Estimation of Residual Exhaust Gas of Homogeneous Charge Compression Ignition Gasoline Engine Operating Under Negative Valve Overlap Strategy

  • Huanchun GongEmail author


To meet the requirements of the homogeneous charge compression ignition gasoline engine’s rapid cylinder exhaust gas rate and accurate control of combustion phasing, a residual exhaust gas rate model was proposed. A heat dissipation model for gas flow in the exhaust passage and exhaust pipe was established, and the exhaust gas was established. Flow through the exhaust valve was considered as an adiabatic expansion process, the exhaust temperature was used to estimate the temperature in the cylinder at the time that the valve was closed, and the cylinder exhaust gas rate was calculated. To meet the requirements of transient operating conditions, a first-order inertial link was used to correct the thermocouple temperature measurement. Addressing this delay problem and modification of the exhaust wall temperature according to different conditions effectively improved the accuracy of the model. The relative error between the calculated results of this model and the simulation results determined using GT-POWER software was within 3.5%.


Residual exhaust gas Negative valve overlap angle Homogeneous charge compression ignition Gasoline engine Numerical estimation 

1 Introduction

Homogeneous charge compression ignition (HCCI) [1] is an efficient low-temperature combustion method that employs advanced variable intake technology, fuel injection technology, and highly diluted combustion. HCCI combustion has significant potential to improve the thermal efficiency and emission performance of gasoline engines. HCCI combustion can be achieved in a variety of ways, including use of a variable compression ratio, intake air heating, exhaust gas pressure or re-suction, and dual fuel. Implementation of the internal exhaust gas recompression strategy requires little modification to the engine, so it is widely used in four-stroke gasoline engines to achieve HCCI combustion [2]. This strategy retains the exhaust gas in the cylinder by closing the exhaust valve in advance to form a negative valve overlap (NVO) angle or by adopting a lower exhaust valve lift. This part of the exhaust gas is enclosed between the exhaust valve and the intake valve. When the process of recompression starts, the fresh gas mixture entering the cylinder during the intake stroke is heated so that it reaches its auto-ignition temperature near the top dead center of compression and achieves self-ignition combustion. Because the residual exhaust gas itself can effectively dilute the mixture and thereby control the autogenous combustion heat release rate, it can also play a role in controlling the pressure rise rate, thereby reducing the tendency for knocking.

One of the difficulties in achieving HCCI is control of the combustion process. Ignition timing of an HCCI engine is determined by the kinetics of chemical reactions, so it is sensitive to changes in the thermal state of the cylinder (exhaust gas rate, in-cylinder temperature) [3]. Exhaust gas is used to fill the cylinder to control the fresh work quality and the engine. The problem of load is more prominent: Under heavy load, the engine is prone to misfire due to additional dilution of the working fluid in the cylinder and thus cannot achieve stable HCCI combustion; under medium and high loads, its energy may not be sufficient for heating the fresh mixture and enabling stable and self-igniting ignition due to reduction in exhaust gas; furthermore, under medium- and high-load conditions, exhaust gas heat can trigger spontaneous combustion due to reduction in the diluted working medium in the cylinder, so the combustion process can be very violent, even causing serious detonation phenomena or engine damage.

In a gasoline engine equipped with a variable valve mechanism, the exhaust gas can be reused by different valve control methods to adjust the distribution and thermal state of the mixture in the cylinder, thereby affecting the combustion process to achieve controlled spontaneous combustion. There are a variety of valve control methods to achieve use of the exhaust gas: In exhaust gas recompression, the exhaust valve is closed early and internal exhaust gas recirculation is realized by the NVO angle; in the exhaust gas re-sucking strategy, the exhaust valve is opened again or the exhaust delayed during the intake process; alternatively, closing of the door to the intake stroke is delayed and the exhaust pipe gas is sucked back into the cylinder. Of these three exhaust valve control strategies, the exhaust gas recompression strategy is the simplest and most feasible, but also has some defects, particularly that of a knock phenomenon at large load. The exhaust gas re-sucking strategy can effectively expand the HCCI load ceiling because of the cooling effect of the re-absorbed exhaust gas on the fresh mixture.

To achieve controllable combustion for an HCCI engine driven by internal exhaust gas, the key is to utilize the heating, dilution, and filling effects of the residual exhaust gas. It is difficult to achieve accurate control of ignition timing and achieve stable HCCI combustion if only the relationship between actuators (valve timing, ignition timing, exhaust gas recirculation (EGR) valve) and combustion information (CA50 (crank angle position where 50% of the heat is released) indicated mean effective pressure (IMEP)) is established. To achieve precise control of the HCCI combustion phase, it is required to be able to quickly and accurately estimate the exhaust gas rate in the cylinder and use the intermediate state quantity to adjust the actuator movement to optimize the combustion phase.

For traditional gasoline engines with fixed valve parameters, the exhaust gas rate is generally determined by experimental calibration MAP; however, when the variable valve mechanism is used to control the ventilation process, the calibration workload of the MAP method is huge and this is not feasible. Modeling is therefore usually used to estimate the engine cylinder exhaust gas rate and temperature state parameters. Yun et al. [4] simplified the process by assuming that all gases remained in the cylinder during the exhaust process, adiabatic compression was performed, and calculation of the exhaust gas volume in the cylinder was accomplished by two simple thermodynamic equations. Although this algorithm is simple, its error is large. Another method [5, 6, 7, 8] calculates the exhaust gas pressure in the cylinder by arranging the exhaust pressure, temperature, or cylinder pressure sensor on the engine and replacing or modifying the temperature and pressure in the cylinder during the recompression phase based on the measured value of the sensor; however, owing to the lack of accuracy and delay of the sensor, this method is mainly used for steady-state conditions. A combustion model was established using methods such as those of Liu Yu et al. [9], Karagiorgis et al. [10] and Koehler et al. [11], where the temperature and exhaust gas volume in the cylinder were obtained by establishing parameters such as the time of ignition and mass heat release rate. Prediction based on the combustion model has high accuracy, but the calculation is complicated and it is difficult to use for real-time control.

In this work, a model to calculate gasoline engine exhaust gas rate based on sensor signals is proposed, based on the application of HCCI combustion under the exhaust gas pressure strategy. Based on the temperature signal measured by a thermocouple, an exhaust gas heat dissipation model was established to estimate the cylinder temperature at the time of closing of the exhaust valve and the exhaust gas volume in the cylinder was calculated [12]. To meet the special requirements of transient working conditions, the model considers the influence of thermocouple delay, uses the time constant of the sensor to modify the signal, and adds a dynamic working condition correction to the wall temperature when establishing the heat dissipation model.

2 Experimental and Simulation Platforms

The object of this study was a 1.8 L supercharged direct injection engine equipped with a double variable valve timing (VVT) system. Figure 1 shows the experimental bench system.
Fig. 1

Homogeneous charge compression ignition multi-cylinder experimental platform

The control system used dSPACE’s MicroAutoBox and RapidPro System [13], which can achieve control of all engine sensors and actuators; the data acquisition and combustion analysis system used NI’s PCB 9234 acquisition card and Labview software. A Kistler 6053B quartz pressure sensor was used to measure the in-cylinder pressure. The sensor was connected to a Charhe Amplifier 5011-type charge amplifier [14] through a high-resistance line to convert the charge signal to a voltage signal. The magnitude of the voltage was proportional to the pressure value. Calibration was achieved using in-cylinder pressure values [15].

The compression stroke injection has little effect on the IMEP output during combustion, but the addition of NVO injection will cause a large drop in IMEP. There are two reasons for the decrease in IMEP: First, NVO injection under large load conditions causes a large advance of the combustion starting point, which leads to the end of the top dead center and negative work; second, a large load, in which the internal temperature is high and the NVO period has a partial exotherm, leads to an increase in heat transfer loss and decrease in thermal efficiency throughout the cycle. NVO injection can maintain a very low level of cyclic fluctuation. Without NVO injection, the addition of a small amount of injection during the compression process can also greatly reduce the cyclic fluctuation. Continuing to increase the compression stroke by fuel injection has little effect on the cyclic fluctuation.

The experimental prototype was a mass-produced four-cylinder turbocharged direct-injection gasoline engine. To achieve dilution of the exhaust gas at low temperature, the intake and exhaust camshafts were replaced on the basis of the prototype. The profile of the camshaft reduces the maximum lift and duration of the intake and exhaust valves. It is easier to store a certain amount of exhaust gas in the cylinder when using the NVO strategy. The basic parameters of the experimental engine are shown in Table 1.
Table 1

Parameters of experimental engine



Cylinder form

Inline four-cylinder four-stroke


1.8 L

Compression ratio


Spray method



Booster + VVT

Intake and exhaust valve number

2 intake valve 2 exhaust valve

Range of rotation/(r/min)


Inlet valve opening duration/(°CA)


Intake valve maximum lift/mm


Exhaust valve opening duration/(°CA)


Exhaust valve maximum lift/mm


The one-dimensional engine cycle simulation software GT-POWER (Gamma Technologies, USA) is widely used in engine design and experimental guidance due to its fast calculating ability, convenient feedback control, and ability to predict engine performance.

The first injection to the cylinder is performed during the overlap of the negative valve, and the second during the intake stroke. Owing to the use of internal EGR and fuel reforming under lean conditions, a large amount of oxygen and active groups are stored in the cylinder and the temperature in the cylinder increases, which makes compression of high-octane fuel easier and weakens the HCCI engine to the intake. Sensitivity to boundary conditions (intake charge and air temperature) is required to achieve stable HCCI combustion.

A four-cylinder gasoline engine model, based on the GT-POWER simulation platform, was established. Combustion was simulated by directly introducing the experimental cumulative heat release rate. Using experimental data to calibrate the model, the validity of the exhaust gas rate model was verified. At 1500 r/min, average effective pressures of 0.2 MPa and 0.3 MPa were indicated.

The experimental and GT-POWER simulation results for the in-cylinder pressure (pi) are shown in Fig. 2. The level of rail pressure directly determines the quality of fuel atomization. Under the same injection pulse width for different rail pressures, the amount of fuel injected into the cylinder is not the same and the atomization effect of the fuel differs significantly. The rail pressure is also influenced by characteristics of the injector, so it is necessary to prevent the injector from entering the nonlinear region and causing deviation of the amount of fuel injected during the experiment. The rail pressure value under various working conditions therefore needed to be determined according to the combustion state and atomization conditions. Whether the rail pressure can be stably maintained at the set value is an important criterion for evaluating the rail pressure control algorithm. In this study, a control algorithm based on model feedforward and feedback was used for control of rail pressure.
Fig. 2

Comparison of experimental and simulation results for in-cylinder pressures (pi) of a 0.2 MPa and b 0.3 MPa

Figure 2 shows that the GT-POWER simulated cylinder pressure was consistent with the experimental cylinder pressure. It is therefore believed that this model can accurately simulate the engine running status and is sufficient as a reference for an exhaust gas rate model.

3 Residual Exhaust Gas Rate Calculation Model

Figure 3 shows the valve timing and profile under the NVO strategy. During the intake phase, the cylinder internal working fluid contains both residual exhaust gas and fresh gas mixture. The composition of the components is unknown. In contrast, during the exhaust phase, the working fluid in the cylinder has a single composition because the working mass in the cylinder from time of the exhaust valve closing to opening of the intake valve comprises only exhaust gas.
Fig. 3

Negative valve overlap strategy of homogeneous charge compression ignition engine

Under a strategy of early closing of the exhaust valve, the intake back pressure and speed are relatively low and there is a long period of front and back recirculation during the intake process, so the actual intake duration is short. Under the strategy of exhaust valve night-off, although the intake and exhaust back pressures are greater than zero when the intake valve is opened, the exhaust valve is in the open state during the larger lift, so exhaust gas is sucked into the cylinder. This amount is very large, which greatly hinders fresh charge from entering the cylinder when the small intake valve is opened at the initial stage of the intake process, which ultimately causes poor mixing in the exhaust valve side of the cylinder.

From theoretical analysis, as long as the cylinder state at any time during the recompression stage is obtained, the exhaust gas volume can be calculated by the ideal gas state equation:
$$ PV = mRT $$
where P is the measured in-cylinder pressure; V is the combustion chamber volume; m is the mass of working fluid in the cylinder; R is the gas constant; T is the temperature in the cylinder. In this study, the exhaust valve closing time was taken as the entry point. The temperature state of the cylinder at this time was estimated from the exhaust gas temperature and the residual exhaust gas volume in the cylinder was calculated according to Eq. (1).

3.1 Thermocouple Temperature Signal Correction

According to the ideal gas state equation, under the condition that the cylinder pressure and piston position are known, calculation of the residual exhaust gas volume in the cylinder when the exhaust valve is closed must first be performed to estimate the cylinder temperature at that time. However, owing to limitations of sensor technology and the difficulty of installation and arrangement, the cylinder temperature is difficult to measure directly. In comparison, exhaust temperature measurement is simple and easy. Based on the correlation between the exhaust temperature and cylinder temperature at the closing time of the exhaust valve, the temperature of the exhaust gas was measured by arranging a temperature measurement point on the exhaust pipe, from which the cylinder temperature was calculated. The exhaust gas temperature sensor was a K-type thermocouple. Because the sensor itself has a delay, direct use of the measured value to estimate the temperature in the cylinder at the closing time of the exhaust valve will inevitably cause an error. To correct this deviation, a first-order inertial link was used to process the measurement signal:
$$ \frac{{T_{\text{ex\_mes}} (s)}}{{T_{\text{ex\_act}} (s)}} = \frac{1}{\tau s + 1} $$
where Tex_mes and Tex_act are the measured and corrected values of the exhaust temperature sensor, respectively; τ is the system time constant, measured experimentally to have a value of 2.1.

3.2 Gas Flow in Exhaust Passage and Exhaust Pipe

Gas flows in the exhaust passage and exhaust pipe. Because there is a certain distance between the exhaust pipe temperature measurement point and the exhaust valve, the high-temperature gas expands when it flows through this distance. Owing to heat exchange, the measured temperature will be lower than the temperature at the exhaust valve at the time of cylinder closing. To accurately estimate the in-cylinder temperature at the time of exhaust valve closing, the corrected thermocouple temperature was used as a starting point. The basic principles of thermodynamics and heat transfer were combined, and a model of exhaust heat dissipation in the exhaust passage and exhaust pipe was established. The temperature of the exhaust gas after passing through the valve was estimated.

During HCCI combustion, the combustion phase may be inappropriate due to various factors, such as the above mentioned change of speed, which may affect the engine performance and emissions. If the combustion phase occurs too late, engine stability will degrade, cyclic fluctuations increase, and there is the possibility of causing a serious fire; if the combustion phase occurs too early, knocking is caused, NOx emissions rise sharply, the output power decreases, and fuel consumption deteriorates. It is therefore necessary to adjust the combustion phase by changing the fuel injection criteria.

In this study, the thermodynamic state of the exhaust gas from the cylinder to the temperature measurement point of the exhaust pipe was simplified into two successive processes: adiabatic expansion from the inside of the cylinder to the outlet of the exhaust valve and convective heat transfer from the outlet of the exhaust valve to the exhaust pipe. The volume of exhaust gas from the exhaust valve outlet to the exhaust pipe temperature measurement point was divided into two parts: flows in the exhaust valve and in the exhaust pipe, as shown in Fig. 4.
Fig. 4

Exhaust system cooling model

On exhaust element area dA, exhaust heat is dissipated to the wall of the exhaust port to reduce the temperature T by dT, obtained by the heat transfer relationship:
$$ \alpha_{\text{epout}} (T - T_{\text{epout}} )dA = - q_{\text{m}} c_{\text{pm}} dT $$
where qm is the exhaust mass flow rate; cpm is the exhaust specific heat capacity; Tepout is the exhaust wall temperature. The difference between the wall and exhaust temperature obtained by simulation changes with operating conditions, as shown in Fig. 5. αepout is the heat transfer coefficient of the exhaust to the exhaust passage, given by [12]:
$$ \alpha_{\text{epout}} = 3.27 \times \left( {1 - 0.797\frac{{h_{\text{v}} }}{{d_{\text{v}} }}} \right)q_{\text{m}}^{0.5} \frac{{T_{\text{B}}^{0.571} }}{{d_{\text{epout}}^{1.5} }} $$
where hv is the lift of the exhaust valve; dv is the exhaust valve diameter; TB is the average temperature of the gas in the exhaust pipe, which was approximately replaced by the measured value of the thermocouple; depout is the diameter of the exhaust port.
Fig. 5

Variation of temperature difference between exhaust gas and exhaust wall with operating conditions

Separating the integral variable from Eq. (3) yields:
$$ \int_{{T_{\text {evc}} }}^{{T_{\text {out}} }} {\frac{dT}{{T - T_{\text{epout}} }}} = \int_{(A)} { - \frac{{\alpha_{\text{epout}} }}{{q_{\text{m}} c_{\text{pm}} }}} dA$$
Solving Eq. (5) results in:
$$ T_{\text{out}} = (T_{\text{evc}} - T_{\text{epout}} )\exp \left( { - \frac{{\alpha_{\text{epout}} }}{{q_{\text{m}} c_{\text{pm}} }}} \right) + T_{\text{epout}} $$
where Tout and Tevc are the temperatures of the working fluid at the outlet and after passing through the exhaust valve, respectively. The temperature of the gas from the exit of the exhaust passage to the outlet temperature sensor was considered equivalent to that in an equal section cast iron cylindrical pipe. Exhaust gas flows in the round tube, forced convection heat exchange occurs with the inner wall of the tube, and the outer wall of the tube conducts natural convection heat exchange with the atmosphere in an infinite space. The convection heat transfer coefficient between the inner wall of the circular pipe and the exhaust gas is not only related to the mass flow rate of the gas, but also affected by its average temperature. The convective heat transfer in the circular tube was deduced according to the constant wall temperature boundary condition:
$$ T_{0} = (T_{\text{out}} - T_{\text{ep}} )\exp \left( { - \frac{{\alpha_{\text{ep}} }}{{q_{\text{m}} c_{\text{pm}} }}} \right) + T_{\text{ep}} $$
Simultaneous solution of Eqs. (5) and (6) yields:
$$ T_{\text{evc}} = \frac{{T_{0} - T_{\text{ep}} }}{{\exp \left( {\frac{{\alpha_{\text{epout}} }}{{q_{\text{m}} c_{\text{pm}} }}} \right)\exp \left( {\frac{{\alpha_{\text{ep}} }}{{q_{\text{m}} c_{\text{pm}} }}} \right)}} + \frac{{T_{\text{ep}} - T_{\text{epout}} }}{{\exp \left( {\frac{{\alpha_{\text{epout}} }}{{q_{\text{m}} c_{\text{pm}} }}} \right)}} + T_{\text{epout}} $$
where Tep is the exhaust pipe wall temperature; T0 corrects the temperature for the sensor delay; αep is the heat transfer coefficient of the exhaust gas to the exhaust pipe.
To calculate the heat transfer coefficient in the exhaust pipe, the flow state in the pipe at that time is considered. The Reynolds number is first calculated:
$$ Re = \frac{4}{{\pi \mu d_{\text{ep}} }}q_{\text{m}} $$
where μ is the dynamic viscosity of the exhaust gas, the value of which is related to the exhaust gas temperature; dep is the diameter of the exhaust pipe.

The Reynolds number determines whether the exhaust gas is in the laminar or turbulent flow state, from which the corresponding characteristic equation is selected to calculate the heat transfer Nusselt number of the exhaust pipe inner wall Nu:

when Re < 2200, the Seider–Tate formula is adopted:
$$ Nu = 1.86\left( {\frac{Re \cdot Pr}{{L/d_{\text{ep}} }}} \right)^{0.33} \left( {\frac{\mu }{{\mu_{\text{ep}} }}} \right)^{0.14} $$
when Re ≥ 2200, the Dittus–Boelter formula is used:
$$ Nu = 0.023\;Re^{0.8} \;Pr^{0.4} \frac{{T_{\text{B}} }}{{T_{\text{ep}} }}\left[ {1 + \left( {\frac{{d_{\text{ep}} }}{L}} \right)^{0.7} } \right] $$
where Pr is the Prandtl number, taken as 0.71; L is the exhaust pipe length; μep is the dynamic viscosity of the exhaust gas at the temperature of the inner wall of the exhaust pipe.
The convection heat transfer coefficient of the exhaust gas and the inner wall of the exhaust pipe is given by:
$$ \alpha_{\text{ep}} = Nu\frac{{\lambda_{\text{f}} }}{{d_{\text{ep}} }} $$
where λf is the thermal conductivity of the exhaust gas, the value of which is related to the exhaust gas temperature.
μ, μep, and λf can be calculated using Eqs. (13) to (15), respectively [13]:
$$ \mu \,{ = 0} . 3 5 5\times 1 0^{ - 6} \times T_{\text{B}}^{0.679} $$
$$ \mu_{\text{ep}}\, { = 0} . 3 5 5\times 1 0^{ - 6} \times T_{\text{ep}}^{0.679} $$
$$ \lambda_{\text{f}}\, { = 2} . 0 2\times 1 0^{ - 4} \times T_{\text{B}}^{0.837} $$

3.3 Adiabatic Expansion of Gas Flowing Through Exhaust Valves

When the exhaust valve is closed, the exhaust gas passes from the cylinder to the outlet of the exhaust valve and undergoes an adiabatic expansion process [14]:
$$ \frac{{T_{\text{\_evc}} }}{{T_{\text{evc}} }} = \left( {\frac{{p_{\text{ex}} }}{{p_{\text{evc}} }}} \right)^{{\frac{\gamma - 1}{\gamma }}} $$
where T_evc is the exhaust gas temperature at the outlet of the exhaust valve when it is closed; Tevc is the temperature in the cylinder at the time of closing of the exhaust valve; pex is the pressure in the exhaust pipe; pevc is the pressure in the cylinder at the time of closing of the exhaust valve; γ is the adiabatic index. The pressures in the exhaust pipe and cylinder can be measured directly by the sensor.

Use of the thermocouple temperature and exhaust pressure signal to estimate the cylinder temperature at the exhaust valve closing time Tevc was thereby achieved. The pressure in the cylinder at the time of closing the exhaust valve was obtained from the cylinder pressure sensor. The gas constant Rg was obtained by the method given in [15], and the exhaust gas volume in the cylinder was calculated. In combination with the known fuel injection and intake air quantities, the residual exhaust gas rate in the cylinder was calculated.

4 Model Validation

The residual exhaust gas calculation model was established based on the exhaust temperature signal. Because experimental measurement of the residual exhaust gas is complex, the CO2 content of the cylinder needs to be sampled and analyzed during the compression process. Accuracy of this measurement is also influenced by the engine operating status; therefore, this study only compared the calculation results of the algorithm with those of the GT-POWER model.

The residual exhaust gas algorithm was verified for both steady-state and transient conditions. The steady-state operating conditions comprised four operating points at 1500 r/min, five at 2000 r/min, and five at 2500 r/min, giving 14 different operating point points in total. The load range was 0.2–0.3 MPa. Verification of transient conditions was undertaken at 1500 r/min, where the load was adjusted by selecting different exhaust valve closing times.

4.1 Steady-State Verification

The key problem in the calculation of exhaust gas volume is correction of the in-cylinder temperature at the time that the exhaust valve closes. For this reason, the exhaust gas temperature used to estimate the temperature at the time of closing of the valve. The results are shown in Fig. 6.
Fig. 6

Verification of in-cylinder temperature at time of exhaust valve closure

Figure 6 shows that, under different operating conditions, the algorithm results of the in-cylinder temperature at the time of exhaust valve closure were generally low. The maximum relative error was 12.1%, and the average relative error was 6.9%. The main reason is the wall temperature of the exhaust passage. The accuracy of this measurement is affected by the operating state of the engine; therefore, an error occurs between the calculated and measured values. In fact, for prediction of the combustion phase, more attention is paid to the in-cylinder temperature when combustion does not occur. The error in estimating the cylinder temperature in this phase is smaller. Estimation of the temperature after combustion does not fully consider the influence of heat dissipation. Although a wall temperature correction was added, it was difficult to obtain the true value, which ultimately leads to estimation deviation of the exhaust valve closing time. However, because calculation of the amount of exhaust gas in the cylinder uses thermodynamic temperature in the calculation, the value generally exceeds 750 K, so the above calculation bias on the exhaust gas results has limited impact. It is precisely for this reason that the correction algorithm for the in-cylinder temperature at the time of closing the exhaust valve can be further simplified to facilitate real-time calculation.

Based on the state of the exhaust valve at the closing time of the cylinder, the amount of gas in the cylinder was easily calculated according to the ideal gas state equation, as shown in Fig. 7. Owing to deviation of the temperature estimation, there is a certain deviation between the calculated and simulation results for the exhaust gas volume. The maximum relative error was 11.9%, and the average relative error was 7.2%. The ultimate goal of the algorithm is to achieve control of combustion phasing, so it is more important to estimate the exhaust gas rate.
Fig. 7

Verification results for exhaust gas volume

Under the premise of known fuel injection and intake air amounts, the exhaust gas rate in the cylinder was readily obtained by calculating the amount of exhaust gas, as shown in Fig. 8. As the EGR rate increases, the maximum burst pressure in the cylinder decreases, so the combustion heat release rate slows down, the high-temperature ignition time is delayed, and the combustion duration is prolonged. EGR is therefore an effective means to control the ignition timing and reaction rate of HCCI. However, this study found that when the octane number was low, such as 0 or 40, even if the EGR rate was as high as 70%, the ignition time in the high-temperature stage occurred before top dead center, resulting in an increase in negative compression work and decrease in IMEP. Lower research octane number (RON) fuels are not suitable for HCCI combustion. Figure 8 shows that the calculation error for the exhaust gas rate is small and the absolute error was within 3.5%. Under heavy load conditions, the amount of exhaust gas in the cylinder is small. Even if there is a large deviation in the calculation of the amount of gas, the impact on the exhaust gas rate in the cylinder is minimal and the effect on the thermal state of the cylinder is also rather weak. At small loads, the cylinder contains a large amount of residual exhaust gas so, for the same load, the amount of exhaust gas will be equivalent to a fresh charge and slight deviation of the exhaust gas rate has little effect on the deviation of the gas calculation and thermal state of the cylinder. Under intermediate loads, the thermal state of the cylinder has a strong influence: Fig. 8 shows that when the operating points are distributed near such loads, the calculation results have high accuracy.
Fig. 8

Verification results for exhaust gas rate

HCCI combustion is achieved by direct injection into the cylinder. In all working conditions, the throttle is fully open and the air–fuel ratio can be varied within a wide range. Therefore, load regulation of HCCI combustion differs from that of a gasoline engine, the combustion characteristics of which are closer to those of a diesel engine. Unlike a diesel engine, when the HCCI combustion load is adjusted by gasoline direct injection (GDI), the cycle, the amount of oil supplied to the ring, and the fuel injection strategy must be changed to meet the combustion requirements of different loads.

4.2 Transient Operating Condition Verification

The transient condition was obtained at 1500 r/min by adjusting the closing time of the exhaust valve. When the exhaust valve is closed, the amount of exhaust gas in the cylinder can be adjusted to regulate the load, as shown in Fig. 9. The estimated exhaust gas and residual exhaust gas rates are shown in Figs. 10 and 11, respectively, for variable conditions. Figure 11 shows that the algorithm can follow changes in the residual exhaust gas rate well during transition of operating conditions and can be used for transient control.
Fig. 9

Exhaust valve closing timing adjustment under transient conditions

Fig. 10

Estimation of exhaust gas volume under transient conditions

Fig. 11

Estimation of residual exhaust gas rate under transient conditions

At 1500 r/min, when the engine load changed from 5 bar to 4 bar and then back to 5 bar, the target value of CA50 varied from 8°CA to 2.5°CA to 8°CA after top dead center (ATDC). The control effect of combustion phase CA50 under changing IMEP conditions is shown in Fig. 12.
Fig. 12

Effect of CA50 control under variable load conditions

Under steady-state conditions, the actual CA50 tracked the expected value. When the load dropped, the CA50 expectation value ranged from 8 to 2.5°CA ATDC, and it took about 13 work cycles to complete tracking to the expected value. When the load rose, the expected value of CA50 varied from 2.5 to 8°CA ATDC, and approximately 10 engine duty cycles were required to achieve this value. The response time was slightly delayed for control, but the absolute value of the deviation between the actual and expected values of CA50 was within 2°CA during load transitions. Considering this aspect, the CA50 control effect is good and the control error is within the allowable range.

5 Conclusions and Outlook

This paper focuses on the application problems in HCCI combustion control and analyzes the influence of negative valve overlap angle control strategy on fuel distribution and spontaneous combustion in the cylinder. The following conclusions are obtained through experimental simulation analysis.

(1) To meet the requirements of transient conditions, the measured temperature signal of the exhaust gas temperature sensor was processed to compensate for thermocouple delay.

(2) The flow process of the exhaust gas in the exhaust passage and exhaust pipe was physically modeled and the corrected exhaust temperature signal was used to estimate the temperature at which the gas flowed out of the exhaust valve. The calculated deviation was within 100 K.

(3) The gas flow through the exhaust valve was modeled as an adiabatic expansion process, which was used to correct the cylinder temperature at the closing time of the exhaust valve, further reducing the estimation error.

(4) The calculation results for the exhaust gas volume and residual exhaust gas rate were compared with GT-POWER simulation results. The average deviation of exhaust gas volume was 7.2%. The relative error of residual exhaust gas rate was within 3.5%, which has adequate accuracy and meets the control requirements.

It should be pointed out that correction of the wall temperature in this study was simply performed by means of simulation. If higher accuracy is required, an estimation model of the wall temperature can be established. The cylinder pressure sensor signal was used to calculate the amount of exhaust gas in the cylinder. When the pressure at the time of closing the exhaust valve is used in actual applications, only the cylinder pressure at a certain time is substituted for the calculation, which may introduce noise and cause deviations in the results. The exhaust valve may also be obtained by modeling the closed cylinder pressure with time.



Hebei Provincial Science and Technology Research Project (Grant No. Z2015092), Langfang Science and Technology Bureau High-Tech Support Project (Grant No. 2016011018), and Yanjing Institute of Technology Research Project (Grant No. 2017YITSRF105) are thanked for joint funding.


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Copyright information

© Society of Automotive Engineers of China (SAE-China) 2019

Authors and Affiliations

  1. 1.College of Engineering, Yanjing Institute of TechnologyLangfangChina

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