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Fig. 10 shows the speed curve of the flywheel intwo cycles of energy-recovery. The figure also shows that the simula-tion model validated the test.Fig. 11 shows the simulation and experimental results for thepressure of accumulator HA1. As seen in the figure, the pressure inthe accumulator increased from the 60th second to the 63rd secondand decreased again until the 65th second. The next recovery cycleoccurred in the interval from the 70th to the 73rd seconds. Obviously,the initial kinetic energy of the flywheel was small, and the cycle ofthe recovery was short. The measured pressure was slightly lowerthan the simulation pressure, so the measured round-trip efficiencyof the test bench would be lower than the estimated value in the pre-vious subsection.Figs. 12 and 13 show additional detail about energy transforma-tion in the system. The pressure in the accumulator increased from120 bar to 130 bar, corresponding to a flywheel speed that decreasedfrom100 rad/s to nearly 0 rad/s at the 63rd second. Then, the speed ofthe flywheel increased again, but the pressure decreased from130 bar down to 120 bar at the 65th second. The flywheel'smaximumspeed was 73 rad/s and the round-trip efficiency was 53%. Fig. 13 in-dicates that the accumulator recovered 8.3 kJ of the flywheel's 12.5 kJof kinetic energy and again generated 6.7 kJ of the flywheel's kineticenergy in the next cycle. Thus, the recovery efficiency in this casewas 66.4%.Fig. 14 shows values of p2 when valve V12 was activated. The pres-sure increased slightly at the beginning due to the delay of the direc-tional control valve. In the system, the use of the check valve CV5prevented a high-pressure peak, illustrating the acceptability of thesystem for practical applications.Fig. 15 shows the measured round-trip efficiency of the systemversus the initial speed and displacement ratio of PM2. In thesetests, accumulator HA2 was pre-charged to 10 bar. From the figure,the maximum and minimum round-trip efficiencies were 59% and22%, respectively. The results were lower than the estimated resultsin the previous section because the braking time of the system inthe case of α=0.5 was greater than 10 s. The round-trip efficiencysignificantly decreased with the low displacement ratio. Fig. 16shows that the round trip efficiency of the system varies accordingto the initial fluid pre-charge of the large low pressure accumulatorHA2. This reflects the limitations of hydraulic technology, becauseonly closed-loop hydraulic devices are employed as secondary units.If the initial fluid pressure is low, it will be inadequate for closed hy-draulic machine function. All drain lines from hydraulic machines areconnected to low pressure accumulator HA2, so the pre-charge fluidpressure should not be too high. The desired value of the initialfluid pressure in HA2 is typically determined by trial and error, andvalues ranging from 10 bar to 20 bar are usually adequate. 5. Assessment of the system5.1. System performanceTo validate the applicability of the system, the system with thecontroller designed in Section 3 performed three representative mis-sion profiles including 10-mode, modified 10-mode, and highwayprofiles. The test bench was controlled using the following threestrategies.1. Traditional hydrostatic transmission (HST): when both valves V11and V12 were inactive, the speed of the flywheel was reduced bythe hydraulic brake system.2. Traditional CPR or non-switching: when valve V11 was active butV12 was inactive, the system became a semi closed-loop CPR sys-tem. The AFSMC controller was employed with regenerativebraking.3. Switching: the scheme of this strategy was presented in Section 3.The experimental parameters of the fuzzy membership functionsof the AFSMC are shown in Table 3. The fuzzy labels used in thisstudy were negative big (NB), negative medium (NM), negativesmall (NS), zero (ZO), positive small (PS), positive medium (PM),and positive big (PB). The output signals from the sensors and thecontrol output to the pump/motor PM2 were scaled into [−1 1].