FIELD

The present invention relates to gas compressors and inlet and discharge valve systems for such.

BACKGROUND

In various machines such as heat pumps, compressors are employed. Many heat pumps have a gas/vapor compressor as a core component. More specifically, for the vapor compression cycle, which is commonly used for, among other, residential heat pumps, refrigerators and air conditioning systems in cars, as well as for industrial heat pumps, the compressor is a central component. The main advantages of using heat pumps as opposed to conventional boiler or furnace systems, is that heat pumps generate several more times the heat than the power which is required to drive them, thereby increasing energy efficiency, and in many cases also improving economics of operation. In addition, for cases where heat pumps can effectively replace fossil- fuelled heat sources, a substantial reduction in climate gas emissions can also be achieved. As reduction in climate gas emissions becomes more and more important, the need for new and further heat pump technologies, which can cover new and further applications, and especially applications within different industries, the need for new and better compressor types is also increasing. There are two main classes of compressor devices used for heat pumps: flow devices (often implemented as centrifugal compressors), and positive displacement devices, with the latter being available in a great variety of different devices based on several different geometrical principles, for example: screw compressors, scroll compressors, vane compressors and reciprocating/piston compressors. Each type has its own advantages and disadvantages, which will not be described in further detail here.

Common to all compressor types, is that they operate on a working fluid, also commonly referred to as a refrigerant, or simply gas. In some examples, the working fluid may be a condensable gas. In some examples the working fluid never changes phase, and hence the working fluid is only operated in a gaseous state. Throughout the present disclosure, any working fluid in any phase, including but not limited to: partly liquid; gaseous; and supercritical is intended when referring to the working fluid.

In the example described herein, the compressor is of the piston-type (also commonly referred to as a reciprocating compressor). However, it should be noted that the present disclosure is not limited to only piston-type compressors, and that it may be equally relevant and useful to other types of compressors.

The performance of heat pump compressors, and more specifically gas/vapor compressors, is mainly a result of the three characteristics:

1 . Mechanical efficiency

2. Volumetric efficiency 3. Isentropic efficiency

Mechanical efficiency is primarily a consequence of internal, mechanical friction, or rather, a lack thereof. Volumetric efficiency is primarily a consequence of the internal, so-called dead volume (also called clearance volume, in the compressor’s working chamber, i.e. the minimum achievable internal volume in the cylinder of a working chamber of a compressor during compression). Lastly, the isentropic efficiency is mainly a consequence of the effectiveness of the gas exchange processes, which include the suction process and the discharge process.

Beyond this, there are further factors that affect the aforementioned characteristics, such as thermal leakage/undesired heat exchange through internal surfaces, e.g., a cylinder wall, which affects the isentropic efficiency, and gas leakage past sealing elements etc., which affects the isentropic efficiency as well as the volumetric efficiency.

In the following, and to explain the principles of operation in an easier way, a reciprocating compressor with a conventional crankshaft is disclosed. In principle, the compressor operation is simple, and consists of four principal processes (or steps) that are performed cyclically. For piston compressors, these are performed once per revolution of the crankshaft:

1 . A suction process.

2. A compression process.

3. A discharge process.

4. A re-expansion process. The suction and discharge processes comprise the gas exchange processes, during which a working fluid, normally in its gaseous/vapor form, is either sucked into or discharged out of the cylinder. Again, this assumes a piston compressor, however, the very same processes also take place in all other positive displacement machines.) For piston compressors in general, the gas exchange processes are governed by the suction and discharge valves: the suction valve controls the inflow of new, uncompressed gas into the cylinder during the suction process, while the discharge valve controls the outflow of compressed gas during the discharge process.

Compressors can have one or more cylinders, and each cylinder has its corresponding set of suction and discharge valves, and the suction/discharge valves can each in turn also consist of multiple valves working in parallel, as is quite common. In conventional compressor designs, the suction/discharge valves are usually in the form of a so-called reed valve, which constitutes a simple, yet effective principal component for many compressor applications.

Reed valves in principle comprise at least a valve plate or a valve port section with a reed element (also called a reed valve blade) and a retainer (sometimes also called a stop plate). The reed element is usually in the form of a thin metal sheet, which sometimes is also made from a spring material, with the reed element performing the actual opening and closing of the valve, by covering or uncovering port openings, slots or similar in the valve plate or valve port section. Sometimes an independent spring element is provided, which is made to continuously push and thus provides a force on the reed element in the closing direction. This is to aid in closing of the valve, and to prevent flow through the valve when it should otherwise be closed. The retainer can typically be a curved, relatively stiff sheet metal piece, which is shaped to let the reed element “roll” against its curved surface, as to limit the movement of the reed element and thus to guide it, and also to prevent damage that could otherwise be caused by excessive bending during operation. Other times the retainer is in the form of a retainer plate, which is fixed at a certain distance from the reed element, and often with a spring element in between. The principal function is the same, but the designs can vary. The fundamental characteristic of reed valves is that they are passively operated, meaning that they open and close only due to a pressure differential (or lack thereof) across the reed element in the opening direction.

Reed valves are compact and lightweight, their design is generally simple, and they provide for easy and affordable manufacturing. However, reed valves have some drawbacks for certain applications: It is difficult to design reed valves with as effective flow areas as for certain other valve types, and it is also difficult to design a compressor with a very low dead volume, especially when the suction valve is of the reed type. The consequence of this is that compressors equipped with reed valves usually have higher dead volumes and smaller effective flow areas than necessary. This in turn results in lower volumetric and isentropic efficiencies, respectively.

Moreover, and as mentioned above, for reed valves to properly close during operation, they are usually provided with a spring force in the closing direction, either by making the reed element itself out of a spring material, or by providing a separate spring element that pushes on the reed element.

Further, reed valves often suffer from so-called valve flutter under certain circumstances, during which the reed element will be exposed to excessive stress, sometimes leading to fatigue or other types of deformation with resulting damage, and in the worst-case catastrophic breakdown. This typically happens in certain frequency ranges, which for a compressor translates into specific speed ranges due to the cyclic operation. To overcome this, a stiffer spring or thicker reed material is sometimes used, and this leads to increased fluid flow resistance, which again leads to flow losses and corresponding reduction in the isentropic efficiency of the compressor.

Modem refrigeration and heat pump systems often use modern synthetic working fluids of the so-called 4 ^th generation. These typically encompass hydrofluoroolefins (HFOs) and hydrochlorofluoroolefins (HCFOs). One difference between these and other fluids is in that their densities are often high for typical conditions that are present in a refrigeration or heat pump system. Especially for industrial heat pumps in the high-temperature category, the relevant modem working fluids have densities during operation in a typical range from 30 - 100 kg/m3 or even higher under relevant operating conditions. Compared to other fluids, this is very high, and this also means that valve performance becomes even more critical in terms of maximizing efficiency and performance. Because of, among other, the high densities, the pressure drops across valves, ports, pipes and other in a heat pump system will usually increase, leading to increased losses, and hence there is a substantial potential performance gain is to be achieved if valve performance can be improved.

Poppet valves are extremely common, they are used in a great variety of applications (for example in combustion engines), and they have a long history and are widely proven and accepted by the industry. Also, there is widely available, low-cost manufacturing of poppet valves around the world.

Since poppet valves are usually actively controlled through mechanical means such as camshafts, cam followers, pushrods and rocker arms, they can be opened actively without causing substantial pressure drops as can otherwise apply to reed valves. Also, their opening and closing timing can be made completely independent from the differential pressures across them, leading to useful flexibility in controlling the gas exchange processes for a compressor.

Further, the flow characteristics of poppet valves are generally very good, primarily due to a geometry that provide little flow resistance, which also applies to denser fluids. They are used in a great number of different applications, spanning from heavy-duty diesel engines to hydraulic circuits, and there is vast empirical experience and knowledge about their workings, limits, and performance in general. Flow characteristics of poppet valves and other properties such as durability have been extensively proven for many years.

UK patent document GB2416196A discloses a valve control system for a reciprocating gas compressor having at least one suction valve between a gas supply and a compressor cylinder. The suction valve opens and closes automatically in response to pressure in the cylinder but has a locking assembly which holds it in the open position for a selected period after it has opened, in synchronisation with the piston strokes of the compressor. This allows gas to enter and exit the compressor via the suction valve during the cycle for which the valve is open so that heat is removed along with the gas thus allowing the cylinder chamber to cool. The valve may be a disk or reed valve and the locking assembly may be electromagnetic or pneumatic. The compressor may have a heat regenerator mounted within the compressor cylinder for enabling transfer of heat out of the cylinder.

US patent document US2019003604A1 discloses a valve assembly including a seat plate having a top surface and a bottom surface, a plurality of first valve modules arranged in a first level relative to the seat plate such that a first seating face of each of the plurality of first valve modules is substantially coplanar with a first plane that is substantially parallel to at least one of the top surface and the bottom surface of the seat plate, and at least one second valve module arranged in a second level relative to the seat plate such that a second seating face of the at least one second valve module is co-planar with a second plane that is substantially parallel to the first plane, wherein the second plane is offset from the first plane by a first distance.

US patent document US4370103A discloses a piston pump mechanism for use in positive pressure and vacuum pumping systems for pumping various gases such as air, freon, natural gas, etc. The pump mechanism incorporates an inlet and discharge valve system that substantially eliminates any unpurged or unswept volume of gas within a pump cylinder during each reciprocating stroke of the pump piston. The pump mechanism also incorporates a control spool that floats within a pump head and has a sealed flexible interconnection with the discharge valve that allows the discharge valve to seek optimum seating engagement about the entire periphery of the cylinder. The valve control spool may be pressure balanced or controllably unbalanced, as desired, for influence of the discharge valve. The spool also provides for inlet of gas to an inlet valve that may be incorporated into the discharge valve.

International patent document W02006115348 discloses intake and exhaust valves for a cylinder type air compressor. The intake valve includes a valve body having a plurality of intake ports, a movable shaft elastically installed in the valve body to open and close the intake ports, a packing for elastically supporting the outer surface of a valve head of the movable shaft when the movable shaft is moved upwards and downwards to reduce striking noise, and an elastic support element for elastically supporting the lower end of a stopper. The exhaust valve includes a plurality of exhaust ports defined in the valve body around the intake ports to be arranged in a circumferential arrangement centered around the movable shaft, and a ring-shaped movable opening and closing element elastically supported to open and close the plurality of exhaust ports.

UK patent document GB2396667A discloses a reciprocating gas compressor is described operating according to an extended cycle of 4, 6 or more strokes, wherein the first two strokes are sequential induction and compression strokes using a low pressure gas as working fluid and compressing it to a high pressure gas, and the remaining strokes are pairs of sequential filling and emptying strokes using more of the low pressure gas as heat transfer fluid for transferring heat from inside the gas compressor to outside the gas compressor. The gas compressor also contains a heat regenerator for absorbing heat from the compressed gas and releasing heat to the heat transfer fluid.

US patent document US 2003/030226 A1 relates to the use of elastomers with the sealing element of reciprocating gas compressor valves to increase the reliability of the gas tight seal within the reciprocating gas compressor valve and to increase the useful life of reciprocating gas compressor valve. The elastomeric material is either used as a coating layer on the sealing element of the reciprocating gas compressor valve, or as the entire sealing element. The elastomeric material acts as a cushion to reduce the wear on the sealing element, provides a superior gas tight seal, and is more tolerant of entrained dirt or liquids in the gas stream thereby increasing the operable life of the reciprocating gas compressor valve.

Chinese patent document CN 111022293A relates to a trans-critical reciprocating piston compressor taking CO2 as a refrigerant.

At least one aim of the present invention is to obviate or at least mitigate one or more drawbacks associated with the prior art.

SUMMARY

According to a first aspect of the invention, there is provided a reciprocating compressor comprising: a suction valve; a discharge valve; and a working chamber; wherein the suction valve is configured to provide fluid communication between a suction channel and the working chamber and the discharge valve is configured to provide fluid communication between the working chamber and a discharge channel; wherein the suction valve comprises at least one non-self-actuated valve and the discharge valve comprises at least one self-actuated check valve.

Embodiments of the present invention can be advantageous in that they may combine some of the best features of non-self-actuated valves and selfactuated valves to create an improved compressor. This may be particularly so when the compressor is in the form of heat pump and refrigeration compressors.

Embodiments of the invention have been based upon the inventors’ efforts in realising that the largest penalty in performance is often seen during the suction process of a compressor. During this process, the fluid flow direction is into the working chamber, and during compression the pressure differential is in the suction valve’s closing direction, and poppet valves are therefore considered a feasible alternative to reed valves. However, for the discharge process, poppet valves are not as suitable, since the direction of pressure is in the opposite direction to the valve’s closing direction, and hence greater measures are sought in various embodiments by the inventors to provide poppet discharge valves that are effective and can manage the pressure differential also when closed.

Furthermore, the inventors have considered that the need for higher valve performance is generally not as great for the discharge process as for the suction process, among other reasons because for reciprocating compressors, the discharge process happens when the piston is approaching top dead centre (TDC), and that is when the volumetric flow becomes lower due to the retardation and hence decreasing instantaneous speed of the piston(s), which again determines the volumetric fluid flow and flow velocity out of the discharge port. Therefore, embodiments using poppet valves or similar for the suction valve system and reed valves or similar for the discharge valve system can provide compromise with good and improved overall performance.

The self-actuated check valve may for example comprise one or more of: a reed valve; a plate-valve; self-actuated poppet valve; self-actuated ball check valve.

The non-self-actuated valve may for example be one or more of: a poppet valve; a rotary valve; a piston valve; a slide valve; a spool valve. The non-self- actuating valve may be one or more of: mechanically controlled; pneumatically controlled; electrically controlled; hydraulically controlled.

The reciprocating compressor may further comprise: a compressor block having at least one cylinder with a piston arranged to reciprocate therein; and at least one cylinder head; wherein the suction valve is arranged in the cylinder head.

The suction valve may be located on a longitudinal centre axis of the at least one cylinder.

The at least one cylinder head may comprise a valve plate.

The at least one cylinder may comprise a first cylinder with a first piston arranged to reciprocate therein and a second cylinder with a second piston arranged to reciprocate therein; the at least one cylinder head comprises a first cylinder head and a second cylinder head; and the first cylinder head is configured to mate with the first cylinder and the second cylinder head is configured to mate with the second cylinder.

The first and second cylinders may be arranged at around 90 degrees relative to each other to form a V-configuration.

The at least one discharge valve may be a plurality of self-actuated check valves.

The discharge valve may be arranged concentrically with the suction valve. The plurality of self-actuated check valves may each be concentric with the suction valve. The suction valve may be arranged such that the central axis of the suction valve is shared with the central axis of the at least one cylinder.

This can maximise the utilisation of the available valve area, and provide for a practical geometry. It should be noted that reed valves usually are limited in size due to the need to reduce their mass at the same time as requiring them to be flexible. Therefore, in reed-valve-based valve systems, many reed valves may be used in parallel. Reed valves may be conveniently configured in parallel. The valves may also conveniently be designed with a geometry that allows for circular patterns of arrangement, corresponding to the circular geometry of the compressor cylinder, and hence flow area utilisation can be maximized.

The discharge valve may be arranged such that the central axis of the discharge valve is shared with the central axis of the at least one cylinder.

The piston may be provided with at least one protrusion and the cylinder head is provided with at least one cylinder head recess, wherein the at least one protrusion and the at least one cylinder head recess are registered such that the protrusion will displace at least a portion of the volume of the cylinder head recess in use.

The at least a portion of the volume of the cylinder head recess may be at least 50% or at least 60%, or at least 70% of the recess volume, or a majority of the cylinder head recess volume.

The at least one protrusion may be a plurality of protrusions and the at least one cylinder head recess is a plurality of cylinder head recesses.

The piston may be provided with at least one piston recess which is shaped and configured to receive a non-self-actuated valve in use.

The piston recess may be configured such that in use the non-self- actuated valve will displace at least 50% or at least 60%, or at least 70% of the piston recess volume, or a majority of the piston recess volume.

This can provide an additional advantage in that the volumetric efficiency of the compressor can thus be further optimised. This advantage is not available when using passive, reed-valve-based valve systems only. The reason is that for suction valves using reed valves, the requirement for having a retainer plate or a stop plate makes it impossible to displace this portion of the dead volume, since the dead volume caused by the retainer/stop plates is hidden on their “back side”, away from the piston, and thus cannot be displaced.

The suction channel may comprise: a first portion formed in the compressor block; and a second portion formed in the at least one cylinder head; wherein the first and second portions are configured to mate at a connection to form a fluid passage between the first portion and the second portion for delivering working fluid to the working chamber in use. Since the suction channel will always see the lowest temperatures compared to the discharge channel, the suction channel can be designed to be part of a compressor block casting, and thereby additional pipes, sealings and flanges can be eliminated, as well having the option to form this part of the suction channel to one’s needs. A combination of good flow characteristics as well as reduced number of parts, including potentially leakage-prone sealings, can then be avoided.

The first portion of the suction channel may comprise a fluid entry path adjacent the connection when the first and second portions are mated in use, wherein the fluid entry path is configured to direct fluid at an angle of between 20 degrees and 80 degrees relative to a normal on a cylinder centre axis. The angle may be around 45 degrees.

By carefully designing the fluid entry path, an improved flow path can be achieved. This flow path is fluid-dynamically optimised and provides reduced pressure drops by promoting shallower turns in the fluid flow path.

The first portion may comprise a buffer reservoir. The buffer reservoir may have a substantially larger volume than the volume of the working chamber such that in use when working fluid is sucked out of the buffer reservoir to fill the working chamber, there is not a substantial pressure drop in the buffer reservoir.

The working chamber may be a first working chamber and the suction channel a first suction channel; wherein the compressor further comprises a second working chamber and a second suction channel; the first and second suction channels join to form a combined suction channel comprising an inlet for the compressor receiving working fluid therethrough.

This is typically advantageous because it reduces the number of parts in the compressor and importantly it reduces the number of sealing surfaces.

The inlet may comprise a manifold.

According to a second aspect of the invention, there is provided a heat pump comprising the reciprocating compressor according to the first aspect of the invention.

The heat pump may be a high-temperature heat pump capable in use of supplying output temperatures above 55 degrees Celsius or above 80 degrees Celsius or above 90 degrees Celsius or above 100 degrees Celsius or above 110 degrees Celsius or above 120 degrees Celsius or above 130 degrees Celsius or above 140 degrees Celsius or above 150 degrees Celsius or above 160 degrees Celsius or above 170 degrees Celsius or above 180 degrees Celsius or above 190 degrees Celsius or above 200 degrees Celsius or above 210 degrees Celsius or above 220 degrees Celsius or above 230 degrees Celsius or above 240 degrees Celsius or above 250 degrees Celsius.

According to a third aspect of the invention, there is provided use of a reciprocating compressor according to the first aspect of the invention in a heat pump.

According to a fourth aspect of the invention, there is provided a method of operating a reciprocating compressor according to the first aspect of the invention, comprising the steps of: sucking working fluid into the working chamber via the at least one non-self-actuated valve; and discharging working fluid from the working chamber via the at least one self-actuated valve.

According to a fifth aspect of the invention, there is provided a reciprocating compressor comprising: a suction valve; a discharge valve; and a working chamber; wherein the suction valve is configured to provide fluid communication between a suction channel and the working chamber and the discharge valve is configured to provide fluid communication between the working chamber and a discharge channel; the compressor further comprising: a compressor block having at least one cylinder with a piston arranged to reciprocate therein; and at least one cylinder head; wherein the suction valve is arranged in the cylinder head, wherein the suction channel comprises: a first portion formed in the compressor block; and a second portion formed in the at least one cylinder head; wherein the first and second portions are configured to mate at a connection to form a fluid passage between the first portion and the second portion for delivering working fluid to the working chamber in use. The suction valve may comprise at least one non-self-actuated valve and the discharge valve may comprise at least one self-actuated check valve. The reciprocating compressor of the fifth aspect of the invention may have any one or more further features as set out in relation to the reciprocating compressor of the first aspect of the invention.

According to a sixth aspect of the invention, there is provided a method of using a reciprocating compressor according to the sixth aspect of the invention.

According to a seventh aspect of the invention, there is provided a reciprocating compressor comprising a working chamber and a buffer reservoir, wherein the buffer reservoir has a substantially larger volume than the volume of the working chamber such that in use when working fluid is sucked out of the buffer reservoir to fill the working chamber, there is not a substantial pressure drop in the buffer reservoir. The reciprocating compressor may further comprise: a suction valve, a discharge valve, and a working chamber, the suction valve being configured to provide fluid communication between a suction channel and the working chamber and the discharge valve being configured to provide fluid communication between the working chamber and a discharge channel. The suction valve may comprise at least one non-self-actuated valve and the discharge valve may comprise at least one self-actuated check valve. The reciprocating compressor may have any one or more further features as set out in relation to the reciprocating compressor of the first aspect of the invention.

According to an eighth aspect of the invention, there is provided a method of operating a reciprocating compressor according to the seventh aspect of the invention, comprising a step of sucking working fluid into the working chamber from the buffer reservoir.

BRIEF DESCRIPTION OF THE DRAWINGS

The prior art and embodiments of the invention will now be described, by way of example only, with reference to the accompanying drawings, in which: Fig. 1 shows a standard prior art compressor;

Fig. 2 shows an improved compressor;

Fig. 3 shows the improved compressor of Fig. 2 with an improved alternative suction channel; Fig. 4 shows the improved compressor of Fig. 2 with an alternative suction channel;

Fig. 5 shows a detailed section view of a piston compressor;

Fig. 6 shows a detailed section view of the piston compressor of Fig. 5;

Fig. 7 shows a detailed bottom view of the cylinder head of Fig. 6;

Fig. 8 shows a limited section view of a discharge valve of the piston compressor shown in Fig. 5;

Fig. 9 shows an optional lubricant return system for the compressor shown in Fig. 5.

Figures 10a and 10b show a first example of a self-actuated valve;

Figures 11 a and 11 b show a second example of a self-actuated valve; and Figures 12a and 12b show a third example of a self-actuated valve.

DETAILED DESCRIPTION

Figure 1 shows a prior art compressor 1 in the form of a piston compressor which utilises a condensable working fluid. The compressor 1 comprises a compressor block 20 having a cylinder 250 with a reciprocating piston 240 arranged inside the cylinder 250, and a cylinder head 10 having a suction channel 40 and a discharge channel 50. The cylinder 250, cylinder head 10 and piston 240 define a working chamber 251 for compression of the compressible working fluid therein. Selection of suitable compressible working fluids is well known in the art.

Working fluid inflow is controlled and admitted from the suction channel 40 into the working chamber 251 through a suction valve 110, and working fluid outflow is controlled and admitted out of the working chamber 251 to the discharge channel 50 through two discharge valves 120. As can clearly be seen in Figure 1 , the suction channel 40 is arranged to suck working fluid into the working chamber 251 . The suction channel 40 is of a substantially vertical and substantially straight configuration. It will be understood by a person skilled in the art that the suction channel 40 connects to other piping and/or manifolds where the working fluid is provided through. Although in the presently described example one suction valve 110 and two discharge valves 120 are provided, any number of suction valves 110 and discharge valves 120 working in parallel may be found in other prior art examples.

The valves 110, 120 shown in Figure 1 are of the reed-type, having at least a reed valve blade 121 and a reed valve retainer plate 122, also called a stop plate. The reed blade 121 controls the opening and thus the fluid communication through a valve opening 143, also called a valve port or a valve slot, in the cylinder head 10. Reed valves function in the way that they are operated by a differential pressure exerted across them, and as such they are “passively controlled”, which means that they do not need any actuator elements to open or close. Opening or closing is performed solely as a function of the differential pressure across the reed blades 121 and in this way such valves can have a simple construction, however, as described above, they also come with drawbacks. In other words, the valve is closed by the reed blade 121 being pushed against the valve slot 143, and thus closing the slot, due to the pressure differential across the reed blade 121 in this direction. Therefore, reed valves also function as check valves, since they only open in one direction.

Figure 2 shows an example of an improved compressor T. Similarly to the prior art compressor 1 described with reference to Figure 1 , the compressor T is a piston compressor which utilises a condensable working fluid. The compressor T comprises a compressor block 20’ having a cylinder 250’ with a reciprocating piston 240’ arranged inside the cylinder 250’, and a cylinder head 10’ having a suction channel 40’ and a discharge channel 50’. The cylinder 250’, cylinder head 10’ and piston 240’ define a working chamber 25T for compression of the working fluid therein.

Working fluid inflow is controlled and admitted from the suction channel 40’ into the working chamber 251 ’ through a suction valve 110’, and working fluid outflow is controlled and admitted out of the working chamber 25T to the discharge channel 50’ through discharge valves 120’. As can clearly be seen in Figure 2, the suction channel 40’ is arranged to suck working fluid into the working chamber 25T with the suction channel 40’ in a substantially vertical and substantially straight configuration. It will be understood by a person skilled in the art that the suction channel 40’ connects to other piping where the working fluid is provided through. Although in the presently described example one suction valve 110’ and two discharge valves 120’ are provided for one cylinder 250’, any number of suction and discharge valves working in parallel may be found in other examples where multiple cylinders 250’ are provided. Simiarly, although shown as one suction valve 110’ and two discharge valves 120’ in the presently described example, in other examples there may be any number of suction valves 110’ and/or any number of discharge valves 120’ for each cylinder 250’.

The suction valve 110’ of the example shown in Figure 2 is a poppet valve, while the discharge valves 120 are reed valves as in the prior art compressor 1 shown in Figure 1 . Due to the geometry of the poppet suction valve 110’, a larger effective opening area can be achieved than with a reed valve, considering the same overall dimension of the compressor T, such as cylinder bore. In addition, the poppet suction valve 110’ can be actively controlled by means of mechanical actuators as is shown in greater detail later with reference to later Figures.

Still referring to Figure 2, it can be seen that in the presently described example, the suction valve 110’ is arranged centrally relative to the working chamber 25T and cylinder 250’, i.e. the suction valve 110’ is on the central axis 252’ of the cylinder 250’. Additionally, it can be seen that the discharge valves 120’ are arranged circumferentially around the suction valve 110’. That is to say, the discharge valves 120’ encircle the suction valve 110’ and the suction valve 110’ and discharge valves 120’ share a central axis with the central axis 252’ of the cylinder 250’, and together the suction valve 110’ and discharge valves 120’ consume substantially the entire area of the top of the working chamber 25V. The suction valve 110’ and the discharge valves 120’ are arranged concentrically as shown in Figure 2. In some alternative examples, the suction valve 110’ and discharge valves 120’ may be arranged concentrically, but not sharing a central axis 252’ with the cylinder 240’. Referring now to Figure 3, an alternative compressor 1” is shown with a suction channel 40” configured to provide a suction fluid flow path 41” therethrough in use. The suction fluid flow path 41” is significantly different from a fluid flow path (not shown) for the compressor T. The suction fluid flow path 41” is formed by connection of the compressor block 20” and the cylinder head 10” at a connection 42”. The suction channel 40” is partly in the cylinder head 10” and partly in the compressor block 20”. The suction channel 40” comprises a common manifold 43” in the compressor block 20” where working fluid first enters the suction channel 40” from auxiliary piping, a buffer reservoir 44” and a fluid inlet channel 45” connecting the buffer reservoir 44” to the working chamber 251”. Still referring to the cross-sectional view shown in Figure 3, the compressor 1” has a front facing towards the viewer in Figure 3, a rear facing away from the viewer and a top where the suction channel 40” and discharge channel 50” connect to additional piping.

As working fluid enters the common manifold 43” of the suction channel 40” at the rear of the compressor 1”, the working fluid travels towards the front of the compressor 1” in the direction of the viewer in Figure 3 and travels downwards to the buffer reservoir 44”. From the buffer reservoir 44” the working fluid is sent upwards towards the top of the compressor 1” through the connection 42” and via the fluid inlet channel 45” to the working chamber 251”. In the presently described example, the working fluid travels across the compressor 1” and then travels upwards towards the top of the compressor 1” through the angle 46” in the buffer reservoir 44”. In the presently described example, the angle 46” is 90 degrees, that is to say the fluid travels across the compressor 1” perpendicular to the central axis 252” of the cylinder along a horizontal axis 47” and then travels parallel to the central axis 252” of the cylinder along a vertical axis 48” and then turns into the working chamber 251” via the suction valve 110” as can be seen in Figure 3.

In other exemplary compressors (not shown) comprising multiple cylinders 250”, there may be provided a single common manifold 43” for receiving working fluid from the auxiliary piping. In this connection, each cylinder 250” may be arranged to receive working fluid from the common manifold 43” through a splitting of the working fluid between the common manifold 43” and each of the buffer reservoirs 44”. This not only reduces the number of external connections that otherwise need to be made (e.g. one for each cylinder head’s 10” suction channel 40”), but also greatly reduces the number of seals/gaskets needed, since the connection 42” is sealed using the same seal as would be required anyway for the interface between the cylinder head 10” and the compressor block 20”, as will be explained in more detail later.

Still referring to Figure 3, it can be seen that in the presently described example, the suction valve 110” is arranged centrally relative to the working chamber 251” and cylinder 250”, i.e. the suction valve 110” is on the central axis 252” of the cylinder 250”. Additionally, it can be seen that the discharge valves 120” are arranged circumferentially around the suction valve 110”. That is to say, the discharge valves 120” encircle the suction valve 110” and share a central axis with the central axis 252” of the cylinder 250”, and together the suction valve 110” and discharge valves 120” consume substantially the entire area of the top of the working chamber 251 ”.

Referring now to Figure 4, another alternative compressor T” is shown with a suction channel 40”’ configured to provide a suction fluid flow path 41”’ therethrough in use. Similarly to the compressor 1” of Figure 3, the suction fluid flow path 41’” is formed by connection of the compressor block 20’” and the cylinder head 10’” at a connection 42’”. The suction channel 40’” is partly in the cylinder head 10’” and partly in the compressor block 20’”. The suction channel 40’” comprises a common manifold 43’” in the compressor block 20’” where working fluid first enters the suction channel 40’” from auxiliary piping, a buffer reservoir 44’” and a fluid inlet channel 45’” connecting the buffer reservoir 44’” to the working chamber 25V”. Still referring to the cross-sectional view shown in Figure 4, the compressor T” has a front facing towards the viewer in Figure 4, a rear facing away from the viewer and a top where the suction channel 40’” and discharge channel 50’” connect to additional piping.

As working fluid enters the common manifold 43’” of the suction channel 40’” at the rear of the compressor T”, the working fluid travels towards the front of the compressor T” in the direction of the viewer in Figure 4 and travels downwards to the buffer reservoir 44’”. From the buffer reservoir 44’” the working fluid is sent upwards towards the top of the compressor T” through the connection 42’” and via the fluid inlet channel 45’” to the working chamber 25V”. In the presently described example, the working fluid travels across the compressor T” and then travels upwards towards the top of the compressor T” through the angle 46”’ in the buffer reservoir 44”’. In the presently described example, the angle 46’” is 45 degrees, that is to say the fluid travels across the compressor T” perpendicular the central axis 252’” of the cylinder 240’” along a horizontal axis 47’” and then travels obliquely to the central axis 252’” of the cylinder 240’” as shown in Figure 4. In other examples, another acute angle may be provided for the angle 46’”. For example, the angle 46’” may be 30 degrees, or 40 degrees, or between 40 degrees and 50 degrees, or 60 degrees.

Reducing the angle 46’” which the working fluid must turn at in the buffer reservoir 44’” results in the fluid flow path 41’” taking a smoother path through the compressor T”, which may result in a reduced pressure drop as the working fluid moves from the common manifold 43’” to the fluid inlet channel 45’”, thereby resulting in increased performance of the compressor T”.

The cylinder head 10’” may in some examples comprise a valve plate (not shown in Figure 4) configured to provide seating and sealing of the suction valve 110’” and discharge valves 120’” in the top of the working chamber 25V”.

Figure 5 shows a detailed section view of an alternative compressor 1””, Figure 6 shows show a detailed section view of the valve system of the compressor 1””, Figure 7 shows a detailed bottom view of the cylinder head of the compressor 1”” and Figure 8 shows a detailed view of the discharge valve of the compressor 1””.

Referring firstly to Figure 5, the compressor 1”” comprises first and second cylinder heads 10A, 10B, a compressor block 20””, a lubricant reservoir 30””, a crankshaft 210”” with counterweights 220”” to reduce vibration and first and second connecting rods 211 A, 211 B connected between the crankshaft 210”” and first and second pistons 240A, 240B arranged in first and second cylinders 250A, 250B to define first and second working chambers 251 A, 251 B. The compressor 1”” further comprises first and second suction channels 40A, 40B configured to provide working fluid to the working chambers 251 A, 251 B in use. The compressor 1”” is designed using a so-called “V” -configuration, having the first and second cylinders 250A, 250B angularly separated by about 90°. Preferably, V-configuration compressors such as in the presently described example may comprise an even number of cylinders 250A, 250B, however it may be envisaged that the compressor 1”” may be configured with an odd number of cylinders 250A, 250B in some examples.

A working fluid has typically undergone pre-heating, evaporation and superheating in one or more heat exchangers (not shown) before entering the compressor 1”” typically through a pipe, a manifold, or a network of pipes (not shown). The working fluid enters the compressor 1”” through a single common manifold 43””, thereby providing only one entry point for the working fluid into the compressor 1””. In the presently described example, as in previous examples, the common manifold 43”” is arranged at the rear of the compressor 1 It will be understood that the front and rear of the compressor 1”” may be reversed in other examples, and indeed that the common manifold 43”” may be located centrally on the compressor in some examples rather than being at the front or the rear. Still referring to Figure 5, after the working fluid has entered the common manifold 43””, the working fluid splits and travels towards the front of the compressor 1”” from the common manifold 43”” in two separate flows (not shown) of working fluid travelling in the direction of the viewer in Figure 5 and each flow travelling downwards with the first working fluid flow travelling to the first buffer reservoir 44A and the second working fluid flow travelling to the second buffer reservoir 44B. In the interest of brevity, the following explanation is in regard to the first fluid flow only. However, it will be understood that the same description applies to the second fluid flow, mutatis mutandis.

The first fluid flow travels from the first buffer reservoir 44A upwards towards the top of the compressor 1”” through a first connection 42A and via a first fluid inlet channel 45A to a first working chamber 251 A. In summary, in the presently described example, the working fluid travels across the compressor 1”” and then travels upwards towards the top of the compressor 1”” via the first buffer reservoir 44A.

When the first piston 240A expands the volume of the working chamber 251 A while moving away from the cylinder head 10A, suction causes working fluid to enter the first working chamber 251 A through a first suction valve 110A, which in the presently described example is a poppet valve.

By the rotary motion of the crankshaft 210””, a translatory motion of the first piston 240A inside the first cylinder 250A is provided, cyclically bringing the pistons 240A “down” to bottom dead center (BDC) position, at which the suction process stops. The compressor 1”” then starts to compress the working fluid in the first working chamber 251 A until the working fluid has reached a pressure exceeding that of the pressure in a discharge channel 50A, and then the discharge process begins, as the first discharge valves 120A, which in the presently described example are reed valves, then open because of the differential pressure across their reed blades, as previously explained.

As previously discussed, the same steps are performed for the second piston 240B in the second cylinder 250B. The first discharge channel 50A may continue into a manifold, pipe or similar where a fluid connection may be made to the second discharge channel 50B such that the working fluid flows can be combined and the combined fluid flow may then be lead further to one or more heat exchangers (not shown), in which the working fluid is typically desuperheated, condensed and optionally subcooled. In other examples, one or more of these processes may be applied, or other processes may be applied to the discharged working fluid.

The first and second inlet valves 110A, 110B are actuated by conventional valve actuator mechanisms, the design and/selection of which will be well within the capabilities of a person skilled in the art.

The first suction channel 40A is partly in the first cylinder head 10A and partly in the compressor block 20”” with connection of the two parts of the first suction channel 40A being provided by the first connection 42A.

In the presently described example, the working fluid in the first fluid flow stream travels across the compressor 1”” and then travels upwards towards the top of the compressor 1””. Although the working fluid path is only described above with reference to the first buffer reservoir 44A, it will be appreciated that the second buffer reservoir 44B is also configured in a similar fashion to as described for the first buffer reservoir 44A. It will also be appreciated that in examples with further cylinders, such as sixteen cylinders, some or all of the buffer reservoirs may be as described above for the first buffer reservoir 44A.

Still referring to Figure 5, the first buffer reservoir 44A provides a larger volume relative to the common manifold 43”” and piping leading to the first buffer reservoir 44A. As can be seen in Figure 5, the first buffer reservoir 44A is located directly adjacent the first working chamber 251 A with only the required fluid inlet channel 45A connecting the first buffer reservoir 44A and the first working chamber 251 A. This allows the first buffer reservoir 44A to maintain a sufficient volume of working fluid such that when working fluid is sucked into the first working chamber 251 A there is not a significant pressure drop, as may be the case if the relatively narrow piping between the common manifold 43”” and the first buffer reservoir 44A were to extend directly to the working chamber 251 A without entering a region of larger volume. In this connection, the first buffer reservoir 44A is of a sufficiently large volume relative to the first working chamber 251 A such that when a volume of working fluid equal to the volume of the first working chamber 251 A when the first piston is at bottom dead centre position, i.e. the full volume of the first working chamber 251 A, is extracted from the first buffer reservoir 44A, there is not substantial pressure drop in the fluid in the first buffer reservoir 44A. In this connection, sufficiently pressurised fluid is always available to the first working chamber 251 A. Additionally, the first buffer reservoir 44A can be continually fed working fluid from the common manifold 43”” to continually ‘top up’ the first buffer reservoir 44A, rather than the first working chamber 251 A sucking fluid from the entire suction channel 40A which would result in a pressure drop. The reduction or elimination of a substantial pressure drop improves the performance of the compressor 1””.

Still referring to Figure 5, when the compressor 1”” is assembled and ready to operate in a heat pump system for example, there are a plurality of interior volumes 60”” in the compressor 1”” which will typically be pressurised due to the properties of the working fluid being used. Therefore the compressor 1”” needs to be hermetically closed against the atmosphere. In order to still have practical access to important internal volumes 60”” of the compressor 1””, the compressor 1”” comprises pressure-reinforced covers 101 A, 101 B, 201 A, 201 B, that enable practical access during assembly, service or for other relevant reasons, while still being strong enough to withstand the internal pressures in the compressor’s 1”” interior volumes 60”” when pressurized.

The lubricant reservoir 30”” is equipped with a plurality of lubricant heaters 301””, that are embedded in a main block 300”” of the lubricant reservoir 30””. These heaters 301 ”” can be activated upon relevant need through a standard control system (not shown).

Figure 6 shows a detailed view of the suction and discharge of the first working chamber 251A and the associated components previously described. Figure 7 shows a detailed section view at the first cylinder head 10A of the compressor 1”” where the first cylinder head 10A meets the compressor block 20””. The first suction valve 110A can be seen encircled by the first discharge valve 120A. The first discharge valve 120A comprises a first valve plate 140A, in which there are a plurality of first discharge valve slots 142A and corresponding first discharge valve bridges 143A between the slots 142A. In the presently described example the plurality of first discharge valve slots 142A are arranged in concentric circles, however it will be understood that other arrangements are possible.

The first discharge valve slots 142A provide fluid communication between the first working chamber 251 A (shown in Figures 5 and 6) and the first discharge channel 50A. In some examples (not shown), the first valve plate 140A may be omitted, and its function replaced by the cylinder head main block (not shown) instead, as will be appreciated by a person skilled in the art.

In the cross-sectional view in Figure 7 the first connection 42A and first discharge channel 50A are also visible.

Figure 8 shows a detailed view of the first discharge valve 120A. The first discharge valve 120A comprises first reed valve blades 121 A which are pressed against the first valve slots 142A of the first valve plate 140A so that the first discharge valve 120A is closed. Separate first reed valve springs 123A provide a closing force to the first reed valve blades 121 A, as the first reed valve springs 123A are pressed against a first valve retainer plate 122A on the opposite side of the first reed valve blades 121 A. In other examples, the spring force provided by the first reed valve springs 123A may be provided by forming the first reed valve blades 121 A out of a spring material and arranging the first reed valve blades 121 A so that they close the discharge valve 120A automatically through their own inherent spring force (such an arrangement is shown in Figures 1 to 4). However, in the preferred example shown in Figures 5 to 8, due to the use of separate first reed valve springs 123A, the first reed valve blades 121 A can be made from a non-spring material.

In the present invention, a first recess 141 A is provided in the first valve plate 140A as can be seen in Figure 8. The first recess 141 A covers the region comprising the first reed valve slots 142A, to decrease the dead volume contribution by the first reed valve slots 142A. The first recess 141 A is formed by machining away this region in which there is no need for the full thickness of the first valve plate 140A, and so material can be machined away without compromising on the required rigidity.

Volumetric efficiency is an important performance factor of any compressor, and this is closely linked to the compressor cylinders’ dead volume: The dead volume is the combination of non-displaceable (by the piston) volumes primarily caused by two factors. Firstly, the clearance volume needed to ensure that the piston does not collide with the cylinder head, valve plate or valves due to production tolerances, valve motion, varying gasket thicknesses etc., and secondly passive volumes that are required as part of ports, slots or other, due to the need to provide a certain material thickness to house these. Referring again to Figure 6, the first piston 240A comprises a first piston crown 241 A comprising a first crown recess 242A and a first crown protrusion 243A. The first crown recess 242A is provided such that the suction valve 110A will enter and displace a major part of the volume of the first crown recess 242A when the first piston 240A is close to top dead center (TDC) during operation. The first crown protrusion 243A is seen on either side of the first crown recess 242A, although it will be easily understood that the first crown protrusion 243A is circular on the first piston 240A. The first crown protrusion 243A is registered with the first recess 141 A such that the first crown protrusion 243A will enter and displace a major part of the volume of the first recess 141 A when the first piston 240A is close to top dead center (TDC) during operation. The provision of a first crown protrusion 243A and a first crown recess 242A reduces the dead volume and thereby increase the volumetric efficiency.

Referring now to Figure 9, the compressor 1”” is again shown but now with an optional lubricant return system 400 comprising an inlet 401”” fluidly connected to a lower surface of the first buffer reservoir 44A. The inlet 401”” communicates lubricant to a tank 402”” where the lubricant is temporarily stored in use. The tank 402”” comprises a valve 403”” which is selectively openable to transfer the lubricant from the tank 402”” to the lubricant reservoir 30””. The valve 403”” may be configured with a timer such that the valve 403”” opens every 20 seconds, 30 seconds, 40 seconds for example, to transfer lubricant back into the lubricant reservoir 30””. Alternatively, or in addition to the timer, the valve 403”” may be configured to be selectively opened by the user of the system or to open automatically when a certain volume of lubricant has gathered in the tank 402””. Although only shown in Figure 9 in connection with the first buffer reservoir 44A, it will be understood that one, all or only some of the buffer reservoirs may be provided with a lubricant return system 400””. For example, if sixteen cylinders are used, there may be provided sixteen lubricant return systems 400””. In some examples where there are multiple buffer reservoirs, the lubricant from each buffer reservoir may be channelled and gathered in one tank, before being channelled back to the lubricant reservoir 30””. In some examples, a valve (not shown) may be provided at the inlet 401 ”” to selectively control the delivery of lubricant to the tank 402””. In other examples, the tank 402”” may be omitted.

In any of the above-described examples, the suction channels 40’, 40”, 40”’, 40A, 40B, and/or discharge channels 50’, 50”, 50’”, 50A, 50B, may be formed during casting of the cylinder head 10’, 10”, 10’”, 10A, 10B and compressor block 20’, 20”, 20’”, 20””, respectively.

The presently described compressors allow a lower pressure drop to be maintained across the suction valve during operation, thereby maintaining a higher efficiency of the compressor when utilizing more demanding working fluids, and especially those with a higher density.

The present disclosure describes poppet and reed valves. It will be understood that the reed valves may be replaced by any self-actuated check valve or valves and the poppet valve may be replaced by any non-self-actuated valve or valves. As non-limiting examples only, known arrangements of self- actuated check valves are now described briefly with reference to Figures 10a to 12b. Firstly referring to Figures 10a and 10b there is shown a first self-actuated check valve 500 which is configured to move from a closed position shown in Figure 10a to an open position shown in Figure 10b. The self-actuated check valve 500 comprises a valve plate 510 and first and second reed elements 521 , 522 which are biased to respectively close first and second apertures 531 , 532 in the valve plate 510 in the absence of sufficient pressure in the direction of arrows A. The self-actuated check valve 500 further comprises first and second curved stop plates 541 , 542 configured to arrest the movement of the first and second reed elements 521 , 522 when the reed elements 521 , 522 are moved to the open position.

As another non-limiting example, another self-actuated check valve 600 in the form of a plate valve is provided which is configured to move from a closed position shown in Figure 11a to an open position shown in Figure 11 b. The selfactuated check valve 600 comprises a valve plate 610 and first, second and third biased closure elements 621 , 622, 623 which are biased to respectively close first, second and third apertures 631 , 632, 633 in the valve plate 610 in the absence of sufficient pressure in the direction of arrow A. The self-actuated check valve 600 further comprises first, second and third stop plates 641 , 642, 643 configured to arrest the movement of the first, second and third biased closure elements 621 , 622, 623 when the biased closure elements 621 , 622, 623 are moved to the open position shown in Figure 11 b. The actual biassing means used is not shown, however it will be understood that the biassing may be implement in a myriad of ways.

As another non-limiting example, another self-actuated check valve 700 is provided which is configured to move from a closed position shown in Figure 12a to an open position shown in Figure 12b. The self-actuated check valve 700 comprises a valve plate 710 and first, second and third spring biased poppet elements 721 , 722, 723 which are biased to respectively close first, second and third apertures 731 , 732, 733 in the valve plate 710 in the absence of sufficient pressure in the direction of arrow A.

As previously mentioned, throughout the present disclosure, reference is made to reed valves and poppet valves. However, it will be appreciated that configurations using a reed valve or valves may be replaced by any self-actuated check valve or valves and configurations using a poppet valve or valves may be replaced by any non-self-actuated valve or valves. The self-actuated valves may be any of the non-limiting examples shown in Figures 10 to 12, or may be any other self-actuated valve.

It will also be appreciated that “self-actuated” is intended to carry the standard meaning within the art. That is that self-actuated valves are passively opened and closed, i.e. they open and close in response to fluid pressure in the system rather than being driven by an external electrical or mechanical source provided for the purpose of opening or closing the valve. Similarly, “non-self actuated” is also intended to carry the standard meaning within the art, i.e. that the valve is actively driven, rather than opening and closing in response to fluid pressure in the system.

It will be understood that at the time of writing, the term “high-temperature heat pump” generally refers to the heat pumps capable of supplying output temperatures above 55 degrees Celsius. It will be appreciated by a person skilled in the art that the definition of “high-temperature” in this context may change over time, and it is foreseeable with advancements in technology that in the future “high-temperature” may be used to refer to heat pumps with an output temperature of above 80 degrees Celsius for example, or even higher.

In the above description there are several features of the described examples that are well-known to a person skilled in the art, and which have either been omitted or at least not described in detail for the sake of brevity. Further, although a preferred example comprises a heat pump compressor, and more specifically a heat pump compressor that utilises a condensable gas, the invention is not limited to this, and may be just as relevant for any other application requiring a compressor, for example an air compressor, natural gas compressor, CO2 compressor etc.

It will be understood that a plurality of non-self-actuated valves may together form a “suction valve”, i.e. there need not be one single non-self- actuated valve element forming the entire suction valve. Likewise, a plurality of self-actuated check valves may together form a “discharge valve”, i.e. there need not be one single self-actuated check valve element forming the entire discharge valve. For example, as shown in the example described with reference to Figure 2, there is provide two self-actuated check valve discharge elements forming the discharge valve. It will be understood that there may likewise be two, three, four or more non-self-actuated suction valve elements forming the suction valve in some examples (not shown).