-1) with resect to time results in; 99

100 d 1 3 dt d2 d 1 (4-5) dt dt For a secific transmission configuration we can now investigate the oeration by combining Equations 4-1 and 4-3 to 4-5 with the aroriate torque relationshis defined in Table 4-1. This aroach is defined here as the Indeendent Analysis Method (IAM), and allows analytical exressions to be obtained which define the oeration of the transmission. The effect of flywheel energy caacity, flywheel DOD (deth-of-discharge) and gear ratios on the transmission efficiency and system mass can be quantified, roviding a fundamental basis for understanding and comaring the erformance of different transmission configurations. This aroach allows exact analytical exressions to be derived to describe the erformance of a brake-controlled transmission with m PGSs in terms of the following key system arameters; i. The PGS characteristic gear ratios n (where n = 1, 2,..., m) ii. The estimated constant efficiency of each PGS iii. The ratio of equivalent inertias at the transmission outut (final drive) to inut (flywheel unit) iv. The ratio of the maximum to minimum transmission inut (flywheel unit) seeds for a full discharge event as defined by the overall deth of discharge (DOD ov ) where; DOD ov KE fw max KE fw 1 fw, min fw, max 2 (4-6) The erformance of secific mechanical flywheel systems is considered in the following sections Analysis of multi-pgs brake-controlled transmission The Transmission Configuration Figure 4-3 illustrates the layout of a generic brake-controlled flywheel transmission using multile PGSs. Sequential braking of branch 2 of each PGS allows a number of hases of transmission oeration to be achieved; by increasing the number of PGSs (and therefore the number of oerating hases) the total amount of energy dissiated in the brakes can be reduced for a given flywheel DOD ov. 100

101 Differential gearing Vehicle Flywheel unit Final drive m 3 eresents a brake acting on branch 2 of each differential Figure 4-3 Schematic diagram of brake-controlled flywheel transmission with m differentials (arrows show the direction of ower flow during flywheel discharge using PGS 1) Examle of Transmission Oeration The oeration of a brake-controlled transmission can be understood by considering a secific configuration. Figure 4-4 is a schematic illustration of a ossible 2-PGS ring-brake transmission, while Figure 4-5 shows the oeration of this system during a full flywheel discharge/charge cycle. PGS gearing Vehicle Flywheel Final drive PGS: 1 2 Figure 4-4 Examle of a ossible configuration for 2-PGS ring-brake controlled transmission (note: 0 < 1 < 2 < 0.5 in this configuration) 101

102 ω 0 Time Seeds Final drive Flywheel ing of PGS 1 ing of PGS 2 Brake alied 2 discharge hases Engine oeration 2 charging hases Conventional braking Figure 4-5 Illustration of simle flywheel discharge/charge cycle with a 2-PGS ring-brake transmission, constant T fd and DOD ov = 75% During flywheel owered vehicle acceleration, braking is alied at the ring rotating with negative seed of the smallest (non-zero) magnitude. Once all ring seeds are ositive or zero the transmission can no longer discharge the flywheel and the conventional ower-train must be used to continue accelerating to the vehicle. egenerative braking is achieved by alying a braking torque to the ring rotating with the smallest ositive (non-zero) seed. Once all the ring seeds are zero or negative conventional braking must be alied to bring the vehicle to a sto. The analysis of the system during both charging and discharging of the flywheel using the generic PGS equations follows. Indeendent Analysis of Transmission Oeration By considering the balance of energy during each hase of transmission oeration we can investigate the overall erformance of the brake-controlled system. The generalised analysis rocedure described earlier is therefore alied to each hase of a multi-pgs brakecontrolled transmission with the aroriate initial conditions. The general exressions defining the oeration of the transmission are as follows. If the transmission consists of m PGSs, the ratio of outut to inut torque during brakecontrol of the n th PGS is as follows; T T fd fw T 3, n (4-7) T 1, n 102

103 Substituting Equations 4-3 and 4-4 into Equation 4-5 obtains; d 2, n 1 T3, n T n 1, n dt 1 n J fd 1 n J fw (4-8) The efficiency of the brake-controlled transmission is related to the total amount of energy dissiated in the brakes located on branch 2 of the m PGSs, E 2, tot. The total energy dissiated in the n th transmission hase, E 2,n, is shown in Equation 4-9 where; t i is the initial time at which ω 2,n = ω 2,i (with ω 2,(n+1) = 0 during charging and ω 2,(n-1) = 0 during discharging) t f is the final time at which ω 2,n = 0 t f 2, nt2 n E 2, n, dt (4-9) ti Substituting Equation 4-8 into 4-9 allows the integration to be exressed w.r.t. the rotational seed of branch 2 of the n th PGS, ω 2,n, as shown in Equation , n 0 n fd fw nt 2, 2, n 2, i fw 3, n n fd 1, n 1 J J E d (4-10) 2, n J T J T Equations 4-7 and 4-10 can be derived for the two cases of flywheel charging and discharging by using the aroriate torque ratios defined in Table 4-2. These results are shown in Table 4-3 below. T T fd fw E 2, n 2 Flywheel discharge Flywheel charging J gs, n gs, n fw fd n 2 2, i J J 2 gs, n J fw n n 1 gs, n 2 1 J J 1 fd 2 gs, n fw fd n 2 2, i J J 2 fw gs, n n n fd 2 Table 4-2 Torque and energy dissiation relationshis for a hase of oeration using brakecontrolled transmission The exressions for E 2, n are for a single hase of oeration. The overall erformance of the transmission can therefore be assessed by considering the total energy dissiated during flywheel charging or discharging, where; 103

104 m E 2, n E2, tot (4-11) n1 These results make it clear that by assuming no losses at the flywheel or vehicle, the amount of energy dissiated at the ring is indeendent of the torque alied at the ring as a function of time, and deends only on the initial rotational seed of branch 2 in each hase of oeration. This exression for the energy dissiated at branch 2 can be substituted into the overall energy balance for the n th PGS during flywheel discharge (Equation 4-12) to obtain Equation gs, ne 1, n E3, n E2, n gs, n 1 2 J fw J 1, i gs, n fw fd n 2 2, i J J 2 gs, n 2 (4-12) J J 1 1, f fd 3, f 3, i (4-13) 2 2 The following relationshis aly for the initial and final seeds in the n th transmission hase. The initial seeds of branches 1 and 3 corresonds to the final oerating oint of the (n-1) th transmission hase. As the final seed of branch 2 of the n th PGS is zero; (4-14) 3, i n1 1, i (4-15) 3, f n 1, f fw n fd The initial seed of branch 2 can be defined in terms of the initial seeds of branches 1 and 3 according to Equation 4-1; 2, i 1 1 n 1 n 1, i n 3, i 1, i 1 n 1 n 1 (4-16) n Substituting Equations 4-14 to 4-16 into Equation 4-13 results in the following exression for the ratio of initial to final flywheel seeds during the n th hase of flywheel discharge. 1, i 1, f n n n 2 n1 gs, n gs, n J fw J fd J fw J fd (4-17) Equation 4-17 shows that the ratio of initial to final flywheel seeds is a function of the PGS ratios, the inertia ratio (defined here as J fd /J fw ) and the efficiency of the PGS. Furthermore, a normalised PGS ratio,, can be defined such that; 104

105 J fd n (4-18) n J fw The erformance of the transmission is then comletely defined by the m values of and η gs. Substituting Equation 4-18 into 4-17, the characteristic equation describing a hase of brake-controlled flywheel discharge is obtained; Characteristic equation for the n th hase of brake-controlled flywheel discharge 1, i 1, f n 2 n n n1 gs, n gs, n (4-19) This analysis can be reeated for the case of flywheel charging, and a similar exression can be derived as shown in Equation Characteristic equation for the n th hase of brake-controlled flywheel charging 1, i 1, f n gs, n 2, n n 1 gs n n1 1 (4-20) For vehicle acceleration (flywheel discharge) from stationary, the initial vehicle seed in the first transmission hase (i.e. braking branch 2 of PGS 1) is zero, and so the value of 0 required in Equation 4-19 is zero (from Equation 4-15). For vehicle braking (flywheel charging) using the m th PGS, the initial vehicle and flywheel seeds can be used to calculate a hyothetical value, h, for m+1 (equal to ω 3,i / ω 1,i ) which allows the final seeds of the vehicle and flywheel to be calculated for the hase of oeration. These two characteristic equations therefore comletely define the oeration of the brakecontrolled flywheel transmission, and allow a thorough investigation of the effect of gearing ratios, PGS efficiency and inertia ratio on system erformance. Exressions for the overall charging and discharging efficiencies can be derived as follows; Overall brake-controlled transmission discharge efficiency discharge KE KE veh fw J fw J fd 2 1 max 2 3 max 2 1 min m 1 max 1 min (4-21) 105

106 Overall brake-controlled transmission charging efficiency KE J 1 max min 1 min 2 2 fw fw 1 max charge (4-22) KEveh J fd 3 max h 2 1 Equations 4-19 to 4-22 are key results, as they show that analytical exressions can be derived defining the normalised PGS ratios and the overall efficiency as functions of the m values of DOD n. This allows the exact calculation of transmission erformance which is indeendent of the alication (defined by the actual vehicle and flywheel inertias), and rovides a basis for the direct comarison of different transmission configurations. The results of this analysis follow for the two distinct cases of single and multile PGS transmissions esults for single PGS brake-controlled flywheel transmission The characteristic equations for flywheel discharging and charging (Equations 4-19 and 4-20 resectively) can now be alied to the case of a brake-controlled transmission consisting of a single PGS with gearing ratio. This allows analytical exressions to be obtained for four key results; i. A normalised gear ratio, ii. The transmission charge, discharge and round-tri efficiencies iii. A flywheel utilisation factor, U iv. A normalised energy dissiation at the control brake These four areas are discussed below. A design method using these results is then described, which rovides a means of identifying aroriate FESS design arameters for any alication. i) Normalised PGS gear ratio From the characteristic discharge equation (Equation 4-19), it is clear that the normalised PGS gearing ratio,, is defined as a function of the PGS efficiency, η gs, and the ratio of maximum to minimum flywheel seeds (a monotonic function of DOD ov ), as shown in Equation

107 1 max gs 1 (4-23) 1 min It is clear from this result that the value of is urely a function of DOD ov. Figure 4- gs 6 illustrates this relationshi. DOD ov Figure 4-6 Value of gs as a function of DOD ov for a single PGS brakecontrolled transmission Substituting the exression for (Equation 4-23) into the characteristic equation for flywheel charging (Equation 4-20) allows an exression to be derived for the normalised hyothetical PGS ratio, h, which defines the initial braking conditions required to fully recharge the flywheel from minimum to maximum seed. h 1 max 1 1 max 1 2 gs 1 min gs 1 min (4-24) This exression for h allows the erformance of the system to be assessed for a full flywheel recharge event. ii) Transmission efficiencies Equations 4-23 and 4-24 can now be substituted into the exressions for discharge and charge efficiency (Equations 4-21 and 4-22) resectively, resulting in Equations 4-25 and 4-26; 107

108 discharge 1 max 1 min 2 2 fw 1 fw gs,max,min 1 (4-25) 2 1 max 1,max fw 1 1 min fw,min (4-26) charge 2 2 h,max 1 fw gs 2 fw,min gs These discharge and charge efficiencies are illustrated in Figure 4-7 as functions of DOD ov and constant PGS efficiency, η gs. η disharge η charge (a) DOD ov (b) DOD ov Figure 4-7 (a) discharge and (b) charging efficiency as a function of η gs and DOD ov for a 1- PGS brake-controlled transmission The roduct of the overall charging and discharging efficiencies is defined as the round-tri efficiency, and is illustrated in Figure

109 η round-tri DOD ov Figure 4-8 ound-tri efficiency of the 1-PGS brake-controlled transmission as a function of and DOD gs ov iii) Flywheel utilisation factor It is aarent from these results that when choosing the PGS gear ratio there is a comromise between achieving a high transmission efficiency or a high DOD ov. This is an imortant consideration as it will affect the useful energy that can be delivered to the vehicle er unit mass of the system, defined here as the secific energy caacity. A flywheel utilisation factor has therefore been defined as the roortion of the kinetic energy of the fully charged flywheel that can be delivered to the vehicle during a flywheel discharge event, as shown in Equation U The value of KE veh max KE discharge fw DOD ov KE veh KE discharge fw DOD ov discharge U is therefore only a function of DOD gs ov as shown in Figure 4-9. (4-27) Figure 4-9 The value of DOD ov U gs as a function of DOD ov for the 1-PGS brake-controlled transmission 109

110 The maximum flywheel utilisation is U 0.25 and occurs when DOD gs ov and gs. For a known inertia ratio and PGS efficiency, the value of required to achieve maximum U can be found. Oerating with maximum U minimises the mass of the flywheel required for a articular useful energy caacity of the system. The requirements of the brake comonents in the transmission will however affect the total transmission mass, and deend on the amount of energy being dissiated in branch 2 of the differential, as discussed below. iv) Normalised energy dissiation at control brake The energy dissiated in the control brake can be evaluated by substituting the exressions for and h (Equations 4-23 and 4-24) into the exressions for E 2 during charge and discharge as defined in Table 4-2. Normalised values of energy dissiation at branch 2 of the differential, E 2, can be defined by dividing E 2 by the maximum kinetic energy of the flywheel, as shown below. E2, discharge 1 min E 2 gs 1 discharge (4-28) max KE fw 1 max 2 1 min 1 min 1 gs E 1 max 1 max 2, charge gs E 2 charge (4-29) maxke fw 2 1 max 1 1 gs 1 min The values of E 2 during both flywheel charging and discharging are functions of η gs and DOD ov, as illustrated in Figure

111 (E2) discharge (E2) charge (a) DOD ov (b) DOD ov Figure 4-10 Normalised energy dissiation at the control brake during discharging and charging of flywheel, as a function of and DOD gs ov for a 1-PGS brake-controlled transmission An assumed worst case for energy dissiation at the control brake is here considered to be the case when a full flywheel charge/discharge cycle occurs in a short sace of time. The following exression therefore reresents the total amount of energy dissiated at the control brake during a charge/discharge cycle er unit of energy delivered to the vehicle during flywheel discharge; E E E E useful E cycle 2, charge 2, discharge 2 KE U 2 veh discharge cycle (4-30) This is an imortant factor, as it allows the brake comonents in the transmission to be sized for a given useful energy caacity of the FESS. It is a function of PGS efficiency and DOD ov as illustrated in Figure

112 (E2)cycle / U DOD ov Figure 4-11 Total energy dissiation in the control brake during a full charge/discharge cycle er unit energy delivered to the vehicle during flywheel discharge as afunction of and gs DOD ov for a 1-PGS brake-controlled transmission The oerating characteristics of the transmission are now comletely defined. In summary, the 1-PGS brake-controlled transmission is seen to oerate with low charge and discharge efficiencies, with a maximum of 25% of the flywheel kinetic energy being delivered to the vehicle. The large amount of energy dissiated at the control branch of the PGS also mean that relatively large brake comonents are required, adding to the mass of the system. This system is therefore unlikely to be a suitable choice for vehicle regenerative braking alications. There are however three imortant conclusions from the analysis of this transmission; i. The analysis method described roduces exact analytical exressions describing the erformance of the flywheel system, and therefore rovides a robust basis for the comarison of different transmission configurations. ii. Normalising the key arameters that define the oeration of the transmission allows the erformance and design requirements to be characterised as functions of the overall flywheel deth of discharge, DOD ov, and the PGS efficiency, η gs. The normalised arameters are therefore indeendent of the actual vehicle or flywheel characteristics. A summary of the arameter deendence for the 1-PGS brake-controlled transmission is resented in Table 4-3 to illustrate these relationshis. 112

113 Parameter Flywheel discharging Flywheel charging PGS gear ratio fn gs DOD ov discharge Efficiency fndod Energy dissiated at control brake Flywheel utilisation E gs ov charge gs fn DOD ov, 2 discharge 2 charge fndod fndod, gs U gs fn DOD ov Table 4-3 Summary of normalised arameter deendence for 1-PGS brake-controlled ov transmission E gs ov gs gs iii. The analysis generates a general design tool which allows arameters to be secified for a articular alication in order to achieve an otimum condition (for examle, maximising the secific energy caacity of the FESS). This is achieved by erforming the following rocedure; Design Method 1. Define J fd, E useful (the energy delivered to the vehicle during full flywheel discharge), max(ω fw ), max(t fd ) and η gs for a articular alication 2. Use E useful to find E 2,cycle (from Figure 4-11) and calculate the mass of the brake comonents as a function of DOD ov 3. Use E useful and U (Figure 4-9) to find max(ke fw ) and calculate the mass of the flywheel as a function of DOD ov 4. Use max(ke fw ) and max(ω fw ) to find the required J fw as a function of DOD ov 5. Use J fd and J fw to find (from Figure 4-6) as a function of DOD ov 6. Calculate the mass of the PGS(s) based on the maximum torque 7. Calculate the total system mass as a function of DOD ov and identify the value of DOD ov that results in minimum mass 8. Identify the ± value required to achieve this DOD ov 9. Use Figure 2-22 to identify suitable differential tye and configuration to achieve the required if the required gear teeth ratios are imractical for both s then adjust the value of max(ω fw ) and reeat stes 5-9 (note that this allows K fw to be identified to allow a ractical PGS to be used with a given maximum flywheel rotor seed) 113

114 10. Use Figure 4-8 to find the value of (E useful / η round-tri ) for the chosen FESS design if this is greater than the tyical amount of energy available during vehicle braking the flywheel is over-sized and will not be fully charged during a tyical braking event. One of the following measures should then be considered; o Decrease the required value of E useful o Oerate at lower DOD ov and therefore higher η round-tri (although this will decrease the secific energy caacity of the system) o Use a more efficient transmission configuration Stes 1-10 are then reeated until a ractical FESS design is identified. The following Section discusses the use of the IAM to investigate brake-controlled transmissions containing multile PGSs. The alication of the design method described here is then erformed for both single and multile PGS brake-controlled transmissions in Section esults for multile PGS brake-controlled transmissions The characteristic charge and discharge equations for a hase of brake-controlled transmission oeration can be alied in an iterative calculation rocedure to investigate the erformance of multile-pgs transmissions. In the case of a transmission with m PGSs, the oeration of the transmission is a function of the m values of and η gs. Once these values are secified, the final flywheel SOC in each hase of brake-controlled oeration can be calculated by secifying the initial value as equal to the final value of the revious hase. This allows the overall transmission efficiency and the normalised brake energy dissiation to be calculated. By secifying the value of the first PGS, 1, a minimising search function can be used to identify the remaining m-1 values of that result in maximum flywheel discharge efficiency. This can be erformed for a range of 1 values. The maximum efficiency (along with the m values of required to achieve this, and the normalised brake requirements) can therefore be defined as a function of DOD ov. Analysis has been erformed using a reresentative value of for all of the m PGS units in the multi-pgs brake-controlled transmissions. This means that in order to achieve the maximum discharge efficiency; gs x 114

115 The values of n, x x discharge U, and E U x r dischar ge the value of x must be secified in order to evaluate are only functions of DOD ov E 2 and A detailed series of results is resented here for the case of a 5-PGS brake-controlled transmission charge U cycle x x x x x DOD ov Figure 4-12 Values of n x required to achieve maximum discharge as a function of DOD ov for a 5-PGS brake-controlled transmission max discharge max U x x ηround-tri DOD ov DOD ov Figure 4-13 discha max, rge x max U x and round-tri efficiency as functions of DOD ov for a 5-PGS brake-controlled transmission (using the n values secified in Figure 4-14) 115

116 (E2)cycle / U DOD ov Figure 4-14 Total energy dissiated at transmission brakes in a full charge/discharge cycle er unit energy delivered to the vehicle during flywheel discharge as afunction of a 5-PGS brake-controlled transmission x and DOD ov for These results show that the 5-PGS brake-controlled transmission can achieve significantly higher efficiency and lower total energy dissiation in the control brakes than the 1-PGS transmission. This is however achieved by increasing the number of transmission comonents which is likely to increase the total mass and cost of the system. In summary, the analysis method resented above allows the key FESS arameters to be generated in a normalised form. A design tool has therefore been created which consists of datasets describing the maximum value of η discharge (with the associated η round-tri and m values of ) and E 2,tot as functions of the flywheel DOD ov and the efficiencies of the m PGSs. This normalised data can then be used to identify the aroriate design oint for a articular transmission configuration and alication. By erforming this analysis for a range of configurations, the most aroriate system for a given alication can be identified. The analysis resented in section is for the alication of brake-controlled flywheel transmissions to regional rail vehicles, and illustrates the comromise between the system mass and erformance egional rail alication of flywheel system with brake-controlled transmission The alication of multi-pgs brake-controlled transmissions to regional rail vehicles can be investigated by considering the mass and efficiency of flywheel systems. Two factors are imortant when assessing the suitability of a articular system; 116

117 i. The mass of the system er unit of useful energy, E useful, that can be delivered to the ii. vehicle from a fully charged FESS The amount of braking energy which is available for regeneration If a system is fully charged, then a given amount of energy can be delivered with minimum system mass by choosing the aroriate gearing ratio(s). It is however ossible that the amount of energy available during vehicle braking may not be enough to fully recharge the system, in which case the flywheel is oversized for the alication. Both of these issues can be illustrated by using the following factors to characterise the total mass (flywheel and transmission) of a given system; Secific torque of a single PGS (a value has been identifed which is consistent for both low seed industrial and high seed aerosace units [ ], although the costs are likely to differ); T gs = 50 Nm/kg Secific energy dissiation for transmission brake disc and calliers (value identified from consideration of conventional rail vehicle brake comonents [132]); E brake = MJ/kg Secific energy of an advanced flywheel energy storage device (intermediate value selected from Figure 2-6); E fw = 0.05 MJ/kg These are reresentative values taken from the literature which when combined with the results of the receding analysis allow the limitations of brake-controlled transmissions to be investigated. This aroach offers a starting oint for analysis, and can be refined as art of a design rocess for articular flywheel units and romising transmission configurations. In the alication of FESSs for regional rail vehicles it is assumed that the standard final drive gear unit (as found in conventional vehicles) is used. The actual PGS ratios used in the flywheel transmissions are constrained to have an absolute value of less than one. This constraint is alied to ensure that the seed of the PGS branch connected to the flywheel is greater than the seed of the branch connected to the final drive of the vehicle, reducing the gear ratio required in the flywheel unit, K fw. The maximum torque in the PGS therefore always occurs at the branch connected to the vehicle final drive. The required maximum torque at this branch has been secified as 10 knm in order to achieve a maximum constant 117

118 vehicle acceleration rate of ±0.5 m/s 2 for a tyical 3-car regional diesel train. For a transmission with a secified number of PGSs, N gs, the mass of the PGS, brake and flywheel comonents (m gs, m brake and m fw resectively) can be estimated as follows; m m gs brake gs N gs max T fd (4-31) T E E useful 2, tot cycle (4-32) E U brake m fw E useful (4-33) U E fw When the system is limited by the available vehicle braking energy, E avail, the following equation defines the useful energy that can be returned to the vehicle; E (4-34) useful E avail roundtri The erformance of the FESS can be characterised by defining a secific energy caacity, SE, as follows; SE useful (4-35) m gs E m brake m fw This defines the amount of energy that can be delivered to the vehicle from a fully charged flywheel er unit mass of the FESS. These equations can be alied to the results of the brake-controlled transmission analysis in order to comare the erformance of different transmission configurations. Figure 4-15 shows a lot of SE against E useful (the energy delivered to the vehicle during full flywheel discharge) using a 1-PGS brake-controlled transmission. Two conditions are illustrated; 1. The system is assumed to be initially fully charged and arameters are secified in order to minimise the mass for a given amount of useful energy delivered to the vehicle (solid line), 2. The energy available during braking is limited to 40 MJ (based on earlier vehicle analysis) and the useful energy that can be recovered and returned to the vehicle is limited by the round-tri efficiency of the transmission. The useful energy can be increased by reducing DOD ov, which increases the transmission efficiency but also increases the required energy caacity (and mass) of the flywheel (dashed line). 118

119 Euseful / FESS mass (MJ/tonne) Max SE of fully charged FESS (occurs when DOD ov 75%) Decreasing DOD ov and U, and increasing η round-tri Max SE when available braking energy is limited to 40 MJ E useful (MJ) Figure 4-15 Limitations on the secific energy caacity (SE) of a 1-PGS brake-controlled flywheel system for a regional rail alication (E avail = 40 MJ, η gs = 100%) It is clear from Figure 4-15 that the 1-PGS brake-controlled transmission is limited in how much braking energy can be recovered and made available for the acceleration of the vehicle (less than 25% of the braking energy). The maximum secific energy of the system is around 8.5 MJ/tonne when the useful energy caacity is 6 MJ (15% of the available energy); this is achieved with DOD ov = 60%. The increased efficiency that can be achieved using multile PGSs allows a higher degree of braking energy recovery. Figure 4-16 shows the results of this analysis for 1-5, 10 and 15 PGS brake-controlled transmissions when E avail = 40 MJ. 119

120 Euseful / FESS mass E useful (MJ) Figure 4-16 Limitations on the secific energy caacity of 1-5, 10 and 15 PGS brake-controlled flywheel systems for a regional rail alication (E avail = 40 MJ, η gs = 100%) Figure 4-16 shows that the highest secific energy caacity for a multi-pgs brake-controlled transmission is aroximately 12 MJ/tonne and occurs when a 4-PGS transmission is used to achieve a useful energy caacity of 20 MJ (i.e. a round-tri efficiency of 50%). Using transmissions with more PGSs (and therefore higher charge/discharge efficiency) allows more braking energy to be recovered at the exense of increased transmission mass and reduced secific energy. In ractical terms, a high number of PGSs (articularly multi-stage tyes) are likely to be difficult to connect together and idling losses in the gearing will be high. These results suggest that multi-pgs brake-controlled transmissions can rovide a viable regenerative braking system when the chosen energy caacity is significantly lower than the available energy exected during a braking event, and the transmission efficiency is therefore not a critical factor. Alternative transmission systems are however required if both a high roortion of braking energy recovery and a high secific energy are to be achieved. 120

121 4.2. Indeendent analysis of clutch-controlled flywheel transmissions The receding analysis shows that a simle brake-controlled transmission consisting of a large number of PGSs can achieve high overall oerating efficiency, and high utilisation of installed flywheel energy caacity. Increased efficiency is however gained at the exense of increased transmission size and weight, limiting the secific energy caacity of the system. This can be addressed by considering ower-slit configurations Clutch-controlled ower-slit transmission modes Power-slit configurations are often defined as inut-couled or outut-couled, which describes whether the ower-slit branch connects to the inut or outut of the transmission. This is an imortant distinction, as these two configurations exhibit different characteristics. In a flywheel transmission, the definition of inut and outut is however confusing as ower flows in different directions during charging and discharging. Also, ower-slit configurations are ossible between two differential gearsets, which do not fall under either category. To ensure clarity, the following terminology is therefore used to describe three different ower-slit configurations analysed in this section. In all cases, the flywheel is assumed to be connected to branch 1 of the differential(s) and the final drive to branch 3. a) Final-drive couled (FDC) ower-slit connection between branches 2 and 3 b) Flywheel couled (FWC) ower-slit connection between branches 2 and 1 c) Dual differential couled (DDC) ower-slit connection between branch 2 of two differentials In all three cases, the ower-slit can be achieved by using fixed gearing with a sliing clutch to control the torque exerted on the vehicle. A detailed analysis of this tye of system is resented below for the FDC configuration. It is shown that the oeration of the clutchcontrolled transmission can be related to the analysis of the brake-controlled transmissions discussed in Section 4.1. This is extremely useful as it allows the same simle characteristic charge and discharge equations (Equations 4-19 and 4-20) to be alied to the more comlex ower-slit transmission configurations once equivalent PGS ratio and efficiency arameters have been secified. The characteristics of the FWC and DDC configurations are 121

122 then briefly discussed, before a comrehensive results table for all three configurations is resented. a) FDC clutch-controlled PST The first ower-slit configuration considered is the final drive couled (FDC) transmission layout which uses a fixed gear ratio and clutch-control to achieve a ower-slit between branches 2 and 3 of the differential. This is illustrated in Figure 4-17, with the arrows defining the direction of ower-flow with no ower recirculation during flywheel discharge. Vehicle Flywheel unit Differential gearing Final drive K K o Control gearbox (CGB) G Fixed Gear atios K = ω 4 / ω 2 K o = ω 6 / ω 3 G = ω 5 / ω 4 Figure 4-17 FDC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) In order to achieve the required ower-slit during flywheel discharge; If > 1, then T 2 /T 3 > 0 (from Table 4-2), therefore ω 2 /ω 3 > 0 for P 2 /P 3 > 0 If < 1, then T 2 /T 3 < 0 (from Table 4-2), therefore ω 2 /ω 3 < 0 for P 2 /P 3 > 0 In either case, the ower transfer from shaft 5 to shaft 6 via the clutch can only be achieved when; (4-36) 122

123 Substituting the exressions for the fixed gear ratios and considering the required sign of (ω 3 /ω 2 ) results in; GK 1 K o (4-37) The clutch therefore locks u when; 3 2 lock 1 GK 1 K o (4-38) At this oint, ower flow through the fixed gear ratio G is no longer ossible. In the configuration illustrated in Figure 4-17 the torques exerted at the transmission inut and outut during flywheel discharge are defined as follows: 1 1 T fw T2 discharge (4-39) K 2 (4-40) 1 cgb o T fd T 1 1 discharge 1 GK T T fd fw discharge cgbk o 1 1 (4-41) GK The torque ratio defined here can also be achieved by using a brake-controlled PGS transmission, as described in Section 4.1. This occurs when the equivalent brake-controlled PGS has a ratio, eq, and efficiency, η eq, such that, eq discharge eq cgbko 1 1 Torque relationshi (4-42) GK This torque relationshi suggests that the analysis method reviously used for brakecontrolled transmissions can be alied to clutch-controlled PSTs. In order to achieve the same overall erformance as a brake-controlled transmission a seed relationshi must however also be defined. This ensures that the seed ratio 3 1 is exactly the same at the end of the clutch-controlled hase (when the clutch locks u) as it is at the end of the equivalent brake-controlled hase (when the seed of branch 2 of the equivalent PGS reaches zero). By considering this final condition of the oerating hase for both the brake-controlled transmission and the clutch-controlled PST, the following seed relationshi can be derived; 123

124 3 1 lock eq 1 K o 1 GK Seed relationshi (4-43) If both the torque and seed relationshis are obeyed, a clutch-controlled PST will therefore exert the same torques on the vehicle and flywheel over the same range of oerating seeds as the equivalent brake-controlled transmission. The erformance of the two systems is then identical, with an equal amount of energy being dissiated in the transmission of each. By substituting Equation (4-43) into Equation (4-42) it is clear that this is achieved when the efficiency of the PGS in the equivalent brake-controlled transmission, η eq, is as follows; eq 1 eq discharge cgb cgb Equivalent PGS efficiency (4-44) There are a number of interesting oints to be drawn from these results; The FDC clutch-controlled PST always has a value of eq less than the actual ratio of the PGS used in the ower-slit, The analysis of clutch-controlled PSTs can be erformed using equivalent values of and η with the characteristic equations reviously derived for brake-controlled transmissions (Equations 4-19 and 4-20) Maximum ossible value of η eq occurs when; eq, eq gs Minimum ossible value of η eq occurs when; eq 0, eq gs cgb The efficiency of the equivalent brake-controlled PGS is therefore always less than the efficiency of the actual PGS in the clutch-controlled PST (due to additional losses that occur in the CGB) The use of a multile seed gearbox (with a range of discrete values of G) in this configuration can allow multile hases of transmission oeration using a single PGS. This otentially allows high transmission erformance to be achieved with a significant reduction in gearing comared to the brake-controlled transmissions discussed earlier. The torque and seed requirements of the CGB must now be considered, as this will affect the mass of the system. With reference to Figures 4-17 and using the seed relationshi (Equation 4-43) the following exression can be derived for the ratio of CGB outut torque to the transmission outut torque in the FDC ower-slit mode; 124

125 T cgb, o discharge K o T T 5 fd cgb K o GK cgbk 1 GK o 1 1 cgb 1 eq 1 1 cgb eq (4-45) This normalised CGB outut torque is therefore a function of the CGB efficiency and the actual and equivalent values. The CGB outut seed can be normalised relative to the maximum seed of the flywheel, as discussed in Aendix A. This normalised seed is defined in Equation 4-46 and is only a function of the actual and equivalent values of and PGS efficiency, and the CGB efficiency. cgb, o 1 K o cgb, o 1, max J J fd fw (4-46) Once the CGB gear ratio, G, is secified the normalised torque and seed at the CGB inut can easily be found from; G T cgb, i T cgb o discharge, discharge and cgb cgb, o cgb, i (4-47) G The equations resented here for η eq and T cgb,o are for the case of flywheel discharge using a FDC clutch-controlled transmission, but the analysis can be reeated for oeration during flywheel charging where the direction of ower-flow in all branches is reversed. The full results (including charging and discharging cases) for eq and η eq are resented in Table 4-5, while the exressions for ω cgb,o and T cgb,o are resented in Table 4-6. These results show that the normalised actual and equivalent PGS gearing ratios and actual comonent efficiencies can be used to comletely characterise the transmission oeration and the normalised torque and seed requirements of the CGB. The effect of the actual and equivalent values on transmission efficiency, flywheel utilisation and normalised CGB ower rating can be investigated using the simle characteristic charge and discharge equations derived for brake-controlled transmission. As in Section 4.1, alying this analysis to a secific transmission configuration roduces a general design tool, which can be used to identify aroriate values of, G, K o and K for a articular alication. Before comlete transmission systems are investigated, the oeration other ower-slit configurations should 125

126 be considered. The derivation of the exressions for eq, η eq, ω cgb,o and T cgb,o for FWC and DDC clutch-controlled PSTs are resented below. b) FWC clutch-controlled PST The flywheel couled (FWC) transmission layout uses a fixed gear ratio and clutch-control to achieve a ower-slit between branches 1 and 2 of the differential, as resented in Figure Flywheel unit Differential gearing Final drive Fixed Gear atios K i K K = ω 4 / ω 2 K i = ω 6 / ω 1 G = ω 5 / ω G 4 Control gearbox (CGB) Figure 4-18 FWC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) By considering the PGS torque and seed relationshis it can be shown that the illustrated ower flow during discharge is only ossible when; (4-48) This configuration cannot therefore be used for the initial acceleration of the vehicle from stationary (when ω 2 /ω 1 = - /(1- )), and a hase of brake-controlled oeration is required to accelerate the vehicle and reduce the seed of branch 2 to zero. The limiting condition in Equation 4-48 is then satisfied and a hase of FWC clutch-controlled oeration is ossible, ending when the clutch locks at the following seed ratio; 126

127 1 2 lock 1 GK 1 Ki (4-49) The ratio between the torques exerted at the transmission inut and outut during flywheel discharge is defined as follows: T T fd fw discharge 1 1 Ki cgbgk eq eq (4-50) The torque ratio defined here can again be achieved by using a brake-controlled transmission, as described earlier. This equivalent differential must have a ratio, eq, and efficiency, η eq, such that, 3 1 eq lock discharge eq 1 1 cgb 1 1 eq cgb Ki GK Seed relationshi (4-51) Equivalent PGS efficiency (4-52) The torque and seed at the CGB can be analysed in the same way as for the FDC configuration, and the full results (including charging and discharging cases) for eq and η eq are resented in Table 4-5, while the exressions for ω cgb,o and T cgb,o are resented in Table 4-6. c) DDC clutch-controlled PST The final configuration considered here is the dual differential couled (DDC) clutchcontrolled PST. This is illustrated in Figure 4-19 where branch 1 of both differentials is connected to the flywheel and branch 3 to the final drive. Arrows again show the direction of ower-flow during flywheel discharge; the direction of the ower-flow between the two differentials now deends on the relative values of and q as illustrated. 127

128 Flywheel unit 1 3 Final drive 2 K Control gearbox (CGB) G 4 Fixed Gear atios K = ω 4 / ω 2 K q = ω 6 / ω q2 5 G = ω 5 / ω 4 6 K q q2 Direction of ower flow during flywheel discharge > q q1 q q3 q > Figure 4-19 Illustration of DDC clutch-controlled PST (arrows show direction of ower-flow during flywheel discharge) In order to achieve the ower flow indicated in Figure 4-19 for flywheel discharge, the following seed relationshi between branch 2 of the two differentials must aly; q 2 2 lock 1 1 q GK 1 1 q K q (4-53) and q 0 to ensure that the sign of3 1 is consistent for each differential; although additional gearing at branch 1 or 3 of one of the differentials could remove this constraint, the condition is required for a ractical coaxial configuration of the two differentials. If then ower can only flow out of branch 2 and into branch q2 during q discharge. The relationshi between the torques at these two branches is then given by; T q2 1 1 q K q cgbt GK (4-54) q 128

129 129 The seed relationshi at lock-u (Equation 4-53) leads to the following seed relationshi between the DDC clutch-controlled PST and an equivalent brake-controlled transmission; q q q q q eq lock GK K GK K Seed relationshi (4-55) The necessary equivalent efficiency of the brake-controlled transmission can be found, as shown in Equation q eq eq eq q cgb eq q eq eq cgb q q discharge eq (4-56) It is clear that the equivalent efficiency is again only a function of the CGB and PGS efficiencies and the normalised PGS ratios,, q and eq. This analysis can be reeated in order to find exressions for η eq during both charging and discharging when q and q, as shown in Table 4-5. An analysis of the normalised CGB torque and seed roduces the results shown in Table 4-6 for the same cases.

130 4. Page 130 PS mode PGS characteristics required to achieve equivalent erformance using brake-controlled transmission eq η eq (charging) η eq (discharging) FDC o GK K 1 1 cgb eq cgb 1 1 cgb eq cgb 1 FWC i GK K 1 1 cgb eq cgb 1 cgb eq cgb 1 1 DDC > q q q q q q GK K GK K q eq eq cgb q q eq eq eq q eq cgb q 1 q eq eq eq q cgb eq q eq eq cgb q DDC q > 1 q eq eq cgb q eq eq eq q cgb q eq q eq eq eq q eq cgb q q eq eq cgb q Table 4-4 Exressions for eq and η eq for FDC, FWC and DDC clutch-controlled ower-slit modes

131 4. Page 131 PS mode Definitions of normalised seed and torque at CGB outut ω cgb,o T cgb,o Exressions for T cgb,o fn(dod ov, η s) fn(dod ov, η s) Charging Discharging FDC fw fd cgb o o J J K max 1, 1 fd cgb o o T T K, 1 1 eq cgb eq eq cgb eq cgb FWC 1 max, 1 o cgb i K fw fd fd cgb o i J J T T K, eq cgb 1 eq cgb DDC > q fw fd cgb o q q J J K 1 max, 1 fd cgb o q q T T K, 1 1 eq q eq q cgb q eq q eq q cgb cgb DDC q > eq q eq cgb cgb 1 1 eq q eq cgb Table 4-5 Exressions for (ω cgb,o ) and (T cgb,o ) for FDC, FWC and DDC clutch-controlled ower-slit modes

132 PS mode Limiting values of the equivalent PGS characteristics eq η eq (charging) η eq (discharging) FDC eq cgb eq cgb eq FWC eq cgb eq cgb eq DDC > q q eq cgb q eq cgb q eq q DDC q > eq q cgb q eq q cgb q eq Table 4-6 The limiting values of eq and η eq for the equivalent PGS brake-controlled transmission There are two imortant oints to note about the results in Tables 4-4 and 4-5; If η gs and η cgb are assumed to be 100% then the value of η eq is 100% for all three ower-slit modes the oeration of the ower-slit modes is then identical to the oeration of a brake-controlled transmission with a PGS ratio of eq. If η gs and η cgb are both less than 100% then the gearing ratios in the transmission must be secified in order to calculate the eq and η eq values and hence assess the erformance of the clutch-controlled transmission In summary, these three different forms of clutch-controlled PST each allow a hase of flywheel discharge to be achieved which is exactly equivalent to a brake-controlled transmission using a PGS with characteristic gear ratio eq and efficiency η eq. A range of transmissions can therefore be imlemented using a conventional counter-shaft CGB with a number of discrete gear ratios to achieve multile hases of flywheel discharge. Using the resective equations for eq and η eq, aroriate values for the CGB gear ratios can be secified in order to achieve identical erformance (with maximum efficiency for a given DOD ov ) to the multi-pgs brake controlled transmission described earlier. The most effective configuration uses a discrete ratio gearbox to achieve clutch-controlled FDC and DDC ower-slit modes as discussed in the following section. 132

133 4.3. Multi-PGS clutch/brake controlled PST The suitability of the three different forms of clutch-controlled PST for use in flywheel transmissions can be considered by using the earlier results for a 15 hase brake-controlled transmission. These results rovide the values required to achieve maximum transmission efficiency for a given flywheel DOD ov. By using multile PGSs the ower in the control gearbox (CGB) can be reduced, but this requires that the gear ratios, G j, be consistent in all ower-slit modes. A transmission consisting of a combination of FDC and DDC owerslit modes is the only configuration able to achieve the maximum efficiency with the same CGB gear ratios in all modes. An examle of this configuration consisting of 3 PGSs and a 4-seed CGB is shown in Figure Vehicle PGS: (A) (B) (C) Sliing clutch Shaft 1 Flywheel unit C 1 C2 C3 C4 C o C B CGB,o Shaft 2 C A C C CGB,i Fixed Gear atios K o = ω 2 / ω 1 K A = ω cgb,i / ω ring,a K B = ω 2 / ω ring,b K C = ω cgb,i / ω ring,c G j = ω cgb,o / ω cgb,i Control gearbox NOTES For the PGS ratios, C > B > A C X refers to an on/off clutch to engage the associated gear (dotted). Fixed gear ratios are defined assuming relevant on/off clutch is engaged. Figure 4-20 Examle of a 3-PGS ower-slit transmission with 4-seed CGB During flywheel owered acceleration of the vehicle from stationary the articular transmission configuration shown in Figure 4-20 is oerated according to the sequence described in Table 4-7. The connecting elements are on/off clutches that engage and disengage the fixed ratio gearing. The control elements are the clutch and brake 133

134 Phase comonents which are used to generate torque at the transmission inut and outut via friction at a sliing interface leading to energy dissiation. Connecting elements CGB gear clutch Shaft clutch Control element PGS brake C 1 C 2 C 3 C 4 C o C A C B C C A B C Sliing clutch 1 X X X X 2 X X X X 3 X X X X 4 X X X X Descrition of oeration Power-slit between FD and PGS A 5 X X X Brake-controlled at PGS A 6 X X X X 7 X X X X 8 X X X X 9 X X X X Power-slit between PGS A and B 10 X X X Brake-controlled at PGS B 11 X X X X 12 X X X X 13 X X X X 14 X X X X Power-slit between PGS B and C 15 X X X Brake-controlled at PGS C Table 4-7 Control sequence for 3-PGS, 4-seed CGB clutch/brake controlled PST during flywheel discharge (X indicates the element is engaged) Once (ω ring, C ) 0 the transmission can no longer discharge the flywheel. It is clear that this tye of configuration allows a large number of transmission hases to be achieved, as defined by the following equation; N N N 1 (4-57) hases PGS CGB A clutch/brake controlled PST consisting of 3 PGSs and a 4-seed CGB therefore has the 15 oerating hases shown in Table 4-7. While all the values of a 15-PGS brake-controlled transmission can be otimised to achieve maximum efficiency (as shown earlier), the owerslit transmission is constrained by the requirement that the same discrete gear ratios in the CGB must be used in all the ower-slit modes. The effect of this constraint on transmission oeration is exlored in the following section CGB kinematic requirements The required values of the CGB gear ratio G j (where j = 1 to 4) during each of the three ower-slit modes of the transmission illustrated in Figure 4-20 can be found by considering 134

135 the aroriate equation for eq (see Table 4-4) alied to each of the four gears. As the values of A, B, C, K o, K A, K B and K C remain constant during transmission oeration in all CGB gears, a general exression for the ratio of the j th gear ratio, G j /G 1 can be derived for each ower-slit mode, as shown in Table 4-8. Power-slit mode 1 (FDC between final drive and PGS A) 2 (DDC between PGSs A and B) 3 (DDC between PGSs B and C) G j G 1 eq, j eq,1 A eq,1 eq, j A A eq,1 eq, j B eq,1 B A eq, j C eq,1 eq, j B eq,1 B C eq, j Table 4-8 Exressions for CGB gear ratios (where eq, j is the equivalent value of the oerating hase using the j th CGB gear in the resective ower-slit mode) The results of Table 4-8 show that the CGB gear ratios are a function of the actual and equivalent values that characterise the transmission oeration. The strength of the articular configuration shown in Figure 4-20 can be illustrated by calculating the values of the CGB gear ratios (relative to G 1 ) using the values that are found to achieve maximum efficiency as a function of DOD ov using a 15 hase brake-controlled transmission. If a constant efficiency, η x, is assumed for all the equivalent PGSs then the values of G j /G 1 are indeendent of the inertia ratio and PGS efficiency, and deend only on flywheel DOD ov. Table 4-9 shows the eq values required to achieve maximum discharge efficiency x for a DOD ov of 80%, and the corresonding values of G j /G 1 that are required for the three ower-slit modes. 135

136 PS mode Phase eq x G j /G A x B x C x - Table 4-9 equired sread of gear ratios in the CGB for DOD ov = 80% (assuming a constant actual or equivalent PGS efficiency of η x for all hases) brake-controlled hases are shown in bold These results show that the required G values are consistent between all three ower-slit modes for a DOD ov of 80%. Further analysis shows that this is true for any DOD ov when the eq values are chosen for maximum discharge efficiency. There is however a slight x variation of G j /G 1 with DOD ov as shown in Figure Figure 4-21 Influence of DOD ov on the CGB gearing ratios required for maximum efficiency 136

137 This variation in G j /G 1 is relatively small (the maximum difference is 3.8% for G 4 /G 1 over the range of DOD ov shown in Figure 4-21), and the overall ratio sread required for the CGB is aroximately equal to N 2 CGB (i.e. 16 for a 4-seed CGB). Secifying the maximum gear ratio, G 4, equal to 4 results in a value of G 1 ¼. The ratio of 4:1 is close to the maximum that can be achieved using a single stage of gearing, justifying the selection of a 4-seed CGB. This allows a comact counter shaft arrangement to be used, as illustrated in Figure This analysis shows that the 3-regime (FDC/DDC/DDC) transmission configuration illustrated in Figure 4-20 can achieve a similar maximum efficiency (as a function of DOD ov ) to the 15-PGS brake controlled transmission using only 3 PGSs and a 4-seed CGB. The kinematic viability of this articular transmission configuration has therefore been illustrated. Using the relationshis defined in Table 4-4 it can be shown any other multile-regime configuration of these comonents consisting of FDC, FWC or DDC ower-slit modes has a lower maximum efficiency due to the constraints of the CGB fixed gear ratios Performance of CGB-controlled PST The erformance of the CGB-controlled PST can be investigated in detail using the following rocedure. 1. Secify value of A, η gs (for A, B and C) and η cgb 2. Estimate values of B, and C 3. Estimate values of eq, j (j = 1, 2, 3, 4) for ower-slit mode 1 and the two values of eq, 1 obtained using CGB gear G 1 in ower-slit modes 2 and 3 4. Use Table 4-8 to find required values of G j /G 1 to achieve the mode 1 eq values 5. Use Table 4-8 to find eq, j (j = 2, 3, 4) for ower-slit modes 2 and 3 6. Use Table 4-5 with actual and equivalent values to calculate all η eq values 7. Use characteristic discharge equations (Equation 4-19) to calculate transmission discharge efficiency, max(t cgb,o ) discharge and max(ω cgb,o ) discharge 8. Iterate stes 2-7 to identify values of B, C, ( eq, 1 ) for modes 1, 2 and 3, and G j /G 1 (j = 2, 3, 4) that result in maximum discharge efficiency for the secified value of A 9. Use characteristic charging equations (Equation 4-20) to calculate transmission charge and round-tri efficiencies, max(t cgb,o ) charge and max(ω cgb,o ) charge 137

138 10. ecord all values of actual and equivalent, DOD ov, η discharge, η round-tri, max(t cgb,o ), max(ω cgb,o ) 11. eeat stes 2-10 for a range of A values This calculation rocedure allows the oeration of the transmission to be comletely characterised as a function of DOD ov with the gear ratios chosen for maximum discharge efficiency. A constant comonent efficiency of 95% has been assumed for all four gears of the CGB and the three PGSs (as suggested by White [133] for tyical oeration in vehicle transmission alications ). The resulting values of eq dischar ge for a range of DOD ov are shown in Figure While the efficiency of all three actual PGSs in the 3-regime PST is 95%, during ower-slit oeration the efficiency of the equivalent brake-controlled PGSs are u to 4% lower, illustrating the effect of the additional losses that occur in the CGB. The CGB losses during DDC ower-slit oeration are seen to increase with DOD ov due to the higher roortion of ower flowing through the CGB. DOD ov Transmission hase Figure 4-22 The values of η eq for a 3-PGS, 4-seed CGB controlled transmission during flywheel discharge with gear ratios chosen for maximum efficiency (η A, η B, η C and η cgb are equal to 95%) While results can also be obtained for eq char ge using the same values of, eq and PGS/CGB efficiencies, these are found to be extremely close to the discharge values in each hase, with a maximum difference of 0.1% over the range of DOD ov considered. The values of, η gs, eq and η eq are used with the characteristic equations derived for the brakecontrolled transmission (Equations 4-16 and 4-17) to calculate the overall charge and 138

139 Efficiency discharge efficiencies of the CGB-controlled PST as a function of DOD ov. These efficiencies are shown in Figure 4-23 for the case when PGS and CGB gear ratios are chosen to achieve the maximum ossible discharge efficiency as a function of DOD ov. DOD ov Figure 4-23 Overall charge, discharge and round-tri efficiencies for the 3-PGS, 4-seed CGB-controlled transmission (with η gs s and η cgb = 95%) Figure 4-24 shows the maximum discharge efficiency of the 3-PGS, 4-seed CGB controlled transmission comared with the 15-PGS brake-controlled transmission. Over the range of DOD ov considered, the discharge efficiencies are very similar with the additional losses in the ower-slit branches of the CGB-controlled transmission reducing efficiency by around 1%. Brake-controlled Max (ηdischarge) CGB-controlled 100% x 95% x gs ' s and cgb 95% DOD ov Figure 4-24 Comarison of maximum ossible discharge efficiency as a function of DOD ov between 15-hase brake-controlled and 3-PGS, 4-seed CGB-controlled transmissions 139

140 While the CGB-controlled PST has been shown to achieve similar (although slightly lower) overall efficiency as the brake-controlled system, its main advantage is that the total mass of the transmission can be substantially lower due to the reduced amount of gearing. In order to quantify this, the torque and seed requirements of the CGB must be characterised CGB torque and seed requirements In order to estimate the mass of the CGB comonent, the maximum torque and seed requirements must be identified. This can be achieved by alying the the exressions for normalised CGB torque and seed defined for the different ower-slit modes in Table 4-5. For the three ower-slit modes the following exressions aly for the normalised CGB torque aly; Power-slit mode 1 (FDC between final drive and PGS A) T cgb, o K o T cgb, o T fd 2 K B cgb, o (DDC between PGSs A and B) 1 B T fd 3 (DDC between PGSs B and C) 1 B T fd T T K B cgb, o Table 4-10 Normalised torque exressions for 3-regime (FDC/DDC/DDC) CGB-controlled PST For CGB-controlled oeration in gear G j the following exressions for the corresonding eq,j in ower-slit modes 1 and 2 aly; eq A, j (4-58) mode1 Ko 1 1 A G jk A 1 A KB A B 1 B G jk A eq, j (4-59) mode 2 1 A KB 1 1 B G jk A 140

141 Combining these equations results in an exression for K o ; A eq, j mode 2 eq, j mode 2 B K A eq, j B mode 1 K (4-60) o 1 B eq, j mode 1 A similar exression can be derived for ower-slit mode 3, where; K o K C 1 C fn ' s, eq ' s This allows the normalised CGB torques and seeds in all three ower-slit modes to be exressed relative to K o. These are still only functions of the PGS and CGB efficiencies and the actual and equivalent values (which are themselves only a function of DOD ov for maximum η discharge ). Figure 4-25 shows the maximum values of T cgb,o and ω cgb,o that occur during charge or discharge as a function of DOD ov when η A, η B, η C and η cgb are all 95%. The roduct of the maximum normalised torque and seed is also shown. This is effectively a maximum normalised CGB ower rating (although the maximum torque and seed do not occur simultaneously during oeration), and is defined as follows; P cgb, rated max T max cgb cgb fd (4-61) T fd fw,max J J fw T max K o T cgb fd 1 max Ko fw cgb,max J J fd fw max P cgb, rated DOD ov Figure 4-25 Maximum values of normalised seed, torque and rated ower at CGB The value of P cgb,rated as a function of DOD ov is very useful, as it allows the required torque caacity of the CGB to be identified once the other variables have been secified for a articular alication. The lot of normalised CGB torque can then be used to find the required value of K o. 141

142 The data in Figure 4-25 has been calculated for the case when the gear ratios (K A /K o, K B /K o, K C /K o, A, B, C and G 1 -G 4 ) have been chosen to achieve maximum discharge efficiency for a given DOD ov. There is however a comromise between choosing gear ratios to maximise efficiency, or to minimise the rated ower of the CGB. The extreme case of minimising CGB rated ower occurs when the transmission gear ratios are chosen so that the ower flow in each ower-slit hase falls to zero, and the transmission becomes a simle 3- PGS brake-controlled configuration. This comromise between transmission efficiency and CGB ower-rating is investigated in the following section Strategy for limiting CGB mass The maximum torque requirement of the CGB can be reduced by adjusting the gearing ratios K A /K o, K B /K o and K C /K o along with the PGS ratios B / A and C / A. While this will reduce the efficiency of the transmission for a given DOD ov it has the advantage of reducing the required mass of the CGB, and can therefore increase the secific energy caacity of the FESS. There is obviously a comromise between reducing the mass and the efficiency of the system; the flywheel utilisation factor rovides a basis for comaring the system erformance and the CGB requirements. A simle aroach to identifying aroriate gearing ratios and the associated system erformance can be imlemented by equalising the maximum CGB torque and seed achieved in each ower-slit modes, as follows; 1. Choose a value of (1- A ) K o /K A in order to achieve equal maximum T cgb at inut and outut during mode 1 oeration 2. Choose values of (1- B ) K o /K B and (1- C ) K o /K C to achieve the same maximum T cgb at inut or outut during mode 2 and mode 3 oeration 3. Perform iteration to find values of B and C that result in minimum P cgb for a given value of A By considering the exressions for normalised CGB torque during mode 1 oeration, it can be seen that max(t cgb,o ) mode 1 = max(t cgb,i ) mode 1 when; G 1 K G o 4 1cgbG 1 A, which equals 1 when G 1 = 1/G 4. (4-62) K A 4 cgb If the CGB and PGS efficiencies are assumed to be 100% the maximum inut and outut CGB torques in modes 2 and 3 are exactly the same as in mode 1 when; 142

143 o o 1 1 and 1 1 B K K B K C (4-63) K C These values are therefore used to achieve an aroximately equal maximum torque in each ower-slit mode when realistic CGB and PGS efficiencies are used. The iterative rocedure has been imlemented using the data secified in Table The resulting values for the maximum torque (which is now indeendent of DOD ov ) and P cgb, rated using this equalised modes aroach are shown in Figure 4-26 as a function of the flywheel utilisation, alongside the values of P cgb, rated obtained when the gearing ratios are selected to maximise discharge efficiency. Parameter G 1 G 2 G 3 G 4 η gs s η cgb Value 1/4 2/3 3/ Table 4-11 Parameters of 3-regime CGB controlled transmission used in equalised modes analysis P cgb, rated (max η discharge ) P cgb, rated (equalised modes) max(k o T cgb /T fd ) (equalised modes) Figure 4-26 Comarison of normalised rated variator ower as a function of U, calculated to achieve maximum discharge efficiency or equal max CGB torque (defined as the equalised modes case) in each ower-slit mode The normalised ower and torque data in Figure 4-26 shows that the equalised modes method of gearing selection achieves the same flywheel utilisation with a lower maximum CGB ower than the maximum η discharge method. The slight reduction in transmission efficiency does however mean that a slightly lower maximum U is ossible. The three normalised PGS gear ratios required for this method are shown in Figure 4-27(a), with the discharge and round-tri efficiencies shown in Figure 4-27(b). 143

144 C B A discharge roundtri (a) (b) Figure 4-27 Values of normalised PGS ratio required to achieve equalised maximum CGB torques in all modes, and the resulting transmission efficiency Figures 4-26 and 4-27 (together with the equalised modes conditions described in Equations 4-62 and 4-63) rovide the information required to imlement a design tool for the 3-PGS CGB-controlled flywheel transmission, allowing a direct comarison with the brakecontrolled transmissions described in Section egional rail vehicle alication of CGB-controlled PST The design tool data for the 3-regime CGB-controlled flywheel transmission can be used to find aroriate system arameters for the secific alication of a regional rail vehicle, as characterised by the data in Table Parameter Value Unit J fd 3670 kgm 2 T fd max(ω cgb) Single PGS secific torque CGB secific torque Flywheel secific energy 10 knm 300 rad/s 50 Nm/kg 5 Nm/kg 0.05 MJ/kg Table 4-12 Values used in analysis of 3-regime CGB-controlled flywheel transmission for a regional rail vehicle The flywheel and transmission arameters are chosen in order to achieve maximum secific energy caacity for a given useful energy caacity, E useful (the energy that can be delivered to 144

145 the vehicle from a fully charged flywheel). This is found by calculating the mass of the flywheel, PGSs and CGB as a function of DOD ov in order to achieve a given value of E useful. The mass of the PGS brake comonents are neglected as the energy dissiated is small. An examle of the comonent masses as a function of DOD ov is given in Figure 4-28 for the case of E useful = 30 MJ. Figure 4-28 Mass of system comonents as a function of DOD ov using a 3-PGS, 4-seed CGB transmission roviding a useful energy caacity of 30 MJ These results are indeendent of the secified maximum flywheel seed (which does however affect the required gear ratios), and deend only on the required torque at the transmission outut, the useful energy caacity and the inertias of the vehicle and flywheel. Increasing the secified maximum CGB seed also affects the CGB mass, reducing the maximum CGB torque that occurs with the same ower flow. For E useful = 30 MJ the minimum total mass (and therefore maximum secific energy caacity) is seen to occur for a DOD ov of around 80%. Figure 4-29 shows the maximum secific energy caacity as a function of useful energy caacity using the 3-PGSs CGB-controlled transmission. 145

146 Euseful / FESS mass E useful (MJ) Figure 4-29 Maximum secific energy of FESS as a function of available energy using a brake/clutch controlled transmission with 4-seed CGB and 3 PGSs, and a 4-PGS ringbrake transmission (E avail = 40 MJ, η gs s and η cgb = 100%) This result shows that when the braking energy is limited to 40 MJ, the maximum secific energy of the 3-regime CGB-controlled transmission is around 17 MJ/tonne which occurs when the useful energy caacity of the system is equal to 32.5 MJ. This is significantly higher than the efficiency and secific energy than can be achieved by a ractical brakecontrolled transmission, illustrating the advantages of the CGB controlled ower-slit. However, if a round-tri efficiency of less than 50% is accetable then the mechanically simler 4-PGS brake-controlled transmission can rovide the same useful energy storage caacity and system mass as the CGB-controlled transmission, and is therefore an attractive otion Summary An analysis method for mechanical flywheel systems has been described, which allows the erformance of secific transmission configurations to be characterised. Simle brakecontrolled transmissions and clutch-controlled PSTs have been investigated using this method, which is indeendent of the alication (i.e. the vehicle tye and drive cycle) and allows a thorough investigation of the interaction between the flywheel, vehicle and transmission. A flywheel utilisation factor has been defined and shown to be a useful measure of the comromise between transmission efficiency and flywheel deth-ofdischarge. Furthermore, in ower-slit transmissions the flywheel utilisation factor also rovides a basis for investigating the comromise between transmission efficiency and 146

147 comonent ower rating. The results of alying the indeendent analysis method (IAM) to a articular transmission configuration roduce a design tool consisting of normalised system arameters as a function of the overal flywheel deth-of-discharge, DOD ov, achieved during oeration. These design tools are extremely useful as they allow an absolute basis for the comarison of different transmission configurations, and also rovide a simle means of secifying aroriate transmission arameters for a articular alication. This has been demonstrated for the case of a diesel regional train, with flywheel energy storage system (FESS) arameters chosen in order to maximise the secific energy caacity of the system. It has been shown that mechanical flywheel transmissions consisting of brake and clutch controlled PGSs can achieve high efficiency with relatively low mass, leading to high secific energy caacity for the FESS. A novel transmission configuration using a control gearbox (CGB) with sliling clutch has been investigated. This CGB-controlled FESS consists of multile ower-slit modes and has been shown to erform articularly well for flywheel alications. An advantage of these mechanical transmissions is the fact that they consist of standard comonents that are currently widely used in conventional automatic transmissions for IC engines. The cost of develoing and manufacturing transmissions for flywheel alications is therefore exected to be relatively low. Transmission systems using variator comonents (a more secialist technology) are more commonly roosed for flywheel alications. A range of variator-controlled transmissions are therefore investigated using the IAM in the following Chater. 147

148 5. Indeendent analysis of variator-controlled transmissions The revious Chater illustrated an indeendent analysis technique for brake and clutch controlled transmissions. The value of this method was demonstrated by roducing results in the form of normalised design tools that allow a comarison of the erformance of different transmission configurations and the identification of aroriate gearing ratios for any alication (as characterised by the inertial loads at the transmission inut and outut). The aim of this Chater is to imlement a similar technique to investigate a range of variatorcontrolled transmissions. The limited range of gear ratios that can be achieved using a variator means that clutch and brake controlled transmission hases are useful for increasing the overall transmission seed ratio range. In the case of flywheel systems for rail vehicles, the resulting decrease in transmission efficiency is accetable as the increased ratio range can eliminate the need for low seed (and therefore inefficient) oeration of the conventional ower-train. The analysis of these clutch and brake controlled oerating modes is identical to that discussed in Chater 4. When analysing the variator-controlled oerating hases, the secific configuration of the transmission is an imortant factor. Three configuration categories are discussed in this Chater. These are: i. A direct variator connection between the flywheel unit and final drive ii. Single regime ower-slit transmissions iii. Multile regime ower-slit transmissions In the context of these transmissions, a regime of variator-controlled oeration relates to a ower-slit transmission hase in which the variator seed ratio varies between the maximum and minimum values. Multile regime transmissions can be achieved by allowing a range of different ower-slit configurations to be imlemented in a single transmission. This increases overall efficiency and/or decreases the required rated variator ower, but it can only be achieved with increased system comlexity. The indeendent analysis method (IAM) reviously alied in Chater 4 is again used to generate a design tool for each of the transmission configurations considered here. This method defines the transmission erformance and allows aroriate design variables to be secified for any alication. An imortant contribution of this research is that it rovides an absolute basis for assessing the strengths and weaknesses of different configurations. Finally, the last section of this Chater illustrates the use of these design tools to identify the most aroriate systems for regional rail vehicle alications. 148

149 5.1. Direct variator-controlled transmission A simle single-ath transmission can be achieved by using a variator to rovide a direct connection between the flywheel and vehicle. The inut of the variator is connected to the flywheel unit via a fixed gear ratio, K i, and the outut to the final drive via a gear ratio, K o, and a clutch as shown in Figure 5-1. As the variator has limited maximum and minimum gear ratios the clutch is required in order to transmit ower during both low seed acceleration and high seed deceleration of the vehicle. The analysis of this tye of flywheel transmission is imortant as it rovides a base case for quantifying the benefits of various ower-slit transmissions and identifying the most aroriate configuration for a given alication. Flywheel unit Clutch Final drive Fixed Gear atios Shaft 1 φ = ω var, o / ω var, i K i = ω var, i / ω fw K o = ω var, o / ω 1 Variator: Inut Outut Figure 5-1 Configuration for a direct (toroidal tye) variator-controlled transmission An analysis has been erformed assuming an overall variator ratio sread, φ t, of 6.25 (a tyical value for toroidal and ush-belt tye variators), and with the following oerating limits; Variator seed ratio limits; max 2. 5 and min 0. 4 A constant variator efficiency, η var, of 85% has been alied throughout the analysis resented in this Chater. The oeration of the direct variator transmission during flywheel discharge consists of the following hases; 149

150 Direct variator oerating hases during flywheel discharge For 0 (K fw ω fd / K fd ω fw ) < 0.4 In this hase of oeration the variator is held at the minimum seed ratio (φ min ) and the transmission is clutch-controlled. The hase ends when the sli seed in the clutch falls to zero. Alying the IAM to this hase results in the following exression relating the initial and final flywheel seeds for flywheel discharge; 2 fw, i mink fw, f clutch 1 Where K is a normalised gear ratio defined as; K K o J fw var (5-1) i fd (5-2) K J For 0.4 (K fw ω fd / K fd ω fw ) 2.5 In this hase the clutch is locked and the variator is controlled to rovide the required torque at the vehicle. The hase ends when the variator reaches its maximum seed ratio. Alying the IAM along with conservation of energy results in the following exression;, fw, fw i f variator var var max min 2 K 2 K Otherwise The transmission cannot be used to discharge flywheel (5-3) The results of Equations 5-1 and 5-3 allow the overall erformance of the direct variator transmission to be assessed in terms of the normalised gear ratio, K, the maximum and minimum variator seed ratios and the variator efficiency. As the final flywheel seed in the clutch-controlled hase equals the initial seed in the variator-controlled hase, the ratio between the maximum and minimum flywheel seeds during discharge can be exressed as; fw,min fw,max var (5-4) 2 2 var maxk var mink Equation 5-4 leads to the following exressions for the overall DOD and discharge efficiency of the transmission; 150

151 var 2 K K var max 2 DOD ov 1 (5-5) var min 2 var maxk 2 K K 2 2 discha rge (5-6) var max var min var It should be noted that the efficiency and DOD ov achieved by the transmission deend on the ratio of the fixed gears, K i /K o, rather than the actual values. The actual values of K i and K o are however imortant in ensuring an aroriate range of oerating torque and seed for the variator. This is aarent in the normalised exressions which can be derived for the maximum variator torque and seed as shown in Equations 5-7 and 5-8. As for the CGB in Chater 5, a normalised rated ower can again be defined as shown in Equation 5-9; T K T var var direct o (5-7) Tfd 2 P var var, rated direct 1 K i var fw,max T max max J var var fd max var direct var T J fd fw,max fw T max K direct (5-8) (5-9) When a constant torque is alied to the final drive, the maximum variator torque will occur at the variator inut during flywheel discharge; max max T var (5-10) direct var If max 1 DODov 1 then the maximum variator seed occurs at the inut and has the value; 1 max var direct (5-11) Otherwise, the maximum variator seed occurs at the outut and has the value; max var max 1 DODov direct (5-12) These equations allow the direct variator transmission to be comletely characterised as a function of the normalised gear ratio, K. Figure 5-2 shows the transmission discharge erformance and the requirements of the variator, assuming constant variator efficiency, η var, of 85%. 151

152 max(t var ) max(ω var ) P var,rated DOD ov η discharge U Figure 5-2 Performance characteristics of the direct variator flywheel transmission as functions of normalised gear ratio, K (η var = 85%) The results show that the maximum U is achieved when K It is also clear that the referred oerating region for the transmission is in the range 0 < K < 0.95, in order to achieve high efficiency and low rated variator ower for a given U. There is no obvious advantage in oerating with K > 0.95, and so this region should be avoided. The roblem with this direct variator-controlled transmission is that the variator is transmitting the full tractive/braking ower between the vehicle and flywheel. This means that large and exensive variators are required to achieve full braking energy recovery (although a down-sized system could oerate in conjunction with conventional braking for limited regeneration). Also, the transmission oerates with relatively low efficiency due to the high ower flow through the variator and the additional losses during clutch-controlled oeration. Alternative systems which can achieve similar or imroved efficiency with smaller variators are therefore attractive otions for flywheel transmissions. Power-slit transmissions can otentially achieve this, and are considered in the following sections Single regime variator-controlled ower-slit transmissions An alternative to the direct variator transmission described above is a variator-controlled ower-slit arrangement. The limited ratio sread of mechanical variators means that variator-controlled PSTs are required to oerate in clutch and brake controlled hases when outside the oerating range of the variator. The analysis of transmission oeration during 152

153 these hases is therefore the same as that described in Chater 4. As with the clutchcontrolled PST, a number of configurations of variator-controlled transmission can be achieved by connecting the variator between different branches of a PGS. Two ossible configurations which avoid ower-recirculation and consist of a single hase of variatorcontrolled oeration are considered in this section; a) FDC variator-controlled PST b) FWC variator-controlled PST Before alying the IAM to these ower-slit configurations, it is useful to first consider the characteristics of the variator-controlled hase using exressions derived for the case with no losses at the PGS or variator. The analysis of these ideal variator-controlled PSTs is resented below Ideal variator-controlled ower-slit hase In the literature [93-94] the simlest form of variator-controlled ower-slit transmissions have been considered and characterised for a single direction of ower-flow as either inut couled (IC) or outut couled (OC) as defined in Figure 5-3. Simle relationshis can be derived for each case relating the overall transmission seed ratio (r) and the ideal ower ratio (P variator /P inut, assuming no PGS or variator losses) to the PGS ratio () and the variator seed ratio (φ) as shown in Table 5-1. Inut Couled Outut Couled V V Inut Outut Inut Outut Figure 5-3 Definition of inut and outut couled configurations with arrows showing direction of ower-flow with no ower recirculation 153

154 Power-slit configuration Inut couled Outut couled Overall transmission ratio r 1 Power-slit ratio P v P i 1 Condition for no recirculation 0 1 r r r 1 Pv P i 1 r r 0 1 Condition for Pv Pi max max r r r rmin Table 5-1 Characteristic relationshis for simle inut and outut couled PSTs [94] If the overall transmission seed range, r t, and variator seed range, φ t, are defined as; r r t r max min and max t (5-13) min both r t and the maximum value of P v / P i can be exressed as a function of /r min for the two cases, as shown in Figure 5-4 using a value of φ t = It is imortant to note that these results only cover the range of values of /r min in which no ower-recirculation occurs. Inut couled Outut couled r t P v / P i r t P v / P i Figure 5-4 Characteristics of simle inut and outut couled PSTs oerating with no ower recirculation (φ t = 6.25) These Figures show that the value of r t varies between the case of a direct variator transmitting 100% of the ower (r t = φ t ) and the case of a fixed gear ratio where no ower flows through the variator branch (r t = 1). Either the IC or OC configuration can achieve any 154

155 secified combination of r t (between 1 and φ t ) and r min resulting in the same maximum value of P v / P i (although the required value of is different). The configurations are therefore kinematically equivalent, as either can be used to achieve the same overall transmission characteristics. The outut couled configuration is however likely to be more aroriate for achieving flywheel discharge in a FESS. This is because if a constant torque is required at the transmission outut, the values of P o and P i will increase with vehicle seed and transmission seed ratio, r. In this situation the maximum actual variator ower, P v, can be minimised by using a system where the maximum value of P v / P i occurs at r min (i.e. low vehicle seed and inut ower) rather than at r max where vehicle seed and inut ower are higher. By considering the reverse case of flywheel charging (where the flywheel connection is the outut and the vehicle connection is the inut), it is clear that the maximum value of P v / P i should now occur at r max (i.e. high flywheel seed and low vehicle seed), which is achieved using an inut-couled arrangement. The transmission configuration can therefore remain the same, while the direction of ower-flow reverses, changing it from outut-couled during owering to inut-couled during regenerative braking. To ensure clarity, the term final-drive couled (FDC) is again used to define this articular configuration without needing to secify the direction of ower-flow. The results of this simle analysis highlight a number of imortant factors; i. Power-slit oeration (without ower recirculation) can only be achieved with a reduction in the transmission seed range relative to the variator itself (i.e. r t is always less than φ t ). Comared to a direct variator transmission delivering the same amount of energy to the vehicle, the reduced ratio range means that a lower roortion of the flywheel energy is transmitted during the variator-controlled hase. This is likely to result in a lower overall efficiency due to losses during the brake and clutch controlled hases. ii. iii. IC and OC configurations are kinematically equivalent as they can be sized to achieve the same values of r min and r max by choosing an aroriate for a given value of φ t. In terms of variator ower, the FDC configuration is likely to be more suited to flywheel transmission alications as the maximum value of P v / P i occurs at the minimum value of P i (assuming constant outut torque). 155

156 This analysis is limited as it considers only the variator controlled hase of oeration, and rovides no way of identifying a suitable comromise between transmission erformance, variator ower caacity and flywheel energy caacity. A thorough investigation is therefore required in order to quantify the effects of comonent efficiencies and gearing ratios on the overall erformance of single regime variator-controlled flywheel transmissions. This is erformed for both FDC and FWC transmissions using the indeendent analysis aroach introduced in Chater 4, as described in the following section Indeendent analysis of single regime variator-controlled PSTs The IAM can be alied to the FDC and FWC variator-controlled PST configurations, with the results again roviding a general FESS design tool. While the brake and clutch controlled hases of these transmissions can be analysed using the method described in Chater 4, a different aroach is required during the oerating hase when the gear ratio of the variator is continuously changing. This is due to the fact that while the total energy dissiated in the variator during this hase is indeendent of time, it is not ossible to derive an exlicit exression for this. It is imortant however, that these variator losses are considered in the analysis as variator efficiencies can be significantly lower than those for fixed gearing, and their effect on the overall transmission oeration should therefore be quantified. This can be done by creating a quasi-static time-ste based model of each variator-controlled PST. A constant efficiency has been assumed for the variator comonent, allowing the losses to be quantified by numerically integrating the ower-flow through the variator. The indeendent analysis method again assumes that there are no losses at the vehicle or flywheel, as justified for brake and clutch controlled transmission in Chater 4. Transmission oeration during flywheel discharge is modelled by identifying the limiting overall transmission seed ratios for oeration in brake, clutch and variator controlled hases. The initial conditions for the analysis are always a vehicle seed of zero, and a fully charged flywheel (i.e. SOC = 100%). A time-ste based rogram has been develoed using the Matlab/Simulink software, and consists of searate modules describing the oeration of each hase (clutch, variator and brake controlled). These modules calculate the aroriate torque being exerted on the vehicle and flywheel during each time-ste, and therefore the acceleration and seed of the vehicle and flywheel. Once the condition is reached where no 156

157 further discharge of the flywheel is ossible, the rogram terminates and the overall discharge erformance of the transmission is assessed. The erformance of the FESS during flywheel charging is assessed by identifying the limiting transmission seed ratios for oeration in each hase. While the oerating limits for the variator-controlled hase remain the same as during flywheel discharge, the reversal of ower flow in the transmission changes the sequence in which clutch, variator and brake controlled oeration is ossible. Using these new oerating limits, a backwards calculation can be erformed using exactly the same modules as for the discharge case to describe each hase of oeration, with the same initial conditions of zero vehicle seed and SOC = 100%, and a final condition of DOD ov = (DOD ov ) discharge. This requires that the PGS and variator efficiencies used during the discharge calculation are inverted, due to the reversal of the direction of ower-flow in all branches of the transmission. This method therefore identifies the minimum initial braking seed required to fully recharge the flywheel, ensuring that the overall charging erformance of the transmission is assessed. This analysis rovides a well defined sectrum of oeration for the system, and allows the erformance and requirements to be characterised. The results therefore rovide a rigorous basis from which a range of other factors can be investigated, including; i. egenerative braking and flywheel owered acceleration outside this range of oeration (for examle at higher vehicle seeds) ii. The effect of losses at the vehicle and flywheel (which will increase the minimum vehicle seed required for full recharge of the flywheel) These issues are considered in more detail in Chaters 6 and 7. The calculation rocedure for the indeendent analysis of ower-slit transmission is therefore as follows; i. Secify values of η gs and η var ii. Secify values of and eq (which can be defined for FDC and FWC configurations as described in Chater 4) iii. un time-ste based comutational model of a articular transmission configuration for flywheel owered acceleration to obtain DOD ov, η discharge, max(t var ) discharge and iv. max(ω var ) discharge un time-ste based comutational model of transmission configuration for flywheel charging to obtain η charge, max(t var ) charge and max(ω var ) charge 157

158 v. Assess η round-tri and P var, rated vi. eeat stes (i) to (v) for a range of and eq values to roduce contour mas of the key erformance arameters This rocedure is used to erform an indeendent analysis of both FDC and FWC variatorcontrolled PSTs as described in the following section. a) FDC variator-controlled PST The first PST considered is the FDC configuration, where a variator, V, is connected between branches 2 and 3 of a PGS is shown below. Vehicle Flywheel unit Differential gearing Final drive Fixed Gear atios K K o K = ω 4 / ω 2 K o = ω 6 / ω 3 φ = ω 5 / ω 4 V Figure 5-5 Schematic diagram of FDC variator-controlled PST configuration (arrows show direction of ower-flow during clutch/variator controlled flywheel discharge) Transmission oerating modes during flywheel discharge i. Clutch-controlled oeration hase This occurs when the seed of shaft 6 is lower than the seed of shaft 5 using the minimum variator ratio; min K K o (5-14) The outut torque from the transmission is controlled by the torque at the clutch. The relationshi between transmission inut and outut torque is given by; 158

159 T T fd fw discharge 1 vark o 1 mink (5-15) ii. Variator-controlled oeration hase This occurs when the clutch is locked and the required seed ratio across the variator lies within the oerating limits; min max K where K o 1 1 The outut torque from the transmission is then equal to; 3 3 (5-16) T T fd fw discharge vark o 1 1 (5-17) K iii. Brake-controlled oeration hase Once the variator ratio required to connect the carrier and ring branches reaches the maximum limit (φ = φ max ) ower can no longer be transmitted; max K K o (5-18) The clutch is then disengaged and braking is alied at branch 2 of the PGS in order to continue discharging the flywheel. The outut torque is therefore; T T fd fw discharge (5-19) FDC esults As with the analysis of the brake/clutch controlled transmission described in Chater 4, the transmission efficiency and flywheel DOD ov of the FDC variator-controlled PST are found to be functions of the normalised PGS ratio,, and a normalised equivalent PGS ratio, eq, relating to the final condition of the clutch controlled ower-slit hase between branches 2 and 3. The value of eq is therefore a function of, K, K o and φ min as defined in the revious chater and restated in Equation 5-20; eq 1 K o 1 mink (5-20) 159

160 It should also be noted that eq reresents the minimum overall transmission ratio that can be achieved during variator-controlled oeration, and therefore corresonds to the variable r min used in Section The term eq is used in the subsequent analysis in order to maintain consistency with the analysis of clutch-controlled PSTs resented in Chater 4. The normalised variator torque, seed and rated ower are exressed in the same form as for the FDC clutch-controlled transmissions; T T var var K (5-21) FDC o T fd var FDC 1 K o var fw, max J J fd fw (5-22) P var, rated T max max J var var fd max var var T J fd fw,max fw T max (5-23) By lotting discharge and round-tri efficiencies, DOD ov, U and P var, rated as functions of and eq a set of indeendent transmission erformance mas can be roduced for secified efficiencies of the PGS and variator. It is clear from Equation 5-20 that eq must always be smaller than in order to revent ower recirculation. The erformance mas are therefore only valid for the region in which > eq. These results again form the basis of a design tool to identify suitable transmission arameters for any alication of the roosed flywheel transmission, and are shown below for the FDC variator-controlled transmission with η var = 0.85, η gs =

161 eq eq eq a) Discharge efficiency b) DOD ov c) ound-tri efficiency (for full charge/discharge cycle)

162 eq eq 1 d) Flywheel utilisation (U) e) Normalised rated variator ower (P var, rated ) Figure 5-6 Contour mas for a FDC variator-controlled PST showing (a) η discharge, (b) DOD ov, (c) η round-tri, (d) U and (e) P var, rated as functions of and eq (with η var = 0.85, η gs = 0.95) dotted line shows min(p var, rated ) w.r.t. U Figures 5-6 (d) and (e) shows that values of and eq can be chosen in order to achieve a articular flywheel utilisation with minimum normalised variator ower. These values are indicated by the heavy dotted line in all the erformance mas of Figure 5-6. Furthermore, it is aarent for Figure 5-6 (a) that these values coincide with the maximum ossible values of transmission efficiency for a given U, which occurs when the losses in the brake and clutch controlled hases are equal as shown in Figure

163 eq CLUTCH DOMINATED C B A BAKE DOMINATED A B C Figure 5-7 Illustration of the effect of eq values on transmission losses (shown as ercentage of transmission inut energy dissiated in comonents) A conclusion of the indeendent analysis is therefore that the FDC variator-controlled flywheel transmission should always be sized to oerate at a oint on this minimised ower curve. It is imortant to note that while the normalised rated variator ower is almost identical during full charge and full discharge events, max(t var ) FDC occurs during charging, and max(ω var ) FDC occurs during discharging. The overall rated variator ower must therefore be calculated using these maximum values. This is encasulated in the following two Figures which form the design tool for a FDC variator-controlled flywheel transmission and rovide all the information required to characterise the gearing ratios, variator size and transmission erformance for a articular alication. 163

164 2 1.5 η discharge η round-tri max(t var ) P var, rated DOD ov Figure 5-8 Characteristics of FDC variator-controlled PST when P var, rated is minimised for a given U 10 1 eq DOD ov Figure 5-9 Values of and eq which minimise P var, rated for a given U Figures 5-8 and 5-9 rovide all the information necessary to define the erformance and design requirements of the FDC variator-controlled PST. b) FWC variator-controlled PST 164

165 An alternative configuration for a single regime PST is to connect the variator between branches 1 and 2 of the PGS, with associated fixed ratio gearing. This forms a flywheelcouled (FWC) variator-controlled PST as illustrated in Figure Vehicle Flywheel unit Final drive Fixed Gear atios K i K K = ω 4 / ω 2 K i = ω 6 / ω 1 φ = ω 5 / ω V 4 Figure 5-10 FWC variator-controlled PST (arrows show direction of ower-flow during flywheel discharge) By considering the torque and seed relationshis for the PGS, it can be shown that the ower-slit oeration illustrated in Figure 5-10 is only ossible when; (5-24) As this ratio is always negative for ω 3 = 0 (i.e. when the vehicle is stationary), an initial hase of brake-controlled oeration (with braking alied at branch 2) is required to accelerate the vehicle from stationary. Once ω 2 = 0, clutch-controlled ower-slit is ossible with the variator held at its maximum ratio, φ max. When the clutch locks, the variator can be used to control the ower flow through the transmission until the minimum ratio is reached, and the transmission can no longer discharge the flywheel. Transmission oerating modes during discharge The three transmission oerating modes are similar to the FDC variator-controlled PST described above, but are imlemented in a different order; i. Brake-controlled oeration hase 165

166 166 While the following relationshi between the seeds of branches 1 and 2 is true, ower is transmitted by alying a braking torque at the ring (5-25) ii. Clutch-controlled oeration hase This is required when the seed of shaft 6 is higher than the seed of shaft 5 using the maximum variator ratio; max i K K (5-26) The outut torque from the transmission is controlled by the torque at the clutch. The relationshi between transmission inut and outut torque is then given by; i rge discha K K T T max var (5-27) iii. Variator-controlled oeration hase This occurs when the clutch is locked and the required seed ratio across the variator lies within the following oerating limits; max min Where i K K (5-28) The outut torque from the transmission is then given by; i rge discha K K T T var (5-29) FWC esults The results from the analysis of the FWC variator-controlled PST can again be used to show transmission efficiency, DOD ov and U as contour lots with axes and eq for a given PGS and variator efficiency. The value of eq now relates to the clutch controlled ower-slit hase between branches 1 and 2, and is therefore a function of, K, K i and φ max as defined for FWC clutch-controlled transmissions in Chater 4, and restated in Equation i eq K K max 1 (5-30)

167 The normalised variator torque and seed can be defined as below, allowing the normalised rated variator ower to again be calculated. T T J var fd var K FWC i (5-31) T fd J fw var FWC 1 K i var fw,max (5-32) Using these contour mas, the minimum rated variator ower required to achieve a articular value of U can again be identified. In contrast to the results for the FDC configuration, this minimised ower curve does not match the conditions for maximum efficiency. At maximum U a large degree of variator-controlled oeration is achieved which results in a relatively high efficiency. As U decreases however, the brake-controlled hase begins to dominate leading to a shar decrease in efficiency. These results are encasulated in the following two Figures which form the design tool for the FWC variator-controlled PST η discharge η round-tri max (ω var ) P var, rated DOD ov Figure 5-11 Characteristics of FWC variator-controlled PST when P var, rated is minimised for a given U 167

168 1 0.8 eq DOD ov Figure 5-12 and eq of a FWC variator-controlled PST which minimise P var, rated for a given U Figures 5-11 and 5-12 rovide all the information necessary to define the erformance and design requirements of the FWC variator-controlled PST. The two single regime variatorcontrolled PSTs can now be comared with the direct variator transmission Comarison of single regime variator-controlled transmissions The strength of the indeendent analysis method resented in this Thesis is the ability to erform a rigorous comarison between flywheel transmissions of different configurations. The comarison of single regime transmissions is erformed on the basis of flywheel utilisation by using the data obtained for the minimised ower lines (i.e. oerating with minimum P var, rated for a given U) of the ower-slit configurations. The key arameters are the transmission efficiency and the required rated variator ower, and these are shown as functions of U for each transmission in Figure 5-13 (note that the definition of P var, rated is the same for all three transmissions, although the definitions of T var and ω var are different). 168

169 Transmission efficiency Pvar, rated Direct variator FDC PST FWC PST 1-PGS brake Flywheel utilisation, U Figure 5-13 Normalised rated variator ower as a function of U for 1-regime flywheel transmissions Flywheel utilisation, U Direct, η discharge Direct, η round-tri FDC, η discharge FDC, η round-tri FWC, η discharge FWC, η round-tri 1-PGS brake, η discharge 1-PGS brake, η round-tri Figure 5-14 Transmission efficiencies as a function of U for 1-regime flywheel transmissions A constant value of U reresents the case where each transmission has been sized in order to deliver the same amount of energy to the vehicle from identical flywheels. This would ideally be achieved with high efficiency (allowing a high degree of energy recovery during braking) and low rated variator ower (reducing the size, cost and mass of the variator). 169

170 Variator seed ratio, φ There are a number of conclusions to be made from the results shown in Figures 5-13 and 5-14, as discussed below. i. Both single regime ower-slit variator-controlled transmissions can achieve the same U with lower rated variator ower than the direct transmission ii. Both FDC and FWC configurations have lower overall efficiency than the direct variator, as ower-slit reduces the ratio coverage of the variator-controlled hase. This is illustrated in Figure 5-15 for three configurations using an identical flywheel to achieve a flywheel utilisation of 60%. Constant and zero variator seed ratio values corresond to hases of clutch and brake controlled oeration resectively. The reduced ratio coverage for the PSTs means that an increased roortion of the flywheel energy has to be delivered to the vehicle (with lower efficiency) during the clutch and brake controlled hases. A breakdown of the ercentage of total transmission inut energy dissiated in the transmission comonents of each configuration is shown in Figure C 1.0 Direct FDC FWC C B B B brake-controlled hase C clutch-controlled hase 0 Time t discharge Figure 5-15 Variator seed ratio as function of time for the three single regime transmissions during flywheel discharge with otimum conditions for U = 60% and the same maximum flywheel KE 170

171 Transmission energy loss Figure 5-16 Normalised rated variator ower and breakdown of transmission losses for single regime variator-controlled transmissions during flywheel discharge with U = 60% (transmission losses exressed as % of total transmission inut energy) As exected from the ideal ower-slit analysis resented in Section the FDC configuration achieves the lowest rated variator ower for a given U, and rovides a reasonable comromise between efficiency and variator size. This can be illustrated by considering the articular case when U = 60%; FDC discharge efficiency is 8% lower than direct transmission FDC rated variator ower is 60% lower than direct transmission The round-tri efficiency of the FDC variator-controlled transmission is around 50% when U = 60%. This is not articularly high, suggesting that there is scoe for imrovement in transmission erformance using multi-regime PSTs Single PGS, 2-regime variator-controlled transmissions The results of the revious section show that ower-slit arrangements can reduce variator losses and size, but that the reduction in the ratio coverage during variator-controlled oerations leads to a relative decrease in overall transmission efficiency for a given U. Multiregime transmissions allow the variator to oerate for a larger roortion of the charge/discharge cycle. This can increase the overall efficiency of the transmission by reducing the roortion of energy dissiated in the brake and clutch controlled hases. The effect on the rated variator ower requirement will deend on the configuration. Two classes of 2-regime transmission consisting of a single PGS are considered; 171

172 Single PGS with combined FDC/FWC ower-slit regimes (described in Section 5.3.1) Single PGS with 2 variator-controlled regimes and synchronous gear shift which can be achieved in FDC and FWC arrangements (both of which are described in Section 5.3.2) In order to assess the efficiency and rated variator ower the indeendent analysis method has been alied to each of these transmissions, and a comarison of erformance is made in Section Combined FDC/FWC 2-regime transmission The first tye of multile regime ower-slit configuration considered is a combination of the two single regime PSTs analysed in Section 5.2. This is ossible due to the limitations required to revent ower-recirculation in the FDC and FWC configurations, as restated below; FDC: FWC: It is clear that ower-slit oeration is ossible with an FDC arrangement at low vehicle seeds when ω 3 0 and ω 2 / ω 1 0. Once the vehicle has reached a seed where ω 2 / ω 1 0 then a FWC arrangement can be used to continue discharging the flywheel. The transmission therefore has 2 regimes of variator-controlled oeration, although clutch and brake controlled hases are still required when oerating outside the seed ratio limits of the variator. A ossible configuration for this tye of transmission is illustrated in Figure

173 PST Vehicle Flywheel unit Shaft 1 Shaft 2 Shaft 2 Shaft 3 Figure 5-17 Schematic diagram of ossible combined FDC/FWC variator-controlled PST configuration The sequence of oerating hases during flywheel discharge is as follows; i. Clutch-controlled FDC (φ = φ min ) ii. Variator-controlled FDC (φ min < φ < φ max ) iii. Brake-controlled at PGS ring iv. Clutch-controlled FWC (φ = φ max ) v. Variator-controlled FWC (φ max < φ < φ min ) Phases 1 to 3 are identical to the single regime FDC PST. The oerating range of the transmission is extended by using these hases to relace the initial brake-controlled hase of the FWC. Phases 4 and 5 are then identical to the single regime FWC PST. The combined FDC/FWC transmission now has 3 DOFs, corresonding to the values of, ( eq ) FDC and ( eq ) FWC. Analysis of this transmission is therefore erformed using the following aroach. Analysis rocedure for combined FDC/FWC 2-regime PST: i. Use the relationshi between and ( eq ) FDC obtained in Section 5.2 for the single regime FDC transmission which gives minimum rated variator ower and maximum discharge efficiency for a given U ii. Vary the values of and ( eq ) FWC and assess the erformance of the combined transmission By combining both FDC and FWC regimes of ower-slit in a single mode-switching 173

174 transmission it can be shown that the transmission efficiency is imroved with little additional gearing. This is due to the extended eriod of variator-controlled oeration, and the reduced ower flow thorough the variator which has the associated benefit of reducing the necessary ower-rating of the variator device (comared to the FWC PST) when oerated over the same range. The erformance of this transmission is comared with the other single- PGS transmissions in Section Synchronous shift 2-regime transmissions The second otion considered for multi-regime flywheel transmissions is the synchronous gear change 2-regime configuration. This is an extension of the basic single-regime FDC and FWC PSTs described in Section 5.2, and allows an increased range of variator-controlled oeration by roviding two searate stages of ower-slit between the flywheel and vehicle. The key feature of these transmissions is that the variator seed ratio at the end of the first stage is the initial value required in the second stage, and that the switch between stages is achieved through the synchronous engagement of clutch elements. The fundamental requirements of this tye of transmission for inut and outut couled configurations have been characterised by White [92], and the efficiency of an inut couled version (described as a 2-stage PS-CVT ) has been indeendently analysed by Mantriota [101]. The alication of these transmissions for flywheel systems has not however been considered. This section therefore alies the indeendent analysis method to quantify the overall erformance (including hases of brake and clutch control) for both FDC and FWC configurations and generate design tools for identification of aroriate system arameters for a given alication. a) FDC synchronous shift configuration The fundamental constraints for a 2-regime synchronous shift FDC transmission are derived by White [92] for the simle case of a variator and PGS with no additional gearing. The key requirement is that the PGS gear ratio, > 1, which is unaffected by the inclusion of the gear ratios shown in Figure These gear ratios are required in order to ensure an aroriate range of oeration for the variator, and that the synchronous gear change between regimes 1 and 2 is achieved at the oint where the variator is oerating at its limiting seed 174

175 ratio. The required value of K 2 must therefore be chosen so that ω 4 = ω 6 when oerating in regime 1 with φ = φ max. This is achieved when; K 2 /K o = 1/φ max equirement for FDC synchronous shift (5-33) Vehicle Flywheel unit 1 3 Clutch to engage egime 1 Final drive 2 Fixed Gear atios K o K = ω 4 / ω 2 K o = ω 5 / ω 3 K 2 = ω 6 / ω fd φ = ω 5 / ω 4 K 4 V 5 K 2 6 Clutch to engage egime 2 Figure 5-18 Schematic diagram of FDC synchronous shift PST configuration (dashed arrows show direction of ower-flow during each regime of flywheel discharge) egime 1 oeration is identical to the single regime FDC transmission considered earlier (see Figure 5-5), as the different osition of the clutch makes no difference to the roortion of energy dissiated during the clutch-controlled hase. For the case of flywheel discharge, the initial clutch and variator controlled hases of oeration are therefore the same. When the variator ratio reaches φ max (at the end of the regime 1 variator-controlled hase) the seeds of shafts 5 and 6 are exactly equal (using the K 2 /K o ratio secified in Equation 5-33) and a synchronous shift to regime 2 is achieved. A second stage of variator-controlled oeration is now achieved as the variator seed ratio is decreased, until it reaches φ min. As the vehicle is now connected to branch 2 of the PGS (via the variator), a final hase of flywheel discharge can be achieved by braking branch 3. The flywheel can no longer be discharged once the seed of branch 3 reaches zero. 175

176 The same analysis method can therefore be used as for the single stage FDC transmission with and ( eq ) FDC values defining an oerating oint, and the conditions required to minimise P var, rated for a given U identified. b) FWC synchronous shift configuration The following constraints aly to the imlementation of a 2-regime FWC synchronous shift PST. In order to achieve the required two hases of ower-slit oeration, 0< <1, as derived by White [92]. Furthermore, to achieve a synchonous gear shift between ower-slit regimes; K 2 /K i = 1/φ min equirement for FWC synchronous shift (5-34) A schematic illustration of the required transmission configuration is shown in Figure Vehicle Flywheel unit Clutch to engage egime Final drive 2 Fixed Gear atios K = ω 4 / ω 2 K 2 K i i 5 K K i = ω 5 / ω 1 K 2 = ω 6 / ω fw φ = ω 5 / ω 4 6 V 4 Clutch to engage egime 2 Figure 5-19 Schematic diagram of FWC synchronous shift PST configuration (arrows show direction of ower-flow during flywheel discharge) This configuration can again be investigated by modifying the existing time-ste based comutation model of the single regime FWC variator-controlled PST to include the second regime of variator oeration. The erformance of the FWC 2-regime PST can therefore be 176

177 Pvar, rated resented in terms of the normalised gear ratios, and ( eq ) FWC, and a minimised ower curve can again be identified Comarison of multile regime variator-controlled transmissions The erformance of the three 2-regime PSTs can now be comared with the earlier results for the single regime PSTs as shown in Figure This is achieved by using the minimised ower curves for each configuration, and comaring the values of discharge efficiency and rated variator ower as functions of U. 1-regime Direct variator FDC FWC 2-regime Combined FDC/FWC FDC synchronous shift FWC synchronous shift ηdischarge U U Figure 5-20 Normalised rated variator ower and discharge efficiency as a function of U for 1 and 2 regime flywheel transmissions Similar values of efficiency as a function of U are obtained for all three 2-regime transmissions. These results illustrate that the efficiency of the transmissions is imroved by increasing the roortion of variator-controlled oeration. This is clearly seen in Figure 5-21 which shows the variator seed ratio during each of the brake, clutch and variator controlled hases for all 2-regime transmissions when sized to achieve U = 60% with the same flywheel unit. 177

178 Transmission energy loss Variator seed ratio, φ 2.5 C C C B B B Time t discharge Direct FDC/FWC FDC sync FWC sync B brake-controlled hase C clutch-controlled hase Figure 5-21 Variator seed ratio as function of time for the three 2-regime PSTs during flywheel discharge with minimised ower conditions for U = 60% using the same maximum flywheel KE A breakdown of the losses which occur in each comonent of the 2-regime transmissions is given in Figure Figure 5-22 Breakdown of losses for 2-regime variator-controlled PSTs during flywheel discharge with U = 60% (losses exressed as % of total transmission inut energy) The increase in efficiency which is achieved by all three 2-regime transmissions considered allows a higher maximum U to be achieved. There is however no significant decrease in rated variator ower when comared with the 1-regime transmission. The FDC synchronous transmission achieves lowest ower, but this requires > 1 which resticts the ractical sizing of the PGS and may require large fixed gear ratios (esecially between the high seed flywheel and the PGS). The FWC synchronous and combined FDC/FWC transmission have otentially more ractical requirements for FESS use, but have significantly higher rated 178