Abstract:
A compressed gas dispensing station having a high pressure gas compressor with a cyclic control system for selective recirculation of cooled, ultra high pressure gas through the compression chamber after the end of the compression stroke for scavenging hot compressed gas from the compression chamber and providing a residual, partially-expanded replacement gas for the expansion stroke which is mixed with the incoming, new charge of gas for a cryogenic gas at the start of compression and a relatively low temperature gas at the end of compression for a single stage compressor. The cyclic control system times the opening and closing of two delivery valves for separate 4000 psi and a 3600 psi branches, the delivery valve for the 4000 psi branch also regulating recirculation of 4000 psi cooled gas through the compression chamber for the 3600 psi branch after the end of the compression stroke to cool the chamber and replace the hot residual compression gas with a cold expanded gas, which is further expanded in the expansion stroke. Compressed gas is collected and stored in two receiver tanks having different pressures for mixing and dispensing at a customer service station according to customer requirements.

Description:
This invention is the subject of provisional application Serial No. 60/049,298, filed Jun. 11, 1997, entitled, “High Pressure Compressor with Internal Cooled Compressor”. This invention further advances the implementation of our initial invention described in patent application Ser. No. 08/379,147 filed Jan. 27, 1995, entitled, “High Pressure Compressor With Internal Cooled Compression,” now U.S. Pat. No. 5,769,610, issued Jun. 23, 1998. 
    
    
     BACKGROUND OF THE INVENTION 
     The invention utilizes a balanced, dual crank reciprocator of the type disclosed in our U.S. Pat. No. 5,674,053, issued Oct. 7, 1997, entitled, “High Pressure Compressor with Controlled Cooling During the Compression Phase,” and U.S. Pat. No. 5,716,197, issued Feb. 10, 1998 entitled, “High Pressure Compressor with Internal Inter-Stage Cooled Compression having Multiple Inlets.” 
     The present invention defines a gas compressor and dispensing station with a new and improved cyclic control system for high and ultra high pressure compressors. The compressor in this system is capable of achieving in one stage, ultra high pressure ratios of over 40/1. The invented system eliminates the need for multi-stage compressors, compressor assemblies, particularly for natural gas compressors, requiring delivery pressures of 3600-4000 psi, for NGV (natural gas vehicle) supply stations and natural gas line transportation systems. 
     This invention relates to a gas compressor with a new cyclic control system that is provided with a control module and sensors for controlling a group of electronically activated, electro-hydraulic valves for regulating pressurized gas flow through the compressor. The electro-hydraulic valves are selectively operated during the reciprocal cycle of the compressor in an electronic-loop of cycle control format for routing gas at two discrete pressures through separate circuits in the compressor. 
     In this specification, the system described in our provisional application is refined with the construction of the electro-hydraulic valves controlling flow of high pressure gases from the compressor to the respective high pressure gas receiving tanks being detailed. 
     The single-stage compressor of this invention is designed to be inexpensively fabricated and operated for alternate fuel vehicles. Natural gas is a relatively clean, burning fuel, and, comprised largely of methane, has advantages over other hydrocarbon fuels in minimizing production of the greenhouse gas, carbon dioxide. Although natural gas is relatively abundant, it has not been widely used as an alternate fuel for vehicles because of the lack of a distribution system. Many cities have an existing infrastructure of gas distribution lines for heating and cooking. However, these are relatively low pressure lines, 30-40 p.s.i. at the street. At this pressure, the gas volume for powering a vehicle is too large to provide the driving range deemed acceptable. 
     Pressurized gas vessels have been designed to contain natural gas at the high pressure necessary for the fuel capacity for the driving range desired in a reasonably sized bottle. One fueling alternative is to replace prefilled gas bottles at a refueling station. It is not economical, however to prefill bottles and deliver such prefilled bottles to fueling stations for exchange with customer bottles. 
     While bottles may be pre-filled on the site of the fueling station, this requires an on-site compressor, and, if a fueling station has an on-site compressor it may as well fill a customer&#39;s fuel bottle already in the customer&#39;s vehicle. For the fuel to be competitively priced compared with gasoline, the on-site compression system must be efficient and productive, requiring minimal storage of compressed gas. 
     The high pressure gas compressor of this invention utilizes a positive displacement compressor with an expansion gas scavenging of the residual gases in the compressor. By strategic timing of the gas flow in the compression and expansion cycle, gas can be compressed in a single stage with a resultant temperature well within the thermal limits of the structural components of the compressor. 
     The gas compression system of this invention is targeted toward the natural gas industry both for high pressure transportation of gas in gas lines, and for destination stations where natural gas is dispensed to customer bottles for use as a vehicle fuel. It is to be understood, however, that the gas compression system can be utilized for gasses other than fuel gas where a cost-effective, high-pressure compression is required. 
     SUMMARY OF THE INVENTION 
     The ultra high pressure gas compressor in the compressed gas dispensing station of this invention is characterized by a control system controlling two high-pressure, electro-hydraulic valves. One valve is a delivery valve for regulating a 3600 psi branch, and the second valve is a delivery and recirculation valve for regulating a 400 psi branch. The compressor is also provided with an automatic or electro-hydraulic intake valve for regulating gas intake into the compressor. 
     The compressor cycle starts with the intake and mixture of an initial remaining charge of precooled, expanded cryogenic gas injected at the end of the previous cycle, followed by the compression stroke achieving 4000 psi. Pressure is monitored by an electronic pressure transducer, which is informing an electronic control module (ECM), that controls the activation of the delivery recirculation valve (DRV). This valve (DVR) is provided with two channels, one conducting the high pressure relative hot gases through a check valve, into a 4000 psi cooled receiver tank, and the second channel conducting a recirculated cooled gas from the cooled receiver tank back into the compression chamber. 
     The recirculation process is started by the activation of the 3600 psi delivery valve, which produces a pressure drop in the compression chamber, which causes the opening of the recirculation check valve, controlling the exit of 4000 psi gas from the cooled receiver tank. In that moment, the scavenging process of purging the hot gases toward the 3500 psi branch, and replacing the displaced gas with cooled high pressure 4000 psi gases is accomplished. 
     The 40-1 expansion of the cooled and high pressure 4000 psi gas, that remains in the compression chamber, produces a very low temperature cryogenic gas, which is mixed with the new intake charge, producing a low temperature mixture, also cryogenic, at the start of the compression cycle. The compression stroke will produce at the end, a relatively low temperature, high pressure delivery gas for the single stage compression. 
     The result will be an equivalent of an isothermic compression cycle. The high pressure compressor of this invention is particularly adapted for use in a gaseous fuel dispensing station. The embodiments described in this specification are designed for natural gas, which is typically a mixture of hydrocarbon gases, primarily methane. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic drawing of the compressor system with a cross section through the head and compression chamber of the compressor. 
     FIG. 2 is a schematic drawing of an alternate configuration of the compression system showing a customer&#39;s fuel bottle. 
     FIG. 3 is a cross-sectional view of a typical electro-hydraulic gas valve assembly for operation under ultra high pressures. 
     FIG. 4 is a cross-sectional view of the actuator control module in the assembly of FIG.  3 . 
     FIG. 5 is a cross-sectional view of the control module taken on a horizontal plane through the piston pusher in FIG.  4 . 
     FIG. 6 is a cross-sectional view of the control spool valve module in the assembly of FIG.  3 . 
     FIG. 7 is a cross-sectional view of the spring return module in the assembly of FIG.  3 . 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring to FIG. 1, a first embodiment of a high pressure gas dispensing station  8  featuring a single stage compressor  10  is schematically illustrated. 
     The compressor  10  has a cylinder  11 , and a piston  12 , and is provided with a cylinder head  13 , having an intake valve  14 , provided with two hydraulic stops  15  and  16  and a spring  17 . The intake valve  14  regulates gas intake through an intake channel  18 . 
     The compressor  10  is provided with a pressure transducer  19 , facing the compression chamber  25  for monitoring the pressure in the compression chamber  25 . The compressor is also provided with an electro-hydraulic discharge valve  20  for the 3600 psi delivery branch  21 . The electro-hydraulic valve module  22 , receives the hydraulic activation fluid from the hydraulic source  23 , and an activating electronic control impulse through the wire  24  from the electronic control module  30 . 
     The 3600 psi gas delivery branch delivers the gas to the cooled receiver tank  31 , which includes a heat exchanger to reduce gas temperature to at least ambient temperature. The discharge valve  20  is controlled by the electronic control module with input from a pressure transducer  32  and a temperature transducer  33  for timely operation of the valve. A final temperature transducer  34 , monitors the final temperature of the gas delivered to the gas dispenser  35 . 
     The compressor  10  is also provided with a valving device  38  having an electro-hydraulic discharge and recirculation valve (HDRV)  40 , controlling the 4000 psi gas for delivered and circulated gas. The valves  20  and  40  are designed for balanced pressure on the valve head shoulder  53  and stem shoulder  54  enabling rapid electro-hydraulic activation. The 4000 psi gas branch is provided with a discharge channel  41  controlled by a check (one way) valve  42 , conducting the hot 4000 psi gas to the cooled receiver thank  43 , which is similar to tank  31 . The discharge and recirculation valve  40  is controlled by the control module  30  with input from the temperature transducer  52 . 
     The electro-hydraulic discharge and recirculating valve  40  receives the hydraulic activation fluid from the source  49  and is electronically connected by the wire  50  with the electronic control module  30  for timely operation. 
     The cooled 4000 psi gas emerging from the cooled receiver tank  43 , is conducted in the passage  44  toward the gas dispenser  45 , and in the gas passage  46 , toward the recirculation “one way” check valve  47 , and through the recirculation channel  48 , back into the port of the electro-hydraulic discharge and recirculation valve (EDRV)  40 . The final temperature of the delivered gas (4000 psi) is monitored by the temperature transducer  49  and used as a control factor for regulation of the operation of the compressor by the control module  30 . 
     The combined gas dispensers  35  and  45  form the gas dispenser cascade for the base station. 
     The compressor cycle control system starts from the moment is which the 4000 psi pressure is reached, close to the end of the compression stroke. The pressure is monitored by the pressure transducer  19 , and the electronic control module  30  signals the activation of the electro-hydraulic discharge and recirculating valve  40 , to discharge the 4000 psi hot gas, to the cooled receiver tank  43 , through the one way check valve  42 . 
     The electro-hydraulic discharge valve  20  is activated after an “angular/time” interval “A”, opening the 3600 psi gas discharge branch  21 , and producing a pressure drop in the compression chamber  51 . In that moment, the check valve  42  is closed, and the check valve  47  is open, starting a flow of cooled 4000 psi gas recirculated from the cooled receiver  43 , to the compression chamber  51 , producing a “scavenging effect” of the hot gases, from the compression chamber  51 , by the open electro-hydraulic discharge valve  20  to the 3600 psi delivery branch  21 . 
     After an “angular-time” internal “B”, the electro-hydraulic valve  20 , is closed by a signal from the electronic control module  30 , and after an “augular/time” interval “C“, the compression chamber  51 , is charged with 4000 psi cooled gas, and the electro-hydraulic discharge recirculation valve  40  is closed. 
     From the “moment C” to the end of the expansion stroke, the remnant gas in the compressor will have a cryogenic temperature producing an “internal cooling fluid of gas”, which will be mixed with the new intake gases. Timing of the sequence is controlled by the electronic control module  30  for optimizing production of high pressure gas within safe operating temperature ranges. 
     The new mixed gas, at the beginning of the compression, will have a very low temperature approaching a cryogenic level, resulting at the end of the compression stroke, in a final relatively low temperature for the delivered high pressure gas. 
     The general compression cycle can be considered an approximation of an isothermic compression cycle, with the lowest energy consumption, obtained in “one single compression stage”. 
     Referring now to FIG. 2, a second embodiment of a high pressure gas dispensing station  60  is schematically illustrated. The dispensing station  60  includes a single-stage gas compressor  62  that utilizes a dual-crank piston assembly  64  that provides a dynamic balance which eliminates side forces of the piston  66  again the cylinder  68 . This enables the ultra high pressures in the range of 4000-5000 to be obtained in a single stage. However, because of the temperature generated in a gas compression of this magnitude, an internal cooling is required to reduce the temperature of the discharged gas to a level within the thermal limits of the system components. Key to the internal cooling is the admission of high pressure cooled gas at the completion of the compression cycle to scavenge residual hot gases and replace the displaced gases with a high-pressure partially expanded gas that cools to cryogenic levels when further expanded during the expansion cycle. Because a portion of the product compressed gas is used for cooling, precise timing of the sequencing is required to maintain efficiencies of the system. 
     System timing is effectively controlled by an encoder  70  that is connected to one of the two crank shafts  72  that feeds a cycle phase signal to a central electronic module  74  that is the universal electronic processor and controller for the dispensing station. It is understood that separate control systems may be employed for the tasks of compressing the gas and dispensing the gas. 
     The central electronic control module  74  receives signals from a variety of sensors and controls the operation of the various electronic components. Because of the partial compressibility of control fluids utilized as an actuating medium and the compressibility of gases in the system, a system program is utilized by the internal processor of the central electronic control module to continually adjust the system to obtain the desired effect of the timed events. The electro-hydraulic regulating valves are designed for precision operation with minimum reaction time and minimized after effects. 
     In the system of FIG. 2, two gas pressure regulator valves  76  and  78  control the discharge of pressurized gas from two storage tanks  80  and  82  maintained at a differential pressure to achieve the cooling objectives of the system during compression. The dispensing station  60  has a pressurized dispenser  84  with a high-pressure gas line  86  that connects to a customer&#39;s high-pressure gas bottle  88  that may remain in the customer&#39;s vehicle (not shown). The use of both a high pressure storage tank  80  and a lower pressure storage tank  82  allows a depleted bottle to be filled first with the lower pressure gas before being topped with the higher pressure gas to the ultimate pressure required by the customer. In this manner, high pressure gas is conserved for final pressurization and in certain instances may not be used for those customers with only lower pressure requirements. 
     It is to be understood that in a gas transportation system, the dispenser  84  is not used and the gas pressure regulator valves  76  and  78  are used to maintain a mix with the desire line pressure in the range between the lower pressure gas in the storage tank  82  and the higher pressure gas in the storage tank  80 . The pressures in the respective tanks  80  and  82  are pre-determined by the system user within certain parameters to insure that for a given high pressure, the differential is sufficient to allow for internal cooling as described. For example, the high pressure tank may be maintained at 20% higher pressure than the lower pressure tank to provide an adequate margin for expansion cooling. The set pressures are maintained by pressure transducers  90  and  92  for the tanks  80  and  82 , respectively. The transducers sense the respective pressure and transmit electrical signals through lines  94  and  96  to the control module  74 . After processing, the control module  74  regulates the operation of the compressor  62  to maintain the tanks within the acceptable storage range and differential pressure. 
     Operation of the compressor  62  is substantially the same as for the compressor  10  in the previously described embodiment. Regulating the operation of the compressor  62  is accomplished by four electro-hydraulic valves  98 ,  100 ,  102 , and  104 . The gas admission valve  102 , is not required to perform at the higher pressures and therefore need not have the complexity of the other valves which preferably have an identical construction, as detailed in FIGS. 3-7. Alternatively, the valve construction as detailed can be used with check valves as a dual valve in the embodiment of FIG.  1 . 
     In operation low pressure gas from a gas source  106  is admitted through intake conduct  108  by electro-hydraulic valve  102  under control of the control module  74  through electronic control line  110 . The gas is compressed on closure of the valve  102  by the piston  66  of the compressor  62 . At the cycle phase that the pressure in the diminishing compression chamber  111  reaches the pressure in the high pressure storage tank  80 , the valve  100  is opened under control of the control module  74  through line  112 , discharging the hot compression gases through outlet conduit  113  and intercooler  114  to storage tank  80  through conduit  116 . Part of the discharged gas to conduit  116  is diverted to a second cooler  118  through conduit  119 , which may advantageously be chilled by otherwise wasted cooling during expansion of gases at the dispenser during customer service. 
     After discharge of the high pressure gas and at the initiation of the expansion of the expansion stroke, the valve  100  under control of the control  74  is closed and electro-hydraulic valves  98  and  104  arranged on opposite sides of the compression chamber  111  are simultaneously opened scanvening the hot gases in the clearance volume remaining in the compression chamber  111 . The scavenged gases are discharged through conduit  120  through cooler  122  to the receiving storage tank  82 . Left in the clearance volume of the compression chamber  111  are the cooled gases from cooler  118 , further cooled by the expansion to the secondary pressure maintained in the storage tank  82 . As the expansion stroke of the piston  66  begins, electro-hydraulic valves  98  and  104  controlled by control module  74  through lines  124  and  126  are closed, allowing the pre-cooled trapped gases to expand to cryogenic levels (minus 250 degrees F.) to mix with the new charge on opening of the electro-hydraulic valve  102 . In this manner the mixture can be prechilled to a low temperature (approximately minus 120 degrees F.) before compression. 
     Since the charge of gas is prechilled before compression, the peak pressure can be well within design limits of the conventional materials used for high pressure compressors. Since the compressor  62  is operated on-site with the dispenser, the storage tanks  80  and  82  can be of minimal size with the dispenser monitored by the control module  74 . 
     A customer request input through a control panel  128  on the dispenser  84  is transmitted through input line  130  to the control module  74 . The control module  74  processes the entry which may be a pressure limit for the customer&#39;s bottle  88 , and operates the electronically controlled gas pressure regulator valves  76  and  78  to efficiently achieve the desired pressure. The dispenser  84 , may include the necessary flow meters to calculate the quantity of gas dispensed and the charge to the customer. 
     In order to instantaneously respond to the commands of the programmed control module, in the ultra high pressure environment of the compression chamber at peak pressure, at least the valves  98 ,  100 ,  104  have the modularized construction as shown in FIG. 3, where a typical electro-hydraulic valve unit  140  is shown. 
     The electro-hydraulic valve unit  140  is an assembly of five modules, a hydraulic connector block  142  for the main hydraulic activation lines; a central spool valve block  144 , detailed in FIG. 6; an actuator control block  146 , detailed in FIGS. 4 and 5; a spring return block  148  detailed in FIG. 7; and, the main valve block  150 . 
     As shown in FIG. 3, the hydraulic connector block  142  has a high pressure intake port  152  connecting a high pressure hydraulic feed conduit  154  to an internal passage  156  that communicates with an internal passage  158  in the coupled spool valve block  144 . The hydraulic connector block  142  also has a low pressure return port  160  connecting a low pressure return conduit  162  to an internal passage  164  that communicates with an internal passage  166  in the spool valve block  144 . 
     The spool valve block  144  has a displaceable spool valve  168  shown in a neutral position in the breakaway portion of the block  144  in FIG. 6, blocking both the hydraulic fluid delivery passage  158  and the return passage  166  to a common passage  170 . The common passage  170  communicate with a piston chamber  172  in the main valve block  150  when the spool valve block  144  and main valve block  150  are coupled as shown in FIG.  3 . 
     The main valve block  150  has an internal bushing  174  that guides a displaceable poppet piston  176  and contains a return spring  178  retained by a spring retainer  180  that biases a valve head  182  to a seated, closed position at the valve port  184  on the connector and  186  of the valve block  150 . The connector end  186  connects with the compressor  62  with the valve port  184  in communication with the compression chamber  111 . 
     Displacement of the poppet piston  176  by hydraulic fluid in the chamber  172  opens an internal gas passage  188  to the compression chamber for communicating ports  190  and  192  and gas conduits  194  and  196  to the compression chamber  111 . 
     Controlling the spool valve  168  and hence the hydraulic actuation and return of the valve head  182  is actuator control block  146  shown in FIGS. 4 and 5. The control block  146  has a connected solenoid actuator  198  that an electronic actuator by the control module  74  attracts a displaceable armature plate  200  connected to a plunger valve  202  biased to closure by a compression spring  204  retained between a stroke limiter  206  and cap plate  208 . The plunger valve  202  is guided by a bushing  210  having a valve seat  212  on which a valve shoulder  214  seats during closure, blocking a high pressure hydraulic conduit  216  connected to feed port  218 . Feed port  218  connects an internal passage  220  to a piston pusher  222  displaceable in a bushing  224  when the plunger valve  202  is electronically actuated unseating the valve shoulder  214  from the valve seat  212 . The displaceable piston pusher  222  is connected to the spool valve  168  in the assembly of FIG.  3 . 
     As shown in FIG. 5 the internal passage  220  to the piston pusher  222  has a relief passage  226  to a relief port  228  connected to a hydraulic fluid return conduit  229 . The relief passage  226  is blocked by a poppet valve  230  on actuation of a solenoid actuator  232  which attracts an armature plate  234  connected to a poppet valve  230  against the action of a spring  236  that on deactivation of the solenoid actuator  232  biases the valve  230  to an open position. 
     Referring to FIG. 7 the spring return block  148  has a bushing  238  for guiding a spring actuated pusher  239  that is connected to the opposite end of the spool valve  168  when the spring return block  148  is connected to the spool valve block  144  as shown in FIG.  3 . The spring actuated pusher  239  is connected to a spring retainer  240  which retains a compression spring  242  in a cavity  244  capped by end cap  246 . The modules  146  and  148  have various bleed passages  248 , such as those capped by set screws  250  in the spring return block and the end cap  252  in the actuator control block  146  shown in FIG.  4 . The bleed passages  248  return hydraulic fluid to the hydraulic return conduit  254  at the bleed line port  256  in the actuator control block  146 . 
     The dual solenoid actuators  198  and  232  are actuated when it is desired that high pressure hydraulic fluid pass from conduit  216  to piston pusher  222  to displace spool valve  168  against spring  242 . This allows high pressure hydraulic fluid from the conduit  154  to pressure chamber  172  displacing poppet piston  176  unseating valve head  182  allowing gas flow into or out of the compression chamber. 
     When deactivated, relief passage  226  is opened providing a sharp cut-off of the control fluid, allowing the return spring  242  to shuttle the spool valve  168  to a position that closes hydraulic feed passage  158 , opening return passage  166  and closing the poppet valve head  182  by action of the spring  178 . 
     While, in the foregoing, embodiments of the present invention have been set forth in considerable detail for the purposes of making a complete disclosure of the invention, it may be apparent to those of skill in the art that numerous changes may be made in such detail without departing from the spirit and principles of the invention.