Combustion Testing and Analysis of an Extreme States Approach to Low-Irreversibility Engines Final Report Investigators Chris F. Edwards, Professor, Mechanical Engineering; Matthew N. Svrcek, Greg Roberts, Sankaran Ramakrishnan, Graduate Researchers, Stanford University Abstract This work explores the use of high-energy-density (extreme) states to improve the efficiency of internal combustion engines for both stationary power and transportation. The basic premise that a fraction of the work lost due to combustion may be recovered by reaction at extreme states was developed under a previous GCEP program. In this program that principle has been used to demonstrate how high efficiency can be achieved in piston engines, and its implications for optimal design of stationary power engines. Introduction In the prior years of the program, a free-piston, extreme-compression engine capable of combustion at compression ratios up to 100:1 was developed. This was used to successfully demonstrate that indicated efficiencies in excess of 60% (LHV) could be achieved. Follow-on work involved development of the methods needed to make direct measurements of combustion efficiency, emissions, and soot under these conditions. In the final year of the project, a new combustion approach based on use of an autoigniting, homogeneous mixture was developed. The details of that development are given below. Also in prior years, a methodology for developing the optimal design of an energy system was investigated. It was this work that first lead to definition of the extreme states principle by The (2007), and realization of the importance of separating reactive engines by classification restrained or unrestrained by Miller (2009). Follow on work extended the approach originally developed by Teh to steady flow engines (from batch), and showed that the key element in the analysis was manipulation of the equilibrium attractor state of the system. In the final year of the project, that methodology has been completed and a robust method now exists for determining the optimal architecture of steady flow engines with work, heat, and matter regeneration. Example calculations show that engines with exergy efficiencies approaching 70% are possible using optimal, regenerative cycles. Experimental Effort The previous Diesel-style emission results showed that a primary obstacle to using a Diesel-style combustion approach at extreme compression ratio is NO x emissions.
Combustion efficiency and the other pollutant species were all within current emissions regulations, even at 100:1 CR and reasonable load. Handling the NO x is complicated, however, by the overall lean equivalence ratio, and hence excess oxygen in the exhaust stream. Aftertreatment systems for NO x exist for this scenario, primarily selective catalytic reduction (SCR). However, these systems are currently expensive in some cases comparable to the cost of the engine itself and require maintenance such as refill of a urea tank. Currently, the least expensive and most well-developed method of reducing NO x is the three-way catalyst commonly employed with automotive sparkignited engines. A defining characteristic of the three-way catalyst is that the engine must be operated in the vicinity of stoichiometric equivalence ratio. This provides incentive for achieving stoichiometric operation of the extreme compression engine. Rather than approaching stoichiometric equivalence ratio from the Diesel-style branch, the work described here investigates the feasibility of premixed combustion at extreme compression ratios. Premixed combustion has the advantage of intimate contact between fuel and oxidizer, resulting in low soot emissions at stoichiometric equivalence ratio. The difficulty in this case lies in achieving proper combustion phasing, with ignition occurring near the minimum volume in the cycle. Given that the compression ratios of interest here are, in all cases, much higher than conventional premixed-charge engines, the problem becomes one of delaying autoignition until the minimum volume is reached. Delaying autoignition is accomplished by keeping the reaction rate low until the desired time, by controlling the temperature of the gas. In this section three methods for controlling gas temperature are explored from a theoretical basis. The efficacy of each technique is discussed, as well as the implications of these methods for the overall engine efficiency. Results from an experimental investigation of two of these methods of temperature-controlled autoignition are reported and discussed. These experiments consist of compression, combustion, and expansion of a premixed, methane-air charge in the extreme compression device, using the temperature-control techniques to achieve ignition phasing at the minimum volume. Methane was chosen as a fuel due to its inherent autoignition resistance. To explore the theoretical basis for temperature-control of autoignition timing, modeling of the compression, combustion, and expansion process was performed. This model assumes a chemically and thermally homogeneous ideal gas mixture, but includes the effects of variable specific heats as well as chemical reaction kinetics from the GRI 3.0 mechanism for methane. Compression and expansion profiles as a function of time are taken from experiments in the extreme compression device. Note that the GRI 3.0 mechanism is not expected to provide a high degree of accuracy at the pressures obtained here, but it serves effectively to illustrate the basic principles of this combustion process. A simple way to adjust the gas temperature, and hence the onset of autoignition, is to start the compression at a lower temperature. The model results of this are shown in Fig. 1. The compression metric of interest in this case is the ratio of gas density during
compression to the density at ambient conditions. This corresponds to the effective compression ratio, whereas the volume ratio corresponds to the geometric compression ratio. In the earlier Diesel-style combustion studies, the effective and geometric compression ratios were identical. In this case, because the charge starts at a lower temperature and hence higher density, they are not the same and the effective compression ratio should be used as the compression metric. An interesting result from this model is that a drop in starting temperature of only 50 K is sufficient to achieve ignition at 100:1 effective compression ratio. This is partly due to the nature of isentropic compression the 50 K decrease in starting temperature results in a decrease of nearly 200 K by the time the density ratio reaches 50:1. This, combined with the fact that temperature has such a direct effect on the ignition process, gives the starting temperature significant leverage on the autoignition process. Figure 1. Temperature as a function of gas density, for compression of stoichiometric methane-air starting at atmospheric pressure and two different temperatures. Heat-Exchange Cooling One way to start compression at a lower temperature is to perform direct heat exchange with the reactant charge. To help understand this method, the left plot of Fig. 2 shows the same compression models from Fig. 1, but magnified near atmospheric conditions. One process for direct heat exchange is to cool the reactant stream prior to induction into the engine. This is shown by moving from State 1 to State 2 at constant pressure, assuming a homogeneous, stoichiometric mixture of methane and air. After reaching State 2, the process follows the normal compression in the engine. Practically speaking this could be
accomplished by a refrigeration cycle, an example of which is shown in the right part of Fig. 2. In this example a refrigerant fluid is compressed, cooled to environmental temperature, and expanded across a Joule-Thompson valve, reaching the temperature of its boiling point. Heat exchange is then performed between the refrigerant and the incoming reactant charge. Given that the charge only needs to be cooled to ~ 250 K, this can be accomplished with existing technology and refrigerants. For example, ammonia, a common refrigerant, has an atmospheric pressure boiling point of 239 K. Figure 2. Left - Temperature as a function of density, with refrigeration cooling process indicated. Right - Example refrigeration schematic. Corresponding thermodynamic states are labeled on the plot and diagram. Figure 3. Left - Temperature as a function of density, with intercooling process indicated. Right - Example intercooling schematic. Corresponding thermodynamic states are labeled on the plot and diagram. A second heat exchange cooling method is available. It is important to understand that as far as the autoignition process is concerned, it only matters that the gas compression follows the correct isentrope. Once the final compression isentrope is reached, the
remaining cycle is identical, regardless of where the cycle is initiated on the isentrope. The cooling process shown in Fig. 3 takes advantage of this to avoid sub-atmospheric temperatures. From the starting State 1, the charge is compressed on the roomtemperature isentrope to State 2, then cooled via constant-pressure heat-exchange with the environment to State 3. State 3 lies on the same isentrope as the post-refrigeration state discussed above. This process is simply compression with intercooling, as shown on the right of the figure. In the context of engines, this is commonly done when using a turbocharger. In that instance, exhaust enthalpy drives a turbine which shares a common shaft with the compressor. The output of the compressor is then cooled to near atmospheric temperature in the intercooler via heat exchange with the environment. The points shown in Fig. 3 assume a post-intercooler temperature of 310 K, resulting in a post-compressor pressure of 2.2 bar. This is well within the range of existing, off-theshelf turbochargers. The efficiency implications of using these methods of autoignition control are discussed below, after the evaporative cooling method is introduced. Evaporative Cooling A third method for lowering the gas temperature and delaying autoignition is to evaporate a liquid in the gas. As the liquid evaporates, sensible energy is extracted from the gas phase to supply the latent heat of vaporization, thus lowering the gas temperature. Water is a good choice because it has a high latent heat of vaporization and a boiling point in the desired temperature range. It is also a product of combustion, and thus does not complicate the existing combustion and emissions measurement systems. As was seen for the direct heat exchange methods, the nature of isentropic compression makes it desirable to lower the gas temperature as early as possible in the compression process in order to exert maximum leverage on the final temperature. Ideally, evaporating water in the gas immediately prior to the start of compression would lower the gas temperature as early as possible, while preventing time for significant heat transfer from the walls. However, the evaporation rate can be limited by saturation of the gas. Vaporization may have to occur some time during the compression stroke after the gas temperature has begun to rise. In that case, a straightforward method for introducing water into the gas during compression is to use existing fuel injection technology. The injection timing should be set as early as possible, while avoiding gas saturation. Chemical kinetic modeling of a homogeneous methane-air charge was again performed, following the same experimental volume-time profile used previously. In this case, vaporization of water was included in the model. Vaporization is assumed to occur homogeneously at a constant rate. The modeled vaporization rate is equal to the experimental rate of water injection in the water vaporization experiments reported later. The vaporization process is modeled as adiabatic mixing of the in-cylinder gas and liquid water. Water in the gas phase mixture is modeled as an ideal gas. The start of injection was chosen to ensure that the gas remains unsaturated. In this example, water evaporation begins when the gas temperature first reaches 400 K.
Figure 4. Pressure as a function of volume for homogeneous, stoichiometric methane-air, with and without water vaporization. The model results are shown in Fig. 4, compared with the earlier model results for methane-air with no water evaporation. Two features are apparent: first, the water injection successfully delays autoignition until 100:1 compression ratio is achieved. Second, the evaporative cooling can be seen in the plot as the compression line drops below the normal isentrope. Once the vaporization is completed, the compression again follows an isentrope. An important result of this exercise is that the total mass fraction of water injected is only 3% of the total charge. This is a relatively small amount less than one-fifth the water produced from combustion and easily achieved with existing injection technology and reasonable injection times. This model result indicates that water injection is promising as a control for autoignition. Efficiency Implications A critical factor in evaluating the above strategies for controlling autoignition is their effect on the cycle efficiency. A dramatic decrease in efficiency as a result of these methods would defeat the purpose of achieving high compression ratios. As discussed in our previous work, exergy destruction during reaction is minimized by performing the chemical reaction at the highest possible internal energy state. The cooling methods for autoignition control run counter to this, in that they function by reducing the gas temperature, and hence internal energy. However, the increased exergy destruction during reaction is offset by decreased exergy remaining in the exhaust. For simple-cycle engines, this exhaust exergy is also destroyed. The reduction in exhaust
exergy is of similar magnitude to the increased destruction during reaction, and thus the net exergy destruction is not significantly affected by these cooling strategies. Efficiency losses due to driving the cooling process itself must also be considered. The refrigeration and intercooling approaches require work to drive a compressor. The water vaporization method uses sensible energy from the gas to drive evaporation. Because the water does not recondense during the expansion, this sensible energy is not returned to the gas and the net work output is reduced. For the refrigeration and intercooling methods, exhaust enthalpy can be used to provide the required compressor work. Turbocharged, intercooled systems, in which a turbine in the exhaust flow provides the compressor work, are commonplace in conventional engines. Exhaust enthalpy could also be used in this way to drive the refrigeration method, although the practical implementation would be more complex. To provide more concrete examples, the efficiency was calculated for two of the model methane-air cases shown earlier the intercooling approach (Fig. 3), and the water injection approach (Fig. 4). The computed efficiencies are shown in Table 1. They are compared to a model cycle with no cooling strategy, in which ignition is artificially delayed until the minimum volume by setting the reaction rate to zero in the model. The efficiencies are calculated using the same chemical kinetics, compression-expansion model described earlier, including the model of evaporation for the last two cases. For the second intercooling case, work is supplied from an external source to a compressor with 0.8 polytropic efficiency. Table 1. Calculated first-law efficiencies with different cooling strategies. In all cases, the effective compression ratio is 100:1, the charge is homogeneous, stoichiometric methane-air, and ignition occurs at the minimum volume. Without use of exhaust enthalpy to drive the cooling process, and with a non-ideal compressor, the intercooling strategy decreases the overall efficiency by ~3%. If exhaust enthalpy is used to drive the compressor, the overall efficiency slightly increases. This is possible because the exergy reduction in the exhaust is greater than the increase in exergy destruction during reaction. The water injection strategy with the earlier injection timing also slightly improves the efficiency, for the same reason. As the water injection is moved later the efficiency is reduced, for two reasons. First, cooling the gas later in the
compression requires more total water injection to achieve the same drop in final gas temperature. Second, the gas pressure decreases below the isentropic value due to the evaporative cooling. If this happens late in the compression stroke, then additional compression work is required. The above results show that these cooling strategies do not significantly affect the efficiency, at least in the theoretical model cases. This, combined with the indications of efficacy of control discussed in the previous sections, provides motivation for experimental exploration of these techniques. Experimental Setup for Premixed Autoignition Two significant additions to the extreme compression apparatus were made to enable experimentation with a premixed charge. First, a system was constructed for filling the cylinder with a homogeneous mixture of fuel and air of known stoichiometry. Second, modifications were necessary to handle the greatly increased thermal loads from stoichiometric premixed combustion most importantly a thermal shield for the piezoelectric pressure transducer. These new systems are described below. Premixed Gas Delivery System For the experiments described below, a homogeneous mixture of gaseous methane and air was used. In order to ensure a completely homogeneous charge, the methane and air were mixed prior to entering the cylinder. Figure 5 illustrates the system used to accomplish this. Figure 5. Schematic of the premixed methane-air delivery system. Methane is supplied from a gas cylinder, regulated to 150 psi. Compressed air is supplied at 120 psi through a filter and dryer. Each gas then passes through a pressure regulator and a precision orifice. The orifice diameters are fixed at 75 micron for the methane and 400 micron for the air. Flow control is achieved by adjusting the pressure upstream of each orifice, via the pressure regulators and high-precision pressure gauges. Downstream of the orifices, the methane and air are mixed. For all flow conditions, the
pressure difference across each orifice is sufficient to maintain choked flow. In this way the flow rates are unaffected by fluctuations in the downstream pressure during the cylinder fill process. The system was calibrated using two independent methods. First, a wet-test meter was used to calibrate the flow rate of each gas. Second, an evacuated tank, weighed on a scale before and after filling, was used to calibrate each gas on a mass basis. These calibrations were performed for a range of upstream pressures, and agreed with each other to within ± 1%. Repeatability between tests was also within ± 1% for each method. The resulting calibration was used to set the stoichiometry of the mixture in the combustion experiments. To fill the cylinder, the piston is moved to a position near the combustor end-wall as shown in Fig. 5. The gas flow is activated with both cylinder valves open. The mixture thus flows across the combustion chamber, scavenging any gas previously existing in the chamber. The two pressure regulators are adjusted with the flow in steady-state to ensure accurate upstream pressures. The vent valve is then closed, and the methane-air mixture pushes the piston to the top. The cylinder-fill valve is closed, and the vent valve is then opened as necessary to achieve the desired final cylinder pressure. Pressure Transducer Thermal Shield A piezoelectric pressure transducer is used for the main combustor pressure measurement. This type of transducer can be affected by thermal shock on the transducer face. Radiative and convective heat transfer causes thermal expansion of the diaphragm and housing. This can relax the pre-load on the piezoelectric element and cause a negative shift in the output. With the bare, unmodified sensor, the time for heat transfer to affect the signal is such that the negative bias occurs in the latter portion of the expansion stroke. This effect is noticeable in the experimental pressure-volume traces. Without a sufficient thermal barrier, the pressure-volume trace decreases significantly below the expected isentrope towards the end of expansion. An artificial decrease in the indicated work and efficiency occurs as a result. For the earlier Diesel-style combustion experiments, a simple vinyl protective layer on the transducer face, as recommended by the manufacturer, was sufficient. This delayed the thermal shock from reaching the diaphragm until after the experiment was completed. Due to the stratified nature of that combustion strategy, the high-temperature combustion regions were away from the transducer. In contrast, with premixed combustion the reaction and high temperatures occur throughout the entire chamber, including near the wall. This situation is further exacerbated by the extremely high pressure obtained here, which compresses the thermal boundary layer. The original vinyl protection was experimentally found to be insufficient for the premixed experiments. The vinyl itself combusted, and was largely consumed, allowing the thermal shock to reach the diaphragm. Silicone RTV survived slightly better, but was still unable to sufficiently delay the thermal shock. In engine applications a metal barrier with a number of small holes is sometimes used, although this has an impact on the
frequency response of the transducer. In the current application the very high pressures correspond to a very small quench distance and hence require very small hole diameters in the thermal barrier. This results in an unacceptable curtailing of the high frequency response of the sensor. Figure 6. Schematic of the pressure transducer thermal shield. An outer Teflon (PTFE) layer was found to successfully survive the premixed combustion, due to the very high resistance of this material to oxidation. The required thickness of Teflon to sufficiently delay the thermal shock was 0.5 mm. Teflon of this thickness is overly stiff, and can affect the frequency response of the sensor. A vinyl inner layer immediately on the diaphragm face, with a thin outer layer of Teflon, was found to have the required combination of flexibility, resistance to thermal shock, and resistance to oxidation. A final difficulty arose with attempts to use adhesives to attach these thermal protection layers the adhesive itself was oxidized and failed. Instead, a thin, stainless steel clamp ring was used to fix the thermal barrier to the transducer face. The resulting thermal protection setup is shown in Fig. 6. As will be seen in the experimental results presented below, the log(p)-log(v) plots follow the expected isentrope at the end of the expansion stroke, indicating that the thermal shock is successfully delayed until after expansion. Also, comparison of pressure versus time for experiments repeated with and without the thermal barrier demonstrated that the frequency response of the transducer is not impacted. Experimental Investigation of Intercooling-Controlled Premixed Autoignition A method was developed for investigating premixed combustion with the intercooling strategy in the extreme-compression apparatus. In the theoretical example for intercooling, the charge was compressed isentropically, then cooled at constant pressure, and then inducted into the engine where it is further compressed isentropically. The actual experimental process differs slightly from this. A comparison between the two is
made in Fig. 7. The theoretical intercooling process followed a path through States 1, 2, and 3, as shown earlier. The experimental process instead follows a path through States 1, 2a, and 3a, ending on the same isentrope as the theoretical process. Figure 7. Comparison of experimental and theoretical intercooling processes. Both are for stoichiometric methane-air. The cylinder is first filled with premixed methane-air as described in the previous section. At the end of the fill process, the piston is located at the top of the cylinder and the gas has come to thermal equilibrium with the walls (355 K). The gas pressure is chosen such that the charge density matches the density at atmospheric pressure and the starting temperature of the target isentrope 250 K in this example. This matching of charge density is necessary to achieve the correct effective compression ratio. At the end of this fill process, the gas state is at State 2a in Fig. 7. With the cylinder fill valve closed, the piston is then moved part way down the cylinder by driver air, thus raising the pressure in the cylinder. During this process, heat transfer occurs such that the gas remains at the wall temperature. The final pressure, and hence piston position, is chosen such that the gas ends at State 3a on the target isentrope. The normal extreme compression experiment is triggered at this point. There are two key differences between the experimental process described above and the theoretical intercooler process. First, the post-intercooler temperature is 45 K higher in the experimental process because the gas temperature is constrained to reach the average cylinder wall temperature. Second, the experimental pre-intercooler compression process is not isentropic. These differences only matter in interpreting the overall cycle efficiency, as discussed in the next section. Experiments conducted using this method are as follows: A set of intercooling experiments was performed with effective compression ratio near 35:1, while varying
equivalence ratio from 0.96 to 1.04. A second set of experiments was performed near 60:1 effective compression ratio, again varying equivalence ratio from 0.96 to 1.04. Experiments were also conducted with 1.0 equivalence ratio at 78:1 and 89:1 effective compression ratio. For the majority of these experiments, the intercooling pressure was chosen to result in ignition phasing at or slightly after the minimum volume. For all of the intercooling experiments, the following data were collected or calculated: pressure, volume, indicated work, efficiency, gas cooling requirement (via starting temperature and pressure), ignition timing, and emissions of NO x, CO, HC, and soot. The cycle performance metrics such as efficiency and cooling requirement are discussed in the next section. Emissions results are discussed in the following section. Efficiency and Autoignition Performance Results Figure 8 shows an example intercooling autoignition result for 78:1 effective compression ratio (62:1 geometric). In the left plot, compression begins at a smaller volume than the full cylinder volume and higher pressure than atmospheric. Rapid reaction at the minimum volume appears as a vertical line on the plot, followed by expansion. The expansion stops prior to reaching the full cylinder volume, in order to avoid the piston colliding with the driver section. The black dashed line shows the expansion isentrope extrapolated to the full cylinder volume. Note that V 0 in the x-axis scale refers to the cylinder volume, not the volume at atmospheric conditions. Thus the x-axis scale in this plot is inversely related to the geometric compression ratio, not the effective compression ratio. Figure 8. Pressure as a function of volume (left plot) and time (right plot), for methaneair with equivalence ratio = 1.0, and 78:1 effective compression ratio (62:1 geometric). Ignition timing is defined as the point where the pressure rises to half way between the compression and expansion isentropes, as indicated by the circle in Fig. 8. This definition is arbitrary, but is consistent across all experiments. A physical interpretation of this definition is similar to the CA50 metric often used in conventional engines [1] it essentially corresponds to 50% completion of the exothermic stage of reaction.
The ignition timing was repeatable from one experiment to the next to within ±0.03 ms, measured relative to the time at which the minimum volume was reached. The actual timing achieved on a given experiment could be determined to within ±0.005 ms. For the 78:1 effective compression ratio experiment shown in Fig. 8, ignition occurred 0.01 ms after the minimum volume was reached. For the experiments near 35:1 effective compression ratio, the timing varied from 0.11 to 0.14 ms after the minimum volume. For most experiments near 60:1 effective compression ratio, the timing varied from 0.02 to 0.06 ms after the minimum volume. To put these values in context with the overall cycle, the time for the piston to travel from the midway point of the compression stroke to the minimum volume, and then return to the midway point on the expansion stroke, was ~5.3 ms for 35:1 effective compression ratio. For 60:1 effective compression ratio, this time was ~2.8 ms. Another feature apparent in Fig. 8 is a large degree of pressure ringing. The ringing is better visualized in the plot of pressure as a function of time on the right. The peak ringing amplitude is ±200 bar. The maximum rate of pressure rise is also extremely high 56.6 bar/μs for this case, which would translate to ~5000 bar/cad for an engine at 1800 RPM. For comparison, maximum rate-of-rise for HCCI engines is targeted to be less than 10 bar/cad. The rate-of-rise is higher in this case partly because no dilution was used, and partly because of the high temperature and pressure. The large pressure ringing is linked to the high rate-of-rise, as the rapid pressure rise does not occur uniformly throughout the cylinder. Pressure waves emanate from the regions of earlier ignition, in a manner very similar to knock in SI engines. Pressure ringing of this amplitude would present practical difficulties in developing an engine, for example with respect to noise, vibration, and harshness (NVH), structural integrity, and heat transfer. One potential solution is to introduce a significant amount of dilution to slow the reaction rate. Another possible solution is to inject water, as discussed in a later section. Experimental results for gross indicated efficiency are shown in Fig. 9. The results are compared to the thermodynamic limit for a stoichiometric, methane-air Otto cycle starting from ambient and with the same effective compression ratio. The indicated net work from the extreme compression experiment is determined as before, by integrating the experimental pressure over volume. The extrapolation of the expansion stroke to the full cylinder volume (shown by the dashed line in Fig. 8), is included in the integration. However, with the intercooling strategy the experiment begins at a point above atmospheric (point 3a in Fig. 7). An assumption must be made about the work required to get to this starting point in other words, the work required for the compressor upstream of the intercooler. Two plausible assumptions are shown in Fig. 9. The higher efficiency is calculated assuming that all of the compressor work comes from a turbine in the exhaust. This case is analogous to a turbocharger with intercooling. The lower efficiency is calculated assuming that the compressor work must be supplied from another source, with a polytropic compressor efficiency of 0.8.
Figure 9. Efficiency as a function of effective compression ratio. The Otto cycle limit is calculated for premixed, stoichiometric, methane-air starting at ambient conditions, and reaching the same effective compression ratio. Variable properties and equilibration of products during expansion are included. The experimental gross indicated efficiency is shown with two different assumptions about compressor work prior to intercooling. The indicated efficiency reaches a maximum of 57%. For comparison, the lean, Dieselstyle combustion from our previous work reached a maximum indicated efficiency of 60%. The theoretical limit is lower for stoichiometric, premixed methane-air 70% at 90:1 compression ratio, as compared to 77% for the lean-diesel cases at the same compression ratio. Thus the indicated efficiency for these stoichiometric-charge experiments is higher relative to the theoretical limit than for the lean-diesel case. This is noteworthy, because the mass and heat transfer losses are likely worse for the stoichiometric, premixed combustion, due to the higher pressure and higher bulk temperatures achieved. Heat transfer is particularly exacerbated in the premixed case, due to reaction occurring near the walls. A potential explanation is that the efficiency benefit from the rapid, premixed reaction offsets the increased heat and mass transfer losses. Especially in the free-piston device, Diesel-style combustion occurs significantly before and after the minimum volume. The premixed strategy avoids this loss mechanism. As pertains to autoignition control specifically, the performance of the refrigeration and intercooling methods is determined by the amount of gas cooling required to achieve autoignition at the desired time. This performance is set by the post-refrigeration temperature required in the refrigeration method, or by the intercooler pressure required in the intercooling method. For example, the earlier kinetics modeling predicted that a refrigeration temperature of 250 K, or an intercooler pressure of 2.2 bar, would be
required to hold off ignition of methane until the minimum volume for 100:1 effective compression ratio. This second metric corresponds, for example, to the outlet pressure of the compressor in a turbocharged, intercooled engine setup. The actual, experimental performance of the intercooling method can be calculated in this same way. The gas pressure and temperature are known at the start of each experiment. Therefore, the isentrope that the experimental compression process followed is also known. From this isentrope, the two cooling requirement metrics can be calculated. The experimental cooling requirements are shown in Fig. 10. As it was for the theoretical discussion, the post-intercooler temperature is assumed to be 310 K. Only experiments for which the ignition timing was similar slightly after the minimum volume are included in the calculation and figures. Figure 10. Experimental cooling requirement to achieve ignition at the minimum volume. Cooling is expressed in terms of the atmospheric starting temperature of a compression isentrope in the left plot, and in terms of the absolute intercooler pressure for a post-intercooler temperature of 310 K in the right plot. Model results are shown for comparison. Model is the same as discussed in earlier section, with volume-time profiles from this set of experiments. The gas cooling requirement increases with effective compression ratio, as indicated by a decreasing refrigeration temperature or an increasing intercooler pressure. For 78:1 effective compression ratio, achieving ignition at the minimum volume required the equivalent of a 240 K refrigeration temperature, or a 2.5 bar intercooler pressure. From these results, it appears that a larger cooling requirement was required experimentally than was predicted in the theoretical discussion. To further this comparison, the same chemical kinetics model used in the theoretical discussion of autoignition control was used here. In this case, the volume-time profile for the model is taken from this data set for the corresponding compression ratio. The charge is again stoichiometric methane-air, with the GRI 3.0 chemical kinetics mechanism. The model cooling requirement is adjusted by changing the starting temperature of the charge, while holding the starting pressure and volume the same as in the experiment. This requires
slightly adjusting the volume profile to give the same effective compression ratio as in the experiment. The cooling requirement thus predicted by the chemical kinetic model is also shown in Fig. 10. The experimental cooling requirements are higher than those predicted by the model for all compression ratios. One likely explanation is that the kinetics mechanism may not accurately predict ignition behavior due to the very high pressures and the rapid compression-expansion profile. Further research is required to understand the accuracy of ignition prediction by the chemical kinetics mechanism under these conditions. Another possibility is that temperature inhomogeneities in the experiments hasten ignition, thus requiring greater cooling. However, inhomogeneities result primarily due to heat transfer to the cylinder walls, and therefore should be at a lower temperature than the isentropic temperature assumed in the model. Note that at 35:1 compression ratio, the intercooler pressure predicted by the model is sub-atmospheric (0.9 bar in Fig. 10). In effect, this means the model predicts that the reactant charge must be heated rather than cooled to achieve the desired ignition phasing at 35:1 compression ratio. Emissions Results Emissions of NO x, CO, CO 2, O 2, and hydrocarbons were measured for this set of experiments. The emissions measurements described here were performed in the same manner and using the same system described in the previous years report for the Dieselstyle combustion study. A carbon balance was performed, comparing the amount of carbon in CO 2, CO, and hydrocarbons in the exhaust to the amount of carbon in the fuel. Contribution of soot was not included in the carbon balance. The carbon balance results are shown in Fig. 11, for all of the intercooling experiments, plotted as a function of equivalence ratio. The level of variability in the balance is ±0.7%. There is, however, a persistent, average offset of -2%. The offset could be of some concern if it comes from the actual methaneair stoichiometry, resulting from the calibration of the premixed gas delivery system. The methane used in these experiments contained up to 0.6% N 2, which could contribute to some, but not all, of the observed offset. An equivalence ratio shift of 2% could be important in interpreting species emissions profiles that change greatly in the vicinity of stoichiometric, such as CO and NO x. However, as discussed above, the fill system was calibrated by two independent methods in close agreement, and no source of error was determined. As such, the results are reported assuming the charge stoichiometry is correct.
Figure 11. Exhaust carbon fraction as a function of equivalence ratio for all intercooling experiments. Measured CO, HC, and NO x emissions are shown below in terms of indicated-specific emissions and mole fraction. Conversion of emissions to a mass basis assumes NO x emissions are entirely NO, and HC emissions have the same H/C ratio as the fuel. Specific emissions are computed using the indicated net work from the experiment. Figure 12. Hydrocarbon emissions as a function of equivalence ratio, for different compression ratios, on an indicated-specific and mole fraction basis. Compression ratios listed in the legend are effective compression ratios. Hydrocarbon emissions are shown in Fig. 12. There appears to be a slight increase in hydrocarbons with increasing equivalence ratio, although the magnitude of the trend is
similar to the repeatability seen in the multiple points at 1.0 equivalence ratio. The hydrocarbons are also significantly lower for the lowest compression ratio. Both of these trends are consistent with an interpretation that the majority of the hydrocarbon emissions come from unburned hydrocarbons stored in crevice volume, as is common in premixedcharge engines. Figure 13. Oxides of nitrogen emissions as a function of equivalence ratio, for different compression ratios, on an indicated-specific and mole fraction basis. Compression ratios listed in the legend are effective compression ratios. Oxides of nitrogen emissions are shown in Fig. 13. Only experiments with similar ignition timing (slightly after the minimum volume) are shown in the these NO x emissions figures. Compression ratio has little effect on the NO x level. At first this appears counterintuitive. However, due to the autoignition control strategy, the peak temperature is maintained relatively constant. Peak temperature is a strong predictor of NO x emissions, consistent with the observed trend. Furthermore, this result indicates that pressure does not have a strong effect on the exhaust NO x level, as the peak pressure changes from 400 bar to 800 bar over this range of effective compression ratio. In general, the rate of NO formation for the thermal NO x mechanism (relevant here) is expected to increase with pressure. However, chemical kinetics modeling (discussed below) suggests that for these experiments the NO x reaches its equilibrium value shortly after ignition, in which case the amount of NO x is no longer limited by the rate of formation. The equilibrium concentration of NO x is weakly dependent on pressure, slightly decreasing with increasing pressure, which is consistent with the experimental observation.
Figure 14. Comparison of experimental NO x mole fraction as a function of equivalence ratio to model prediction. Model and experimental results are for 60:1 effective compression ratio. The model assumptions are the same as those used to predict the cooling requirement for Fig. 10. As expected, the NO x emissions decrease with increasing equivalence ratio, due to the scarcity of oxygen. Of particular note is the steepness of this decline, with NO x levels changing by a factor of 5 over an equivalence ratio change of only 8%. For comparison, predicted NO x levels were computed using the compression-expansion model with GRI 3.0 kinetics, which includes a NO x mechanism. The model setup was the same as that used to produce the model results in Fig. 10, with the volume-time profile taken from the experiments, and the starting temperature adjusted to achieve ignition phasing at the minimum volume. The model and experimental results for 60:1 effective compression ratio are shown in Fig. 14. The steep experimental NO x trend is predicted by the model, although the amount of NO x is consistently higher in the experiments than in the model. This result suggests that the observed trend is not an experimental artifact, but rather is a property of this combustion process in the extreme compression apparatus. Furthermore, in a study of a conventional spark-ignited natural gas engine, Einewall et. al. report a linear decrease in NO x emissions of ~30% when varying equivalence ratio from 1.00 to 1.02 [1]. This trend agrees well with the results from the present study. The appearance of the same NO x trend in the conventional engine suggests that it is a general feature of methane-air engine combustion, rather than a result of the free-piston profile, high compression ratio, or HCCI combustion strategy.
Figure 15. Carbon monoxide emissions as a function of equivalence ratio, for different compression ratios, on an indicated-specific and mole fraction basis. Compression ratios listed in the legend are effective compression ratios. Mole fractions of CO predicted by the chemical kinetic model are shown for comparison. Model assumptions are the same as those used in the NO x discussion above. Carbon monoxide emissions are shown in Fig. 15. The CO emissions are low for lean and stoichiometric equivalence ratio, and then increase significantly for slightly rich equivalence ratios. Increasing compression ratio appears to reduce the CO emissions, particularly for slightly rich equivalence ratios. The chemical kinetic model predictions of CO mole fraction are also shown in the plot on the right. The model is the same as was used for the NO x emissions discussed above. The CO mole fraction values predicted by the model are similar to the experimental results. However, the experimental CO increases non-linearly for rich equivalence ratios, while the model predicts a linear increase with increasing equivalence ratio. The model also predicts higher CO for lower compression ratio, although the effect is less than that seen in the experimental results.
Figure 16. Combustion efficiency as a function of equivalence ratio, for different compression ratios. Combustion efficiency was calculated from the measured emissions of CO and unburned hydrocarbons, as well as a computed level of H 2 emissions assuming an equilibrium constant for the water-gas shift reaction of 3.5 (as is common practice [1]). The results are shown in Fig. 16 as a function of equivalence ratio. At the leanest equivalence ratio, hydrocarbons are the primary contributor to fuel energy remaining in the exhaust. For this reason the combustion efficiency is higher for the lower compression ratio. At stoichiometric, CO contributes on an equal level with hydrocarbons, while for rich equivalence ratios the CO dominates. Thus the combustion efficiency follows the inverse CO emissions trend for equivalence ratios above stoichiometric. A primary goal of exploring stoichiometric combustion was to enable the use of a threeway catalyst. Recently, a significant amount of research has focused on operating conventional spark-ignited (SI) natural gas engines with stoichiometric charge and a three-way catalyst. For example, Chiu et. al. [2] and Saanum et. al. [3] both report preand post-catalyst emissions for modern, heavy-duty, stoichiometric, natural gas engines. To help put the emissions results into context, specific emissions from this study at 60:1 effective compression ratio are compared to those from the SI natural gas engines in Table 2. Because catalyst performance is dictated by species concentration rather than specific emissions, the results are also compared on a mole fraction basis. Approximate exhaust mole fractions for the Chiu and Saanum papers were calculated using their reported specific emissions, brake efficiency, brake power, and EGR fraction.
Table 2. Comparison of emissions levels from this study to those of spark-ignited, stoichiometric, heavy-duty natural gas engines from the literature. Three-way catalyst conversion efficiencies from the literature are used to calculate hypothetical post-catalyst emissions from this study. Chiu and Saanum both chose an equivalence ratio that minimized post-catalyst emissions of CO and NO x. In both studies, the optimal equivalence ratio was reported to be slightly rich of stoichiometric (although an exact value was not given). The CO/NO x ratio from this study most closely matches their results at an equivalence ratio of 1.028 likewise slightly rich. The emissions results from this study are interpolated to 1.028 equivalence ratio in Table 2 for comparison. Specific emissions of both CO and NO x were lower for this study than for the SI engines. Partly this is due to the use of indicated work for the present study as opposed to brake work. Hydrocarbon emissions from the current results were dramatically lower. This is potentially due to the unusually long stroke and large clearance volume of the extreme compression apparatus. This geometry creates a relatively small ratio of crevice volume to clearance volume in comparison to conventional engines. On a mole fraction basis, the CO and NO x emissions from this study are nearly identical to the SI engine exhaust, while the HC emissions remain lower. The catalyst conversion efficiencies reported from the natural gas engine studies were greater than 90% for all species, using conventional three-way catalysts. Because the exhaust gas composition from the current study is very similar to that reported for the SI engines, it is reasonable to expect similar catalyst performance. Hypothetical postcatalyst emissions from the extreme compression experiments are therefore calculated in the right column of Table 2, using the average catalyst efficiencies from the two SI engine studies. The resulting post-catalyst emissions would be significantly below the upcoming Tier 4 CARB and EPA emissions regulations. This is a promising result, as it indicates that meeting emissions regulations with this combustion strategy might be accomplished with existing catalysts.
Experimental Investigation of Vaporization-Controlled Premixed Autoignition An experimental investigation of water vaporization for autoignition control was also conducted. To accomplish this, a stock Bosch common-rail Diesel injector was used to inject water into the cylinder during compression. The injector was connected to a stainless steel accumulator volume, which in turn was supplied with 1450 bar pressurized water from an air-driven, high-pressure pump. In order to enable early injection without significant wall-wetting, the injector was mounted axially in the end-wall of the combustor. Injection was through a single orifice along the axis of the cylinder, with an orifice diameter of 360 μm. The water vaporization experiments were only performed at stoichiometric equivalence ratio, and emissions were not measured. Because emissions were not measured, heating the cylinder walls was not required. Experiments were first performed with a wall temperature of 330 K, with an effective compression ratio up to 41:1. Further experiments were performed with a wall temperature of 298 K, up to an effective compression ratio of 61:1. As will be discussed below, achieving proper ignition phasing at compression ratios significantly higher than 61:1 was not practical with the current experimental setup. For each of these experiments, the cylinder was filled with a stoichiometric, homogeneous methane-air charge at atmospheric pressure and the cylinder wall temperature, using the process described previously for the intercooling methane-air experiments. The timing of the start of water injection was varied. The optimal start-of-injection timing was found to be near an instantaneous compression ratio of 2:1 (i.e. the midway point of the compression stroke). Figure 17. Pressure as a function of volume for stoichiometric methane-air with waterinjection cooling. An intercooling experiment at 60:1 effective compression ratio is shown for comparison. The temperature and compression ratio shown in the legend are