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Future specification of automotive LPG fuels for modern turbocharged DI SI engines with today's high pressure fuel pumps.


Due to the worldwide tightening of C[O.sub.2] legislation and increasing requirements regarding a reduction of pollutant emissions, internal combustion engines have gone through a tremendous progress during the last decades. Therefore, modern spark ignition engines feature a large number of innovative technologies, such as direct injection, exhaust gas turbocharging, variable valve trains, thermal management, start & stop, intelligent generator management systems and friction reduction, for instance by using demand-controlled oil and water pumps. However, to achieve the drastic C[O.sub.2] reduction, which is required to meet future targets, further optimization has to be made in combustion processes. Technologies like exhaust gas recirculation, lean combustion, innovative ignition systems, water injection and variable compression ratio have to be considered.

Moreover, an enormous potential in C[O.sub.2] reduction can be realized through the use of alternative fuels. Gaseous fuels like LPG are a sensible short- and mid-term alternative. In Europe, LPG is already the most-utilized alternative fuel for automotive use with an already adequate core infrastructure in place [1, 2, 3]. According to the European fuel standard EN 589 [4], automotive LPG mainly consists of [C.sub.3]/[C.sub.4] paraffins, but also mono-olefins are allowed as fuel components. In changing phase from gaseous to liquid state, these molecules undergo a significant volume reduction, which results in considerable benefits in terms of transportation and storage aspects. From a combustion perspective, LPG is an interesting alternative fuel for modern spark ignition engines, due to its high knock and pre-ignition resistance compared to gasoline, which is because of the significantly enhanced auto-ignition characteristics in the low and mid temperature regime, especially for propane [5]. This enables higher compression ratios and therefore higher combustion efficiencies. The thermodynamic efficiency of the LPG fueled engine can be improved by direct injection of LPG. This concept allows high torques at low engine speeds in comparison to external LPG mixture formation systems [6]. Furthermore, LPG DI reveals reduced pollutant emission characteristics. Gaseous emissions, such as HC, can be reduced significantly [7, 6] mainly due to severe flash-boiling and a better mixture formation of LPG compared to gasoline [8]. To meet future emissions legislation for DI SI engines, it is noteworthy that direct injection of LPG offers tremendous particulate emissions reduction [4, 7, 9, 10, 11, 12, 13]. Besides all these numerous positive aspects, the biggest challenge for the direct injection of LPG is the hot fuel handling. Considering state-of-the-art gasoline direct injection engines, high fuel pressure is built up using mechanically driven high pressure fuel pumps upstream of the fuel rail. When using such a high pressure fuel pump to pressurize LPG fuels, the high vapor pressure, especially of propane and propene, can result in evaporation of LPG upstream or inside the HPP, since the thermal conditions of these pumps are considerably influenced by heat transfer from the engine. As a result, the engine stalls, due to the fact that the HPP cannot compress a gas or a vapor. Furthermore, the supercritical state can be reached. This physical state is defined by the critical point, in which the saturated liquid and saturated vapor curves meet. It is heavily dependent on the components of the fuel mixture. Pure propane/propene transits to the supercritical state at a pressure of 42.5/46.7 bar, which is lower than usual fuel pressures in the high pressure part of direct injection systems. The corresponding critical temperatures are 96.7[degrees]C/92.4[degrees]C. Therefore, this physical state is particularly relevant under hot soak and hot idle conditions, in which such high fuel temperatures within the fuel system can be reached. As a result of a rapid fuel density drop close to supercritical point, the high pressure pump is not able to keep the rail pressure constant and the engine stalls. Countermeasures to avoid LPG boiling and the supercritical state in the fuel system at engine operation and hot soaks are required for an increased usage of LPG direct injection applications in the future. On the one hand, cooling measures can be considered, which keep LPG in the liquid state. On the other hand, the fuel pressure at inlet of the HPP can be increased, in order to counteract the high compressibility at high fuel temperatures. Moreover, an optimized HPP design, such as an increased compression ratio of the pump, can be reasonable. However, there is still little knowledge about the effectiveness and suitability of one of these methods or even a combination of them, depending on the gas mixture components.

This paper is believed to be the first study on LPG hot soak and hot idle investigations performed on a modern turbocharged 4-cylinder DI SI in a climate cell. In order to enable wider use of LPG fuels and to assist the process of standardization, a limitation for maximum content of [C.sub.3] fuel components will be recommended. For these purposes, four different LPG fuel formulations, which all are in accordance to the current LPG fuel standard EN 589, have been tested. Besides engine warm-up and hot soak tests, variations of the fuel pressure at inlet of the HPP have been investigated. Comparisons of time durations to build up fuel rail pressure are made as a function of the [C.sub.3] content. Furthermore, the injection and combustion behavior is studied in stationary tests on a single cylinder engine for different physical states.


Table 1 summarizes properties of the fuels investigated in this study. All LPG fuels are market relevant fuels according EN 589 [4]. A conventional RON 95 E5 according to EN 228 [14] serves as reference fuel for all investigations. The selection of the LPG fuels was made in order to cover a wide range of possible fuel components. LPG 1 has the highest [C.sub.3] content according to EN 589, LPG 2 represents a typical winter grade LPG, whereas LPG 3 has a high propane content and LPG 4 a high butane content. The last two fuel formulations represent homologation fuels according to [15].

The motor octane number of the LPG fuels is calculated based on the mass fractions according to EN 589, while the methane number is calculated in accordance to the so called AVL algorithm partially standardized in DIN 51624 [16] and established in [17, 18, 19]. Comparing the critical points, the strong dependence on the fuel mixture components is evident. Isobutane/n-Butane has a critical temperature of 134.7[degrees]C/152[degrees]C and a critical pressure of 37/38 bar. Therefore, the higher the content of [C.sub.4] paraffins, the higher is the critical temperature and the critical pressure. Furthermore, compared to gasoline, the lower density of all LPG fuels can be noted. While the heating value is higher for the LPG fuels, the lower densities lead to a lower volumetric heating value in comparison to conventional gasoline. Due to high vapor pressures, high propane and especially high propene contents are reflected in high enthalpies of vaporization approximated via Clausius-Clayperon equation.


Series Production Engine Hardware

As a modern downsized engine with direct injection, the 1.6 l inline turbocharged gasoline engine of the Ford Motor Company was chosen, which is presented in detail in previous publications [5]. A short summary of the technical data of the engine is given in Table 2.

The engine is quantity controlled with a stoichiometric combustion system and a conventional valve train with variable cam phasing on intake and exhaust side. The four valve combustion chamber roof with central arranged injector ([Q.sub.stat] = 12.5 [cm.sup.3]/s at 10 MPa with n-Heptane) and spark plug aligned with the crankshaft axis has a compression ratio of CR = 10 [20]. For the control of the engine, an RCP hardware system, consisting of a dSPACE MicroAutoBox II and a RapidPro was chosen. Measurements were performed with the ControlDesk NG experiment software. A base engine model was used employing a calibration based on the series ECU calibration, due to the fact that the main focus of the investigations was set on the behavior of the injection system and the fundamental functionality of the HPP under high fuel temperatures.

For the direct injection of LPG, the conventional components of the gasoline high pressure injection system of the engine were used (high pressure pump, fuel rail, fuel injectors). No modifications of the fuel-carrying components have been made, except for the fact that a thermocouple has been installed in the fuel line between the HPP and the fuel rail, in order to be able to measure the fuel temperature at rail inlet. Further thermocouples were positioned at the surface of the HPP, the fuel rail and the solenoid housing of the injector of cylinder 3 (counted from flywheel side). An additional middle pressure pump has been applied upstream the HPP in order to be able to vary the HPP inlet pressure. To keep the fuel temperature at the inlet of the HPP constant, an electrical heater was mounted to the fuel line upstream the HPP with a distance of about 20 cm to the pump to avoid direct heat transfer. For fuel changes, the system was completely drained and flushed with the new fuel to minimize possible contamination by other fuels.

Climate Cell Instrumentation

For the hot soak investigations, the engine was mounted into an idle rig. The cylinder pressure was measured with Kistler 6041 pressure transducers mounted flush in the combustion chamber roof of each cylinder. Sampling was performed via Kistler 5011 charge amplifiers and a FEVIS 3.2.4. Dynamic intake and exhaust gas pressures were measured with Kistler 4045 A5 pressure transducers and sampled in 1[degrees] CA resolution. Measurements of static pressures and temperatures were performed with piezoresistive pressure transmitters and thermocouples. Engine oil and coolant was not conditioned by any external conditioning system. The coolant circuit contains two controllable valves and a thermostat, in order to allow different cooling strategies. The required air mass flow to cool down the coolant was delivered by an electric fan. The fan was located in front of the cooler in a manner that a coolant temperature of up to 120[degrees]C could be reached. The temperature in the climate cell was set-up to ~30-35[degrees]C to create warm ambient conditions, representing a hot summer day in northern Europe. Therefore, the temperature of the intake air, which is sucked into the intake snorkel, equals the temperature of the air inside the climate cell.

Experimental Procedure

All tests were performed in homogenous operation mode with single injection into the intake stroke. Valve timing was set to maximum negative valve overlap of 38[degrees] CA. Pre-investigations had shown that a constant idle phase at high engine speed of 4000 1/min, followed by a hard shutdown, results in fuel temperatures of only ~60[degrees]C, ~25 minutes after shutdown. This is also known as hot soak because heat is "soaking" from the engine to the fuel. Higher fuel temperatures were found to be a result of long constant idle phases. Therefore, the engine was warmed up at a constant idle speed of 800 1/min until the high temperature level stabilized. The results of this warm-up phase are presented in the result chapter exemplarily for LPG 1. This warm-up was applied right before each HPP inlet pressure variation or hot start.

Single Cylinder Research Engine


The steady state idle tests were conducted on a homogeneously operated, DI SI single cylinder research engine, presented in previous publications [5, 21]. Injector and spark plug are placed in central cross position in the combustion chamber roof, which means that the spark plug is installed between the exhaust valves, while the injector is installed between the intake valves. The tumble ratio of the symmetrical filling ports was increased by an additional tumble flap up to 4.1 [22]. Table 3 summarizes the most important engine parameters.

The direct-injection gasoline injector ([Q.sub.stat] = 13.5 [cm.sup.3]/s at 10 MPa with n-Heptane) is a six-hole solenoid actuated injector and was also used for LPG direct injection. Concerning the fuel system, the HPP used was volume flow controlled. To reach the supercritical state, the fuel was heated up by an electrical heater mounted to the complete fuel line downstream the high pressure pump including the fuel rail and pipes upstream the injector.

Test Bench Instrumentation

The single cylinder engine instrumentation was similar to that used for the series production engine and includes only minor differences, such as the cylinder pressure transducer Kistler A6043 A100. In total, 200 consecutive cycles were measured and averaged. The engine was coupled to an eddy-current brake and an electric motor to maintain the desired engine speed with an accuracy of [+ or -] 1 1/min. The intake air mass flow was measured with an ultrasonic air mass meter. The gaseous exhaust gas composition was determined from a partial mass flow of exhaust gases. The exhaust gas sample was fed to the following emission analyzers via a transfer line heated to 193[degrees]C:

* HC: Flame ionization detector (Rosemount NGA 2000, calibrated with [C.sub.3][H.sub.8])

* [O.sub.2]: Paramagnetic oxygen analyzer (Rosemount NGA 2000)

* CO and C[O.sub.2]: Infrared gas analyzer (Rosemount NGA 2000)

* N[O.sub.x]: Chemiluminescence analyzer (Eco Physics 700 EL ht)

Experimental Procedure

The investigation on the single cylinder engine focuses on the injection, combustion as well as the emission behavior of LPG direct injection during idle. Therefore, the engine speed and the load were set to 800 1/min and 0.7 bar IMEP as for the hot idle investigations. The selected settings are summarized in Table 4.

A rail pressure of only 40/50 bar was chosen in order to investigate the effect of decreasing density close to the critical point more clearly. Valve timing was set to maximum negative valve overlap of 45[degrees]CA. With these settings spark timing sweeps for different fuel temperatures were performed for all LPG fuels and compared to the reference fuel.



Preliminary investigations had shown, that long, constant idle operation represents the most severe operation mode regarding changes of physical states of LPG fuel components. During such a period of constant idle, the engine got fully warmed up and a rail pressure drop was observed at certain fuel temperatures inside the HPP dependent on the LPG fuel. Due to the fact, that the critical temperatures of the tested LPG fuels differ significantly, the rail pressure drop is expected to occur at different fuel temperatures during the investigated warm-up phases. Figure 1 exemplarily shows the warm-up phase with LPG 1.

This fuel can be seen as the "worst case" fuel with regard to hot idle conditions due to the high [C.sub.3] content. It is evident, that high oil and coolant temperatures in the range of 110[degrees]C are reached during constant idling at 800 1/min. At 337 seconds the thermostat opens, cooled water out of the radiator enters the engine, while hot water flows out of it. This results in a slight increase of the coolant temperature at engine outlet. The highest surface temperatures during the warm-up phase were observed at the injector solenoid housing (cylinder 3). At the inlet of the HPP, the fuel temperature was conditioned to 50[degrees]C, whereas the fuel pressure was set to 50 bar. Therefore, the physical state is either liquid or supercritical depending on the fuel temperature and outgassing of the fuel is prevented, neglecting occurrence of local cavitation inside the high pressure fuel system. While the fuel flows through the HPP and the pipe between the HPP and the rail, it is heated through the heat transfer from the engine and the fuel temperature increases steadily with continuous engine warm-up. This can be seen with regard to the fuel temperature at the inlet of the fuel rail, which reached 94[degrees]C at the end of the measurement. The required fuel rail pressure was set to 110 bar. In Figure 1, the fuel pressure increase from the low pressure level of 50 bar at the inlet of the HPP to the rail pressure is illustrated exemplarily by three arrows at different times in the diagram of the fuel pressure. Up to 1125 seconds, the HPP is able to deliver the required rail pressure. However, a clear rail pressure drop can be seen, which starts at this time and continues till the end of the measurement. This can be explained using Figure 2, in which the isobaric density diagram of propene is shown.

Propene is suitable for an estimation of a change of state of LPG 1 due to the fact, that 50.2 % (m/m) propene is contained in this fuel and the critical temperature of propene is lower than that of propane. Besides the isobars and the two-phase region, which ends in the critical point at 92.4[degrees]C, the isentropes are depicted. It is obvious, that the isobars of pressures higher than 46.6 bar show a rather constant density gradient in a temperature range of 0[degrees]C to ~80[degrees]C. However, as temperature increases the isobars of 46.6 bar to 80 bar show a heavily nonlinear behavior between 80[degrees]C and 140[degrees]C. For instance, the density changes about -5 %/[degrees]C at 50 bar and 100[degrees]C. The three highlighted arrows in this Figure correspond to the arrows of the warm-up in Figure 1. They represent three estimated changes of state of the fuel due to compression by the HPP. The thermodynamic changes of state are estimated under the following assumptions. The compression is considered to be isentropic. Furthermore, at the beginning of the compression the state of the fuel is described by the estimated fuel temperature inside the HPP and the pressure at the inlet of it. The temperature inside the HPP was estimated to be an average from oil and HPP surface temperature, since the HPP is exposed to the temperature of the climate cell and its piston is driven by fuel pump cams of the exhaust camshaft, which are in turn lubricated by the warm oil of the engine. Thus, the fuel is heated up severely inside the pumping chamber, because the upper part of the piston is in contact with the fuel, while the lower part is in contact with the warm engine oil. Moreover, the fuel flows through the inner drillings of the HPP, which wall temperatures are influenced by the heat transfer from the valve cover. Due to this fuel heating process the estimated temperature of LPG 1 inside the HPP is 70.5[degrees]C at the beginning of the warm-up process. The changes of state at this temperature correspond to the left arrow in Figure 1 and Figure 2. Thereby, the HPP is able to pressurize the fuel from 50 bar to the required rail pressure of 110 bar. The middle arrow illustrates the change of state at the beginning of the rail pressure drop at an estimated fuel temperature inside the HPP of 95[degrees]C. Further temperature increase results in progressing rail pressure drop. Finally, the right arrow shows the change of state at an estimated fuel temperature inside the HPP of 100.5[degrees]C. At this time, the HPP is only able to compress the fuel from 50 bar to 89 bar. The rail pressure drop with increasing temperatures can be explained by two reasons. Firstly, the strong decrease of density results in a reduced filling of the cylinder of the HPP. The reduced filling in turn limits the maximum deliverable mass flow to the fuel rail. Secondly, the isentropic compressibility is increasing as it can be seen by the increasing distance between each isobar at temperatures higher than the critical one. The isentropic compressibility [K.sub.S] limits the pressure at the end of compression. It is defined as [25]:

[K.sub.S] = -1/V x [([partial derivative]V/[partial derivative]p).sub.S] (1)

In order to examine the effect of an increasing isentropic compressibility with increasing temperatures according to Figure 2, an exemplary calculation will be presented. For this purpose, the maximum achievable pressure at the end of compression is determined for two different temperatures, 80[degrees]C and 100[degrees]C, in case of propene. The temperatures refer to the beginning of compression.

The initial pressure is 50 bar. Assumptions are an isentropic compression, a constant pump compression ratio [[epsilon].sub.HPP] of 1.15, a constant pump displacement and a constant mass of substance. At a temperature of 80[degrees]C and a pressure of 50 bar, the density of propene is 395.14 kg/[m.sup.3]. In Figure 2 this state can be found at the end of the linear density range of this isobar. At the end of compression the density [rho] can be calculated to 454.3 kg/[m.sup.3] with the following equation.

[[rho].sub.2] = [[epsilon].sub.HPP] x [[rho].sub.1] (2)

Subscript 1 refers to the state at the beginning of compression, whereas subscript 2 indicates the state at its end. Based on the assumptions the pressure at the end of compression is found to be 200 bar. Thus, a pressure increase [DELTA][p.sub.80[degrees]C] of 150 bar is reached. For a temperature of 100[degrees]C and a pressure of 50 bar at the beginning of compression, the density of propene is 156.4 kg/[m.sup.3]. This state is located in the heavily nonlinear range of density in Figure 2. At the end of compression, according to Equation 2 the density is 179.9 kg/ [m.sup.3]. Based on the stated assumptions, the pressure at the end of compression is only 58 bar, which results in a pressure increase [DELTA][p.sub.100[degrees]C] of only 8 bar in this case. However, the HPP is able to compress the fuel from 50 bar to 89 bar for an estimated temperature in the HPP of 100[degrees]C. The main reasons for this difference are that the real compression is not isentropic and LPG 1 also contains 48.3 % (m/m) propane. Also, the estimation of the fuel temperature inside the HPP lacks on accuracy. For a higher accuracy of determination of this temperature, a thermocouple was applied into the pumping chamber of the HPP afterwards. Unfortunately, strong leakage occurred and the tests had to be stopped.

In order to show the relative increase of the isentropic compressibility, a ratio is determined in the following equation. As estimation, the interval from the beginning to the end of compression is considered.


The estimation shows an increase of the isentropic compressibility by a factor of ~19 for a pressure of 50 bar and an increase of temperature from 80[degrees]C to 100[degrees]C for pure propene. The example calculation shown in Equation 3 indicates that the isentropic compressibility limits the maximum achievable pressure increase inside the HPP.

For the other fuels, the rail pressure drop during the warm-up phase was expected to occur at higher fuel temperatures according to their components and their critical temperatures. The results of all investigated warm-up phases are summarized in Figure 3 in terms of the estimated fuel temperature inside the HPP to the time, when the first rail pressure drop was observed during the measurement. The HPP inlet pressure was always kept constant to 45-50 bar. It is evident, that the estimated temperatures are very close to the critical temperatures of the fuels according Table 1.

A decreasing [C.sub.3] content results in an increasing estimated fuel temperature inside the HPP, when a rail pressure drop can be observed. For LPG 3 ([T.sub.LPGinHPP,estimated] = 104[degrees]C) and LPG 1 the fuel temperature at the inlet of the HPP was set to ~50[degrees]C. In case of LPG 2, the temperature at the inlet of the HPP had to be increased continuously up to 88[degrees]C in order to observe a rail pressure drop. For LPG 4, even an increase up to -150[degrees]C was needed, due to the high critical temperature of this fuel. This warm-up phase can be found in the appendix as an example for a fuel with high [C.sub.4] content. Overall Figure 3 shows a linear relation between the estimated fuel temperature inside the HPP at first rail pressure drop during warm-up phase and the [C.sub.3] content of the fuel.

Besides design optimization aiming at an increase in compression ratio of the HPP, one possible measure to increase the reachable rail pressure at high engine and fuel temperatures is an increase of the HPP inlet pressure. This possibility has been analyzed and will be discussed in the following section.

Fuel Pressure Variation at the Inlet of the HPP

The procedure of the experiment consists of a variation of the HPP inlet pressure at high fuel and engine temperatures. The results for LPG 1 are depicted in Figure 4. The engine speed, fuel temperatures, fuel pressures and the injection duration of each cylinder is shown as a function of time. The fuel and engine temperature are set prior to the experiment by the warm-up phase at 800 1/min, which was shown in Figure 1. The fuel temperature inside the HPP is estimated to be 107[degrees]C.

All temperatures remain fairly constant during the experiment. The fuel temperature in the HPP inlet is conditioned to 55[degrees]C and the fuel temperature at the rail inlet stays between 95[degrees]C and 106[degrees]C. Therefore, it can be assumed that LPG 1 is injected in supercritical state into the combustion chamber. During the experiment, the HPP inlet pressure is reduced incrementally from 70 bar to 50 bar and then increased to 70 bar. The set value of the rail pressure is 131 bar during the entire measurement. Clearly the HPP is not able to deliver the desired rail pressure, if the HPP inlet pressure drops below 70 bar. The rail pressure follows the incremental pressure variation of the HPP inlet pressure. As already used for the warm-up phase, three arrows in the fuel pressure diagram illustrate again the pressure increase by the HPP. A comparison of their lengths shows that the possible pressure increase by the HPP decreases with the reduction of the HPP inlet pressure. This can be explained using the density temperature diagram of propene again, which is shown in Figure 5 with the three arrows corresponding to those in Figure 4. At three time stamps they represent estimated changes in thermodynamic state of LPG 1 during the experiment of the HPP inlet pressure variation. The thermodynamic changes are again estimated as isentropic compressions. The temperature of LPG 1 at the beginning of the compression is assumed to be the estimated fuel temperature in the HPP of 107[degrees]C. It is shown as a dashed isotherm in the following Figure.

Depending on the pressure, this temperature is located in the range of highly nonlinear density behavior. At a HPP inlet pressure of 70 bar, the HPP can pressurize the fuel to the required rail pressure of 131 bar (left arrow in Figure 4). By stepwise decreasing the HPP inlet pressure to 60 bar, the reachable rail pressure decreases to 102 bar (middle arrow in Figure 4). With a further decrease to 50 bar, only a rail pressure of 66 bar can finally be reached (right arrow in Figure 4). These pressures can be found at the end of the arrows in Figure 5. Lowering the HPP inlet pressure results in a severe decrease of density and an increasing compressibility. Therefore, the HPP is not able to reach the required rail pressure with low inlet pressures close to the supercritical point.

For evaluation of an LPG DI concept using a state-of the-art high pressure fuel pump, the necessary HPP inlet pressure for reaching a rail pressure of 100 bar is used, based on the fuel pressure variation. For LPG 1, this corresponds to a minimum pressure of 60 bar, read off at the second arrow in the fuel pressure diagram in Figure 4. Increasing the HPP inlet pressure in the second part of the experiment again, results in gas dynamic effects and rail pressure fluctuations. Therefore, the rail pressure of -110 bar, reached at 280 seconds for a HPP inlet pressure of 60 bar, was not considered to be steady state.

Dependent on the fuel components, different required HPP inlet pressures were expected for each LPG fuel. In contrast to LPG 1 with almost 100 % (m/m) [C.sub.3] content, LPG 4 contains a high content of [C.sub.4] paraffins. In order to assess the required HPP inlet pressure for a fuel with high [C.sub.4] content, the inlet pressure variation with LPG 4 is presented in Figure 6. Prior to the measurement, the fuel and engine temperature are set by the warm-up phase at 800 1/min as already shown for LPG 1. The fuel temperature inside the HPP is estimated to be 108.5[degrees]C, as average of the HPP surface and the oil temperature. All temperatures remain fairly constant during the experiment. The fuel temperature at the HPP inlet is conditioned to 55[degrees]C and the fuel temperature at the rail inlet stays between 96[degrees]C and 100[degrees]C. In contrast to LPG 1, it can be assumed that LPG 4 is injected in liquid state into the cylinder due to its high critical temperature of 133.6[degrees]C. During the test, the HPP inlet pressure is reduced incrementally from 50 bar to 25 bar and then increased to 60 bar. The rail pressure set value is again 131 bar as already set for LPG 1. The HPP is able to deliver this rail pressure for an inlet pressure of 30 bar. Further decrease of the inlet pressure results in a rail pressure drop, as it can be seen from 70 seconds. At 100 seconds, the rail pressure drops to 36 bar, while the pressure at the HPPs inlet is only 24 bar. The pressure increase achieved by the HPP is illustrated with an arrow at this time. As a result of the rail pressure drop, the injection durations increase steeply as the required fuel mass cannot be delivered anymore. Thus, the engine speed decreases from 90 seconds and the engine stalls. During the following engine restart at 116 seconds, the desired rail pressure is reached with 30 bar inlet pressure. This is illustrated by the right arrow. The injection duration is again ~0.8 ms.

The low rail pressure at 100 seconds is caused by outgassing of the fuel inside the HPP in the two-phase region, since the critical temperature of this mixture is not reached during the measurement. This can be explained using the density temperature diagram of n-butane, which is shown in Figure 7.

In comparison to the density diagram of propene it is evident, that the high critical temperature of n-butane shifts the range of the strong density drop and high compressibility to temperatures of ~140[degrees]C. The temperature of LPG 4 at the beginning of the compression is assumed to be the estimated fuel temperature in the HPP of 108.5[degrees]C. It is shown as a dashed isotherm. For estimation, an isentropic compression is assumed again. The illustrated arrow corresponds to the right arrow in Figure 6. It represents the estimated changes in thermodynamic state of LPG 4 during steady state idling with 30 bar inlet pressure of the HPP. Since this pressure is already very close at the boiling curve at 108.5[degrees]C, a further reduction of the HPP inlet pressure results in vapor lock and consequently engine stall.

In direct comparison with the illustrated changes of state of LPG 1 in the isobaric density diagram of propene in Figure 5, the compression ratio of the HPP seems to be smaller. This can be attributed to the fact, that LPG 4 contains a relatively high [C.sub.3] content of 24.3 % (m/m), which results in a reduced density and an increased compressibility of the fuel, compared to pure n-butane.

Using a rail pressure of 100 bar as requirement for an LPG DI concept again, a minimum pressure of 30 bar, based on the fuel pressure variation in Figure 6, can be stated for LPG 4. Comparing the results of all pressure variations of the tested LPG fuels, it is evident that the required inlet pressure strongly depends on the [C.sub.3] content of the LPG fuel. A fairly linear trend can be noted in Figure 8. The higher the propane/propene content, the higher is the required pressure upstream the high pressure pump, due to the fact that the compressibility increase and density drop of the contained [C.sub.3] hydrocarbons are in a range of ~100[degrees]C, in contrast to these of the butanes/butenes. Thus, for the given boundary conditions of the performed hot idle tests, a maximum content of 70 % (m/m) propane/ propene can be recommended for an LPG direct injection concept using a state-of-the-art high pressure fuel pump. This corresponds to LPG 2, which is the typical winter fuel. In that case, the pressure at pump inlet has to be set to ~45 bar.

However, it has to be noted, that HPPs ability to face high inlet pressures is limited and an optimized design for LPG DI might be highly recommended.

Hot Start

A great relevance for LPG DI has to be assigned to the hot start behavior. In order to demonstrate the feasibility of the found solution in terms of the required HPP inlet pressures, a hot start with LPG 2 is presented exemplarily in this section. The focus of interest is the rail pressure build-up as well as the combustion behavior. The engine is warmed-up before the hot start investigations. At the end of this heating procedure the engine is shut down and stopped for ~30 s by deactivating the injection and the ignition. The pressure at the inlet of the HPP was set to 45 bar, in order to have a 5 % safety margin, regarding the determined inlet pressure of 43 bar according to Figure 8. The corresponding rail pressure for this inlet pressure was found to be 110 bar during the pressure variation at the inlet of the HPP with LPG 2. Since the rail pressure control was not deactivated before shut down, the rail pressure during the following hot soak equals the set rail pressure during the warm-up phase, except a loss of ~5 bar due to a little fuel temperature drop from ~94[degrees]C to 85[degrees]C at the inlet of the fuel rail after shut down. The results of the successful hot start are depicted in Figure 9 in terms of engine speed, fuel temperatures and pressures, as well as the injection duration and the relative air-fuel ratio. Right before the starter was activated at ~7 seconds, the HPP inlet pressure was increased at ~1.5 seconds to reach 45 bar.

The engine cranks at approximately 300 1/min and fires after one working cycle. Afterwards, the engine speed increases up to ~1800 1/ min, followed by a short period of 880 1/min till it reaches its set point of 800 1/min. This behavior is caused by the non-optimized throttle calibration. During the actual hot start, the fuel temperature at the rail inlet was between 84[degrees]C and 91[degrees]C, while the temperature at the inlet of the HPP was conditioned to ~50[degrees]C. As an average of oil and HPP surface temperature, the fuel temperature inside the HPP was estimated to be 106.5[degrees]C during the entire test. When the engine has been started, the rail pressure drops slightly to 90 bar, due to an steep increase in the injection duration up to 5 ms. After 7.9 s the HPP is able to reach the rail pressure set value the first time. Right before, the air-fuel ratio, measured by the [lambda]-sensor, already reaches its stoichiometric set value, even with the base calibration of the injection factors for start and after-start. Therefore, the injection duration is reduced to 1.2 ms at this time and remains fairly constant until the end of the measurement. The evaluation of the indication data, demonstrate that each fuel injection is followed by an ignition during the entire hot start. The cylinder pressure curve, the injection and the ignition signal of current clamps are shown in the appendix. Overall, the hot start can be assessed as successful.

For all other fuels, the hot start investigations were also performed with the determined HPP inlet pressures of Figure 8. The results are summarized in Figure 10 in terms of the required times for rail pressure build-up.

A clear trend can be seen as a function of the C3 content in the fuel. With increasing propane/propene content in the fuel, the HPP needs more time to build up the rail pressure set value, because it has to face a higher compressibility of the fuel at high engine and fuel temperatures.

Injection and Combustion Behavior during Idle

In order to evaluate the injection and combustion behavior during idle in steady state, investigations were performed on the single cylinder engine at an engine speed of 800 1/min and an indicated mean effective pressure of 0.7 bar with a spark timing of 15[degrees]CA BTDC. The results are shown in Figure 11 for all fuels and three different settings of rail pressure and fuel temperature before injector inlet. The fuel temperature was measured ~30 mm upstream the injector inlet. Rail pressures of 40 bar and 50 bar were studied in order to see the effect of the decreasing density more clearly. These low rail pressures are applicable for an alternative LPG DI concept without a HPP and only a powerful in-tank pump, which can deliver a rail pressure of ~50 bar. For a rail pressure of 40 bar, the percentages of the increased injection durations are given for each LPG fuel compared to RON 95 E5. It is evident, that all LPG fuels show increased injection durations compared to the base fuel. In general, the density of LPG fuels is lower than that of RON 95 E5. Already at 50[degrees]C and 40 bar the density of the butanes is ~23-26 % lower compared to RON 95 E5, which can be considered as incompressible fluid. The density of propane is even ~35 % lower than that of RON 95 E5 at 50[degrees]C and 40 bar. Moreover, the density of LPG fuel components shows strong temperature dependence as shown exemplarily in Figure 2 for propene. Thus, as the temperature along the injector up to the nozzle tip is increased, the density differences become even more severe. Therefore, especially fuels with high propane/propene content show heavily increased injection durations. While a pressure of 40 bar is lower than the critical pressure of LPG 1, LPG 2 and LPG 3, it can be assumed that these fuels are injected in gaseous state into the cylinder. Their injection duration is increased by more than ~200 % compared to RON 95 E5 at 40 bar rail pressure and 50[degrees]C before injector inlet. LPG 4 contains 75 % (m/m) butanes. Therefore, the increase in the injection duration amounts only to 31 %. Similar results were found for a rail pressure of 50 bar while keeping the fuel temperature at injector inlet constant at 50[degrees]C. LPG 2 and LPG 4 show rather moderately increased injection durations compared to RON 95 E5. Even though the rail pressure is higher than the critical pressure of the [C.sub.3] hydrocarbons and the gaseous state is avoided in the injector, twice the injection duration of RON 95 E5 is required for LPG 1 and LPG 3 under these conditions. Since the critical temperatures of these fuels are close to the temperature of the coolant of the engine, these fuels can reach the critical state inside the injector, on the one hand by heat transfer from the cylinder head and on the other hand from the combustion chamber. A rapid drop in density close to the critical point then results in significantly increased injection durations. For a rail pressure of 50 bar, all LPG fuels were also heated upstream of the injector inlet to ensure the supercritical state in the injector. LPG 1 was heated up to 120[degrees]C, while LPG 2, LPG 3 and LPG 4 were heated up to 175[degrees]C prior to injection in order to ensure a safety margin of more than ~30[degrees]C and to be able to compare also different temperatures for fuels with high [C.sub.3] contents. The effect of decreasing density becomes even more obvious under these conditions. As it can be seen, following an isobaric density curve, the gradients of the density of propene (Figure 5) have their maximum between ~100[degrees]C and 150[degrees]C dependent on the pressure, while the density of n-butane (Figure 7) rapidly drops between 152[degrees]C and 180[degrees]C. However, its density at 175[degrees]C and 50 bar is still ~54 % higher than that of propane at 120[degrees]C and 50 bar, due to the larger range of linearity, which prevents low densities and thus severe increase in injection durations, even for a transcritical injection.

When comparing the standard deviation of the indicated mean effective pressure in Figure 11, it can be seen that all LPG fuels show similar combustion stabilities below the limit set to 0.25 bar. The gasoline base fuel performs slightly worse at 40 bar rail pressure, which can be assumed to be caused by mixture formation drawbacks due to lower vapor pressures of the fuel components.

Regarding emissions, a clear advantage of all LPG fuels is evident for unbumed hydrocarbon emissions compared to RON 95 E5 (Figure 11). These findings can be explained by flash-boiling primary atomization and good mixture homogenization of the gaseous fuels. When heating up the fuels, all LPG fuels show a small increase of HC emissions. The fuel spray can collapse at high fuel temperatures and the penetration length is heavily increased, leading to deteriorated mixture formation [8]. Nevertheless, for CO emissions, no distinct difference was found between the fuels.

N[O.sub.x] emissions of the LPG fuels are mostly lower compared to RON 95 E5. Only LPG 1 shows a similar N[O.sub.x] emission level. While spark timing is set to 15[degrees] CA BTDC and the enthalpy of vaporization is comparable for all LPG fuels, the higher adiabatic flame temperature of propene can be considered as a reason of increased N[O.sub.x] emissions. Since idle operation is the most severe operation mode regarding changes of physical states of LPG fuel components, due to low fuel mass flows and long duration inside the injector, the fuel density is the focus of interest. Using the following injection equation [26], in which [A.sub.B] is the injection valve orifice cross-section and [DELTA][p.sub.B] is the pressure difference at the injector, the density of the fuel has been calculated based on the measured injection duration for all fuels exemplarily for a rail pressure of 40 bar and a fuel temperature before injector inlet of 50[degrees]C.

[[??].sub.B,0] = [A.sub.B] x [[alpha].sub.B] x [square root of (2 x [DELTA][p.sub.B] x [[rho].sup.B])] (4)

The fuel mass flow during injection [[??].sub.B,0] was calculated from the measured fuel mass flow using the injection duration. The flow factor [[alpha].sub.B] was calculated based on RON 95 E5 and was assumed to be constant for all fuels, due to a lack of empirical data for flow factors of LPG fuels. Thus, cavitation effects are neglected for this approach. The density [[rho].sub.2] gained hereby, is compared to the density [[rho].sub.1] derived from inlet temperature of the injector based on material properties taken from [23]. Both densities are compared in Figure 12.

Using this comparison, a rough estimation of density drop along the injector can be made. In case of RON 95 E5, both densities are equal because this fuel was used for the calculation of the flow factor. In case of the LPG fuels, the tremendous differences between the two densities are evident. For LPG 1, the density at the inlet of the injector is reduced by 91 % as the fuel is heated up to the nozzle tip. This fuel formulation and also LPG 3 have the highest content of propane/propene and therefore show the strongest drop in density. The smallest drop between the two densities was found for LPG 4. The reason is the high content of butanes of ~75 % (m/m) in this fuel.


The target of this study was to explore hot idle conditions, as well as injection and combustion behavior of LPG direct injection, in order to enable wider use of LPG fuels for modern downsized spark ignition engines and to assist the process of standardization.

The following conclusions can be drawn from the performed investigations with four different LPG fuels burned in a modern turbocharged 4-cylinder DI SI engine at hot start conditions:

* A maximum content of 70 % (m/m) C3 fuel components as an upper limit for an LPG direct injection concept is recommended, when a HPP, which is based on state-of-the-art HPP technology for gasoline engines, is used.

* For a maximum propane content of 70 % (m/m), pump functionality can be maintained with the fuel pressure of approximately 45 bar upstream the HPP at fuel temperatures of about 110[degrees]C during hot idle.

* For lower pressure levels upstream the HPP, cooling measures need to be implemented, which keep the fuel rail temperatures at lower levels permanently, in particular during the most critical hot idle conditions.

* A comparison of the time duration to build up fuel rail pressure during hot start shows a significant rise with increasing C3 content.

* Besides an increase of the inlet pressure of the HPP and cooling measures, an optimized HPP design with an increased pump compression ratio, may enable a possibly higher C3 content and a faster rail pressure build-up.

* Engine calibrations will have to take into account the severe changes in density of LPG fuels with high C3 content at high temperatures in order to effectively execute LPG DI.

Furthermore, the injection and combustion behavior was studied in stationary idle tests on a single cylinder research engine for different physical states of LPG fuels, which lead to the following findings:

* Especially for LPG fuels with high C3 content, twice the injection duration of RON 95 E5 was found, due to a change of state of the fuel inside the injector, although liquid state was conditioned at the inlet of the injector.

* No drawbacks in combustion stability were observed for high fuel temperatures.

* All LPG fuels showed a clear advantage regarding HC emissions compared to RON 95 E5. However, a small increase was observed, when heating up the fuels.

Martin Krieck, Marco Gunther, and Stefan Pischinger

VKA, RWTH Aachen University

Ulrich Kramer

Ford-Werke GmbH

Thomas Heinze

IAP, Saarland Univ. of Applied Sciences

Matthias Thewes



[1.] N. N. Future Transport Fuels "Report of the European Expert Group on Future Transport Fuels", 2011

[2.] Heil, V. "SNG and LPG from biogenic waste materials - technical feasibility and market potential" Fraunhofer Institute for Environmental, Safety and Energy Technology UMSICHT, 2011

[3.] Kankariya, N., C, V., R, M., L, S. et al., "Development of 1.2L MPI Bifuel LPG Engine for Indian market application," SAE Technical Paper 2009-24-0119, 2009, doi:10.4271/2009-24-0119.

[4.] "DIN EN 589 Automotive fuels - LPG - Requirements and test methods"; German version EN589:2008+A1:2012 Deutsches Institut fur Normung e.V. Beuth Verlag GmbH, Berlin, 2012

[5.] Krieck, M., Gunther, M., Pischinger, S., Kramer, U. et al., "Effects of LPG Fuel Formulations on Knock and Pre-Ignition Behavior of a DI SI Engine," SAEInt. J. Engines 9(1):237-251, 2016, doi: 10.4271/2015-011947.

[6.] Gunther M., Nijs M., Pischinger S., Kramer U. "Effects of LPG fuel formulations and mixture formation systems on the combustion system of a boosted SI engine" 22. Aachen Colloquium Automobile and Engine Technology, 2013

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[8.] Lacey, J., Poursadegh, F., Brear, M., Petersen, P. et al., "Optical Characterization of Propane at Representative Spark Ignition, Gasoline Direct Injection Conditions," SAE Technical Paper 2016-01-0842, 2016, doi:10.4271/2016-01-0842.

[9.] Krail M., Schade W., Fiorello D., Fermi F. et al. "Outlook for Global Transport and Energy Demand", Deliverable 3 of TRIAS, Funded by European Commission 6. RTD. Programme, 2007

[10.] Beer B., Grant T., Watson H., Olaru, H. "Life-cycle Emissions Analysis of Fuels for Light Vehicles"Report to the Australian Greenhouse Office, Australian Government, 2004

[11.] Boretti, A. and Watson, H., "Development of a Direct Injection High Efficiency Liquid Phase LPG Spark Ignition Engine," SAE Int. J. Engines 2(1):1639-1649, 2009, doi: 10.4271/2009-01-1881.

[12.] Arcoumanics C. "A Technical Study on Fuels Technology related to the Auto-Oil II Programme", Final Report, Volume II: Alternative Fuels Prepared for European Commission Directorate--General for Energy, 2000

[13.] N.N. "Propane Basics" US Department of Energy, Office of Energy Efficiency & Renewable Energy, National Renewable Energy Laboratory, Vehicle Technologies Program, 2010

[14.] "DIN EN 228:2014-10 Automotive fuels--Unleaded petrol Requirements and test methods"; German version EN228:2012 Deutsches Institut fur Normung e.V. Beuth Verlag GmbH, Berlin, 2014

[15.] N.N "Regulation No 83 of the Economic Commission for Europe of the United Nations (UN/ECE)--Uniform provisions concerning the approval of vehicles with regard to the emission of pollutants according to engine fuel requirements" Official Journal of the European Union, 2012

[16.] "DIN 51624:2008-02 Automotive fuels--Compressed natural gas Requirements and test methods" Deutsches Institut fur Normung e.V. Beuth Verlag GmbH, Berlin, 2008

[17.] Cartellieri W. "Erweiterung der Energieerzeugung durch Kraftgase. Teil 1--Literaturrecherche", Forschungsvorhaben 2-235, FVV Forschungsbericht 2-235/1, 81 FVV e.V., Frankfurt am Main, 1968

[18.] Taucar G., Cartellieri W. "Erweiterung der Energieerzeugung durch Kraftgase. Teil 2--Untersuchungen am CFR-Motor", Forschungsvorhaben 2-235, FVV Forschungsbericht 2-235/2, 82 FVV e.V., Frankfurt am Main, 1968

[19.] Cartellieri W., Pfeifer U. "Erweiterung der Energieerzeugung durch Kraftgase. Teil 3--Untersuchungen zur Ubertragbarkeit der am CFRMotor gefundenen Ergebnisse auf andere Motoren, Forschungsvorhaben 90 und 102, FVV Forschungsbericht, 120 FVV e.V., Frankfurt am Main, 1971

[20.] Weber C., Brumley A., Felipe D., Whiston P et al. "1.6 SCTI: The New EcoBoost DI-Turbo Engine with Central Direct Injection for Ford's Volume Carlines" 19. Aachen Colloquium Automobile and Engine Technology, 2010

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[22.] Adomeit, P, Jakob, M., Pischinger, S., Brunn, A. et al., "Effect of Intake Port Design on the Flow Field Stability of a Gasoline DI Engine," SAE Technical Paper 2011-01-1284, 2011, doi:10.4271/2011-01-1284.

[23.] NIST, "Thermophysical Properties of Fluid Systems", Chemistry WebBook, 2015,

[24.] Center for Applied Thermodynamic Studies, University of Idaho, Program Allprops

[25.] Richter, D. "Mechanik der Gase", Springer 2010

[26.] Pischinger S. "Internal Combustion Engines Volume I" - Lecture Notes, Institute for Combustion Engines, VKA RWTH Aachen University, 7th edition, 2015


Contact person for questions:

Martin Krieck

Tel.: +49-241-80-48128


This work was performed as part of a research project assigned by Forschungsvereinigung Verbrennungs-kraftmaschinen e.V. (FVV, Frankfurt) and conducted at the Institute for Combustion Engines, VKA of RWTH Aachen University.


ATDC--After top dead center

BTDC--Before top dead center

CA--Crank angle

CO--Carbon monoxide

C[O.sub.2]--Carbon dioxide

DI--Direct injection

EVC--Exhaust valve closing

ECU--Electronic control unit


HPP--High pressure pump

IMEP--Net indicated mean effective pressure (calculated 720[degrees] CA)

IVO--Intake valve opening

LPG--Liquefied Petroleum Gas

MN--Methane number

MON--Motor octane number

N[O.sub.x]--Oxides of nitrogen

PFI--Port fuel injection

RCP--Rapid control prototyping

RON--Research octane number

SI--Spark ignition

ST--Spark timing


[alpha]--Flow factor

[epsilon]--Compression ratio


[lambda]--Rel. air-fuel ratio


A--Injection valve orifice cross-section

[??]--Mass flow

n--Engine speed


Q--Volume flow










Table 1. Fuel properties

                          RON 95 E5   LPG 1   LPG 2   LPG 3   LPG 4

Motor octane number         86.3      89.4    94.6    93.3    91.5
Methane number               --        26      28      31      15
Density (15[degrees]C)/      749       515     526     520     560
Critical                     --       93.6    113.7   102.8   133.6
Critical press./bar          --       44.7    42.6    43.8    41.7
Ethane/% (m/m)               --        0.4     0.3     0.8     0.1
Propane/% (m/m)              --       48.3    67.4    82.1    16.1
Propene/% (m/m)              --       50.2     0.2     0.2     8.2
n-Butane/% (m/m)             --        0.4    21.2     2.3    50.7
Isobutane/% (m/m)            --        0.7    10.8     1.6    24.6
C4 olefins/% (m/m)           --        --     0.03    13.0     0.2
n-Butene/% (m/m)             --        --      --      7.9     --
Isobutene/% (m/m)            --        --     0.03     0.3     --
(Z)-2-Butene/% (m/m)         --        --      --      1.3     --
(E)-2-Butene/% (m/m)         --        --      --      3.4     --
C5 paraffins/% (m/m)         --        --      0.1     0.1     0.1
Spec, enthalpy of           29.5      28.2    25.2    26.1    23.9
Stoichiometric air          14.17     15.19   15.56   15.52   15.4
Lower heating               41.99     46.08   46.17   46.20   45.82

Table 2. Series production engine hardware specification

Bore/mm                                   79
Stroke/mm                                81.4
Number of cylinders                        4
Cylinder displacement/[cm.sup.3]          399
Valves per cylinder                        4
Compression ratio/1                       10
Injection system                    Bosch Multihole
Maximum fuel pressure/bar                 150

Table 3. Research engine hardware specification

Bore/mm                           75
Stroke/mm                        82.5
Displacement/[cm.sup.3]          364
Number of valves                  4
Compression ratio/1             10.93
Injection system            Bosch HDEV 5.2
Maximum fuel pressure/bar        200

Table 4. Single cylinder engine calibration settings

Engine speed/(1/min)                        800
IMEP/bar                                    0.7
Start of injection/[degrees]CA BTDC         280
Fuel rail pressure/bar                     40/50
IVO (1mm)/[degrees]CA ATDC                  13
EVC (1mm)/0 CA BTDC                         32
Relative air/fuel-ratio/1                    1
Air temp, in intake manifold/[degrees]C     25
Coolant and oil temperature/[degrees]C      90
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Author:Krieck, Martin; Gunther, Marco; Pischinger, Stefan; Kramer, Ulrich; Heinze, Thomas; Thewes, Matthias
Publication:SAE International Journal of Fuels and Lubricants
Article Type:Report
Date:Nov 1, 2016
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