Electric-driven Zonal Hydraulics in Non-Road Mobile Machinery

The goal of this research is to apply direct-driven hydraulics (DDH) to the concept of zo‐ nal (i.e., locally and operation-focused) hydraulics, which is an essential step in the hy‐ bridization and automation of machines. DDH itself aims to combine the best properties of electric and hydraulic technologies and will lead to increased productivity, minimized energy consumption and higher robust performance in both stationary and mobile ma‐ chines operating in various environments. In the proposed setup, the speed and position control of a double-acting cylinder is implemented directly with an electric motor drive in a closed-loop system without conventional control valves and an oil tank. The selection of the location of the hydraulic accumulator and connection of the external leakage lines will also be part of this study. Simulations and experimental research to study the details of the hydromechanical and electrical realization of the DDH are performed.


Introduction
The climate for industrial and economic growth is changing as resource scarcity pushes up costs, and enterprises are more closely scrutinized because of changing expectations about their public and social responsibilities. Governments have set tight CO 2 emission rules [1] and new exhaust limits have just been implemented, such as Tier Emission Standards. The next known step will be the 2019/2020 Tier V, which is a regulation imposing an ulterior sharp tightening of exhaust limits, especially in terms of particles. Now, there is a four-year time window to prepare engines for the upcoming regulations through solutions for peak power shaving and the downsizing of diesel engines. Electric and hybrid vehicles are a suitable solution for reaching these target environmental requirements; such vehicles have a huge potential for application in the non-road mobile machinery (NRMM) industry, with its market in mining, process and goods manufacturing, forest harvesting and construction work. Figure  1 illustrates typical examples of NRMMs. 1

Introduction
The climate for industrial and economic growth is changing as resource scarcity pushes up costs, and enterprises are more closely scrutinized because of changing expectations about their public and social responsibilities. Governments have set tight CO 2 emission rules [1] and new exhaust limits have just been implemented, such as Tier Emission Standards. The next known step will be the 2019/2020 Tier V, which is a regulation imposing an ulterior sharp tightening of exhaust limits, especially in terms of particles. Now, there is a four-year time window to prepare engines for the upcoming regulations through solutions for peak power shaving and the downsizing of diesel engines. Electric and hybrid vehicles are a suitable solution for reaching these target environmental requirements; such vehicles have a huge potential for application in the non-road mobile machinery (NRMM) industry, with its market in mining, process and good manufacturing, forest harvesting and construction works. Figure 1 illustrates typical examples of NRMMs. The NRMM industry and its customers are traditionally quite conservative in putting new ideas and technologies into use. When they opt for "mild hybridization", the machine manufacturers do not need to change the construction frame of the whole machine immediately. However, some changes in layout design concepts, machine types and business models should be considered in the long term. Therefore, we address several challenges to improve NRMMs by bringing zonal hydraulics into machines. Combining the best properties of traditional hydraulics and electric intelligence allows the following benefits to be achieved in NRMMs:  easy electrification of NRMMs  higher efficiency compared to conventional machines  electrohydraulic power pack with no tank and pipelines  reduction of potential leakage points  power-on-demand control  sensorless position control The NRMM industry and its customers are traditionally quite conservative in putting new ideas and technologies into use. When they opt for "mild hybridization", the machine manufacturers do not need to change the construction frame of the whole machine immediately. However, some changes in layout design concepts, machine types and business models should be considered in the long term. Therefore, we address several challenges to improve NRMMs by bringing zonal hydraulics into machines. Combining the best properties of traditional hydraulics and electric intelligence allows the following benefits to be achieved in NRMMs: • easy electrification of NRMMs • higher efficiency compared to conventional machines • electrohydraulic power pack with no tank and pipelines • reduction of potential leakage points • power-on-demand control • sensorless position control

State of the art
During recent years, the NRMM industry has had a tendency towards integrated, compact electrohydraulic systems that deliver powerful, linear movement with either valve-or pumpcontrolled systems. These technological steps are considered important stages towards a solution to commonly identified problems, such as the reduction of CO 2 emissions and improved performance, productivity, reliability and controllability [2,3], as well as good options for forcing a smaller diesel engine to operate within its optimum efficiency area since, in general, the diesel engines of NRMMs operate far from their optimum efficiency range [4].
Restrictive environmental regulations are not the only factor imposing the need for innovation in NRMMs. Today, as the search for natural resources expands to areas previously considered too remote and environmentally hostile to support viable extraction and processing operations, researchers all around the world are facing the challenge of establishing and maintaining industrial activities in extreme environments. In these, components of the drive face thermal issues, primarily at start-up, when everything in the drive has cooled down to low, even Arctic temperatures. Under these conditions, conventional hydraulic fluid in all hydraulic systems solidify, so when the NRMM starts up, there is no fluid circulating to protect the pump and other components from extreme wear. For successful operation, a relevant start-up procedure is crucial, as the risk of equipment failure is unacceptable in remote locations.
Zonal or decentralized hydraulics-as we will call it-is an approach first introduced in the aircraft industry [5]. Adapting this time-proven design from the aircraft industry, simplifying its design as well as broadening its orientation and performance options will support a similar development in NRMMs. In a fully zonal system, the hydraulic pumps are removed from the engine and replaced within hydraulic power packs distributed throughout the NRMM system. In this architecture, multiple hydraulic power sources may be utilized in each zone in order to achieve energy savings and work with a power-on-demand approach.   In this study, an electrohydraulic actuator (EHA) will be used to achieve high power density, low noise and high performance in a compact package in order to take the development of a power unit onto the next level. The electrohydraulic actuator allows decoupling from the main hydraulic system, thus enabling a zonal approach and, at the same time, reducing parasitic losses in order to obtain better fuel efficiency and lower operating costs. The robust, leak-free, one-piece housing design delivers system simplicity and lowers both installation and maintenance costs. The electrohydraulic actuator reduces space and weight demands. It can eliminate hoses, fittings, valves and fixtures and is easy to integrate into larger systems.
Currently, electrohydraulic actuators are mostly installed and developed for aircraft applications [6,7], where the price level and reliability requirements are very high. Most of the research studies related to electrohydraulic actuators have been conducted to adjust the state of the servo valve [8,9]. However, this approach results in low energy efficiency because of the flow via the pressure relief valves of the hydraulic pumps and throttle losses at the control valves. Consequently, several techniques have been developed to overcome this drawback in order to achieve higher efficiency. In ref. [10], the concept of a pump-controlled electrohydraulic actuator is introduced as an advanced hydraulic system where the proposed structure of the electrohydraulic actuator is directly operated by a bidirectional pump. There are already commercially integrated power packages on the market, but they are based on conventional technology and they do not respond to the above-mentioned challenges created by the hybridization problem. Examples of commercial pump-controlled electrohydraulic actuators are, for instance, the mini-motion package from Kayaba Industry Co. [11] and the intelligent hydraulic servo drive-pack from Yuken Kogyo Co. [12]. The disadvantages of the architecture evidenced in ref. [10] are the use of complex and expensive pumps and the lower dynamic properties of the system in general. In addition, none of these commercial solutions are designed for power-on-demand control, sensorless positioning or even energy regeneration, all highly desirable improvements in a competitive product.
Some studies have been conducted to introduce the concept of electrohydraulic actuators as zonal hydraulics in NRMM. In ref. [13], an electrohydraulic actuator is used for the power steering of heavy vehicles. In refs. [14][15][16][17], a compact drive for automation of all kinds of linear motions was introduced and investigated from the thermal point of view. In these sources, even if they proposed direct pump control similar to the direct-driven hydraulic approach, a set of valves is used to balance the flow and to ensure the direction of the electrohydraulic actuator motion. In refs. [18,19], direct-driven hydraulics (DDH), an electrohydraulic actuator, was introduced without conventional directional valves. The DDH drive combines the best properties of electric and hydraulic drive technologies in one: • direct control of flow, as well as the velocity and position of the actuator Therefore, the idea of implementing the DDH drive in the powertrain of NRMMs as an application of the zonal hydraulic concept was born. In this research, direct-driven hydraulics (DDH) is seen as a tool to convert existing NRMMs to hybrids combining the best properties of traditional hydraulics and electric intelligence. Figure 3 illustrates the hybrid proposal for NRMMs. With this approach, several challenges that traditionally manifest themselves in the hybridization process will be solved. In the proposed setup, the speed and position control of a double-acting cylinder is implemented directly with a motor drive in a closed-loop system without conventional control valves and an oil tank. The selection of the optimal size and the location of the hydraulic accumulator will also be part of this study. Simulations and experimental research to study the details of the hydro-mechanical and electrical realization of the DDH will be performed. The next chapter introduces the experimental DDH setup.

Description of test setup
The experimental setup uses a speed-controlled electric servo motor drive rotating two hydraulic pumps to directly control the amount of hydraulic oil pumped to the asymmetrical double-acting cylinder MIRO C-10-60/30× 400. The simplified circuit diagram of the experimental test setup is illustrated in Figure 4. The hydraulic pump/motors P 1 and P 2 create an input and output flow that depends on the rotating speed of the servo motor. The oil pressure rises to the required level as determined by the payload. During lowering motion, the potential energy of the payload creates a flow that rotates the hydraulic machine P 1 as a motor and the hydraulic machine P 2 as a pump, and the mechanically connected electric motor acts as a generator, which is controlled by the frequency converter. The speed-controlled generator controls the amount of fluid flow and the position of the payload. The program for the electric drive controls both the electrical and hydraulic sides of the system as there is no conventional valve control. For easier understanding of the real displacements of the pumps, the following ratios R A and R Q will be defined. In this work, the ratio between the cylinder areas is defined by the following equation: is the cylinder area from the rod side. The diameter of the cylinder piston head is d 1 = 0.06 m and d 2 = 0.03 m is the diameter of the piston rod, which determines the piston surface areas A 1 and A 2 . According to the datasheet, the cylinder stroke is l = 0.4 m.
As previously stated, the pump/motors P 1 and P 2 are mounted on the same axis, so their speeds are identical. If the pump leakage is ignored, the ratio of the flow rates R Q can be produced directly from the displacement of the pump/motors D 1 and D 2 . Figure 4 illustrates the first prototype of the DDH setup with R Q ≈ 0.63 and R A = 0.75. To test the concept, the components for this prototype were taken off the shelf.
As previously stated, the pump/motors P 1 and P 2 are mounted on the same axis, so their speeds are identical. If pump leakage is ignored, the ratio of the flow rates R Q can be produced directly from the displacement of pump/motors D 1 and D 2 . Figure 4 illustrates the first prototype of the DDH setup with and . To test the concept, components for this prototype were taken from the shelf.  Two XV-2M internal gear pump/motors by Vivoil with displacements of 14.4 and 22.8 cm 3 /rev were used, P 2 P 1 , respectively [20]. The position feedback from the motor is given by means of its in-built incremental enc (4096 pulses per revolution, resolution 14 bits), and read with the Unidrive SP1406 drive software [21]. It conv the AC power supply from the line and allows the speed of the permanent magnet brushless servo motor, Unim 115U2C manufactured by Emerson Control Techniques, to be set, taking advantage of the information obtaine the feedback device fitted to ensure the rotor speed is exactly as demanded [22]. This experimental setup was te with a payload of 150 kg at motor speed ranges from 300 to 500 rpm. Figure 5 shows an example of measured for a motor speed of 400 rpm and a payload of 150 kg: speed, torque and pressure. wire-actuated encoder, c) pressure sensor, d) reversible gear pump/motor P 1 , e) pressure sensor, f) gearbox, g) PMSM motor/generator, h) current sensors, i)frequency converter, j) tank, k) pressure sensor, l) reversible gear pump/ motor P 2 and m) pressure sensor in the tank line; (c) the crane is used for the DDH setup as a test platform (side view).
Two XV-2M internal gear pump/motors by Vivoil with displacements of 14.4 and 22.8 cm 3 /rev were used, P 2 and P 1 , respectively [20]. The position feedback from the motor is given by means of its in-built incremental encoder (4096 pulses per revolution, resolution 14 bits), and read with the Unidrive SP1406 drive software [21]. It converts the AC power supply from the line and allows the speed of the permanent magnet brushless servo motor, Unimotor 115U2C manufactured by Emerson Control Techniques, to be set, taking advantage of the information obtained by the feedback device fitted to ensure the rotor speed is exactly as demanded [22]. This experimental setup was tested with a payload of 150 kg at motor speed ranges from 300 to 500 rpm. Figure 5 shows an example of measured data for a motor speed of 400 rpm and a payload of 150 kg: speed, torque and pressure. In Figure 5, during the lifting, which lasts from 1 to 7 s, the pressure in the pump/motor P 1 is about 2 MPa, which rises by the end to 4 MPa. It is worth remarking that the rise appears to be due to the difference between R Q and R A . A similar rise also occurs in the pump/motor line P 2 , from atmospheric pressure to 2.6 MPa.
During the lowering, performed from 7 to 13 s, a drop in the pressure from high pressure to about 1 MPa happens. During the lifting, the motor torque is 7 Nm; during the lowering, the torque is around 1 Nm. The pressure in the tank line is close to atmospheric pressure, as it is supposed to be in an open system. In order to overcome the rise in pressure at the end of the movement and investigate the tankless approach, a new prototype was designed. Figure 6 illustrates the experimental test setup of the second DDH prototype. The system is closed by giving up the tank completely (compare to Figure 2b) and replacing it with a hydraulic accumulator A. Within the framework of this work for the second prototype and . It can be seen that the ideal R Q = R A was not achieved. Therefore, an accumulator B is added to compensate for the displacement difference between the available pump/motors and cylinder areas.

Components used
In the second prototype, DDH setup an electric motor, an IndraDyn T MST130A-0250-N torque motor from Bosch Rexroth, is used. The parameters of the electric motor are shown in Table 1. Two Parker AD100B20T9A1 diaphragm accumulators are utilized as the initial choices. Table 4 shows the Within the framework of this work for the second prototype R Q ≈ 0.73 and R A = 0.75. It can be seen that the ideal R Q = R A was not achieved. Therefore, an accumulator B is added to compensate for the displacement difference between the available pump/motors and cylinder areas.

Components used
In the second prototype, DDH setup an electric motor, an IndraDyn T MST130A-0250-N torque motor from Bosch Rexroth, is used. The parameters of the electric motor are shown in Table 1.   [23].

New Applications of Electric Drives
Bosch Rexroth AZMF-12-011U and AZMF-12-008U external gear motors are used as hydraulic pumps/motors. The parameters of the pump/motors are shown in Tables 2 and 3.  Two Parker AD100B20T9A1 diaphragm accumulators are utilized as the initial choices. Table  4 shows the parameters of the hydraulic accumulators.  According to the manufacturer, the pump/motors used have an external leakage-oil line with a pressure limitation of 0.3 MPa. During the start, the maximum allowed short-term pressure is 1 MPa. Thus, a detailed investigation of the hydraulic connection of the external leakage line is required.

Investigation of the hydraulic connections
This chapter introduces calculations for pump/motor leakage flow in order to identify a suitable connection for the external leakage line. The DDH system can be divided into three separated fluid volumes: the piston side of the cylinders A 1 and the rod side of the cylinders A 3 with the hydraulic accumulator B and the hydraulic accumulator A between the pump/ motors. The accumulation of hydraulic fluid in any of these volumes will raise the pressure, as was shown in Figure 5 for the first prototype.
This chapter contains a simple model in which the volumetric efficiency of the pumps η v, p , determines the actual volume produced and the leakage flow of the pump/motor. For simplicity, it is assumed that this efficiency is the same for both pumps/motors. This model is a rough simplification as in reality, the leakage flow is distributed both as internal and external leakage flow [25] and volumetric efficiency varies as a function of the differential pressure and speed. This case study aims to find out how leakages should be controlled so that the system is as insensitive to fluctuations in the volumetric efficiency as possible.
The calculations were performed by determining the change in the volume of fluid during the lifting and lowering with the maximum stroke length of the cylinder l [26]. The volume changes are determined during a single lifting-lowering cycle.
Next, the theoretical flow Q Vt , the actual flow Q Vr and the leakage flow Q Vl for both pumps are determined. It is assumed that the speed of the electric motor is n in rev/s. In this case, the pumps produce theoretical flows of: where D 1 and D 2 are the theoretical pump displacements for the pump/motors P 1 and P 2 , respectively.
The actual flows Q Vr1 and Q Vr2 and leakage flows Q Vl1 and Q Vl2 in accordance with the volumetric pump efficiency η (D, p) are During the lifting, the side flow Q Vc1,up of the cylinder A 1 is equal to the actual flow Q Vr1 produced by the pump/motor P 1 . Theoretical flow Q Vt1 leaves the pump/motor P 1 during the lowering from A 3 cylinder's side. Therefore, The cylinder losses are assumed to be negligible, so the required flows Q Vc2,up and Q Vc2,down are determined as It is assumed that the cylinder is moved at a constant velocity v, which can be determined as follows for lifting and lowering, respectively: In this case, the raising (t up ) and lowering (t down ) times can be calculated by combining Equations (1), (3) and (11): As was stated earlier, the external leakage line of the pump/motors that are used should be connected to the line under 0.3 MPa. The next section will determine the structure. The least sensitive case will be selected as the system structure. The review will be carried out for a total of four different alternative cases. Figure 7a illustrates Case I, where both the leakage lines are connected to the line with hydraulic accumulator A. In Case II in Figure 7b, both external leakage lines are connected to hydraulic accumulator B. Figure 7c shows Case III, where the external leakage lines of the pump/motors P 2 and P 1 are connected to the line with hydraulic accumulators A and B, respectively. Figure 7d illustrates Case IV, where the external leakage lines of the pump/motors P 1 and P 2 are connected to the lines with hydraulic accumulators A and B, respectively.

Case I
In Case I, during the lifting, the A 3 side of the cylinder should remove flow equal to Q Vc2,up , and the pump P 2 is able to remove the theoretical flow of Q Vt2 . Thus, during the entire lifting motion, the volume ΔV up,I can be obtained with Equations (2), (9) and (12): (14) During the lowering, the A 3 chamber of the cylinder should receive the flow of Q Vc2,down . The pump/motor P 2 is able to produce the actual flow of Q Vr2 . Thus, during the whole cycle in the A 3 chamber of the cylinder, the following difference in volumes ΔV down,I occurs, which is obtained with Equations (4), (10) and (13): The entire cycle change of volume ΔV t,I in Case I is:

Case I
In Case I, during the lifting, the A 3 side of the cylinder should remove flow equal to Q Vc2,up , and the pump P 2 is able to remove the theoretical flow of Q Vt2 . Thus, during the entire lifting motion, the volume ΔV up,I can be obtained with Equations (2), (9) and (12): During the lowering, the A 3 chamber of the cylinder should receive the flow of Q Vc2,down . The pump/motor P 2 is able to produce the actual flow of Q Vr2 . Thus, during the whole cycle in the A 3 chamber of the cylinder, the following difference in volumes ΔV down,I occurs, which is obtained with Equations (4), (10) and (13):

New Applications of Electric Drives 154
The entire cycle change of volume ΔV t,I in Case I is: Figure 7b illustrates Case II, where both of the leakage lines are connected to the line with hydraulic accumulator B and A 3 chamber branch. During the lifting, the chamber A 3 should remove the flow Q Vc2, up , as well as the leakage flows of the pumps Q Vl1 and Q Vl2 . The pump/ motor P 2 is able to remove the theoretical flow Q Vt2 . In this case, the volume ΔV up, II is obtained with Equations (2), (5), (6), (9) and (12):

Case II
During the lowering, the A 1 chamber of the cylinder should remove the flow Q Vc2,down . The pump/motor P 2 is able to provide the flow Q Vr2 . In addition, the same line will also have the external leakage flows Q Vl1 and Q Vl2 . In this case, the change in the volume ΔV down, II can be obtained with equations (4), (5), (6), (10) and (13): The entire cycle change of volume ΔV t,II in Case II is:

Case III
In Case III, the leakage flow of the pump/motor P 2 is connected to the hydraulic accumulator A line and the P 1 leakage flow to the A 3 chamber line of the cylinder, as shown in Figure 7c. During the lifting, the A3 chamber of the cylinder should pump out the flow Q Vc2,up and the pump/motor P 1 the leakage flow Q Vl1 , but at the same time the pump/motor P 2 can remove the flow of Q Vt2 . Thus, during the whole stroke of the lift, the change in the volume is derived from Equations (2), (5), (9) and (12): During the lowering, the A 3 chamber of the cylinder should provide the flow Q Vc2,down . The pump/motor P 2 is able to pump out a flow equal to Q Vr2. In this case, the total volume of ΔV down,III is obtained with Equations (4), (5), (10) and (13): The entire cycle volume change ΔV t, III in Case III is:

Case IV
In Case IV, the leakage flow from the pump/motor P 1 is directed to the hydraulic accumulator A and the leakage flow of the pump/motor P 2 is directed to the hydraulic accumulator B. During the lifting, A 3 chamber flow of the cylinder Q Vc2,up and leakage flow Q Vl2 should be removed by the pump/motor P 2 , but the pump/motor P 2 is able to actually pump out the flow of Q Vt2 . Thus, during the whole stroke, the change in the volume of the fluid chamber ΔV up,IV is obtained with Equations (2), (6), (9) and (12): During the lowering, the A 3 chamber of the cylinder should pump out the Q Vc2,down volume flow. The required pumping flow rate of P 2 is Q Vr2 and the additional leakage flow is Q Vl2 . In this case, the change in the volume ΔV down,IV is obtained with Equations (4), (6), (10) and (13): The entire cycle change in the volume ΔV t, IV in Case IV is:

Summary
Four cases of the locations of pump/motor leakages were investigated. The different implementation variants were determined by the changes in the volume of the cylinder as a function of the volumetric efficiency, and the results are shown in Figure 8. As a result, Case IV appears to be the most independent case [26]. However, this method did not investigate the pressure built up with various payloads. Neither did this investigation consider the practical aspects of realizing such a system. Therefore, the system structure will be investigated further in Matlab Simulink. Figure 8. Results of the DDH system with four cases of external leakage connections [26].

Simulink modeling
Matlab Simulink R2015a and Simscape SimHydraulics and the SimPowerSystems component library were used for creating the simulation model. The following assumptions were made: • The effect of temperature on fluid properties is neglected.
• External leakages in the hydraulic pump/motor are neglected. Volumetric efficiency is assumed to be 1.
• Losses in the hydraulic accumulator are ignored.
The simulation model for the test setup (introduced in Figure 2) is shown in Figure 9.   The movement of the cylinder is smooth and identical to the first prototype demonstrated in Figure 5. The flows in volume A (the line with accumulator A), B (the line with accumulator B) and C (line between the pump/motor P 1 and the cylinder) are illustrated in Figure 11. Flows B and C are mirrors of each other, which corresponds to correct behavior. Conversely, flow A has an initial drop which corresponds to sucking in oil from accumulator A during the initial lifting motion.   The system pressure is illustrated in Figure 13, where p 1 is the P 1 pump/motor pressure, p 3 is the P 2 pump/motor pressure, and p 2 , p 4 are the "tank" line pressures (line with accumulator A).

Figure 13.
Simulation results of the pressure in the DDH system: p 1 is the P 1 pump/motor pressure, p 3 is the P 2 pump/ motor pressure, and p 2 , p 4 are the "tank" line pressures (accumulator A).
The pressure p 3 ranges between 0.93 MPa and 1 MPa during the whole lifting-lowering cycle with a maximum payload of 150 kg. The tank pressure (p 2 and p 4 ) varies between 0.38 MPa and 0.18 MPa. Thus, according to the simulation, Cases II, III and IV do not fulfill the requirements of the leakage line where the maximum allowed constant pressure is 0.3 MPa and in the short term, it is 1 MPa. According to the results shown in Figure 13, from the proposed connection cases in Figure 7, only Case I can be used for the realization of the DDH setup, where both external leakage lines are connected to line A (accumulator A).

Experimental investigation
The essential parameters, such as currents, voltage, pressure, flow and height, were measured from the DDH setup. The location of the utilized sensors is illustrated in Figure 4. The experimental efficiencies and energies of the DDH test setup system for lifting movements were calculated as shown below: where E pot is the potential energy of the payload where m is the mass of the payload in kg and g is a gravitational constant in m/s 2 . H is the position of the cylinder piston in m.
E mech is the energy of the shaft and is calculated as the integral of the power at the shaft (P shaft ).
where P shaft is the output energy of the shaft in W, T is the motor torque in Nm and Ω is the angular speed in rad/s. A motor control algorithm was utilized to measure the angular speed and estimate the motor torque.
The output energy of the hydraulic part E hydr is calculated as the integral of the output hydraulic power: where p is the pressure in Pa, v c is the velocity of the cylinder piston in m/s and A is the cross area of the cylinder piston in m 2 .
Depending on the operating point of a hydraulic pump/motor unit in its performance curve, the relationship between the flow and hydraulic losses in a system varies significantly. During the lifting, the hydraulic pump/motor unit operates as a pump. The input energy of the pump is mechanical energy and the output is hydraulic energy. Figure 14 presents an example of a Sankey diagram for the measured losses and efficiencies of the DDH system.
In Figure 14, hydro-mechanical losses equal 33.4%. Hydro-mechanical losses include shaft and hydraulic losses. Hydraulic losses in DDH systems are composed of pipe friction losses and other fittings, entrance and exit losses and losses from changes in the pipe size resulting from a reduction in the diameter, pump/motors and cylinder losses. The overall cylinder efficiency is mostly dependent on the frictional losses encountered by the piston and the rod during its stroke. Frictional losses depend on the pressure difference across the seal, sliding velocity, seal material, temperature, time, wear and direction of the movement. Measured electrical machine losses are 20.7%. Electrical machine losses are composed of the following elements: stator and rotor resistive losses, iron losses, additional losses and mechanical losses. Mechanical losses include friction in the motor bearings. Bearing losses depend on the shaft speed, bearing type, properties of the lubricants and the load. The converter losses take place mostly in the semiconductor switches and in the auxiliary power systems. In this study, the losses of the frequency converter are not included.
It was challenging to create the DDH test setup due to the asymmetrical double-acting cylinder used and the difficulty of finding matching displacements of the pump/motors in order to fulfill R Q = R A . Moreover, another challenge was the location of the hydraulic accumulators and external leakage pump/motor lines. The Sankey diagrams showed that electrical machine losses are in second place after the dominant hydro-mechanical losses. Taking pressure balancing into consideration, the hydro-mechanical losses in the system will be reduced in the second DDH prototype.

Conclusion
This paper described a DDH setup and its potential for applicability to NRMMs. It investigates the compensation of pump/motor displacement for an asymmetrical double-acting cylinder and the location of hydraulic accumulators and external leakage pump/motor lines. The review was carried out for four different alternative cases. The mathematical model used in the review suggested Case IV as the least sensitive case. Further investigations were performed in Matlab Simulink. According to the simulations, Cases II, III and IV did not fulfill the requirements of a leakage line where the maximum allowed constant pressure was 0.3 MPa and, in the short term, it was 1 MPa. Therefore, only Case I can be used for the realization of the DDH setup, in which both external leakage lines are connected to line A (accumulator A). The experimental tests demonstrated that direct-driven hydraulics (DDH) has the advantage of a fully self-contained electrohydraulic actuator, which combines the high power density of hydraulics and the accuracy of an electric motor. The measured energy efficiency of the DDH varies by up to 46% with the direction of the motion of the cylinder and the motor speed. As the Sankey diagram showed, the hydro-mechanical losses dominate in the original DDH setup. As expected, with regard to the efficiency of the DDH setup, the weak link in the chain is found in the losses of the hydro-mechanical components of the system. Therefore, further studies are required on the improvement of DDH.