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The pump and thepump/motor in the proposed system were pressure coupled to im-prove the efficiency of the primary power source. Because of thenovel structure of the proposed system, the efficiency of the second-ary unit could be improved by using high efficiency pump/motors,such as bend axis pump/motors work in the positive region. Thepump supplied only a load, which could improve the efficiency andperformance of the load [17].To operate the proposed system, a hierarchal control system in-cluding a supervisory control level and a plant control level wasdesigned. The supervisory control determined the strategy controlof the total system and generated calculated reference values forthe plant control level. The plant control level included the prima-ry, the secondary, and the valve controllers, which controlled thedirectional control valve, the pump, and the secondary unit, re-spectively. We only focused on energy-recovery and the controlla-bility of the system, allowing for simple design of the supervisorycontroller, pump, and valve controllers. The sophistication of thecontroller design was due to the synthesis of the secondarycontroller.For the speed control of the secondary unit, some studies haveused a PID [16], a bilinear control [18], a robust control [19], or a ro-bust with nonlinear control [20]. However, in the above studies, thecontrollers were designed for small model uncertainties and werenot suitable for use in the proposed system. Thus, an adaptive controlscheme was proposed.An adaptive fuzzy sliding mode control (AFSMC) was developedbecause it combined the robust characteristic of a traditional slidingmode control and the online tuning of the adaptive fuzzy control[21,22]. The parameters of the fuzzy rules base were tuned accordingto an adaptive mechanism so that the performance of the system wasguaranteed. The use of a sliding surface as the input allowed for a sim-pler controller design [23], and the system possessed Lyapunovstability.This paper was organized as follows: Section 2 discusses the prop-erties of the system based on mathematical equations derived viaphysical modeling, Section 3 describes the control system design,Section 4 presents the test bench system and the experimental pro-grams used for experimental validation of the basis system character-istics. An assessment of the system is presented in Section 5, and theconclusions are presented in Section 6.2.
Proposed system2.1. Principle of operationFig. 1 shows the schematic of the proposed system. In comparisonwith a traditional HST system, a directional control valve and two hy-draulic accumulators (HA1,HA2) were added. The high-pressure ac-cumulator HA1 functioned as a storage system or a power supply,and the low-pressure accumulator HA2 functioned as a low pressure,high-flow source for the hydraulic pump/motor PM2 during recovery,and the boots system while driving. The flywheel simulated the prac-tical load. The secondary unit and the pump/motor PM2 could func-tion as a hydraulic pump or a hydraulic motor depending on thespeed, displacement and pressure difference between the two portsof PM2. Fig. 2 shows the four-quadrant operation of the secondaryunit. If PM2 operated in quadrant I or III, the flywheel would be pow-ered by the system; however, if PM2 operated in quadrant II or quad-rant IV, the system recovered the kinetic energy of the flywheel.In the proposed system, driving and braking were distinguished viacontrol of the directional valve. Fig. 3 shows the driving phase of thesystemwhere p1 became the high pressure and either the accumulatorHA1 or the electric motor M powered the flywheel. The pilot checkvalve PCV2 was opened, and the system was in a semi closed-loop cir-cuit. PM2 was controlled to activate during quadrant I operation. Dur-ing braking, solenoid V12 was engaged and pump P1 was deactivated.Pressure p1 was decreased and the pilot check valve PCV2 was closed.The pressure p2 increased and PM2 operated in quadrant IV. Therefore,the kinetic energy of the flywheel was transformed into hydraulic en-ergy and was stored in the high pressure accumulator HA1.Theuseof check valves CV1,2,3,4,5 prevented cavitation and absorbed the peakpressure in p2 when the directional was switched.Whenever the system receives a brake signal the solenoid V12 isactivated, p1 decreases and p2 is pressured directly from the accumu-lator HA1. Hence, the proposed system is able to handle any brake ac-tion for which the braking time is greater than the response time ofthe directional control valve. When the braking time is shorter thanthe response time of the valve, the system is unable to brake the fly-wheel. However, the response time of a directional control valve,which is the time needed for the spool of the valve to totally changeits position, is often less than 0.1 s, which is also less than the time re-quired for a brake action. Furthermore, the use of the control valve re-duces the time required for regenerative braking compared to a CPRsystem, in which the displacement of the pump/motor reverts fromthe positive to the negative region to generate a braking action(Fig. 4). 2.2. Modeling2.2.1. Hydraulic pump/motorsThe volumetric and mechanical losses in a general axial piston hy-draulic machine were used in this study in the same manner as in[24]. In a practical hydraulic machine, the volumetric losses causeinlet restriction, leakage, and fluid compressibility. Viscous, Coulomb,and hydrodynamic frictions cause mechanical losses. The volumetriclosses Qloss and mechanical losses Tloss of a hydraulic machine areexpressed in Eqs. (1)–(8).Δq1 ¼ 1:76 10−7ω=2π ðÞþ 1:7 10−13Δp ð1ÞΔq2 ¼ 5:10 10−14−2:83 10−14α ω=2π ðÞΔp ð2ÞΔq3 ¼ 5:80 10−13ω=2π ðÞDmaxΔp ð3ÞQloss α;ω;Δp ðÞ¼ Δq1 þ Δq2 þ Δq3 ð4ÞΔT1 ¼ 4:27 10−2ω=2π ðÞ ð5Þ ΔT2 ¼ 5:84 10−2αω=2π ðÞ þ 8:02 10−4α2ω=2π ðÞ2ð6ÞΔT3 ¼ 2:5 10−7Δp−8:1 10−15Δp2ð7ÞTloss α;ω;Δp ðÞ¼ ΔT1 þ ΔT2 þ ΔT3 ð8Þwhere, α, ω,Dmax, Δp, Δq1,2, Δq3, ΔT1, ΔT2 and ΔT3 are the displace-ment ratio, speed, maximum displacement, pressure difference be-tween two ports, volumetric losses due to leakage, volumetric lossdue to the compressibility of fluid, torque loss due to viscosity, torqueloss due to leakage, and the torque loss due to compressibility of fluidof the pump/motor, respectively.2.2.1.1. Hydraulic pump. Eqs. (9)–(12) express the ideal flow rate offluid through a pump or motor, the mechanical torque at the pumpshaft, the pump volumetric efficiency, and the torque efficiency, re-spectively:Qi¼ αωDmax; ð9ÞηvP ¼ Qi−QlossQi; ð10ÞTi¼ αΔpDmax; ð11ÞηtP ¼ αDmaxΔpαDmaxΔp þ Tloss: ð12Þ2.2.1.2. Hydraulic motor. Eqs.