COMBINED HEATING AND COOLING SYSTEMS

We describe a combined heating and cooling system, the system comprising: a working fluid circuit comprising a compressor, a gas cooler and an evaporator; a heating circuit, thermally coupled to said working fluid circuit via said gas cooler; and a cooling circuit, thermally coupled to said working fluid circuit via said evaporator; wherein said heating circuit further comprises a thermal storage tank, in particular a stratified thermal storage tank, controllably coupled to said heating circuit to controllably store heat for said heating circuit.

Skip to: Description  ·  Claims  · Patent History  ·  Patent History
Description
FIELD OF THE INVENTION

This invention relates combined heating and cooling systems, in particular using carbon-dioxide as a working fluid, and to control schemes for such systems.

BACKGROUND TO THE INVENTION

Using carbon dioxide as a working fluid in integrated heating and cooling systems that operate on the basis of a vapour-compression cycle has particular environmental benefits, but also poses some special problems. In the vast majority of vapour-compression cooling systems the energy contained within the discharge gas of a refrigeration compressor is jettisoned as if it were a waste product. Generally it has been considered that the jettisoned energy is insufficiently warm for it to be worth re-using in its entirety

A carbon-dioxide-based cooling and heating apparatus can be found in US2008/0245505. This describes an apparatus and method with two refrigerating cycle circuits and a sublimation heat exchanger. A heat pump using carbon-dioxide as a refrigerant and natural water such as well water and the like as a heat source is described in US2007/0261432. A cooling-heating device for an ice rink facility is described in CA2,748,027A. In broad terms this addresses the problem of maintaining a ratio of recovered heat to generated heat, and describes a system with multiple heat exchangers in the CO2 refrigerant line and a valve to control the flow rate of the CO2 refrigerant. A hot water supply and air conditioning system using a carbon-dioxide heat pump is described in US2009/0145149. This describes a system with low temperature and high temperature water tanks in a circuit including a gas cooler, and having a heat exchanger for recovering heat from river water and the like. However this system is not controllable and is relatively inefficient. Further background prior art can be found in: US2009/113911, US2007/056302, U.S. Pat. No. 6,668,572, JP2007/085685, JP2010/112657A, JP2008/224206A and JP2001/108317A, and in “CO2 Heat Pump For Domestic Hot Water”, Fornasieri E., et al, 8th IIR Gustav Lorentzen Conference on Natural Working Fluids, Copenhagen, 2008.

There is a need to improve upon these approaches, in particular to facilitate improvements in their efficiency.

SUMMARY OF THE INVENTION

According to a first aspect of the invention there is therefore provided a combined heating and cooling system, the system comprising: a working fluid circuit comprising a compressor, a gas cooler and an evaporator; a heating circuit, thermally coupled to said working fluid circuit via said gas cooler; and a cooling circuit, thermally coupled to said working fluid circuit via said evaporator; wherein said heating circuit further comprises a thermal storage tank, in particular a stratified thermal storage tank, controllably coupled to said heating circuit to controllably store heat for said heating circuit.

Embodiments of the above described system are adapted for control to facilitate efficient operation of the system. In particular by providing a controllable thermal store it is easier both to operate the system in an efficient regime when simultaneously heating and cooling and also to achieve stable control of the combined heating and cooling system.

In preferred embodiments the thermal storage tank is configured such that the tank is stratified, in particular into one or more of layers of fluid separated by one or more thermoclines (although a mechanically stratified tank may alternatively be employed). In one embodiment, as described later, a single tank is employed with a thermocline which moves up and down within the tank according to the amount of heat stored in the tank; optionally multiple such tanks may be provided “in parallel”. Additionally or alternatively however the stratified thermal storage tank may comprise a plurality of tank vessels or chambers coupled “in series”, for example stacked one above another or side-by side, and coupled together by one or more fluid flow conduits to allow fluid to move between them. Where the vessels or chambers are coupled side-by-side a conduit from the top (or an upper portion) of one vessel/chamber may couple to the bottom (or a lower portion) of the next vessel/chamber. Optionally two (or more) conduits may couple each vessel to each adjacent vessel, optionally one conduit allowing fluid flow in one direction (for example, up) and a second conduit allowing fluid flow in a second direction (for example, down). In a tank comprising a set of vessels the moving thermocline may be approximated by a change in the number of vessels containing hot (warmer) as opposed to cold (cooler) fluid (for example water). Suitable devices are known to those in the art and are also available for purchase.

Use of a stratified thermal storage tank is advantageous as this separates relatively warmer and cooler portions of the heating circuit fluid. This in turn facilitates achieving a low temperature for the input to the gas cooler, this low temperature in the heating circuit facilitating efficient operation of the working fluid circuit. Thus, paradoxically, employing stratified thermal storage not only facilitates obtaining a higher temperature for the portion of the heating circuit used for heating, it also facilities achieving a lower temperature in a different portion of the heating circuit which facilitates efficient operation of the overall system.

Further, and importantly, use of a stratified thermal storage tank allows the thermal storage tank to be used as a gauge in which the degree of stored thermal energy can be determined from a set of temperature sensors at different levels within the tank. This in turn facilitates control of the system based upon stored energy rather than on temperature per se. In the case of a tank comprising a set of vessels/chambers the stored thermal energy gauge may be provided by a count of the number of vessels/chambers containing hot (warmer) as opposed to cold (cooler) fluid. It is not, however, essential to employ stratified thermal storage as in principle the advantages of such an approach may be achieved in other ways, for example by means of multiple smaller thermal storage tanks.

In some preferred embodiments the system also includes a thermal dump system to enable heat to be dumped from said heating circuit, again to facilitate efficient overall system control, counter-intuitively dumping thermal energy facilitating an overall energy saving. In embodiments the heating circuit is configured, for example using valves, to direct or switch flow in the circuit between the thermal storage tank and the thermal dump, for example when the thermal store does not need to be replenished.

In preferred embodiments a heating side control system controls storage and dumping of heat for the heating circuit, in particular based on a level of stored energy. In embodiments this is measured by a set of temperature sensors in the stratified thermal storage tank, that is control is effectively based upon thermal energy stored in the heating circuit. In embodiments the heating side control system controls the compressor with the aim of maintaining the stored thermal energy at a substantially steady state, which may be defined by a target range. For example the virtual temperature gauge provided by the stratified thermal storage tank may be controlled to a target percentage/percentage range (of full capacity).

Preferred embodiments of the system also include a measure of stored coolth (the term of art “coolth” is explained in more detail later). The stored coolth may be measured by one or more temperature sensors in the cooling circuit. In preferred embodiments the heating side control system is responsive to both stored heat energy and to stored coolth, in particular to control the compressor and/or to dump heat from the heating circuit.

Preferred embodiments of the system also include a controllable coolth dump to allow coolth to be dumped from the cooling circuit. This may comprise, for example, a controllable heat exchanger to exchange heat with an ambient, typically external environment. Such a heat exchanger may also be used for “warmth dumping”, that is for free cooling where the ambient temperature is less than a target temperature for the cooling circuit. A cooling side control system is preferably included to control the coolth dump responsive to a sensed temperature of coolant within the cooling circuit, in particular to control a heat input to the evaporator provided by the coolant. In some preferred embodiments the coolant temperature for the cooling side control system is measured in a coolant flow path to an input to the evaporator, more particularly in the mixer header described later.

In some preferred embodiments of the system the cooling circuit has a coolth output (the ‘ring-main’ in the example described later), and is controllably reconfigurable between parallel and series modes of operation. In the parallel mode the controllable coolth dump is coupled in parallel with the coolth output, and in the series mode these are coupled in series. Optionally a configuration may be provided in which a portion of the output of the coolth dump may be mixed with a portion of the output from the evaporator. In preferred embodiments control is provided, preferably by the cooling side control system, to switch the cooling circuit between the series and parallel modes of operation dependent upon an ambient temperature (of the controllable coolth dump). More particularly the series mode may be selected when the ambient temperature is below a target temperature for the coolth output. Optionally a solar thermal energy capture system may be coupled to the heating side of the circuit to facilitate use of ‘free’ solar heating when available.

Although preferred embodiments of the systems and control techniques we describe are applied in the context of a combined heating and cooling system, in particular using carbon-dioxide as the working fluid, in principle some benefit is obtainable by applying the techniques we describe separately to one or other side of a system, that is to a system which heats but does not cool and vice versa.

In a related aspect the invention provides a method of controlling a combined heating and cooling system, in particular as described above, the method comprising: determining one or both of a stored heat in said heating circuit and a stored coolth in said cooling circuit; and controlling one or both of said compressor and said coupling of said stratified thermal storage tank to said heating circuit responsive to said determination of stored heat/coolth to maintain one or both of said stored heat and said stored coolth in a steady state, more particularly within a respective target range.

Again the stratified thermal storage tank may comprise a single tank or a set of stacked vessels (tanks/vessels may be coupled “in series” and/or “in parallel”). In embodiments of the method the heating circuit is controlled to control the stored heat and/or stored coolth, for example to maintain this is a steady state, in embodiments to achieve a target or target range of stored heat and/or stored coolth. In embodiments this is achieved by controlling the compressor and preferably also by controlling dumping of heat from the heating circuit. Whilst it might be thought that it would be most efficient to store all the heat generated, counter-intuitively the overall system efficiency can at times be increased by dumping heat.

In a further related aspect the invention provides a method of controlling a combined heating and cooling system, in particular as described above, the method comprising: determining a temperature of coolant circulating in said cooling circuit; and controlling dumping of coolth from said cooling circuit responsive to said coolant temperature to control a heat input to said evaporator provided by said coolant.

Again counter-intuitively, the overall system efficiency can at times be increased by dumping coolth from the cooling side of the system, so as effectively to heat the input to the evaporator (from the cooling circuit) and increase the efficiency of the working fluid circuit. In embodiments the coolth may be dumped by controlling the rate at which a heat exchanger operates (in embodiments, to exchange heat with an ambient, generally external environment); and/or by selecting a series or parallel mode of operation for the cooling circuit. In embodiments the method further comprises selecting a series mode of operation when ‘free’ cooling is available, that is when an ambient environment of the coolth dumping device is less than a target desired temperature for the coolant in the coolth output of the cooling circuit.

In preferred implementations both of the above control methods are implemented in a combined heating and cooling system of the type we describe. Preferably the working fluid comprises carbon-dioxide. In preferred embodiments the working fluid circuit is configured to operate (by default) with a transcritical cycle.

In a further aspect the invention provides a carbon dioxide-based combined heating and cooling system, the system comprising: a working fluid circuit comprising a compressor, a gas cooler and an evaporator; a heating circuit, thermally coupled to said working fluid circuit via said gas cooler and having heat output; and a cooling circuit, thermally coupled to said working fluid circuit via said evaporator and having a coolth output; and a first control system to control said working fluid circuit responsive to one or both of heat stored in said heating circuit and coolth stored in said cooling circuit to partly satisfy heat and coolth demands from respective said heat and coolth outputs with stored heat and/or coolth.

Thus in preferred embodiments the system is controlled based (partly) on stored heat and/or coolth, and preferably operates to control storage and/or dumping of heat and/or coolth. In embodiments the control is further facilitated by controlled dumping of either heat or coolth, for example to facilitate achieving an approximately steady state of stored heat and coolth (even with varying heat and coolth demands). In embodiments such a steady state may involve the stored heat and coolth being approximately in the middle of their respective ranges (for example 50%+/−30%), but alternatively the system may be biased towards greater or lesser storage of heat and/or coolth.

In preferred embodiments a further control system, preferably operating with a shorter cycle time, operates to control the cooling side circuit to control the temperature of the coolant flowing into the evaporator, again preferably by controlling at least dumping of coolth from the cooling side circuit. In embodiments the second control system may operate to control the coolant temperature up towards a target temperature. In broad terms this is advantageous because this raises the pressure of the working fluid (carbon dioxide) in the evaporator, which in turn increases the mass flow of working fluid through the evaporator and compressor, without causing a proportionate rise in the shaft power of the compressor. This helps to raise the “Coefficient of Performance” (COP) of the refrigeration cycle.

Preferably the second control system also controls the coolant temperature down towards a target to achieve a target cooling effect (from a coolth output of the cooling side circuit). As previously described the second control system may also selectively couple the coolth dump (heat exchanger) either in series or in parallel with the coolth output responsive to a sensed ambient temperature. That is a parallel configuration may be selected if the ambient temperature is greater than a target temperature of the coolant.

As previously described in preferred embodiments the system controls the heating side of the system in response to a determined stored heat gauge and/or a determined stored coolth gauge. As previously described, preferably the heating circuit includes a controllable heat dump to (indirectly) control the efficiency of operation of the working fluid circuit, in particular to facilitate dumping heat from the working fluid (carbon dioxide) circuit.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other aspects of the invention will now be further described by way of example only, with reference to the accompanying Figures, wherein like numerals refer to like parts throughout, and in which:

FIGS. 1A and 1B show a cooling circuit of an integrated heating/cooling system according to a preferred embodiment of the present invention, configured respectively into a parallel mode of operation and a series mode of operation;

FIG. 2 shows a heating side of the integrated heating/cooling system of FIG. 1;

FIG. 3 shows a control system for the integrated heating/cooling system of FIGS. 1 and 2;

FIG. 4 shows a pressure-enthalpy curve illustrating operation of part of the integrated heating/cooling system of FIGS. 1 and 2;

FIG. 5 shows a stratified thermal storage tank instrumented with a temperature sensors for use with the integrated heating/cooling system of FIGS. 1 and 2;

FIGS. 6A and 6B show, respectively, a compressor control procedure a thermal store/dump control procedure for the integrated heating/cooling system of FIGS. 1 and 2;

FIG. 7 shows regions of a lookup table for controlling operation of the integrated heating/cooling system of FIGS. 1 and 2;

FIGS. 8A, 8B and 8C illustrates a first example of operation of the integrated heating/cooling system of FIGS. 1 and 2; and

FIGS. 9A, 9B, 9C, 9D and 9E illustrate a second example of operation of the integrated heating/cooling system of FIGS. 1 and 2.

 DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

We will now describe details of a preferred embodiment of an integrated heating and cooling system (also referred to as a heatpump-chiller) which employs a water to water heat pump and which uses CO2 as the primary refrigerant. We also describe details of an associated control system that automatically regulates the heating and cooling of the heatpump-chiller.

We will consider an application where both a chilled water circuit and a heated water circuit are provided, for example a chilled water ring main and a low temperature hot water (LTHW) ring main. In the description below we generally refer to chilled water but the skilled person will understand that this may include an antifreeze such as glycol or salt (in which latter case the chilled water may be brine). However the skilled person will also recognise that the techniques we describe are not limited to use with any particular working fluid such as water (in the heating or cooling circuits), or even carbon dioxide. Similarly the skilled person will recognise that although in the description below we refer to a “ring main”, this merely refers to any form of circuit (which may include series and/or parallel sub-circuits).

We will describe techniques for employing a common, potentially single heat pump-chiller as the thermal engine. However to implement such a system in practice involves considerations relating to both the apparatus and the control system. For example, to cope with year round varying load profiles it is desirable to implement a control strategy that facilitates the heat pump-chiller adapting to out of balance situations. In broad terms this occurs where the quantum of warmth or coolth required by the LTHW & brine ring-mains respectively does not correlate with the achievable ratio of warmth and coolth that can be cogenerated by the heatpump-chiller. (In the context of a refrigeration system the heat input is sometimes termed the coolth output, measured in Joules or Jules/second).

We describe further below what is meant by “free cooling”, but in outline some preferred embodiments of the system employ a heat exchanger, more particularly an air-blast heat exchanger in the chilled water circuit, located in the external (outside) environment. In this way, when the external ambient temperature is below a target for the chilled water circuit, the ambient air can be used for “free” cooling of the chilled water.

In embodiments the heating/cooling system uses a vapour-compression refrigerating principle and is fitted with a gas cooler that is controlled to operate in a transcritical range of CO2 pressures. In some preferred embodiments the integrated heating/cooling system includes a system for heat (hot water) storage, in embodiments a thermally layered storage tank. This storage facilitates control of the system, in particular facilitating the joint operation of the heating and cooling portions of the system in a manner which provides some tolerance to varying loads on either side of the system (heating and cooling), such that the operation of the system can be overall more efficient. In some preferred embodiments the integrated heating/cooling system additionally or alternatively includes a mechanism to effectively provide cold (chilled water) storage, for similar reasons. In practice this can be considered to be provided by the circulating chilled water in combination with a measure of a cooling requirement to be provided by the circulating chilled water. The measure may comprise a temperature difference between a target cold temperature and the temperature of the circulating chilled water at a point prior to its use for cooling, that is, for delivering coolth to a process. Thus in embodiments a measure of such a temperature difference may be used as a proxy for the stored coolth in the chilled water (brine) circuit. We will sometimes refer to this later as a “virtual” coolth store.

Referring now to FIGS. 1A, 1B, 2 and 3, these show a preferred embodiment of the integrated heating/cooling system 100. A control system 300 is shown schematically; in embodiments it comprises a general purpose computer system or dedicated controller 310 including stored program code to control the operation of the system. The control system 300 also includes one or more interfaces 302 for controlling the heating and cooling circuits, more particularly valves, optionally pumps, and for controlling a CO2 refrigerant compressor of the combined heating/cooling system. It preferably also includes one or more sensor interfaces 304 for sensing temperatures and/or pressures within the system, and optionally for monitoring correct operation of components of the system. The control system may still further comprise a user interface 306 for interacting with, monitoring, configuring, and/or controlling the heating/cooling system; and optionally a wired or wireless network interface 308 for external communications and remote monitoring/control.

FIGS. 1A and 1B show a cooling (chilled brine) circuit 100a of the system, with valves controlled to put the circuit into parallel and series modes of operation respectively. Here, broadly speaking, “parallel” refers to an air blast heat exchanger and chilled ring-main configured so that fluid flows through both systems take place in parallel (ie fluid enters both systems from a common source at the same time), whilst “series” refers to an air blast heat exchanger and chilled ring-main configured so that fluid flows in series, firstly through the air-blast heat exchanger and then after that into a chilled water circuit.

FIG. 2 shows a heating (LTHW) portion of the circuit 100b. FIGS. 1 and 2 together show a complete system—evaporator 102 is common to FIGS. 1 and 2. The evaporator has a relatively lower pressure side 102a connected to the cooling circuit, more particularly through which the chilled water (brine) circulates, and a relatively higher pressure side 102b connected to the CO2 circuit, more particularly through which the CO2 liquid circulates.

Referring to FIGS. 1A and 1B, in the symbols for the valves open triangles indicate the flow is allowed through that part of the value and solid triangles indicate that flow is blocked. Thus the dashed (blue) lines indicate liquid (water/brine) paths in which there is substantially no flow whilst solid (red/green) lines indicate paths of liquid (water/brine) flow. Thus in the parallel mode of FIG. 1A air-blast heat exchanger 104 is connected in parallel with a chilled ring main 107 thus allowing chilled water flows from the evaporator to be distributed through the chilled water circuit and through the air-blast heat exchanger 104 at the same time whereas in the series mode of FIG. 1A the air-blast heat exchanger 104 is connected in series with chilled ring main 107 so that the air-blast heat exchanger can be used as a means of supplying extra coolth into the chilled water (brine) circuit

The chilled ring main typically comprises one or more cooling devices such as (air) coolers 107a-c which may be used, for example, for comfort cooling (air conditioning), for food conservation (to cool a food chiller cabinet), or the like. Air-blast heat exchanger 104 is preferably located outside a building or in some other environment external to or outside the regions heated/chilled by the system 100. A typical air-blast heat exchanger used in embodiments of the invention may comprise a coiled or meandering piped liquid path in combination with one or more fans to force air over the piped liquid.

The external air-blast heat exchanger may be used to dump (jettison) excess coolth from the chilled water (brine) circuit in what might be termed a “sky cooling mode” of operation. This mode is operable when the chilled water (brine) temperature inside the air-blast heat exchanger is more than a threshold colder, for example 2 or more degrees Celsius colder, than the ambient air. Similarly when the chilled water (brine) temperature inside the air-blast heat exchanger is more than a threshold warmer, for example 2 or more degrees Celsius warmer, than the ambient air then the air-blast heat exchanger may be used to dump (jettison) excess warmth from the chilled water (brine) circuit

Depending upon the geographical location, time of year, and operating environment the range of values of the ambient/outside temperature, perhaps −20° C. to +30° C., may overlap with the target temperature range for the chilled water circuit, perhaps −5° C. to +15° C. (the higher end encompassing air conditioners). These values are merely illustrative but are intended to convey the inventor's recognition that, at times and with a suitable control strategy, ambient air may provide an important part of the overall system control. By way of example, a chilled water circuit operating with a 12° C. flow and 7° C. return temperature can be maintained solely by an air-blast heat exchanger system without the need for mechanical refrigeration, so long as a sufficiently cold flow of ambient air is passed across the air-blast heat exchanger so that the temperature of the water at the exit of the air-blast heat exchanger is reduced to within the maximum allowable flow temperature of 12° C. (in this case).

In the illustrated embodiment the chilled liquid circuit comprises two main pumps, a first pump 106 and a second pump 108. The first pump 106 may be termed a chilled circuit pump because it circulates chilled liquid through the air coolers 107a-c of ring main 107, either directly (parallel mode) or indirectly (series mode). The second pump 108 may be termed an evaporator pump because it draws chilled liquid through the evaporator.

In the parallel mode first pump 106 circulates the cooled liquid through cooling devices (air coolers) 107a-c and the second pump 108 circulates the cooled liquid through air-blast heat exchanger 104 to chill the ambient air (dump excess coolth). In the series mode the first pump 106 circulates chilled water (brine) sequentially through the air-blast heat exchanger 104 and next through air coolers 107a-c; this may be a substantially constant circulation of liquid. The second pump 108 may operate intermittently or not at all to circulate liquid in the cooling circuit, in particular to mix liquid in a mixing tank, mixing header 110.

Mixing header (or “discharge header”) 110 is, essentially, a junction box so that the circulating flows meet and mix prior to flowing back through evaporator 102 and/or air coolers 107. In a similar manner preferably a suction header 112 is included in the path of liquid exiting the evaporator 102, to provide a “junction box” with parallel flow outputs on this side of the evaporator (in operation the “suction header” is substantially entirely full of liquid).

In addition, as previously outlined, the cooling circuit 100a of FIG. 1 includes a number of controllable valves, which may be controlled by the control system to configure the circuit for either a series or a parallel mode of system operation. The valves thus operate as mode selection valves; in embodiments the valves are controlled electrically by control system 200 to select fluid flow paths through the valves.

Thus in one embodiment the lower pressure side 102a of evaporator 102 has an output fluid flow path or conduit 130 coupled to (suction) header 112, which in turn has first and second output conduits 132, 134. (Here “conduit” is used broadly to mean any fluid flow path). Conduit 132 provides a first input to a first controllable (mode selection) valve 114, which has a second input from conduit 150 and an output conduit 136. Valve 114 is controllable to direct either the first or second input to output 136.

The second output conduit 134 of evaporator 102 is coupled to an input (suction side) of evaporator pump 108. The output conduit 136 provides a first input to an optional (but preferable) mixing valve 116, which has a second input conduit 138, and an output conduit 140. Conduit 140 provides an input to chiller circuit pump 106. The second input 138 to mixing valve 116 is from liquid which has passed through the chilled circuit 107, and which has thus been warmed by this passage, thus helping to warm the input to the chilled “ring main” in a condition in which there is excess coolth in the suction header. Valve 116 is preferably controllable to selectively mix input 138 with the flow between input 136 and output 140. We describe control of valve 116 later, but in embodiments if the chilled water (brine) temperature inside the suction header is greater than a threshold difference lower than a desired target temperature some of the output of the ring main may be mixed in.

An output conduit 142 of chiller circuit pump 106 provides an input to a second controllable (mode selection) valve 120, which has a first output conduit 144 to air-blast heat exchanger 104 and a second output conduit 146 to chilled “ring main” 107. Valve 120 is controllable to direct the input 142 to either the first or second output 144, 146.

The output conduit 146 of valve 120 is coupled to ring main 107, to provide a chilled water input to the air coolers 107a-c of ring main 107. An output conduit 148 from ring main 107 provides a first input to (mixing) header 110.

Pump 108 has an output conduit 152 which provides an input to a third controllable (mode selection) valve 118. Valve 118 has a first output conduit 154 joining conduit 144, to provide an input to (external) air-blast heat exchanger 104, and a second output conduit 156 which provides a second input to (mixing) header 110. Valve 118 is controllable to direct the input 152 to either the first or second output 154, 156.

An output conduit 158 from air-blast heat exchanger 104 provides an input to a fourth controllable (mode selection) valve 122, which has a first output conduit 160, and a second output conduit 162 which provides a third input to (mixing) header 110. The first output conduit 160 is coupled to conduit 146, so that circulating liquid from air blast heat exchanger 104 can provide an input to ring main 107. Valve 122 is controllable to direct the input 158 to either the first or second output 160, 162 (in series and parallel mode respectively).

The (mixing) header 110 has a first output conduit 150, which provides an input to valve 114 to enable circulation of liquid through the external air blast heat exchanger 104 and ring main 107 in series mode. The (mixing) header 110 has a second output conduit 164 which provides an input to the higher temperature side 102a of evaporator 102 and, in embodiments, provides a continuous path through evaporator 102 to output conduit 130. As described further below, evaporator 102 acts as a heat exchanger and the path 164, 130 through evaporator 102 provides one part of the flow through this heat exchanger. The skilled person will appreciate that within the heat exchanger the fluid conduit may be defined in many ways, for example by tubes, plates, baffles and the like.

The skilled person will appreciate that the configuration of the valves and conduits may be varied whilst still providing a cooling circuit which can be switched between parallel and series modes of operation. As previously described in such a system in the series mode liquid circulates sequentially through the external air-blast heat exchanger and chilled “ring main” (optionally but preferably water is also circulated through evaporator in a second circuit). In the parallel mode liquid from the evaporator is circulated in parallel through the external air-blast heat exchanger and chilled “ring main”, optionally mixing the output from the ring main with the input to the ring main, so that the ring-main feed-in temperature is not excessively cold.

It will be appreciated that the temperatures of the circulating liquid depend upon the application. Nonetheless it is helpful for understanding embodiments of the invention to provide illustrative example temperatures.

Thus for a system in which the chilled ring main cooling devices 107a-c comprise air coolers for chilling food the input 146 to these may be at around −5° C. whilst the output 148 from these may be at around −1° C. (for an air conditioning circuit the input temperature may be higher, for example around +10° C.). In this example an air cooler has air at, say, +1° C. blown over a conduit, radiator or the like carrying the chilled liquid (brine), so that the air is cooled to, say, around −2° C. whilst the brine is heated from, say, −5° C. to −1° C.

In this example in parallel mode the output 130 from the evaporator 102 may have become significantly colder than the targeted chilled ring-main feed-in temperature (it may have reached −8° C. for example, at the same time brine being fed into the air-blast heat exchanger 104 would also be at −8° C.). Parallel mode is selected when the external air temperature is high enough to allow dumping of coolth, for example −8° C. coolth can easily be dumped into ambient air of +9° C., and thus the output 158 from the air-blast heat exchanger could be substantially warmer than the water exiting air coolers 107 a-c. Thus in a food chilling application the mixing header may receive one flow at around −1° C. and another flow at around +5° C. (assuming 9° C. ambient).

By enabling unwanted coolth to be dumped through the air-blast heat exchanger in this way it is possible to impose a heat load on the evaporator 102 during times that require the CO2 refrigerating system to be operated primarily for the production of hot water, even though an excessive cooling effect (relative to that required by the chilled ring-main 107) may be emanating from the low pressure side of the CO2 refrigerating system.

In series mode the outlet 148 from coolers 107 a-c on the chilled water (brine) ring-main may again be at around −1° C. However the ambient temperature is sufficiently low in this mode of operation for the air-blast heat exchanger 104 to be able to at least partially cool the output from the chilled ring-main 107. In preferred embodiments of series mode the cooling for the chilled ring-main can at times be provided solely by the air-blast heat exchanger 104.

Referring now to FIG. 2, this shows a heating (LTHW) portion of the circuit 100b, in which evaporator 102 is common to the circuits shown in FIGS. 1A, 1B and 2. In embodiments the heating circuit side of system 100 includes a vapour (gas) compression refrigeration system. Thus the evaporator 102 of FIG. 2 is the same evaporator 102 as illustrated in FIGS. 1A and 1B. We will describe a preferred implementation of the system which employs carbon dioxide as the vapour but in principle other gases may also be employed.

In more detail, conduit 202 carries high pressure dense gas, in particular carbon dioxide, towards an expansion valve 204, which allows the pressure to reduce, with a concomitant reduction in CO2 temperature. In embodiments expansion valve 204 may be an automatically adjustable small bore needle valve. Expansion valve 204 converts the gas, which at this point is a dense mist, into a cold, mostly liquid state in conduit 206 leading away from expansion valve 204. By way of example, the temperature of the liquid in conduit 202 may be around 30° C. whilst the temperature of the fluid in conduit 206 may be around −10° C. For efficient operation it is preferred that there is some turbulence in the flow within conduit 206. This is facilitated, in particular, by a mixture of “wet” and “dry” fluid, which in turn can be achieved by arranging for the vapour (carbon dioxide) in conduit 206 to be in a region of the pressure enthalpy curve in which the vapour is not fully wetted but is also not subject to a vapour fraction in excess of 0.5

Conduit 206 provides a continuous path through evaporator 102 to output conduit 208. During its path through evaporator 102 the carbon dioxide is boiled, for example at 10° C. to the point where 5K or more of superheat is generated by the counter-flowing liquid (brine/glycol) in the conduit 164 to 130. As energy is passed from the brine to the CO2 then the brine is reduced in temperature eg. from −1° C. at inlet to 102 to say −5° C. at the outlet from evaporator 102.

During its passage through the evaporator the carbon dioxide is also “dried”, although there may still be some residual dampness. For this reason conduit 208 is preferably provided with a droplet separator 210 to remove residual liquid droplets from the flow prior to compression. Thus in embodiments an output 212 of (optional) droplet separator 210 is provided to a compressor 214, and this in turn has an output 216 to an optional oil separator 218 (to remove residual compressor oil), and thence to conduit 220. Compressor 214 raises the pressure of the carbon dioxide, and also the temperature of the carbon dioxide, for example to around 85° C.

Conduit 220 provides an input to a heat exchanger 222, which may be referred to as a “gas cooler”: During its passage through this element the temperature of the carbon dioxide is reduced prior to its passage through expansion valve 204, using the preceding figures to around 30° C. (the temperature in conduit 202). A mechanical filter 226 is preferably provided between an output conduit 224 of heat exchanger 222 and expansion (needle) valve 204.

Heat exchanger 222 is coupled to a water heating circuit. For convenience we will sometimes refer to this circuit as an LTHW (low temperature hot water) circuit, but the skilled person will appreciate that other types of heating circuit may also be implemented. Thus in embodiments heat exchanger 222 has an input conduit 228 and an output conduit 230, together forming part of the LTHW (heating) circuit. By way of example, input conduit may carry water at around 28° C., which cools the counter-flowing carbon dioxide gas flowing through conduits 220, 224, in turn heating the water so that output conduit 230 may, for example, be at around 55° C. The skilled person will appreciate that the degree of heating depends both on the temperature of the counter-flowing gas and also on the mass flows of the gas and water.

The output conduit 230 from heat exchanger 222 provides a source of hot water for thermal storage device 232, in preferred embodiments a layered (stratified or thermocline) thermal storage device. Conduit 230 provides an input to an upper, high temperature region of the tank where the water temperature may be, for example, around 55° C. Stratification within tank 232 maintains a temperature differential between upper and lower regions of the tank and thus a lower region of the tank may at the same time be at a much lower temperature, for example around 30° C., without substantial mixing between the stratified layers. An output conduit 234 from a lower region of tank 232 provides a lower temperature outlet and a lower temperature return to conduit 228, an input to heat exchanger/gas cooler 222. In embodiments this return is via a liquid pump 236 (although it will be appreciated that the pump may be located elsewhere); this pump may be termed a gas cooler pump.

In preferred embodiments the LTHW circuit also includes a heat exchanger 240 which is usable to dump excess heat, and which may therefore be termed a hot side dump exchanger. This heat exchanger may be a liquid-to-liquid heat exchanger or a liquid-to-gas heat exchanger; in the latter case it may be situated in any convenient location to dump heat, for example external to a heated environment/building. An input to heat exchanger 240 is provided by a conduit 248, a branch of conduit 230. A return conduit 242 from thermal-dump heat exchanger 240 is coupled a first input of a (dump) changeover valve 244. Valve 244 has a second input from conduit 234, and an output conduit 246 which provides an input to pump 236. Thus valve 244 is controllable to either permit or inhibit a heated water flow through thermal-dump heat exchanger 240. Valve 244 is likewise controllable to either permit or inhibit a heated water flow into (through) the thermal store 232, and thus effectively to switch flow between the thermal store and thermal dump. One or more additional shut-off valves may be operated in coordination with valve 244 so that gas passing through 222 is cooled either via the thermal store 232 or via the hot-side dump exchanger 240.

Optionally a solar water heater 260 may also be coupled to thermal store 232, as shown in simplified form, to enable solar thermal heating input to the system. In this configuration the solar thermal panel(s) would serve (instead of the gas cooler) as a means of generating heat into the thermal store.

A heating circuit 250, in embodiments an LTHW “ring main”, is coupled to thermal storage tank 232. More particularly in embodiments a first conduit 252 from an upper, heated region of tank 232 provides an input to heating circuit, which has a second, output conduit 254 providing an input to a lower, cooler region of the tank. A pump 256 pumps water through heating circuit 250, which includes one or more heating devices 256a-c, for example radiators.

Referring again to the carbon dioxide vapour (gas) compression refrigeration circuit, in preferred embodiments this operates in a transcritical mode, that is as it circulates through the refrigeration circuit the carbon dioxide defines a transcritical cycle (enclosing the critical point) on a pressure (p) enthalpy (H) graph for the carbon dioxide. We use the terms vapour and gas interchangeably herein.

FIG. 4. shows such a pressure-enthalpy curve with labelled points A to D corresponding to labelled locations A to D in FIG. 2. In FIG. 4 the critical point is labelled X and lies at the top of the dome defined by the saturated liquid line (to the left) and saturated vapour line (to the right). It can be seen that the closed curve defining the carbon dioxide cycle extends above the critical point encompassing part of the supercritical region.

We now describe an example cycle: Point D labels the output from expansion valve 204, the input to evaporator 102. As the (mostly liquid) carbon dioxide passes through the evaporator at substantially constant pressure it is boiled at −10° C. (at the saturated vapour line), and then heated-up further, to −5° C. (in the superheated region), to arrive at point A where it has 5° C. of superheat. The skilled person will appreciate that the extent of the superheating at point A is variable.

The vapour is then compressed by compressor 214, moving from point A to point B at the output of the compressor and heating the gas to, for example, 85° C. At point B the gas is beyond the critical point X, in a supercritical region. During its passage through gas cooler 222 the vapour is cooled at approximately constant pressure, for example to 30° C., to arrive at point C, the input to the expansion valve 204. Expansion valve 204 reduces the pressure, down to point D, reducing the temperature, for example to around −10° C., liquefying the vapour. It should be noted that whilst transcritical operation is the default mode (when the gas cooling pump connects via the thermal storage tank) it may be the case that the CO2 refrigerating cycle operates entirely below the critical point when the hot-side dump exchanger is engaged.

Control Schemes

We next describe some preferred examples of control schemes for efficient operation of the above-described system. In broad terms it is desirable to provide thermal storage on the heating and cooling sides of the system. This may be inherent, provided by the heated/cooled liquid circulating within the respective circuits and/or additional thermal storage may be provided, for example using a layered thermal storage tank as previously described. Thus we refer below to stored heat and stored coolth. Efficiency may still further be increased by taking advantage of effectively free cooling which may be provided by a cold external ambient environment (by comparison with a target temperature) at certain times of year.

Consider, for example, a system which is providing cooling for air conditioning units. If, say, the system makes “10 units” of cooled liquid but the air conditioners need only 6, there is excess coolth. Thus it is beneficial in such a situation to dump some excess coolth. Where the external ambient temperature is high enough this can be achieved by using air-blast heat exchanger 104 to “refrigerate the sky”, mixing the warmed brine return from air-blast heat exchanger 104 in mixing header 110.

The above example illustrates an example of an unbalanced condition which the control system should aim to address. We now describe preferred embodiments of suitable control schemes. In embodiments two control schemes operate, one for the CO2 refrigerating system and one for the air-blast heat exchanger 104. Although these systems are essentially distinct it is the case that changes made by one control system may lead to changes needing to be made by the other control system.

Heat Storage—Temperature Sensing

In embodiments the stratified thermal storage tank 232 is instrumented with a set of temperature sensors at different levels within the tank, so that the stored energy in the tank can be determined. Referring to FIG. 5 there may, for example, be four temperature sensors sensing temperatures T1 to T4 at successively lower levels within the tank. The stored energy level (SEL) may then be defined as 25%, 50%, 75% or 100% according to whether T1; T1 and T2; T1, T2 and T3; or T1, T2, T3 and T4 are at greater than a threshold temperature (preferably the same but potentially different for each of T1 to T4). Optionally a 0% level may be defined if T1 is not greater than the threshold. An example table for determining the stored (heat) energy level is shown below (Table 1). Conceptually this provides a stored heat or energy level “fuel gauge” for the system, as shown in the inset to FIG. 5. It will be appreciated that this approach may be adapted for other numbers of temperature sensors.

TABLE 1 Stored (heat) energy level (%) 100 75 50 25 T1 >threshold >threshold >threshold >threshold T2 >threshold >threshold >threshold <=threshold T3 >threshold >threshold <=threshold <=threshold T4 >threshold <=threshold <=threshold <=threshold

In preferred embodiments the control system 300 measures temperatures T1 to T4 at intervals, for example every five minutes, and determines the stored (heat) energy level, for example in terms of a percentage, as above.

Coolth Storage—Temperature Sensing

In a similar manner, the cooling (chilled brine) circuit 100a is instrumented with at least one temperature sensor to measure a temperature of liquid circulating within the circuit, in embodiments a temperature, TSH, of the input to chilled ring main 107 (FIGS. 1A and 1B), measured by a sensor located in the suction header 112.

A difference between this temperature and a target temperature Ttarget (see below) is used as a measure of the stored coolth. This can be seen by considering a target temperature of −5° C.; where TSH is, say, −8° C. the chilled brine ring-main could in this circumstance run for some time without further cooling of the brine because of the coolth stored in it, whereas if TSH is at a temperature of, say, −4° C. then the chilled brine ring-main could not operate for long (if at all) without additional coolth being provided by the evaporator 102. Preferably, but not essentially, when the target temperature is approximately the same as the measured TSH the stored coolth is deemed to be at a 50% level. In embodiments the target temperature is a target temperature for the circulating liquid (brine) but in principle other temperatures may be used, for example a target temperature of fresh foods being chilled by an air cooler

An example table for determining a stored coolth level is shown below (Table 2), where ΔT=TSH−Ttarget. Conceptually this provides a stored coolth gauge for the system. It will be appreciated that the threshold temperature values may be varied, and that this approach may be adapted for finer or coarser gradations of stored coolth. In embodiments a stored coolth level of 0% is not used.

TABLE 2 Stored coolth gauge (%) 100 75 50 25 ΔT <= −3° C. ΔT <= −1° C. −1° C. < ΔT < +1° C. ΔT >= +1° C.

In preferred embodiments the control system 300 measures TSH at intervals, for example at around the same time as T1-14, say every five minutes, and determines the stored coolth level, for example in terms of a percentage, as above. This gives information about the level of cooling that is (or is not) being delivered whilst heat is also being (or is not being) delivered by the compressor to the thermal storage tank.

Heating-Side/Compressor Control

In preferred embodiments of the system data from the stored energy level (SEL) and stored coolth gauge (SCG) are processed jointly to control the compressor 214, gas cooling pump 236, and changeover valve 244 (to control heat dumping).

In preferred embodiments the control system operates with the aim of keeping both the stored heat and stored coolth (“gauges”) at around 50% full. In this way there is heating potential available in the heating circuit and cooling potential available in the cooling circuit. This approach also tends to optimise (power) efficiency. Alternatively, however, the system may be run with a bias towards either the heating or the cooling circuit.

As described above, embodiments of the system may operate with a control cycle which operates at intervals of, say, 300 seconds, to monitor the SEL and SCG and in response control the heating circuit side of system, more particularly the compressor, gas cooling pump and dump changeover valve. However in embodiments the control cycle is adaptive, and may change depending upon the SEL and/or SCG.

Preferred embodiments of the control system define a set of (discrete) compressor speeds and then control the compressor by incrementing or decrementing the compressor speed between one level to another (although a similar concept may also be applied to a continuously variable compressor speed). In the description below, increasing the speed is referred to as “loading” the compressor and decreasing the speed as “unloading” the compressor; “stay” denotes leaving the compressor speed unchanged.

In one, preferred approach the control system employs a lookup table to determine whether to load, stay or unload the compressor, as shown in the table below (Table 3). Table 3 also defines whether the dump changeover valve 244 is controlled so that pump 236 pumps to the layered thermal store 232 (“thermal store”) or to the hot-side dump heat exchanger 240 (“hot dump”).

TABLE 3 SCG SEL Pump 236/valve 244 (%) (%) Compressor directs heat to 25 25 Load Thermal store 25 50 Load Thermal store 25 75 Load Hot dump 25 100 Load Hot dump 50 25 Load Thermal store 50 50 Stay Thermal store 50 75 Unload Thermal store 50 100 Unload Hot dump 75 25 Load Thermal store 75 50 Unload Thermal store 75 75 Unload Thermal store 75 100 Unload Hot dump 100 25 Load Thermal store 100 50 Unload Thermal store 100 75 Unload Thermal store 100 100 Unload Hot dump

The underlying logic of Table 3 may also be implemented by the procedures shown in FIGS. 6A and 6B (although use of a lookup table is potentially more flexible).

Thus referring to FIG. 6A, which shows a compressor control procedure, at step S604 the system reads the SEL and SCG gauges and, at S606, determines whether either is at 25%. If so the compressor is loaded (sped up), S608, and the procedure loops; otherwise if both are at 50% the compressor speed is unchanged (S612). At step S614 one or both of SEL and SCG is at 75% or 100% and neither is at 25%; in this situation the compressor is unloaded (the compressor speed is decreased).

FIG. 6B shows a control procedure for the thermal store/dump: At step S624 the procedure reads the SEL gauge and if this is less than or equal to 50% (S626), heat is stored (S628), and if the gauge is at 100% (S630) heat is dumped (S632). Otherwise SEL is at 75% and the procedure stores heat if the compressor is being unloaded by the compressor control procedure and dumps heat if the compressor is being loaded by the compressor control procedure.

Optionally the above described control logic/procedures may optionally be state-dependent, that is the control applied may depend upon a previously applied control output. In particular if the compressor was unloaded (decreased) in a previous control cycle, and a subsequent unload (decrease) is indicated then it may be unloaded (decreased) faster. A similar approach may be applied when loading (increasing) the compressor. Additionally or alternatively the cycle period may be shortened in such situations.

It will be appreciated that the two control loops of FIGS. 6A and 6B interact as the SEL gauge of FIG. 6A is affected by the thermal store/dump control of FIG. 6B, and the control loop of FIG. 6B is affected by the compressor speed adjusted by the procedure of FIG. 6A. In broad terms, embodiments of the combined procedures aim to keep one or both the SEL and SCG gauges at a steady state. This may, for example, be at around 50% or some other value—for example if providing a constant summer air conditioning load, say during times of low demand for low temperature hot water, then the SEL might read say 75% all day long whilst the SCG oscillates between say 25% and 50%.

Nonetheless, although joint control of the compressor has been described based upon both the SEL and SCG “gauges”, in principle (but less preferably) compressor control may be based upon just one of these by simply assigning a false value to the gauge that is not needed. For example the SEL parameter could be assigned a value of 100% to inhibit the compressor from altering its speed for any reason other than for the maintenance of temperatures in the suction header (chilled brine circuit).

Cooling-Circuit Control

In broad terms the air-blast heat exchanger control operates to facilitate a mixing header temperature that is consistent with energy efficient operation. It also preferably (but not essentially) switches the brine circuitry between series and parallel modes of operation, in such a way that the mixing header (110) temperature can be raised (parallel) or lowered (series)

In embodiments the air-blast heat exchanger control measures a temperature of the circulating coolant (for example water/brine/glycol) in the mixing header (110), and an ambient (air) temperature at entry into air-blast heat exchanger 104, TMH and Tamb respectively. The circulating coolant temperature is preferably measured at a location in the return path to the evaporator 102, conveniently at the mixing header 110. The air-blast heat exchanger control system operates to control a rate of exchange of heat with the ambient air by controlling the air-blast heat exchanger fan speeds, series-parallel switchover valves, and on/off operation of pumps 106, 108. In one, preferred approach the control system employs a lookup table to define a control strategy for controlling these elements.

Table 4, below, shows an example lookup table, where T_glycol refers to TMH (noting that the coolant need not be glycol). As described further below, Table 4 encodes information for mode, pump and air-blast heat exchanger control but in practice multiple separate tables may be employed. Table 4 is also referred to later as the Glycol Ambient Matrix (GAM).

TABLE 4a T. glycol T. amb −37 −35 −33 −31 −29 −27 −25 −23 −21 −19 −17 −15 −13 −11 −9 −7 −5 30-33 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 27-30 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 24-27 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 21-24 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 28-21 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 15-18 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 12-15 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0  9-12 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 6-9 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 3-6 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 0-3 100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 −3-0  100 100 100 100 100 100 100 100 100 100 80 60 40 20 15 10 0 −6 to −3 100 100 100 100 100 100 100 100 100 100 100 80 60 35 10 0 0 −9 to −6 100 100 100 100 100 100 100 100 100 100 100 80 60 35 10 0 0 −12 to −9  100 100 100 100 100 100 100 100 100 55 10 0 0 0 0 0 0 −15 to −12 100 100 100 100 100 100 100 100 55 10 0 0 0 0 0 0 0 −18 to −15 100 100 100 100 100 100 55 10 0 0 0 0 0 0 0 0 0 −21 to −18 100 100 100 100 55 10 0 0 0 0 0 0 0 0 0 0 0 −24 to −21 100 100 100 55 10 0 0 0 0 0 0 0 0 0 0 0 0 −27 to −24 100 100 55 10 0 0 0 0 0 0 0 0 0 0 0 0 0 −30 to −27 100 10 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0

TABLE 4b T. glycol T. amb −3 −1 1 3 5 7 9 11 13 30-33 0 0 0 0 0 0 0 0 0 27-30 0 0 0 0 0 0 0 0 0 24-27 0 0 0 0 0 0 0 0 0 21-24 0 0 0 0 0 0 0 0 0 28-21 0 0 0 0 0 0 0 0 0 15-18 0 0 0 0 0 0 0 0 0 12-15 0 0 0 0 0 0 0 0 0  9-12 0 0 0 0 0 0 0 0 0 6-9 0 0 0 0 0 0 0 0 0 3-6 0 0 0 0 0 0 0 0 0 0-3 0 0 0 0 0 0 0 0 0 −3-0  0 0 0 0 0 0 0 0 0 −6 to −3 0 0 0 0 0 0 0 0 0 −9 to −6 0 0 0 0 0 0 0 0 0 −12 to −9  10 55 100 100 100 100 100 100 100 −15 to −12 10 45 80 100 100 100 100 100 100 −18 to −15 10 35 60 80 100 100 100 100 100 −21 to −18 10 25 40 60 80 100 100 100 100 −24 to −21 10 15 20 40 60 80 100 100 100 −27 to −24 10 15 20 40 60 80 100 100 100 −30 to −27 10 15 20 40 60 80 100 100 100

The different regions of Table 4 above may be appreciated by cross-referencing to FIG. 7. Table 4a shows regions A, C and E of FIG. 7; Table 4b, which should be considered as lying horizontally to the right of Table 4a, shows regions B and D of FIG. 7.

FIG. 7 shows a version of the table in which regions A, B and C are regions in which the system operates in a parallel mode (FIG. 1A) and in which region D denotes a region of series operation (FIG. 1B). Region E denotes a region in which the system may be configured for either series or parallel operation—for example this region may be selected to be one or the other mode (by default, series), or hysteresis may be applied so that the region denotes parallel operation if approached from a parallel region (typically region C) or series operation if approached from series region D.

The numbers in Table 4 denote a percentage rate of fan speed of the air-blast heat exchanger fans. The quantum of heat exchange in air-blast heat exchanger 104 is a function of fan speed and temperature difference (between the ambient air and the brine flowing through the air-blast heat exchanger).

The evaporator pump 108 is preferably ON in all operational regions except for region B (orange “0”s in Table 4), where evaporator pump 108 is preferably OFF.

Table 4 above is an example table for a target coolant temperature, TMH, of −1° C. (suitable for merchandising refrigerators). The target coolant temperature is preferably, but not essentially, measured in the mixing header. In general different target temperatures will employ different lookup tables. For example for an air conditioning circuit TMH may be around +14° C., whilst for (industrial) process cooling TMH may be higher, for example around +20° C. We describe below the principles underlying the table, which enable an adapted table to be constructed for different target coolant temperatures. Nonetheless in embodiments the use of a lookup table is preferred over a procedural approach, again, because it provides flexibility.

Series/Parallel Mode

In Series Mode air-blast heat exchanger 104 is controlled so as to reduce the value of TMH until it is equal to (or close to) its target temperature. Air-blast heat exchanger 104 is able to work in this way so long as the ambient air being passed across the air blast heat exchanger is colder (by a margin, for example of 2 or more degrees Celcius) than the targeted TMH. The heat exchange capacity of air-blast heat exchanger 104 is proportional to the temperature difference between the brine inside the heat exchanger and the temperature and mass flow of the air flowing across the outside of the heat exchanger.

In parallel mode air-blast heat exchanger 104 is controlled so as to increase the value of TMH until it is equal to (or close to) its target temperature. If the ambient air is colder than the targeted TMH then the air-blast heat exchanger fans are operated only to the point where they can no longer increase actual TMH (regions A and C of FIG. 7). For example, if the actual TMH is −8° C. and if the air ambient air temperature is −3° C. then it might be possible to warm the brine inside the mixing header to a temperature of say −6° C. (allowing for a temperature difference between the ambient air and the brine inside the mixing header). There would be no benefit to be gained by attempting to warm the brine any further and so fans would at this point be switched off even if the actual TMH was colder than the target TMH.

Coordinated Control

As described, there are two control loops; one that operates according to a set of temperature measurements taken in the thermal storage tank and in the suction header (112) and which affects the behaviour of the compressor and the heat storage/dumping system and another that acts according to prevailing temperatures in the mixing header (110) and which influences the behaviour of the fans fitted to air-blast heat exchanger 104, pumps 106 and 108 and a set of mode selection valves for parallel and series configuration These control loops interact indirectly in that the temperature inside the mixing header (which may be influenced by the control of the air blast heat exchanger) and the temperature inside the suction header (which may be influenced by the CO2 compressor) will affect one another. For example it can be the case in series mode that the mixing header is being cooled by the air-blast heat exchanger whilst the suction header is simultaneously being cooled due to the compressor. In this situation it can be considered that the air-blast heat exchanger inhibits the build-up of excessively warm water at the exit of the cooling ring-main 107 whilst the compressor system inhibits the build-up of excessively warm water at the entry into cooling ring-main 107.

Achieving balanced operation in which the two control loops reinforce (rather than work against) one another is facilitated by setting appropriate time intervals for the operation of the two control loops. For example the air-blast heat exchanger (“GAM control”) may operate every 90 seconds, whilst the compressor (SEL/SCG control) loop may operate once every 300 seconds. Preferably the air-blast heat exchanger control operates 5 or more times more frequently than the SEL/SCG control; preferably the intervals at which each control scheme operates are adjustable. Balanced operation is further facilitated by assigning appropriate target temperatures to TMH and TSH.

For example, in an application where the temperature set point for air inside an air-conditioned room is 22° C. and the volume of air and brine flowing through the coolth emitters in the room is fixed then a practical method of controlling the thermal output of the coolth emitters is to maintain a stable brine temperature at exit from the emitters whilst at the same time allowing the brine temperature at inlet to the emitters to float.

For example the brine exit temperature (TMH) from a chilled brine ring-main could be held at say 14° C. whilst the brine inlet temperature (TSH) might in practice be allowed to fluctuate between 9° C. and 12° C. Thus in the case of the heatpump-chiller described here it is feasible to task the air blast heat exchanger 104 to maintain 14° C. (TMH) at the same time as assigning the SCG a set point of say 10° C. (TSH).

In this example if the ambient air temperature was 4° C. and in the absence of exceptionally high loads being imposed on the coolth emitters then the air blast heat exchanger 104 would by targeting a brine exit temperature of 14° C. effectively be delivering a brine temperature at entry to the coolth emitters of 9° C. to 12° C. Allowing for dead band effects on temperature set points it can be seen that in this example a SCG set point of say 10° C. may well result in the SCG reading a stored coolth value of 50%, which in the absence of a low SEL reading for stored warmth would not necessarily require for the compressor to be put into operation. In other words a free-cooling service via the air blast heat exchanger 104 acting alone would be sufficient to keep the air conditioning system operating within its required operating range.

If in the same example the compressor was later compelled by a low SEL reading to operate (at the same time as which the air blast heat exchanger 104 was operating) then in practice the air blast heat exchanger would need to reduce its cooling output and its fans would be therefore slowed down. Indeed if the compressor was operated at high enough speed for long enough then the air blast heat exchanger 104 fans would be switched off and parallel configuration might be engaged, eventually leading to the fans being switched back on so that the air blast heat exchanger could dump excess coolth emanating from the evaporator as a result of the compressor being in operation.

Broadly speaking we have described how a LTHW system and a chilled brine system may be driven from a common heatpump-chiller that automatically switches its emphasis between LTHW production (in which case the heatpump-chiller focuses on providing load to the thermal storage tank) and heat dumping (in which case heat is deliberately dumped in order to generate extra cooling effect in the brine circuit).

We now give some examples of operation of the cooling side control.

Example A Series-Mode Circuit

Referring to FIGS. 8A-8B, which corresponds to FIG. 1B, this example illustrates the effect of using an air blast heat exchanger 104 for cooling during colder weather. FIG. 8A shows temperatures in the system at t=0 minutes and FIG. 8B at t=2 minutes. FIG. 8C corresponds to a version of FIG. 7 and illustrates the main features of a lookup table for control of the (fan of) cooler 104. FIG. 8A corresponds to point a1 in FIG. 8C and FIG. 8B to point a2.

The target coolant temperature (in the mixing header) is, in this example, 21° C., a temperature suitable, for example, for an industrial process. The ambient temperature is significantly lower than this and if the temperature of the mixing header becomes excessive this can quickly be brought under control by the control of the cooler fan(s), even if the compressor is not providing much cooling effect. As previously discussed, in embodiments the scan rate of the GAM control procedure may be 90 seconds as compared to 600 seconds for the heating side (SCG+SEL) control. Thus the GAM control “fine tunes” the step changes implemented by the heating side control.

Example B Parallel-Mode Circuit

Referring to FIGS. 9A-9D, which corresponds to FIG. 1A, this example illustrates operation of the system with the cooling side in a parallel mode of operation. Thus FIGS. 9A to 9D show temperatures in the system at, respectively t=0, 6, 12, 18 minutes, and FIG. 9E illustrates the main features of a lookup table for control of the (fan of) cooler 104. In FIG. 9A the recirculation port 138 of mixing valve 116 is closed; in FIG. 9B it is partially open; in FIG. 9C it is wide open; and in FIG. 9D it is partially open once again. The SEL and SCG “gauges” are also illustrated as insets. The target coolant temperature (in the mixing header) is, in this example, 15° C., and the target suction header temperature is 10.0° C. In broad terms, the air blast heat exchanger fan(s) and valves are operated in such a way as to influence the temperature of the mixer header 110. The SCG uses the temperature of the suction header 112 as a means of determining how much or how little coolth is available to the pumps which draw from the suction header.

A short explanation of the operation at each step is given below (where b1 to b4 refer to the points in FIG. 9E):

t=0 mins (b1)—FIG. 9A
TAMB=23° C., TMIX=22° C., Cooler fan(s)=NIL. The cooling side circuit is set in parallel mode.
The set point for TSUC-HDR=10° C.; TSUC-HDR (ACTUAL)=14.5° C.
TSUC-HDR (ACT-TGT)>1 therefore the stored Coolth Gauge (SCH) reads 25%
t=6 mins (b2)—FIG. 9B

Due to the low cooling potential of the suction header (SCG=25%) the compressor speed has been incremented. After a short time the suction header became colder and a value of TSUC-HDR=9.0° C. was read, which is within the target range for TSUC-HDR—and so SCG=50%. At the same time it was noted that TMIX HDR=14.5° C., which is still in the no fan zone (allowing for a dead band).

t=12 mins (b3)—FIG. 9C

Strong demand for heating has depleted the thermal store. The Stored Energy Level (SEL) in the thermal store reads 25%. The combination of SEL=25% and SCG=75% sends a “load signal” to the compressor. This forces a low value of TMIX HDR=6.0° C. and the cooler 104 is set to run at “Parallel-Mid Speed Fan” (FIG. 9e).

t=18 mins (b4)—FIG. 9D

TMIX HDR=14° C., which is close to ideal, so the cooler fan(s) are slowed down. Since increasing the compressor speed (see above) the thermal store has been replenished and SEL=50%. At the same time TSUC-HDR=7.0° C. which equates to SCC=75%. At SEL=50% and SCG=75% the compressor is unloaded.

No doubt many other effective alternatives will occur to the skilled person. It will be understood that the invention is not limited to the described embodiments and encompasses modifications apparent to those skilled in the art and lying within the spirit and scope of the claims appended hereto.

Claims

1. A combined heating and cooling system, the system comprising:

a working fluid circuit comprising a compressor, a gas cooler and an evaporator;
a heating circuit, thermally coupled to said working fluid circuit via said gas cooler; and
a cooling circuit, thermally coupled to said working fluid circuit via said evaporator;
wherein said heating circuit further comprises a thermal storage tank, in particular a stratified thermal storage tank, controllably coupled to said heating circuit to controllably store heat for said heating circuit.

2. A combined heating and cooling system as claimed in claim 1 wherein said heating circuit further comprises a thermal dump controllably coupled to said heating circuit to controllably dump heat from said heating circuit.

3. A combined heating and cooling system as claimed in claim 1 wherein said stratified thermal storage tank comprises a set of heat storage temperature sensors at different stratified levels within the tank, the combined heating and cooling system further comprising a heating side control system responsive to said storage temperature sensors to control storage of heat in said thermal storage tank and dumping of heat from said heating circuit.

4. A combined heating and cooling system as claimed in claim 3 wherein said heating side control system is configured to control said compressor to maintain thermal energy stored in said stratified thermal storage tank within a target range.

5. A combined heating and cooling system as claimed in claim 4 further comprising at least one coolth storage temperature sensor, and wherein said heating side control system is configured to control said compressor responsive to said at least one coolth storage temperature sensor to maintain coolth stored in said cooling circuit within a range.

6. A combined heating and cooling system as claimed in claim 1 wherein said cooling circuit further comprises a controllable coolth dump to controllably dump coolth from said cooling circuit.

7. A combined heating and cooling system as claimed in claim 6 further comprising a cooling side control system to control said controllable coolth dump responsive to a sensed temperature of coolant within said cooling circuit to control a heat input to said evaporator provided by said coolant.

8. A combined heating and cooling system as claimed in claim 6 wherein said cooling circuit has a coolth output and is controllably reconfigurable between a parallel mode in which said controllable coolth dump is coupled in parallel with said coolth output and a series mode in which said controllable coolth dump is coupled in series with said coolth output.

9. A combined heating and cooling system as claimed in claim 8 further comprising a cooling side control system to control said controllable coolth dump responsive to a sensed temperature of coolant within said cooling circuit to control a heat input to said evaporator provided by said coolant, and wherein said cooling side control system is configured to control said cooling circuit between said series mode and said parallel mode dependent upon an ambient temperature of said controllable coolth dump.

10. A combined heating and cooling system as claimed in claim 1 further comprising a solar thermal energy capture system coupled to said stratified thermal storage tank; and/or wherein said working fluid comprises carbon dioxide.

11. A method of controlling a combined heating and cooling system as claimed in claim 1, the method comprising:

determining one or both of a stored heat in said heating circuit and a stored coolth in said cooling circuit; and
controlling one or both of said compressor and said coupling of said stratified thermal storage tank to said heating circuit responsive to said determination of stored heat/coolth to maintain one or both of said stored heat and said stored coolth in a steady state, more particularly within a respective target range.

12. A method of controlling a combined heating and cooling system as claimed in claim 1, the method comprising:

determining a temperature of coolant circulating in said cooling circuit; and
controlling dumping of coolth from said cooling circuit responsive to said coolant temperature to control a heat input to said evaporator provided by said coolant.

13. A carbon dioxide-based combined heating and cooling system, the system comprising:

a working fluid circuit comprising a compressor, a gas cooler and an evaporator;
a heating circuit, thermally coupled to said working fluid circuit via said gas cooler and having heat output; and
a cooling circuit, thermally coupled to said working fluid circuit via said evaporator and having a coolth output; and
a first control system to control said working fluid circuit responsive to one or both of heat stored in said heating circuit and coolth stored in said cooling circuit to partly satisfy heat and coolth demands from respective said heat and coolth outputs with stored heat and/or coolth.

14. A carbon dioxide-based combined heating and cooling system as claimed in claim 13 further comprising a second control system to control said cooling side circuit to control temperature of coolant in said cooling side circuit flowing into said evaporator to control an efficiency of said working fluid circuit.

15. A carbon dioxide-based combined heating and cooling system as claimed in claim 14 wherein said second control system is configured to control said coolant temperature up towards a target temperature.

16. A carbon dioxide-based combined heating and cooling system as claimed in claim 15 wherein said second control system is further configured to control said coolant temperature down towards said target temperature to achieve a target cooling effect.

17. A carbon dioxide-based combined heating and cooling system as claimed in claim 15 wherein said second control system is further configured to control a controllable coolth dump in said cooling circuit to control said coolant temperature.

18. A carbon dioxide-based combined heating and cooling system as claimed in claim 17 wherein said second control system is further configured to control said cooling circuit to selectively couple said coolth dump in series and in parallel with said coolth output responsive to a sensed ambient temperature.

19. A carbon dioxide-based combined heating and cooling system in as claimed claim 13 further comprising one or both of a stored heat gauge and a stored coolth gauge, wherein said stored heat gauge and a stored coolth gauge are configured to determine stored heat and stored coolth dependent upon one or more sensed temperatures in said heating circuit and said cooling circuit respectively, and where said first control system is responsive to said stored heat gauge and/or said stored coolth gauge to satisfy said heat and coolth demands.

20. A carbon dioxide-based combined heating and cooling system as claimed in claim 13 wherein said heating circuit further comprises a controllable heat dump and wherein said second control system is configured to control said heat dump to control said stored heat to within a target range, to control a combined heating and cooling effectiveness of said heating and cooling system.

21. A method of controlling a carbon dioxide-based combined heating and cooling system, the combined heating and cooling system comprising a heating circuit and a cooling circuit each coupled to a shared carbon dioxide-based working fluid circuit, the combined heating and cooling system further comprising first and second heat exchangers to dump, respectively, heat and coolth from the working fluid of the system, the method comprising controlling the system in accordance with the following table: Greater work Dump heat from the Dump coolth from heating or cooling working fluid the working fluid Heating No Yes Cooling Yes No

22. A method as claimed in claim 21 further comprising controlling said coolth dumping responsive to a sensed temperature of said working fluid circuit.

23. A method as claimed in claim 21 further comprising controlling a rate of said coolth dumping by controlling a rate of operation of said second heat exchanger.

24. A method as claimed in claim 21 wherein said working fluid circuit comprises a compressor, a gas cooler and an evaporator, the method further comprising controlling said compressor responsive to one or both of a measure of stored heat and a measure of stored coolth in said system.

25. A method as claimed in claim 21 further comprising controlling dumping of heat from the working fluid of the system responsive to one or both of a measure of stored heat and a measure of stored coolth in said system

26. A method as claimed in claim 21 further comprising selectively coupling said first heat exchanger into said cooling circuit in parallel or in series dependent upon an ambient temperature of said first heat exchanger.

27. A method as claimed in claim 21, used for providing combined heating and cooling for commercial premises, the method comprising: using said cooling circuit to cool a merchandising refrigerator in the commercial premises; and

providing a combined heating and cooling system comprising a heating circuit and a cooling circuit each coupled to a shared carbon dioxide-based working fluid circuit, the combined heating and cooling system further comprising first and second heat exchangers to dump, respectively, heat and coolth from the working fluid of the system;
using said heating circuit to heat the commercial premises;
controlling said carbon dioxide-based combined heating and cooling system using the method of claim 21.
Patent History
Publication number: 20170010051
Type: Application
Filed: Jul 8, 2016
Publication Date: Jan 12, 2017
Inventor: STEVE CONNOLLY (CAMBRIDGE)
Application Number: 15/205,567
Classifications
International Classification: F28D 20/00 (20060101); F28F 27/02 (20060101); F25B 49/02 (20060101); F25B 9/00 (20060101); F25B 30/02 (20060101);