Variants of Secondary Control with Power Recovery for ...

CHINESE JOURNAL OF MECHANICAL ENGINEERING Vol. 28,aNo. 3,a2015

·618·

DOI: 10.3901/CJME.2015.0408.038, available online at www.springerlink.com; www.cjmenet.com; www.cjme.com.cn

Variants of Secondary Control with Power Recovery for Loading Hydraulic Driving Device

LI Wanguo*, FU Yongling, CHEN Juan, and QI Xiaoye

School of Mechanical Engineering and Automation, Beihang University, Beijing 100191, China

Received October 12, 2014; revised January 12, 2015; accepted April 8, 2015

Abstract: Current high power load simulators are generally incapable of obtaining both high loading performance and high energy

efficiency. Simulators with high energy efficiency are used to simulate static-state load, and those with high dynamic performance

typically have low energy efficiency. In this paper, the variants of secondary control (VSC) with power recovery are developed to solve

this problem for loading hydraulic driving devices that operate under variable pressure, unlike classical secondary control (CSC) that

operates in constant pressure network. Hydrostatic secondary control units are used as the loading components, by which the absorbed

mechanical power from the tested device is converted into hydraulic power and then fed back into the tested system through 4 types of

feedback passages (FPs). The loading subsystem can operate in constant pressure network, controlled variable pressure network, or the

same variable pressure network as that of the tested device by using different FPs. The 4 types of systems are defined, and their key

techniques are analyzed, including work principle, simulating the work state of original tested device, static operation points, loading

performance, energy efficiency, and control strategy, etc. The important technical merits of the 4 schemes are compared, and 3 of the

schemes are selected, designed, simulated using AMESim and evaluated. The researching results show that the investigated systems can

simulate the given loads effectively, realize the work conditions of the tested device, and furthermore attain a high power recovery

efficiency that ranges from 0.54 to 0.85, even though the 3 schemes have different loading performances and energy efficiencies. This

paper proposes several loading schemes that can achieve both high dynamic performance and high power recovery efficiency.

Keywords: load simulator, variants of secondary control, power recovery efficiency, energy regeneration, hydraulic driving device,

simulation, AMESim

1 Introduction

A load simulator, which is also known as a loading system, is used to test a device under a given load of force or torque. The rotation of the tested object drives the rotation of a corresponding load simulator, which absorbs all of the mechanical power of the tested object while simultaneously applying a torque load to it. These types of load simulators commonly use power dissipation regulators, which include hydraulic valves with throttle regulation[1], hydraulic dynamometers, electric eddy current dynamometers, etc. Throttle regulation is also used in the current electro-hydraulic servo loading technique, which exhibits a superior dynamic performance (but actually the lowest energy efficiency)[2–3]. All of the power in these load simulators is dissipated into waste heat, which results in a large amount of wasted energy and heavy cooling requirements. Consequently, power recovery and reuse should be implemented in high power load simulators.

In the fields closely related to mechanical engineering, Energy recovery and reuse have been widely used in

* Corresponding author. E-mail: [email protected]

© Chinese Mechanical Engineering Society and Springer-Verlag Berlin Heidelberg 2015

applications, such as braking energy recovery in electric rail vehicles[4–5] and gravitational potential energy recovery in electric elevators[6] and forklifts[7]. These recovery and reuse methods are also used in excavator[8–9], drilling rig[10], electric vehicles[11], hybrid vehicles[12–13], etc. Hydraulic transmission can be used as an effective means for energy recovery. A closed-loop hydraulic energy-regenerating system was proposed and proved effective[14]; hydraulic system is also used to develop mechanical roadway which can capture and recover waste energy of vehicles[15]; and moreover hydrostatic secondary control system can be used in hybrid vehicles to increase energy efficiency[16].

Power recovery and reuse have also been frequently implemented in load simulators; however, it is difficult to attain both good dynamic performance and high energy efficiency, as with the aforementioned electro-hydraulic servo load simulator, for example. The electric servo loading system that is generally used for low power tested objects exhibits good dynamic performance but its energy efficiency has generally not been studied[17]. An electric dynamometer can provide a large torque and feed the absorbed power back into the electric network[18]. This device can therefore be used in a high power loading test, usually for a static or low response dynamic loading demand in which the dynamic performance has not been

CHINESE JOURNAL OF MECHANICAL ENGINEERING

·619·

confirmed. In a test bench for a hydraulic pump and motor, the pump and the motor can be used for loading and tested elements each other[19–20]; power recovery and reuse can also be implemented through the hydraulic line or the mechanical line, but high energy efficiency is generally obtained for static loading operation. Hydrostatic secondary control exhibits good dynamic performance and promotes power recovery and can therefore be used in a loading system with power recovery[21]. Within this scheme, secondary units can be used as loading elements to absorb and transmit mechanical power from the tested hydraulic driving device, convert the form of the power from mechanical→hydraulic→mechanical→electric, and finally feed the power into the electric network. However, an excessive number of energy conversion processes reduce the power recovery efficiency. The most typical example of secondary control applied to a loading system is the assembly of loading and tested elements in a hydraulic network under constant pressure[22], which can produce both better dynamical performance and high energy efficiency because the recovered power can be fed directly back to the tested elements in the hydraulic line; however, this scheme can only be applied to tested or driving devices that operate at constant pressure. Many hydraulic-driven devices operate under a variable pressure following a variable load, such as a hydraulic-driven chassis or the driving head of vehicles and construction machineries. For these devices, the load simulator can also feed the recovered power through the hydraulic line, but a pressure valve or another unit is needed between the tested and loading elements. We performed a preliminary design of such a system, which was initially named the variable pressure secondary control (a term that has since become inappropriate as stated in section 2.3); however, poor simulation results were obtained because of the low performance of the designed pressure valve and the control strategy[23]. In this study, we develop variants of hydrostatic secondary control that improves the design scheme and results in much better performance, and we also consider and evaluate more suitable schemes.

2 Load Simulator and Variants of Secondary Control

2.1 Tested objects In this study, the tested or loaded objects are limited to

hydraulic driving devices that operate under variable pressure. These devices include hydraulic motors with DA (speed-related), HA2 (pressure-related) or other controls; mechanical transmission mechanisms such as a gearbox, etc; and corresponding hydraulic pumps that use DA, LR2 (constant power control) or other control schemes. Under these control modes, these devices generally operate at lower (higher) pressures with higher (lower) flow rates and output higher (lower) speeds and lower (higher) torques; thus, the engine can work in a high efficiency regime, and

even though the operating pressure varies following the loads, the range of variation is not wide. It is not necessary to study all of the tested objects to explore every combination of the aforementioned control modes: it is sufficient to study only one mode. As the preliminary research reported in Ref. [23], the hydraulic driving head of the rotary drilling rig is selected for the tested device, for which LR2 control is used for the pump, and HA2 control is used for the motor.

The abovementioned control modes of the variable delivery pump and motor are introduced briefly as follows. DA is mainly used for the pump but can also be used for the motor. The pump DA is different from the motor DA, but they can work together. The motor DA must work with the pump DA, but the pump DA can also work with the motors of other control modes. The pump DA controls the pump displacement q1, depending on the pump speed n, supplied by the engine, which is also related to the operating pressure p. q1 increases when n increases but decreases when p increases. The motor DA controls motor displacement q2, by introducing n, which is converted into a pressure signal by the DA pump. q2 decreases when n increases. q2 is also related to the operating pressure, p. If n is constant, q2 increases when p increases. In this working mode, q1 and p increase simultaneously, but q2 decreases when n increases and the load torque of the motor remains constant. Details of DA are discussed in Ref. [24] and Ref. [25]. LR2 controls pump displacement q, that is a function of the operating pressure p, so that a specified drive power is not exceeded with a constant drive speed, which can be stated as Constant.p q = Details of this factor are discussed in Ref. [26]. HA2 controls the motor displacement q, according to the working pressure p, and makes q proportional to p. In this mode, q can be regulated from 0 to qmax when p increases 10 MPa. Details on these factors are discussed in section 3.2 and in Ref. [27].

2.2 Principle of the load simulator

Fig. 1 shows the principle on which the load simulator operates. As a control mode of the pump 2, HS1 controls pump displacement with the use of a built-in servo variable displacement device. The pump swivel position can be fed back through a positional transducer. DS1 is a type of servo variable displacement control for a secondary unit that can work both as a motor and a pump. The matching controller can control the speed and torque of the secondary unit. Details on HS1and DS1 are discussed in section 4, as well as in Ref. [28] and Ref. [29].

The arrangements of the gear increaser (8) and the secondary unit (9) are shown at the right portion of Fig. 1. The arrangements of the gear reducer (6) and motor (5) are similar. MS in Fig. 1 is the main shaft.

High energy efficiency and better performance are obtained for this simulator using secondary control units which are selected as the loading elements. Classical hydrostatic secondary control (CSC) uses a constant

Y LI Wanguo, et al: Variants of Secondary Control with Power Recovery for Loading Hydraulic Driving Device

·620·

pressure network in which the secondary units work in parallel and the primary units hold the constant pressure network together using accumulators. CSC cannot be used for tested devices that operate under variable pressure, and thus, variants of secondary control are used to meet the loading demands.

The tested device consists of 3 hydraulic motors (5) with HA2 control that drive in parallel a gear reducer (6), which can output a lower speed with a higher torque on the output shaft of the gear reducer (6) known as main output shaft. The loading subsystem consists of a torque and speed sensor (7), a gear increaser (8), 4 secondary units (9) and a power feedback passage (FP). The secondary units (9) are central loading components consisting of axial piston pumps/motors that function as variable displacement pumps and can create torque by regulating their swash plate angles under certain operating pressure: the best performance can be obtained under a constant pressure. However, the rapid regulating ability of the secondary unit produces good performance even under a variable pressure

as long as the variations are not severe. Power in mechanical form is absorbed from the tested device by the secondary units (9) and is fed in hydraulic form back into the hydraulic network of the tested device through the feedback passage (FP) that separates the tested device and the loading subsystem and enable them work under different pressures: the FP is the key component in variants of secondary control (VSC). Because of the transmission power loss, the feedback power must be less than the driving power required by the tested device: thus, a supplementary power subsystem is required, which serves as the primary external power supply of the entire system and consists of a variable displacement pump (2) and an electric motor (1). For the hydraulic motors and secondary units to operate at high speeds, the inlets of these units must be maintained at certain pressure by a low pressure hydraulic source (4). The rated speed of the selected secondary units is also lower than that of the hydraulic motors; thus, the number and delivery of the secondary units (9) are larger than those of the hydraulic motors (5).

Fig. 1. Operating principle of hydraulic system

Ⅰ. Supplementary power subsystem;Ⅱ. Hydraulic-driving device; Ⅲ. Loading subsystem; 1. Supplementary electric motor; 2. Supplementary power pump; 3. Hydraulic source for control mechanisms; 4. Low pressure source; 5. Hydraulic motor; 6. Gear reducer; 7. Torque and speed sensor; 8. Gear increaser; 9. Secondary unit; 10. Proportional/servo pressure valve;

11. Pressure sensor; 12. Accumulator; 13. Hydraulic transformer; 14. Pressure vessel In this system, the hydraulic motors and secondary units

operate in one direction, and the load torque is also applied in a fixed direction, resulting in power flow in one direction, as shown in Fig. 1. Despite this configuration, the load torque can still vary at higher frequencies.

In this study, 4 types of feedback passage are presented. The first type of FP is referred to as a hydraulic transformer (HT) because a hydraulic transformer (13) is used together with an accumulator (12) for constant pressure control, which creates a constant pressure network in the loading subsystem. The second type of FP is known as a constant pressure valve (CP) and consists of a proportional/servo pressure valve (10) and an accumulator (12), which are also controlled to create a constant pressure network in the loading subsystem. The third type of FP is known as a

variable pressure valve (VP) and is similar to the second type of FP, except that a pressure vessel (14) is used rather than an accumulator to maintain variable pressure control following a loaded torque to reduce the energy loss from the pressure valve. The fourth type of FP is known as a no pressure valve (NP) and consists of only an oil line and a pressure vessel (14) without a pressure valve. The tested device and the loading subsystem are connected by the NP and operate in a single pressure network created by the two subsystems. These 4 systems are validated later in the paper.

2.3 Variants of hydrostatic secondary control

The FP and the secondary units can create a pressure network, such that the secondary units are coupled to the


·621·

pressure and work independently of each other as well as directly regulating the load torque: thus, this system, in a manner, corresponds to hydrostatic secondary control. However, this loading system is not classified under classical secondary control because the FP is not a typical primary unit, such as a constant pressure variable displacement pump, but functions as a pressure regulating mechanism, i.e., a hydraulic transformer (HT), a proportional/servo pressure valve (PV) or even an oil line. Thus, these types of systems are known as variants of a hydrostatic secondary control system (VSCS) and are classified as HTVSCS, CPVSCS, VPVSCS and NPVSCS, depending on which of the 4 types of FPs are being used, i.e., HT, CP, VP and NP, respectively; the corresponding regulator modes are known as variants of hydrostatic secondary control (VSC) and are classified as HTVSC, CPVSC, VPVSC and NPVSC.

In view of the abovementioned reason and that VPVSCS works under variable pressure, which is also different from the classical secondary control., VPVSCS should be more appropriately called variant of secondary control rather than variable pressure secondary control.

3 Analysis of the System

3.1 Simulating the work state of the tested device To reuse the feedback power, the original hydraulic

source for the tested device, which operates with a constant power control (LR2), is eliminated and replaced by a supplementary power subsystem such that the hydraulic motors (5) are driven by the 2 power flows from the feedback passage FP and the supplementary pump (2). The LR2 operates by regulating the delivery of the supplementary pump (2) with an approximately constant rotating speed input from the electric motor (1). However, the LR2 is originally designed to maintain a constant output power for the hydraulic source: to realize this operation under the work condition, the pressure and overall flow rate of the hydraulic subsystem must be measured in the tested device, resulting in a complex system. For simplicity, the torque and rotating speed signal from the sensor (7) on the main output shaft can be used as a feedback signal to form a speed closed-loop control subsystem, while the value of the desired speed is calculated from the constant power demand. However, the power on the main output shaft of the tested device is maintained constant rather than that of the driving hydraulic power. This inconformity with the original work state has little effect on the test results; thus, this design can be applied in practice.

3.2 Main components and static operation points

The secondary units (9) supplied by Rexroth use DS1 control, which produces the best regulating performance of the loading torque. The supplementary power pump (2) uses HS1 control, which is also a servo variable displacement control, and the corresponding driving

electric motor (1) is a common induction motor with an approximately constant speed. The hydraulic motors (5), as the main components of the tested device, generally use HA2 control. The hydraulic source (3) for the control mechanism uses a constant pressure variable displacement pump. The main parameters are given below.

The rated power on the main output shaft Pc is 270 kW. The maximum torque on the main output shaft T is 367 kN • m; the rotating speed of the main output shaft n ranges from 7 to 22.8 r/min; the delivery of the hydraulic motor (5)

mq ranges from 80 to 160 mL/r; the delivery of the secondary units (9) Lq ranges from 0 to 250 mL/r; the delivery of the supplementary power pump (2), sq , ranges from 0 to 500 mL/r; the maximum rotating speed of hydraulic motor (5) ranges from 3100 to 4900 r/min (speed-related); the maximum rotating speed of the secondary units (9) range from 1500 to 1800 r/min (speed-related); the rated rotating speed of the supplementary pump (2) is 1500 r/min; the operating pressure of the tested device, mp ≤28 MPa; the operating pressure of the loading subsystem, Lp ≤31.5 MPa; the gear ratio of the gear reducer, mi , is 185/1; and the gear ratio of the gear increaser is 1/86.

The proportional/servo pressure valve (10) is a key component of the VPVSCS and the CPVSCS and can support a large flow rate, create a low pressure loss and exhibit a high dynamic performance. The D665 servo-proportional control valve supplied by MOOG is compatible with this application, which has a rated flow rate over 1000 L/min under a pressure loss of 0.5 MPa per land and a response time of 10 ms. A pressure closed-loop can be constructed using this valve and a pressure sensor.

Maintaining a constant set of operating parameters results in a constant load torque; thus, a set of static operating points can be calculated from the working curve of the driving motor in the HA2 control mode and a rated constant power. The loaded torque Ts in N • m is the primary parameter, and the other parameters are functions of Ts including the rotary speed sn on the main output shaft in r/min, the pressure msp of the tested device in MPa, the delivery msq of the hydraulic motors (5) in mL/r, the pressure Lsp of the loading subsystem in MPa and the delivery Lsq of the secondary units (9) in mL/r, etc. In this study, the subscript s of the variables denotes a static value. For example, mp is a transient value during operation and

msp is a static value in static state. Thus, these two values are different concepts.

Evidently, For the NPVSCS, L m.p p» Where Lpdenotes the pressure of the loading subsystem, mp denotes the pressure of the tested device; for CPVSCS and VPVSCS, L m ;p p≥ but for HTVSCS there is no thus constraints. The value of Lp is set constant for the HTVSCS and the CPVSCS. For the VPVSCS, Lp should vary regularly with Ts, because msp varies regularly with Ts, Lsp is determined by adding approximately 0.5 to 1 MPa bar to the value of msp .


·622·

The working curve of the hydraulic motors (5) is shown in Fig. 2.

Fig. 2. Working curve of the hydraulic motors (5)

maxq and minq are the maximum and minimum actual delivery of the hydraulic motors (5), respectively; 0p , bp , and ep are the pressure values at a delivery of 0, minq , and

maxq , respectively. The system operates at a constant power in the region m16 MPa 26 MPap≤ ≤ , which is restricted to m21 MPa 26 MPap≤ ≤ when the delivery of the hydraulic motor (5) is restricted from 80 to 160 mL/r. Thus

msp can be calculated out by

ms m ms b

160,

( ) 80,

80,

q k p p

éêê= - +êêë

ms

ms

ms

26,

21 26,

21,

p

p

p

>

<

≤ ≤ (1)

where km is the slope of the working curve of the hydraulic motor, as shown in Fig. 2, and

max max minm

e 0 e b

016 mL r MPa.

q q qk

p p p p

- -= = = / /

- -

In this example, all of the hydraulic motors (5) and their gear ratios to the main shaft are identical; the same is true for the secondary units (9). Under an ideal static state, if the energy efficiencies of all transmission segments are equal, then the following equation is derived:

ms lo Vms s s Ls lo VLs c( ) ( ),

0.06 30 π 0.06

p p q T n p p q P

M M M M

- -= = =

/

(2)

where lop is the pressure of the low pressure line in MPa, which is approximately constant at 1.5 MPa; Vmsq is the total flow rate of the driving hydraulic subsystem in L/min;

VLsq is the total flow rate of the loading subsystem in L/min; M is the number of driving hydraulic motors at 3; and Pc is the rated output power of the driving subsystem in W, which is constant under the constant power control mode at 270 kW. Thus:

m s msVms ,

1000

i M n qq =

(3)

where im is the total transmission ratio of the gear reducer

(6) at 185:1. Combining and solving Eqs. (1)–(3) yields the following:

ms p s

sms

m

0 lo

20 0 lo s ms

m m

msm

( )

π, 26,

80

1[( )

20

80( ) 4 , 21 26,

π, 21.

40

lo

p f T

Tp

i M

p p

p p p p T pi k M

Tp

i M

= =

ìïï >ïïïïïïï + ïïïïí ùïï ú+ - +ï úïï úûïïïïï <ïïïî

≤ ≤

(4) Considering the mechanical efficiencies of hydraulic

motors (5) and gear reducer (6), Eq. (4) can be rewritten as follows:

ms pm smm gr

sms

m

0 lo

20 0 lo s ms

m m

msm

1( )

π, 26,

80

1[( )

20

80π( ) 4

, 21 26,

π, 21,

4

0

lo

p f T

Tp

i M

p p

p p p p T pi k M

Tp

i M

= = ´

ìïï >ïïïïïïï + ïïïïí ùïï ú+ - +ï úïï úûïïïïï <ïïïî

≤ ≤

(5) where mm is the mechanical efficiency of the hydraulic motors (5); gr is the mechanical efficiency of the gear reducer (6). Thus Lsp can be determined by

Ls pm s ms

ms

28, for HTVSCS and CPVSCS,

( ) (0.5 1), for VPVSCS,

, for NPVSCS.

p f T p

p

éêê= = + ~êê»êë

(6)

In addition, other equations as follows are easily

obtained:

cs n s

s

30( ) ,

π

Pn f T

T= = (7)

s gi gi mpps q s

p Ls lo

2π( ) ,

( )

T iq f T

M p p

= =

- (8)

where gi is the transmission efficiency of the gear increaser (8); mp is the mechanical efficiency of the secondary units (9); igi is the gear ratio of the gear increaser (8) at 1/86; and Mp is the number of the secondary units at 4.

The expressions for the other variables are omitted here.


·623·

Note that Eqs. (1)–(8) only apply to the static state. In this study, the tested driving head of the rotary drilling

rig is provided by a manufacturer, in which the hydraulic motors with HA2 control are used as the driving elements. HA2 control can quickly increase the driving torque when the load increases and decrease the rotary speed so that the driving power remains approximately constant. This control mode fits the drilling work and provides the best engine output. The minimum delivery of motor (5) qmin is set at a value of more than 0 (e.g., 0.5qmax,) to prevent exceeding speed under a slight load because of the limited flow rate and sufficient preliminary torque.

Note that Pc is the rated power of the tested driving head given by its manufacturer. Under the constant power mode, Pc is constant, and the given main shaft rotating speed, ne, is calculated from the given Pc and Te by Eq. (7). n is obtained by control on the pump (2) and the motors (5) during operation. In the system shown in Fig. 1, Pc is obtained by controlling T and n, which attempt to make T and n consistent with Eq. (7).

3.3 Accumulator and pressure vessel

The accumulator in CPVSCS is of bladder-type and is used to smooth the loading subsystem pressure L.p The charging-discharging process of the accumulator is considered an adiabatic process because it runs fast. The volume of the accumulator can then be calculated by

0 1/ 1/ 1/a0 1 2( )k k k

VV

p p p- -=

-

,

where V0 is the volume of the accumulator in m3; V is the effective volume in m3; pa0 is the charged gas pressure in Pa; p1 is the minimum operating pressure in Pa; and p2 is the maximum operating pressure in Pa, k is the adiabatic index at k=1.4. p2 and p1 are determined on the basis of the control accuracy of T. The set operating pressure is 28 MPa, and the pressure error is limited to 0.5 MPa based on experience, therefore p2 and p1 are 28.5 and 27.5 MPa, respectively. The charged gas pressure pa0 is determined as 0.85p1 (i.e., 23.4 MPa). The key parameter is V, which is dependent on all dynamic processes, including the dynamic regulations of n, T, and pL. Thus, some important data required by general static-state calculations cannot be estimated with good accuracy without dynamic simulations or tests. Rough estimates results in large errors. Thus, simulation is the most effective method. Simulations show that under the most severe working conditions, the rise step of given load torque is 175 kN • m and the maximum pressure error is 0.5 MPa when V0=0.015 m3; when V0=0.04 m3, the error is 0.32 MPa; when V0=0.063 m3, the error is 0.25 MPa; and when V0=0.080 m3, the error is 0.22 MPa. A rough estimate was performed, and the results showed that under a pressure error of 0.5 MPa, V=0.001 2 m3 and V0=0.054 m3, which are more than thrice the simulation

result under the same pressure error. The calculation process is omitted at this point.

Determining the volume of the pressure vessel (14) is highly dependent on all dynamic processes, implying that simulation is still the best approach. In the simulations, the volume of the vessel (14) is regulated under different operation frequency zones; thus, good performance is obtained. In practice, it should be ranked several grades of volume depending on the working frequency zones because the oversize vessel may restrain the fast regulation of pressure. In general, the vessel formed by the pressure lines has smaller volumes than the requirement; therefore, a number of independent vessels should be mounted near the high pressure ports of the secondary units (9). The structure is shown in Fig. 3. The three pressure vessels with different volumes can provide four different volumes through four different combinations by switching on or off the electromagnetic directional valves; this number of volumes is enough for adapting to different operating frequencies. The loading subsystem can be equipped with only one centralized set of pressure vessels to simplify the structure.

Fig. 3. Pressure vessels (14)

1, 2, 3. Pressure vessels with different volumes; 4. Electromagnetic directional valve

3.4 Loading performance and power recovery

efficiency The CPVSCS exhibits the best dynamic and static

performance because the approximately constant pressure causes less interference in the loading regulator than a variable pressure does. The CP performance is better than that of the HT under the aforementioned work condition of this system although both FPs use a constant pressure network for correctly set work parameters, because the proportional/servo valve of the CP exhibits better dynamic performance as a control valve than the hydraulic transformer which is actually a set of complicated transmission systems involving moving parts with large inertia. Effective control of the VP results in a better loading performance than that of the NP, because the VP can partially regulate the regular variation in the pressure, whereas the NP cannot regulate the pressure. The loading performance can be used to rank the 4 types of VSCSs from high to low as CPVSCS → HTVSCS → VPVSCS →

NPVSCS: this rank can change slightly for low frequencies as we will show later in the simulation results. Certainly, different results will be obtained if the parameters of the


·624·

system are set inappropriately or the control is ineffective. The power recovery efficiency is the ratio of the

feedback power to the absorbed power of the loading subsystem and can be defined for a constant load as follows:

f

ob in

,P

P P =

+ (9)

where is the power recovery efficiency; fP is the feedback power from the loading subsystem; obP is the power absorbed by the loading subsystem from the tested device; inP is the power externally supplied to control the regulating mechanisms of the loading subsystem.

Fig. 1 shows this system:

f bP P= , ob c dP P P= + , where bP , cP , and dP denote the power flow through points b, c, and d, respectively. Thus,

in

.b

c d

P

P P P =

+ + (10)

Eq. (9) indicates that the power recovery efficiency

also varies dynamically under a dynamically variable load because the transient power at every point varies dynamically, making it difficult to evaluate the efficiency characteristic. In this case, the power recovery efficiency can be defined as follows:

f

ob in

,E

E E =

+ (11)

in

,b

c d

E

E E E =

+ + (12)

where

2

1d ,

t

i it

E P t= ò

the subscript i corresponds to f, ob, in, b, c, and d in Eqs. (11) and (12). The time from t1 to t2 lies within a stable run period and for periodic operation the span from t1 to t2 should equal integer multiple of a cycle. Of course, this definition also applies for a constant load. Ei is energy that the respective transmission element or unit transmits or converts, and this energy is accumulated over a period of time. Thus, unlike power, Pi, this value is not an instantaneous quantity, but an integral quantity of Pi.

The recovery efficiency can also be defined using the following Eq. (13) or other equations that are not discussed in this study:

f inan

ob

.P P

P

-= (13)

It can easily be seen that the power recovery efficiency increases with the transmission efficiency of the gear reducer (8), the secondary units (9), the hydraulic oil lines of the loading subsystem and the FP, and as the power for the control mechanisms inP decreases. Different are obtained for the 4 types of VSC loading subsystem because of the different FPs. Different FPs produce different inP values; however, these differences in inP have a small effect and therefore, the FP transmission efficiency (which is defined below) should primarily be considered to compare the values of the 4 different systems:

FPct

,b

a

P

P P =

+ (14)

FPct

,b

a

E

E E =

+ (15)

where aP is the power flowing through point a at the FP inlet; ctP is the externally supplied FP control power; and E is defined in Eqs. (11) and (12).

The NP is an oil line, which exhibits very low power losses and thus has the highest transmission efficiency. The VP and CP inevitably produce throttling energy losses, where the VP transmission efficiency is higher than that of the CP, because the VP can control the pressure of the loading subsystem following the load to maintain a lower pressure drop through the pressure valve than that for the CP. In addition, the pressure of the tested device varies within a narrow range during normal operation; therefore, the pressure drop through the pressure valve (10) is not high, and the CP transmission efficiency remains high. There is no such throttling power loss with the hydraulic transformer as that with pressure valve, However, the hydraulic transformer is still being developed for use as a transmission element[30] and has an overall transmission efficiency that is no higher than 80% and below 75% in most cases. Under the working condition of this system, the HT efficiency is lower than that of the other 3 FP types, as the simulation results will later show. The 4 types of VSCSs can be ranked from high to low in terms of the power recovery efficiency as follows: NPVSCS → VPVSCS→CPVSCS→HTVSCS.

In addition to the low energy efficiency, a hydraulic transformer has a much higher volume, weight and price than a proportional/servo valve for the same rated power capacity, and consequently, the hydraulic transformer is not selected for further study.

3.5 Control strategy

All of CPVSCS, VPVSCS and NPVSCS are multivariable systems. The VPVSCS and CPVSCS have 3 controlled variants, T, Lp and n, and the NPVSCS has 2 variants, T and n. There are 6 executive elements in the VPVSCS and CPVSCS and 5 executive elements in the NPVSCS, among which 4 displacement regulating


·625·

mechanisms of 4 secondary units (9) operate together to control T; under an approximate constant driving rotary speed of the electric motor (10), the displacement regulating mechanism of supplementary power pump (2) controls n. This condition coordinates the feedback hydraulic flow from the loading subsystem. The speed, n, signal is provided by the torque and speed sensor (7). The proportional/pressure valve (10) controls Lp in CPVSCS and VPVSCS. The expected values of ne, Lep , and the feedforward value, qLp, are determined by Eqs. (6)–(8), which have the common independent variable Te. Note that

Lep is determined by Eqs. (6) and (5), so its value is related to the mechanical efficiencies of the hydraulic motors (5) and gear reducer (6). These efficiencies are also variable under different working conditions, thus their values cannot be accurately determined. However, they are provided with constant values in the following simulations, which results in an error to Lep and even the control of

Lp . This condition does not affect CPVSCS as long as the constant Lep is set high enough. Therefore, it mainly affects VPVSCS, and the following simulations show that the control performance of Lp is not ideal even though

Lep is accurate. Considering that the efficiencies are set at typical values, the error of Lep brings less trouble than the control error of Lp . Lp is also not the control objective in this project, and controlling Lp in VPVSCS increases the power feedback efficiency, and moreover the efficiency is truly increased. The errors of Lep and Lp also do not bring serious effects on the control of T and n because of the superior performance of the secondary control and speed control system property, which is also true to NPVSCS. Consequently, the errors of the mechanical efficiencies of hydraulic motors and gear reducer do not extensively harm the system.

As for the control of n, because of the transmission energy loss and L mp p> , the feedback flow rate must be less than the requirement. Thus, decreasing the delivery or flow rate of the pump (2) can generally decrease n and vice versa, although the variation of T and pm may result in interference through the HA2 control of the motors (5). Therefore, n can be effectively controlled.

These variables couple and interfere with each other, creating difficulties in control. Fig. 4 shows the control strategy.

Fig. 4. Control strategy

eT — Expected T; Lep — Expected Lp ; en — Expected n; Lpq — Feedforward value of qL The torque load is exerted by 4 secondary units (9)

together that play the same role; thus, the same control is applied to all of the 4 secondary units to obtain the desired T. A double closed-loop and a feedforward element are used to obtain accurate control of T, where the inner loop controls the displacements of the secondary units (9), the outer loop controls T. The control performance of T can be increased by calculating the displacements qpe in advance from eT and Lep and inputting these values into the inner loop, which is implemented by the feedforward element. The most serious interference for T is the severe variation in Lp , which is sensitive to mp . Introducing the pressure difference between Lep (or Lsp ) and Lp to compensate for the closed-loop rotating speed can smooth

mp and even Lp and improve the control for T, which is

validated using simulations later in the paper. For the VPVSCS and NPVSCS, a pressure vessel of a certain volume can restrain high frequency fluctuations in Lp and improve the control for T; this effect is exploited by installing 4 pressure vessels near the outlets of the secondary units (9).

4 Simulation Model

The CPVSCS, VPVSCS and NPVSCS systems are designed in detail and simulated by AMESim; the simulation model of the VPVSCS is shown in Figs. 5–7. In Fig. 5, the CPVSCS and NPVSCS models are similar to the VPVSCS model and are not shown here for space considerations. units 1 to 14 are the same as in Fig. 1; 2-1,


·626·

5-1, 9-1, 10-1 and the control system are “supercomponents” that are shown in Figs. 6 and 7; “integrator set 1” and “integrator set 2” are also “supercomponents” and are integrator sets that record the energy flowing through

certain points by integrating the corresponding power from the power meters; and E1 to E11 are power meters used to analyze the energy efficiency and are only used in the simulation.

Fig. 5. Overview of the AMESim simulation model for the VPVSCS

Fig. 6. Mechanisms for the “supercomponents” in Fig. 5 H-1—Hydraulic servo valve; H-2—Servo cylinder; SP1—Hard spring; SP2—Soft spring;

SE1—Displacement sensor; SE2—Displacement transducer


·627·

Fig. 7. Control system for the “supercomponent” in Fig. 5

Particular attention is paid to the modeling of key units such as the hydraulic motors (5), the secondary units (9), the main transmission shaft between (5) and (9), the servo-proportional valve (10), and the supplementary power pump (2) to realistically simulate their dynamic performance and energy expenditure. The variation in the rotating speed of the supplementary power pump (2) following the load is also taken into account, and therefore the electric motor (1) rather than a constant speed source is used to drive the pump (2).

Fig. 6(a) shows the HA2 control mechanism, which is a closed loop control subsystem. The hydraulic servo valve H_1 controls the servo cylinder H_2 by introducing the pressure of the hydraulic motors (5) in Fig. 4 and the feedback displacement signal of cylinder H-2 (which indicates the delivery of the hydraulic motors) by the soft spring SE1. The hard spring SP1 is used to set the open pressure, i.e., p0 in Fig. 2. The displacement transducer SE2 transforms the displacement of the servo cylinder H-2 into the delivery of the hydraulic motor.

The exact actual product structure does not need to be modeled in the simulation of the servo-proportional valve (10), only the relevant performance needs to be simulated. Fig. 6(b) shows the 2 units used, i.e., a 2-way servo valve and a valve-controlled cylinder (10-1): the first unit is used to simulate the output dynamic performance and the work state of the whole valve, and the second unit is used to simulate the energy expenditure of the preamplifier stage. The dynamic performance is neglected for the less important units, such as the relief valves, the pumps for the low pressure hydraulic source (4) and the hydraulic source for the control mechanism (3).

Figs. 6(c) and 6(d) show the variable displacement control mechanisms of the pump (2) and the secondary unit (9), respectively, which are all electro-hydraulic vale-control-cylinder displacement subsystems.

Fig. 7 shows the control system model. To simulate the actual performance of the secondary units (9) and the

supplementary power pump (2), the parameters of the proportional components inside their displacement closed-loops are adjusted using their actual performances.

Here, the performances of the key components and units are in agreement with that of the actual products; however, the efficiency of the transmission components and the units is set to a constant, which also indicates that the frictions on the main transmission processes are modeled with low accuracy; and the gear clearance and the lubricating power for the gear boxes are neglected.

The simulation range is the normal operating interval over which the delivery of the hydraulic motors (5) can vary from 80 to 160 mL/r following a loaded torque from 120 to 300 kN • m.

The primary simulation parameters are shown in Table 1.

Table 1. Primary simulation parameters

Parameter Value

Reducer (6) output shaft inertia Jma/(kg • m2) 15 000

Increaser (8) input shaft inertia Jpb/(kg • m2) 5000

Main shaft stiffness KSR/(N • m • (°)–1) 2.38´108

Pump (2) shaft inertia Ja/(kg • m2) 5.6

Motors (5) delivery settling time tsmc/ms 133

Secondary units (9) delivery settling time tsp/ms 55

Pump (2) delivery settling time tss/ms 212

Servo valve (10) natural frequency fNsv/Hz 70

Motors (5) mechanical efficiency mm 0.94

Motors (5) volumetric efficiency vm 0.95

Secondary units (9) mechanical efficiency mp 0.96

Secondary units (9) volumetric efficiency vp 0.94

Gear increaser (8) transmission efficiency gi 0.94

Gear reducer (6) transmission efficiency gr 0.93

Pump (3) overall efficiency pc 0.9

Electric motor E-MOTOR-2 efficiency emc 0.945

Gear reducer (6) transmission ratio 185/1

Gear increaser (8) transmission ratio 1/86

Note: a) Jm is the equivalent inertia, which equals 0.146 1 kg • m2 per shaft, where the shafts are those of the driving motor (5); b) Jp is the equivalent inertia, which is 0.169 0 kg • m2 per shaft, where the shafts are those of the secondary units (9); c) corresponding to a 1/2 stroke.


·628·

5 Results and Discussion

5.1 HA2 control The HA2 variable displacement control mechanism is a

relatively independent control subsystem, indicating that its control performance should be simulated. Therefore, a test system with the HA2 control mechanism is designed and simulated. Fig. 8 shows the test system. In this subsystem, the pressure of the motor (5) is the input signal, whereas the delivery of the motor (5) is the output signal.

Fig. 8. Test system of the HA2 control mechanism

TH_1-step element for pressure; TH_2-relief valve; TH_3-flow source; TH_4-torque source; TH_5-step element for torque

A flow source TH_3 other than a pressure source is used to prevent exceeding speed, with the relief valve TH_2 and step element TH_1 to set the pressure. The torque source TH_4 and step element TH5 provides sufficient load to form the pressure. The main parameters are shown in Table 2, and the simulation curves are shown in Fig. 9.

Table 2. Main parameters of HA2 mechanism

Parameter Value

H_1 valve diameter Dv/m 0.01 H_2 piston diameter Dp/m 0.042 H_2 piston rod diameter Dr/m 0.03 H_2 cylinder half an stroke L/m 0.06 SP1 hard spring stiffness kh/(kN • m–1) 6.545 SP2 soft spring stiffness ks/(kN • m–1) 250

Fig. 9(a) shows the given pressure values of pb and pc at 21 and 26 MPa, respectively. The corresponding relative delivery values should be 0.5 and 1, respectively and Fig. 9(b) shows good simulation results. Fig. 9(c) shows the step response from 0.5 to 1, with a settling time of 0.133 s.

5.2 Step responses

The same control parameters, except for the pressure loop, and the same structural parameters are used for the 3 types of systems, to compare their step responses which are shown in Fig. 10.

Ref-CPVSCS and Ref-VPVSCS in Fig. 10(b) are the reference values used to control Lp in the CPVSCS and the VPVSCS, respectively. As is shown in Fig. 10(a), the systems exhibit similar step responses for the loaded torque, for which the VPVSCS responses are closer to those for the

NPVSCS than to those of the CPVSCS. The given step value is also 175 kN • m, whereas the given final value is 300 kN • m for the increasing step. The time when the step response of T enters the span of 296.5 kN • m to 303.5 kN • m is the settling time by a 2% error band. A similar rule is observed for the decreasing step. Fig. 10(a) shows that the settling times of the 3 system range from 0.05 to 0.27 s. The three curves almost overlap each other for the static state.

Fig. 9. Operation curves of HA2

The pressure Lp of the CPVSCS is effectively

controlled for a static load, whereas those of the VPVSCS and the NPVSCS fluctuate slightly with a light static error. Note that in the NPVSCS there is no control valve to regulate the pL, which only varies following the load torque and the regulations of the hydraulic motors (5) and the supplementary pump (2). The 3 systems also exhibit similar speed responses as is shown in Fig. 10(c), and the CPVSCS exhibits the best performance because the approximately constant Lp produces less interference than for the other 2 systems. The speed responses are also obviously slower than the loaded torque because the transmission subsystem has a very high inertia.


·629·

Fig. 10. Step responses

Controlling speed is to maintain a constant main shaft output power P4 on the main output shaft, the response curve is shown in Fig. 11. Fig. 11(a) shows the given torques and Fig. 11(b) shows P4. It can be seen that P4 is approximately constant for different static loaded torques. However, P4 varies sharply when the loaded torque steps, because the speed has a slower step response than that of the loaded torque.

Fig. 11. Output power under different static loads

5.3 Sine wave responses 5.3.1 Loaded torque and pressure

As in the step response simulations, the same control parameters, except for the pressure loop, and almost the same structural parameters are used for the 3 types of systems for the simulations at the same frequency: for different frequencies, the parameters are adjusted to obtain the best response. Especially for the VPVSCS and NPVSCS, the volumes of the hydraulic vessels (14) decrease at higher frequencies because an oversized volume can weaken the pressure response.

The sine wave responses are simulated when provided with the sine signal with the same amplitudes of 90 kN • m at the corresponding frequencies of 1, 5, 10, 15, 20, 25, and 30 Hz. Fig. 12 shows the sine wave response curves for the torque and the associated transient errors, and Fig. 13 shows the pressure Lp curves for the load subsystem. The 3 types of systems exhibit good wave shapes overall, and the amplitudes decrease, while the dynamic errors and the phase lags increase with the frequency. The torque curves of 1, 25, and 30 Hz are presented, whereas the others are omitted considering the space. As shown in Fig. 12(a), the three response torque curves coincide approximately with the given torque, suggesting that they coincide with the given torque at 0 Hz and that the maximum frequency when the response amplitude is more than 63.7 kN • m (0.707 times given torque amplitude) is the -3 dB bandwidth. The maximum frequency when the lag phase is lower than -90° is the -90° bandwidth. In Figs. 12(b) and


·630·

12(c), at 25 and 30 Hz, the lowest amplitude is approximately 70.67 kN • m (-2.1 dB) and 56.79 kN • m (-4 dB), respectively. Consequently, the -3 dB bandwidths are between 25 and 30 Hz. The maximum phase lags are approximately -11° and -76°, which are all lower than -90°, so the -90° bandwidths are more than 30 Hz. The bandwidths are only estimated roughly, so the accurate calculation of frequency domain is unnecessary.

Fig. 12. Sine wave response curves for the loaded torque

and the associated dynamic errors

For the amplitude of 90 kN • m, the deliveries of the

secondary units (9) vary within a small span that is less than the maximum possible span. Consequently, the abovementioned high bandwidth can be achieved.

The rule for the dynamic loading performances of the 3 systems can be roughly summarized as follows: at low frequencies (<10 Hz), the transient errors are ranked from low to high, i.e., VPVSCS→NPVSCS→CPVSCS; at intermediate and high frequencies (10 Hz<frequency<30 Hz), the transient errors are ranked as CPVSCS→ VPVSCS→NPVSCS. At high frequencies (>25 Hz), the VPVSCS results are very close to the NPVSCS results. This behavior can be explained by considering Fig. 13.

Fig. 13. Loading subsystem pressure pL

In Fig. 13, Reference is the reference value used to

control pressure Lp in the VPVSCS and has the same phase as the given load torque. For the VPVSCS and the NPVSCS at low frequencies, as is shown in Fig. 13(a), the pressure Lp varies smoothly in the same direction as the given load torque, enhancing torque control; however, when


·631·

the frequency increases above a certain value (approximately 10 Hz), as is shown in Figs. 13(b) and 13(c), the variation in pL lags significantly behind the given load torque, weakening torque control. At low and intermediate frequencies, effective regulation of the pressure pL by the VPVSCS results in a higher performance than that of the NPVSCS; however, at high frequencies, the VPVSCS regulates pL poorly, resulting in similar performance to that of the NPVSCS. Fig. 13 also shows that pL is well-regulated by the CPVSCS but poorly regulated by the VPVSCS, i.e., constant pressure control is much easier than variable pressure control in the presence of strong interferences.

5.3.2 Rotating speed and output power

Fig. 14 shows the rotating speed under a sine loaded torque. As previously mentioned, the rotating speeds are controlled effectively under static loads; however, as is shown in Fig. 14(b), the speed responses under a sine load are poor, even at a frequency as low as 5 Hz, of course the effect is better at lower frequency as is shown in Fig. 14(a). This behavior can be primarily attributed to the high inertia of the transmission system and is similar to the performance of the main shaft output power shown in Fig. 15. Therefore, the output power is not constant under dynamically varying loads. Note that the given speeds are not sine waves but have the same frequencies and phases as the given sine load.

Fig. 14. Rotating speed under sine loads

Fig. 15. Transient output power under a 1 Hz load

5.4 Power recovery efficiency Figs. 16 and 17 show that the 3 systems have different

power recovery efficiencies under corresponding loads and working conditions, where the power recovery efficiency and the FP efficiency are defined in Eqs. (9)–(15).

Fig. 16. Efficiency under static loads Fig. 16(a) shows that the power recovery efficiency of

the 3 systems ranges from 0.54 to 0.85, which can be ranked from high to low as NPVSCS → VPVSCS →

CPVSCS, and the VPVSCS result is closer to that of the NPVSCS than to that of the CPVSCS, especially under static loads. Under increasing static loads, is approximately constant for the NPVSCS and the VPVSCS, but increases for the CPVSCS. This result can be attributed to the given approximately constant transmission efficiency of each process for the NPVSCS and the VPVSCS; however, the FP efficiency increases with the load for the


·632·

CPVSCS, as shown in Fig. 16(b), which can be explained by the pressure of the tested subsystem mp increasing with the load, whereas the pressure of the loading subsystem Lp is approximately constant, and the FP power loss decreases. Fig. 17(a) shows that under sine loads, decreases as the load frequency increases, this is because unlike with static loads, the control power for the loading subsystem increases as is shown in Fig. 17(c), and the FP efficiency varies slightly as is shown in Fig. 17(b). Thus, a high dynamic system performance is obtained at the cost of the control power or the energy efficiency. The control power is very low under static loads and can be neglected. In most cases, the efficiencies for the 3 FP types are above 0.8 and are higher than the HT efficiency.

Fig. 17. Efficiency and control power under sine loads

6 Conclusions and Future Work

Variants of secondary control with power recovery are

developed for high power loading hydraulic driving devices to obtain high loading performance along with high power recovery efficiency. After analyzing the technical merits of 4 schemes, 3 schemes are designed and simulated: CPVSCS, VPVSCS and NPVSCS. The results show that the chosen schemes function effectively. In addition to simulating the given loads, the power recovered by the loading subsystem is fed back to drive the tested device and realize the work conditions of the tested device.

Under a given load step of 175 kN • m, the settling times (2% error band) of the loaded torque range from 0.05 to 0.27 s. Under a given sine load of amplitude 90 kN • m, the -3 dB bandwidth is above 25 Hz, and the -90º phase bandwidth is above 30 Hz. The power recovery efficiency ranges from 0.54 to 0.85, which is higher under static loads than under alternating loads and increases with the load for the CPVSCS but is approximately constant for the VPVSCS and the NPVSCS. The power recovery efficiency decreases as the frequency of the load increases for all of the 3 systems. Overall, the 3 systems can be approximately ranked in terms of their loading performance from high to low as CPVSCS→VPVSCS→NPVSCS and as VPVSCS→ NPVSCS→CPVSCS for low frequencies (<10 Hz). The 3 systems can be ranked in terms of their power recovery efficiency from high to low as NPVSCS→VPVSCS→ CPVSCS.

For all of the loading performances or power recovery efficiencies, the VPVSCS results are closer to those of the NPVSCS than to those of the CPVSCS. Thus, in general, the CPVSCS exhibits the highest loading performance, the NPVSCS has the highest energy efficiency, and the VPVSCS combines the advantages of the CPVSCS and the NPVSCS.

In this study, we use constant efficiencies for the transmission components and the units and neglect the gear clearance and the lubricating power of the gear boxes, which can affect the precision of analysis. In future work, we will use more precise modes for the systems and carry out more accurate analyses.

References [1] RADPUKDEE T, JIRAWATTANA P. Design of an engine load

simulator[C]//2005 ASME International Mechanical Engineering Congress and Exposition, Orlando, Flowrda, USA, November 5–11, 2005: 1–9.

[2] LI Jianying, SHAO Junpeng, WU Zhongwen, et al. Study of the electro-hydraulic load simulator based on double servo valve concurrent control[C]//The Ninth International Conference on Electronic Measurement & Instruments, Beijing, China, Aug. 16–19, 2009: 3-699–3-705.

[3] YAO Jianyong, JIAO Zongxia, SHANG Yaoxing, et al. Adaptive nonlinear optimal compensation control for electro-hydraulic load simulator[J]. Chinese Journal of Aeronautics, 2010(6): 720–733.

[4] YU J G, ERCOLINO M B. Measurement and analysis of Acela express regenerative power recovery[C]//Proceedings of the ASME/IEEE Joint Rail Conference and the ASME Internal Combustion Engine Division, Spring Technical Conference, Pueblo, CO, United States, March 13–16, 2007: 720–733.


·633·

[5] IBAIONDO H, ROMO A. Kinetic energy recovery on railway systems with feedback to the grid[C]//14th International Power Electronics and Motion Control Conference, Ohrid, Macedonia, Sept. 6–8, 2010: T994–T997.

[6] HE Jinping, MAO Chengxiong, LU Jiming, et al. Design and implementation of an energy feedback digital device used in elevator[J]. IEEE Transactions on Industrial Electronics, 2011, 58(10): 4636–4642.

[7] MINAV T A, LAURILA L I E, IMMONEN P A, et al. Electric energy recovery system efficiency in a hydraulic forklift[C]//IEEE EUROCON2009, St. Petersburg, Russia, May 18–23, 2009: 758–765.

[8] CHEN Mingdong, ZHAO Dingquan. Research on boom energy recovery system with closed circuit in hydraulic excavators[C]//2011 International Conference on Transportation, Mechanical, and Electrical Engineering, Changchun, China, Dec. 16–18, 2011: 954–957.

[9] LIN Tianliang, WANG Qingfeng. Hydraulic accumulator-motor- generator energy regeneration system for a hydraulic excavator[J]. Chinese Journal of Mechanical Engineering, 2012, 25(6): 1121–1129.

[10] ZHANG Lujun. An energy-saving oil drilling rig for recovering potential energy and decreasing motor power[J]. Energy Conversion and Management, 2011, 52: 359–365.

[11] XU Guoqing, LI Weimin, XU Kun, et al. An intelligent regenerative braking strategy for electric vehicles[J]. Energies, 2011, 4(9): 1461–1477.

[12] RAMBALDI L, BOCCI E, ORECCHINI F. Preliminary experimental evaluation of a four wheel motors, batteries plus ultracapacitors and series hybrid powertrain[J]. Applied Energy, 2011, 88(2): 442–448.

[13] ZHANG Zhongliang, CHEN Jie, WU Bofu. The control strategy of optimal brake energy recovery for a parallel hydraulic hybrid vehicle[J]. Journal of Automobile Engineering, 2012, 226(11): 1445–1453.

[14] HO T H, AHN K K. Design and control of a closed-loop hydraulic energy-regenerative system[J]. Automation in Construction, 2012, 22(3): 444–458.

[15] TING C, TSAI D, HSIAO C. Developing a mechanical roadway system for waste energy capture of vehicles and electric generation[J]. Applied Energy, 2012, 92(4): 1–8.

[16] SUN Hui, JIANG Jihai, WANG Xin. Parameters matching and control method of hydraulic hybrid vehicles with secondary regulation technology[J]. Chinese Journal of Mechanical Engineering, 2009, 22(1): 57–63.

[17] WANG Xingjian, WANG Shaoping, WANG Xiaodong. Electrical load simulator based on velocity-loop compensation and improved fuzzy-PID[C]//IEEE International Symposium on Industrial Electronics 2009, Seoul, Korea, July 5–8, 2009: 238–243.

[18] ZHA Hongshan, ZONG Zhijian. Emulating electric vehicle’s mechanical inertia using an electric dynamometer[C]//2010 International Conference on Measuring Technology and Mechatronics Automation, Changsha, China, March 13–14, 2010: 100–103.

[19] WANG Huabing, HU Junke. Design of testing system with power recycle for hydraulic pumps durability[J]. Machine & Hydraulics, 2012, 40(18): 76–78. (in Chinese)

[20] JIANG Xingfang. Testing system with power recycle for hydraulic motor’s durability[J]. Journal of Hunan Industry Polytechnic, 2012,

12(6): 6–10. (in Chinese) [21] LI Wanguo, WANG Zhanlin. Realizing project for electric network

hydraulic-secondary-control system running reversibly[J]. Chinese Journal of Mechanical Engineering, 2006, 42(7): 76–84. (in Chinese)

[22] XUE Hua, CHEN Jingru, LUAN Menggui, et al. Coupling influence and decoupling control of the secondary regulation loading system for the drive axle of vehicle[C]//2010 International Conference on Mechanic Automation and Control Engineering, Wuhan, China, June 26–28, 2010: 3246–3249.

[23] LI Wanguo, FU Yongling, QI Xiaoye. Variable-pressure secondary-control with power recovery for loading hydraulic driving head of rotary drilling rig[C]//2013 International conference on Advances in Materials Science and Manufacturing Technology, Xiamen, China, May 18–19, 2013: 1881–1888.

[24] Rexroth general catalogue[G]. Bosch Rexroth AG. RE92003/02.98: 12/44–13/44.

[25] Rexroth general catalogue[G]. Bosch Rexroth AG. RE91604/05.99: 8/40.



[28] Rexroth general catalogue[G]. Bosch Rexroth AG. RE92076/05.93: 6–8.

[29] Rexroth general catalogue[G]. Bosch Rexroth AG. RE92055/04.96: 5/26.

[30] LI Xiaojin, YUAN Shihua, HU Jibin, et al. Mathematical model for efficiency of the hydraulic transformer[C]//2009 Asia-Pacific Power and Energy Engineering Conference, Wuhan, China, March 27–31, 2009: 1–5.

Biographical notes LI Wanguo, born in 1969, is currently a PhD candidate at School of Mechanical Engineering and Automation, Beihang University, China. He received his master degree on mechatronics engineering from Beihang University, China, in 2006. His research interests include control and energy conservation of mechatronic systems, fluid power transmission. Tel: +86-10-82339159; E-mail: [email protected]

FU Yongling, born in 1966, is currently a professor and a PhD candidate supervisor at School of Mechanical Engineering and Automation, Beihang University, China. He received his PhD degree from Harbin Institute of Technology, China, in 1993. His main research interests include mechatronics engineering, fluid power transmission. Tel: +86-10-82317307; E-mail: [email protected]

CHEN Juan, born in 1973, is currently a lecturer at Beihang University, China. Her main research interests include mechatronics engineering and reliability.

QI Xiaoye, born in 1961, is currently an associate professor at Beihang University, China. His main research interests include mechatronics engineering and fluid power transmission. Tel: +86-10-82339707; E-mail: [email protected]

Variants of Secondary Control with Power Recovery for ...

Documents

Transcript of Variants of Secondary Control with Power Recovery for ...