|Publication number||US20030213246 A1|
|Application number||US 10/146,006|
|Publication date||Nov 20, 2003|
|Filing date||May 15, 2002|
|Priority date||May 15, 2002|
|Also published as||US20040226296|
|Publication number||10146006, 146006, US 2003/0213246 A1, US 2003/213246 A1, US 20030213246 A1, US 20030213246A1, US 2003213246 A1, US 2003213246A1, US-A1-20030213246, US-A1-2003213246, US2003/0213246A1, US2003/213246A1, US20030213246 A1, US20030213246A1, US2003213246 A1, US2003213246A1|
|Inventors||John Coll, William Thompson Hanna, Donald Anson, George Henry Stickford, John Edward Lambert, Jan Beryl Yates, Raymond Zaborski|
|Original Assignee||Coll John Gordon, William Thompson Hanna, Donald Anson, George Henry Stickford, John Edward Lambert, Jan Beryl Yates, Raymond Zaborski|
|Export Citation||BiBTeX, EndNote, RefMan|
|Patent Citations (5), Referenced by (68), Classifications (10), Legal Events (2)|
|External Links: USPTO, USPTO Assignment, Espacenet|
 The present invention relates to cogeneration systems and, more particularly, to microprocessor-based control sub-systems for controlling the thermal and electrical output of integrated micro-combined heat and power generation systems used to supply domestic electrical power, domestic space heating (SH) water, and domestic hot water (DHW).
 The concept of cogeneration, or combined heat and power (CHP), has been known for some time as a way to improve overall efficiency in energy production systems. With a typical CHP system, heat (usually in the form of hot air or water) and electricity are the two forms of energy that are generated. In such a system, the heat produced from a combustion process can drive an electric generator, as well as heat up water, often turning it into steam for dwelling or process heat. Most present-day CHP systems tend to be rather large, producing heat and power for either a vast number of consumers or large industrial concerns. Traditionally, the economies of scale have prevented such an approach from being extrapolated down to a single or discreet number of users. However, increases in fuel costs have diminished the benefits of centrally-generated power. Accordingly, there is a potentially great market where large numbers of relatively autonomous, distributed producers of heat and electricity can be utilized. For example, in older, existing heat transport infrastructure, where the presence of fluid-carrying pipes is pervasive, the inclusion of a system that can provide CHP would be especially promising, as no disturbance of the adjacent building structure to insert new piping is required. Similarly, a CHP system's inherent multifunction capability can reduce structural redundancy.
 The market for localized heat generation capability in Europe and the United Kingdom (UK), as well as certain parts of the United States, dictates that a single unit for single-family residential and small commercial sites provide heat for both SH (such as a hydronic system with radiator), and DHW (such as a shower head or faucet in a sink or bathtub), via demand or instantaneous system. Existing combination units are sometimes used, where heat for DHW is accumulated in a combination storage tank and boiler coil. In one configuration, SH water circulates through the boiler coil, which acts as the heating element for the water in the storage tank. By way of example, since the storage capacity required for instantaneous DHW supplying one to two showers in a single family residence (such as a detached house or a large apartment) is approximately 120 to 180 liters (roughly 30 to 50 gallons), it follows that the size of the storage tank needs to be fairly large, sometimes prohibitively so to satisfy thermal requirements of up to 25 kilowatts thermal (kWt) for stored hot water to meet such a peak shower demand. However, in newer and smaller homes there is often inadequate room to accommodate such storage tank volume. In addition to the need for instantaneous DHW capacity of up to 25 kWt, up to 10 kWt for SH is seasonally needed to heat an average-sized dwelling.
 Furthermore, even in systems that employ SH and DHW in a single heating system to consolidate spacing, no provision for CHP is included. In the example given above, it is likely that the electrical requirements concomitant with the use of 35 kWt will be between 3 and 5 kilowatts electric (kWe). The traditional approach to providing both forms of power, as previously discussed, was to have a large central electricity generating station provide electricity on a common grid to thousands or even millions of users, with heat and hot water production capacity provided at or near the end-user on an individual or small group basis. Thus, with the traditional approach, the consumer has not only little control over the cost of power generation, as such cost is subject to prevailing rates and demand from other consumers, but also pays more due to the inherent inefficiency of a system that does not exploit the synergism of using otherwise waste heat to provide either additional electric generation or heating capacity.
 Large-scale (in the megawatt (MW) range and up) cogeneration systems, while helpful in reducing the aforementioned inefficiencies of centrally-based power generation facilities, are not well-suited to providing small-scale (below a few hundred kW) heat and power, especially in the small-scale range of a few kWe and below (micro-based systems) to a few dozen kWe (mini-based systems). Much of this is due to the inability of the large prime mover systems to scale down, as reasonable electrical efficiency is often only achieved with varying load-responsive systems, tighter dimensional tolerances of key components, and attendant high capital cost. Representative of this class are gas turbines, which are expensive to build for small-scale applications and which sacrifice efficiency when operating over varying electrical load requirements. Efficiency-enhancing devices, such as recuperators, tend to reduce heat available to the DHW or SH loops, thus limiting their use in high heat-to-power ratio (hereinafter Q/P) applications.
 A subclass of the gas turbine-based prime mover is the microturbine, which includes a high-speed generator coupled to power electronics. This microturbine could be a feasible approach to small-scale cogeneration systems. Other shortcomings associated with large-scale CHP systems stem from life-limited configurations that have high maintenance costs. This class includes prime movers incorporating conventional internal combustion engines, where noise, exhaust emissions, lubricating oil and spark plug changes and related maintenance and packaging requirements render use within a residential or light commercial dwelling objectionable. This class of prime mover also does not reject a sufficient amount of heat for situations requiring a high Q/P, such as may be expected to be encountered in a single family dwelling. Other prime mover configurations, such as steam turbines, while generally conducive to high Q/P, are even less adapted to fluctuating electrical requirements than gas turbines. In addition, the steam-based approach typically involves slow system start-up, and high initial system cost, both militating against small-scale applications.
 In view of the limitations of the existing art, the inventors of the present invention have discovered that what is needed is an autonomous system that integrates electric and heat production into an affordable, compact, efficient and distributed power generator, and the ability to control the thermal and electrical output of such a system.
 The above mentioned needs are meet by providing microprocessor-based control sub-systems which control the thermal and electrical output of integrated micro-combined heat and power generation (M-CHP) systems used to supply domestic electrical power, domestic space heating (SH) water, and domestic hot water (DHW). The M-CHP system uses a microprocessor controller to control the internal operating conditions, such as pump speeds, gas flow rate, and evaporator outlet temperature. Controlling these parameters enables setting the capacity of the system at any instant in time, thereby permitting load following, using a variable capacity operation. The controller also monitors a number of additional safety controls and system protection devices, such as relays/contactor of the alternator to grid, and electrical trips to the feed pump, the oil pump, the hydronic pump, the blower, the gas valve, the expander bypass valves, and other electrically powered devices in the system.
 Overall, system capacity control emphasizes simplicity, reliability, and low cost. For the most part, the most basic control system is a single flow rate system with on/off control from a space thermostat, wherein the instantaneous capacity of a hydronic heating system of M-CHP system is fixed. In a more preferred system, working fluid flow rate and system thermal capacity are ultimately determined by the processor reading the outdoor ambient temperature. The outdoor ambient temperature is used to determine a set-point for the supply (hot-water) temperature in the hydronic heating system. The controller uses a look-up table or algorithm to establish the set-point according to a linear scale, such as varying the set-point linearly from 25° C. at an outdoor temperature of 20° C. to 75° C. at an outdoor temperature of −20° C. The controller uses this set-point to operate the M-CHP in a variable capacity mode by modulating the pump flow rate to maintain the actual hydronic supply temperature at the desired set-point.
 Additionally, the controller coordinates the burner fuel flow rate (heat input rate) with the feed pump flow rate to provide optimal thermodynamic performance by preventing liquid from entering the expander. This is accomplished by maintaining a fixed evaporator exit temperature, such as, for example, 310° F. Furthermore, the controller controls the hydronic pump of the heating system at a speed that keeps the pressure difference between the supply and return headers of the condenser at a preselected value for optimal thermodynamic performance of the hydronic heating system.
 In normal operation the M-CHP system capacity will vary as needed to maintain the supply temperature at its set-point, wherein the burner operates to provide the needed heat input. System mass flow varies essentially with pump speed, but since the expander operates at nearly a fixed speed, the volume flow of high-pressure vapor is essentially constant. Thus, the evaporator pressure varies with load, allowing the vapor density at the evaporator exit to vary to maintain a constant volume flow at each mass flow setting. Preferably, a variable frequency drive is used for the feed pump motor. However, a motor with discrete speed steps may also be used, as well as any other variable speed or flow approach. In normal operation the system operates at higher capacity to attain the higher hydronic supply temperatures required by colder weather. With the higher hydronic supply temperature, the capacity of the indoor radiators/convectors increases to handle the added heat flow. After M-CHP shut down, the hydronic pump may run on to deliver residual heat in the hydronic system to the radiators and maintain the flow of valid control data to the controller from the hydronic system.
 In one embodiment, coordinating the flow rates consists of open loop speed control of the feed pump with an induction motor. Proportional integral differential (PID) control of the boiler gas flow to maintain an outlet superheat between about 20° F. and about 30° F. using a closed loop control system may also be provided. A liquid condensate reservoir ahead of the feed pump and the liquid level is maintained above a predetermined level to ensure adequate net positive suction head (NPSH) to the pump. Preferably, no active control is necessary as a sufficient minimum level can be maintained under all operating conditions by filling the loop with a minimum refrigerant quantity for most operating conditions. However, if for certain operating conditions, the liquid level at the pump inlet cannot be maintained within a desired range, then an active means of maintaining liquid level control may be used. For example, level sensing at the liquid reservoir along with some means of speed control of the feed pump motor may be used.
 High and low side pressures float depending mainly on boiler heat input and hydronic loop return temperature to the condenser. Additionally, no active control of the power module speed is necessary. Connection of the generator to the grid loads the alternator sufficiently to limit the speed to about 3150 RPM (50 hz application). However, should the controller sense a power module overspeed (e.g., loss of grid connection, by sensing, for example, the rate of change of frequency), it will quickly command the opening of a bypass loop around the expander. Preferably, the sensor to detect expander overspeed is built into the power module.
 Typically, it is unnecessary for the controller to control the generator output as long as the generator output is less than the dwelling load. However, the controller can be programmed to monitor and deal with the situation when that generator output does exceeds dwelling load. In particular, the controller can be programmed to switch the extra power on the public grid or use the extra power for a resistance heater to preheat boiler feed.
 The following detailed description of the preferred embodiments of the present invention can be best understood when read in conjunction with the following drawings, where like structure is indicated with like reference numerals and in which:
FIG. 1 shows a schematic diagram of an integrated micro-CHP system including a control subsystem according to the present invention;
FIG. 2 shows a schematic diagram of a PLC control subsystem according to one aspect of the present invention;
FIG. 3 is a start-up/operating flow chart of the operating logic of one embodiment of the integrated micro-CHP system according to the present invention;
FIG. 4 is a shutdown flow chart of the operating logic of the integrated micro-CHP system according to one embodiment of the present invention;
FIG. 5 shows a schematic diagram of a sensor that may be used by a control subsystem of one embodiment of the present invention to sense the saturation pressure of an organic working fluid of an integrated micro-CHP system;
FIG. 6 is a chart showing measured outside ambient temperature versus desired hydronic set point; and
FIG. 7 is a chart showing net power output at various hydronic supply temperatures, which varies approximately linearly with mass flow.
 With reference to the figures, an embodiment of the invention is illustrated incorporated into a number of the major components of an illustrated organic rankine cycle (ORC) micro-cogeneration system used for the supply of domestic electrical power, domestic space heating (SH) water, and domestic hot water (DHW). Fluid supply lines connecting these components together are not drawn for ease of illustration. However, it is to be understood that the present control subsystems and control method of the present invention may advantageously be used with a number of micro-cogeneration systems. Exemplary micro-cogeneration systems are disclosed by co-owned U.S. patent application Ser. No. 09/998,705 filed Nov. 30, 2001, the entire disclosure of which is herein incorporated fully by reference.
 Referring initially to FIG. 1, a number of system components of an exemplary micro-CHP system 100 is schematically shown, such as an expander 101, a condenser (heat exchanger) 102, feed pump 103, heat source (boiler) 104, and evaporator 105. The expander 101 may be of any design, such as for example, a positive displacement expander and a scroll expander. In a direct fired system, the feed pump 103 circulates an organic working fluid (such as naturally-occurring hydrocarbons or halocarbon refrigerants, not shown) through a loop at least defined by the fluidly-connected expander 101, condenser 102, feed pump 103 and evaporator 105. In an indirectly fired system, the feed pump 103 circulates the working fluid through a first or inner loop, and at least defined by fluidly-connected expander 101, condenser 102, feed pump 103, and a first loop portion of evaporator (heat exchanger) 105, which is illustrated by dashed-line 105 a. In such an arrangement, the first loop portion or interloop portion 105 a receives heat from a second loop portion of the evaporator or heat pipe which is directly heated by heat source 104. More detailed information is provided in co-pending U.S. patent application Ser. No. 09/998,705 regarding directly and indirectly micro-CHP system arrangements, to which reference is made.
 Optionally, a desuperheater 106 may be included between the outlet of the expander 101 and the inlet of condenser/heat exchanger 102. In this alternative embodiment, a regulated by-pass valve 107 is used to circulate the working fluid exiting feed pump 103 through the coil loops (not shown) of the desuperheater 106 extracting heat energy from the vaporized working fluid exiting expander 101 before being circulating into evaporator 105 or interloop portion 105 a, if so configured. Since the size of the desuperheater is fixed, the amount of heat removed from the working fluid vapor exiting expander 101 will vary with working fluid flow rate and temperature as the system modulates. However, it is to be appreciated that controllable by-pass valve 107 permits the ability to control the flow of liquid working fluid through the desuperheater. By controlling the liquid working fluid flow, the amount of de-superheating is controllable, and therefore the heat-to-power ratio of the micro-CHP is controllable.
 When maximum thermal output is required, by-passing the desuperheater 106 will divert all the thermal energy to the hydronic fluid in condenser 102. When maximum electrical power output is desired, directing all the liquid refrigerant coolant first to the desuperheater 106 will remove the maximum amount of energy from the working fluid vapor in the desuperheater prior to it entering the condenser 102. By controlling the liquid working fluid flow rate between zero and 100 percent mass flow to the desuperheater 106, the thermal and electrical outputs can be tailored to better match the space heating load and the electrical power load of the house or building.
 A generator 108 (preferably induction type) is coupled to expander 101 such that motion imparted to it by expander 101 generates electricity. While the expander 101 can be any type, it is preferable that it be a scroll device. The scroll expander can be a conventional single scroll device, as is known in the art. The generator 108 is preferably an asynchronous device, thereby promoting simple, low-cost operation of the system 100, as complex generator speed controls and related grid interconnections are not required. An asynchronous generator always supplies maximum possible power without controls, as its torque requirement increases rapidly when generator 108 exceeds system frequency. The generator 108 can be designed to provide commercial frequency power, 50 or 60 Hz, while staying within close approximation (often 150 or fewer revolutions per minute (rpm)) of synchronous speed (3000 or 3600 rpm). The load on the expander 101 imposed by the grid ensures that mechanical speeds in the expander 101 are kept within its structural limits.
 Inherent in a micro-CHP (cogeneration) system is the ability to provide heat in addition to electricity. Excess heat, from both the heat source 104 and the expanded working fluid, can be transferred to external DHW and SH loops. The nature of the heat exchange process is preferably through either counterflow heat exchangers (e.g., for either or both the DHW and SH loops), or through a conventional hot water storage tank (e.g., for a DHW loop). In one embodiment, for example, a simple SH loop 109 may include fan 110 which provides for the drawing and blowing of air heated by heat source 104 to a space 111. For DHW services, an external heating loop 112 (shown partially) may be coupled to condenser 102. As an option, a preheat coil (not shown) can be inserted into the external heating loop 112 such that the hydronic fluid (typically water) flowing therethrough can receive an additional temperature increase by virtue of its heat exchange relationship with the heat exchange fluid flowing through a second circuit (not shown).
 The hydronic fluid flowing through external heating loop or DHW system 112 is circulated with a conventional hot water pump 113, and is supplied as space heat via a radiator or related device (not shown) to, for example, space 111. As an example, hydronic fluid could exit the condenser 102 at about 50° Celsius and return to it as low as 30° Celsius. It will be appreciated by those of ordinary skill in the art that while the embodiments depicted in the figures show separate DHW and SH heat exchangers, it is within the spirit of the present disclosure that series, parallel, sequential and/or the same heat exchange configurations could be used. Additionally, although the capacity of the system 100 is up to 60 kWt if desired, larger or smaller capacity units could be controlled by the herein disclosed control subsystem and method of the present invention.
 Combustion chamber 114, enclosing heat source 104, includes at least a flumed exhaust duct 115 to the outside of the building, and an exhaust gas fan 116. Heat at least to the evaporator 105 is provided by heat source or burner 104, which is supplied with fuel by a gas train 117 having a shut off valve 118 and variable flow gas valve 119.
 Other devices may be used with the exhaust duct 115, such as an exhaust gas recirculation device with an exhaust duct heat exchanger (not shown) which can be used to improve the thermal efficiency of the system 100 by lowering the temperature of the exhaust gas that is pulled away and vented to the atmosphere by fan 116. The heat given up by the exhaust gas in an exhaust gas heat exchanger may be used to provide additional heat to other parts of the system 100. More detailed information is provided by co-pending U.S. patent application Ser. No. 09/998,705 regarding exhaust gas recirculation devices with or without an exhaust duct heat exchangers, to which reference is made.
 Block valve 120 and bypass valve 121 are situated in the organic working fluid flow path defined by piping 122 (of which bypass conduit 123 is part). In response to a no load condition (such as a system start up/shut down or grid outage) on the system, the superheated vapor in piping 122 is permitted to bypass around expander 101 through bypass conduit 123, thereby avoiding overspeed of expander 101. In this condition, the rerouted superheated vapor is fed into the inlet of condenser 102 or if so configured, desuperheater 106. Under normal operating conditions, where there is a load on the system, the superheated vapor enters the expander 101, causing the orbiting involute spiral wrap to move relative to the intermeshed fixed involute spiral wrap. As the superheated vapor expands through the increasing volume crescent-shaped chambers of the expander 101, the motion it induces in the orbiting wrap is transferred to the generator 108, via a coupled shaft or an integral rotor/stator combination on the expander 101.
 Controller Sub-Systems
 As illustrated by FIG. 2, provided are controller sub-systems, generally indicated by 124, to control the thermal and electrical output of integrated micro-combined heat and power generation (M-CHP) systems, such as illustrated by FIG. 1. The controller sub-system 124 are microprocessor-based, wherein the M-CHP system 100 uses a microprocessor controller 125, which is preferably a programmable logic controller (PLC) or alternatively, any other conventional microcomputer, to monitor and control the operating conditions of system 100. It is to be appreciated that a PLC based subsystem simplifies setup and troubleshooting and may accommodate programming modifications from future developments and refinements to such micro-CHP systems. In a preferred embodiment, the controller is an Allen Bradley MicroLogix 1500 PLC with a 1764-LRP Processor and includes other components, such as for example, shown below in Table 1 for it operation, wherein the controller 125 is programmed using conventional programming software and techniques. In a preferred embodiment, the controller 125 ladder logic may be programmed with Allen Bradley RS Logix 500 software, wherein a graphical maintenance program running Allen Bradley RS View on a laptop PC can be used for troubleshooting and setup in the field.
 The controller subsystems 124 further include a plurality of process control sensors 126, a plurality of control outputs 127, a protection relay subsystem 128, a system trip reset 129, and the necessary user and service personnel interfaces, such a maintenance subsystem 130 and a data acquisition and communications subsystem 131. Since the user and service personnel interfaces are conventional, for brevity no further discussion is provided.
 For the purpose of controlling and monitoring the operation of cogeneration system 100, the plurality of process control sensors 126 and control outputs (I/O devices) 127 are individually called out in Table 2 and shown in FIG. 1, by sensors S0-S19, control points A1-A11, and protection relays M1-M4, for the system at various points therewithin. As illustrated in FIG. 2, the sensors and I/O devices shown in Table 2 are interfaced with the controller 125 for control of the micro-CHP system 100. Additionally, the controlled and monitored parameter of sensors S0-S19 and control points A1-A12 are also listed in Table 2, as well as the preferred interface and type for each.
TABLE 1 Component Mfg Part No Description Processor 1764LRP MicroLogix 1500 Processor Base 176424BWA MicroLogix 1500 Base 24 I/O Analog Voltage 17691F4 Analog Voltage Input Module Input Module Analog Voltage 17690F2 Analog Voltage Output Module Output Module Right end cap 1769ECR Right end cap terminator terminator Cable 1761CBLPM02 Cable Thermocouple 1769-IT6 1769-IT6 Thermocouple Input Module Input Module Real time clock 1764-RTC Real time clock Remote access MICRORAD Remote access modem modem
 The information obtained from sensors S1-S19 is used by the controller 125 to provide detailed system control through actuation and/or modulation of the various control points A1-A11. Such controlled and monitored parameters include pump speeds, gas flow rate, inlet and outlet temperatures and/or pressures to, from, and within the expander 101, condenser 102, evaporator 105, combustion chamber 114, desuperheater 106, and throughout various points of the space heat, domestic hot water, and generation loops 109, 112, and 122, respectively. Controlling and monitoring at least three of the above mentioned parameters (as will be discussed in later sections) enables setting the capacity of the system 100 at any instant in time, thereby permitting load following, using a variable capacity operation.
 All of the pumps and fans 103, 110, 113, and 116, are responsive to input signals from controller 125 via the their associated control points, namely A10, A8, A7, and A11, respectively. The controller 125 uses appropriate program logic, as will be explained in a later section, to control the turning on/off the fans, running and varying the speeds of the pumps, and also opening and closing valves, such as for example, by-pass valves 107, 121, shut off valves 118 and 120, and variable flow gas valve 119, in response to predetermined conditions. Such predetermined conditions include the demand for maximum heat or power output and an electric grid outage. Valves 107, 118, 119, 120, and 121 are responsive to input signals from controller 125 via the their associated control points, namely A4, A1, A6, A3, and A2, respectively.
 The controller 125 also monitors a number of additional safety controls and system protection devices, such as relays/contactor 132 (FIG. 1) of the alternator to grid having monitors M1-M4 listed in Table 2. Additionally, the controller 125 monitors and reports the electrical trips to the feed pump, an oil pump, the hydronic pump, the blower, the gas valve, the expander bypass valves, and any other electrically powered device in the system. Furthermore, the controller may monitor and activate a contactor of a crankcase heater for the power module via control point A9 and switch the output of the generator 108 to the power grid via control point A5.
 In addition to controlling and monitoring system operations, the controller 125 possesses functions to optimize the performance, efficiency, and safety the system 100. The functions of the controller 125 include automatic start up and shut down, modulation control of hydronic supply temperature, monitoring and tripping on safety and abnormal operating parameters, and electrical grid connection interface. With reference also to FIGS. 3-4, these functions of the controller 125 are explained in further detail.
FIG. 3 is a flow chart of the operating logic of the system 100. In operation, the controller 125 in one embodiment is programmed to operate the system in a quasi-steady state in response to a need for heat that is keyed to a specified hydronic supply temperature set point. This automatic operation is a heat load following mode. In other embodiments, such a call for heat, for example, may be from a thermostat, such as sensor S0, which demands heat when the space temperature falls below a user set-point, or according to an on-off timer, used for overnight shutdown. In the latter case, a timer would enable the thermostat to signal the controller 125 to initiate startup.
 In the heat load following mode, the call for heat occurs whenever the controller 125 determines that the temperature differential between the actual hydronic supply temperature of the DHW system 112, sensed via sensor S2, and a desired hydronic supply temperature is greater the a predetermined value, such as, for example 0.5 to 5° Celsius.
TABLE 2 Sensor-I/O Device Parameter PLC Interface Type S0 Space Temperature Digital Input Thermostat Or Analog Signal Thermometer S1 Hydronic Condenser Inlet Temperature Thermocouple Input Thermocouple S2 Hydronic Supply Temperature Thermocouple Input Thermocouple S3 System Trip Data Logger Error Signal S4 Expander Inlet Temperature Thermocouple Input Thermocouple S5 Expander Inlet Pressure Analog Input Pressure Transducer (0-500 Psi) S6 Feed Pump Inlet Temperature Thermocouple Input Thermocouple S7 Feed Pump Inlet Pressure Analog Input Pressure Transducer (0-200 Psi) S8 Power Module Power Output Digital Input Amp/Watt Meter S9 Condenser Internal Pressure Analog Input Pressure Transducer (0-200 Psi) S10 or 10′ Evaporator Coil Temperature Thermocouple Input Thermocouple S11 Gas Flow Digital Input Flow Rate Meter S12 Expander Outlet Temperature Thermocouple Input Thermocouple S13 Feed Pump Set-point Digital Input Flow Rate Meter S14 Heat Source High Limit Thermocouple Input Thermocouple S15 Flame Detection Digital Input Electro Optical S16 High Feed Level Digital Input Level Indicator S17 Low Feed Level Digital Input Level Indicator S18 Outside Temperature Digital Input Thermostat Or Analog Signal Thermometer S19 Hydronic Fluid Flow Digital Input Flow Rate Meter A1 Main Gas Valve Relay Output Solenoid Valve Actuator A2 Expander Bypass Relay Output Solenoid Valve Actuator A3 Expander Shutoff Relay Output Solenoid Valve Actuator A4 Desuperheater Bypass Relay Output Solenoid Valve Actuator A5 Power Module Run/Stop Relay Output Power Module Contactor A6 Evaporator Bumer Firing Rate Analog Output Gas Proportioning Valve Actuator A7 Hydronic Pump Run And Speed Relay Output Analog Output Pump Variable Frequency Drive A8 Space Heat Fan Run/Stop Relay Output Motor Contactor A9 Crank Case Heater On/Off Relay Output Crank Case Heater Contactor A10 VFD - Feed Pump Run And Speed Relay Output Analog Output Pump Variable Frequency Drive A11 Forced Draft Fan Run/Stop Relay Output Motor Contactor M1 Net Failure Digital Input Protection Relay M2 Over/Under Voltage Digital Input Protection Relay M3 Over/Under Frequency Digital Input Protection Relay M4 Power Module Internal Temperature Digital Input Digital Thermometer
 It is to be appreciated that the hydronic pump 113 operates continuously so there is always a flow through the DHW system 112 enabling the controller 125 to continuously monitor the actual supply temperature and return temperatures, via sensors S2 and S1, respectively. In other embodiments, the controller 125 may use the internal temperature of the condenser 102 and hydronic flow rate, via speed control of hydronic feed pump 113, to determine a hydronic supply temperature using predetermined heat exchange values and/or algorithms.
 The desired hydronic supply temperature is set by the controller 125 sensing the outdoor ambient temperature, via thermostat S18, and correlating the sensed outdoor ambient temperature to a desired hydronic supply temperature for the boiler. The correlation between outdoor ambient temperature and desired hydronic supply temperature may be, for example, according to the illustrated linear relationship shown by FIG. 6. In this illustrative embodiment, for example, on cold days, say −20° C. ambient, the hydronic set point is 75° C., but on warm days, when only a little heat is needed, i.e., 20° C. ambient, the set point is 25° C., and at a 0° C. ambient the set point is 50° C. However, for other embodiments other scalar relationships between outdoor ambient temperature and the desired hydronic supply temperature may be used such as, for example, logarithmic, exponential, and other nonlinear functions. To avoid the influence of sunshine on cold days, a single measuring point on a north facing side of a building or home should be used for thermostat S18.
 In response to the call for heat in step 200, the controller 125 is programmed to modulate the thermal output of the evaporator 105 in order to have the actual hydronic supply temperature match the desired hydronic set point for the given outdoor ambient temperature. Accordingly, the controller 125 in step 202 determines whether the evaporator 105 is already at its operating temperature via thermocouple sensor S10 or S10, (FIG. 1). Assuming the temperature of the evaporator 105 is below the minimum operating temperature (e.g., cold start-up), the controller 125 verifies, and if necessary closes the expander shutoff valve 120 and opens the expander bypass valve 123 as shown in steps 204 and 206, respectively. Once verified, the controller 125 checks to see if the burner is on via flame detector S15 and if not, purges the combustion chamber 114 with force draft blower 116 and activates the burner 104, as shown at blocks 208 and 210.
 The burner 104 is activated by controller 125 opening the gas valve 118 and metering its flow with flow rate valve 119, via actuator A1, to nominally between 40 and 80% of full flow when ignition is expected. An igniter (not shown) is turned off by a timer and the gas valve is turned off when no flame is proven via flame detector S15. The burner 104 will then typically be set to come on to about 50% of its capacity to warm up system 100. If desired, an off-the-shelf combustion controller can initiate burner firing upon receiving an enabling signal from the controller 125. If so arranged, the combustion controller will provide flame and combustion air detection, light the burner, and provide high temperature limit protection for the evaporator 105. The controller also drives a variable speed draft fan to provide the approximately correct air flow rate for the burner.
 Next, the controller 125 checks to see if the evaporator working fluid level is in the desired operating range, via level sensors S16 and S17 in step 212 and if not, in step 214 will activate feed pump 103. It is to be appreciated that both burner 104 firing and feed pump 103 flow may be controlled in part, and conventionally by room temperature and its user determined set-point, as well as outdoor temperature, via sensor S18. Additionally, feed pump 103 comes on to a speed predetermined by the controller 125 to coincide with the flow requirements established by the initial burner firing rate and the design response of the system 100. In particular, the controller 125 runs feed pump 103 fast enough to keep the organic working fluid liquid level between level low sensor S17 and level high sensor S16. Accordingly, the feed pump is turned on by the controller 125 at the appropriate time to keep the evaporator filled, but not over-filled with working fluid.
 When the system is operating, superheated working fluid is moving past sensor S4, which is able to provide a valid signal to the controller 125 so the heat source or burner 104 firing rate and feed pump 103 flow can be adjusted for both the safe operation and needed output. However, when the system is just starting, the controller 125 must be given some initialized state which can be used as a safe operating condition until such time as working fluid is flowing past temperature sensor S4.
 It is desirable to have a minimum amount of working fluid flow during startup, so that the fluid heats up as rapidly as possible. However, some flow is needed to prevent local overheating of the fluid in evaporator 105, and to provide the controller 125 with an indication that the burner 104 is indeed firing. Accordingly, the burner gas rate is set to provide the longest possible run time for the system, consistent with measured outdoor temperature and rate of change of indoor temperature. Feed pump 103 operates to keep the evaporator 105 supplied with the working fluid at the factory-preset value for temperature sensor S4. When temperature sensor S4 gets to about 50% of the thermostat set-point, feed pump 103 speed is increased until the temperature reading in temperature sensor S4 reaches its set-point, at which time the burner 104 modulates for constant values of at least temperature sensor S4, and the feed pump speed is modulated to maintain the desired hydronic supply temperature.
 The controller will abort the start up sequence of steps 202-214, if burner 104 fails to heat the evaporator 105 or if any of the parameters listed in Table 3 are exceeded. If the startup sequence fails, the controller 125 will attempt to re-start again for a maximum of three re-start attempts. If three re-start attempts are completed with no successful re-start, the controller 125 will be locked out from re-starting until manually reset, via trip reset 129 (FIG. 2). The re-start attempt counter will reset after a successful start or a manual reset.
 After the startup sequence is complete the system 100 will go into run mode, wherein working fluid pressure is allowed to build and feed pump 103, burner 104, and evaporator 105 are controlled by controller 125 via two separate feed back loops.
 At the appropriate time, the expander 101 is connected to receive the now superheated vapors of the working fluid, wherein the controller 125 in step 216 opens the cutoff or block valve 120 to expander 101 and after a short delay (i.e., 1 second), closes the bypass valve 121. It is to be appreciated that liquid is prevented from entering into the expander by maintaining the evaporator's working fluid exit temperature at a fixed value for all operating conditions. As mentioned previously, in one embodiment, the evaporator exit temperature is set to operate at about 154° C. (310° F.), which has been found to give good overall system efficiency regardless of system load.
TABLE 3 Parameter Threshold Fault Response Expander inlet temperature 340° F. (171.1° C.) Full shut down Expander inlet pressure 420 psia (29.5 Kg/cm) Full shut down Feed pump inlet 200° F. (93.3° C.) Full shut down temperature Feed pump inlet pressure 200 psia (14 Kg/cm) Full shut down Protection relay trip Suitable trip values Power module shut down only Power module temperature 300° F. (148.9° C.) Power module shut switch down only Startup sequence failure 3 attempts Full shut down
 Further protection to the scroll expander is provided during start-up and while running by controlling the displacement of the feed pump, which in one embodiment is by feed pump speed, wherein the rate of change in feed pump speed is limited by an output ramp in the controller logic which provides time for the evaporator PID control to modulate the evaporator burner for steadier operation and to prevent over temperature of the expander inlet. In this control arrangement, a negative feed forward value will be added to the feed pump speed control variable when ramping down to better control temperature of the expander inlet.
 The system 100 will warm up and quickly come to a near-steady-state operating point for the feed pump flow setting. After a short start-up delay (i.e., 1 second) in step 218 wherein the controller verifies that the output of the generator is within set parameters via sensor S8, the electrical output is then connected to the grid in step 220 via contactor 132 responding to control signal A5 from the controller 125. It is to be appreciated that there is no need to control the generator output as long as the generator output is less than the dwelling load. Preferably, connection of the generator 108 to the grid loads the alternator sufficiently to limit the speed to about 3150 RPM (50 hz application).
 However, should the controller 125 sense a power module overspeed (e.g., loss of grid connection), it will quickly command the opening of a bypass loop 123 around the expander 101. Preferably, the sensor to detect expander overspeed is built into the power module. Furthermore, the controller 125 can be programmed with at least two options for dealing with the case when the generator output exceeds the dwelling load. The controller 125 can put the extra power on the grid or the extra power can be used for a resistance heater at condenser 102.
 Additionally, it is to be appreciated that the power module 108 will trip on over/under voltage, over/under frequency and loss of mains conditions detected by the protection relay 128 (FIG. 2). To prevent nuisance tripping of the heating system, only the power module 108 will trip on voltage and frequency grid disturbances. The controller 125 will log the fault and re-start the power module 108 after the over/under voltage or over/under frequency condition has dissipated. If a loss of mains occurs, the entire system 100 will shut down with all latches/relays reset to start ready condition. The controller 125 will then automatically re-start on the restoration of power if heat is called for by the thermostat/timer conditions.
 In step 222, the controller checks to see if the call heat has been satisfied via continuously reading sensor S0 and/or S18. If the hydronic supply temperature is not at the desired step point as checked in step 224, the controller 125 in step 226 reads the outdoor ambient temperature via sensor S18 and uses a look-up table or algorithm to establish the desired set-point according to the linear relationship illustrated by FIG. 6. The controller 125 will then use the continuously updated desire set-point to operate the evaporator 105 in a variable capacity mode by modulating the gas valve 119 on the burner 104 to maintain the actual hydronic supply temperature, sensed via sensor S2, at the desired set-point.
 In step 228, the controller 125 checks to see whether the expander inlet temperature via sensor S4 is in a desired operating range. If not, then in step 230 the controller 125 modulates the burner fuel flow rate (heat input rate) to provide optimal thermodynamic performance and to prevent liquid from entering the expander. Additionally, in this quasi-steady state the controller 125 may control the hydronic pump 113 of the heating system at a speed that maintains the pressure difference between the supply and return headers of the condenser at a preselected value for optimal thermodynamic performance of the hydronic heating system.
 With the controller subsystems 124 operating the system 100 in the above described heat load following mode, it is to be appreciated that the system will operate for as many hours as possible during the heating season. The controller subsystems 124 will run the system 100 just hard enough to maintain the hydronic supply temperature at the correct value for the nominal heating load. When the system 100 operates at less than the maximum supply temperature, more power is generated than at the maximum temperature, the controller 125 can automatically and passively maximize the power which can be made and sold.
 Further it is to be appreciated that in the above described heat load following mode, the greater the temperature differential error, the larger the thermal output response that will be initiated by controller 125 in order to achieve the desired hydronic supply temperature. As illustrated in FIG. 7, since the thermal output of the system varies approximately linearly with the mass flow of the system 100, the controller 125 controls its thermal response by controlling both the displacement of the feed pump 103, via a PID speed control, and the heat input to the evaporator 105, via adjusting the gas flow rate to the burner 114 to maintain a 310° F. (154.4° C.) temperature into the expander.
 In particular, the controller 125 is programmed to increase or decrease the mass flow rate in proportion to the system heating capacity required to match the actual and the desired set point for the hydronic supply temperature. For example, the desired set point of the hydronic supply temperature is 50° C. for a given sensed outdoor ambient temperature of 0° C., if while operating in a quasi-steady state, the actual hydronic supply temperature suddenly drops to 45° C. (e.g., a door or window of the space or building is open to the outdoors), the controller 125 will increase the percentage of system mass flow to meet this heating demand. Should in this example after a predetermined time period the temperature error between the actual and desired hydronic supply temperatures decrease, the controller 125 will further increase the percentage of system mass flow in order to meet this heating demand. Accordingly, it is to be appreciated that in normal operation the system 100 must operate at higher capacity to attain the higher hydronic supply temperatures required by colder weather and with the higher hydronic supply temperature the capacity of the indoor radiators/convectors increases to handle the added heat flow.
 In run mode, the controller 125 monitors the operating parameters shown in Table 1 and performs a full or partial shut down if any of the parameters exceed preset thresholds. An individual trip event on the automatic trip settings is defined as a Level 1 trip. In the event of a Level 1 trip, the controller 125 will self reset and automatically attempt to re-start. If the re-start fails, the controller 125 will attempt to re-start again for a maximum of three re-start attempts. If the three re-start attempts are completed with no successful re-start, this case is defined as a Level 2 trip. In the event of a Level 2 trip, an alarm notification message will be sent by controller 125, via the data acquisition and communications subsystem 131 and the controller 125 will be locked out from re-starting until manually reset, via trip reset 129 (FIG. 2). The re-start attempt counter will reset after a successful start or a manual reset.
 Normal shutdown of the micro-CHP will occur if the outdoor thermostat signal indicates that the outside temperature has gone above the temperature that corresponds to the lowest hydronic supply temperature set-point and remains for 30 minutes. The system will re-start when the thermostat signal indicates that the outside temperature is below the normal startup temperature. The controller 125 will also perform the normal shutdown sequence when switched off manually by the user or service personnel.
 In the former case, when the system load falls below about 30 to about 40% of a full load and 30 minutes has expired, the controller 125 is programmed to shutdown the system 100 and cease making both heat and power to ensure economical use of the m-CHP system. Since the hydronic pump is kept running at all times, even at a low flow rate, the controller 125 continuously monitors the error signal between the hydronic actual and set point values. When this error is large enough, (i.e. the actual temperature is below the set point by a preselected value) the controller starts the system for another on-cycle. Should the controller also sense that during operation of the system at the minimum system mass flow, the actual supply temperature begins to exceed the set point, it is programmed also to shutdown the system when this error exceeds a predetermined value. In either of these conditions of heat load following mode, or upon receiving a heat satisfied message from a space thermostat, if not configured in the heat load following mode, the controller 125 will follow a normal shutdown procedure according to the program control logic illustrated by FIG. 4.
 In step 300, normal shutdown begins with the controller 125 turning off burner 103. After a time delay in step 303 (i.e., 15 seconds), the controller 125 will shutdown (disconnect) the power module 108, open the bypass valve 121 and close the shutoff valve 120 to the expander 101 in step 304. After another time delay (i.e., 5 seconds) in step 306, the controller 125 stops the feed pump 103 in step 308. After another time delay (i.e., 60 seconds) in step 310, the controller 125 may slow the hydronic pump 113 to its minimum speed, completing a normal shutdown of the system 100.
 However, in step 300 the time delay may be conditioned on available reservoir heat energy in heating system 112, wherein if desired, the hydronic pump 113 may run on to deliver residual heat in the hydronic system to the radiators until below a certain set-point. In such a case, the controller 125 will initiate a partial shutdown if the hydronic supply temperature exceeded the set-point by 5° C. for a period of 30 minutes when the hydronic set-point is at the minimum setting. The evaporator 105, feed pump 103, expander 101, and power module 108 will shutdown while the hydronic pump will continues to run. When the hydronic supply temperature error signal reaches −0.5 to 5° C., the system will re-start.
 Data will be logged using the logging capability of the controller 125. Preferably, a remote access modem will interface with the controller 125 to download any system data. The modem may be self-dialing and use either land line or cellular service. In the event of a Level 2 trip, the controller may use the modem to send an alarm notification message to service personnel. Data points logged by controller 125 are shown in Table 4.
 Evaporator Heat Control
 In normal operation, no control of the evaporator heat rate is required as the ideal heat input rate is set by a given mass flow rate out of the feed pump. If the heat input rate is greater than the ideal rate, then the evaporator outlet fluid will be more superheated than desired. This will lead to increasing evaporator pressure until the density at the expander inlet is sufficient to provide a match between expander and feed pump mass flow rate. Thus, the results of excessive heat input rate will be excessive evaporator pressure and evaporator outlet superheat.
 If there is too little evaporator heat input for a given feed pump flow rate, evaporator pressure will be reduced, and some liquid may be admitted to the expander 101. Liquid admission to the expander 101 is not expected to result in damage to the expander because of the increasing chamber volume at all times. However, partial liquid admission is likely to result in reduced pressure ratio across the expander 101 and, therefore, less thermodynamic work by the expander. Accordingly, to obtain a better match between the evaporator heat input rate and the feed pump flow rate over all operating conditions, better control over the evaporator heat output is desired.
 The control concept involves sensing the level of superheat at the evaporator outlet and maintaining this level in the range of 20° F. to 30° F. (−6° C. to −1° C.). This will tend to minimize the evaporator pressure while maintaining some margin from liquid entering the expander 101. The evaporator pressure will then float until the density at the expander inlet is such that the expander and feed pump flow rates match.
TABLE 4 Channel Parameter Signal Type 1 Hydronic condenser inlet Thermocouple temperature 2 Hydronic condenser outlet Thermocouple temperature 3 Hydronic fluid flow Analog 4 Expander inlet temperature Thermocouple 5 Expander inlet pressure Analog 6 Feed pump inlet temperature Thermocouple 7 Feed pump inlet pressure Analog 8 Power module power output Analog 9 Evaporator inlet temperature Thermocouple 10 Gas flow Analog 11 Expander outlet temperature Thermocouple 12 Feed pump speed (drive frequency) Analog 13 Protection relay trip Digital 14 Level 1 trip failure Digital 15 Level 2 trip failure Digital (failed to restart)
 The superheat exiting the evaporator 105 may be sensed by either measuring the evaporator outlet temperature and saturation temperature, measuring the evaporator outlet temperature and outlet pressure, or sensing the saturation pressure. The first approach requires measuring the temperatures of the working fluid in its saturated state at the evaporator inlet with inlet temperature sensor S10 (indirectly fired) or S10, (directly fired) and also in its vapor state such as with expander inlet temperature sensor S4. The difference between the temperatures is then the superheat. For an improved superheat sensing, in addition to the temperature, expander inlet pressure sensor S5 may be used to sense the evaporator outlet pressure. The controller 101 may then use a look-up table and/or calculation to determine the saturation temperature from the pressure and temperature readings, wherein the superheat is the difference between the measured temperature and the computed saturation temperature. Furthermore, pressure sensor S5 can also provide a safety feature to protect both the evaporator 105 and the expander 101 from potentially harmful overpressure.
 The third approach is illustrated by FIG. 5, wherein the evaporator outlet pressure is applied to one side of a diaphragm 500 of a superheat sensing device 501. The other side of the diaphragm sees the same working fluid, but from a sensing bulb 502 at the same temperature as the evaporator outlet gas. The pressure of this latter side is the saturation pressure corresponding to the evaporator outlet temperature. If there is positive superheat, the pressure from the sensing bulb 502 will exceed the evaporator outlet pressure and the diaphragm 500 will compress a spring 504 by an amount proportional to the degree of superheat. A potentiometer (or other position sensor) 506 attached to the diaphragm 500 then provides an electrical output to the controller 125 proportional to the degree of superheat. The controller 125 can then use the output sign from the superheat sensing device 501 to minimize the evaporator pressure while maintaining some margin from liquid entering the expander 101.
 Although some aspects of the present invention are identified herein as preferred or particularly advantageous, it is contemplated that the present invention is not necessarily limited to these preferred aspects of the invention. For example, an alternative mode of providing heat following control may be as follows. As before, this alternative heat following mode uses an outdoor temperature sensor (i.e., sensor S18) to determine the set point of the hydronic supply temperature. Controller 125 after sensing the actual hydronic supply temperature (i.e., via sensor S2), adjusts the burner firing rate, via the gas valve 118 and/or 119, via actuator A1, to bring the actual hydronic supply temperature into line with the desired set point. If the actual hydronic supply temperature is too low, the controller 125 will increase the firing rate, and if the temperature is too low, the controller 125 will decrease the firing rate.
 Further, in this alternative heat following mode, to maintain the evaporator outlet temperature of the superheated working fluid vapors within a desired operating parameter, such as, for example 310° F. (154.4° C.), the controller 125 will adjust the feed pump flow rate and/or speed of the organic working fluid past the burner in order to control the evaporator exit temperature of the superheated vapor. If the exit temperature of the superheated working fluid vapors from the evaporator is above its desire operating parameter, the controller 125 will increase the working fluid's flow rate. Should the exit temperature of the superheated working fluid vapors from the evaporator be less than its desired operating parameter, the controller 125 will decrease the working fluid's flow rate.
 Having described the invention in detail and by reference to preferred embodiments thereof, it will be apparent that modifications and variations are possible without departing from the scope of the invention defined in the appended claims.
|Cited Patent||Filing date||Publication date||Applicant||Title|
|US2151733||May 4, 1936||Mar 28, 1939||American Box Board Co||Container|
|CH283612A *||Title not available|
|FR1392029A *||Title not available|
|FR2166276A1 *||Title not available|
|GB533718A||Title not available|
|Citing Patent||Filing date||Publication date||Applicant||Title|
|US7055328 *||Nov 21, 2003||Jun 6, 2006||Siemens Aktiengesellschaft||Method and device for regulating the power output of a combined-cycle power station|
|US7067933 *||Nov 10, 2003||Jun 27, 2006||Terry Edgar Bassett||Waste oil electrical generation system|
|US7260934 *||Apr 5, 2006||Aug 28, 2007||John Hamlin Roberts||External combustion engine|
|US7279800||Jan 9, 2006||Oct 9, 2007||Bassett Terry E||Waste oil electrical generation systems|
|US7290393||May 6, 2004||Nov 6, 2007||Utc Power Corporation||Method for synchronizing an induction generator of an ORC plant to a grid|
|US7454912 *||Dec 22, 2005||Nov 25, 2008||Denso Corporation||Device for utilizing waste heat from heat engine|
|US7469540 *||Aug 30, 2005||Dec 30, 2008||Brent William Knapton||Energy recovery from waste heat sources|
|US7848853 *||Mar 25, 2009||Dec 7, 2010||Solarlogic, Llc||System and method for controlling hydronic systems having multiple sources and multiple loads|
|US7950230 *||Sep 9, 2008||May 31, 2011||Denso Corporation||Waste heat recovery apparatus|
|US8041462 *||Mar 22, 2011||Oct 18, 2011||Solarlogic, Llc||System and method for controlling hydronic systems having multiple sources and multiple loads|
|US8093733 *||May 20, 2009||Jan 10, 2012||Honda Motor Co., Ltd.||Cogeneration system|
|US8126595 *||Nov 10, 2010||Feb 28, 2012||Solarlogic, Llc||System and method for controlling hydronic systems having multiple sources and multiple loads|
|US8132409||May 7, 2008||Mar 13, 2012||Solar Turbine Group, International||Solar collection and conversion system and methods and apparatus for control thereof|
|US8344528||Jun 29, 2010||Jan 1, 2013||Terry Edgar Bassett||Waste oil electrical generation systems|
|US8407998||Dec 1, 2010||Apr 2, 2013||Cummins Inc.||Waste heat recovery system with constant power output|
|US8418466||Dec 10, 2010||Apr 16, 2013||David Hardgrave||Thermodynamic amplifier cycle system and method|
|US8544274 *||Jul 23, 2009||Oct 1, 2013||Cummins Intellectual Properties, Inc.||Energy recovery system using an organic rankine cycle|
|US8554377 *||Nov 12, 2010||Oct 8, 2013||Terrafore, Inc.||Thermal energy storage system comprising optimal thermocline management|
|US8577507 *||Oct 17, 2011||Nov 5, 2013||Solarlogic, Llc||System and method for controlling hydronic systems having multiple sources and multiple loads|
|US8590307 *||Feb 25, 2010||Nov 26, 2013||General Electric Company||Auto optimizing control system for organic rankine cycle plants|
|US8613195 *||Oct 21, 2011||Dec 24, 2013||Echogen Power Systems, Llc||Heat engine and heat to electricity systems and methods with working fluid mass management control|
|US8627663||Sep 2, 2009||Jan 14, 2014||Cummins Intellectual Properties, Inc.||Energy recovery system and method using an organic rankine cycle with condenser pressure regulation|
|US8635871||Jan 31, 2013||Jan 28, 2014||Cummins Inc.||Waste heat recovery system with constant power output|
|US8646272 *||May 27, 2010||Feb 11, 2014||Korea Institute Of Energy Research||Rankine cycle system and method of controlling the same|
|US8656720||May 12, 2011||Feb 25, 2014||William David Hardgrave||Extended range organic Rankine cycle|
|US8674525 *||Jul 9, 2008||Mar 18, 2014||Universiteit Gent||Combined heat power system|
|US8683801||Aug 13, 2011||Apr 1, 2014||Cummins Intellectual Properties, Inc.||Rankine cycle condenser pressure control using an energy conversion device bypass valve|
|US8707914||Feb 28, 2012||Apr 29, 2014||Cummins Intellectual Property, Inc.||Engine having integrated waste heat recovery|
|US8752378||Aug 9, 2011||Jun 17, 2014||Cummins Intellectual Properties, Inc.||Waste heat recovery system for recapturing energy after engine aftertreatment systems|
|US8752381 *||Apr 22, 2010||Jun 17, 2014||Ormat Technologies Inc.||Organic motive fluid based waste heat recovery system|
|US8776517||Aug 5, 2011||Jul 15, 2014||Cummins Intellectual Properties, Inc.||Emissions-critical charge cooling using an organic rankine cycle|
|US8794001 *||May 15, 2012||Aug 5, 2014||Kobe Steel, Ltd.||Power generation apparatus|
|US8800285||Jan 5, 2012||Aug 12, 2014||Cummins Intellectual Property, Inc.||Rankine cycle waste heat recovery system|
|US8813498 *||Jun 18, 2010||Aug 26, 2014||General Electric Company||Turbine inlet condition controlled organic rankine cycle|
|US8826662||Dec 22, 2011||Sep 9, 2014||Cummins Intellectual Property, Inc.||Rankine cycle system and method|
|US8844287||Mar 25, 2013||Sep 30, 2014||William David Hardgrave||Thermodynamic amplifier cycle system and method|
|US8893495||Jul 16, 2012||Nov 25, 2014||Cummins Intellectual Property, Inc.||Reversible waste heat recovery system and method|
|US8919328||Jan 20, 2012||Dec 30, 2014||Cummins Intellectual Property, Inc.||Rankine cycle waste heat recovery system and method with improved EGR temperature control|
|US9003798 *||Mar 15, 2012||Apr 14, 2015||Cyclect Electrical Engineering Pte.||Organic rankine cycle system|
|US9021808||Jan 10, 2012||May 5, 2015||Cummins Intellectual Property, Inc.||Rankine cycle waste heat recovery system|
|US9062898||Oct 3, 2012||Jun 23, 2015||Echogen Power Systems, Llc||Carbon dioxide refrigeration cycle|
|US9091278||Aug 19, 2013||Jul 28, 2015||Echogen Power Systems, Llc||Supercritical working fluid circuit with a turbo pump and a start pump in series configuration|
|US20040093864 *||Nov 10, 2003||May 20, 2004||Bassett Terry Edgar||Waste oil electrical generation system|
|US20040129002 *||Nov 21, 2003||Jul 8, 2004||Manfred Huengerle||Method and device for regulating the power output of a combined-cycle power station|
|US20050247059 *||May 6, 2004||Nov 10, 2005||United Technologies Corporation||Method for synchronizing an induction generator of an ORC plant to a grid|
|US20100194111 *||Jul 9, 2008||Aug 5, 2010||Van Den Bossche Alex||combined heat power system|
|US20100307155 *||Feb 12, 2009||Dec 9, 2010||Junichiro Kasuya||Waste Heat Utilization Device for Internal Combustion Engine|
|US20110016863 *||Jul 23, 2009||Jan 27, 2011||Cummins Intellectual Properties, Inc.||Energy recovery system using an organic rankine cycle|
|US20110048404 *||Feb 2, 2009||Mar 3, 2011||Faith Louise Limited||Heating system|
|US20110079012 *||Apr 7, 2011||Young Jin Baik||Rankine cycle system and method of controlling the same|
|US20110083437 *||Apr 14, 2011||General Electric Company||Rankine cycle system|
|US20110120127 *||Jul 3, 2009||May 26, 2011||Uhde Gmbh||Low energy process for the production of ammonia or methanol|
|US20110203278 *||Feb 25, 2010||Aug 25, 2011||General Electric Company||Auto optimizing control system for organic rankine cycle plants|
|US20110259010 *||Apr 22, 2010||Oct 27, 2011||Ormat Technologies Inc.||Organic motive fluid based waste heat recovery system|
|US20110308252 *||Dec 22, 2011||General Electric Company||Turbine inlet condition controlled organic rankine cycle|
|US20120047892 *||Oct 21, 2011||Mar 1, 2012||Echogen Power Systems, Llc||Heat Engine and Heat to Electricity Systems and Methods with Working Fluid Mass Management Control|
|US20120090321 *||Nov 9, 2011||Apr 19, 2012||Gaertner Jan||Exhaust gas heat utilization in motor vehicles|
|US20120096857 *||Apr 26, 2012||Gaertner Jan||Heat utilization device an operating method|
|US20120118554 *||Nov 12, 2010||May 17, 2012||Terrafore, Inc.||Thermal Energy Storage System Comprising Optimal Thermocline Management|
|US20120271466 *||Oct 25, 2012||Solarlogic, Llc||System and method for controlling hydronic systems having multiple sources and multiple loads|
|US20120312021 *||May 15, 2012||Dec 13, 2012||Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.)||Power generation apparatus|
|US20140224469 *||Feb 11, 2013||Aug 14, 2014||Access Energy Llc||Controlling heat source fluid for thermal cycles|
|US20140373542 *||Mar 12, 2012||Dec 25, 2014||Cyclect Electrical Engineering||Organic rankine cycle system|
|WO2005108749A1 *||May 6, 2005||Nov 17, 2005||United Technologies Corp||A method for synchronizing an induction generator of an orc plant to a grid|
|WO2010054724A2 *||Oct 6, 2009||May 20, 2010||Daimler Ag||Clausius-rankine cycle|
|WO2012059752A2 *||Nov 2, 2011||May 10, 2012||Energetix Genlec Limited||Heating system - modular|
|WO2012085093A1 *||Dec 21, 2011||Jun 28, 2012||Robert Bosch Gmbh||Waste heat recovery installation|
|WO2014078838A2 *||Nov 19, 2013||May 22, 2014||Heat Assured Systems, Llc||System and methods for controlling a supply of electric energy|
|U.S. Classification||60/653, 60/670|
|International Classification||F01K25/08, F01K17/02|
|Cooperative Classification||F01K25/08, F24H2240/00, Y02E20/14, F01K17/02|
|European Classification||F01K25/08, F01K17/02|
|Jul 15, 2002||AS||Assignment|
Owner name: INERGIA, INC., OHIO
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNORS:COLL, JOHN (JACK);HANNA, WILLIAM THOMPSON;ANSON, DONALD;AND OTHERS;REEL/FRAME:013252/0393;SIGNING DATES FROM 20020517 TO 20020529
Owner name: INERGIA, INC., OHIO
Free format text: INVALID RECORDING;ASSIGNORS:COLL, JOHN(JACK) GORDON;HANNA, WILLIAM THOMPSON;ANSON, DONALD;AND OTHERS;REEL/FRAME:013080/0073;SIGNING DATES FROM 20020517 TO 20020528
|Jan 21, 2003||AS||Assignment|
Owner name: BATTELLE MEMORIAL INSTITUTE, OHIO
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNOR:INERGIA INCORPORATED;REEL/FRAME:013670/0717
Effective date: 20030107