HEAT PUMPS AND FLUID PUMPS THEREFOR

Abstract

A fluid pump for pumping a fluid. One or more piston-cylinder arrangements each include a respective cylinder portion, a respective head portion and a respective piston portion together defining a respective pumping chamber. One or more connecting arrangements connect a crank member to the one or more piston-cylinder arrangements to drive the respective piston portion of each of the one or more piston-cylinder arrangements. A housing and the respective piston portion of each of the one or more piston-cylinder arrangements together house the crank member in an interior. The interior is sealed to capture blow-by. A transmission is arranged to transmit power, for rotating the crank member, into the housing magnetically, electrically or both.

Claims

1. An apparatus comprising: a fluid pump, the fluid pump including: one or more piston-cylinder arrangements each including a respective piston portion and a respective pumping chamber; a crank member; a housing; and a transmission; the housing together with the respective piston portion of each of the one or more piston-cylinder arrangements defining an interior in which the crank member is housed; the interior being sealed to capture blow-by; and the transmission being arranged to transmit power, for rotating the crank member, into the housing magnetically, electrically or both magnetically and electrically.

2. The apparatus of claim 1 wherein the transmission comprises a magnetic coupling including: a rotatable portion outside the housing; and a portion inside the housing magnetically co-operable with, to be rotated by, the rotatable portion.

3. The apparatus of claim 2 including an electric motor for driving the rotatable portion.

4. The apparatus of claim 1 including an electric motor having a stator and a rotor, which are located inside the housing.

5. The apparatus of claim 4 wherein the transmission includes at least one wire for the electric motor and which enters the housing via a sealed aperture.

6. The apparatus of claim 1 wherein the transmission includes an electric motor comprising: a stator outside of the housing; and a rotor inside the housing.

7. The apparatus of claim 1 including two or more of the piston-cylinder arrangements.

8. The apparatus of claim 7 including one or more connecting arrangements connecting the crank member to the respective piston portions, wherein the crank member and the one or more connecting arrangements are configured to move the respective piston portions in unison to hold substantially constant a volume of the interior.

9. The apparatus of claim 8 wherein the one or more connecting arrangements include a member having: two ends; and a respective one of the piston portions at each of the ends.

10. The apparatus of claim 1 comprising reed valves each including: a respective port; a respective ridge surrounding the respective port; and a respective petal arranged to bear against the respective ridge to close the respective port; wherein each respective one of the pumping chambers has a respective first of the reed valves defining one of an inlet to the respective pumping chamber and an outlet from the respective pumping chamber.

11. The apparatus of claim 10 wherein each respective ridge is shaped to contact a respectively corresponding respective petal along a respective line of contact not more than 0.5 mm thick.

12. The apparatus of claim 11 wherein each respective one of the pumping chambers has a respective second of the reed valves defining the other of the inlet to the respective pumping chamber and the outlet from the respective pumping chamber.

13. The apparatus of claim 10 wherein pistons have a combined displacement of not more than 10 cc (0.61 ci).

14. The apparatus of claim 1 wherein the apparatus is a heat pump and includes a fluid circuit; and the fluid pump is arranged to pump fluid about the fluid circuit.

15. The apparatus of claim 14 wherein the apparatus is a reverse cycle heat pump and includes a reversing valve arrangement; and the reversing valve arrangement includes: first-mode flow paths through which fluid flows when the pump is in the first mode of operation; second-mode flow paths through which fluid flows when the pump is in the second mode of operation; one or more valve elements; and one or more electromechanical drives; and the one or more electromechanical drives are arranged to electromechanically drive the one or more valve elements: from one or more respective first-mode positions at which the one or more valve elements occlude the second-mode flow paths; to one or more respective second-mode positions at which the one or more valve elements occlude the first-mode flow paths; to reverse a direction of flow about the fluid circuit and thereby switch from the first operating mode to the second operating mode.

16. The apparatus of claim 14 configured for evaporation of the fluid when the heat pump is in at least one operating mode of the heat pump, and wherein: the fluid circuit includes a fluid circuit portion; the fluid circuit is configured for at least most of the evaporation to occur along the fluid circuit portion; and the fluid circuit portion is arranged for the fluid to flow upwardly, or at least horizontally, along substantially all of the fluid circuit portion when the heat pump is in the at least one operating mode of the heat pump.

17. The apparatus of claim 16 wherein the fluid circuit is configured for at least 90% of the evaporation to occur along the fluid circuit portion.

18. The apparatus of claim 16 wherein the fluid circuit portion includes a serpentine conduit portion along which the fluid flows.

19. The apparatus of claim 14 including: a sensor to sense at least one of pressure or temperature, of the fluid on a high-pressure side of the fluid circuit; and a control arrangement configured to vary a speed of the fluid pump in response to the sensor and in positive relation to the at least one of the pressure or temperature.

20. The apparatus of claim 19 wherein the fluid circuit comprises: a downstream heat exchanger downstream of a fluid pump; and a refrigerant-cooling fan for driving air through the downstream heat exchanger; and the control arrangement is configured to vary a speed of the refrigerant-cooling fan in response to the sensor and in positive relation to the at least one of the pressure or temperature.

21. The apparatus of claim 19 wherein: the fluid circuit comprises an upstream heat exchanger upstream of the fluid pump; and the apparatus comprises: a refrigerant-heating fan for driving air through the upstream heat exchanger; and a temperature sensor for providing an indication of a temperature of the material to be cooled by the heat pump; and the control arrangement is configured to vary a speed of the refrigerant-heating fan in response to the temperature sensor and in positive relation to the temperature of the material to be cooled by the heat pump.

Description

BRIEF DESCRIPTION OF DRAWINGS

[0078] FIG. 1 schematically illustrates a heat pump;

[0079] FIG. 2 is a perspective view of a portion of a compressor;

[0080] FIG. 3 is a cross-section view of the portion of FIG. 2;

[0081] FIG. 4 schematically illustrates a reversing valve in a first operating mode;

[0082] FIG. 5 schematically illustrate the reversing valve in a second operating mode;

[0083] FIG. 6 is a side view of a valve plate;

[0084] FIG. 7 is a cross-section view of a magnetic coupling;

[0085] FIG. 8 is a side view of a petal;

[0086] FIG. 9 is a cross-section view of a pair of ports;

[0087] FIG. 10 is a front view of a connection arrangement;

[0088] FIG. 11 is a top view of the connection arrangement;

[0089] FIG. 12 is a rear view of a pump-and-valve arrangement;

[0090] FIG. 13 is a flow chart illustrating a start-up cycle;

[0091] FIG. 14 charts cabin temperature over time;

[0092] FIG. 15 charts the speed of selected components;

[0093] FIG. 16 illustrates an evaporator of the prior art; and

[0094] FIG. 17 illustrates an evaporator according to an exemplary embodiment of the present application.

DESCRIPTION OF EMBODIMENTS

[0095] The inventor's earlier compressor 5 includes a mechanical transmission for transmitting power into the housing 13. Shaft power is conveyed by the input shaft 31. The sealing arrangement 39 incorporates a lip seal acting on the shaft 31 to prevent the escape of refrigerant from the interior 41 to atmosphere. The inventor's studies have shown that: [0096] a) pressure within the volume 41 can exceed 600 kPa (87 psi) when the compressor 5 is working hard on a hot day; [0097] b) the seal 39 cannot reliably seal against these pressures and as such refrigerant is sometimes lost to atmosphere; [0098] c) power is lost to the friction between the seal 39 and the shaft 31; and [0099] d) the friction and the power loss increase in positive relation to the pressure within the volume 41.

[0100] Accordingly, a compressor having a magnetic transmission in the form of the magnetic coupling 49 in place of a mechanical transmission is proposed. The magnetic coupling 49 incorporates an external rotor 51 and an internal rotor 53 separated by housing portion 55. The housing portion 55 includes a radial mounting flange carrying an array of bolt holes and by which the housing portion 55 is mounted and sealingly engaged with a front of the body 15. The housing portion 55 defines a closed cup including a cylindrical wall 55b extending from the flange 55a, and an end face 55c.

[0101] The external rotor 51 is shaped to embrace the cup 55b, 55c and includes magnets 51a positioned to revolve about and in close proximity to the cylindrical wall 55b. The internal rotor 53 includes magnets 53a positioned to revolve within and in close proximity to the cylindrical wall 55b.

[0102] The internal rotor 53 forms part of a crank member 29′ (FIGS. 10 and 11). The external rotor 51 is connected to a source of shaft power. This source is preferably an electric motor. The magnets 51a, 53a are magnetically co-operable such that the magnets 53a follow the magnets 51a as a result of the magnetism operating through the housing portion 55. As such shaft power is magnetically transmitted into the housing 13 (from the rotor 51 to 53) without friction losses or the risk of leakage associated with a seal akin to the seal 39.

[0103] The coupling 49 includes both external magnets 51a and internal magnets 53a. Other variants are possible. By way of example, the magnets 51a might be replaced with steel (or other material) magnetically co-operable with the magnets 53a. For the avoidance of doubt, ‘magnetically co-operable material’ and similar terms as used herein take in both magnets and non-magnets.

[0104] The coupling 49 incorporates rotationally-driven external magnets 51a to create a rotating magnetic field by which the internal rotor 53 is rotated. In another variant of the compressor, the external rotor 51 is replaced by a stator co-operable with the rotor 53 to form an electric motor whereby the stator electrically creates a rotating magnetic field by which power is magnetically transmitted into the housing 13.

[0105] Another possibility is to place the stator inside the housing 13, in which case a magnetic transmission in the form of a wireless power transfer may be employed to transmit power into the housing 13. Alternatively, an electrical transmission in the form of simple wires (or other conductor(s)) may be employed. Of course, sealing about one or more wires is simpler than sealing about the input shaft 31 and does not entail the friction losses associated with the seal 39.

[0106] The improved sealing offered by an electrical and/or magnetic transmission helps to minimise the weight of the compressor. Some existing compressors are mounted within a sealed canister to capture any refrigerant leaking therefrom. The improved sealing allows for the elimination of the canister and its weight. It also contributes to the reliability of the heat pump, particularly on hot days when the pump is working hardest. On hot days, refrigerant pressure can built up within the canister of canister-sealed systems such that refrigerant is effectively lost from the fluid circuit, reducing the performance and reliability of the system.

[0107] The addition of a sealant, such as LOCTITE™, to the metal-to-metal interfaces between the body 15, plate 25b and head portion 25 is another possible precaution against leakage.

[0108] Preferred forms of the compressor, including the electric drive motor, weigh less than 6 kg. To the inventor's knowledge, this is much lighter than commonly available compressors delivering comparable output.

[0109] FIGS. 10 and 11 illustrate a preferred connection arrangement 57 by which the pistons 23′ are driven. The connection arrangement 57 is a form of Scotch Yoke including a single connecting member 59 having a respective one of the two pistons 23′ at each of its ends.

[0110] For the avoidance of doubt, as various terminology is used herein: [0111] ‘member’ and similar terms refer to a single unitary body which may be integrally formed or made up of separate, rigidly-connected, bodies of material; [0112] ‘integrally formed’ refers to a member formed of a single continuous body of material; and [0113] components may be integrated by welding but not by typical mechanical fastening techniques.

[0114] One advantage of the connection arrangement 57 over the connection arrangement 45 is that it enables the cylinders 21′ to be moved into coaxial alignment to enable the sleeve-receiving bores to be machined in a single operation from one end of the block 15. This reduces manufacturing costs. The block 15 may be formed by metal injection moulding. The co-axial alignment is advantageous in this context because it enables the sleeve-receiving bores to be cored from one end of the block.

[0115] Whilst, in principle, the member 59 and pistons 23′ could be parts of a common integrally formed member, preferably they are separate members. In this example, each piston 23′ is connected to the connecting member 59 via a respective gudgeon pin. Whilst there is no need for the gudgeon pin to work through a range of angles associated with the movement of a conventional conrod, a very small degree of movement is desirable in that it enables each of the pistons 23′ to self-align within its respective cylinder bore 21′ to account for any misalignment between the bores 21′ as a result of manufacturing tolerances and/or deformations in use resulting from thermal and/or mechanical stresses.

[0116] Separately forming the components also enables different materials to be used. The pistons 23′ and cylinders 21′ are preferably formed of substantially the same material, or at least of materials selected to have substantially identical thermal coefficients, so as to expand at the same rate as the components heat up in operation.

[0117] In this case, the components 19′, 23′ are formed of cast iron selected for its self-lubricating properties. The cylindrical surfaces of these components are ground to form accurate finishes to provide a satisfactory compromise between sealing and friction without the need for piston rings or any similar sealing arrangements. In one variant, the cylinder is ground to a diameter in the range of 16.002 mm to 16.005 mm (0.63 in to 0.63012 in) and to a surface roughness not greater than 0.8. The exterior of the piston is ground to a diameter in the range of 15.997 mm to 16.000 mm (0.62980 in to 0.62992 in) and within the same roughness tolerance.

[0118] The connecting member 59 has at its centre an oval-shaped aperture 61 in which the pin 35′ is mounted. The pin 35′ is mounted eccentrically (relative to the axis about which the crank member 29′ rotates) to describe a circular path. Whilst following the circular path, the pin 35′ races around the periphery of the oval-shaped opening 61 to drive the member 59 and pistons 23′ connected thereto to reciprocate.

[0119] The shape of the hole 61 and the shape and eccentricity of the pin 35′ may be varied to vary the stroke and acceleration curves of the pistons. The crank members 29, 29′ and connection arrangements 45, 57 are but two examples of a range of possibilities. ‘Crank member’ is used herein to refer to a rotationally-driven member having at least one eccentric portion by which the piston(s) is/are driven to reciprocate. By way of example, the crank member could be a cam and the connection arrangement could be a cam-following arrangement.

[0120] In each of the described arrangements, the crank member and the connection arrangement co-operate to drive the pistons in unison so as to hold substantially constant the internal volume 41. Other arrangements are possible. By way of example, with reference to FIG. 3, the crank 29 may be modified to define a pair of connecting rod journals 180° apart, whereby the pistons 23 are driven in opposition to each other so that (with the addition of a suitable inlet and outlet) the volume 41 becomes another pumping chamber. For the avoidance of doubt, such a pumping chamber is ‘sealed to capture blow-by’ as those words and similar words are used herein.

[0121] The present inventor has recognised that most refrigerants have lubricating properties adequate to lubricate the pistons and cylinders and provide some resistance to blow-by. Blow-by into the interior 41 provides sufficient lubrication for the crank member 29′ to operate without additional lubricating oils. Accordingly, preferred forms of the compressor 5 operate without an additional volume of oil within the housing 13 and as such losses associated with moving this oil are avoided. R134A and HR30 are the preferred refrigerants.

[0122] The inventor has also found room to significantly improve the efficiency of the reed valves. In the inventor's earlier variants, the petals 46 sealed against the planar face of the valve plate 25b in conventional fashion. The inventor's studies suggested that the petals 46 adhere to the planar surface and that, with each stroke of the compressor, work is required to break the petals away from this adherence. The studies also suggest that some petals twisted along their long axes rather than fully lifting away from their ports and therefore imposed some restriction to flow.

[0123] To address this issue, the modified valve seats illustrated in FIG. 9 have been developed. FIG. 9 illustrates a cross-section view through the inlet and outlet ports. Each of the ports is surrounded by a respective ridge 63 to present the petal 46 with an edge to seal against in place of a planar surface. Desirably, the contact area (between the petal and the ridge 63) is not more than 20% of the area of the port.

[0124] The valve plate 25b is formed of 3 mm thick stainless steel and the ports 25c′, 25d′ have finished internal diameters of 6 mm. The plate is laser-cut. The holes for the ports are initially cut undersized and then subject to a drilling operation. The drill is operated at a slow rpm and rapid rate of advance whereby consistent burrs respectively surrounding each end of each hole are formed. This drilling operation takes the holes to their finished size(s). The resultant downstream burr is then lapped to define a consistent sealing edge for the petal 46.

[0125] The burr is then removed from the upstream edge of the outlet port 25c′ so as not to interfere with the petal 46 for the inlet port 25d′. The burr may also be removed from the upstream side of the inlet port 25d′.

[0126] Another possibility, instead of the described drilling and lapping operations, is to fit the plate with suitable ridge-defining inserts. This may be preferable for mass production. Metal injection moulding is another possibility for the valve plate. Optionally, the apexes of the ridges may be subject to a post-moulding material-removal operation (e.g. lapping).

[0127] The inventor's testing identified the petal 46 as a potential failure point and it has been discovered that this can be addressed by taking advantage of the anisotropic properties of the spring steel. Early failures of the petal were associated with cracking in the vicinity of the root portion of the petal. Preferably the petals are cut so that their long axis (which is perpendicular to the axis of bending) is at least approximately parallel to the axis along which the steel is rolled. The axis along which the steel is rolled is believed to be coincident with the orientation of the steel's grain structure.

[0128] Testing suggests that the disclosed sealing ridges also improve the reliability of the petals by at least reducing torsional loads on the petals associated with twisting about their long axes.

[0129] The inventor's early testing has focused on two variants of the compressor both of which have a stroke of 8 mm (0.31 in). One variant has a bore of 16 mm (0.63 in) corresponding to a displacement of 3.2 cc (0.20 ci) for the compressor. The other variant has a bore of 17.5 mm (0.69 in) corresponding to a displacement of 3.9 cc (0.24 ci). Both variants are configured to operate at about 2,100 rpm.

[0130] The pump 5 is an example of a positive displacement pump. Other positive displacement pumps are possible. Indeed, some aspects of the invention may be implemented with non-positive displacement pumps such as centrifugal pumps.

[0131] The inventor has discovered that conventional reversing valves (such as the reversing valve 47) are not reliable when used with small compressors (say up to 10 cc). Practical design considerations dictate a less than perfect seal between the pistons 51a, 51b and the interior of the housing 49 such that the valve does not reliably change mode when driven by such a small compressor. By way of example, with reference to FIG. 4, when the solenoid is switched to pressurise the chamber 51c, the high-pressure fluid leaks past the valve 51a without building sufficient pressure in the chamber 51c to move the shuttle 51. This problem came as a surprise to the inventor.

[0132] To address this non-obvious problem, the inventor proposes to replace the refrigerant-driven valve element 51 with one or more electromechanically-driven valve elements, e.g. a solenoid could be added to the valve 47 to electromechanically drive the shuttle 51. This would allow the small three-way solenoid and the associated capillary tubes to be eliminated.

[0133] FIG. 12 illustrates another implementation of this aspect of the invention. FIG. 12 is a rear perspective view of a pump-and-valve unit 65 including the compressor 5 and a valve arrangement 67.

[0134] The valve arrangement 67 incorporates ports 5.sub.in′, 5.sub.out, 7″ and 11″. The ports 5.sub.in′, 5.sub.out′ are respectively connected to the compressor's inlet and the outlet. Plumbing 69 in the form of copper tubing defines a T-piece connecting the port 5.sub.out′ to each of the ports 7″, 11″. One arm of the T-piece carries a first first-mode solenoid S.sub.FM1 which, when open, defines a first first-mode flow path akin to the flow path FP.sub.FM1 and mutually connecting the ports 5.sub.out, 7″. The other arm of the T-piece carries a first second-mode solenoid S.sub.SM1 which, when open, defines a first second-mode flow path akin to the flow path FP.sub.SM1 and mutually connecting the ports 5.sub.out′, 7″.

[0135] The unit 65 further includes further plumbing 71 defining another T-piece equipped with solenoid valves S.sub.FM2, S.sub.SM2 for alternately connecting the port 5.sub.in′ with one or the other of the ports 7″, 11″.

[0136] This arrangement advantageously thermally separates the first-mode flow paths from each other and the second-mode flow paths from each other, to avoid heat exchange therebetween (as may occur within the valve 47) to lead to further efficiency gains.

[0137] In operation, only a diagonally-opposed two of the four solenoid valves are open at any one time. All four valves change state to change the operating mode. By way of example, to switch from the first operating mode to the second operating mode, the first-mode solenoid valves S.sub.FM1, S.sub.FM2 are closed and the second-mode solenoid valves S.sub.SM1, S.sub.SM2 are opened.

[0138] The valve arrangement 67 incorporates four two-way solenoid valves. Another possibility entails one or more three-way valves. Any adjacent two of the four two-way solenoid valves could be replaced by a suitable three-way valve.

[0139] The described arrangements of one or more electromechanically-driven valve elements enables smaller than conventional heat pumps to reliably reverse cycle.

[0140] The solenoid valves are preferably 12V solenoid valves to operate from the same power supply as is the compressor. For the avoidance of doubt, the movable fluid-contacting portion of a conventional solenoid valve is an electromechanically-driven valve element as those words and similar words are used herein.

[0141] The present inventor has also discovered that, in the context of extraordinarily small heat pumps, the evaporator can be a source of inefficiency. FIG. 17 illustrates a preferred form of evaporator 311 including an inlet 313, an outlet 315, and a fluid path portion 317 connecting the inlet to the outlet. The evaporator 311 is essentially the evaporator 211 turned on its side so that the vertical runs 219 become horizontal runs 319, and downward bends 221a become upward bends 321a. The inventor's tests suggest that dramatic improvements in efficiency can be realised through this simple reorientation and that this efficiency gain is achieved by eliminating the gas-traps defined by the downward portions of the fluid path 217.

[0142] The inventor has recognised that as the refrigerant evaporates along the path 217 it becomes relatively buoyant, and that along the downward runs 219b this buoyancy provides a resistance against which the compressor must work.

[0143] In the heat exchanger 311 the buoyancy of the vapor urges the refrigerant to flow in the same direction as it is urged to flow by the fluid pump 5. More specifically: [0144] through the upward portions of the bends 317 the buoyancy of the vapor urges the refrigerant to flow in the same direction as it is urged by the pump 5; and [0145] along the horizontal runs 319 the buoyancy at least does not resist the urging of the fluid pump 5.

[0146] In the illustrated example, a single serpentine path connecting the inlet 313 to the outlet 315 is disclosed. Other options are possible. By way of example, the relevant portion of the fluid circuit may be made up of two parallel flow path portions.

[0147] Preferably all mixed-phase portions of the low-pressure side of the fluid circuit are upward or at least horizontal. Likewise, preferably there are no downward portions in the fluid path portion within the evaporator, although it is possible that a downward run of conduit might be added to a single-phase portion of the flow path within the evaporator, e.g. for more convenient connection to the expansion valve or the compressor, without a dramatic reduction in efficiency.

[0148] The inventor's studies into the reliability of heat pumps have also revealed that many failures occur at start-up as opposed to when the heat pump is pumping heat. Many of these failures have been found to relate to the torque necessary to start the compressor. In an inactive heat pump, significant pressure can build in the fluid circuit, including in the pumping chamber, such that the torque required to move the piston is more than the electric motor is capable of, e.g. the current draw from the motor may exceed an enforced current limit, e.g. may blow a fuse. As such, automotive air conditioners are notorious for being least reliable when they are most needed, i.e. on the hottest days.

[0149] The present inventor's adoption of two smaller pistons (having their compression and suction strokes out of phase to each other) as opposed to a single larger piston is a first step to addressing the start-up problems. Having a variable-speed drive unit (e.g. electric motor) dedicated to the compressor is another step in the right direction and is in stark contrast to the conventional approach of driving a compressor by selective engagement with a vehicle's combustion engine. The dedicated drive allows the compressor to be driven at a speed that is most efficient for the particular operating conditions rather than to be driven at a speed corresponding to the combustion engine, which may well lead to sub-optimal operation of the compressor and the heat pump.

[0150] Providing a dedicated variable-speed drive allows for the convenient implementation of new control strategies to improve the performance of the compressor and/or heat pump and to address the risk of failure at start-up.

[0151] The present inventor has recognised that the start-up problems can be largely avoided by operating the compressor at an idle speed rather than allowing the compressor to shut down completely (e.g. at the whim of a thermostatic controller). The inventor's prototypes have been tested at an idle speed of 1,000 rpm at which their electric motors each draw about 5 amps at about 12 volts (i.e. at about 60 watt). When the compressor is operated at this speed, the heat pump 1 provides negligible heating performance. The fans 7a, 11a may be deactivated or at least slowed or operated on a low-duty cycle.

[0152] Whilst the about 60 watt idle-speed power draw results in some waste relative to simply deactivating the system, the inventor has found that under typical automotive operating conditions this waste is more than offset by the other efficiencies of the disclosed methods and apparatus.

[0153] The 60 watt/1000 rpm idle speed is achieved using Allied Motion's PJ2P021Q electric motor. Of course, other electric motors (and indeed other drive arrangements more generally) are possible. Direct current 12V, 36V or 48V drives are preferred to suit automotive applications.

[0154] The inventor's studies have shown that many existing 12V vehicle electrical systems are capable of supplying the compressor with 20 amps for a sustained period. Preferred variants of the disclosed heat pumps have a coefficient of performance of at least 4 whereby utilising the available 20 amps at 12V (240 watt) they are capable of delivering about 900 watts or more of heating and/or cooling within the 20 amp current limit.

[0155] The start-up problems are most significant on hot days during which the heat pump may well operate almost continuously. Accordingly, to suit some applications, it may be appropriate to enter the idle-speed mode based on an indication of the ambient temperature. By way of example: [0156] if the ambient temperature is above a predetermined threshold, when the heat pump is likely to be operating at least most of the time, the heat pump may enter the idle-speed mode if and when the pump is deactivated (e.g. deactivated by a thermostatic controller); and [0157] if the ambient temperature is below a predetermined threshold (and restarting is unlikely to be problematic), the heat pump may be deactivated in conventional fashion.

[0158] To implement these control strategies, the heat pump 1 may include a controller 71, an ambient temperature sensor 73, a cabin (or other target) temperature sensor 75 and a pressure sensor 77. The control arrangement 71 is configured to receive data from the sensors 73, 75, 77 and to send control signals to the components 5, 7a, 11a. The various components are preferably connected by suitable wired links, although wireless connections are also possible. Various of the components may be integrated and/or commonly housed.

[0159] FIG. 13 illustrates a preferred control logic implemented by the control arrangement 71. FIGS. 14 and 15 illustrate the results of that logic. Again, by way of example, the invention is described with reference to cooling the interior of a truck cabin, although of course the disclosed principles are applicable to other contexts. At step 101, the control process is initiated, e.g. upon ignition of the truck's combustion engine. At step 103, the output of the sensor 75 is checked to determine whether the sensed temperature is above a threshold, e.g. 45° C. If not, the control arrangement 71 moves on to pressure control mode at step 105.

[0160] On hotter days, the control arrangement moves on to increment a counter at step 107, and then to check whether the counter is above the threshold corresponding to, for example, 15 minutes. Whilst the counter remains below the threshold, the controller moves on to operation in an overspeed mode at step 111.

[0161] In the overspeed mode, the compressor is operated faster than its components are intended for continuous operation, to maximise the cooling effects. For the same reason, the condenser fan 7a is also operated at about 20% above its typical operating speed. The internal fan is also operated at about 10% more than its peak efficiency speed. The peak efficiency speed is the speed at which the air it moves rejects the most heat (as measured by power) to the refrigerant for a given temperature of the air being moved.

[0162] Overspeeding the fan serves to stir the air within the truck cabin. In particular, the hottest pools of air around the ceiling of the cabin, which can be as hot as 80° C. (176° F.) are moved. When the air is recirculated within the truck cabin, the displacement of these hot pools of air leads to hotter air moving through the evaporator 11 and thereby to improved cooling. The illustrated cycle continues to iterate until the end of the preset period for the overspeed mode, whereupon (at step 109) the control 71 moves on to the pressure control mode at step 105.

[0163] Other variants are possible. By way of example, between steps 107 and 111 the sensor 75 may be rechecked. The control arrangement may move to the pressure control mode if the cabin temperature drops below a threshold which may or may not be the same threshold as the overspeed mode entry threshold (at step 103).

[0164] The pressure control mode is so named because the speed of the condenser is controlled in positive relation to the pressure on the high-pressure side of the fluid circuit 3, e.g. in response to the sensor 77. This control arrangement advantageously leads to efficient steady-state operation of the heat pump, in stark contrast to conventional stop-start thermostatic control and sub-optimal compressor speed.

[0165] The pressure on the high-pressure side of the fluid circuit provides an indication of the load on the heat pump. Other indicators of load may be relied upon to control the heat pump.

[0166] For the avoidance of doubt, ‘in positive relation’ means ‘to increase with’. Over the domain of positive real numbers, y=ax, y=ax.sup.2, y=a.sup.x and y=log x are examples of positive relationships.

[0167] The speed of the external fan 7a is also controlled in positive relation to the pressure of the high-pressure refrigerant. In other examples of the invention, the air flow to a heat exchanger may be controlled in other ways, e.g. in the context of a truck, instead of accelerating the fan 7a, a vent may be opened to allow for more air movement as a result of movement of the truck.

[0168] The speed of the internal fan 11a is controlled in positive relation to the cabin temperature, e.g. controlled in positive relation to an amount that that temperature exceeds a set desired temperature. The desired temperature is 22° C. in this example. Preferably the speed of the fan 11a is also limited to the peak efficiency speed.

[0169] FIG. 14 illustrates the internal temperature of a truck that has been activated after it has been left stationary in the sun for some time. At start-up (i.e. at step 101), the sensor 75 indicates a temperature of 60° C. As described, the pump 1 (or more specifically its controller 71) enters the overspeed mode for a period of 15 minutes until step 109′ whereat the pump switches to the pressure control mode. During the overspeed mode, the temperature within the cabin decays at an exponential rate. Upon switching to the pressure control mode, the efficiency of cooling increases and as such there is a shallow knee in the graph. Thereafter the temperature in the cabin continues to decay at an exponential rate until about the 30 minute mark whereat the temperature approximates a steady 22° C. and the heat pump is continuously operating to provide efficient cooling matched to the heat entering the cabin from other sources.

[0170] The described heat pump is particularly advantageous in and for the context of vehicles wherein its compact size, light weight and reliability are particularly advantageous.

[0171] As noted, electrically-driven variants of the heat pump allow for the compressor to be operated at speeds independent of the crank speed of the vehicle. It also enables the heat pump to be moved away from the engine. The costly and often problematic drive belt for driving the compressor can be eliminated.

[0172] The conventional placement of automotive air conditioners towards the front of the engine bay presents other problems. Since the compressor is positioned to be driven off the engine's crank shaft, it typically follows that the condenser is positioned foremost in the engine bay in front of the engine's radiator. At this location, the condenser is typically in prime position to be damaged during minor road accidents. Vast quantities of refrigerant are lost to atmosphere in this way. This causes environmental damage. The damage to the air conditioner also adds to the cost of the crash repair. The adoption of an electrically-driven compressor enables the entire heat pump to be conveniently relocated to a safer position, e.g. towards the rear of the vehicle whereto suitable cooling air may be ducted.

[0173] The overall efficiency of the heat pump also lends itself to the ongoing operation of the heat pump whilst the vehicle's combustion engine is deactivated. In many parts of the world, it is not uncommon to see unoccupied vehicles idling simply to run the air conditioner. This is of course grossly inefficient and thus expensive. It also leads to unnecessary pollution.

[0174] Whilst the invention is particularly advantageous in the context of self-propelling land vehicles, it may also be advantageously applied elsewhere, such as in the context of yachts and other water-going vessels. In the context of some water-going vessels, using existing technologies, it is not uncommon for the vessel's large diesel engines to be idling simply to power the air conditioners. Of course, this is especially inefficient.

[0175] The application of the described heat pump, and in particular reverse-cycle variants of the disclosed heat pump, to heat the occupant-carrying interior of a vehicle is particularly advantageous. At present, many vehicles have essentially separate heating and cooling systems. The heating systems have a ‘water core’ through which heat is transferred from the water of the combustion engine to air to heat the interior. Of course, these cores entail some cost, some weight and some risk of failure. Advantageously these cores (and the associated fans, etc) can be eliminated by adopting the disclosed heat pumps. As described, the same heat exchanger that cools the cabin can be employed to heat it when the vehicle is exposed to colder weather.

[0176] Early variants of the compressor 5 were developed to provide air conditioning to the helmets of race-car drivers racing at an elite level whereat high performance, minimum weight and absolute reliability are essential under very challenging conditions. As such, the compressor 5 has been proven to inertial loadings up to 8G. With the continued development of the concept disclosed herein, the pump's performance continues to surprise and impress. A test of a variant of the heat pump powered by a N70ZZ battery and a 250 watt solar panel to cool a 5.5 m (18 ft) caravan was highly successful.

[0177] The integration of the described heat pump with battery solar systems is particularly advantageous. In such systems, the efficiency of the disclosed heat pump is a clear advantage. The improved start-up characteristics of the compressor are also important. Many such systems which cannot supply the electrical current necessary to start a conventional heat pump of similar scale (or can only do so when their batteries are in good condition and fully charged) can reliably start up the disclosed heat pump over a wide range of battery conditions. This is particularly advantageous in the context of water-going vessels and dwellings that are not connected to a mains power supply.

[0178] Some vehicles may be equipped with two or more of the disclosed heat pumps. By way of example, a limousine may be fitted with separate heat pumps for the driver and each of two or more passengers in the rear of the vehicle to enable each of the occupants to have separately controlled heating and cooling.

[0179] Optionally, a vehicle may be connected to mains power whilst it is stationary, to operate the heat pump as required. By way of example, in colder climates the heat pump may be operated overnight to provide a comfortable interior for the driver in the morning.

[0180] Various examples have been described. The invention is not limited to these examples. Rather, the invention is defined by the claims. By way of example, the disclosed heat pump may be employed for heating substances other than air. Likewise, whilst a fluid pump for pumping fluid about the fluid circuit of a heat pump has been described, the disclosed fluid pumps may find application elsewhere. In particular, the compressor 5 is particularly advantageous when used as a vacuum pump to evacuate the fluid circuit of a heat pump.

[0181] For the avoidance of doubt, a typical combustion engine is an example of a fluid pump, driven by the release of chemical energy from fuel, as ‘fluid pump’ and similar terminology is used herein albeit that preferred forms of the disclosed fluid pumps are driven by sources of more organized power, such as sources of shaft or electrical power.

[0182] The term ‘comprises’ and its grammatical variants has a meaning that is determined by the context in which it appears. Accordingly, the term should not be interpreted exhaustively unless the context dictates so.