1. Field of the Invention
The invention relates to a high pressure gas cycle for use in an engine. The invention also relates to an electrical power generating plant.
2. Background of Related Art
The Otto cycle shown diagrammatically in FIG. 1 [Prior Art], with pressure as the ordinate and volume as the abscissa, is a cycle that closely represents the explosion and compression stages of a gasoline automobile engine. In the Otto cycle, the air is drawn in at atmospheric pressure, shown at 1, then compressed isentropically to a high pressure with fuel mixed in the air, shown at 2. The fuel air mixture explodes at a theoretically constant volume to form a combustion gas having an increased pressure, shown at 3. The combustion gas expands isentropically back to the original inlet volume, shown at 4, where the combustion gas is then discharged. As shown in FIG. 1, the combustion gas discharges at a pressure above atmosphere, shown at 4, and then the pressure drops to atmosphere as it exhausts from the engine, not shown. The energy that could be available by expanding from exhaust pressure to atmospheric pressure, represented by the difference between 4 and 1 shown in FIG. 1, is wasted.
The theoretical efficiency of the Otto Cycle is defined by the relationship 1-(V.sub.2 /V.sub.1).sup.(K-1) where V.sub.1 /V.sub.2 is the ratio of the volume before compression to the volume after compression, commonly called the compression ratio. K=ratio of specific heat at constant pressure to specific heat at constant volume=C.sub.p /C.sub.v.
Another cycle commonly used in automobile engines is the Diesel cycle. The pressure vs. volume diagram for a Diesel cycle is shown in FIG. 2 [Prior Art]. In a Diesel cycle, air is drawn in at atmospheric pressure, shown at 5, and then the air is compressed, shown at 6. The fuel is injected after the air is compressed and burns at somewhere near constant pressure to form a combustion gas, shown at 7. The combustion gas expands isentropically back to the original inlet volume, shown at 8, where the combustion gas is then discharged. As shown in FIG. 2, the combustion gas discharges at a pressure above atmosphere, shown at 8, and then the pressure drops to atmosphere as it exhausts from the engine, not shown. The energy that could be available by expanding from exhaust pressure to atmospheric pressure, represented by the difference between 8 and 5 shown in FIG. 5, is wasted.
The advantage of the Diesel cycle compared to the Otto cycle is that the compression ratio can be made much higher than that for the Otto Cycle, because the fuel is not mixed in the air, and therefore the rise during compression in temperature will not ignite the fuel until after the fuel is injected at the high pressure. In general, this increase in compression ratio for the Diesel Cycle enables the Diesel cycle to achieve higher efficiencies than are possible with the Otto Cycle.
The compression ratio for the Otto Cycle is usually limited to about 10 to 1, corresponding to a pressure ratio of about 25 to 1. The reason for this is that at higher ratios the fuel air mixture becomes so hot that the explosion occurs before the mixture is fully compressed. This preignition or detonation actually decreases the power output.
In the Diesel engine the fuel injection occurs after compression, and only air is being compressed during the compression cycle. Therefore, typical compression ratios are 23 to 1, corresponding to pressure ratios of 82 to 1. This is the basic reason why the Diesel cycle is more efficient than the Otto cycle.
A further cycle is the Brayton or Joule cycle, as shown in FIG. 3 [Prior Art]. In the Brayton cycle, air is drawn into a compressor at atmospheric pressure, shown at 9, and then compressed to a high pressure, shown at 10. Fuel is injected into the compressed air in the combustor, where it burns at nearly a constant pressure (except for friction losses in the combustor) to form a combustion gas, shown at 11. The combustion gas expands isentropically back to atmospheric pressure, shown at 12. In this case, the expansion to atmospheric pressure is advantageous, and the theoretical efficiency is like the Otto cycle above in that the theoretical efficiency is equal to 1-(V.sub.2 /V.sub.1).sup.(K-1) where V.sub.1 /V.sub.2 is again the volume ratio of specific volume at atmospheric pressure divided by the specific volume at the pressure at which burning starts. The Brayton cycle is the cycle commonly used in gas turbines, and is limited in efficiency by the fact that the temperature of the gas entering the turbine is nearly the same as the combustion temperature. Therefore the combustion temperatures possible in a gas turbine system are usually limited to approximately 2300 to 2600.degree. F. However, the Brayton cycle does have an advantage over the Otto cycle in that complete expansion back to essentially atmosphere is achieved.
A disadvantage of the Brayton cycle is that as pressure ratios or compression ratios are increased the temperature leaving the compressor and entering the combustor becomes higher. Therefore, less fuel energy can be added because of the temperature limit of the turbine. For this reason, although efficiency can be increased by increasing the pressure ratio in a gas turbine cycle, the output gradually decreases as higher pressure ratios are used. Therefore, it is common practice to limit the pressure ratio in industrial gas turbines to about 10 to 20 atmospheres.
Table 1 shows typical theoretical performance calculations for the Brayton cycle, based on air standard data and constant mass flow through the cycle.
TABLE 1 A B C D E F G H I 1 PS/P1 32 32 32 32 32 32 32 32 2 T1 520 520 520 520 520 520 520 520 3 T3 2810 2810 2810 2810 2810 2810 2810 2810 4 P1/J 2.7201 2.7201 2.7201 2.7201 2.7201 2.7201 2.7201 2.7201 5 V1 13.089 13.089 13.089 13.089 13.089 13.089 13.089 13.089 6 EFF COMP 1 0.9 0.85 0.9 0.9 0.9 0.9 0.9 7 COOL FACT 1 1 1 0.9 0.8 0.7 0.6 0.5 8 N/(N - 1) CO 3.463 3.1167 2.94366 3.463 3.895875 4.452429 6.1945 6.2334 9 (N - 1)/N CO 0.288757 0.320852 0.339726 0.288767 0.256682 0.224597 0.192511 0.160428 10 N COMP 1.406009 1.472434 1.514522 1.406009 1.345319 1.289651 1.238407 1.19108 11 EFF TURB 1 0.9 0.85 0.9 0.9 0.9 0.9 0.9 12 N/(N - 1) TU 3.463 3.847778 4.074118 3.847778 3.847778 3.847778 3.847778 3.847778 13 (N - 1)N TU 0.288767 0.25989 0.245452 0.25989 0.25989 0.25989 0.25989 0.25989 14 N TURB 1,406009 1.351151 1.325297 1.351151 1.351151 1.351151 1.351151 1.351151 15 W IN 212.1196 251.5701 276.9102 235.6885 221.0261 207.4814 194.9616 183.3819 16 V1/V2 11.76285 10.52493 9.858518 11.76285 13.14636 14.6926 16.4207 18.35205 17 T2/T1 2.72043 3.0404 3.245924 2.72043 2.434134 2.177968 1.94876 1.743674 18 T2 1414.624 1581.008 1687.88 1414.624 1265.75 1132.543 1013.355 906.7106 19 HEAT IN 330.8437 291.3941 266.0545 330.8437 366.1417 397.725 425.9845 451.2699 20 V3 2.210342 2.210342 2.210342 2.210342 2.210342 2.210342 2.210342 2.210342 21 W EXPAN 421.3532 395.5742 381.6844 395.5742 395.5742 395.5742 395.5742 395.5742 22 NET WORK 209.2335 144.0041 104.7742 15.6857 174.6481 188.0928 200.6127 212.1923 23 EFF. 0.632424 0.49419 0.393807 0.483267 0.478723 0.472922 0.470939 0.470212 A J K L M N O P 1 P2/P1 32 32 32 32 32 32 32 2 T1 520 520 520 520 520 520 520 3 T3 2810 2810 2810 2810 2810 2810 2810 4 P1/J 2.7201 2.7201 2.7201 2.7201 2.7201 2.7201 2.7201 5 V1 13.089 13.089 13.089 13.089 13.089 13.089 13.089 6 EFF COMP 0.9 0.9 0.9 0.9 0.9 0.9 0.9 7 COOL FACT 0.4 0.3 0.2 0.1 0.05 0.04 0.03 8 N/(N - 1) CO 7.79175 10.389 15.5835 31.167 62.334 77.9175 103.89 9 (N - 1)/N CO 0.128341 0.096256 0.06417 0.032085 0.016043 0.012834 0.009626 10 N COMP 1.147237 1.106508 1.068571 1.033149 1.016304 1.013001 1.009719 11 EFF TURB 0.9 0.9 0.9 0.9 0.9 0.9 0.9 12 N/(N - 1) TU 3.847778 3.847778 3.847778 3.847778 3.847778 3.847778 3.847778 13 (N - 1)N TU 0.25989 0.25989 0.25989 0.25989 0.25989 0.25989 0.25989 14 N TURB 1.351151 1.351151 1.351151 1.351151 1.351151 1.351151 1.351151 15 W IN 172.6651 162.7406 153.5437 145.0155 140.9852 140.197 139.4146 16 V1/V2 20.51057 22.92296 25.6191 28.63234 30.26937 30.60784 30.9501 17 T2/T1 1.560171 1.39598 1.249068 1.117617 1.057174 1.045484 1.033922 18 T2 811.289 725.9096 649.6154 581.1609 549.7306 543.6515 537.6396 19 HEAT IN 473.8944 494.1378 512.2509 528.4577 535.9099 537.3512 538.7766 20 V3 2.210342 2.210342 2.210342 2.210342 2.210342 2.210342 2.210342 21 W EXPAN 395.5742 395.5742 395.5742 395.5742 395.5742 395.5742 395.5742 22 NET WORK 222.9091 232.8337 242.0305 250.5587 254.589 255.3772 256.1596 23 EFF. 0.470377 0.471192 0.472484 0.474132 0.47509 0.475252 0.475447
In Table 1 the following definitions are used:
P.sub.2 /P1 is the pressure ratio PA1 T1=air inlet temperature in .degree. R (degrees Rankin) PA1 T3=combustion temperature in .degree. R PA1 P1/J=inlet pressure in lbs/ft.sup.2 /778.2 PA1 V1=inlet specific volume in ft.sup.3 /lb. PA1 EFF COMP is polytropic compression efficiency PA1 COOL FACT is a factor that is multiplied by (N-1)/N, where (N-1)/N=(k-1)/K/comp EFF., and (K-1)/K=0.2888. This factor shows a new value of (N-1)/N that simulates continuous intercooling during the compression process. Actual intercooling is a step-by-step process, but this simulation shows the approximate effect of intercooling. PA1 (N-1)/N CO=(K-1)/K/EFF COMP.times.COOL FACT PA1 N Comp. is the polytropic exponent used in the equation for work PA1 EFF TURB is the polytropic efficiency of the turbine PA1 (N-1)/N TU=(k-1)/K.times.EFF TURB PA1 WIN=Compressor work in Btu/lb PA1 V1/V2=Compression ratio PA1 T2/T1=Ratio compressor discharge temperature/inlet temperature PA1 T2=Compressor discharge temperature .degree. R PA1 Heat IN=Heat Input from T2 to T3, assuming specific heat=0.2371 Btu/lb.degree. F. PA1 V3=Specific volume at turbine inlet in cu.ft./lb PA1 W EXPAN=Turbine work output in Btu/lb PA1 NET WORK=W EXPAN-WIN PA1 EFF=NETWORK/HEAT IN=cycle efficiency PA1 a wall structure defining an interior chamber; PA1 a first reflecting surface for reflecting a pressure wave within said interior chamber; PA1 a second reflecting surface for reflecting said pressure wave within said interior chamber, wherein said first and second reflecting surfaces being constructed and arranged to resonate said pressure wave in said interior chamber; PA1 at least one first inlet for introducing a first gas into said interior chamber; and PA1 at least one outlet from said interior chamber for drawing off a pressurized gas from said interior chamber. PA1 at least one combustion chamber; PA1 at least one compressor constructed and arranged to provide a compressed gas to said at least one combustion chamber; and PA1 at least one turbine blade constructed and arranged to be driven by a pressurized gas formed in said combustion chamber; wherein said combustion chamber comprises: PA1 at least one turbine engine; PA1 at least one electrical generator connected to said turbine engine; wherein said turbine engine comprises: PA1 introducing a combustible gas into a combustion chamber having first and second reflecting surfaces that are constructed and arranged to provide a resonating pressure wave reflecting between said first and second reflecting surfaces, said combustible gas being introduced into said combustible chamber at a frequency such that said resonating pressure wave ignites said combustible gas to thereby form a resonating pressure wave; PA1 introducing a second gas into said combustion chamber at a location and frequency such that said pressure wave compresses and combines with said second gas to form a pressurized gas having a temperature lower than a combustion temperature of said combustible gas; and PA1 withdrawing said pressurized gas from said combustion chamber.
In column C, the Brayton cycle efficiency is listed as 0.494. This is higher than the actual efficiency of a gas turbine because leakage losses, cooling air losses, pressure drop in the combustor, and losses due to kinetic energy of the gases leaving the turbine column D, where efficiencies of 85% are used for compressor and turbine, were excluded.
It would be advantageous if one could combine the constant volume or explosion cycle as shown in the Otto cycle at the high pressure end, and at the same time expand the volume all the way to atmospheric pressure at the exhaust end, as shown FIG. 4. FIG. 4 is a theoretical complete expansion cycle.
An almost complete expansion cycle was made by Sargent, in which the air inlet to the engine is throttled to take in less air volume and thereby allow for an increase in volume in the exhaust. However, this Sargent cycle was not a success in a reciprocating engine because of the high mechanical friction losses.
Thus, there is a need for a complete expansion cycle that is suitable for use in a reciprocating engine, which substantially avoids wasting energy due to exhaust pressures that are greater than atmospheric pressure.
There is also a need for an improved combustion chamber that is capable of supplying compressed gas to a turbine blade at temperatures significantly below the combustion temperature of the fuel being burned.
There is a further need for an improved apparatus for supplying a compressed gas having significantly reduced friction losses.
Electrical power plants utilizing turbine engines to drive electrical generators produce large amounts of combustion gasses which contain carbon dioxide and byproducts such as nitrogen oxides. Furthermore, the exhaust gas from conventional turbine engines usually has a temperature of about 700.degree. F. to about 1240.degree. F. Typically a Rankine cycle system is used to recovery valuable energy from the exhaust gas. However, efficient low temperature vapor turbines usually cannot be used because the exhaust temperature from a conventional turbine engine is too high. Exhaust temperatures from conventional turbine engines usually require the use of an expensive steam turbine to recover the energy.
Thus, there is a need for an electrical generating power plant comprising more efficient turbines to reduce the quantity of combustion gasses produced, and for turbine engines having exhaust temperature suitable for use in driving low temperature vapor turbines.