Abstract:
A system and method for controlling the temperature of a process tool uses the vaporizable characteristic of a refrigerant that is provided in direct heat exchange relation with the process tool. Pressurized refrigerant is provided as both condensed liquid and in gaseous state. The condensed liquid is expanded to a vaporous mix, and the gaseous refrigerant is added to reach a target temperature determined by its pressure. Temperature corrections can thus be made very rapidly by gas pressure adjustments. The process tool and the operating parameters will usually require that the returning refrigerant be conditioned and processed for compatibility with the compressor and other units, so that cycling can be continuous regardless of thermal demands and changes.

Description:
REFERENCE TO PRIOR APPLICATIONS 
   This invention relies for priority on Provisional Patent Application No. 60/546,059 filed Feb. 19, 2004, entitled “Transfer Direct of Saturated Fluid System”, and Provisional Application No. 60/576,705 filed Jun. 2, 2004, entitled “Transfer Direct Heat Exchanger System”, both naming Kenneth W. Cowans, Glenn Zubillaga and William W. Cowans as inventors. 

   BACKGROUND OF THE INVENTION 
   Thermal control units (TCUs), such as heating and chilling systems are widely used to establish and maintain a process tool or other device at a selected and variable temperature. Typical examples of a modem thermal or temperature control unit are found in highly capital intensive semiconductor fabrication facilities. Stringent spatial requirements are placed on the TCUs, in order to preserve expensive floor space as much as possible. Reliability must be assured, because the large capital equipment costs required do not tolerate downtime in operation if profitable performance is to be obtained. The target temperature may be changed for different fabrication steps, but must be held closely until that particular step is completed. In many industrial and common household refrigeration systems the purpose is to lower the temperature to a selected level, and then maintain the temperature within a temperature range that is not highly precise. Thus even though reliable and long-lived operation is achieved in these commercial systems, the performance is not up to the demands of highly technical production machinery. 
   In most modem TCUs actual temperature control of the tool or process is exercised by use of an intermediate thermal transfer fluid which is circulated from the TCU through the equipment and back again in a closed cycle. A thermal transfer fluid is selected that is stable in a desired operating range below its boiling temperatures at the minimum operating pressure of said fluid. It also must have suitable viscosity and flow characteristics within its operating range. The TCU itself employs a refrigerant, usually now of an ecologically acceptable type, to provide any cooling needed to maintain the selected temperature. The TCU may circulate the refrigerant through a conventional liquid/vapor phase cycle. In such cycles, the refrigerant is first compressed to a hot gas at high pressure level, then condensed to a pressurized liquid. The gas is transformed to a liquid in a condenser by being passed in close thermal contact with a cooling fluid; it is either liquid cooled by the surrounding fluid or directly by environmental air. The liquid refrigerant is then lowered in temperature by expansion through a valve to a selected pressure level. This expansion cools the refrigerant by evaporating some of the liquid, thereby forcing the liquid to equilibrate at the lower saturation pressure. After this expansive chilling, the refrigerant is passed into heat exchange relation with the thermal transfer fluid to cool said thermal transfer fluid, in order to maintain the subject equipment at the target temperature level. Then the refrigerant is returned in vapor phase to the pressurization stage. A source of heating must usually be supplied to the thermal transfer fluid if it is needed to raise the temperature of the circulated thermal transfer fluid as needed. This is most often an electrical heater placed in heat exchange with the circulated fluid and provided with power as required. 
   Such TCUs have been and are being very widely used with many variants, and developments in the art have lowered costs and improved reliability for mass applications. In mass produced refrigerators, for example, tens of thousands of hours of operation are expected, and at relatively little cost for maintenance. However, such refrigeration systems are seldom capable of operating across a wide temperature range, and lower cost versions often use air flow as a direct heat exchange medium for the refrigerated contents. 
   In contrast, the modern TCU for industrial applications has to operate precisely, is a typical requirement being ±&lt;1° C., at a selected temperature level, and shift to a different level within a wide range (e.g. −40° C. to +60° C. for a characteristic installation). Typical thermal transfer fluids for such applications include a mixture of ethylene glycol and water (most often in deionized form) or a proprietary perfluorinated fluid sold under the trademark “Galden” or “Fluorinert”. These fluids and others have found wide use in these highly reliable, variable temperature systems. They do not, however, have high thermal transfer efficiencies, particularly the perfluorinated fluids, and impose some design demands on the TCUs. For example, energy and space are needed for a pumping system for circulating the thermal transfer fluid through heat exchangers (HEXs) and the controlled tool or other equipment. Along with these energy loss factors, there are energy losses in heat exchange due to the temperature difference needed to transfer heat and also losses encountered in the conduits coupling the TCU to and from the controlled equipment. Because space immediately surrounding the device to be cooled often at is a premium, substantial lengths of conduit may be required, which not only introduces energy losses but also increases the time required to stabilize the temperature of the process tool. In general the larger the volume of the TCU the farther the TCU needs to be located remotely from the device to be controlled. The fluid masses along the flow paths require time as well as energy to compensate for the losses they introduce. Any change in temperature of the device to be controlled must also affect the conduits connecting the TCU and the controlled device along with the thermal transfer fluid contained in said conduits. This is because the thermal transfer fluid is in intimate thermal contact with the conduit walls. Thus, the fluid emerging at the conduit end nearest the controlled device arrives at said device at a temperature substantially equal to that of the conduit walls and these walls must be changed in temperature before the controlled device can undergo a like change in temperature. 
   Under the continuing demand for improved systems and results, there is a need for a TCU which minimizes these losses. If possible, the system should also be compact, of low capital cost, and preserve or even increase the long life and reliable characteristics which have become expected. 
   To the extent that straightforward refrigeration systems may have hitherto employed a refrigerant without a separate thermal transfer fluid, it has been considered that the phase changes imposed during the refrigeration cycle prohibit direct use of the refrigerant at a physical distance outside the cycle. A conventional refrigerant inherently relies on phase changes for energy storage and conversion, so that there must also be a proper state or mix of liquid and vapor phases at each point in the refrigeration cycle for stable and reliable operation of the compressor and other components. Using a saturable fluid such as a refrigerant directly in heat exchange with a variable thermal load presents formidable system problems. 
   The present application teaches for the first time a system which directly employs the high thermal transfer efficiency of a refrigerant mixture of liquid and vapor in a highly efficient system capable of very fast temperature change response. It eliminates the need for substantial delay times to correct temperature levels at the device being controlled, as well as for substantial energy losses in conduits and HEXs, and the need for substantial time delays in shifting between target temperatures at different levels. 
   SUMMARY OF THE INVENTION 
   Systems and methods in accordance with the invention employ a variable phase refrigerant directly as a cooling or heating source throughout a wide temperature range and with high speed response and high thermal efficiency. The refrigerant is maintained as a saturated mix of liquid and vapor during the principal part of its thermal control range and in direct contact with a controlled unit functioning as a variable heat load. The temperature of controlled equipment can be adjusted very rapidly by variation of the pressure of the saturated fluid mix. The energy losses in conduits, HEXs and fluid masses are minimized and the delay in temperature response of the cooled device due to the change in temperature of these components is substantially eliminated. 
   Systems and methods in accordance with the invention, in more specific examples, compress a cycling refrigerant to a high temperature, high pressure state, but provide proportional control of a hot gas flow, as well as a separate flow of a condensed liquid/vapor mist. The liquid/vapor mist initially comprises an expanded flow of condensed refrigerant, but is combined with a proportioned flow of hot gas, determined by a controller, in accordance with a chosen set point for the controlled device. To this end the two flows are brought together in a mixing circuit, at which the saturated fluid is brought to a target temperature and pressure and a pressure drop is introduced in the expanded flow to compensate for flow nonlinearities inherent in the expansion valve device. The saturated fluid itself is then transported directly through the controlled process or equipment. The temperature of the controlled process or equipment is sensed and sent to the controller, which can vary the temperature of the controlled system rapidly simply by pressure change. By thus changing the temperature of the medium effecting the cooling or heating, such change in temperature can be made available to the controlled device nearly immediately following the pressure change. This eliminates many thermal energy losses and temperature changes arising from use of a separate thermal transfer fluid in contrast with the controlled device. 
   The invention herein disclosed thus effectively can apply cooling or heating to a controlled device rapidly enough so as to counteract the effects of a change in power applied to the controlled device and thereby keep the controlled device at an invariant temperature. 
   In moving the refrigerant through a complete continuous cycle for ultimate direct heat transfer, a number of novel expedients are utilized to assure that the phases of the refrigerant are stable throughout. At the compression step, for example, a balance of input temperature and pressure is maintained at the compressor by employing a desuperheater valve responsive to the compressor input temperature, and a feed-through loop with an electrical heater and heat exchange system is incorporated so as to assure that the input flow at the compressor input is raised to the proper range if necessary. This balance also assures that refrigerant returned to the compressor input is free of liquid as well as in a selected pressure range. In addition, input pressure to the compressor is limited by a close-on-rise valve in the return flow path from the controlled process. 
   The path for flow of condensed refrigerant includes an externally stabilized conventional refrigeration thermostatic expansion valve (TXV), while the hot gas bypass flow path to the mixing circuit includes a proportional (or proportioning) valve. The proportional valve is responsive to control signals from the controller system, which commands the proportions of flow to be such as to achieve the desired pressure and temperature of the delivered mixture. 
   The system can also heat outside the mixing range by utilizing only hot gas at the upper end of the temperature range. When a high control temperature is needed that is attainable by using hot pressurized gas only, the proportional valve is opened more fully and the thermal expansion valve is shut down by action of a spring-loaded check valve with a predetermined pressure relief load on the check valve&#39;s spring. The refrigerant may alternatively be heated externally to raise the temperature even more. In this latter case a counter-current HEX can also be employed to further extend the heating range upward in temperature in an efficient manner. 
   The system is arranged to enable the control of a unit across a range of temperatures in not only the mixed fluid and hot gas modes, but also in a chilling mode using only thermal expansion of pressurized ambient refrigerant. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     A better understanding of the invention may be had by reference to the following description, taken in conjunction with the accompanying drawings, in which: 
       FIG. 1  is a block diagram of a temperature control unit in accordance with the invention; 
       FIG. 2  is a block diagram of an alternate temperature control unit in accordance with the invention using a different method of introducing electrical heat to the system; 
       FIG. 3  is a flow chart of steps followed in practicing methods in accordance with the invention; 
       FIG. 4  is a graphical chart of variations in pressure vs. enthalpy during an energy transfer cycle in the system and method showing a cycle effective at −20° C.; 
       FIG. 5  is a graphical chart of variations in pressure vs. enthalpy during an energy transfer cycle in the system and method showing a heating cycle effective at over 120° C.; 
       FIG. 6  is a graphical chart of variations in pressure vs. enthalpy during an energy transfer cycle in the system and method, showing a cycle effective at +40° C.; 
       FIG. 7  is a block diagram of details of a system feature for use in heating the output of the TCU above 120° C. employing an auxiliary electric heater and a counter current HEX in a system such as  FIG. 1 , and 
       FIG. 8  is a graphical chart of variations in pressure vs. enthalpy during an energy transfer cycle in the system and method, showing a cycle effective in heating at +40° C. using a heat pump capability of the unit. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   A block diagram of a temperature control unit (TCU)  110  is depicted in  FIG. 1  for operation principally in the range of approximately −50° C. to +140° C., by way of example only. Other temperature ranges may be utilized, depending upon the refrigerant and to some extent the load, but the example given assumes the use of refrigerant R507 as an example. TCU  110  can be a compact unit and is characterized by low cost as well as moderate size, enhanced economy and rapid response. Temperature levels are to be held stable at different target levels irrespective of the lengths of the lines coupling associated devices. The TCU  110  in this example is intended for the purpose of controlling the temperature of a tool  112 , such as a cluster tool for semiconductor fabrication. Such tools have internal passageways for passage of a thermal control fluid. The TCU is intended to establish different target temperatures of the tool for operating cycles during different fabrication steps. 
   The system incorporates a controller  114 , such as a proportional, integral, differential (PID) controller of the type described in U.S. Pat. No. 6,783,080 of Antoniou and Christofferson, which is suitable for receiving a number of different types of commands and includes a user-friendly setup system. In the TCU  110 , a compressor  158  is employed which may be a highly reliable yet low-cost commercial refrigeration compressor providing a pressurized output of hot gas refrigerant at approximately 120° C. at 400 psi or more at the output line  102 . The temperature at the tool  112  is sensed by a transducer  118  located at the tool  112  and a measurement signal is returned to the controller  114 . This temperature signal is used in the controller  114  for different purposes. For example, it can control both the opening of a controllable proportional valve  144  which supplies hot gas directly from the compressor  58  output, and the flow of saturated fluid after liquefaction of the hot compressor output in condenser  156 , so as to provide a mix of liquid and gas at a desired temperature to the controlled device  112 . 
   For these purposes, the hot gas flow from the compressor  158  branches into two flow paths, one of which enters a compressor control system  120  including a conventional condenser  156  including a heat exchanger (HEX)  104  that is liquid cooled by a facility water source  154 . An air cooled condenser could equally well be employed, and liquid cooling is chosen as an example only. Water is supplied to HEX  104  in condenser  156  through either a controlled water valve  106  responsive to the output pressure of compressor  158  or a controllable bypass valve  105  that is responsive to the controller  114 . Bypass valve  105  is activated whenever a maximum cooling effort is needed. Opening valve  105  assures that the condenser  156  is supplied with the coldest water possible. This provides the system with the maximum cooling output by assuring that the condensing temperature is as low as possible. The output pressure is measured by a transducer which is contained in the coolant flow controller valve  106 , which is a commercially available unit called a compressor head pressure regulator. This is conventionally applied to refrigeration systems used in applications in which the supply of cooling water may be too cold or too abundant for one significant reason or another. One typical application would use such a coolant flow controller to limit the supply of cooling water for reasons of economy or efficiency. In this invention the controller  106  is used for this purpose as well, but controller  106  primarily functions so as to maintain the output of the compressor  158  at a high pressure level for most operational modes. This high pressure is required for the compressor to be available as a strong source of heat. 
   An auxiliary benefit to the use of coolant flow controller  106  is so that the presently disclosed system can be a very efficient user of cooling water. This water is typically supplied in semiconductor fabrication plants from a source refrigerated by a cooling tower or other approach. The power needed to run such cooling source is a significant part of the total power used by the fabrication installation. The supply of cooling water from the source  154  to the condenser HEX  104  is varied inversely in accordance with compressor  158  output pressure so as to maintain a substantially constant compressor output pressure. The compressor control system  120  also includes an interaction with a countercurrent subcooler  130 . When such subcooler is used, said interaction includes the injection of the output from a desuperheater valve  134  into the outgoing path of said subcooler combining the output of valve  134  with refrigerant gas being returned from the tool  112 , thereby cooling said outgoing return flow in said subcooler  130 . This incoming opposite flow into an incorporated subcooler (which is optional for some applications) is directed into expansion and control circuits, described below. The incoming flow to control the temperature of the tool through subcooler  130  is completed to the return flow input side of the subcooler  130  via the desuperheater valve  134 . This arrangement and its purposes are in accordance with U.S. Pat. No. 6,446,446 by William W. Cowans. 
   Also, a hot gas bypass valve (HGBV)  164  is placed between the compressor output and the compressor input. The HGBV allows flow to pass directly from the compressor output to its input if the input pressure falls below a preset level. The HGBV is a standard commercial refrigeration control component. The pressure at the input to the compressor  158  cannot be allowed to fall below a certain level, which level is determined by the compressor design. This is because refrigeration compressors are lubricated by oil carried mixed in the refrigerant. At some low pressure the carryover of oil is inadequate to lubricate the compressor machinery. Refrigeration compressors are also limited in the compression ratio that can be experienced without damage occurring. This occurs due to the adiabatic heating undergone by the gas as it is compressed. At discharge gas temperatures over around 120° C. refrigeration compressors can give trouble. The HGBV  164  alleviates this problem. 
   The mechanism described above includes some standard approaches to compressor management in commercial refrigeration equipment but include unique approaches to the invention discussed herein, as shown in the section describing operation of the system. 
   The fluid in the liquid line  132  from the subcooler  130  is paralleled by the separate hot gas flow in hot gas line  159 , and both lead to a mixing circuit  140 . The hot gas flow in line  159  traverses a proportional valve  144 , which valve is controlled by controller  114  signals which assure selected reduction in pressure in the hot gas flow provided into the mixing circuit  140 . The valve  144  varies the mass flow, which ultimately varies the pressure. A separate input provided to the mixing circuit  140  from the vapor/liquid line  132  is controlled via a thermal expansion valve (TXV)  157 . This operates as a normal refrigeration valve of the thermostatic expansion type. TXVs are diaphragm operated valves, one side of which diaphragm is maintained at the refrigerant pressure at a suitable point in the low pressure refrigeration circuit which the other side is at the saturation pressure of the temperature at substantially that same pressure point. A sensing bulb  124  placed at the latter point in the circuit is filled with the refrigerant gas and thus exists at a saturation pressure corresponding to the point at which the bulb is mounted to supply this saturation pressure. In the TCU circuit shown in  FIG. 1  conduit  149  communicates with output line  161  at a location proximate to the bulb  124  and thus equalizes the pressure to that pressure in the low pressure level proximate to bulb  124 . This is called external equalization. 
   If proportional valve  144  were to be fully closed, the TCU circuit shown in  FIG. 1  would function as a normal vapor cycle refrigeration system. In this normal operation the TXV regulates the refrigeration output so as to produce the maximum refrigeration at which the system is capable. The action of the diaphragm-regulated TXV  157  throttles the flow of high pressure refrigerant liquid through the line in such manner as to supply the maximum amount of expanded liquid-vapor mix that can be boiled completely to pure vapor. In the principal operating mode, however, TXV  157  supplies a selected proportion of misted liquid vapor for combination with the hot gas from valve  144  when valve  144  is not fully closed. As stated above, the TXV  157  is externally equalized by the pressure communicated via the conduit  149  with the return line from the tool  112 . The TXV  157  output flows through a delta P valve  155 , which comprises a spring-loaded check valve establishing a fluid pressure drop (delta p) between the output of the TXV  157  and the mixing Tee  165 . The total pressure across the delta P valve  155  is greater than the pressure drop across a fully open proportional valve  144  in the hot gas line when all the output of the compressor  159  is diverted to flow only across proportional valve  144 . This establishes smooth control of the flow mixing from 100% hot gas to 100% expanded liquid, and overcomes the non-linear characteristics of the TXV and the fact there is always a pressure drop across the proportional valve  144  no matter how far it is opened. If the hot gas flow is full open, the check valve closes off the TXV. The output from the TXV  157  and the delta P valve  155  is therefore a saturated fluid whose temperature is essentially determined by the pressure at the output of the delta P valve  155 . The pressure can be varied rapidly by changing the setting of the proportional valve  144 , which changes the mass flow and thus the pressure. Thus the temperature can almost instantaneously be adjusted to correct the temperature of the tool  112 , as measured by a temperature sensor  118  responsive to the tool temperature and signaling the controller  114 . 
   The system also includes a “Close on Rise” (COR) valve  150  in the return line from the tool  112  to act as a safeguard against excessive pressure buildup in the pressure input at the compressor  158 . This is a commercially available refrigeration component and is traditionally used for this purpose. In the subject invention it serves the same purpose but also allows the TCU to act as a heat pump as will be explained below. 
   Solenoid valve  121  is shown in the hot gas line  159  leading to the proportioning valve  144 . Valve  121 , which has a rapid response time, is included because in some systems it is desirable that the flow of hot gas be interrupted instantaneously to achieve cooling without the delay that might be incurred in the process of closing the proportioning valve  144 . There are also some requirements for TCU systems to control loads which need to be heated instantaneously as well. To accommodate these, a solenoid valve  122  can also be used to shunt the operation of proportional valve  144 . To aid in the operation of those systems in which heating needs to be applied suddenly another solenoid valve  109  can be included in the line to the TXV  157  for the purpose of shutting flow through TXV  157  substantially instantaneously. For systems needing instantaneous cooling another solenoid valve  111  can be included to shunt the operation of TXV  157 . 
   A receiver  108  is shown in  FIG. 1 . This is a relatively small reservoir for refrigerant and is needed in some systems that have a requirement to hoard cooling potential while the process of heating proceeds apace. A receiver is a device that takes the liquid output of the condenser  156  and stores the condensed liquid if an amount of such liquid is produced in excess of that used by the TXV. 
   Downstream of the outputs of the proportional valve  144  and the TXV  157  in the mixing circuit  140  the two streams of refrigerant are combined at the mixing Tee  165 . After such mixing has occurred the output flow travels through supply line  113  to cool or heat the tool  112 . After leaving tool  112  the mix of vapor and liquid returns to the TCU through return line  160 . 
   The first processing or conditioning of returning refrigerant that occurs in the TCU is electrical heating. This is driven by heater  117 . In  FIG. 1  it is shown as immersed in the liquid within a heated accumulator  116 : This is one embodiment of the invention. In an alternative version shown in  FIG. 2  the heater is immersed in a HEX  216 , placed in good thermal exchange relation with the refrigerant passing through HEX  216 . The difference between heated accumulator  116  and HEX  216  is that the accumulator has capacity for a significant amount of liquid storage and the HEX has only capacity for that amount of refrigerant necessary to carry out the heat transfer function. 
   As the refrigerant passes out of either accumulator  116  ( FIG. 1 ) or HEX  216  ( FIG. 2 ) it passes through return line  161  to which line are attached sensor bulb  124  and equalization line  149 . Return line  161  thence couples with the return passage of subcooler  130 . Emerging from subcooler  130  the refrigerant passes into suction line  162  through which the refrigerant returns to the suction input  163  of compressor  158 . 
   Operation of the System 
   A counter-intuitive refrigeration cycle has thus been disclosed, which focuses on maintaining a transitional phase of saturated fluid (misted liquid and vapor) in a heat exchange relation with a system whose temperature is to be controlled, as shown sequentially in the flow chart of  FIG. 3 . The use of the saturated phase together with appropriate internal manipulation enables a refrigerant fluid and cycle to be employed directly for temperature control, while phase change and stability barriers not previously surmounted are overcome. By establishing liquid droplets and vapor mist in equilibrium at a selected pressure, the temperature is predetermined. Moreover, the capacity for thermal energy interchange is substantially higher than in a pure liquid or pure gaseous phase, because the dynamics of evaporation and liquefaction enhance the ability to transfer heat to a surface, as opposed to the strictly heat conductive effects existing in both the pure liquid and pure gas phases. 
   A temperature change with a fluid in the pure gas phase and a temperature change in the purely liquid phase are both dependent solely upon thermal energy conduction. In the intermediate region, between these pure mono-phase states a mixed liquid/vapor exists. Transport of vapor into and out of the liquid droplets can be viewed as strictly dependent on pressure or temperature, with the lower the pressure the lower the temperature of evaporation. From an equilibrium temperature, however, heat is supplied to a cooling source until all of the vapor is liquefied, or heat is taken up in evaporation, at a substantially constant temperature, until the entire mass is evaporated or condensed. This means that a liquid/vapor mix can be used as a constant temperature sink or source and, contrary-wise, that by varying the pressure, the temperature of a unit in thermal exchange relation with the liquid/vapor mix can be varied. It is significant that this variation can be extremely rapid because of the fact that pressure changes are transported through a fluid at the speed of sound; hundreds of meters per second. 
   Referring to  FIG. 1  the crucial mixing zone comprises the elements within mixing system  140  which includes the hot gas output from the proportional valve  144  and the output of liquid and vapor from TXV  157 , both of which branch from the compressor  158  output line  102 . Using an output pressure of 400 psi, by way of example, the liquefied output from the cooled condenser  156  to the TXV  157  will be at a substantially like pressure. After expansion at the TXV  157 , as commanded by controller  114 , the TXV  157  provides a misted liquid flow. This can be viewed classically as a dispersion of droplets within a surrounding atmosphere of liquid vapor. The heat exchange characteristics of this misted liquid are in accordance with the equation set out by McAdams, W. H. in the book “ Heat Transmission ”, Third Edition, McGraw-Hill Book Company, New York, 1954, p. 335 &amp; 402. Combination of the misted liquid with a controller-determined hot gas flow also incoming at the mixing head  165  results in diminution by a controlled amount of the size of the droplets brought about by the need to equilibrate the temperature within the total mix of liquid and vapor from TXV  157  with the hot gas from proportional valve  144 . This process of mixing hot gas from  144  with liquid/vapor from TXV  157  thus can supply a controlled temperature and output pressure of refrigerant at the input to controlled tool  112 . Mixing circuit  140  further includes the delta p valve  155 , which introduces a pressure drop substantially no greater than the inherent drop in the proportional valve  144 , when said proportional valve  144  is wide open. Furthermore, the mixing head  165  and delta p valve  155  prevent back-flow of the mix into the liquid/vapor line  132  when valve  144  is wide open. 
   A typical refrigeration circuit (with subcooler) is shown as operating in the classical thermodynamic cycle  401  to  402  to  403  to  404  to  405  and back to  401  in  FIG. 4 . By plotting pressure against enthalpy in the circuit in this manner, one can see that the compressor  158  of  FIG. 1  drives the pressure upward and also drives the enthalpy higher, giving the line  401  to  402  a slope showing increases in both amplitudes. Condensation of the compressed gas lowers the enthalpy, while maintaining the pressure, as shown by the constant pressure line  402 – 403 . This shift moves the refrigerant through the liquid dome shown on the PH chart, causing liquefaction of the refrigerant while maintaining the pressure. The evaporation point of the refrigerant is about 45° C. at 400 psi. In the classical refrigeration cycle, the pressure is dropped to a selected level, without changing the enthalpy, as the refrigerant is expanded, as shown from points  404 – 405 . The expanded refrigerant, released as liquid/vapor mixture, moves through the liquid dome transition in the line from  405 – 401 , and is directed through the heat exchange area. The gas is recompressed following point  401  and the cycle is repeated. 
   The present invention modifies the basic refrigeration cycle to accomplish the objectives of a modern TCU with more flexibility. The Mollier diagram (a display of enthalpy versus temperature in the vicinity of the liquid dome) of refrigerant (type R 507) shown in  FIG. 4  shows the operation of the refrigerant in providing a flow of liquid and vapor at −20° C., which temperature is chosen as an example. The invention provides for a variation in the heating or cooling capabilities of the fluid under rapid control of the unit. The refrigeration cycle is shown from point  401  which is taken at the compressor input  163  ( FIG. 1 ). The gas is compressed to point  402 , which point is about 30 KPa (ca. 400 psig) at a temperature of about 120° C. Gas that enters the condenser  156  is cooled and liquefied to point  403  at a temperature of around 60° C. This liquid is passed through subcooler  130 . In this component the liquid is cooled by exchanging heat with the refrigerant returning from the tool  112  in line  161 . Liquid refrigerant thus cooled in subcooler  130  is then expanded through TXV  157  to point  404 . At this point the refrigerant is at a temperature of around −20° C. and consists of about 50% gas and 50% liquid in the current example. This is mixed with hot gas expanded through proportional valve  144  and depicted on  FIG. 4  by the dotted line path from point  402  to point  406  which, in the present example would be at a temperature of about 85° C. The addition of heat from gas at 85° C. mixing with the liquid/vapor at point  405  results in a total mix at point  407 . This controlled mix is about 70% gas and 30% liquid. The addition of the hot gas has boiled off the difference of 50% liquid at point  405  and the hot gas added has been cooled to −20° C. In the example given the mixture boils off liquid in cooling the tool  112  and further heats as it gains heat from the surrounding environment to point  408 . This gas then enters subcooler  130  and is heated close to ambient temperature by absorbing heat from the counterflowing liquid refrigerant being cooled from  403  to  404 , then drops in pressure and increases in enthalpy to point  401  wherein the cycle is repeated. 
   In consequence, as one can deduce from a study of  FIG. 4 , there is a range of operation in which the liquid/vapor mix, dependent upon the pressures maintained, stabilizes the temperature of the tool  112 . If the tool is giving up heat to the fluid, and is to maintain a given temperature T, shown as −20° C. in  FIG. 4  as an example, as determined at the tool  112  by the sensor  118 , pressure is adjusted in the flow of vapor/liquid in supply line  113  by adjusting the opening of valve  144  to change the mass flow rate. This alters the temperature accordingly in line  113  as vapor and liquid equilibrate at the adjusted saturation temperature in order to hold the temperature of the tool constant. In cases of extreme heating the flow from the TXV  157  can be shut off entirely by fully opening the proportional valve  144 . In this case the entire flow though the tool  112  is derived (see  FIG. 4 ) from the flow of gas at point  406 . This gas is at a temperature around 80° C. and thus can heat tool  112  rapidly. 
   The system is further stabilized by the external equalization feedback path from a pressure bulb  124  at the tool  112  output. As is known with thermal expansion valves, transmission of the pressure return to TXV  157  from the pressure line  149  helps to assure that there is no offset because of any pressure losses in the lines or in the tool  112 . 
   The invention can be used to provide heat at an elevated temperature outside the bounds limited by the liquid dome.  FIG. 5  shows the operation of the invention in this mode, in which the system operates both inside and outside the zone of liquefaction in adjusting thermal energy. The operation depends on the addition of a heater at the output of the mixing circuit  140 .  FIG. 7  shows alternatives to the basic system presented in  FIG. 1  that are employed to incorporate this ability. In the supply line  113  to the tool  112  downstream of mixing circuit  140  an electrical heater  702  is placed in good thermal contact with the supply line  113 . A counter current HEX  701  is also placed in line  113  and additionally intercepts return line  160  to receive the outgoing flow from tool  112 . The use of a counter current HEX to isolate the temperature of line  113  on the side of HEX  701  that is closest to the tool from the line  160  into the other side of HEX  701  allows the attainment of even higher temperatures. With this feature refrigerant gas at temperatures as high as 260° C. or even higher can be supplied to tool  112 . 
     FIG. 5  shows the thermodynamic performance of the TCU fitted with a subsystem such as in  FIG. 7 . Refrigerant gas enters the compressor at point  507 . It is then compressed to about 30 KPa at point  501  where the gas enters HEX  701 . In the countercurrent HEX  701  the input gas is heated to point  502  in absorbing heat from the outgoing gas as it is cooled from point  508  to point  505 . The electrical heater  702  then heats the input gas from point  502  to point  503  which is the temperature at which the gas enters tool  112 , assuming there is negligible loss in gas temperature as it passes from heater  702  to tool  112  through line  113 . The gas is cooled in the process of heating tool  112  from point  504  to point  508 . At point  508  the gas enters HEX  701  and cools to point  505  in heating the input gas on the other side of HEX  701 . The gas then passes through COR valve  150  and drops to a pressure at point  506  suitable for the compressor  158  input. The gas would be ready to restart its cycle and be compressed again except it is too hot for successful operation of the compressor. In the system of  FIG. 1 , however, the hot gas mixes with the output of desuperheater valve  134  which is opened in response to the sensor  126  at the compressor  158  input. This action adds a fraction of condensed refrigerant at the return side of the subcooler  130 . The combination of the condensed liquid fraction from the condenser  156 , which is at point  511  in  FIG. 5  with the returning hot gas (lowered in pressure to point  510 ), provides the input gas at a temperature appropriate for compressing at point  507 . The system therefore operates, after compression and condensation, in a hot gas mode outside the thermodynamic liquid dome demarcated in pressure and enthalpy parameters. 
   The operation of COR valve  150  can come into play at lower temperatures under particular circumstances. If the TCU is called into operation at a temperature significantly over 10° C. the pressure at which liquid and gas equilibrate in the refrigerant will be too high for successful compression in conventional compressors. Referring to  FIGS. 1 and 5 , COR valve  150  protects the compressor  158  when, as is shown in  FIG. 6 , the TCU is being called on to cool a load at 40° C. (in a manner similar to that shown in  FIG. 4  wherein the tool was cooled with refrigerant at −20° C.).  FIG. 6  shows the gas being compressed from point  601  to point  602 . Some of this gas then is condensed at the same pressure to point  603  and expanded through TXV valve  157  to point  604 . The remainder of the compressed gas is allowed to pass through the proportional valve  144  to point  605 . The two streams are then combined in mixing circuit  140  to exit at an intermediate pressure and enthalpy point  607 . The liquid in this mixture, which is supplied to the tool  112 , is then evaporated in cooling tool  112  to point  608 . The gas at this point is then processed automatically in COR valve  150  to expand to a lower pressure suitable to enter compressor  158  at point  601 . The cycle then repeats. 
   The TCU can perform as a heat pump in supplying heat at a desired point. This is shown in  FIG. 8  which should also be considered along with  FIG. 1 . The operation shown herein is for supplying close to maximum heating at a temperature around 40° C. After being compressed from point  801  to point  802  most of the hot gas from compressor  158  is passed through the proportional valve  144  to a lowered pressure at point  805 . Controller  114  mixes an amount of condensed but high pressure liquid at point  803  that has been expanded through TXV  157  to point  809  with the gas at point  805  to provide a mixture at point  810 . The combination is then passed through tool  112  giving up heat to tool  112  in condensing liquid to point  804 . If the mix at point  804  were to be passed through COR valve  150  and input for compression in compressor  158  there would be so much liquid in the mix at point  804  that the compressor energy would be dissipated in evaporating liquid and the output pressure of compressor  158  would be too low. A pressure switch  168 , shown in  FIG. 1  senses this and activates heater  117  whenever the pressure sensed by switch  168  is below the threshold value. This action heats the liquid/vapor mix at point  811  and heats it to point  808  outside the liquid where it enters COR valve  150  and expands to point  801  where it then is all gas and ready to be recompressed. 
   There are, however, a number of other factors that can arise, particularly with respect to improvement of energy efficiency and safe, reliable operation. At the input to the compressor  158  the return line from the tool  112  passes through the subcooler  130 , acting to exchange heat energy between the condensed fluid from the condenser  156  and minimize loss of thermal energy by further cooling the fluid in the liquid line  132 . To assure that the mass flow at the compressor  158  input is sufficient, and above a potentially damaging minimum a loop from the output of the compressor  158  is fed through the HGBV valve  164  which ensures that the input to the compressor does not fall below a fixed pressure. The desuperheater valve  134  with a sensing bulb  126  at the compressor input ensures that the input to compressor  158  is cool enough for proper operation. The output of the desuperheater valve  134  is first passed through the liquid in the receiver  108 , when a receiver is used, and then feeds back to the return line into subcooler  130 , which passes through to the compressor  158 . 
   A separate control is effected at the condenser  156 . When the compressor  158  output is sensed by the pressure sensor  118 , and a signal is returned to the controller  114 , the consequent variation of the facility water source  154  assures that the condenser  156  is cooled sufficiently by the HEX  104  to maintain the refrigerant flow in the liquid line  132  substantially constant. 
   This system therefore provides a highly efficient heat exchange system in which the refrigerant is used directly under variable load conditions but maintained in a controlled, misted liquid/vapor phase when in contact with the tool  112 . This control in a principal mode is maintained by the controller  114  adjusting the proportions of the hot gas and the expanded liquid refrigerant at a selected pressure as determined by the heating or cooling needs of the tool  112  at a specific target temperature. Subsequent heat exchange in the tool itself may well occur, and the system and method stabilize or condition the refrigerant throughout the cycle. In the hot gas mode, with no flow in the liquid line  132 , the proportional valve  144  is opened to create the flow rate and temperature at the tool needed for maintenance of the target temperature, which with R507 refrigerant is thereby approximately 150° or more. For employing refrigerant in the lowest temperature range, only liquid line  132  need be used, and the TXV  157  is controlled to provide a cooling output to the tool  112  down to about −40° C. 
   As previously noted, other refrigerants can be used and the system can be designed to operate in a different mixed mode cycle of higher or lower value than the figures given. 
   It is to be appreciated that although different flow and variations have been disclosed the invention is not limited thereto but encompasses all alternatives and expedients within the scope of the appended claims.