INTRODUCTION
Temperatures under -65.5 C are considered to be in the cryogenic region. All gases, including natural gas (methane) can be liquefied at extremely low temperatures. Such liquid gasses are commonly called " cryogenic fluids ". Once liquefied, the gasses will remain in a liquid phase even at atmospheric pressure, provided that it is kept at its saturation temperature.
Basic concepts in the liquefaction techniques are isoenthalpic and isoentropic phenomena. In the case of using isenthalpic expansion (Joule-Thompson effect), the decrease in the fluid temperature is provided by the pressure drop through the valve. The cooling which is the result of expansion or throttling of a gas from a high to low pressure is called Joule Thompson cooling. A gaseous pure substance has a positive Joule-Thompson coefficient could go through a continuous temperature decrease and finally it becomes liquefied by using throttling processes. A positive Joule-Thompson coefficientindicates that the temperature drops during such a throttling process.
It can be approximated;
An isentropic process can also be used. In this thermodynamic scenario, a turbo-expander is used to effect a pressure drop. In this case work is done to the turbine and the gases expand and cool. These systems can be installed at the large 'city gate' pressure regulators where the pipeline gas is distributed to the city grid. A part of the total flow through the pressure reducing turbo-expander can be refrigerated to the point of liquifaction. This would lead to an inexpensive source of LNG that would utilize the embedded horsepower in the pipeline gas for refrigeration energy.
The saturation temperature of natural gas is -163 C. Fuel volume increases approximately 600 times when LNG becomes gas. This reduction in volume provides significant economic advantages in storing and transporting of natural gas. LNG boils at -163 C and vaporizes rapidly when exposed to higher temperatures. The gas itself is much lighter than air and therefore rises from spills. Since at this time the LNG and its vapor is not odorized, it is completely odorless and not dedectable.
The liquefaction process increases fuel quality by removing many impurities (water in particular) that are commonly associated with pipeline gas. LNG also provides an opportunity for cogeneration by cooling various mechanical components enroute to the engine, where it arrives and is utilized in its gaseous form.
LNG in USA
The greatest concentration of LNG plants are located along the East coast, The Midwest and South Central Regions of the country. Totally 26 LNG plants, 0.14 % of total natural gas deliveries (Source: American Gas Association, Issue Brief 1990-9). LNG plants condense natural gas from its normal gaseous state to liquid phase at -163 C and subsequently revaporize the gas. Domestic LNG is liquefied from domestic natural gas supplier and does not contain the percentage of higher hydrocarbons seen in some imported LNG.
At present, there are four LNG import terminals in the US. The Distrigas terminal at Everett, Mass; Panhandle Eastern Terminal at Lake Charles, LA; Columbia`s terminal at Cave Point, MD and Southern Corp`s Terminal at Elba Island, GA. Depending upon the design of the liquefaction train of various export stations, the time spent at sea in LNG ships and other considerations, the gas composition of imported LNG varies. It tends to have a higher heating value than domestic natural gas.
DIESEL CONVERSIONS TO NATURAL GAS ENGINES
Initiation of combustion in diesel engines is achieved by the self ignition property of the fuel. Cetane number of a fuel is associated with the performance of that fuel in compressed ignition engine. The cetane number of natural gas is very low.
For a given diesel engine, bigger cetane number gets shorter ignition delay period and it means smaller amount of fuel in the combustion chamber when the fuel ignites. Consequently high cetane number provides lower rates of pressure rise and lower peak pressure. Both decrease combustion noise and provides improved control of combustion, as a result increased engine efficiency and power output are achieved. Easier starting, particularly in cold weather, faster warm-up, reduced exhaust emission and odor are results of higher cetane number. A high speed diesel engine, normally is supplied with a fuel which cetane number changes between 45-50. This number trends to decrease to 35. [17] Due to the low cetane number of natural gas, its performance is very poor when directly inducted into diesel engines. Some modifications are needed to use these fuels in compression ignition engines. Two methods are available to use gaseous fuels in compression ignition engines.
Conversion to SI engines
There are some necessary modifications to convert the compression ignition engine to spark ignition engine.
It is clear that these modifications needs much effort and expense, especially for the aftermarket conversion of an engine. The conversion of Cummins B and Navister DTA-360 engines to operate on LPG in an SI configuration costs $ 5,000 plus the fuel tank and installation costs to Greenway Environmental Research. [4]. It should be noted that this type of conversion provides a significant advantage over compression ignition engines regarding exhaust emission.
Dual-fuel operation or fumigation is achieved by the burning of both a gaseous fuel and diesel at the same time in a CI engine. An ignition source must be provided to start the combustion when gaseous fuels are used in CI engine due to their low cetane number. Injection of small amount of diesel fuel can be used as an ignition source. In dual fuel operation, the gaseous fuel-air mixture is on the lean side. This mixture is compressed during the compression stroke. Near the end of the compression stroke, diesel fuel is injected in order to initiate the combustion of gaseous-air mixture. Due to its function to start the combustion, the diesel in fumigation type engine is referred to as pilot diesel. The changeover of the mode from duel-fuel to straight diesel or vice versa can take place while the engine runs.
The combustion characteristics of dual-fuel operation are different than those of straight diesel operation. Mainly, the difference stems from the presence of the premixed fuel-air mixture. The process in dual-fuel operation makes it kind of hybrid compression ignition/spark ignition engine. (The spark plugs are replaced by diesel fuel injector as an ignition source)
In diesel process, the control of the combustion is achieved by mixing. The fuel injected to combustion chamber splits into droplets which turn in vaporize and auto-ignite after mixing with high temperature compressed air. Mixture formation is changed by the introduction of gaseous fuel. Gaseous fuel/air mixture is initiated by the help of diesel fuel combustion. Flames which is generated from the pilot fuel propagate to varying degrees and rates throughout the surrounding gaseous.
However, the combustion process in dual-fuel process is controlled by flame propagation. During the intake stroke, a nearly uniform mixture of gas and air is introduced into intake manifold and this mixture is compressed to high temperature and pressure. The flame fronts which is initiated by injected pilot diesel which propagate through the gas-air mixture. The hybrid nature of the engine provide some advantages. Combustion occurs quickly in a premixed charge. This provides the engine run on a cycle closer to Otto cycle. Otto cycle is assumed to be more efficient compared with Diesel cycle at the same compression ratio. The heat release in Diesel engine is limited because of diffusion burning which causes larger time losses and a decrease in efficiency during combustion.
The most concerned problems with this type of engine is knock tendency. There are two types of knock problem occur in dual-fuel engines. The autoignition of the end gas causes the type of knock which is referred to as spark knock. This phenomenon is the one in spark ignition engine. It has been observed in fumigation type engine. Diesel knock is because of the increase in the cylinder pressure, shortly after the initiation of combustion. Too early injection causes this type of knock. When the injection is achieved, the conditions in the cylinder are insufficient to provide the autoignition of the diesel fuel. This causes to form a mixture with high fraction of the diesel fuel. Finally, when it burns, undesirable rapid rate of pressure increase occurs. These two knock types could take place in the same cycle.
The problems with fumigation engine can be defined:
The requirements for this conversion is to supply a gaseous fuel metering system and modification of the diesel fuel metering and delivery system. A primary design parameter of fumigation is the amount of gaseous fuel consumption regarding the engine operation range. The major objective in conversion from diesel to dual-fuel is the replacement of maximum amount of diesel fuel with the alternative fuel. The advantageous of the conversion:
According to the model by Liu and Karim:
The pilot fuel is atomized and distributed within its spray cone after its injection to combustion chamber . The entrainment of gaseous fuel and air into spray cone modifies the ignition and combustion process. The rate and amount of gaseous fuel and air are dependent on injection conditions, the quantity of the pilot fuel and the concentration of the gaseous fuel in the cylinder. It can be assumed that there are four zones.
The first one is "pilot fuel unburned zone" which has too rich mixture to burn immediately. "Gaseous fuel unburnt zone" is compressed and heated by the movement of piston and flame front.
The diffusion combustion of the pilot fuel and part of the gaseous fuel propagates towards the core of the pilot fuel. This zone is called "diffusion burned zone". The combustion is assumed to be achieved stoichometricly in this region. The entry of fresh mixtures of diesel and gaseous fuel to this zone takes place during the combustion. "Propagation burned zone" is the flammable region which the flame propagation towards to it. When the fuel charge is entrained to burnt zone from unburnt zone, the release of its energy is realized at the edge of burnt zone. Different zones of cylinder charge cause different temperature and different combustion process, consequently charge temperature and combustion products vary within the cylinder.
The reacting zone within the flame front appears very thin for normal flame propagation due to the fast reaction rate of gaseous fuel-air mixture. This fast reaction rate provides mixture to release its energy immediately at the edge of flame front and converts it to exhaust product through the reacting zone. But at light load, combustion is confined within the fuel spray zone because of difficulty in propagating of the flame through out the very lean mixture. While the expansion process improves the reaction rates of the lean gaseous fuel-air mixture in the reacting zone of the flame front decrease dramatically due to drop in charge temperature initially slowly, later rapidly. The result of this phenomena is to collect more unconverted gaseous fuel and carbon monoxide in the reacting zone which increases emission level.
Figure 1. Combustion zones in dual-fuel operation
In terms of diesel engine combustion, the release of energy can be considered in two parts.
The combustion of gaseous fuel-air mixture consists of three parts.
There are two important points which needs to be pointed out:
Figure 2. Scheme of duel-fuel operation
An increase in the concentration of the gaseous fuel provides the flammable zone enlarge and the mixing rates from the unburned zone to burned zone are increased. Further increase in the concentration of the gaseous fuel causes the flammable zone to extend into the whole gaseous fuel region. Autoignition of the gaseous fuel in the surrounding zone may occur before the flame front reaches.
Although dual fuel operation provides high thermal efficiency comparable to diesel operation and low smoke production, at lower loads lower thermal efficiency and higher HC emission occur.
Natural gas due to its poor cetane number resists burning even in the presence of relatively high temperatures and pressures found in diesel engine. Direct contact with a spark or flame is usually necessary to ignite natural gas. Good quality natural gas (high methane content) has a high octane number and is suitable for high compression ratio engines. Diesel fuel is a mixture of various hydrocarbons. When diesel fuel is introduced into an engine cylinder, it will easily ignite due to temperature and pressure without the aid of a spark or other ignition source. But the presence of particulate or agglomerations of carbon and partially oxidized fuel must be considered when diesel engine runs.
The already developed and tested Detroit Diesel 6V92 dual fuel engine with side port injection requires a fuel pressure of 325 psig, Fig 3.
These engines use natural gas the main source of power and diesel fuel as the ignition source. They are hybrid engines combining the characteristics of both diesel and spark ignition engines. A small quantity of diesel fuel injected into the cylinder autoignites with the effects of cylinder temperature and pressure. The natural gas surrounding the burning diesel droplets ignites and burns like in a spark ignition engine. One important difference between this type of engine and a spark ignition engine is the quality of the ignition source. The diesel pilot engine has a powerful multipoint ignition source. Generally the diesel pilot level is fixed and natural gas is used to increase power above a base power level produced by the diesel pilot.
There is one major problem in two-stroke diesel engine as a pilot ignition engine from the point of view of efficiency and emissions. In two-stroke engines fresh air is used to purge exhaust gas from the previous cycle. A significant percentage (30%) of the fresh air is swept through the cylinder out along with exhaust gases during each cycle. In the situation of premixing natural gas with air flowing into the engine, a large fraction (30 %) would pass through the engine without ever introducing the combustion process. This decrease efficiency and increase the emissions dramatically. In order to eliminate this problem, timed injection of natural gas is utilized. Natural gas is introduced to the cylinder late in the scavenging process so that loss of natural gas is minimized. It is clear that timed injection improves trapping of the natural gas significantly, Table 1.
| Engine |
Peak Brake Efficiency |
| Original Diesel |
39% |
| Premixed Pilot Ignition |
24% |
| Pilot Ignition with NG Injection |
30% |
At idle, the engine operates on 100 % diesel fuel only. Above idle depending upon the driving cycle, the engine switches to burn primarily natural gas, using a small amount of diesel fuel for continued combustion. This is known asthe pilot ignition mode.
In these engines high-speed natural gas valve is capable of delivering natural gas into the cylinder in 0.7 milliseconds to 4.5 milliseconds, depending on the amount of gas required.
Figure 3. Natural Gas Injector in Dual-Fuel.
Currently, there are low pressure and high pressure LNG vehicle fuel systems in the automotive industry. The low pressure type already has been operated in light-duty vehicles with spark ignition type engines. Low pressure LNG fuel system is very similar to those which run on CNG with two exceptions:
Because fuel tank pressure is not allowed to exceed approximately 60 psi, depending on the tank. Both CNG and low pressure LNG systems usually circulate engine coolant through the final regulator to maintain the natural gas at a constant temperature, this feature is very important with LNG due to the extremely low temperature of LNG.
This feature is also used with CNG systems, reduction of the pressure in the regulators typically causes the gas to be cooled due to the expansion it undergoes.
The high pressure type fuel system, Figure 2, is intended to fuel modified diesel engines which require a much pressure fuel gas supply. Generally, high pressure LNG fuel system contains the following components:
Figure 4. LNG FuelSystem
The fuel tanks ( commonly known as " dewars" ) are fabricated with an inner shell surrounded by super insulated vacuum space enclosed in an outer shell. The tanks are cylindrical, which offers the optimum pressure boundary shape with respect to weight, volume and cost. Both ends are closed with ASME flanged and dished ( or torispherical ) heads. To hold the maximum volume with a minimum heat leak, the diameter/length ratio should be as large as can be accommodated within the vehicle chassis dimensions. The fuel tanks and their supports are designed to withstand 8 g acceleration in all directions.
In general, the cryogenic inner vessel is fabricated from austenic stainless steel, 9 % Ni steel, aluminum alloys or composites materials. These materials exhibit an increase in tensile and yield strength without loss of ductility at low temperatures. Outer shell can be made from low carbon pressure vessel quality steel which conforms to ASME specifications. Piping is usually stainless steel both inside and outside the tank. Valving is required to be stainless steel and all liquid lines must have safety shut off actuators for local and remote control.
The vacuum space between outer and inner tank is evacuated at the time of manufacture through a vacuum pump-out valve, which also serves as a relief valve in case of inner shell leak into the vacuum space. Normally, the vacuum space is conditioned to hold static vacuum for many years. Vacuum decay due to damage or accident can be detected when the outer shell becomes noticeably colder than the ambient temperature. Baffles in the inner tank minimize LNG sloshing when vehicle is in motion. This helps to keep the pump suction submerged. Short term pump starvation is compensated by the residual high pressure liquid and gas volume within the vaporizer and buffer, which then provides the fuel to the engine until the liquid pumping resumes.
The pump discharge operating pressure is intentionally kept at 500 psig level to provide this temporary reserve capacity above the 300 psig, minimum engine gas supply requirement.
During the LNG filling operation, the required fuel level is determined from the level gauge. Accidental overfilling is prevented by locating the vapor return pipe inlet at the maximum LNG high level mark in order to assure the necessary vapor space above the liquid in the tank. If the liquid is forced above this level, the excess liquid will discharge through the vapor return connection
The LNG pump consists of two main subassemblies
The entire pump assembly can be mounted either upright or in an inclined position depending on the vertical space availability. When discharge pressure reaches 500 psig, a three way solenoid valve diverts the hydraulic fluid flow back to reservoir. One way check valve at the pump discharge prevents the back flow of high pressure LNG. As the pressure decays below 500 psig, the motor starts again.
The LNG vaporizer serves to vaporize the LNG and to warm it through heat exchange with the coolant. The high pressure LNG from the LNG pump discharge enters the vaporizer tubes surrounded by the engine coolant at 71 C and exits near that temperature. To prevent vaporizer coolant freeze-up the engine coolant should be circulating before, or simultaneously with starting the LNG pump.
At any time when continuos fuel demand from the engine is interrupted, the cold LNG remaining in the vaporizer circuit will expand and rise in pressure. The CNG buffer space, in addition to the internal volume of the vaporizer, is provided to limit such pressure increase to less than 1000 psig without venting.
In dual fuel engines, the products from the combustion of diesel pilot along with the emittent from the combustion of the main premixed charge contribute to pollutant formation in dual fuel engines.
CO emissions with dual fuel engines is higher than that with diesel engines, Figure 5. Because, dual fuel engines operate unthrottled, This causes air fuel mixture becomes leaner as the load is reduced. The result of this is high concentration of incomplete combustion products.
Figure 5. Comparison of NOX emissions from 5 PING and 10 #2 D vehicles.
Figure 6. Comparison of CO emissions from5 PING and 10 #2 D vehicles
The combustion of the biggest part of the fuel is achieved under lean premixed conditions which flame temperatures are less than in diesel engines. This provides the NOX emission lower with dual fuel engine than that with diesel engine, Figure 4. But it should be noted that NOX emission with dual fuel engine is higher than that with spark ignition engine. Because the combustion of diesel pilot takes place under near stoichometric conditions at shortly before top-dead center. Relatively long residence time contribute to NOX formation more.
Hydrocarbon emissions from dual-fuel engines are much higher than those from diesel engines, Figure 7. HC emissions mainly stem from the unburned mixture forced into crevice volumes during compression. Piston land in other word the space between the side of the piston and the cylinder wall is one of these crevice volumes. This section is tended to be relatively large in diesel engines, although this approach is changed lately due to emission concern. When diesel engine is converted to dual-fuel engine, this crevice volume holds relatively large amount of unburned mixture, increasing HC emission.
Figure 7. Comparison of HC emissions from 5 PING and 10 #2 D vehicles.
Figure 8. Comparison of PM emissions from 5 PING and 10 #2 D vehicles.
Another resaon for the increase in HC emission is due to overscavenging of the cylinder with dual-fuel engines. Usually, high rated diesel engines there is overlap between the intake valve opening and exhaust valve closing. This approach provides air to blow through the cylinder, increasing the efficiency by removing the exhaust products, the lowering the cylinder temperature and increasing the power output. But in dual-fuel engines the same process causes the unburned air-fuel mixture blow out the exhaust. Finally, under light conditions too lean burn air-fuel mixture fails to complete combustion, resulting in unreacted and partially reacted fuel in the exhaust.
CO2 emission with dual-fuel engine tends to be lower compared to diesel engine, Figure 9.
|
No |
Test Number |
Manufacturer |
Vehicle Type |
Engine Type |
Vehicle |
Mileage |
Gross
Weight |
Testing
Weight |
Engine
Power |
Engine
Speed |
Fuel Type |
| PING1 |
HM-2114-D2/LNG |
GMC |
Bus |
DDC 6V-92TA PING |
1983 |
202254 |
36900 |
32856 |
277 |
2100 |
LNG |
|
PING2 |
HM-2579-D2/LNG |
IK |
Bus |
DDC 6V-92TA PING |
1991 |
8293 |
40200 |
35686 |
277 |
2100 |
LNG |
| PING3 |
HM-2579-CD/LNG |
IK |
Bus |
DDC 6V-92TA PING |
1991 |
8311 |
40200 |
35686 |
277 |
2100 |
LNG |
|
PING4 |
HM-2579-CD/CNG |
IK |
Bus |
DDC-6V-92TA PING |
1991 |
8329 |
40200 |
35686 |
277 |
2100 |
CNG |
| PING5 |
HM-2579-D2/CNG |
IK |
Bus |
DDC 6V-92TA PING |
1991 |
8347 |
40200 |
35686 |
277 |
2100 |
CNG |
|
D1 |
HM-1995-D2 |
GMC |
Bus |
DDC 6V-92TA |
1983 |
24031 |
36900 |
32160 |
277 |
2100 |
#2 D |
| D2 |
ADEQ-4433-D2 |
TMC |
Bus |
DDC 6V-92TA |
1988 |
15858 |
36900 |
31339 |
277 |
2100 |
#2 D |
|
D3 |
PaT-2077-D2 |
ORN |
Bus |
DDC 6V-92TA |
1991 |
1064 |
39221 |
34046 |
277 |
2100 |
#2 D |
| D4 |
PTTW-421-D2 |
BIA |
Bus |
DDC 6V-92TA |
1990 |
8668 |
38000 |
33437 |
277 |
2100 |
#2 D |
|
D5 |
RTA-9001-D2 |
TMC/RTS |
Bus |
DDC 6V-92TA |
1990 |
128901 |
36900 |
31499 |
277 |
2100 |
#2 D |
| D6 |
BSDA-8455-D2 |
FLX |
Bus |
DDC 6V-92TA DDEC II |
1988 |
39500 |
33757 |
277 |
2100 |
#2 D |
|
|
D7 |
BSDA-8466-D2 |
FLX |
Bus |
DDC 6V-92TA |
1988 |
178798 |
39500 |
33828 |
277 |
2100 |
#2 D |
| D8 |
PaT-3510-D2 |
NPLN |
Bus |
DDC 6V-92TA |
1983 |
132234 |
37680 |
33669 |
277 |
2100 |
#2 D |
|
D9 |
MDTA-9068-D2 |
FLX |
Bus |
DDC 6V-92TA |
1990 |
206506 |
39500 |
33137 |
277 |
2100 |
#2 D |
| D10 |
MTC-2207-D2 |
GLG |
Bus |
DDC 6V-92TA DDEC II |
1993 |
55948 |
39600 |
33234 |
277 |
2100 |
#2 D |
Figure 9. Comparison of CO2 emissions from5 PING and 10 #2 D vehicles.
Table 3. Emissions Data
|
No |
Test Number |
Runs of Test |
CO(Avg.) |
NOx(Avg.) |
HC(Avg.) |
PM(Avg) |
CO2(Avg.) |
| PING1 |
HM-2114-D2/LNG |
4 |
22.2 |
29.9 |
40.52 |
1.10 |
2493 |
|
PING2 |
HM-2579-D2/LNG |
4 |
39.3 |
24.7 |
62.7 |
0.97 |
3059 |
| PING3 |
HM-2579-CD/LNG |
3 |
42.7 |
29.9 |
73.67 |
0.76 |
2886 |
|
PING4 |
HM-2579-CD/CNG |
4 |
38.7 |
31.5 |
68.45 |
0.71 |
2881 |
| PING5 |
HM-2579-D2/CNG |
3 |
31.8 |
39.9 |
51.47 |
0.70 |
3096 |
|
D1 |
HM-1995-D2 |
3 |
9.5 |
41.2 |
2.97 |
1.86 |
2974 |
| D2 |
ADEQ-4433-D2 |
5 |
19.2 |
46.6 |
1.49 |
4.15 |
3011 |
|
D3 |
PaT-2077-D2 |
3 |
13.1 |
30.9 |
2.53 |
1.32 |
3869 |
| D4 |
PTTW-421-D2 |
7 |
10.4 |
20.4 |
1.82 |
3.47 |
3048 |
|
D5 |
RTA-9001-D2 |
6 |
20.9 |
27.2 |
1.85 |
1.81 |
3129 |
| D6 |
BSDA-8455-D2 |
4 |
25.4 |
41.4 |
2.14 |
1.09 |
2977 |
|
D7 |
BSDA-8446-D2 |
4 |
22.3 |
38.3 |
3.2 |
3.1 |
3226 |
| D8 |
PaT-3510-D2 |
5 |
10.9 |
44.1 |
3.46 |
1.42 |
3427 |
|
D9 |
MDTA-9068-D2 |
4 |
12.6 |
27.5 |
1.74 |
1.85 |
3189 |
| D10 |
MTC-2207-D2 |
4 |
9.5 |
25.8 |
2.77 |
1.65 |
3761 |
Figure 10. Comparison of average emissions from5 PING and 10 #2 D vehicles.
In contact with hot surfaces, natural gas will ignite at much higher temperatures ( 540 C ) than diesel fuel (225 C). However the ignition energy required ( such as from a spark ) is only about 20 % higher for the natural gas when burning, LNG quickly vaporizes and is rapidly consumed without residue and produces somewhat lower flame temperature. LNG does not pool or cling to surfaces. Even if an LNG pool should exist for a short time, the surface pol burning rate is high.
The LNG vapors, when present in air within their flammability limits from 5 to 15 % can be ignited and, if in confined space, will cause an explosion. Therefore, every precaution must be taken to avoid creating this combination of ignition source, vapor concentration and confinement.
Skin contact with LNG ( or bare metal surfaces ) will cause frost burns.
The allowable threshold limit for LNG vapor concentration in breathing air is much higher (750 times) than that for diesel fumes and 35 times higher than that for gasoline fumes.