Automotive Innovation (2018) 1:300–310https://doi.org/10.1007/s42154-018-0039-3

Design of a Hydraulic Control Unit for a Two-Speed Dedicated ElectricVehicle Transmission

Xiangyang Xu1,2 ·Wenbo Sun1,2 · Tianyuan Cai1,2 · Yanfang Liu1,2 · Xiao Han1,2

Received: 12 August 2018 / Accepted: 3 November 2018 / Published online: 1 December 2018© China Society of Automotive Engineers(China SAE) 2018

AbstractTwo-speed automatic transmission is one solution to increase the economic efficiency and dynamic performance of batteryelectric vehicles (BEV). Hydraulic control unit (HCU) is a key component in automatic transmissions, which determinesthe quality of shifting directly. Based on the structural scheme and shift logic of a two-speed dedicated electric vehiclestransmission (2DET) with two wet clutches, we designs a 2DET hydraulic control unit composed of three subsystems:pressure regulating and flow control system, shift operated and control system and cooling and lubrication system. The resultsof the experiments, including the valve body bench test, transmission bench test and vehicle test, show that the design ofhydraulic control unit meets the requirements.

Keywords Battery electric vehicle · Automatic transmission · Hydraulic control unit · Dynamic simulation

Abbreviations

BEV Battery electric vehicleHCU Hydraulic control unit2DET Two-speed dedicated electric vehicle transmission

1 Introduction

With energy shortages and environmental pollution becom-ing ever more serious, battery electric vehicles (BEVs)occupy an increasingly important position in the currentmarket because of their superior energy-saving and environ-mentally friendly features. Many companies and universitieshave carried out research on the power characteristics ofbatteries, charging technology, powertrains and electroniccontrol systems for BEVs.

Most BEVs currently on the market, including the BAICEU260, Tesla Model S and BYD e6, are equipped with a

B Yanfang Liuliuyf@buaa.edu.cn

1 School of Transportation Science and Engineering, BeihangUniversity, Beijing 100191, China

2 Beijing Key Laboratory for High-efficient PowerTransmission and System Control of New Energy ResourceVehicle, Beihang University, Beijing 100191, China

single-gear reducer [1]. Our simulation analysis results showthat utilizing a two-speed automatic transmission can bothmarkedly lower the requirements for the motor in the elec-tric vehicle while not significantly increasing the complexityor cost of the powertrain and will keep the motor working ina more efficient operating range. Shifting gear improves theeconomic efficiency and dynamic performance of the entirevehicle by more than 10% [2]. Oerlikon Graziano of Italyhas developed a two-speed transmission designed for a smallelectric vehicle. AVL, an engineering company based inAus-tria, has designed a two-speed automatic transmission withfixed-shaft gears and the German branch of GKN, a Britishglobal engineering group, has also designed a two-speedautomatic transmission based on synchronizer shifting thatis currently being used on the BMW i8 hybrid sports vehicle.Prof. Song Jian from Tsinghua University has developed anew type of dual-block uninterrupted mechanical transmis-sion (UMT) with simple planetary rows, centrifugal frictionclutches and band brakes that is not yet in mass production.

The paper is organized as follows. Section 2 introduces thestructure of our two-speed dedicated electric vehicle trans-mission (2DET) and outlines the design concept proposedfirstly in [3]. Section 3 presents the parameter design ofthe hydraulic control unit (HCU). Section 4 describes thethree-dimensional modeling of the hydraulic valve body, andSect. 5 presents the simulation results from the SimulationXsoftware of Germany. Sections 6 and 7 describe the experi-

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Design of a Hydraulic Control Unit for a Two-Speed Dedicated Electric Vehicle Transmission 301

1C1 0 C2

2

34

5

67

Fig. 1 Transmission schematic of the 2DET

Table 1 Shift pattern of the2DET

Speed Clutch

C1 C2

R ●

N

1 ●

2 ●

mental results of the bench tests and vehicle tests. Section 8presents our conclusions.

2 Design Concept for the HCU and 2DET

Figure 1 shows the structure of the 2DET. Only two multi-disk wet clutches are used as the shift elements. With clutchC1 or clutch C2 engaged, power is transmitted to the interme-diate shaft via gear set 3 or gear set 4 to obtain two differentspeed ratios.

All simple shifts can be performed by changing the twoshift elements (shown in Table 1). For reverse gear, clutchC1 is engaged and first gear selected, and the motor rotatesin reverse.

The HCU in the 2DET must meet the following require-ments [4–7].

(1) The automatic shift function must be guaranteed. TheHCU receives information on the actual working condi-tions and controls the shift solenoid valve that controlsthe position of the shift valve’s spool and activates sep-aration and engagement of the clutch.

(2) Sufficient cooling and lubrication flow for each trans-mission system must be provided and the flow must bedistributed to ensure an appropriate oil pressure in themain line and other systems.

(3) Safety must be taken into consideration to prevent thetransmission from hanging when two gears are engaged

Fig. 2 Schematic of the design of the 2DET

at the same time. When the electronic control systemfails, the HCU should be able to avoid invalidation; thatis, the car should be able to work even if the electroniccontrol system fails.

Figure 2 shows a schematic of the 2DET hydraulic controlsystem we have designed to satisfy the above three designrequirements.

Table 2 shows the components of each subsystem in theHCU and the functions of each component. DAV_1 andDAV_2 in Fig. 3 are two large flow solenoid valves [8–12]that directly control the oil pressure to the clutches. Thesevalves must be designed to meet the requirements of theclutches. Compared with the two-stage valve structure usedin the conventional hydraulic system of an automatic trans-mission consisting of a proportional valve and a mechanicalvalve, adopting large flow solenoid valves simplifies thestructure of the valve body and improves the rate of responseand the accuracy of shift. Valve SV_1 consists of a solenoidvalve and a mechanical valve and is not activated when theHCU is operating normally. However, when there is a faultin the HCU, it is designed to cut off the oil circuits for clutchC2 to ensure that two gears are not engaged at the same time.

3 Parameter Design

Figure 4 shows a flowchart of the detailed development pro-cess of the HCU for an automatic transmission.

First, the design objectives of the vehicle should be speci-fied, and the relevant requirements and the design parametersof the 2DET should be determined to match the performancerequirements of the powertrain, including the design of themechanical and electronic systems that will directly deter-mine the design requirements of the HCU as well as theoverall space boundaries that will affect the layout of theHCU.

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Table 2 Subsystem componentsof the HCU

Subsystem Components Function

Pressure regulation and flowcontrol

Filter Keeps hydraulic oil clean

Electronic oil pump Supplies the oil to HCU

Accumulator D_A Reduces fluctuations in the mainoil pressure

Valve HP_CV Regulates the main line pressure

Cooling and lubrication Valve OF_CV Cooperate to regulate the flow ofcooling and lubricationaccording to the actual needs ofthe gearbox

Long throttle hole

Cooling and lubrication orifice

Shift operation and control DAV_1\DAV_2 Controls the clutches

Valve SV_1 Prevents the transmission fromhanging when two gears areengaged simultaneously

One-way valve Opens when disengaging theclutch

Wet clutches C1\C2 Carry out the shift

Fig. 3 Direct shift control

The design of the mechanical system of the automatictransmission includes design of the gears and selection ofthe components. According to the actual design scheme, therequirements for the cooling and lubrication flow of the HCUraised by the internal working friction pairs of the clutchescan be obtained. The design of the electronic system includesthe shift logic, safety mode and regulatory requirements. Theoverall design of theHCU is determined by its design require-ments, which are based on the design requirements of themechanical and electronic systems.

The clutch is the main control object in the hydraulic sys-tem.The total flowandoil pressure requirements for theHCUin the transmission can be obtained by calculating the coolingand lubrication flow demand for clutches, shift flow demandand shift oil pressure demand of the hydraulic system com-bined with the cooling and lubrication flow demands of themechanical friction pairs. This allows the component design,HCU structure design, hydraulic valve body layout and oilcircuit design to be decided in sequence.

After the hydraulic valve body design has been completed,its strength needs to be analyzed. When the strength condi-

Vehicle design objectives

Mechanical system design requirements Electronic system Space limit

HCU design requirements

HCU principle design

Function definition

Clutch design

Cooling and lubrication flow Shift flow Shift pressure

Total flow

Cooling and lubrication flow

Total cooling and lubrication Flow

Components design of HCU

HCU structure

Hydraulic valve body layout

Oil circuits design

Strength analysis Structure optimization

Valve body test

Transmission test

Mechanical system design

HCU spatial arrangement

verification verification

Fig. 4 HCU development process

tion is met, valve body tests and structural optimization canbe carried out simultaneously. Tests are performed to verifythe design of the mechanical and hydraulic systems, and if

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Design of a Hydraulic Control Unit for a Two-Speed Dedicated Electric Vehicle Transmission 303

there are problems, design modifications are made until thehydraulic system meets the design requirements.

3.1 Clutch Design

The 2DET uses two identical wet clutches, which are themain control objects in the hydraulic control system.The totalflow and oil pressure demand for the entire hydraulic controlsystem is obtained by calculating the shift flow demand, theshift oil pressure demand and the cooling and lubricating flowdemand for clutches combined with the mechanical systemcooling and lubricating flow demand. The detailed design ofthe HCU can then proceed. Equations (1)–(8) are formulasfor calculating the clutch-related parameters [13].

Fn � Tmaxrm · μ

(1)

Fclamp � FnZR

(2)

Pk � Fclamp

Ak(3)

Pe � Fseal + FsmaxAk

(4)

P � Pk + Pe (5){sp � ssum + ssinssum � ZR · s (6)

Voil � sp · Ak (7)

Qflow � Voilts−min

(8)

where Fn is the total pressing force, Tmax is the maximumtorque that needs to be transmitted, rm is the equivalent radiusof the friction plate, μ is the friction coefficient, Fclamp is themaximumpressing force each friction plate actually receives,ZR is the number of friction plates, Ak is the cylinder pistonarea, Pk is the oil pressure at nominal torque, Pe is the pistonpressure to balance the spring and friction, Fseal is the sealedfriction, Fsmax is the maximum spring force, sp is the totalpiston stroke, ssin is the stroke of the reset spring, s is theseparation distance of a single friction plate, Voil is the vol-ume of hydraulic fluid needed for engagement and Qflow isthe maximum volume flow required.

Calculation and analysis give the required maximum oilfilling flow for the 2DET as 4.3 L/min. The clutch requiresan oil pressure of 12 bar to transmit a torque of 300 Nm.

To prevent the oil pressure generated by the HCU frombeing too high and damaging the clutches, the hydraulic pres-sure safety system is designed to activate at 20 bar.

3.2 Flow Distribution for the Coolingand Lubrication System

3.2.1 Clutch Cooling and Lubrication Flow

The flow required for clutch cooling and lubrication is deter-mined by Eqs. (9)–(11) [12].

Q̇gen � Mk(ω10 − ω20) (9)

Qgen � (ω10 − ω20)∫

Mk

(1 − t

ts

)dt (10)

Qoil � Qgen

kA (11)

where Qgen is the heat produced by the clutch sliding fric-tion, Mk is the friction torque during clutch engagement,ω10 andω20 are the angular velocities before and after theclutch is engaged, k is the coupling coefficient and A is theclutch friction plate engagement area.

Because the two clutches in this project are identical, thetotal lubrication flow required by the clutches is 3 L/min.

3.2.2 Bearing Cooling and Lubrication Flow

For the bearing, using the calculated rotational speed that canbe achieved as the calculation condition, the required flow forcooling and lubrication is determined by Eqs. (12)–(14) [12].

Qgen � k4n · M (12)

M � k1 · G1 · (nμ)0.62 · (Peq

)0.3 (13)

Qoil � Qgen

kb · (T2 − T1)(14)

where Qgen is the heat produced by the bearing, k4n is thesize factor,M is the operating torque, n is the rotational speedof the bearing, k1 is the geometric factor, Peq is the equiva-lent dynamic load, k1 is the bearing torque constant, k1 is acoefficient, T1 is the temperature of the bearing and T2 is themaximum operating temperature of the bearing.

Calculation gives the required flow for cooling and lubri-cation of all the bearings in the entire machine as 1.09 L/min.

3.2.3 Gear Cooling and Lubrication Flow

The flow required for gear cooling and lubrication can becalculated as follows [12]:

Qoil � P · kc · ρ · t (15)

where P is the power loss, k is a coefficient, ρ is the densityof the hydraulic oil, c is the specific heat capacity and t is thetemperature drop. The design goal of transmitting a torque

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of 300 Nm requires a flow for cooling and lubrication of1.3 L/min.

3.3 Electronic Pump

Because the engine in a BEV does not have an idle state,the design of the electronic oil pump must be matched tothe HCU to ensure that the HCU still works normally whenthe vehicle is stationary. In this project, a gear pump wasselected. The main performance parameters are the nominaloil pressure and displacement.

3.3.1 Nominal Pressure

The nominal pressure is the highest oil pressure that the oilpump can reach under normal operation. The actual outputpressure of the oil pump depends on the size of the back-endload and the pressure loss of the entire system, regardlessof the pump displacement. In this project, the maximumoil pressure of the HCU was the oil pressure required fora clutch transmission torque of 300 Nm when shifting takesplace, which is 12 bar. Taking into account oil pressure lossesduring the flow of hydraulic oil through the oil circuits, thenominal pressure of the oil pump should be slightly higherthan the maximum clutch oil pressure. Therefore, the nomi-nal pressure of the oil pump was set at 15 bar.

3.3.2 Output Volume

Disregarding leakages, the displacement of the hydraulicpump is the volume of hydraulic oil that the hydraulic pumpoutputs in one revolution. The displacement is determinedbased on the peak flow of the HCU, which in this projectwas 12 L/min. Thus, the electronic oil pump needs to meet amaximum displacement requirement of 12 L/min.

The actual oil pump speed flow measured by experimentis shown in Fig. 5. The flow and speed are basically linearbecause a quantitative gear pump is used and the performanceof the electronic pump meets the requirements.

4 Three-Dimensional Modeling

Figure 6 shows the three-dimensional structure of the 2DEThydraulic valve body developed in this project. It is designedwith an integral structure that differs from the structureadopted by traditional automatic transmissions (AT).

The three-dimensional model of the HCU uses an inte-grated method that emphasizes the sharing of product char-acteristics during the design process to reduce modificationsand improve the design efficiency. The detailed design flowis shown in Fig. 7. The design of the oil circuits for onehydraulic valve body is based on the layout of the hydraulic

Fig. 5 n−Q map for the pump

Fig. 6 Model of hydraulic valve body

valve body. First, a 2D layout of the hydraulic valve body ismade, and then, the 3D layout is made to incorporate the lim-itations of the oil circuits, minimumwall thickness limits andother factors.When the 3D layoutmeets all requirements, theindividual oil circuits can be designed. Because the hydraulicvalve body has an integral structure, there are some princi-ples that should be observed in designing the oil circuits,including the method of plugging the circuits and the circuitconnections. A draft design of the oil circuits should be pro-duced and then optimized, an operation that is simpler withan integral valve body than with a valve body consisting ofthree or more components. Finally, all the design processesare completed when the optimized oil circuits meet all therequirements.

Following are the four main advantages of the integralvalve body compared with the traditional design.

(1) Reduced size and volume. The structure is one-thirdsmaller than that of the traditional design.

(2) Simple inner structure and low cost. By using oil circuitscomposed of holes that are perpendicular to the wall ofthe valve body rather than complex oil sinks, the valvebody is easier to manufacture.

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Design of a Hydraulic Control Unit for a Two-Speed Dedicated Electric Vehicle Transmission 305

Structural analysis of hydraulic components

HCU structure

Hydraulic valve body 2-d layout

Oil circuits design Minimum wall thickness limit Boundary dimension

Hydraulic valve body 3-d layout

Check whether the requirements are met

Hydraulic componentsoilports classification

Oil circuitsplugging method

Oil circuits connection method

Principle of oil circuits design

Oil circuits design

Oil circuits optimization

Check whether the requirements are met

2DET hydraulic vale body

false

False

True

True

Fig. 7 Design flow for the hydraulic valve body

(3) Reliable sealing and low leakage. Leakage between twovalve components is eliminated.

(4) Fast response and high accuracy. Reduced leakage andthe simple inner structure contribute to improved per-formance.

The above advantages make the integral structure moresuitable for a transmission with a low number of gears.

5 Simulation Results

The 2DET HCU is modeled with the SimulationX softwareof Germany.

The large flow solenoid valve acts as a direct shift valve intheHCUand is thus of paramount importance. In this project,

Fig. 8 Large flow solenoid valve test bench

we considered the multi-physics coupling effect, and simu-lated and analyzed the dynamics of the large flow solenoidvalve. First, the electric field and the magnetic field werecombined to simplify the complex physics coupling of thelarge flow solenoid valve into an “electromagnetic–mechan-ical–hydraulic” triple coupling effect. That is, the key pointsextracted from the “electromagnetic–mechanical” were airgaps and the key points extracted from the “mechanical–hy-draulic” were valve structures. The key points were thencoupled to perform complete dynamic modeling and sim-ulation analysis of the valve.

The large flow solenoid valve test bench shown in Fig. 8was set up to verify the simulation model. According to thecharacteristics of the large flow solenoid valve, the output oilpressure of the valve was approximately linear with the valvecurrent. The response curve of the pressure of the large flowsolenoid valve as a function of the current was obtained byvarying the input current signal as shown in Fig. 9. The tem-perature of the hydraulic oil was maintained at 90 °C, and theinput oil pressure was 2.0MPa. To ensure that the simulationresults and the experimental test results were comparable, thesame environmental parameters as used in the actual exper-imental tests were input to the simulation model. Figure 10shows a comparison between the oil pressure response curveobtained in the experimental test (blue curve) and the simu-lation results (red curve) for the same parameters.

The figure shows clearly that under the same externalenvironmental conditions, the maximum oil pressure valueson the pressure response curves obtained by simulation andexperiment are close to each other at around 1.26 MPa. Inaddition, the slopes of the two oil pressure curves are almostidentical when increasing and decreasing the current. Thesecurves verified that themodel used for themulti-physics cou-pling anddynamics of thevalvewas accurate,which indicatesthat the correct modelingmethodwas selected. The completeHCU model could then be built.

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Fig. 9 Current signal

Fig. 10 Comparison of simulation and experiment

The shifting processes of the HCU were simulated withSimulationX. During the simulation, the main oil pressure(17 bar) and total flow (12 L/min) remained unchanged. Fig-ure 11 shows the current signal to the two large flow solenoidvalves. To ensure that the control pressure of the large flowsolenoid valve corresponds to the main oil pressure, the cur-rent curve of the DAV_1 is offset upward.

Figure 12 shows the oil pressure curve of the shiftingprocess [14]. In the process of upshifting, clutchC1 is depres-surized and the oil pressure to C1 changes smoothly withoutfluctuation. Clutch C2 is boosted and the oil pressure to C2fluctuates by a certain amount. At the same time, the main oilpressure drops slightly. However, with the rapid completionof the oil-charging process, the oil pressure to clutch C2 andthe main oil pressure are quickly stabilized and the clutch oilpressure begins to rise steadily. The oil pressure fluctuationin clutch C2 during the oil filling process is caused by thechange in the load of the clutch control oil passage when theclutch piston cylinder starts to establish the control oil pres-sure andwhen the piston starts tomove. At the same time, theoverall load on the HCU also changes due to changes in theload on the clutch control oil circuits. However, the fluctua-

Fig. 11 Current signal

Fig. 12 Oil pressure curve during shifting

tion in the main oil pressure is small and transient, so it canbe ignored. When downshifting, the change in the clutch oilpressure is the same as when upshifting, and is not describedhere.

6 Experimental Results

After the design had been completed, a series of experimentswere carried out on the HCU, including a bench test on thevalve body and a bench test on the whole transmission, inwhich the various functions of the HCU were verified.

6.1 Valve Body Tests

The layout of the test rig for the valve body bench tests isshown in Fig. 13. According to the design goals and exper-imental data, the HCU design mainly needs to satisfy tworequirements at the same time: a reasonable oil pressure in themain oil line to ensure full compression of the wet clutch and

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Design of a Hydraulic Control Unit for a Two-Speed Dedicated Electric Vehicle Transmission 307

Fig. 13 Valve body bench test

Fig. 14 Oil pressure changes at the outlet of the oil pump

sufficient cooling lubrication flow to prevent critical compo-nents such as bearings and gears from burning out.

Because the viscosity characteristics of the hydraulic oilare affected by the temperature, the outlet oil pressure of theoil pump and the clutch oil pressure are greatly affected bythe temperature. Figure 14 shows how the oil pressure at theoutlet of the oil pump changes at different temperatures, andFig. 15 shows how the oil pressure to the clutch changes.Because the electronic oil pump used is a quantitative gearpump, the flow from the pump depends on the speed of rota-tion.Therefore, at different temperatures, theflow in theHCUbasically remains the same. Figure 16 shows the changes inthe valve body inlet flow, and Fig. 17 shows changes in thecooling and lubrication flow.

Figures 18 and 19 show further oil pressure tests on theclutches. For the 2DET HCU, because the oil pressure tothe clutches is controlled directly by the large flow solenoidvalves, we can input a certain input signal to the large flowsolenoid valves and verify the response of the output oil pres-sure to this input signal.

Fig. 15 Oil pressure changes in the clutch

Fig. 16 Changes in the valve body inlet flow

Fig. 17 Changes in the cooling and lubrication flow

Figure 18 shows the response of clutch C1, which is con-trolled by DAV_1, a normally high large solenoid valve, to astep signal (black curve) at under 90 °C. In the phase where

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Fig. 18 Valve experiment withstep signal

Fig. 19 Valve experiment withslope signal

the control oil pressure started to decrease, the oil pressure tothe clutch began to respond to the current increase at 0.5A,corresponding to an oil pressure value of 1 MPa. Subse-quently, the control oil pressure followed the changes in thecurrent steps, and the output response to each current stepwas instantaneous and stable, forming oil pressure responsesteps.

Figure 19 shows the response of clutchC1 to a slope signal(black curve) at under 90 °C. When the current was increas-ing, the clutch oil pressure began to respondwhen the currentreached 0.58 A, corresponding to an oil pressure value of1.25 MPa. The oil pressure continued to decline after thecurrent reached 1 A for 1 s, and after a lag in the current of0.23 s, it started to rise gradually to a maximum.

The valve experimentswith a step signal and a slope signalshow that the clutch oil pressure followed the changes in thecurrent to the large flow solenoid valve well, although thecharacteristics of the large flow solenoid valve produced lagsin the results.

6.2 Transmission Test

The output oil pressure and flow satisfied the requirementsof the bench test on the transmission. The hydraulic valvebodywas installed to carry out functional bench tests, includ-ing a static test, no-load test, no-load temperature-increasingtest, loading experiment, efficiency test and durability test.Among these, the static test further verified the functions ofthe HCU. Figure 20 shows the test bench for comprehensiveperformance testing.

The static test results for the hydraulic valve body wereobtained from the data in Table 3, where Npump is the speedof the pump, Toil is the temperature of the hydraulic oil, Pis the oil pressure in the main circuit and PC1/C2 is the oilpressure to clutches C1 andC2. As the speed of the electronicpump increased, the HCU could provide sufficient oil pres-sure for the clutches, but the oil pressures to clutches C1 andC2 showed a large drop in comparison with the main circuitoil pressure. The main reason for considering the flow dis-tribution of the HP_CV valve is that part of the hydraulic oil

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Design of a Hydraulic Control Unit for a Two-Speed Dedicated Electric Vehicle Transmission 309

2DET

Fig. 20 Comprehensive performance test bench with 2DET

Table 3 Static test results

Npump/rpm Toil/◦C P/bar PC1/bar PC2/bar

400 52 2 1.2 1.1

600 51.3 2.9 2.1 2.1

800 51.1 4.4 3.1 3.1

1000 51.1 6.3 4.5 4.8

1200 51.1 8.8 5.9 6.6

1400 51.1 10.7 7.8 8.3

1600 51.1 14.1 9.6 10.6

1800 50.8 16 11.8 12.6

Fig. 21 2DET installed on the BAIC EU300

in the main line flows to the cooling lubricating oil circuits,leading to a decrease in the clutch oil pressure.

The tests showed that the indicators of the HCU met thebasic requirements, and in the loading experiment, the HCUcan engage the clutch under an input torque of 300 Nm.

7 Vehicle Test

To further test the performance of the 2DETHCU, the single-stage reducer in the BAIC EU300 was replaced by the 2DET,as shown in Fig. 21.

Fig. 22 Power-on upshift

Fig. 23 Power-on downshift

Figure 22 shows the power-on upshift curve for the wholevehicle, showing that the main oil pressure stabilized ataround 14 bar at the beginning. When the shift operationstarted, the main oil pressure fluctuated when clutch C1 dis-engaged and clutch C2 engaged, and then stabilized at about14 bar. The oil pressure of clutch C1 and clutch C2 fol-lowed with the target oil pressure well, but the oil pressure ofclutch C2 jumped by about 3 bar near KP point. Because themotor speed could not be controlled, motor speed can onlybe dragged down by increasing the oil pressure of clutchC2. Thus, the lock-up phase of clutch C2 coincided with theinertia phase of clutch C1.

The power-on downshift profile is shown in Fig. 23. Sim-ilarly, the main oil pressure was stable at more than 12 bar,meeting the requirements. However, when the shift operationstarted, the main hydraulic pressure fluctuated and was diffi-cult to control. The actual oil pressure to clutches C1 and C2basically followed the target oil pressure change, but the oilpressure jump at the KP point still existed. Because themotortorque could not be controlled, the torque phase of clutch C2overlapped the inertia phase of clutch C1.

8 Conclusions

This project developed an HCU for a 2DET that is suitablefor use in other two-speed transmissions for BEVs.

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(1) Our design for the 2DET hydraulic control unit adoptedthe working principles of the mechanical system andshift logic and the valve body was designed with anintegral structure that reduced leakages and the size ofthe unit.

(2) The functional tests on the valve body and the trans-mission bench tests verified the functionality of thehydraulic valve body. The results of the vehicle testsshow that the shifting quality was satisfactory. Largeflow solenoid valves were used to control the clutches,and these are known to have many advantages, includ-ing reducing the number of valves and improving therate of response and the shifting accuracy.

(3) The hydraulic control unit developed in this project isthe first prototype version for the 2DET, and there aremany factors that remain for future study, such as theefficiency of the HCU and shift quality.

Acknowledgements This work is supported by the National NaturalScience Foundation of China (No. 51405010) and Beijing Key Labora-tory for High-efficient Power Transmission and System Control of NewEnergy Resource Vehicles.

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