International Journal of Automotive Technology, Vol. 9, No. 5, pp. 509−521 (2008)
DOI 10.1007/s12239−008−0061−2
Copyright © 2008 KSAE 1229−9138/2008/042−01
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN FOR FOUR-STROKE ENGINES P. K. WONG , L. M. TAM and K. LI *
Department of Electromechanical Engineering, University of Macau, Macao S.A.R., China (Received 26 March 2007; Revised 2 November 2007)
ABSTRACT−In modern four-stroke automotive engine technology, variable valve timing and lift control offer potential
benefits for making a high-performance engine. In this paper, a novel design named dual-mode electrohydraulic fully variable valve train (EHFVVT) for both engine intake and exhaust valves is introduced. The system is mainly controlled by either proportional flow control valves or proportional pressure relief valves, and hence two different families of valve displacement patterns can be achieved. The construction of the mathematical model of the valve train system and its dynamic analysis are also presented in this paper. Experimental and simulation results show that the dual-mode electrohydraulic variable valve train can achieve fully variable valve timing and lift control, and has the potential to eliminate the traditional throttle valve in the gasoline engines. With the proposed system, the engine performance at various speeds and loads will be significantly improved.
KEY WORDS : Fully variable valve train, Dynamic model, Simulation, Electrohydraulic technology
NOMENCLATURE A1 A2 Ac Ar Cd Ctp1 Ctp2 D d dp Fst1max Fst2max Fco1 Fco2 Fr1 Fr2 ΔF1 ΔF2 F F0 Fpre
g Kg K1 K K' l2 m1 m2 Po P1 P2 PL ΔP
: cross-section area of the master cylinder piston (m2) : cross-section area of the valve cylinder piston (m2) : cross-section area of check valve orifice (m2) : cross area of pressure relief valve orifice (m2) : discharge coefficient of valve orifice : leakage coefficient of the master cylinder (m3/ s·Pa) : leakage coefficient of the valve cylinder (m3/s·Pa) : tappet displacement (m) : hydraulic pipe diameter (m) : poppet valve diameter (m) : maximum static friction force on the master cylinder piston (N) : maximum static friction force on the valve cylinder piston (N) : sliding friction on the master cylinder piston (N) : sliding friction on the valve cylinder piston (N) : friction force on the master cylinder piston (N) : friction force on the valve cylinder piston (N) : resultant force on the master cylinder piston (N) : resultant force on the valve cylinder piston (N) : load on poppet valve (N) : preload in disk-spring stack (N) : preload in valve spring (N)
Q1 Q2
Re r u V1 V2 v2 w X1 X2 X3
*Corresponding author. e-mail:
[email protected] 509
: acceleration of gravity=9.81 m/s2 : stiffness of valve spring (N/m) : stiffness of disk-spring stack (N/m) : gain of proportional flow control valve (m/V) : gain of proportional pressure relief valve (Pa/V) : length of hydraulic pipe (m) : mass of the master cylinder piston (Kg) : mass of the valve cylinder piston (Kg) : set pressure (Pa) : operating pressure in the master cylinder (Pa) : operating pressure in the valve cylinder (Pa) : residual gas pressure in combustion chamber (Pa) : pressure drop along the pipe to the valve cylinder (Pa) : hydraulic flow generated by the master cylinder (L/min) : hydraulic flow into the valve cylinder (L/min) : reynolds number : base circle radius of input cam (m) : proportional valve control signal (V) : initial volume of the master cylinder (m3) : initial volume of the valve cylinder (m3) : flow velocity in the pipe to the valve cylinder (m/ s) : valve area gradient (m) : cam lift (m) : piston displacement of the master cylinder (m) : piston displacement of the valve cylinder (i.e. displacement of engine poppet valve) (m)
510 Xv
βe θ λ ν ρ τ
P. K. WONG, L. M. TAM and K. LI : displacement of proportional valve spool (m) : bulk modulus of hydraulic medium (Pa) : crank angle (rad) : hydraulic flow drag coefficient : kinematic viscosity of hydraulic medium (m2/s) : density of hydraulic medium (kg/m3) : time delay constant of proportional flow control valve (s)
1. INTRODUCTION
Variable valve timing and lift technology provides the possibility to control the valve events, i.e. timing, lift and duration. Various studies have shown that an engine using variable valve timing allows the reduction of pumping loss, control of the internal residual gas recirculation and emissions, along with improvement of performance over a wide operating range (Ebner , 1991). All of these factors contribute to a considerable potential improvement in fuel economy. Many researches have been conducted on different types of valve trains, including mechanical, electromagnetic, and electrohydraulic. Up to now, a lot of mechanical variable valve trains for four-stroke automotive engines have been developed. Some designs have also been transferred to the commercial products, such as Honda’s VTEC and i-VTEC, Toyota’s VVTL-i and BMW’s VANOS and Valvetronic, but the products generally cannot provide fully variable valve timing & lift control. Even though some research grade mechanical valve trains (Pierik and Burkhard, 2000; Flierl ., 2005; Sellnau , 2006) are able to support fully variable valve timing & lift control, they usually offer one family of valve lift curves, hence the engine can only be optimized at specific operating conditions (Milovanovic ., 2005). Overall speaking, the general disadvantages of the mechanical variable valve actuators are complex (Maas , 2004) and bulky cylinder head, resulting in high engine center of gravity (CG). All of the mechanical variable valve systems are also limited in their flexibility of output valve profiles and strategies due to mechanical constraints (Milovanovic ., 2005). Researches on electromagnetic variable valve train (EMVVT) have recently been started (Wang ., 2002; Qiu , 2004; Cope and Wright, 2006). This type of valve actuator can theoretically perform any valve lift profiles. The EMVVT offers considerably higher degree of flexibility in comparison to mechanical variable valve train systems. However, an EMVVT system introduces a difficult motion control problem. Control of accurate valve timings, fast transitions, low seating velocities, load variation and hysteresis issues are challenge topics to the EMVVT (Giglio , 2002; Wang ., 2002). The durability is also a serious problem because the solenoid systems in the EMVVT are required to be switched on and off at each engine cycle. Owing to the above problems, the EMVVT is still being in research stage now. et al.
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The electrohydraulic variable valve train (EHVVT) has been developed for many years. Perhaps it is conceptually the simplest type when it is compared with the other two types of variable valve actuators (Giglio ., 2002). Moreover, the power density of hydraulic is the best for valve motion (Exner ., 1991). The EHVVT, with and without positional feedback control, shows higher degree of flexibility than the EMVVT system since the EMVVT has possibility to vary valve lift- something not easily achieved with electromagnetic “catch and hold” approach. This makes it difficult for the EMVVT system to realize all possible benefits of improved combustion promised by electrohydraulic systems (Milovanovic ., 2005). Although the EHVVT has many attractive advantages, its weaknesses are high cost and low durability as compared with the mechanical variable valve trains, and hence the EHVVT is yet not available commercially. Improving the weaknesses of the current EHVVT systems seems to be a challenge research topic, so the paper focuses on the development of electrohydraulic variable valve actuation systems. Lotus Engineering developed an active electrohydraulic fully variable valve train (Lumley, 1999). The system consists of a hydraulic piston attached to each engine poppet valve, which moves inside a hydraulic cylinder. Movement of the piston, and thus the engine poppet valve, is controlled by flow of hydraulic fluid either above or below the piston. The hydraulic flow control is the function of the high speed electrohydraulic servo valve. Position feed-back from the piston and engine valve assembly, is continuously provided by a fast linear displacement transducer, which allows actual valve profiles to be continuously monitored and corrected from cycle-to-cycle. Although this electrohydraulic system can theoretically perform any valve lift profiles and has good soft touch-down capability, it can only be operated below the engine speed of 3500 rpm (Allen and Law, 2002). One of the reasons is that the hydraulic servo valves are required to be switched on and off at each engine cycle. This requires the hydraulic servo valve has a high operating frequency and an extremely fast respond time. It is believed that the operating limits of the existing servo valve technology determine the rpm limit. Moreover, this kind of delicate servo valve is around $1000 per each (Lumley, 1999). Since each poppet valve requires one delicate servo valve, the total system cost will be much more expensive. Recently, Fiat has published its new electrohydraulic fully variable valve system named UNIAIR. It is the first hydraulic variable valve system can be functional up to engine speed of 8000 rpm (Ellison and Maiano, 2004). The major breakthrough is that UNIAIR system uses a rotating camshaft to cyclically push a tappet piston to generate consistent hydraulic pulses. UNIAIR is applied to the intake valves only, mainly due to the high cost of applying it to the exhaust valves. The system construction is that each engine poppet valve is connected to one tappet piston through an oil chamber, which is controlled by a normally open, on-off et al
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MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 511 hydraulic solenoid valve. When the solenoid valve is closed (activated) the intake valve essentially follows the cam motion (full lift). Early intake valve closing is obtained by opening (deactivating) the solenoid valve at a certain cam angle. Oil flows out of the high pressure oil chamber into the low pressure channel. The motion of the valve is decoupled from that of the tappet and, forced by the valve spring, the valve closes earlier than in the full lift mode. During the following re-filling of the high pressure oil chamber, oil flows back through the open solenoid valve also thanks to the presence of a spring-loaded accumulator. The function of the accumulator is to store the discharged hydraulic fluid when full valve lift is not required. Similarly, late intake valve opening can be achieved by retarding the solenoid valve activation. However, it still relies on high speed solenoid valves as Lotus’s design to realize variable valve timing and lift control, because the solenoid valves are necessary to be switched on and off at each engine cycle too. Switching the solenoid valves at high frequencies also causes durability problem. In view of the above deficiencies, this paper proposes a novel design named dual-mode electrohydraulic fully variable valve actuator in order to simply achieve fully variable valve timing and lift control together with high engine operating speed limit and durability, but it does not use the high cost electrohydraulic valves. 2. SYSTEM DESIGN
Figures 1 shows the schematic design of the novel electrohydraulic fully variable valve train (EHFVVT). The system is mainly controlled by either a common hydraulic proportional flow control valve or a proportional pressure relief valve. Therefore, two types of variable valve strategies, (1)
Early-Valve-Closing + Variable-Max-Valve-Lift and (2) Late-Valve-Opening + Early-Valve-Closing + VariableMax-Valve-Lift, can be achieved. Each of the valve strategy can optimize the engine performance under certain engine speed and load conditions. Both valve strategies may be applied to one engine, but each strategy should be individually fit to specific engine load and speed region. The selection of the valve strategies actually depends on the engineer’s experience and the engine characteristics. It is like Fiat’s UNIAIR, the novel design is also a closed circuit. The hydraulic fluid is pumped directly into a specially designed valve cylinder from a master cylinder, and it is also drawn back by the master cylinder itself. If no control signal is applied to Solenoid A of the mode selection valve, the system only operates with proportional flow control valve (i.e. proportional flow control mode). When Solenoid A is energized, the system operates with proportional pressure relief valve (i.e. proportional pressure control mode). Whenever the proportional flow control valve is operating, the oil passage between the pressure relief valve and the master cylinder is blocked, and vice versa. To control the valve timing and lift, the dual-mode EHFVVT system uses the cam and disc-spring stack for pumping different amount of hydraulic media into the valve cylinder to realize variable valve timing and lift control according to the back pressure created by the proportional valves. Unlike Lotus’s EHFVVT system which is necessary to switch its hydraulic servo valves on and off at each engine cycle to produce the hydraulic pulses, the proportional valves in this novel design are unnecessary to do that, because the rotating cam cyclically pushes the master cylinder to generate the hydraulic pulses. As a result, a higher system operating speed and longer service life can be obtained. In the general operating cycle, when the cam lobe comes
Figure 1. Schematic diagram of dual-mode electrohydraulic fully variable valve actuation system.
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down, the lobe pushes down the master cylinder piston. Thus the hydraulic oil flows into the valve cylinder through one of the proportional valves. The pressure drop across each proportional valve is adjustable by the current passing through its solenoid. The pressure in the master cylinder affects the compressed length of the disc-spring stack, and the pressure drop governs the output hydraulic flow to the valve cylinder. The electronic control unit (ECU) in Figure 1 provides the proportional valve control current according to engine speed and load signals. When the cam lobe passes out, the tappet spring forces the tappet to return. The valve spring in the valve cylinder then pushes the piston of valve cylinder backward and closes the engine poppet valve. Meanwhile the oil flows back to the master cylinder through a check valve. The proportional flow control valve produces back pressure for reverse flow and the proportional pressure relief valve does not permit reverse flow, so they are paralleled with the check valve. The check valve provides an oil return path with minimum restriction. Just like the conventional mechanical valve trains, this design still uses a cam to provide the piston displacement, but this design allows the design engineer to place the camshaft mechanism away from cylinder head. It is because the valve train components are connected via hydraulic pipes which permit the components to be distributed in the other places of the engine. This feature is able to lower the engine CG and provide more design flexibility. 2.1. Principle of Fully Variable Valve Timing and Lift Control The working principles of these two types of hydraulic proportional control systems are discussed one by one in the following parts. 2.1.1. Principle of proportional flow control EHFVVT system As shown in Figure 1, the tappet and the master cylinder piston are connected by a hard disc-spring stack. When the hydraulic pressure changes, the disc-spring stack is compressed into different lengths. The spool displacement of the proportional flow control valve is finely controlled by the current passing through its solenoid. The smaller the control current that passes through the valve solenoid, the smaller the valve throttle opening becomes. A smaller valve throttle opening creates higher pressure drop for the same volume of fluid to pass through. In other words, the smaller the control current, the higher the back pressure rises in the master cylinder. The back pressure affects the compression of the disc-spring stack. The higher the disc-spring compression, the lesser the hydraulic medium that can flow into the valve cylinder. The reason is that part of tappet displacement is spent for compressing the disc-spring stack instead of pushing the hydraulic medium into the valve cylinder. Therefore, the maximum valve lift and its timing can be changed. The
above principle is applicable to the case that while the poppet valve is extending due to the forward motion of the master cylinder piston. If no control current is applied to the proportional flow control valve, the poppet valve is deactivated. When the cam lobe is passing out, the system pressure gradually decreases and let the poppet valve start to close by discharging the oil via the check valve. In the system parameters, the preload in the valve spring is larger than that in the disc-spring stack, and the cross-section area of the master cylinder piston is larger than the valve cylinder piston. When the cam has just passed its maximum lift point, the resultant valve spring force (i.e. preload + instantaneous valve spring force) pushes the valve cylinder piston backward. According to the above design parameters and the basic principle in hydrostatics, the pressure force acting on the master cylinder piston, which is produced by the valve spring, is enlarged strong enough to push back the piston gradually and keep the disk-spring being compressed as before, until the poppet valve is fully closed. If the poppet valve is fully closed and the cam lobe has not completely passed out, the master cylinder piston is no longer to be pushed backward. At that moment, the disc-spring stack is allowed to release and let the tappet follow the remaining part of the cam lobe to return. Therefore the poppet valve can be closed first without following the cam lobe profile. The duration between the fully-closed point of the poppet valve and the termination point of the cam lobe is proportional to the compressed length of discspring stack. The higher the disc-spring stack being compressed, the earlier the poppet valve closes. This action results in early-valve-closing. 2.1.2. Principle of proportional pressure control EHFVVT system The proportional pressure mode uses a hydraulic proportional pressure relief valve to adjust the compressed length of the disc-spring stack in the master cylinder to control the hydraulic flow. The basic working principle of this mode is quite similar to the flow control mode. The major difference is that the hydraulic medium can pass the proportional flow control valve when the cam lobe starts to come down, whereas the pressure in the master cylinder in the proportional pressure control mode is necessary to take time to reach a specific set pressure level. When the pressure at the relief valve inlet exceeds the set level, the relief valve poppet is forced off its seat, and the valve is opened and let hydraulic medium flow into the valve cylinder. This action causes a delay between the cam lobe starting point and the opening point of the engine poppet valve (i.e. late-valveopening). This delay time is adjustable by varying the set pressure electronically via the ECU. The larger the control current, the higher the set pressure. When the tappet reaches maximum displacement, the peak pressure occurs in the master cylinder. The poppet valve then obtains maximum valve lift.
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 513 When the cam lobe is passing out, the poppet valve also achieves early-valve-closing according to the same principle in the flow control mode. Such kind of variable valve strategy, Late-Valve-Opening + Early-Valve-Closing + Variable-Max-Valve-Lift, is helpful for some engine operating conditions. If no control current is applied to the proportional pressure relief valve, a full valve lift is obtained. 2.2. Dynamic Model The general dynamic model of the dual-mode system is shown in Figure 2 and the system equations are derived one by one as follows. The model starts with representation of the input cam profile using a mathematical equation. Actually the cam profile is engine dependent. Nowadays many cam profile synthesis methods are available for cam design, such as: polydyne method, multi-polynomial method, and spline methods, etc. The selection of methods depends on the cam profile complexity and accuracy requirements, so the model of this part is case dependent. After selection of cam profile synthesis method, the cam lift function of crank angle can be obtained. The next consideration is the relationship among the cam lift X1, crank angle θ and the tappet displacement D, and they can be expressed as: D ( θ ) = [ X 1 ( θ ) + r ] cos θ − r (1) Equilibrium equation of the master cylinder piston: F 0 + ( D – X 2 ) K 1 + m 1 g = P 1 A 1 + X·· 2 m 1 F 1 (2) r
In Equation (2), the piston friction force F 1 is expressed as: ⎧ · X· 2 ≤ 0 ⎪ F 1 signX 2 F 1=⎨ (3) · ⎪ F signX X· > 0 r
st
r
2
c o1
⎩
Δ F 1 > F s t 1max Δ F 1 ≤ F s t 1max
⎧ F s t 1max ⎩ Δ F1
F 1=⎨ st
2
(4)
The friction force model is based on the Restoration Integral Friction Force Model (Haessig and Friedland, 1991). This model is able to simulate the friction force between the piston and the hydraulic cylinder wall more accurately. Flow continuity equation in the master cylinder Q1: · ( V 1 – X 2 A 1 ) P-1 Q = X· A − C P − -------------------------------(5) 1
2
1
tp 1
1
βe
For proportional flow control mode:
Flow equations across the proportional flow control valve spool (engine valve open) and the check valve (engine valve closed) (Jiang et al., 2002): ⎧ ⎪ C d wX v --2- ( P 1 – P 2 ) ρ ⎪ Q1= ⎨ ⎪ C ( A + wX ) --2- P v ⎪ d c ρ 2 ⎩
valve open valve closed
(6)
The relationship between the control voltage and the proportional valve spool displacement X is (Wang, 1999): (7) X· = 1--- ( Ku – X ) v
v
τ
v
For proportional pressure control mode:
Flow equations across the proportional pressure relief valve (engine poppet valve open) and the check valve (engine poppet valve closed) are expressed as (Jiang et al., 2002): ⎧ ⎪ 2 ⎪ C d A r --ρ- ( P 1 – P 2 ) P 1 > P 0 valve open ⎪ Q1= ⎨ 0 P 1 < P 0 valve closed ⎪ ⎪ 2 valve closed ⎪ C d A r --ρ- P 2 ⎩
Figure 2. General dynamic model of dual-mode EHFVVT system.
(8)
A linear relationship between the control voltage and the set pressure is usually assumed. P0 = K′u (9) Flow continuity equation of the valve cylinder Q2: · ( V 2 + X 3 A 2 ) P-2 (10) Q = Q − C P − --------------------------------2
1
t p2
2
βe
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Q 2 = X· 3 A 2
(11)
Equilibrium equation of the valve cylinder piston: P 2 A 2 + m 2 g = K g X 3 + X·· 3 m 2 + F r 2 + F + Δ PA 2 (12) In Equation (12), the piston friction force Fr2 is also expressed as (Haessig and Friedland, 1991): ⎧ · ⎪ F st 2 signX 3 Fr2= ⎨ · ⎪ F co 2 signX 3 ⎩ ⎧ F st 2max ⎩ Δ F2
F st 2 = ⎨
X· 3 ≤ 0
X· 3 > 0
Δ F 2 > F st 2max Δ F 2 ≤ F st 2max
(13) (14)
For the load F on the engine poppet valve, it is given by: P L ⋅ π d -2p F = -----------------+ F pre
(15) 4 Actually the load on the poppet valve is the force produced by the residual gas pressure PL in the combustion chamber. When the poppet valve is opened, this force should be overcome. The pressure in the combustion chamber can be approximately estimated to be 15 bars (Lumley, 1999). Furthermore, the pressure drop ΔP along the pipe in Equation (12) refers to Equation (16) (Jiang et al., 2002). l ρ v2 Δ P = λ -----2 ----------2d 2
(16)
For the general hydraulic copper pipe (Jiang et al., 2002) (the most widely used in automotive hydraulic system): λ =0.3164Re –0.25 (17)
Figure 3. Input cam profile. 2d Re= v------ν 4 X· 3 A-2 v 2 = -------------2
πd
(18) (19)
Equations (1)~(19) are useful for system simulation and dynamic analysis.
3. SIMULATION AND EXPERIMENTAL RESULTS 3.1. Simulation MATLAB SIMULINK 6.5.1 was selected to be the simulation software. Actually the input cam profile is engine dependent. For demonstration propose, a proper gasoline car engine cam profile was designed for prototype testing, and multi-polynomial method was selected to represent the cam profile for simulation. In this cam, the intake cam lobe is the same as the exhaust cam lobe. Figure 3 shows the input cam profile used in the prototype system. The dynamic
Figure 4. Simulation block diagram of proportional flow control mode EHFVVT.
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 515 Table 1. Simulation parameters of the prototype system. A1 (m2) m1 (kg) V1 (m3) k1 (N/m) 0.07 2.355×10−6 16000 1.7×10−4 Ctp1 (m3/Pa·s) Ctp2 (m3/Pa·s) F0 (N) βe (Pa) −15 −15 1.5×10 1.5×10 0 7×10−8 A2 (m2) m2 (kg) V2 (m3) Cd −5 −7 8.5×10 0.08 1.7×10 0.6 3 ρ (kg/m ) l2 (m) K (m/v) Fpre (N) −5 870 0.03 2.08×10 50 Fst1max (N) Kg (N/m) τ (s) Fco1 (N) 0.03 15 20 8000 2 Fco2 (N) ν (m /s) d (m) Fst2max (N) −4 0.006 25 15 3.8×10 w (m) Ac (m2) PL (MPa) r (m) −4 1.5 0.02 0.0157 0.02×10 2 Ar (m ) K' (Pa/ν) dp (m) −2 −7 3×10 1.26×10 4500 equation of the input cam displacement against crank angle is shown in Equation (20), which is obtained by multiple regression of the cam lift data in Figure 3 with 6 degrees, because a polynomial of degree five or six can usually regress the general automotive engine cam profiles (Xue, 2001). It is also found that a polynomial of degree 6 can fairly represent the cam profile in Figure 3, even though the degree of the polynomial has been tried to be increased over 6, the improvement is not obvious. The term θ /2 in Equation (20) is used for converting the cam angle into crankshaft angle according to the four-stroke cycle. Figures 4 and 5 illustrate the simulation block diagrams for the valve displacement in each mode based on Equations (1~7,10~20) and Equations (1~5,8~20) respectively.
⎧ –1.12 × 10 θ--- +0.0063 θ--- − 0.5432 θ--- + ⎪ 2 2 2 ⎪ ⎪ θ θ θ ⎪ 1.3474 --2- − 5.4292 --2- +10.9554 --2- − X (θ) = ⎨ (20) ⎪ 5.5264 0.5 < θ < 4.5 ( Valve open duration ) ⎪ others ( Valve close duration ) ⎪0 ⎪ ( θ ∈ [ 0,4 π ] ) ⎩ The other system parameters used in the simulation are shown in Table 1 where the geometrical and mechanical parameters come from the prototype EHFVVT presented in Section 3.2, the hydraulic specific parameters come from engineering handbooks and experiments. As the control voltage u for the proportional valves represents the control program in the ECU, and the values are determined by the experience of the automotive engineer according to different engine loads and speeds, a series of value u was tried in the simulation in order to test if the valve timing and lift could be adjustable under different engine loads and speeds. The simulation results are presented in Sections 3.3~3.5. –4
3
6
5
2
4
2
1
3.2. Experiments In order to verify the accuracy of the mathematical model developed and the feasibility of the dual-mode design, a prototype EHFVVT and its test rig were fabricated and some experiments were done. The prototype was designed based on some components and configuration of Toyota 1600c.c. 4A-series DOHC gasoline engine. Figure 6 shows the system prototype and the test rig. Figure 7 shows the interconnection between the prototype system and the input/output devices of the test rig. In the test rig, the engine speed sensor is a magnetic crankshaft-position sensor that tells the ECU the crankshaft position and how fast the engine crankshaft is turning. The ECU counts the voltage pulses generated from the sensor to determine the crankshaft speed and position. The crankshaft signal is simulated by the electric motor unit. In the unit, the three-phase AC motor provides the driving torque
Figure 5. Simulation block diagram of proportional pressure control mode EHFVVT.
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Figure 6. System prototype and test rig.
Figure 7. Interconnection between the test rig for the prototype dual-mode EHFVVT system. to the camshaft. In the prototype system, the ratio of the camshaft to the crankshaft speed signal is 1:2. The frequency inverter is used for regulating the three-phase motor speed in order to simulate different engine speeds. Besides, a throttle position sensor is used for simulating various engine load signals. The hydraulic proportional pressure relief and flow control valves were built by modification of commercial proportional valves.
MoTeC Programmable Engine Management System was selected to be the ECU. With the programming software of MoTeC, three control maps were created for individual control of the mode selection valve, the proportional flow control valve and the proportional pressure relief valve. Each control map is expressed as a three-dimensional pulsewidth-modulation (PWM) look-up table and stored in the ECU. The duty cycles of the PWM were programmed according to the crankshaft speed and the engine load
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 517
Figure 8. Valve displacement profile. signals which are the indexes to the tables. The duty cycle in the look-up table for the mode selection valve is either 0% or 100%, whereas the duty cycle range for the proportional valves is from 0% to 100%. With adjustment of the duty cycle, various effective control voltages can be obtained and hence different values of control currents can pass through the solenoids of the proportional valves. For example: at an engine speed of 4500 rpm and at 60% engine load, the relevant duty-cycle in the table for the mode selection valve and the proportional pressure relief valve is 100%, and 50% respectively, and the output voltage of the ECU is set to be 14V. If the engine approximates 60% load at 4500 rpm, then the solenoid of the mode selection valve is fully energized and switches the system to the pressure control mode. At the same time, the effective control voltages to the proportional pressure relief valve is 14V × 50%= 7 V. The duty cycles of the look-up table are actually determined by the experience of the automotive engineer according to different engine conditions and road environments. In this study, an example look-up table was designed to illustrate the feasibility of the design, but in fact the valve timing and lift can be reprogrammed according to the user requirements. The programmable ECU can also log the data from the sensors and download the data to a computer. Furthermore, a linear variable differential transformer (LVDT) was also employed for recording the real-time displacement of the poppet valve. The LVDT is mounted at the bottom of the valve spring housing and the core of the LVDT, which is extremely light in weight, is attached to the poppet valve. When the valve moves back and forth, the LVDT transforms the valve displacement into voltage signals which are received and recorded by the computer via a data acquisition board. The experiments were conducted under the combinations of various control parameters. Figure 8 displays an oscilloscope trace of the flow control mode taken at a real-time operation which is equivalent to an engine speed of 7000
Figure 9. Valve profiles under proportional flow control mode at 60% engine load and different engine speeds. rpm, 7.5 mm valve lift with valve duration of 210o crank angle. The figure also depicts an important capability of soft touch-down. Experimental data shows that a touch down velocity of less than 0.1 m·s−1 can be achieved in this prototype. The touch down velocity is controlled by the check valve orifice. With this check valve, the pressure control mode also has soft touch-down capability. This characteristic is important to avoid noise, vibration and harshness issues. 3.3. System Response at Constant Load and Different Engine Speeds The simulation and experimental results in Figures 9 and 10 show the valve duration and lift can be varied at constant load and different engine speeds under the proportional flow control mode and pressure control mode respectively. As engine speed is increased, the valve lifts in both cases go deeper, and more air-fuel mixture can be drawn into the combustion chamber. Therefore the engine volumetric efficiency will be increased. This variable valve lift control is able to improve the volumetric efficiency at different engine speeds. Moreover, as engine speed is increased, the valve open duration for both cases is changed as well. For an intake valve, longer opening time can maximize the air-fuel mixture entering the cylinder, particularly at medium to high engine speeds. So a higher volumetric efficiency is achieved at high engine speeds. On the other hand, the valve can keep short open duration at low speeds so as to prevent excessive valve overlap, which reduces volumetric efficiency (Pulkrabek, 2004). Figures 11 and 12 present the experimental and simulation results of the poppet valve acceleration at 4000 rpm and 60% engine load under the proportional flow control
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Figure 13. Simulation of P1 and P2 under proportional flow control mode at 4000 rpm and 60% engine load.
Figure 10. Valve profiles under proportional pressure control mode at 60% engine load and different engine speeds.
Figure 14. Simulation of P1 and P2 under proportional pressure control mode at 4000 rpm and 60% engine load.
Figure 11. Valve acceleration under proportional flow control mode at 4000 rpm and 60% engine load.
Figure 12. Valve acceleration under proportional pressure control mode at 4000 rpm and 60% engine load.
mode and pressure control mode respectively. In Figures 11 and 12, the mean poppet valve acceleration in its retraction period is less than that in its extension period. It is because the hydraulic media flow velocity is controlled by the check valve orifice during poppet valve retraction period, resulting in soft touch-down. Figures 13 and 14 also show the simulation results of P1 (master cylinder pressure) and P2 (valve cylinder pressure) at 4000 rpm and 60% engine load under the proportional flow control mode and pressure control mode respectively. Actually, the pressure differential between P1 and P2 is due to the compression of the disc-spring stack in order to absorb the excessive hydraulic medium in the master cylinder as the cam lobe comes down. In Figure 13, the maximum pressure differential under the proportional flow control mode is found at maximum valve lift, where the maximum excessive hydraulic medium is stored. In Figure 14, the maximum pressure differential at the pressure control mode occurs before opening the proportional pressure relief valve. It is because the front part of the cam displacement is spent by compressing the disc-spring stack before pressure
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 519 so more air-fuel mixture can be drawn into the combustion chamber for combustion, hence additional torque can be provided by the engine.
Figure 15. Valve profiles under proportional flow control model at 5000 rpm engine speed and different engine loads. relief, and hence no hydraulic medium flows into the valve cylinder to build up the pressure P2.
3.5. Simulation of Valve Strategies for Fully Variable Valve Timing and Lift Control 3.5.1. Valve strategy of proportional flow control mode The simulation of a set of valve profiles for both intake and exhaust valves based on various control voltages is presented in Figure 17. Figure 17 depicts that the valve lift, timing and valve overlap can be adjustable through the proportional flow control system, and hence it forms a reasonable valve strategy which matches with the general valve profiles for throttleless four-stroke spark ignition engines (Allen and Law, 2002), resulting in reduction of pumping loss (Sellnau and Rask, 2003). It shows that the dual-mode EHFVVT system has the potential to eliminate the traditional throttle valves in the gasoline engines. 3.5.2. Valve strategy of proportional pressure control mode Figure 18 is a set of valve profiles for intake and exhaust valves based on the proportional pressure control system. It shows that the valve lift, timing and overlap can be adjust-
3.4. System Response at Constant Engine Speed and Different Loads Figures 15 and 16 present the simulation and the experimental results of the valve profiles at constant speed and different engine loads under the proportional pressure control mode respectively. When engine load is increased, the valve open duration and lift are increased accordingly,
Figure 17. Valve strategy of proportional flow control mode at 100% engine load.
Figure 16. Valve profiles under proportional pressure control mode at 5000 rpm engine speed and different engine loads.
Figure 18. Valve strategy of proportional pressure control mode at 100% engine load.
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P. K. WONG, L. M. TAM and K. LI
Table 2. Comparison between the dual-mode system and typical EHFVVT systems. Lotus’s EHFVVT Fiat’s UNIAIR system system Valve lift Fully variable Fully variable Variable, but it somewhat Valve timing depends on the input cam Not constrained profile Test engine speed Up to 8000 rpm Up to 3500 rpm Cost High High Individual valve actuation Yes Yes Control system complexity Complex and demanding Complex and demanding able. The higher the relief pressure setting, the smaller the valve overlap. Hence it forms a reasonable valve strategy, Late-Valve-Opening, Early-Valve-Closing and VariableMax-Valve-Lift, for the four-stroke engines. Such valve strategy can promote better combustion and fuel economy as well as reduction of HC emission and pumping loss from idle to medium engine speeds (Law , 2001) (Sellnau and Rask, 2003).
Novel dual-mode EHFVVT system Fully variable Variable, but it somewhat depends on the input cam profile Up to 8400 rpm Relatively low Yes Relatively simple
3.6. Discussion of Results Figures 9~12 & 15~16 show that the experimental results which are in good agreement with the simulation results. These results are able to prove that the mathematical model developed is correct. One of error sources may be that when the cam profile is regressed for simulation using Equation (20), there would be a little bit difference between the actual cam profile and the profile generated using Equation (20). The imperfect tune-up of the simulation parameters, deformation of the test rig and the measurement error in the experiment are also possible error sources. Overall speaking, both experimental and simulation results illustrate the dual-mode EHFVVT system can really achieve fully variable valve timing and lift control.
Basically, the current valve strategies offered by Lotus’s EHFVVT system and Fiat’s UNIAIR are quite similar to the output valve strategies (Figures 17~18) of the dualmode system. However, in terms of flexibility of valve timing & lift control, Lotus’s system is the best because it is independent of camshaft. As far as maximum operating speed is concerned, the dual-mode system is higher than Lotus’s design and slightly higher than UNIAIR system. At the same time, the estimated cost of the dual mode system is the cheapest. It is because the proportional valves in the dual-mode system are not required to be switched on and off at each engine cycle for variable valve timing and lift control. That simplifies the control system and lets the system attain high durability and reliability. So it can use common and low cost proportional valves to control, whereas the other two systems adopt high cost hydraulic valves and complicated closed-loop control systems. As a matter of fact, it is quite difficult to point out which system is the best based on the current information collected from the literatures, but one point can be sure that the dual-mode EHFVVT system provides an alternative solution to the problem of variable valve timing and lift control.
4. DESIGN COMPARISON
5. CONCLUSIONS
This section discusses the comparison between the dualmode system with two typical EHFVVT systems, Fiat’s UNIAIR and Lotus’s EHFVVT. Fiat’s UNIAIR is the most practical electrohydraulic variable valve train. Lotus EHFVVT is one of the most famous active electro-hydraulic variable valve trains. Table 2 shows the comparison between the dual-mode system and these two EHFVVT systems. Since UNIAIR has already patented and no mathematical model and experimental data has been released, the comparison cannot include any quantitative comparison about simulation and experimental results. So it focuses on their design features only. It is noted from Table 2 that the dual-mode system almost shares the same characteristics with Fiat’s UNIAIR.
A dual-mode electrohydraulic fully variable valve timing and lift control system for four-stroke engine and its mathematical model for valve train dynamic were successfully developed. The design is the first attempt to use common hydraulic proportional flow control valves and proportional pressure relief valves to provide two families of variable valve profiles. The proportional valves are much cheaper than the fast speed electrohydraulic servo/solenoid valves which are conventionally adopted in the EHFVVT systems. Unlike the typical EHFVVT systems, the proportional valves in the dual-mode system are not required to be switched on and off at each engine cycle. That simplifies the control system and lets the system attain high durability and reliability. Another new idea of the design is to use a disc-spring stack for fully variable valve displacement
et al.
MODELING AND SIMULATION OF A DUAL-MODE ELECTROHYDRAULIC FULLY VARIABLE VALVE TRAIN 521 control which can make the design more compact and simple. These two innovations make the idea of the novel EHFVVT system more practical. Because of possible separation of the master and the valve cylinder via hydraulic pipes, it will massively reduce the height of the engine. Therefore the vehicle size can be reduced and the engine design flexibility can be also increased as well. Based on the mathematical model developed, system simulation was done; two families of valve profiles under the proportional flow control and pressure control modes were simulated. Besides, a prototype EHFVVT system and test rig were successfully fabricated and tested. To demonstrate the flexibility of the system, the prototype was designed to be programmable, so that the engineers are allowed to select and test different valve strategies and profiles for various engine requirements. By comparing the simulation results with the experimental results, it is observed that the dynamic model developed is valid. Experimental and simulation results also show that the dual-mode EHFVVT system can really achieve fully variable valve timing and lift control, and has the potential to eliminate the traditional throttle valve in the gasoline engines. If the proposed design is implemented on the real fourstroke automotive engines, it is believed that the engine performance at different speeds and loads will be significantly improved. Nevertheless, the dual mode EHFVVT system is still in preliminary study; many investigations and experiments are required to be carried out in the future.
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