SEP4 Project Report Heat Pump · 2017-11-15 · SEP4 PROJECT REPORT HEAT PUMP Tamas Pinter...

[Type here] [Type here] [Type here]

po

SEP4 PROJECT REPORT

HEAT PUMP Tamas Pinter (239855), Gergo Szemeti (239935), Julian McIntosh (239934),

Behnam Pariz (240542), Ben Gallienne (239930)

SEP4 PROJECT REPORT

HEAT PUMP Tamas Pinter (239855), Gergo Szemeti (239935), Julian McIntosh (239934),

Behnam Pariz (240542), Ben Gallienne (239930)

Supervisors:

Bo Leander Gylling

Pia Weber Pedersen

Tamas Pinter, Ben Gallienne, Julian McIntosh, Gergo Szemeti, Behnam Pariz

1 | P a g e

Abstract The team created a vertical ground-source heat pump to be used as a climate control system in medium-large detached houses in Denmark. A variety of planning methods were used to organise the project, including Gantt-charts, Work Breakdown Structures, Project Logs and Concentric Circles. Different approaches were discussed and evaluated for the project, weighing up the pros and cons of different types of climate control system, components and coolants against each-other. The economics of the project were also discussed, covering Market Analysis, Budgeting suggestions for the associated company, and costing suggestions for the project. The project resulted in a mechanically successful design, although the economic analysis showed that the system could only truly be efficient when scaled up to be used in larger buildings or several houses at once, as in its existing form it is not a worthwhile purchase for most of its target markets.


2 | P a g e

Table of Contents List of Tables and Figures ........................................................................................................................ 5

Executive Summary ............................................................................ Fejl! Bogmærke er ikke defineret.

Introduction ............................................................................................................................................ 6

Collaboration with UniClimat .................................................................................................................. 6

Project Planning & Structuring........................................................... Fejl! Bogmærke er ikke defineret.

Decision-Making Processes ............................................................ Fejl! Bogmærke er ikke defineret.

Work-breakdown Structure ........................................................... Fejl! Bogmærke er ikke defineret.

Concentric Circles ........................................................................... Fejl! Bogmærke er ikke defineret.

Research .................................................................................................................................................. 9

Heat pump types ................................................................................................................................. 9

Air source heat pump - ................................................................................................................... 9

Ground source heat pump .............................................................................................................. 9

Exhaust air heat pump - .................................................................................................................. 9

Water source heat pump ................................................................................................................ 9

Hybrid heat pump ........................................................................................................................... 9

Solar-assisted heat pump ................................................................................................................ 9

Advantages of the Heat Pumps........................................................................................................... 9

Disadvantages of the Heat Pumps .................................................................................................... 10

Horizontal Ground Source Heat Pump ............................................................................................. 10

Vertical Ground Source Heat Pump .................................................................................................. 10

Project Limitations & Assumptions ....................................................................................................... 10

The House ..................................................................................................................................... 10

The Environment ........................................................................................................................... 10

System Operation ......................................................................................................................... 11

Final Design ........................................................................................................................................... 11

System Layout ................................................................................................................................... 11

Components Summary ..................................................................................................................... 12

Compressor ................................................................................................................................... 12

Heat Exchanger ............................................................................................................................. 13

Accumulator .................................................................................................................................. 13

Thermostatic Expansion Valve ...................................................................................................... 13

Control Panel..................................................................................................................................... 14

Materials ............................................................................................................................................... 14


3 | P a g e

Surface Heat Exchanger Piping ......................................................................................................... 14

Comparison of physical properties ................................................................................................... 15

Geothermal Heat Exchanger Piping .................................................................................................. 16

Copper ........................................................................................................................................... 17

Polyethylene (PE) .......................................................................................................................... 17

High Density Polyethylene (HDPE) ................................................................................................ 17

Poly Vinyl Chloride (PVC) .............................................................................................................. 17

Chlorinated Poly Vinyl Chloride (CPVC) ........................................................................................ 18

Composite Piping .......................................................................................................................... 18

Comparison of Physical Properties ................................................................................................... 18

Heat Exchanger Case ......................................................................................................................... 19

Comparison of Physical Properties ................................................................................................... 20

Production Methods ............................................................................................................................. 21

Review of the Thermodynamic and Mechanic Calculations and Interpreting the Results ................... 21

Heat Requirement of the Building .................................................................................................... 22

Heat Exchanger and Compressor Properties .................................................................................... 23

Indoor Heat Exchanger Dimensioning .............................................................................................. 24

Heat Transfer Coefficient .............................................................................................................. 24

Plate Dimensioning in the Heat Exchanger ................................................................................... 25

Compressor Calculations .................................................................................................................. 26

Ground Loop Pipe Dimensioning ...................................................................................................... 28

Energy Requirement from the Ground ......................................................................................... 28

Mass Flow Rate and Mechanics Calculation ................................................................................. 28

Outdoor Heat Exchanger Dimensioning ........................................................................................... 29

Heat Transfer Coefficient .............................................................................................................. 29

Plate Dimensioning in the Heat Exchanger ................................................................................... 29

Calculating the Required Length of the Pipes in the Ground Loop .................................................. 31

Pump Dimensioning .......................................................................................................................... 31

Heat Pump Losses ............................................................................................................................. 32

Coefficient of Performance Calculation ............................................................................................ 33

Decision Between Plate Type Heat Exchanger and Pipe Type Heat Exchanger................................ 33

Refrigerant Selection (R134a) ............................................................................................................... 34

Regulations ........................................................................................................................................... 36

Regulations for Internal Parts ....................................................................................................... 36

Testing of Complete System ......................................................................................................... 37

Safety Considerations ........................................................................................................................... 38


4 | P a g e

Economics ............................................................................................................................................. 39

Introduction ...................................................................................................................................... 39

Analysis of Survey and Results ........................................................................................................... 39

Improvements for Survey Questions ............................................................................................. 41

Market Research ............................................................................................................................... 41

Budget ............................................................................................................................................... 43

Investment and Financial Consequences .......................................................................................... 43

Conclusion ............................................................................................................................................. 47

Bibliography .......................................................................................................................................... 47


5 | P a g e

List of Tables and Figures Figure 1 - Diagram for system ............................................................................................................... 11

Figure 2 - Grundfoss VTZ compressor ................................................................................................... 13

Figure 3 - Grundfoss thermostatic expansion valve ............................................................................. 14

Figure 4 - Pedrollo PK-60 pump ............................................................................................................ 31

Figure 5 - Specifications chart from pump manufacturer .................................................................... 31

Figure 6 - Distribution curve of heat consumption ............................................................................... 39

Figure 7 - Interest in heat pumps .......................................................................................................... 40

Figure 8 - Payback period ...................................................................................................................... 43

Figure 9 - Comparison of cost and NPV for houses .............................................................................. 44

Figure 10 - Payback period .................................................................................................................... 45

Figure 11 - Investment for all energy sources ...................................................................................... 45

Figure 12 - NPV for energy sources ...................................................................................................... 46

Figure 13 - Price for heat pumps ........................................................................................................... 46

Table 1 - Pipe materials advantages vs disadvantages ......................................................................... 15

Table 2 - Properties of pipe materials ................................................................................................... 15

Table 3 - Advantages and disadvantages of ground loop materials ..................................................... 17

Table 4 - Material properties ................................................................................................................ 18

Table 5 - Casing material advantages vs disadvantages ....................................................................... 20

Table 6 - Specifications of casing materials .......................................................................................... 21

Table 7 - Summary table for data presented in the section ................................................................. 23



Table 10 - Summary table for data presented in the section ............................................................... 26










Table 20 - Properties of pipes ............................................................................................................... 34

Table 21 - Table describing vapor saturation temperatures at two different pressures ..................... 35

Table 22 - Heat usage according to source ........................................................................................... 42

Table 23 - Prices for unit ....................................................................................................................... 42

Table 24 - Consumption of energy sources .......................................................................................... 44

https://viaucdk.sharepoint.com/sites/sepm4/Delte%20dokumenter/SEP4%20Project%20Report.docx#_Toc484148446









6 | P a g e

Introduction Having a warm home has always been important, especially in colder climates. These days, with the drive towards green energy and lowering ones carbon footprint, the importance of having a functional, efficient heating system in ones home is increasingly vital. The drive towards efficiency and eco-friendliness has, through new regulations and laws, begun to push certain types of heating systems out of the market (such as oil-based heating systems), thereby opening up opportunities for new products and innovations to take their place.

Collaboration with UniClimat Sergey Muraveynikov

This project began with the involvement in the INNOVA Project, a Nordic-Russian cooperation

project between students and staff from VIA University College in Denmark, ITMO

University and Lomonosov Moscow State University in Russia and HAMK University in Finland. This

project aims to achieve full collaboration between these universities, in the hopes to share helpful

information for mutual growth and to learn about different teaching methods that will

potentially improve the educational experience of the students.

In collaboration with this project is Sergey Muraveynikov from ITMO University - he started his

own small company called UniClimat. UniClimat works with Air Handling Units, Dehumidifiers, Heat

Exchangers and other such devices that control the climate inside buildings. Although the company

is young it has already made great advancements in the technology used in these devices.

UniClimat provided the project with the task of creating a plan for the creation of a Heat

Exchanger. UniClimat has previously taken an Air Handling Unit and made certain experimental

changes to it, to make it more optimised for smaller buildings. There was no concrete information

regarding the Heat Exchanger, just a few idea generation sketches and a parts

list. Therefore, this project was tasked to provide a thorough plan and procedure for this process, by

fully understanding the mechanisms of this device and making relevant calculations to optimise the

process and costs.

This project was then modified and expanded slightly by the group, in collaboration with UniClimate,

to encompass an entire heat-pumping system. This change had to be made as the group was unsure

whether a single heat exchanger would provide a challenging enough project for 5 engineers to

tackle.

At the end of this project, a lot of useful information was provided to UniClimat. They received a full

parts list also including fittings and materials, a production procedure with details on how the device

should be assembled, 3D drawings for a detailed look at the mechanisms of the device, a new

refrigerant that was not harmful or a potential hazard and a full analysis of the market for this device

and of course budget plan with all costs and investments considered.


7 | P a g e

Background In 2017, most houses and apartment blocks contain some form of climate control. This involves a

variety of systems such as radiators, heat exchangers and central heating. Many of these heat

exchange systems can, when used in small buildings or apartments, be relatively inefficient and

expensive, meaning that many households and building companies will settle for a model that is not

sufficient for their needs so as to cut costs.

Sergey Muraveynikov from ITMO University suggests that we work to improve on an existing

product, the Flaktwoods ECONET, attempting to improve on its design in terms of either efficiency,

cost or complexity.

Purpose The purpose of this project is to develop a heat exchanger system that is designed for individual

households, making it moreoptimized for application in small buildings

Problem Formulation

Design Air Handling Units (AHUs) contain many parts and therefore require efficient use of space and

proper housing design to compliment the function of the system. The ductwork systeminvolved is

very large and may cause problems in small buildings since they require a larger ventilation chamber

and more technical space in buildings.At the same time such a larger system is not needed for a

smaller household.

Our design will:

• Use space efficiently, therefore making it a smaller product.

• Be cost efficient in terms of production and maintenance.

Market Considering the information given to us regarding the specific market that is targeted with our

product - we would do mostly secondary research, since companies in this line of work are

somewhat lacking. This in turn opens many opportunities for our product to be received more

openly.

Our market analysis will include:

• Economics calculations regarding income and expenses.

• Environmentally friendly solutions.

• Research into closest competitors and existing similar products.


8 | P a g e

Delimitations

Software: Because we do not have any relevant programming knowledge, any eventual software, meant to

control any component in the system, will only be delimitated. Only the overall function of the

software will be stated.

Civil engineering limitations: Our system will be designed with ducts supplying air to the individual rooms of the household. The

overall dimensioning of the ducts, together with their effect on airflow will be established, however,

because we lack relevant civil engineering knowledge, their specific layout in a house and their effect

on the house structure and walls will not be calculated or specified.

House and room size limitations: Because we intend to design this device for individual household rather than large factories or

supermarkets, our system will be designed to work in a household with a floor space of up to 200m2.

Fluid property limitations: We expect our device to function on altitudes lower than 1000 m, with clean (not heavily polluted)

air, and that clean water is supplied to the system.

Extreme weather conditions: Our system will not be designed to withstand extreme weather conditions such as floods, typhoons

or extreme wind speeds (tornados).

Choice of Model and Method We strongly believe that the key for successful project is cooperation between team members. That

is why we are going to discuss all the problems and possible solutions. Every important decision will

be brainstormed and input of each team member will be appreciated. In the process of choosing the

ultimate solution we will use various tools. Mostly we will use systematic evaluation. It provides a

wide overview of the problem, and due to grade and weight system it enables to select the most

suitable solution. We plan to use it for such problems like choosing the best design based on

selected criteria. The main theme of the semester is Economics and Thermodynamics. To make sure

that we won’t extend project too much, we are going to use concentric circles with categories must

be included in the project, should be included in the project and might be in the project. Acquired

knowledge from previous semesters and from this semester, will be put in use of development of

the product. Economics should help in regards of calculating costs and analysing the market for the

product, whilst Thermodynamics will help with the fluid dynamics, thermodynamics and heat

transfer. Autodesk Inventor is used to make designs of our project while MathCAD and Matlab are

used for the calculations. Any graphs required for the report will be created in Matlab, MathCAD and

Graph.


9 | P a g e

Gantt Chart

Research The instructions from UniClimat were to create a highly efficient heating system. The group’s

research showed that some variant of heat-pump would be best suited for this, as they generally

have high efficiency compared to alternative heating methods. The next approach was therefore to

analyse the different types of heat-pumps currently in existence and select an approach for the

project.

Heat pump types Air source heat pump - A cheap and quick-to-install system with a relatively short lifespan (20

years). However, it only operates under reduced efficiency during the winter, due to the low

ambient temperatures outdoors. 52000-114000DKK (ICAX, 2015)

Ground source heat pump - Requires a high initial investment cost, but is one of the most efficient

heating systems available. Usually a 20-30 year lifespan, 114000-175000 DKK (Comfort Pro, 2015)

Exhaust air heat pump - Extremely cheap to operate with a relatively long lifespan (25 years), and can improve the quality of the air indoors. 52000,- DKK (Designing Buildings Ltd., 2016)

Water source heat pump - High installation cost with an extremely high performance efficiency., up to 20 years lifespan, 60000,- DKK (The Green Age, 2015)

Hybrid heat pump - Ideal for cold winters but not freezing temperatures. No external heat source needed. Short Lifespan (10-15 years). 18000,- DKK (American Heating & Air Conditioning Co., 2015)

Solar-assisted heat pump - Provides hot water whenever it is required and can work in cold areas. Low installation costs, low maintenance, and requires no external heat pump. (25 years lifespan), 7000,- DKK (Clear Sky Energy, 2017)

Advantages of the Heat Pumps

• Worth to invest in long term

• Low maintenance cost

• Long lifespan

• Can provide cooling whenever it is needed


10 | P a g e

Disadvantages of the Heat Pumps

• High initial cost

• Difficult to install

• Less efficient in cold areas but there are solutions for this problem

(GreenMatch, 2017)

Horizontal Ground Source Heat Pump

The horizontal version of the system consists of an arrangement of pipes covering a wide surface area only a short distance under the ground. Its heat is drawn from the ambient temperature of the ground as it is heated by the sun.

The main advantages of this system is that it is slightly cheaper than the vertical system. It does, however, render the ground above it useless for construction, and is therefore (as our survey showed) less marketable than the vertical version.

Vertical Ground Source Heat Pump This system consists of a single pipe extending deep underground, drawing its temperature from the

ambient temperature of the earth.

This deck requires a greater initial investment than the horizontal version due to the drilling

required, however it has a greater consistency of efficiency and causes less disruption (Johnson,

2015).

It was finally decided that, judging by the survey and market analysis results (discussed in the economics section), the vertical ground source heat pump should be selected as the preferable option.

Project Limitations & Assumptions An important consideration for the project was to decide on a target for it. What limitations

could/should the project be allowed to have?

The House The ‘target’ house for this project was decided to have a floorspace of approximately 200m2, as this

is the around the average size for a detached house in Denmark. It was assumed to have 4 rooms,

and have a yearly energy consumption of 75kWh per m2 (or 15000kWh if new). (Dongenergy, 2017)

The Environment Several assumptions were made about the environment in our target region (Denmark). The

minimum temperature that could be reached in Denmark was assumed to be -15oC, whilst the

maximum was assumed to be 30oC. As a safety measure, however, -20 oC and 35 oC were

calculated. This meant that the project would be optimised to work at maximum efficiency between

these temperatures, and should the environmental temperature fall or rise below these limits, the

system could not be guaranteed to work at maximum performance.


11 | P a g e

System Operation The system was assumed to be in operation for between 1700 and 2300 hours per year.

Final Design System Layout

Figure 1 - Diagram for system

The main function of the heat pump is to transfer heat from one place to another. In this case, the

main heat source is the ground. The overall system, as shown on Figure 1, contain numbers

indicating which part being described.

(1) Piping runs from a water pump inside the heat pump to down into the ground, collects heat, and

then transfers it back into the heat pump.

(2) Here the warm water is run into a heat exchanger, where the heat is collected and the water is

once again cooled down before being sent to the ground again.

(3) In this loop, there is also a pressure relief valve in place, to prevent damage in case the pipes get

clogged up. A pressure gauge is also in place, for the operator to determine if the system is

functioning properly.

(2) The heat collected in the heat exchanger is transported around by a refrigerant under pressure.

(4) This system starts from a compressor to a relief valve to a 4-way valve. The main purpose of the

relief valve, is to be a safety factor in case the system gets clogged up. The refrigerant is relieved to

the atmosphere, as it is completely safe. The 4-way valve is in place to be able to change the

direction of flow. This is especially important when going from colder temperatures in the winter to

warmer temperatures in the summertime, as this is switched and the entire system after the valve is

reversed. The arrows marked with green on Figure 1 is reversible.

1

2

3 4 5

6


12 | P a g e

(5) If the temperature inside the house is required to be higher in the winter, the refrigerant is first

directed to the radiator placed inside the house. The radiator then acts as a condenser and a fan is

now able to transport the heat away from the radiator. From the radiator, a pipe now transports the

refrigerant back to the heat pump, where it is connected to an expansion valve. This ensures that

only a set flow rate of refrigerant is able to pass through at any given time, so that the pressure is

just not high enough for the refrigerant on the other side of the valve is exactly not superheated.

This improves efficiency of the system and protects the compressor from being flooded. The thermal

sensing probe is connected to the return part of the 4-way valve.

(6) The temperature is sensed here at this point, will always be the temperature after the

evaporator, no matter which direction the flow is directed by the 4-way valve. From this point, the

refrigerant goes to the heat exchanger connected to the ground loop. The heat is collected from this

point and carried forward to the 4-way valve. From this point, the flow is directed to the

accumulator, where excess, liquefied refrigerant is stored and dried. The type of accumulator used

here is of a suction type. The compressor, which is placed right after the accumulator, sucks just the

amount of refrigerant out it needs to compress.

(6) If servicing needs to be done, there is a filling valve and a bleeding valve in place. These make it

possible to fill the system with new refrigerant whenever needed. If the system is no longer wanted

to be taken in use, the bleeding valve can be used to drain the remaining refrigerant for recycling

purposes.

When reversing the system through the 4-way valve in the summertime, the refrigerant first goes to

the ground loop where the temperature is colder than on ground level, and the remaining system

simply runs in reverse.

Components Summary

Compressor

When selecting the compressor, several factors had to be determined first. Most important of all, was the overall capacity of the needed device. This value was calculated based on how much heating and compressing the refrigerant needed, in order to enable the system to function as intended at a high COP. Furthermore, to ensure lower energy consumption and flexibility for the consumer, the inverter type of compressor was selected. This variant enables the compressor to vary its output according to a specified need at any given time, opposed to traditional compressors, which only have an on or off state. The worst-case scenario, where the given house must be heated to normal room temperatures in the winter, the capacity on the compressor was designated to be 2,9 kW, according to (Appendix pg. 19).

The compressor series VTZ from Grundfoss provides a range of different, adjustable compressors, which are all operated by specialized controllers, to maximize efficiency according to consumer needs. Looking at the product range, the GTZ 038-G (Danfoss), provided a capacity which exceeded the worst-case scenario by a margin, but at the same time being able to go sufficiently under that target, in order to make the entire system function seamlessly according to different consumer needs all year round.


13 | P a g e

Figure 2 - Danfoss VTZ compressor

Heat Exchanger

The main heat exchange unit, which main function is to exchange the heat between the ground loop, and the refrigerant loop. According to calculations in (Appendix pg. 14), the cooling capacity of the unit must be at least 18.4 kW. According to the manufacturer Kaori (Kaori), who is the manufacturer of the selected exchange unit, their K-series are specially well suited to transfer heat between refrigerants. Hereby the selected system refrigerant, R134a, is also stated to be especially suitable for application. As seen in (Appendix pg. 17), the dimensions of the wanted unit can be determined by a set of factors, set by the manufacturer, according to wanted capacity. This includes

the overall length of the unit, determined by number of plates. The manufacturer was consulted with the desired length and pressures used in the system, and the unit variant K-105 with 445 plates was approved.

Accumulator The accumulator unit is responsible for collecting and drying the gas coming from the main heat exchanger. This is done in order to protect the compressor from being flooded with liquid, hence maximizing efficiency and potentially preventing premature failure. The minimum size of accumulator was calculated in (Appendix pg. 12), which set a minimum volume of 0,8L. The selected accumulator is able to accommodate that amount, by containing 1L. The unit is able to withstand the desired temperatures ranging from -20 up to 100oC and pressures up to 210bar.

Thermostatic Expansion Valve When the refrigerant is under pressure and at superheated temperatures, right after the evaporator, the superheat is no longer needed when going to the condenser. The thermostatic expansion valve, is a device which regulates flow in such a way that the superheat is kept on only one of the sides of the valve at a time. This is done by measuring a temperature from a point right after the condenser, with a capillary tubing with pressurized vapor inside. The vapor expands and contracts according to the measured temperature, which in turn presses or contracts a membrane inside the valve itself. This membrane is what controls the allowed flow. By adjusting the valve according to refrigerant properties, it is possible to have one side, which is exactly not superheated, which is what is wanted, as this optimizes the heat dissipation in the main heat exchanger, which in turn brings up the overall CoP of the system. The selected expansion valve is manufactured by Grundfoss and is from their TGEN 4,5 series. This was selected based on the calculations in (Appendix pg. 13). What makes this series from Grundfoss especially suited for this system, is the fact that both openings are the same size, along with the internals which are designed in such a way, that it can be used in both directions. This is important because the system is designed to be reversible.


14 | P a g e

Figure 3 - Danfoss thermostatic expansion valve

Control Panel Due to the safety regulations in most developed countries, an emergency stop button had to be

included in the system’s controls. This was a simple component to include, simply included

somewhere close to the power supply to allow it to break the circuit when required. The Schneider

40mm Electric Harmony Emergency Button (Manufacturer Part Number ZB4B-S844) was selected,

due to its relatively low cost and ease of installation.

The remaining control for the system was provided by the compressor itself, with the integrated

control panel that came as part of the system. According to the manufacturer, the panel allows for

manual control over pressure, as well as allowing an automated adjustment of performance based

on an external temperature sensor. Should the user wish to utilise this feature, temperature would

be measured by an RS Pro K-Type Thermocouple (RS Part Number 814-0131). This sensor was

selected due to its low cost, alongside the fact that it reaches its optimal resolution at around room

temperature. These types of sensors also perform well in atmospheric conditions. This would be

contained in an AN Electric Box (Manufacturer part number AN-2851).

Materials Surface Heat Exchanger Piping It was extremely important to use the right materials for the system’s components to reduce the risk of the it failing. Therefore, se-veral criteria were set to help to mitigate the chances of this occurring. The section of piping positioned inside the heat exchanger had to meet the following criteria:

• Easy to install

• Corrosion resistant

• Cost effective

• Heat conductivity

• Low maintenance

Therefore, Stainless steel, Copper, Silica Glass, HDPE, Titanium, PVC were analysed for their effectiveness according to these conditions.

Using the software Granta CES EduPack 2016, which was provided by VIA University College, we could do more research about these materials.


15 | P a g e

Stainless

steel Copper High Density

Polyethylene (HDPE)

Titanium Poly Vinyl Chloride (PVC)

Advantages High corrosion resistance, resistance to fire and heat

Reduced cost for installation, Conductivity, Less material, Efficient

Resistance to environment, High pressure resistance, Can be fused, Great for all types of loops

Chemical and corrosion resistance, high strength

Price, Resistance to environment, Widely available

Disadvantages High cost May corrode and leak

Odor/taste from, Upfront cost

High cost, reactive at high temperatures

Poor bending (no slinky), Hard to seal

Table 1 - Pipe materials advantages vs disadvantages

Comparison of physical properties The numbers in the table show the average value for the properties across any variants of the material.

Data from Granta CES EduPack 2016

Stainless steel

Copper Titanium PVC

Thermal conductivity [W/m*°C]

18 250 10.5 0.208

Density [kg/m^3]

7850 8940 4600 1430

Price [DKK/kg] 44.1 46 146.5 12.7

Specific heat capacity [J/kg*°C]

490 380 560 1400

Recycle yes yes yes Yes

Maximum service temperature [°C]

785 251 400 64.8

Tensile strength [MPa]

1360 235 965 51.5

Yield strength [MPa]

735 122 750 42.9

Thermal expansion coefficient [µstrain/°C]

16.5 17.4 9.45 122

Young’s modulus [GPa]

199.5 130 105 3.14

Table 2 - Properties of pipe materials


16 | P a g e

Thermal conductivity is obviously by far the most important factor in this material. A high thermal conductivity will significantly reduce the size of the heat exchanger required, thereby decreasing the production, material and installation costs of the part, as well as making it smaller and more compact to allow it to fit into a small building as per our project description.

The workability of the material was also important, as an exceptionally hard material would mean that the product incurs additional costs during the production stage. The material shouldn’t, however, be too soft, as this would mean that the system incurs a greater chance of breaking or failing, which could cause physical harm to the user.

The maximum service temperature of the material should be above around 60°C to ensure that the material doesn’t experience any negative effects due to the high temperatures in the heat exchanger. Theoretically, all the selected materials should perform safely in the environment, however it may be possible that an uneven temperature distribution causes damage to materials with maximum service temperatures too close to this value, meaning that the plastic is potentially a risky choice as far as materials are concerned.

Copper was finally chosen as the material for the heat exchanger piping. Its high conductivity was ideal, and its price isn’t prohibitively high compared to the heat transfer efficiency increase caused by its inclusion. Titanium was, as a material, too insulative and too expensive to be used in the heat exchanger. While its high strength and low density are useful properties, they are not high priority, and therefore Titanium is poorly suited for the required purposes. The plastics had similar issues, their low thermal conductivity being a huge barrier to creating a compact and efficient heat exchange system. Stainless steel was not a bad choice as far as materials went, however it had no significant advantages over copper aside from its strength, and again its relatively low thermal conductivity meant that it was unsuitable choice for a component so reliant on a fast, efficient heat transfer.

Geothermal Heat Exchanger Piping It is extremely important to use the right materials for the parts that are being used, to avoid failing of the system. Hence some criteria were set to be considered. The part which is going into the ground should meet as many of the following criteria as possible:

• Efficient

• Easy to install

• Corrosion resistant

• Low maintenance

• Cost effective

• Long life-span

Therefore, these conditions were used to research the materials, and found out the main pipe types are PE, HDPE, PVC, cPVC and copper for the geothermal heat exchanger (Geothermal Pros and Cons, 2012). Fortunately all of them are corrosion resistant, so they are all ideal for the piping system.

Using the software Granta CES EduPack 2016, which was provided by the school, more relevant research about these materials was done. Unfortunately HDPE and cPVC are not included in this software.

Copper

Polyethylene (PE)

High Density Polyethylene (HDPE)

Poly Vinyl Chloride (PVC)

Chlorinated Poly Vinyl


17 | P a g e

Chloride (cPVC)

Advantages Reduced cost for installation, Conductivity, Less material, Efficient p.62

Slow corrosion, Can be fused

Resistance to environment, High pressure resistance, Can be fused, Great for all types of loops

Price, Resistance to environment, Widely available

Pressure, Price, Availability

Disadvantages May corrode and leak

Not high pressure (no vertical loops), Expensive

Odor/taste from, Upfront cost, Bending radius (no slinky) p69

Poor bending (no slinky), Hard to seal

Resistance to environment Hard to seal

Table 3 - Advantages and disadvantages of ground loop materials

(Geothermal Pros and Cons, 2012)

Copper Copper, with its high thermal conductivity, was a good choice for the material. Despite its high cost

per kg, a system made of copper would actually have a lower cost than one made of most other

materials, as its high conductivity meant that it could be shorter in length, and therefore use less

material in its creation. Copper is, however, vulnerable to corrosion, especially in damp and

acidic/alkaline environments (such as underground) and may therefore require some kind of coating

to protect it from the elements.

Polyethylene (PE) While PE can be manufactured with different geometries, and is widely used for horizontal and helical structures, PE is not able to withstand the higher pressures present in a vertical loop. Furthermore, the resistance to corrosion is fairly good, yet not as good as some other materials.

High Density Polyethylene (HDPE) Compared to normal PE, HDPE cannot be bent, or moulded with any special geometries. This makes it impossible to use for a helix configuration. But looking at other properties, it is especially good at withstanding large pressures, which makes it ideal to use for vertical loops. Furthermore, HDPE excels in withstanding environmental impacts, which enables it to function without maintenance for at least 50 years. Noteworthy to mention that the pipes are not able to be glued together, but need to be assembled through heat fusion, which calls for specialized personnel equipment. This results in essentially one single piece of pipe. Though because of these superior properties, the initial cost is higher than other types of material. But taking longevity into consideration, over time it will earn the initial premium back through the lack of servicing and exchanging for new pipes.

Poly Vinyl Chloride (PVC) Cheap, easy to assemble and superior resistance to different environments. Where this pipe type falls short is in bending radius and ability to withstand high pressures. This makes it only suitable for horizontal layouts consisting of straight parts.


18 | P a g e

Chlorinated Poly Vinyl Chloride (CPVC) CPVC is similar to standard PVC pipes, but trades environmental resistance with the ability to carry pressurized substances due to its increased strength.

Composite Piping Another potential solution was to use a composite of different materials on the piping, such as using a copper interior with a thin PVC layer over the outside surface. This would, in theory, allow the application of the copper’s superior thermal conductivity, whilst still protecting it from corrosion and damage.

Several combinations of materials were analysed, however it was eventually decided that the negatives of using this kind of pipe construction far outweighed the positives. A composite pipe would likely be more expensive, due to it being a more complex construction. Several of the combinations of materials also raised new issues such as galvanic corrosion or significantly reduced thermal.

Comparison of Physical Properties As two of the beforementioned materials are variants on two of the other ones, the comparison below is only done for the overall types, as there are only minor differences between the variations. The numbers in the table show the average value for the properties across any variants of the material.


Copper PVC PE


250 0.208 0.419

Density [kg/m^3] 8940 1430 949

Price [DKK/kg] 46 12.7 15

Thermal conductor or insulator

Good conductor Good insulator Good insulator


380 1400 1880

Recycle yes yes Yes


251 64.8 99.5

Tensile strength [MPa] 235 51.5 30.5

Yield strength [MPa] 122 42.9 22.8


17.4 122 158


130 3.14 0.7585

Table 4 - Material properties

The price of the material is naturally an extremely important factor in the material selection. While not wholly representative of the final cost of the pipes, it will be reflected in it. The copper alloys, at 46dkk/kg, are the most expensive of the three choices by a large margin.

Thermal conductivity was one of the most important factors in the calculation, as it would impact heavily on the performance of the heat pump, and thereby affect how long the underground pipe


19 | P a g e

system would have to be, which would in-turn increase the costs of the system’s installation and manufacture.

Recyclability of the material was very important to the project, as it would heavily impact the scrap value of the material at the end of the project’s lifecycle, thereby affecting its profitability. A recyclable material will have a higher, likely positive scrap value, whereas a non-recyclable material will have a lower or potentially even negative one.

Density should only have a small impact on the project, largely impacting the installation costs of the piping system by potentially requiring more heavy-duty machinery to complete the installation.

The thermal expansion coefficient is naturally important to the material, as a material that frequently expands and contracts when heated or cooled could potentially cause damage to itself due to trying to expand in a confined space and warping as a result. This could lead to leaks or holes in the piping system, reducing the potential lifespan of the project.

The tensile strength and yield strength of the chosen material could be fairly important, so as to help it avoid damage through shifting earth, tree roots and similar circumstances. The exact chances of damage occurring this way were unknown, however judging by existing products on the market it was assumed that it was worth a little consideration when selecting materials.

The maximum service temperature of the material was worth considering. It was important to consider that when the system was reversed during the summer, hot water would instead be sent through the pipes, reaching temperatures of around 40-60 degrees. Therefore the maximum service temperature had to be considered to ensure that no damage was done to the pipes.

PVC was finally selected as the material of choice, due to its low cost and compliance with the other qualities previously stated.

Heat Exchanger Case The casing of the heat exchanger was obviously an important material to select, as it was the final barrier between the heat gathered by the pump and the outside world. Naturally this means that this material could have a large effect on the energy efficiency of our project, as well as the safety of its user.

• Low conductivity (both thermal and electrical) • Corrosion resistant • Cost effective • Long life-span • Resilient to external forces • Easy to install • Easy to manufacture

Therefore, these conditions were used to research the materials PE, PVC, PP, ABS, Stainless steel for the heat exchanger case.

Using the software Granta CES EduPack 2016, which was provided by the school, more research could be done about these materials.

Selection of casing material for heat exchanger


20 | P a g e

Polyethylene (PE)

Poly Vinyl Chloride (PVC)

Stainless steel Polypropylene (PP)

Acrylonitrile butadiene styrene (ABS)

Advantages Slow corrosion, Can be fused

Cheap, Resistance to environment, Widely available

High corrosion resistance, resistance to fire and heat

Low cost, excellent moisture and chemical resistance

High heat resistance, Good impact resistance

Disadvantages Expensive, Flammable, Subject to stress cracking

Hard to seal, Intolerance of high temperatures

High cost Poor weathering resistance, oxidizes readily (Jpgreene, 2000)

Moderate heat, moisture and chemical resistance (Ashraf, 2014)

Table 5 - Casing material advantages vs disadvantages

Comparison of Physical Properties The numbers in the table show the average value for the properties across any variants of the material.


PE PVC Stainless steel PP ABS


0.419 0.208 18 0.14 0.2715

Density [kg/m^3]

949 1430 7850 900 1110

Price [DKK/kg] 15 12.7 44.1 28.7 18.8


1880 1400 490 1915 1655

Recycle Yes Yes yes Yes yes


99.5 64.8 785 107.5 69.4

Tensile strength [MPa]

30.5 51.5 1360 34.5 41.4

Yield strength [MPa]

22.8 42.9 735 28.95 34.75


158 122 16.5 151 159.3


0.7585 3.14 199.5 1.223 2


21 | P a g e

Table 6 - Specifications of casing materials

Once again, one of the most important factors in this decision was the thermal conductivity. To ensure that the casing did not become dangerously hot to the touch, it was important to select a material that would not allow an easy throughflow for such temperatures. The maximum service temperature would also be important, so as to ensure that the panel did not sustain damage due to the potentially high temperatures in the interior.

The strength of the material was also important, as any damage to the casing could present a threat to the user or the system itself. Fortunately, as the system should be static and immobile, the risk of heavy impacts is reduced somewhat, however it was still a possibility and therefore should be taken into account.

The appearance of the material, despite not being featured in the table, was also an important factor to consider. This would be one of the few features of the system that would be visible to the user, and it was therefore important to ensure that it would not look unappealing.

Stainless steel was ultimately selected for the casing, as it combined all the major properties required and, most importantly, great resistance and durability.

Production Methods Many of the production methods involved in this project would have to be custom-made or

purchased based on the purchaser’s house. The exact method of running pipes to-and-from the heat

exchanger would vary from house to house depending on, for example, the house’s layout and

building material. However, the construction elements deemed possible to predict were matched

with the most effective production methods to suit their purposes. These elements are detailed

below and in the relevant appendix.

The pipes used in the geothermal heat exchanger would be heat fused together, because HDPE does

not react well with glues. A socket fusion method would likely be used, as this would ensure a more

secure connection between the pipe sections for a relatively small price increase.

The pipes in the surface heat exchanger would be arc-welded together. This process was selected

due to its low cost and relatively high speed, thereby minimising the costs for the task and the

project as a whole.

Review of the Thermodynamic and Mechanic Calculations and

Interpreting the Results The thermodynamic calculations had a large influence on all the different aspects of the product: it

allowed the correct dimensioning of the system and the individual components in order to withstand

the loads resulting from the high pressures and to supply sufficient heating energy to the building,

provided data on the expected energy loss of a newly built building, influenced the standard

component and refrigerant selection, the geometry, the shape, the design layout of the heat pump

and the length of the ground loop and provided values for the energy consumption of the heat

pump, thus showing the highest available Coefficient of Performance.


22 | P a g e

Heat Requirement of the Building In order to calculate the necessary heat- and ground pump dimension, the energy requirement of

the building needed to be determined first. To achieve this, the following factors had to be

considered: where will the heat pump operate, what are the highest and lowest expected outdoor

temperatures in the summer and in the winter, and if the house is newly built or older with bad

insulation. Since the primary target location for marketing this device is Denmark, the lowest

temperature was selected to be -20 degrees Celsius in the winter and the highest to be 35 degrees

Celsius in the summer. These were selected as the worst-case scenarios, not as average values. This

does not mean, however, that the heat pump cannot operate with temperatures higher or lower

than those temperature values, as because an adjustable compressor is used, these values can still

be exceeded. These values were only used to optimise the system for this range, as significantly

lower or higher temperatures cannot be expected for more than a few days yearly, and can be

considered extreme temperatures in this climate. It was decided that the heat pump will be used

with newly built houses, so the corresponding insulation coefficient was used.

The overall energy requirement of the heat pump was the sum of two losses: the losses due to

insulation imperfection, and the losses due to building ventilation, which is regulated in Denmark.

The ventilation losses were, however, calculated to be significantly lower than the losses due to the

insulation, when the worst-case scenario was inspected. This means that the largest amount of

energy consumption was expected due to the temperature difference between the inside and

outside temperature of the building, which is influenced by the insulation efficiency.

The most demanding operation for the heat pump was the heating operation when it was -20

degrees Celsius outside: that is when the temperature difference is the largest between the

expected room temperature of 21 degrees Celsius and the outside temperature, thus this is the

situation when the largest amount of energy will be lost. The heat pump was therefore dimensioned

using this situation (Appendix pg. 8). However, as a safety measure, the air conditioning situation

was calculated as well (Appendix pg. 19), to ensure that the refrigeration cycle can be reversed as

expected and the system will still function on a high efficiency.

The highest power demand for a building that has an area of 200 m2 is, based on these

requirements, 20.8 kW (Appendix pg. 11), from which the most significant portion, 17.2 kW is the

non-ventilated energy dissipation due to the temperature difference. In the summer, when air

conditioning is required, the power requirement goes down to 6.7 kW in total. This decrease is due

to the smaller temperature difference between the indoors and outdoors temperature. It is

important to note, that under these conditions as well, it is the non-ventilation related heat

dissipation that causes more power loss, in that case 5.6 kW. The ventilation related losses were

calculated using the standard ventilation value for Danish households.

Afterwards, the mass flow rate of the air inside the building was determined. The heat pump needs

to circulate the air inside the building in order to keep the rooms at the desired temperature. For

these calculations, once again the worst-case scenario was calculated, meaning that the air has to be

blown through a room with large cross-sectional profile. Because the mass flow rate is constant in

the loop, this allowed the calculation of the increased air velocity inside the reduced cross-sectional

profile of the heat exchanger, where it is heated up or cooled down to the desired temperature by

heat dissipation between the air and the refrigerant inside the exchanger. As the air reaches the

optimal room temperature, it leaves the heat exchanger, decelerates with the increase of the cross-

sectional profile, and dissipates the heat in the building. Once the heat is dissipated, the air returns


23 | P a g e

to the heat exchanger. The optimum room temperature was assumed to be between 21 and 23

degrees Celsius. It was determined that with the current selected mass flow rate, the building can

sufficiently be heated up from 5.8 degrees Celsius (Appendix pg. 11).

Lowest expected outdoor temperature -20 oC

Highest expected outdoor temperature 35 oC

Room temperature 21 oC

Power loss in the winter 20.8 kW

Power loss in winter due to temperature difference 17.2 kW

Power loss in winter due to ventilation 3.6 kW

Power loss in the summer 6.7 kW Table 7 - Summary table for data presented in the section

Heat Exchanger and Compressor Properties Once all properties of the air flow were determined, including the Reynolds number that indicates

the flow type, the design and the dimensioning of the heat exchanger was possible. Based on the

refrigerant selection, some properties were already known. It was known that the desired air

temperature after the compressor is 50 degrees Celsius at 10 bar, and that decreases in the

condenser to 35 degrees Celsius at the same pressure. On the other side of the loop the pressure

should be 2 bar, with a temperature of 10 degrees Celsius (which matches the ground temperature)

before the compressor and -10 degrees Celsius after the expansion valve. The selection of these

values is further explained in the part Refrigerant selection.

From this data the enthalpy values for the points before the compressor, after the compressor,

before the expansion valve and after the expansion valve were determined, that were used to find

the most optimal mass flow rate this heat dissipation. This mass flow rate, however, is only a

guideline for the further calculations. If the mass flow rate is increased, the temperature difference

in the condenser will be smaller for a given power output, and if the flow rate is decreased, the

temperature difference will increase. This allowed to change the mass flow rate to a value that is

more suitable for further parts of the calculations, while only influencing the temperature difference

on the condenser at a constant power output. Using the mass flow rate and knowing the outlet

diameter of the compressor, the required compressor velocity flow rate was calculated to be 0.006

m3/s (Appendix pg. 12). This value is the maximum flow rate requirement for the compressor.

Mechanics calculations were also performed on the tube coming out of the compressor (Appendix

pg. 10) in order to ensure that the sizing of the wall thickness is properly done and the tube will not

yield or break under the stresses that are present because of the high-pressure fluid moving through

the pipe. This was calculated using the yield strength of the material (copper) selected for our piping

and the highest allowable pressure in the system, which occurs on the condenser side. This way, the

smallest allowable wall thickness was calculated, which is still able to support the high pressures.

Pressure in condenser 10 bar

Temperature change in condenser 15 oC

Pressure in evaporator 2 bar

Temperature change in evaporator 20 oC

Flow rate of compressor 0.006 m3/s


24 | P a g e

Minimum tube wall thickness 5.2*10-5 m Table 8 - Summary table for data presented in the section

Indoor Heat Exchanger Dimensioning

Heat Transfer Coefficient Based on the required mass flow rate in the compressor and knowing the number of plates in the

heat exchanger, it was possible to calculate the flow through each individual plate in the heat

exchanger (Appendix pg. 12). The flow coming from the compressor pipe is distributed through the

plates: meaning if there are 50 plates, the flow coming out of the compressor is divided into 50

separate flows through the plates, before the flow is once again unified at the end of the heat

exchanger, where the fluid flowing through the plates are directed into a tube, leaving the heat

exchanger and directing the flow towards the expansion valve. Because this way the fluid flow is

divided into n (n represents the number of plates) parts and therefore the area is increased as well,

the fluid velocity is significantly reduced.

The main goal of the heat exchanger calculation was to determine the required surface area, with

which the required energy can be dissipated (in the winter) or absorbed (in the summer). Because

the required energy dissipated is larger, this was used to determine the minimum size of the heat

exchanger.

One of the most important factors in determining heat dissipation was the heat transfer of the fluid.

The faster the heat transfer, the less time it needs to stay in the heat exchanger or the smaller the

heat exchanger can be. Therefore, the goal was to have a heat transfer coefficient as high as possible

to allow high heat transfer. The factors influencing this coefficient were the thermal conductivity of

the refrigerant (which was similar between the refrigerants considered for our application), the

diameter of the flow and the Nusselt number, which is primarily dependent on the Reynolds

number. Therefore, the higher the Reynolds number, the higher the heat transfer coefficient is, if

the diameter of the tube is the same. That means that a turbulent flow is more advantageous in a

hydraulic diameter that is as small as possible. This was the main consideration is selecting the shape

and dimension of the plates in the heat exchanger: to increase the heat transfer coefficient.

Of course it is not only the refrigerant that has to dissipate or absorb heat at a high rate. The same

property is expected from the air as well, that flows through the heat exchanger, absorbs the heat

that is dissipated from the plates or dissipates it to the colder plates. A high heat transfer coefficient

is required here as well, in order to ensure that the air can gather the heat while it is inside the heat

exchanger, and reaches the required temperature (room temperature). The air flow also depends on

the Reynolds number and the hydraulic diameter, meaning the design criteria here is the same as for

the refrigerant: turbulent flow and small hydraulic diameter are advantageous. Because the air has

different density and dynamic viscosity (properties that influence the Reynolds number),

conductivity and Prandtl number than air, it meant that the air flow rate and velocity had to be

different than that of the refrigerant to achieve similar effect.

Based on these considerations, the desired heat transfer coefficient values in the system (for both

the heating purpose in the winter and the cooling purpose in the summer) are presented in the

summary table below (Appendix pg. 13).

Heat transfer coefficient for refrigerant in the summer 91.58 W/(m2*K)


25 | P a g e

Heat transfer coefficient for the air in the summer 122.45 W/(m2*K)

Heat transfer coefficient for refrigerant in the winter 39.76 W/(m2*K)

Heat transfer coefficient for the air in the winter 277.27 W/(m2*K) Table 9 - Summary table for data presented in the section

Plate Dimensioning in the Heat Exchanger The main goal of the plate dimensioning was to calculate the necessary total surface area of the heat

exchanger in order to select a suitable heat exchanger for this application (Appendix pg. 14). An own

heat exchanger design with the most suitable dimensioning for this application is also presented in

this section. This showed promising results and managed to be smaller thus more efficient than

standard parts for this purpose, however, because of the limitation of time and because the project

focuses on the overall design, this design was not implemented in the final design, but the findings

will be presented in this section as a possibility to further improve the system in the future.

The required surface area of the heat exchanger was calculated based on the shape, dimensions,

average temperature difference, heat transfer coefficient and thermal conductivity of the

refrigerant, air and plate material and the required heat dissipation.

The average temperature difference is the difference in temperature between the average inside

temperature of the fluid in the condenser and the average air temperature running through the heat

exchanger. The largest temperature difference is in the winter, when it can be as high as 28.1 oC. In

the summer it is not expected to go over 23.8 oC. Knowing the average temperature difference and

the heat requirement of the building, the total thermal resistance for the building was calculated.

The lower the resistance is, the larger power dissipation is possible using the same temperature

difference.

This total thermal resistance was divided up into 3 components: the overall resistance consists of the

convection resistance of the refrigerant, the conduction resistance of the metal and the convection

resistance of the air. Each of these resistance values depend on the dimension of the system and

material properties as well: for the convection resistance, the heat transfer coefficient of the fluid

and the surface area is required – the larger these values are, the smaller the resistance, thus more

heat can be dissipated. In terms of conduction, the surface area and the metal conduction

influenced the resistance – similarly, the larger these values are, the lower the resistance is.

The heat transfer coefficient and the conductivity was already calculated or given, and the surface

area also had to remain the same: even though increasing this property would have further

decreased the resistance, because it also affects the heat transfer coefficient, simply selecting a

larger area wouldn’t have been advantageous. After trying out various shapes and dimensions for

the heat exchanger, the most advantageous shape was found to be a rectangular cross-section (this

was preferred over a circular tube), where one dimension of the cross-section is significantly larger

than the other. This resulted in a thin, long rectangular plate with the refrigerant running through it.

It was calculated that in the winter, the necessary surface area for the heat exchanger is 14.13 m2

(Appendix pg. 14). Instead of creating a single large plate with this surface area, the dimensions of

the plates were constrained, meaning that the thickness and the width of the plate were already

given, and a decision was made to limit the length in 0.5 m in order to avoid having an extremely

long heat exchanger that doesn’t fit in the building. This way, it was possible to calculate how many

pieces of plates are required in the heat exchanger by just dividing the overall length with the


26 | P a g e

limited length. If the length is limited to 0.5 meters, a minimum of 28.15 plates are required,

meaning 29 plates will be required for the heat exchanger element. This way, the same required

surface area will be reached but in a more compact, square layout to allow easy storage and small

space requirement in the building. It will be a 0.5x0.057x0.5 meter box with a volume of 0.014 m3.

This was the largest required dimension for the heat exchanger. In the summer, a much smaller unit

with the dimensions of 0.5x0.033x0.5 meters and a volume of 0.0083 m3 would have been sufficient

with 17 plates (Appendix pg. 25), however, the heat exchanger was designed to perform under the

worst-case scenario in the winter, meaning the larger dimensions were selected.

The required minimum plate thickness was also calculated (Appendix pg. 10) to ensure that the

plates in the heat exchanger can support the pressure of the fluid without yielding or breaking. For

these calculations, the yield strength of the plate material (copper) and the maximum system

pressure of 10 bar was used to ensure a sufficient wall thickness for the used cross-section. Based on

these calculations, a wall thickness of just 1.633*10-5 meters would be enough to support the

pressure. Having determined this data, the outside dimensions of a plate element were calculated.

Total thermal resistance of the plate 1.35*10-3 oC/W

Minimum plate wall thickness 1.633*10-5 m

Required surface area of heat exchanger 14.13 m2

Number of plates required in the heat exchanger 29

Height of heat exchanger 0.5 m

Length of heat exchanger 0.057 m

Width of heat exchanger 0.5 m

Volume of heat exchanger 0.014 m3

Table 10 - Summary table for data presented in the section

Compressor Calculations The compressor calculations (Appendix pg. 13) were necessary in order to calculate the required

maximum compressor output power for the heat pump. The power requirement of the compressor

was one of the most important technical data of our system, because this value directly influenced

the value of the Coefficient of Performance, which indicates the efficiency of the system: the higher

the CoP, the higher energy dissipation is possible while using the same compressor power to drive

the system. This means that the compressor power consumption should be as low as possible for the

heat pump.

This number was calculated based on two factors: first and foremost, the power output is necessary

to increase the enthalpy (the pressure and the temperature) of the refrigerant, so the energy stored

in the refrigerant can be dissipated with higher temperature differences. Secondly, due to the high

velocity fluid flow in the pipes, the pressure losses due to the friction between the fluid and the pipe

has to be recovered by the compressor. The minor losses resulting primarily from the bends in the

pipes were excluded from the calculations, because the constant required to calculate this constant

can only be determined using experiments: due to the lack of data of this constant, no exact

calculations were possible. However, it is worth noting that these minor losses are usually negligible

compared to the frictional losses, and because the current heat exchanger layout consists of plates,

there are far fewer minor losses than in a traditional coil type heat exchanger.


27 | P a g e

As mentioned in the paragraph above, the primary source of the power requirement is the need to

increase the enthalpy of the cold refrigerant at low pressure at the compressor inlet so that the

compressor can deliver the required high pressure and high temperature refrigerant on the outlet.

The compressor power output requirement is the same in the summer and in the winter: because

the refrigerant flow is simply switched around using a 4-way valve, the enthalpies before and after

the compressor remain the same, only the condenser and evaporator switch roles. The enthalpy

values were optimized for the winter, because that is when the highest power output and therefore

the greatest efficiency are required: as mention in the section Refrigerant selection, the goal was to

create a large temperature difference at relatively low pressure. This way the refrigerant was hot

enough to allow fast and efficient heat transfer to the cold air, while its enthalpy still hasn’t

increased drastically, because the pressure difference hasn’t increased drastically compared to other

refrigerants. And lower enthalpy differences result in lower compressor output power, increasing

the system efficiency. And because the evaporator side has sufficiently low temperatures (the

refrigerant evaporates at -10 oC at 2 bars) this meant sufficient temperature differences between the

refrigerant and the room temperature in the summer, when the flow is switched and the air

temperature of over 20 oC has to be cooled using the refrigerant at -10 oC. The same 30 oC

temperature difference occurs on the ground loop side, where the fluid at over 40 oC is cooled down

using ground water of 10 oC.

Therefore, the enthalpy difference that the compressor has to create is 2.11*104 J/kg both in the

summer and winter season. In order to determine the total power requirement, this value has to be

multiplied by the mass flow rate, which shows how much energy input is necessary when a specified

amount of mass (measured in kg) flows through the compressor in 1 second. The required mass flow

rate was already calculated before in order to ensure the required energy dissipation in the heat

exchanger. Using the enthalpy difference and the necessary mass flow rate, the required compressor

output power was calculated to be 2.416 kW (Appendix pg. 13).

This is, however, only the useful power output of the compressor. It also needs the compensate the

losses that result from the friction between the fluid and the surface. This decreases the pressure in

the pipe, which the compressor has to regenerate constantly. These frictional losses increase with

increasing surface roughness, increasing velocity and a large length-diameter ratio. An easy way to

represent these losses is to calculate the head loss in meters, which shows how many meters of fluid

flow are lost. The total head loss in the two heat exchangers is 0.34 meters. From this, the total

power loss of 532 W can be calculated (Appendix pg. 19).

The summation of the useful power output and the frictional losses results in the required maximum

compressor output power of 2.948 kW, which will be sufficient to power the system. This is the

power that has to be supplied to the system at the specified maximum volumetric flow rate of 0.006

m3/s.

Enthalpy difference 2.11*104 J/kg

Useful power output 2.416 kW

Head loss in the system 0.34 m

Frictional losses in the system 532 W

Total power output of compressor 2.948 kW

Maximum volumetric flow rate in compressor 0.006 m3/s Table 11 - Summary table for data presented in the section


28 | P a g e

Ground Loop Pipe Dimensioning

Energy Requirement from the Ground The outdoor heat exchanger is the heat exchanger that has the role to absorb or dissipate the heat

coming from the water in the ground loop. In the winter, it absorbs heat from the water, the

compressor increases its enthalpy, and this heat that was gained from the ground and amplified by

the compressor is dissipated to the air in the house. In the summer, the cycle is reversed: the hot,

high pressure fluid dissipates energy to the water and the cold refrigerant flows towards the house,

where it is cooled even further by passing through the evaporator that decreases its pressure and

temperature even further. Therefore, this unit fulfils the function of the inlet unit the uses heat from

the ground to gain or dissipate energy. The total energy output on the heat exchanger inside the

house is the summation of the energy coming from the ground-side heat exchanger plus the

compressor output power that further increases the gained energy. Therefore, the power

requirement from the ground (the energy that needs to be absorbed) is the required energy output

in the house minus the power output of the compressor. The difference has to be filled out by the

energy absorbed from the ground or dissipated to the ground. The power requirement from the

ground was calculated to be 18.4 kW (Appendix pg. 14). This value was used for further

dimensioning of the system, since this can be considered the worst-case scenario, because the

power requirement in the summer is significantly lower. In the summer the ground only needs to

contribute with 6.1 kW to the total power requirement: this is possible because the required power

output on the indoor heat exchanger is significantly lower than in the winter.

Maximum power requirement in the winter 18.4 kW

Maximum power requirement in the summer 6.1 kW Table 12 - Summary table for data presented in the section

Mass Flow Rate and Mechanics Calculation Once the power requirement was determined, the mass flow rate could be calculated. The mass

flow rate is determined by the required enthalpy change with which the energy output should be

created. It was calculated that a temperature difference of 1oC was already enough to dissipate

enough energy to the heat exchanger with a reasonable mass flow rate. The enthalpy change

therefore resulted only from the temperature change of the water while it flows through the heat

exchanger, because the pressure remains constant during the process. This way, the mass flow rate

in the ground loop was calculated to be 4.37 kg/s (Appendix pg. 15). Due to the selected larger pipe

diameter, this resulted in a flow velocity of 5.55 m/s in the pipes. Using the area of the pipes

underground and the previously calculated flow velocity, the volumetric flow rate was calculated to

be 2.813*10-3 m3/s. Using this volumetric flow rate, the flow velocity between the plates of the heat

exchanger was calculated to be 0.049 m/s due to the increase in the overall area.

The mechanics calculations were also performed for the pipes of the ground loop. The goal of these

calculations was to ensure that the underground pipes are adequately dimensioned, meaning that

their wall thickness is large enough to withstand the pressure of the water flowing through it. To

determine the minimum wall thickness, the maximum expected system pressure of 1 bar, the yield

strength of polyethylene and the required inside diameter were used. Based on these calculations, it

was determined that a wall thickness of just 3.175*10-5 meters (Appendix pg. 15) is enough to

withstand the low-pressure fluid flowing through without yielding or breaking.


29 | P a g e

Mass flow rate in ground loop 4.37 kg/s

Flow velocity in ground loop 5.55 m/s

Volumetric flow rate 2.813*10-3 m3/s

Flow velocity of water in the plates 0.049 m/s

Minimum wall thickness of plates 3.175*10-5 m Table 13 - Summary table for data presented in the section

Outdoor Heat Exchanger Dimensioning

Heat Transfer Coefficient The principle of determining the heat transfer coefficient in the heat exchanger that is connected to

the ground loop is the same as the procedure of determining the heat transfer coefficient for the

heat exchanger inside the house. The larger the heat transfer coefficient, the better the heat

transfer through the fluid is, meaning it can dissipate energy more efficiently. The factors influencing

the heat transfer were discussed in detail at the section Indoor heat exchanger design. The most

advantageous situation is when high velocity fluid with good thermal conductivity flows through a

small diameter. These result in large Reynolds and Nusselt numbers. It can be determined the heat

transfer coefficient is lower in the winter than in the summer, when the heat exchanger is used a

condenser. This means that the heat exchanger had to be designed for winter operation when it acts

as an evaporator for two reasons: the heat transfer is less efficient and the system requires more

energy intake from the ground.

The heat transfer coefficients for the water were determined using the same method. The goal was

to allow high velocity fluid flow in small hydraulic diameter between the plates to increase the heat

transfer coefficient. The heat transfer coefficients for the heat exchanger in the ground loop are

presented in the table below (Appendix pg. 16).

Heat transfer coefficient for refrigerant in the summer 81.456 W/(m2*K)

Heat transfer coefficient for water in the summer 3.244*103 W/(m2*K)

Heat transfer coefficient for refrigerant in the winter 44.713 W/(m2*K)

Heat transfer coefficient for water in the winter 345.297 W/(m2*K) Table 14 - Summary table for data presented in the section

Plate Dimensioning in the Heat Exchanger In the ground loop heat exchanger, different dimensions were required than in the house heat

exchanger. Smaller dimensions are adequate for this heat exchanger because the energy output in

the house is the summation of the energy that the refrigerant absorbs from ground loop heat

exchanger and the compressor. This way, the required power intake at the ground loop heat

exchanger is the power requirement in the house minus the compressor power output. This reduced

power requirement was already calculated to be 18.4 kW in the winter.

The sizing of the heat exchanger was done using the same method as for the dimensioning of the

heat exchanger in the building. The average temperature difference was calculated between the

average refrigerant temperature in the plate heat exchanger and the average water temperature

flowing around the plates in the heat exchanger. This temperature difference was calculated to be


30 | P a g e

9.5 oC. Using the average temperature difference and the known power intake requirement of the

plate type heat exchanger, the total thermal resistance of 5.164*10-4 oC/W was calculated (Appendix

pg. 17) for the heat exchanger. The smaller the total thermal resistance is, the larger the energy that

can be absorbed using the same average temperature difference, therefor a smaller resistance is

only advantageous.

The total thermal resistance consists of the convection resistance of the refrigerant flowing inside

the plate, the conduction resistance of the plate material (copper) and the convection resistance of

the water flowing through the heat exchanger. The convection resistance is influenced by the heat

transfer coefficient and the surface area, and the conduction resistance is influenced by the thermal

conductivity of the material and its surface area. The larger these values are, the smaller the

resistance to heat transfer is. This means that the same design consideration were used as in the

heat exchanger used inside the building. In a short summary, that means that the shape of the cross-

section played an important role here as well. Because this heat exchanger has to fulfil the role of

the condenser in the summer, the same thin plate type heat exchanger was used for this design as

well. This allowed a large surface area while keeping the hydraulic diameter of the refrigerant flow

as low as possible and so the refrigerant dissipates heat on a large surface area.

The calculations resulted – as expected – in the conclusion that the heat exchanger has to be

designed for the operation in the winter when it fulfils the role of the condenser, because this is the

case when the largest surface area, thus the largest amount of plates are required due to the lower

heat transfer coefficient and the higher power requirement. In the winter 97.449 is the minimum

number of plates (Appendix pg. 17) required to provide the required energy to the system. This

number has to be rounded up to 98 for practical use. This result was determined from the minimum

surface area requirement of 48.9 m2 using the plate dimensioned determined in the design of the

indoor heat exchanger. From this data, a compact heat exchanger with a volume of 0.05 m3 can be

designed with the following properties: the height of the heat exchanger is 0.5 meter, the length is

0.198 meter and the width is 0.5 meter.

In the summer, the required surface area and thus the number of plates in the heat exchanger is

lower, mainly because of the larger temperature difference between the average temperature of the

refrigerant inside the plates and the average temperature of the water in the ground loop. In the

summer, that difference is 32.31 oC, more than 20 degrees higher than in the winter. Using the same

equations, it was determined that the required surface area in the summer is just 2.39 m2 (Appendix

pg. 29), which is significantly smaller than the requirement in the winter. This also mean that only 5

plates are enough to dissipate the energy to the water. However, because the heat exchanger has to

be designed for the worst-case scenario in order to ensure that it can work adequately in the winter

when it works as the evaporator in the system, the larger dimensions (that were calculated for the

winter operation) will be used with the larger plate number was used.

Average temperature difference in the winter 9.5 oC

Average temperature difference in the summer 32.31 oC

Total thermal resistance of the heat exchanger 5.164*10-4 oC/W

Required total surface area 48.9 m2

Required number of plates 98

Height of the heat exchanger 0.5 m

Length of the heat exchanger 0.198 m

Width of the heat exchanger 0.5 m

Volume of the heat exchanger 0.05 m3 Table 15 - Summary table for data presented in the section


31 | P a g e

Calculating the Required Length of the Pipes in the Ground Loop One of the most important factors influencing the initial investment for the ground source heat

exchanger was the required length of the pipes underground. The required pipe length was

calculated using the design guidelines of Tennessee Valley Authority. (Geokiss, 2017) This design

guideline was written for vertical ground source heat pumps using high-density polyethylene pipes,

and because the designed ground source heat pump uses the same material and the same layout,

this paper was found to be suitable to base our pipe length on.

As a first step, the coil type was selected from the table on page 1. The ground source pipe

dimension of 0.032 meter corresponds to the Vertical U-tube (1¼” Pipe) in the table. Afterwards, the

expected ground temperature was selected. In Denmark, the expected ground temperature in

adequate depth is a constant 10 oC throughout the year. This value corresponds to 50 Fahrenheit,

meaning the second column in the ground temperature table has to be used. Then the length of

bore had to be multiplied by the trench to find the required length of the pipe per ton. The power

requirement on the ground loop was converted from Watts to Refrigeration Tons. The energy

requirement in the winter was used for this calculation, because that is when the required energy

consumption is the highest. The result of this conversion was 5.23 Refrigeration Tons. In order to

calculate the total length of the required ground pipes, the length per ton was multiplied with the

value of Refrigeration Tons. Given that the requirement is 300 Feet per Ton, the overall length

needed to be 1569 Feet long. Because the heat exchanger is dimension in SI Units, this value had to

be converted to meters. Using unit conversion from Feet to meters, it was determined that the

required total length is 478.348 meters in the ground (Appendix pg. 18).

Required ground loop length 478.348 m Table 16 - Summary table for data presented in the section

Pump Dimensioning In the ground loop of the system, it was necessary to ensure that the water is circulated through the

loop at the same constant volumetric flow rate to allow the water to gain energy from the ground

and then dissipate it to the heat exchanger.

Figure 5 - Specifications chart from pump manufacturer

In order to calculate the necessary dimensioning, it was important to describe the function of the

pump in the closed loop. The velocity of the water before and after the pump were the same, the

height before and after the pump remained the same, and the pump wasn’t required to increase the

pressure of the fluid, because uniform pressure is required. Therefore, the only factor determining

the power consumption of the pump is the frictional losses in the pipes through the loop. Because of

the friction, the water loses energy, and the pump needs to regenerate this energy in order to avoid

Figure 4 - Pedrollo PK-60 pump


32 | P a g e

that the water stops as a result of friction. The friction coefficient was calculated based on the

Reynolds number. It was calculated previously, that the Reynolds number is 2098 for the fluid flow in

the ground, which means that the flow is laminar. The friction coefficient was calculated to be 0.031

accordingly, using the Reynolds number. The head loss of the pump was calculated to be 5.012

meters (Appendix pg. 30) based on the friction coefficient, velocity and length to diameter ratio. The

higher these values are, the higher the losses in the system. This way the best result can be achieved

for a fluid flow that has low friction coefficient, low flow velocity and low length to diameter ratio.

The size of the pump was then chosen based on the head loss of the system and the volumetric flow

rate that the pump needs to provide.

Reynolds number of water flow 127.594

Friction coefficient 0.031

Volumetric flow rate through pump 4*10-3 m3/s

Head loss due to friction 5.012 m Table 17 - Summary table for data presented in the section

Heat Pump Losses The heat pump losses were necessary to calculate in order to calculate the required total

compressor power output. The total compressor output consists mainly of the work used to increase

the enthalpy of the refrigerant at a constant mass flow rate and secondly of the work needed to

overcome the losses in the system. The summation of these two values provide the overall

compressor power output requirement, which was presented in the section Compressor

calculations. In this section, the focus will be on interpreting the losses.

There are two types of losses in the system: frictional losses and minor losses. The minor losses were

not calculated for two reasons: the main reason was the lack of data. The minor losses result mainly

from the bends in our piping, however, it is important to note that these losses depend on a factor

(k) that is only possible to determine using experiments. Because of lack of experimental data, it was

not possible to calculate minor losses in the system.

The frictional losses were, on the other hand, calculated using the data known from calculations.

Similarly to the process of calculating the head losses for the pump, the head loss depends on the

friction coefficient, velocity of fluid flow and length to diameter ratio. The smaller these number are,

the lower the head loss in the system. The friction coefficient depends on the Reynolds number

which shows what type of fluid flow is present in the system, if it’s laminar or turbulent flow. The

losses were calculated for the operation in the winter, because that is when the highest losses occur.

The losses were calculated separately for the fluid flow in the compressor and in the evaporator. The

summation of these two frictional losses resulted in the overall head loss of the system. The fluid

flow in both the evaporator and condenser were considered to be laminar flow based on the

Reynolds number. The friction coefficients were calculated to be 0.019 and 0.032, respectively

(Appendix pg. 19). Knowing the friction coefficient, the dimensions of the plates and the flow

velocity, the frictional head losses were calculated to be 0.05 meter on the indoor heat exchanger

and 0.291 meter on the ground loop heat exchanger.

Then the power loss was calculated based on the head loss in meters. By multiplying the head loss

by the refrigerant density, the gravitational acceleration and the velocity of the refrigerant, the

power loss in Watts was calculated for the heat exchanger at the ground loop to be 462.726 W and

for the indoors heat exchanger to be 69.293 W. The summation of these two losses resulted in a


33 | P a g e

total power loss of 532.019 W (Appendix pg. 19), that the compressor needs to regenerate. These

losses are significantly higher than in the summer, when the total power loss is just 66.5 W.

Friction coefficient in heat exchanger indoors 0.019

Friction coefficient in heat exchanger outdoors 0.032

Head loss in heat exchanger indoors 0.05 m

Head loss in heat exchanger outdoors 0.291 m

Power loss in heat exchanger indoors 69.293 W

Power loss in heat exchanger outdoors 462.726 W

Total power loss in the heat exchanger 532.019 W Table 18 - Summary table for data presented in the section

Coefficient of Performance Calculation The main goal of the heat exchanger design was to improve the Coefficient of Performance value,

which shows how much work the compressor has to deliver with respect to the overall energy

output of the system. The higher the CoP value, the higher the efficiency of the system is, because it

indicates that the heat pump can provide higher output power using the same compressor input

power. This ratio basically shows how efficiently the heat from the ground source is utilized, and

how much the compressor has to work to achieve the required result. The coefficient of

performance also highly depends on the temperature difference the system needs to overcome: the

higher the difference, the lower the coefficient of performance. Therefore, the coefficient of

performance was calculated both for the summer and winter operation, calculating with the highest

possible power consumption in both cases. The coefficient of performance in the summer is 9.628

(Appendix pg. 31) using the maximum power output requirement of the system and the maximum

compressor input power requirement, and in winter the coefficient of performance value was

calculated to be 7.061 using the maximum power output requirement of the system and the

maximum compressor input power requirement. As a result, it can be concluded that the system will

always operate with a Coefficient of Performance value of at least 7.061 (Appendix pg. 19) or higher.

Thus, it can be concluded that this optimised design significantly improves the usual Coefficient of

Performance values for heat pumps, which usually ranges between 3 and 4.

Coefficient of Performance in the summer 9.628

Coefficient of Performance in the winter 7.061 Table 19 - Summary table for data presented in the section

Decision Between Plate Type Heat Exchanger and Pipe Type Heat Exchanger Before deciding on the plate type heat exchanger design for the system, a more traditional heat

exchanger design using pipes was also considered. Instead of the fluid flowing through long, thin

plates to dissipate heat, it would have run through a coil consisting of circular pipes. Even through

the required surface area was reasonable, even lower than that of the plate type heat exchanger,

requiring 7.9 m2 indoors and 10.7 m2 outdoors, the reason for not selecting this design was a result

of the high losses present in the system. As described before, the frictional losses in the system are a

function of the friction coefficient, the length to diameter ratio and the velocity of the flow. When

pipes are used, the length to diameter ratio is large – for practical reasons and in order to provide as

much surface area as possible for the heat transfer, the pipe needs to have a small diameter, but


34 | P a g e

large length to allow heat dissipation. Due to this large length to diameter ratio, the frictional losses

are extremely high: in total, they would add up to 6.9 kW, which is larger than the useful power

output of the compressor to increase the enthalpy of the system. Due to the increased losses, the

coefficient of performance is decreased significantly to merely 2.215 (Appendix pg. 41), which is

lower than that of average heat exchangers. Therefore, it was decided to use plate type heat

exchangers, where this problem is solved thanks to the fact that the length to hydraulic diameter

ratio is significantly lower than that of pipes, because it is possible to have a large hydraulic diameter

while still remaining practical in terms of size and allows efficient heat transfer because of the large

surface area present in a thin plate. This solution also resulted in a significantly larger coefficient of

performance. The table below presents the most important data for comparison of the losses and

CoP values.

Total frictional losses in pipes 6.9 kW

Total frictional losses in plates 532.019 W

Coefficient of Performance of pipes 2.215

Coefficient of Performance of plates 7.061 Table 20 - Properties of pipes

Refrigerant Selection (R134a) The refrigerant selection directly influenced the efficiency of the system through the pressure and

temperature properties, which can be described using the enthalpy. Therefore, from a

thermodynamic point of view, it was crucial to select a refrigerant that provides the highest

efficiency for the system: that means that the compressor work required is as low as possible.

To achieve this, the required temperature in the system should be achieved at as low pressures as

possible, so the compressor’s required power output remains low. The reason behind this can be

explained by the thermodynamic calculations. In the heat exchanger, the required power output

depends on the average temperature difference between the average temperature of the

refrigerant in the heat exchanger and the average temperature of the fluid flowing through the heat

exchanger, and on the total thermal resistance. The total thermal resistance is only influenced by the

heat conductivity of the refrigerant. The compared refrigerants, however, had very similar thermal

conductivity values, which only influenced the thermal resistance to a minor extent. A much larger

difference could be achieved if the temperature difference between the refrigerant and the fluid

outside is large. That mean that by using the same thermal resistance, the larger the temperature

difference is, the larger the power output of the heat exchanger is. This became one of the most

important criteria while selecting a refrigerant.

The other important consideration, which directly affected the coefficient of performance of the

system, was the output power of the compressor. This was determined using the mass flow rate of

the refrigerant and the enthalpy change the compressor needs to create. The mass flow rate is not

directly affected by the refrigerant selection, however, the lower the enthalpy difference that needs

to be created, the lower the output power that the compressor needs to provide. And because only

the temperature part of the enthalpy change is useful for increasing the power output of the heat

exchanger, the goal was to keep the pressure and the pressure difference at as low level as possible,


35 | P a g e

to avoid unnecessary work. The useful work of the enthalpy increase is only the increase in

temperature.

Therefore, a refrigerant was required that is able to create high temperature differences in the heat

exchanger with small pressures and pressures differences. In order to ease this process, a

comparison was made in Excel. Some of the most common refrigerant types were considered,

including R134a, R404a, R407c, R410a, R507. Two given pressures were selected for the comparison,

2 bar and 10 bar. The vapor saturation temperature was written up for all of the above mentioned

refrigerant at these two pressures. For R134a, the temperature of saturated vapor is -10 oC at 2 bars

and 40 oC at 10 bars. (actrol, 2017) For R404a, the temperature of saturated vapor is -30.2 oC at 2

bars and 17.2 oC at 10 bars. For R407c, the temperature of saturated vapor is -21.3 oC at 2 bars and

24.9 oC at 10 bars. For R410a, the temperature of saturated vapor is -37.1 oC at 2 bars and 7.8 oC at

10 bars. For R507, the temperature of saturated vapor is -31.7 oC at 2 bars and 15.8 oC at 10 bars.

(actrol, 2017)

Table 21 - Table describing vapor saturation temperatures at two different pressures

Because the heat exchanger is designed for cold climates, where it needs to work under

temperatures of down to -20 oC, but summer temperatures are not expected to exceed 35 oC, the

primary focus of the designed heat pump is to provide sufficient heating at high efficiency, while it is

still able to cool down the building in the summer. Both the heating and the cooling operation are

required to be efficient, but the heating efficiency is prioritized over the cooling efficiency, because

this operation will consume the most power, and because the target is to work in cold climates, the

heating function can be expected to dominate over the cooling function.

In the table it is visible that using the pressure of 2 bars, most of the refrigerants evaporate at very

low temperatures. While it is advantageous to have a large temperature difference between the

refrigerant and the fluid flow outside, when the system is switched from cooling to heating, at 10 bar

pressure only 2 of these refrigerants would be able to heat up the building: R134a and R407c. For

the other refrigerants, the pressure would need to be increased even further to provide heating for

the building. This, however, would result in larger required enthalpy difference of the refrigerant,

meaning the compressor would need to provide more power output to achieve the same

temperature that other refrigerants can already deliver at lower temperatures.

Because the heating output is the primary focus of the heat exchanger designed for cold climates,

R134a was selected , because it can provide large heating temperature differences (between the

refrigerant vaporizing at 40 degrees C and the room temperature of 21 degrees C) at low pressures,

meaning the compressor doesn’t have to provide as much energy as it would be necessary for other

refrigerants to increase the pressure, a property that is not useful in terms of energy transfer and

mechanics dimensioning. When the pressure is decreased to 2 bar, however, it reaches -10 degrees,

which provides adequate temperature difference in the summer between the evaporator and the


36 | P a g e

room temperature. This means that it can work as efficiently in the summer as in the winter under

these climate conditions. However, the system has lower pressures as it would be required when

using other refrigerants. Because of these advantages, R134a was selected as the refrigerant for the

heat pump. The non-flammability and non-toxicity of the refrigerant are also important advantages,

since the heat pump is designed to operate in private households.

Regulations Regulations for Internal Parts

According to DS/EN 378 (Danig, 2014), which describes the law and regulations regarding heat pumps, several factors must be met before a system can legally be used within the EU. The first factor that was considered was during the design phase, where it was decided that the system would use a direct layout, where the refrigerant is in the piping going inside the house. The rooms to be heated were also determined to be of category A, which indicates that people will stay in the room for longer periods of time. This is determining the requirements for the selected refrigerant. When describing refrigerants, two factors are considered; the flammability and toxicity of the refrigerant. The flammability is rated from 1-3, where 3 is the most flammable, while the level of poisonousness, is determined by the letters A and B, where A is low (>400 ppm) and B is high (<400 ppm). Because of this requirement, the refrigerant selection was narrowed down to a few options since they are all non-toxic and non-flammable: (Danfoss, 2017)

• R134a • R404A • R407C • R410A • R507 • R22

According to thermodynamic calculations in appendix pg. 30, R134a was selected as the most optimal refrigerant, because of its low toxicity, non-flammability, and thermodynamic properties. Another factor which had to be taken into consideration, was the Refrigerant Concentration Level (RCL), which is a value calculated for the room with refrigerant in it. If this value exceeds the Practical Limit for the refrigerant, it is not allowed to be used.

RCL[kgm3]=mrefrigerant[kg] Vroom[m3]

Equation 1 - Calculation of refrigeration concentration level (Danig, 2014)

As can be seen in Equation 1, the RCL factor can be determined by the amount of refrigerant currently in the room at any given moment, and the size of the room. If this value is smaller or equal to the Practical Limit for the selected refrigerant, it is fine for that particular room. The selected refrigerant R134a, has an RCL factor of 50000 PPM, according to (Calm, 2000), which translates to 50 kg/m3. The dimensioning of the system decides that the maximum mass of the refrigerant in the entire system will go up to 70 kg.

50[kgm3]=70[kg] Vroom[m3]→Vroom[m3]=1,4m3

Equation 2- Exact calculations for minimum room size

According to Equation 2, the smallest room volume allowed to contain the system, is 1,4 m3.


37 | P a g e

Operating temperatures in which the system might operate was determined according to DS/EN 378 (Danig, 2014). The system has an air cooled, high pressure side and a low-pressure side with contact to the outside, in form of the ground loop. In the summertime when the system is switched around, the high-pressure side will be at the ground loop heat exchanger side. This also has a designated maximum temperature. The climate zone was determined to be up to 32oC, which is true for the Scandinavian climate. If the temperature where the heat pump is put up rises further than that, a higher climate zone must be used.

Climate zone <32oC <38oC

High pressure side, air cooled 55oC 59oC

Low pressure, contact with outside 32oC 38oC

High pressure, water cooled 32oC 38oC

Table 1 - Maximum operating temperatures according to climate zone (Danig, 2014)

As can be seen on Table 1 (Danig, 2014), temperatures for the high-pressure side must never exceed 55oC, while the low-pressure side cannot exceed 32oC. In the summertime where the high-pressure side has been switched around, the maximum temperature remains the same. This determined the maximum pressures the system may operate at. Properties for the refrigerant R134a (Ohio EDU, 2008), can be taken as an example. In a heating situation, where the temperature can only go up to the point where it is equal to the high-pressure side (55oC). Knowing this, properties for the refrigerant can be looked up, and the maximum allowed pressure can be determined. In the case for R134a, the maximum allowed pressure can be determined to be 1500 kPa, as also seen on Table 2. To put on a safety factor, 1400 kPa can be set as the PS value (maximum allowed pressure for system). When the system is reversed, the maximum allowed temperature is 32oC, which calls for a maximum pressure of 800 kPa. The PS value can be set to 750 kPa.

Pressure Temperature

800 kPa 31,3oC

1500 kPa 55oC Table 2 - Properties for refrigerant R134a (Ohio EDU, 2008)

The PS value is now used to determine the PED component categories for the fluid containers and pipes. It is not needed for compressors and pumps. The main deciding factor of the PED rating is the liquid being used. If it is hazardous in any way, the liquid is categorized in group 1, and the rating must be evaluated further, taking physical sizes of the system into consideration. Group 2 is used for classification of liquids which are completely non-hazardous. In the selected system, where R134a was used, which is considered non-hazardous, hence could be placed in group 2. Being in group 1 it meant that many extra precautions didn’t have to be made. But to be able to stand behind the system, it had to be assembled by a professional, according to DS/EN 378 article 3, §3. Knowing this, it was possible to determine how much documentation the purchased components need to have. Because all parts of the system got a PED rating below category 1, according to DS/EN 378 article 3, §3, no parts inside the machine needed to be CE certified, have traceability documented nor any other kinds of documentation. (Danig, 2014)

Testing of Complete System

Safety measurements include pressure valves according to EN 13136 and a control panel with switches according to EN ISO 13849-1, DS/EN 12263 and IES 60364.


38 | P a g e

Before the system is ready to use, it must go through a series of tests. If the given component isn’t CE certified, it must pass the following criteria:

• 1,43*PS for every component, without permanent deformation • 3*PS for every component, without breakage • Testing for fatigue according to the specifications of DS/EN 378. • Pipes are tested with 1,1*PS

At the place of assembly, the system must pass a constant pressure of 1*PS without any leakage finding place. A visual inspection must be carried out after installation, to see if everything is as it should be visually. Documentation and user manuals should also be produced for the end user. In case of leakage inside inhabited rooms where the coolant is passed through, an oxygen sensor must be installed. Lastly the surface of the heat pump must visibly contain the following information:

• Name and address of manufacturer • Model, series number and reference • Year of production • Type of refrigerant (according to ISO 817) and capacity (kg) • PS, maximum allowed pressure (operating, stand still) • CE certificate (etc.) • Electrical data according to EN standards

Safety Considerations Purely in terms of reachable temperatures, the system presented very little risk to the user. The

highest temperature achievable by the system was approximately 55oC, with average operating

temperatures of between 30 oC and 40 oC. The group’s research suggested that according to

governmental standards a ‘safe’ temperature could be approximated to somewhere in the region of

44oC. It was therefore important to calculate the required material and thickness to sufficiently

insulate these high-temperature areas, to ensure that no harm could come to the user.

The other safety consideration was naturally due to pressure. According to regulations, the pressure

inside the pipes could not exceed 15 bar. The calculations made in response to this resulted in a wall

thickness capable of resisting any spikes in pressure, so as to ensure that no failures occurred within

the above-ground sections of the system while it was in operation.

Further regulations stated that an emergency stop button had to be included somewhere in the

system. Schneider Emergency Button (RS Stock number 330-9388) was selected, due to its low cost.

The button would be mounted between the compressor and its connection to the mains electricity,

where a single press would be able to sever the connection, thereby breaking the circuit and

stopping the heat exchanger instantly.


39 | P a g e

Economics Introduction The economics of this project has had a great deal of importance in the procedure and progress.

When considering how one might go about putting together a device, important questions need to

be asked, such as: What materials should be used for the best economical outcome? This meant that

not only did the mechanical properties have to be taken into consideration, but also the raw

material’s cost, its processing cost and the quality and marketability of the finished product. Other

considerations were directly associated to the company’s financial side, such as budgeting,

establishing a selling price, calculating costs and having an investment and financial consequences

timeline for the customer to consider.

All of this was considered with the interests of UniClimat and their goals concerning this project and

their product. This section of the project will therefore provide UniClimat with a clear target market

and financial plan to confidently proceed with.

Analysis of Survey and Results Primary research was conducted in the form of a survey with many residents of Horsens walking on

the streets, and it generated 20 conclusive results ready for analysis (Appendix pg. 56). The survey

aimed to gain a better understanding of the market and to see if people would be willing to make a

relatively large investment in a heat pump that would replace their current heating system. The

respondents were informed that the investment would generate a payback due to related savings,

but the payback period would depend on how much they were currently paying for their source of

heating, and therefore how much money they would save.

It was hypothesised that not many people would be willing to make the investment due to many

factors, such as the amount of money involved, the long payback period (estimated to be about 10

years), the disruption of installation and the general complacency of people. This hypothesis turned

out to be somewhat true, but it was a pleasant surprise to realize how people were very conscious

about saving energy and switching to green energy sources. (Refer to the “Survey Data” sheet in the

“Heat Pump Economics” excel document to see how the data was collected, sorted and analysed)

Figure 6 - Distribution curve of heat consumption

0

1

2

3

4

5

6

10000 12000 14000 16000 18000 20000 22000 24000 26000

Nu

mb

er o

f R

esp

on

den

ts

Heat Consumption (DKK)

Distribution Curve of Respondents' Heat Consumption Bills

Respondents

Poly. (Respondents)


40 | P a g e

Figure 6 - Distribution curve of heat consumption is a visual representation of the distribution of

respondents according to how much they pay for heat consumption in a year (refer to the “Survey

Data” sheet in the “Heat Pump Economics” excel document to see this data). This graph depicts how

many people fit into a certain price range for their heat consumption bill. It is clear when considering

the polynomial trendline that most of the respondents pay between 16,000 and 20,000 DKK every

year.

Whilst further analysing the data, it was calculated that out of everyone who answered the survey,

they all spent on average 17,700 DKK every year on heating, which is supported by the Distribution

Curve shown in the graph above - as one can see the trendline peaks just below 18,000 DKK. This

data is also complimented by the findings from SparEnergi.dk, which indicates that the average

140m2 house uses about 15,000kWh each year, and therefore calculates to be 22,500 DKK each year

when using the provided rate of 1.5 DKK/kWh (SparEnergi, 2016). It is believed that the 4,000 DKK

difference between the survey’s average yearly consumption and SparEnergi’s data is due to the fact

that the survey only covers a small group of people living in Horsens and not the whole of Denmark.

The graph in Figure 7 depicts the fact that only 20% (“No” - Blue) of our respondents said that they

would definitely not want to buy the heat pump due to how expensive it is and the disruption it may

cause. Whereas the remaining 80% (“Yes and Maybe” – Blue) indicated that the proposal was

interesting and they might be willing to buy the product if the payback period was short and the

amount of disruption during installation was negligible. This information shows that people are in

fact very willing to consider this kind of investment since it offers a green solution to energy

consumption and it also offers to cut down costs on energy consumption.

It was also found that 45% of the respondents would not buy the horizontal heat pump due to the

damages the installation would cause to their garden and 30% said they would definitely still go

through with it, meaning the rest were unsure about how committed they were to their previous

answer that they would like to make the investment. This shows a big shift in the result of this

Figure 7 - Interest in heat pumps

0

2

4

6

8

10

12

Yes Maybe No

Nu

mb

er o

f R

esp

on

den

ts

Response

Comparison of the Respondents' Interest in the Heat Pump with Their Willingness to Ruin Garden

Interested in Heat Pump Willing to Ruin Garden


41 | P a g e

survey, since now it is noted that people are in fact less willing to buy the Heat Pump due to the

disruption the installation may cause.

Given that these results are all from a very small number of respondents, it is quite clear that there

is in fact a potential market for Heat Pumps. It is believed that the market will only continue to grow

stronger as long as the public continues to become more aware of the environment and the effects

that certain energy sources have on it. Therefore, going into the market with a Heat Pump is a great

business opportunity, since the competition is currently very small.

Improvements for Survey Questions It was noted from the survey results that simply asking “What type of heating system do you use?”

was too vague and the answers did not actually provide any helpful information. A better question

to ask would have been “What is your source of heating?”, since this would have generated answers

like; “Electricity”, “Natural Gas”, “Oil”, “Central Heating” or maybe even “heat pump”. Instead, most

respondents said that they use radiators, which can use either central heating, electricity or even a

heat pump.

Asking about garden size turned out to be a very irrelevant question, whereas it would have been

much more beneficial for this project if respondents were asked what was the size of their house.

This would have made it possible to see how heating costs changed in houses of different sizes.

Market Research This section of data is all secondary research that concentrates on the current and potential market

for Heat Pumps, this is necessary information for further analysing whether there is in fact a market

for these devices in Denmark. By using the information provided by respective companies it is

possible to estimate financial values that are relevant for the budgeting and investment calculations,

such values include; expected number of sales, costs of production, salaries, selling price of finished

product, maintenance costs and consumption savings.

The biggest consumer of energy in Danish homes is heating, which makes up about 40% of energy

consumption costs. In 2013 the Danish government at the time decided that it would be in the best

interest of residents to use Central Heating instead of Oil or Natural Gas whenever possible, due to

the high costs and the impact these methods have on the environment (Green Match, 2017). In fact,

from 2013 new houses were no longer allowed to have oil or gas furnaces installed in them, and

since 2016 Oil furnaces have not been allowed to be installed into any house, new or old. These

regulations are an attempt to make people consider new methods of heating that use green energy,

especially in areas where Central Heating is not available. This means that there has been a huge

shift to use clean Electrical Heating, Central Heating and Heat Pumps. Since these three energy

sources are all considered green (depending on your electricity provider), the market is generally in

favour of clean Electrical Energy because it is readily available everywhere and Central Heating

because is a very cheap in comparison to other sources. The market for Heat Pumps is also in

competition with these other sources, but the costs regarding the initial costs needs to be developed

further, since it is known to be a very expensive investment. Apart from this, Heat Pumps will

continue growing in the market as long as people are becoming more conscious about using green

energy (Green Match, 2017).

In Denmark, residents are rewarded for switching from their current heating source to a Heat Pump.

This reward comes in the form of reduced tax - residents can expect a discount of about 0,3

DKK/kWh (GreenMatch, 2017). It is also relevant to note that the tax on electricity is reduced by 0,5


42 | P a g e

DKK/kWh for residents that use electricity as their heating source. Electrical heating is also the most

common method of heating used residents in Denmark. With the cost of electricity being about 2

DKK/kWh depending on supplier, the heating costs for electrical and heat pump heating are 1,5

DKK/kWh and 1,2 DKK/kWh respectively (Apaere, 2015). (Appendix pg. 60).

Method unit/year unit DKK/unit Ratio DKK/m2 m2 DKK/year

Heat Pump 3604 kWh 1,2 30,89 140 4324,8

Central Heating 18,1 MWh 700 90,50 140 12670

Natural Gas 2000 m^3 7,25 103,57 140 14500

Electrical Heating 15000 kWh 1,5 160,71 140 22500

Oil 2300 L 9,5 156,07 140 21850 Table 22 - Heat usage according to source

The table above shows how the respective costs of energy sources compare to each other. The

yearly consumption costs (DKK/year) for the different energy sources are calculated by using the

provided costs per unit consumption (DKK/unit) and multiplying them by the unit consumption per

year (unit/year). The ratio shown in the “DKK/m2”, is a just an approximate value for comparing how

the costs of different heating methods vary, but these costs for heating different size houses

depends on the efficiency of the method used and the number of people living in the house. It is

therefore necessary to assume that the efficient and number of residents is the same for every size

house. This information, makes it possible to see how much money the customer is spending by

using a certain product, and therefore they can also see how they will save by switching to a Heat

Pump.

Although the data shown above clearly indicates that Heat Pumps offer a great saving on the

consumption bill, Heat Pumps are also extremely expensive initially and are therefore not so

desirable. The average Heat Pump costs anywhere around 80,000 – 130,000 DKK not including

Installation costs (Appendix pg. 58), and generally Ground Source Heat Pumps are just 115% of the

price of other Heat Pumps. However, the installation costs of Ground Source Heat Pumps are much

higher than other Heat Pumps and can therefore become up to 200% more expensive (GreenMatch,

2017). This radical increase of price is due to the fact that the installation involves drilling deep into

the ground to install 480 metres of piping in vertical area of about 22m2.

Full Price Example (DKK) 94625 100%

Wholesale Price 44600 47%

Installation 31000 33%

Tax 18925 20% Table 23 - Prices for unit

Table 23 that shows how the costs of buying a Ground Source Heat Pump are distributed for the

average Heat Pump. It clearly indicates that out of the whole sales price (94,625 DKK) the installation

costs make up 33% and tax is at 20%. This information is useful for setting a sales price that the

customer will pay for the product since it provides a tax rate, but is not relevant for the installation

costs of a Ground Source Heat Pump.


43 | P a g e

Budget After analysing the market this product would be competing in, it was time to put this information

into financial terms and therefore make a budget plan for UniClimat. Before making the budget plan

the goals and fixed costs of UniClimat had to be considered, so that a selling price could be

calculated. UniClimat had no need to make any real investment since their staff are made up almost

completely of volunteering students, they are renting their own warehouses and they are already

generating an income in Russia. Therefore, a Cost-Plus Pricing method was used, because Target

Rate of Return was not relevant for this application. This method required an Overhead Absorption

Rate to be calculated with respect to the fixed costs and the expected number of units that would be

sold. UniClimat provided information that suggested that their competitors were all selling a

combined amount of about 150 Heat Pumps in Russia. Therefore, 50 units each year seemed to be a

realistic expectation in this market. (Appendix pg. 62) (Refer to the “Price Setting” sheet in the “Heat

Pump Economics” excel document to see how the selling price of 287843,16 DKK was calculated).

Once the selling price was calculated, a quarterly budget plan was created to help UniClimat plan

their spending and see how much money they will be dealing with on a yearly basis if sales proceed

as estimated (Appendix pg. 62). As the calculations show, UniClimat will have an annual revenue of

11,993,465 DKK and make a total profit of 1,564,365 DKK every year. (Refer to the “Yearly Budget”

sheet in the “Heat Pump Economics” excel document to see how these calculations in detail).

The next step was to put these values into an Investment timeline, and therefore see how

UniClimat’s investment pays-back (Appendix pg. 63). The graph below shows that the Net Present

Value is positive and the discounted payback period is 3.58 years. This information indicates that if

sales take place as expected, UniClimat will greatly benefit from this project.

Figure 8 - Payback period

Investment and Financial Consequences This section of the project focusses on what the investment portfolio looks like for the costumer.

This gives the customer more tangible information to help them decide whether the initial

investment of buying the Heat Pump is worth going through with in order to save more money in

-6000000,00

-4000000,00

-2000000,00

0,00

2000000,00

4000000,00

6000000,00

0 1 2 3 4 5 6 7 8 9 10 11

Pay

bac

k (D

KK

)

Payback Period (DKK)

Payback Period for UniClimat's Investment


44 | P a g e

later years. The data considers the selling price of the device and the installation costs as the Capitol

Investment, whereas the Cash Flow is made up of the money saved on heating consumption and the

cost of the energy required to operate the Heat Pump.

This investment calculation was not only performed on the ground source heat pump, but also

considers the related cash flows for conventional heating systems, such as; Natural Gas, Electrical

Heating, Central Heating and Oil. With respect to consumption costs, the Heat Pump is considerably

cheaper to operate, and therefore offers a great opportunity for users to make a saving by switching

to a heat pump (Appendix pg. 64-93).

The table below shows how the yearly consumption costs for each year were calculated using the

data from SparEnergi.dk (SparEnergi, 2016). This data is for the average 140m2 house.

Consumption of Different Energy Sources

Method unit/year unit DKK/unit Ratio DKK/m2

m2 DKK/year

Heat Pump 3604 kWh 1,2 30,89 140 4324,80

Central Heating 18,1 MWh 700 90,50 140 12670,00

Natural Gas 2000 m^3 7,25 103,57 140 14500,00

Electrical Heating 15000 kWh 1,5 160,71 140 22500,00

Oil 2300 L 9,5 156,07 140 21850,00

Table 24 - Consumption of energy sources

The data above is used to calculate the yearly savings residents would make by having a Heat Pump.

The investment timelines are calculated for houses of 100m2 up to 200m2 in order to clarify exactly

who fits into the target market of this project.

-200000,00

-150000,00

-100000,00

-50000,00

0,00

50000,00

100000,00

150000,00

100 120 140 160 180 200

(DK

K)

Size of House (m2)

Comparison of Consumption Costs and NPV for Different Size Houses (Electrical)

Heat Pump Eletrical NPV

Figure 9 - Comparison of cost and NPV for houses


45 | P a g e

The graph above is representation of the different NPVs for different size houses and clearly shows

the customer that the Heat Pump is much more efficient and offers a better outcome in larger

houses. This is further supported by the graph below depicting the payback period for different size

houses that would switch from electrical heating source to heat pump source. (Refer to the 4

different “Investment ###” sheets in the “Heat Pump Economics” excel document to see precise

calculations regarding these investments).

The NPV and Payback Period graphs of the Electrical Heating calculations were chosen to be

displayed since they depict the best outcome for these calculations (refer to the “Graphs” sheet in

the “Heat Pump Economics” excel document to see details on how the other energy sources

compared in these calculations). Although the larger houses offer a better outcome, by observing

the graphs below it is very clear that no matter which energy source a resident was previously using,

they will never generate a Payback or a positive NPV.

-350000

-300000

-250000

-200000

-150000

-100000

-50000

0

0 5 10 15 20 25 30

Pay

bac

k (D

KK

)

Payback Period (years)

Comparison of Investment Payback for All Energy Sources (200m2)

Central Heating

Natural Gas

Oil Furnace

Electrical Heating

-350000

-300000

-250000

-200000

-150000

-100000

-50000

0

0 5 10 15 20 25 30

Pay

bac

k (D

KK

)

Payback Period (years)

Payback Periods for Different Size Houses (Electrical)

100m2

120m2

140m2

160m2

180m2

200m2

Figure 10 - Payback period

Figure 11 - Investment for all energy sources


46 | P a g e

Unfortunately, these graphs are clearly indicating that there is in fact no target market to be justified

with the calculations made for this product. This downfall in the project is completely due to the

installation costs of a Ground Source Heat Pump being so high. As seen in the graph below, the unit

alone (SEP4 Heat Pump) is reasonably priced in comparison to normal Heat Pumps, but it is the

installation costs that cause the selling the price to over double. All this indicates is that Ground

Source Heat Pumps are better suited for larger buildings complexes, since their cost efficiency

increases in relation to increasing area of buildings (Appendix pg. 94-96). Therefore, it would be

unadvisable for potential customers to buy this product if they’re aiming for personal, but it is still a

source of clean energy and therefore environmentally beneficial.

-300000,00

-250000,00

-200000,00

-150000,00

-100000,00

-50000,00

0,00

100 120 140 160 180 200

Aks

etit

el

Aksetitel

Comparison of NPVs for the Different Energy Sources in Different Size Houses

Central Heating

Natural Gas

Oil Furnace

Electrical Heating

0 50000 100000 150000 200000 250000

DAIKIN ALTHERMA 8.43 kW

Mitsubishi ECO-DAN W50

Panasonic Aquarea T-CAP

Nibe F2030

Bosch 10 AW

SEP4 Heat Pump

Selling Price (DKK)

Dif

fere

nt

Hea

t P

um

ps

Price Comparsion for Different Heat Pumps

Selling Price Including Installation Selling Price of Unit

Figure 12 - NPV for energy sources

Figure 13 - Price for heat pumps


47 | P a g e

Conclusion Considering the results from the economics calculations it has become quite clear that it was not

possible to find a target market with the set building size limitations. As indicated before, the best

economical outcome was the one where residents switched from Electrical Heating to the Ground

Source Heat Pump, but still did not generate an outcome that would benefit the potential costumer.

And therefore, with no costumers, UniClimat would not make many sales in the market of houses

sized between 100m2 and 200m2. But, there is still potential for a market in office blocks,

apartments, and even multiple houses, since they cover a much larger area. This means that the

Ground Source Heat Pump created in this project would, therefore, have to be reconsidered and

multiple changes would be made in order to upscale the system.

Although the factor restricting the success of this project was the installation cost, it is also seen as a

positive factor in the solution of marketing to larger buildings. This is because when creating a

Ground Source Heat Pump for larger buildings the cost of installation are estimated to change a

negligible amount and only the direct material costs will be changed considerably. This change will

be very small in comparison to the change of the building area, and therefore means that there is

potential for significant improvement in economical outcome of the calculations performed for the

current market data.

Mechanically speaking, the project can be considered a success. The system fulfils its purpose, and

would theoretically provide adequate heating to the target market. It is significantly more energy

and cost efficient than many of the competitors over the long term and could, with a small amount

of redesigning, potentially become a truly marketable product.

Bibliography actrol. (2017, 5 28). Refrigerant technical data. Retrieved from actrol:

http://www.actrol.com.au//Global/Assets/Howto/forane-R134a-technical-data.pdf

Apaere. (2015). Hvad Koster 1 kWh. Retrieved from Apaere: http://www.apaere.dk/hvad-koster-1-

kwh/

Ashraf, R. (2014). Retrieved from http://textileinsight.blogspot.dk/2014/08/advantages-and-

disadvantages-of.html

Geokiss. (2017, 5 29). Guidlines of Tennessee Valley Authority. Retrieved from Geokiss:

http://www.geokiss.com/res-design/GSHPDesignRec2.pdf

Geothermal Pros and Cons. (2012). Retrieved from http://geothermalprosandcons.net/geothermal-

pipe-what-to-use-and-what-to-stay-away-from/

Green Match. (2017). Opvarmning. Retrieved from Greenmatch:

http://www.greenmatch.dk/opvarmning

GreenMatch. (2017, March 21). Hvad er Prisen på en Luft til Vand Varmepumpe. Retrieved from

Varmepump Tilbud: http://www.varmepumpetilbud.dk/blog/hvad-er-prisen-paa-en-luft-til-

vand-varmepumpe

Jpgreene. (2000). Retrieved from

http://www.csuchico.edu/~jpgreene/itec041/m41_ch06/sld022.htm


48 | P a g e

SparEnergi. (2016). Dit Varmeforbrug. Retrieved from SparEnergi:

http://sparenergi.dk/forbruger/varme/dit-varmeforbrug

Dong Energy (2017), Varmeforbrug [online] DongEnergy. Available at:

https://www.dongenergy.dk/privat/f%C3%A5-en-lavere-energiregning/tjek-dit-

energiforbrug/gennemsnitsforbrug/varmeforbrug

HSE (2012), Managing the risks from hot water and surfaces in health and surfaces in health and

social care [online]. Available at: http://www.hse.gov.uk/pubns/hsis6.pdf

Dong Energy (2017), Energiregning [online]. Available at:

https://www.dongenergy.dk/privat/f%C3%A5-en-lavere-energiregning/tjek-dit-

energiforbrug/gennemsnitsforbrug/varmeforbrug

Actrol sources for other refrigerants

http://www.actrol.com.au//Global/Assets/Howto/forane-R507-technical-data.pdf

http://www.actrol.com.au//Global/Assets/Howto/forane-R410A-technical-data.pdf

http://www.actrol.com.au//Global/Assets/Howto/forane-R407C-technical-data.pdf

http://www.actrol.com.au//Global/Assets/Howto/forane-R404A-technical-data.pdf

SEP4 Project Report Heat Pump · 2017-11-15 · SEP4 PROJECT REPORT HEAT PUMP Tamas Pinter...

Documents

Transcript of SEP4 Project Report Heat Pump · 2017-11-15 · SEP4 PROJECT REPORT HEAT PUMP Tamas Pinter...